Proceedings of the 28th International Cryogenic Engineering Conference and International Cryogenic Materials Conference 2022: ICEC28-ICMC 2022, ... in Science and Technology in China, 70) [1st ed. 2023] 9819961270, 9789819961276

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Advanced Topics in Science and Technology in China

Limin Qiu Kai Wang Yanwei Ma  Editors

Proceedings of the 28th International Cryogenic Engineering Conference and International Cryogenic Materials Conference 2022 ICEC28-ICMC 2022, Hangzhou, China

Advanced Topics in Science and Technology in China

70

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Limin Qiu · Kai Wang · Yanwei Ma Editors

Proceedings of the 28th International Cryogenic Engineering Conference and International Cryogenic Materials Conference 2022 ICEC28-ICMC 2022, Hangzhou, China

Editors Limin Qiu Institute of Refrigeration and Cryogenics Zhejiang University Hangzhou, China

Kai Wang Institute of Refrigeration and Cryogenics Zhejiang University Hangzhou, China

Yanwei Ma Institute of Electrical Engineering Chinese Academy of Sciences Beijing, China

ISSN 1995-6819 ISSN 1995-6827 (electronic) Advanced Topics in Science and Technology in China ISBN 978-981-99-6127-6 ISBN 978-981-99-6128-3 (eBook) https://doi.org/10.1007/978-981-99-6128-3 Jointly published with Zhejiang University Press The print edition is not for sale in China Mainland. Customers from China Mainland please order the print book from Zhejiang University Press © Zhejiang University Press 2023 This work is subject to copyright. All rights are reserved by the Publishers, whether the whole or part of the material is concerned, specifically the rights of reprinting, reuse of illustrations, recitation, broadcasting, reproduction on microfilms or in any other physical way, and transmission or information storage and retrieval, electronic adaptation, computer software, or by similar or dissimilar methodology now known or hereafter developed. The use of general descriptive names, registered names, trademarks, service marks, etc. in this publication does not imply, even in the absence of a specific statement, that such names are exempt from the relevant protective laws and regulations and therefore free for general use. The publishers, the authors, and the editors are safe to assume that the advice and information in this book are believed to be true and accurate at the date of publication. Neither the publishers nor the authors or the editors give a warranty, expressed or implied, with respect to the material contained herein or for any errors or omissions that may have been made. The publishers remain neutral with regard to jurisdictional claims in published maps and institutional affiliations. This Springer imprint is published by the registered company Springer Nature Singapore Pte Ltd. The registered company address is: 152 Beach Road, #21-01/04 Gateway East, Singapore 189721, Singapore Paper in this product is recyclable.

Preface

The 28th International Cryogenic Engineering Conference and International Cryogenic Materials Conference 2022 (ICEC28-ICMC 2022) was successfully held from 25 to 29 April 2022 on Yaotai® as an immersive virtual event like “metaverse”. The ICEC-ICMC series conference is one of the largest and most influential international conferences in the cryogenics field held every two years alternately between Europe and Asia. The conference was originally scheduled to be held offline by Zhejiang University in Hangzhou, China, in the early autumn of 2020, but was cancelled due to the outbreak of the COVID19 pandemic and finally postponed to this year as a virtual conference. After four years of waiting, the international cryogenics community finally had an opportunity to spend a valuable week together to share information, news and ideas on cryogenic engineering and materials. The conference attracted more than 300 attendees from nearly 20 countries, and the scale is comparable to previous offline conferences. A total of 1 short course, 5 plenary talks, 20 oral sessions and 22 poster sessions were arranged during the conference. 229 papers in total were presented, including 114 oral and 115 poster presentations. Among them, a total number of 148 papers were finally published in the proceedings. The topics covered a wide range of focuses in cryogenic engineering and materials, e.g. large-scale refrigeration and liquefaction, cryogenic components/systems/facilities/tests, cryogenic heat transfer and thermal insulation, cryocoolers, space cryogenics, cryogenics for superconducting materials/devices/systems, superconducting materials and devices, etc. Despite the form of virtual event, the conference attracted wide interest and great enthusiasm from industry, with 17 cryogenics-related companies from across the globe as the sponsors and exhibitors. As the first group of people who dared to eat the “crab” of the immersive international academic conference, we found a way to integrate the virtuality and reality together for a brand-new meeting experience, and a long list of praises from worldwide attendees were given. As the first virtual event in the ICEC-ICMC history, the organizers, attendees and sponsors made this conference indeed a full success.

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Preface

Group photo of the local organizers of the ICEC28-ICMC 2022 Limin Qiu

Organization

International Cryogenic Engineering Committee Milind Atrey Lars Blum Marcel ter Brake John Brisson Maciej Chorowski Pascale Dauguet Carlo Ferdeghini Alain Girard Sangkwon Jeong Takanobu Kisu Ercang Luo Antonio Morandi Hirotaka Nakai Holger Neumann Limin Qiu Mike Sumption Laurent Tavian John Vandore Vitaly Vysotsky John Weisend Liye Xiao Yifeng Yang

Indian Institute of Technology Bombay Linde-Kryotechnik University of Twente Massachusetts Institute of Technology Wrocław University of Technology Air Liquide Advanced Technologies CNR-SPIN CEA Grenoble, Laboratory of Cryogenic Engineering (SBT) Korea Advanced Institute of Science and Technology Kyushu University Technical Institute of Physics & Chemistry, CAS University of Bologna KEK Karlsruhe Institute of Technology Zhejiang University The Ohio State University CERN Cryox Russian Scientific R&D Cable Institute European Spallation Source Institute of Electrical Engineering, CAS Southampton University

International Cryogenic Materials Committee David A. Cardwell Takanobu Kiss David Evans K. Theodore Hartwig Eric E. Hellstrom Juan Knaster Richard P. Reed

University of Cambridge Kyushu University University of Bristol Texas A&M University Florida State University Fusion for Energy, ITER Cryogenic Materials, Inc.

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Organization

Kenichi Sato Sonja Schlachter Michael D. Sumption Herman H. J. ten Kate Bob Walsh Yanwei Ma Balu Balachandran Hyung Seop Shin Jun-ichi Shimoyama Timothy Haugan Klaus Peter Weiss

Sumitomo Electric Industries, Ltd. Karlsruhe Institute of Technology Ohio State University CERN National High Magnetic Field Laboratory Institute of Electrical Engineering, CAS Argonne National Laboratory Andong National University Aoyama Gakuin University AFRL/RQQM Karlsruhe Institute for Technology

Organizing Committee

Limin Qiu Zhejiang University

Yuan Zhou Technical Institute of Physics & Chemistry, CAS

Organization

Yanzhong Li Xi’an Jiaotong University

Shuiming Shu Huazhong University of Science and Technology

Guobang Chen Zhejiang University

Ercang Luo Technical Institute of Physics & Chemistry, CAS

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Organization

Yu Hou Xi’an Jiaotong University

Tao Jin Zhejiang University

Maoqiong Gong Technical Institute of Physics & Chemistry, CAS

Peng Zhang Shanghai Jiaotong University

Organization

Xiaoqin Zhi Zhejiang University

Liye Xiao Institute of Electrical Engineering, CAS

Yonghua Huang Shanghai Jiaotong University

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Organization

Programme Committee

Kai Wang Zhejiang University

A. T. A. M. de Waele Eindhoven University of Technology

Ho-Myung Chang Hong Ik University

Organization

Laurent Tavian CERN

Xiaobin Zhang Zhejiang University

Holger Neumann Karlsruhe Institute of Technology

Tripti Sekhar Datta Indian Institute of Technology, Kharagpur

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Organization

Yanwei Ma Institute of Electrical Engineering, CAS

Marcel ter Brake University of Twente

Kun Liang University of Sussex

S. Mostafa Ghiaasiaan Georgia Institute of Technology

Organization

Michael D. Sumption Ohio State University

John Pfotenhauer University of Wisconsin-Madison

Wei Guo Florida State University

Takanobu Kiss Kyushu University

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Organization

Awards and Prizes Mendelssohn Award The Mendelssohn Award was established in memory of Kurt Mendelssohn (1906– 1982), the founder of the International Cryogenic Engineering Committee. Persons to be honoured by the Mendelssohn Award are selected by their work in cryogenic engineering: – New and promising solutions to difficult problems. – Promotion and encouragement of work in new fields of low-temperature applications, for stimulating the cryogenic community’s interest in these fields and helping to establish them. – Long-standing contributions to cryogenics. As the ICEC28 was postponed from 2020 to 2022, we have two Mendelssohn awardees:

Dr. Guy Gistau Baguer 2020 Mendelssohn Award winner

Dr. Ray Radebaugh 2022 Mendelssohn Award winner

Organization

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Gustav and Ingrid Klipping Award The International Cryogenic Engineering Committee awards a young researcher for outstanding work in cryogenic engineering. The award is named after Gustav and Ingrid Klipping to commemorate their enormous contributions to the field of cryogenics and more specifically to recognize their active role in involving young researchers. As the ICEC28 was postponed from 2020 to 2022, we have two Gustav and Ingrid Klipping awardees:

Dr. Andrew May 2020 Gustav and Ingrid Klipping Award winner

Dr. Shiran Bao 2022 Gustav and Ingrid Klipping Award winner

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Organization

Sponsors and Exhibitions Platinum Sponsors

Gold Sponsors

Silver Sponsors

Cloud Service

Organization

Conference Highlights

Plenary talks at the Theatre Hall

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Organization

Oral presentations at the Lecture Rooms

Organization

Posters and sponsor exhibitions at the Exhibition Halls

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Contents

Plenary Helium Refrigeration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Guy Gistau Baguer

3

Thermodynamics and Heat Transfer in Cryogenics: A Personal View . . . . . . . . Ray Radebaugh

16

Large-Scale Refrigeration and Liquefaction Design and Preliminary Test of a Helium Cryogenic Refrigerator for the High Energy FRagment Separator of HIAF . . . . . . . . . . . . . . . . . . . . . . . . Shaoqi Yang, Wei Pan, Rui Xue, Gang Zhou, Dongsheng Ni, Xiujuan Xie, and Wei Wu Operation of the Upgraded Cryogenics Infrastructure of SM18, CERN Main Facility for Testing Superconducting Magnets, Power Links, and RF Cavities . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Nicolas Guillotin, Thierry Dupont, Frederic Ferrand, and Antonio Perin The Design Consideration and Optimization for the CEPC Cryogenic System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Mei Li, Rui Ge, Shaopeng Li, Ruixiong Han, Miaofu Xu, and Zhengze Chang Safety Relief System Design for the DALS Test Facility Cryogenic System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xu Shi, Zheng Sun, Bihui Lai, Liangbing Hu, Guanglong Cui, Lei Xu, and Xilong Wang Upgrade of the Ex-LEP Helium Refrigerator for HL-LHC at CERN LHC Point 4 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Emmanuel Monneret, Serge Claudet, Vanessa Gahier, and Robert Herrmann Analysis of the Optimal Operating Temperature of Dalian Advanced Light Source . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zheng Sun, Xu Shi, Lei Xu, Liangbin Hu, and Xilong Wang

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Key Design Parameters for Improving Oil Removal in Helium Compressors Stations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Krzysztof Brodzinski, Vanessa Gahier, and Udo Wagner Investigation on Dry Vacuum Pumps Suitable for Superfluid Helium Cryogenic Refrigeration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xiujuan Xie, Shaoqi Yang, Xiangdong Xu, Wei Pan, Rui Xue, and Linghui Gong Overview of the Maintenance and Consolidation Activities for CERN Cryogenics During LHC Long Shutdown 2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . V. Iuga, N. Bonetti, K. Brodzinski, F. Ferrand, C. Fluder, L. Herblin, B. Ivens, S. Junker, A. Perin, O. Pirotte, and M. Pezzetti Commissioning of the Large-Scale 2 K Helium Refrigeration System at ESS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . J. Zhang, P. Arnold, N. Kolev, and A. Rueegge Numerical Simulation on Flow Resistance of Helium Cryogenic Transfer Lines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sheng Xu, Mengjia Chen, Lei Zhang, Jiuce Sun, Yuzhe Lin, Zhengrong Ouyang, and Shaowei Zhu Comparison and Analysis of Two Hydrogen Liquefaction Processes Based on Helium Expansion Cycle Integrating with Mixed Refrigerant Pre-cooling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Yujing Bi and Yonglin Ju Flow Process Analysis of the Cryogenic Distribution System for S3 FEL Project . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xinbo Dong, Zheng Sun, Liangbing Hu, Shen He, Bihui Lai, and Xilong Wang Fabrication, Cryogenic Test and Installation of the ESS Cryogenic Moderator System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . H. Tatsumoto, Y. Beßler, E. Rosenthal, P. Arnold, M. Kickulies, M. Boros, A. Horvath, M. Segerup, P. Tereszkowski, and D. Lyngh The Economic Analysis of Helium Liquefaction Plant and Helium Recovery System After Improvement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Li Zhun, OuYang Zhengrong, Luo Yuchen, Zhang Xuehua, and Wang Zezhang

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Contents

Analysis of Large-Scale Energy Storage Technology for Renewable Energy Based on Liquid Hydrogen . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Ming He, Han Zhou, Cui Lv, Wenhui Cui, Jihao Wu, Meimei Zhang, and Linghui Gong Performance Investigation of the Cryogenic Packed Bed Regenerator with Different Parameters for Liquid Air Energy Storage Systems . . . . . . . . . . . Wei Ji, Luna Guo, Xiaoyu Fan, Jiyun Liu, Liubiao Chen, and Junjie Wang Study on Flow Equalization in Solid Phase Packed Bed Regenerator of Liquid Air Energy Storage Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xiaoyu Fan, Wei Ji, Luna Guo, Jiyun Liu, Liubiao Chen, and Junjie Wang Preparation and Preliminary Test on the Integrated Test of the Accelerator Cryoplant and the Cryogenic Distribution System at ESS . . . . . . . . . . . . . . . . . . . J. Zhang, P. Arnold, J. Fydrych, and J. G. Weisend II Development of a 10 TPD Hydrogen-Refrigerated Hydrogen Liquefier . . . . . . . Jing Li, Gang Zhou, Liqiang Liu, Kun Yang, Lianyou Xiong, Jihao Wu, Hailing Qin, and Linghui Gong Investigation on Performance of Large-Scale Hydrogen Liquefiers Based on Collins and Claude Cycles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Rui Xue, Shaoqi Yang, Xiujuan Xie, Ningning Xie, Wei Wu, Kunyin Li, Guoqiang Shen, and Linghui Gong Design, Analysis and Optimization of Refrigerant Cycle in Cryo-compressed Hydrogen Storage Process . . . . . . . . . . . . . . . . . . . . . . . . . . . Jingyao Yang, Haocheng Wang, Xueqiang Dong, Yanxing Zhao, Maoqiong Gong, and Jun Shen Study of Pressure and Temperature Fluctuations Applied to the ESS Cryogenic Moderator System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . H. Tatsumoto, P. Arnold, M. Segerup, P. Tereszkowski, Y. Beßler, and D. Lyngh

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Cryogenic Components, Systems, Facilities, and Testing Multi-field Coupling Rotor Characteristics Analysis for High Speed Turbo Expander Generator Brake . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Shun Qiu, Nan Peng, Liangwei Zheng, Han Yan, Changlei Ke, Kongrong Li, Xiaohua Zhang, Bing Dong, and Liqiang Liu

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Contents

Valves for Helium Cryogenics: Short Review and Experience on Installation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sergiy Putselyk Design, Fabrication and Installation of the Cryogenic Distribution System for Re-Configured FRIB A1900 Fragment Separator . . . . . . . . . . . . . . . . Nusair Hasan, Mathew Wright, Venkatarao Ganni, Fabio Casagrande, Shelly Jones, Brandon Laumer, Chinh Nguyen, Adam Fila, and Nathan Joseph Sub-Kelvin Cryogenics for Experimental Cosmology . . . . . . . . . . . . . . . . . . . . . . Andrew J. May Commissioning and Operational Experience from FRIB Target and Fragment Pre-separator Superconducting Magnet Quench Management System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Nusair Hasan, Venkatarao Ganni, Peter Knudsen, Fabio Casagrande, Brandon Laumer, and Stuart Bott Design and Test Results of the Vacuum Barrier for SHINE Linac . . . . . . . . . . . . Yawei Huang, Zhitao Yao, Yuefeng Liu, Yanfei Zhai, and Yiyong Liu Conceptual Design of the Helium Guard System for the Shenzhen Superconducting Soft X-ray Free Electron Laser Test Facility Cryoplant . . . . . . Guanglong Cui, Bihui Lai, Xingzhong Sun, Xu Shi, Liangbing Hu, and Xilong Wang Theoretical and Experimental Study for Static Evaporation Rate of a Self-Developed Liquid Hydrogen Storage Tank . . . . . . . . . . . . . . . . . . . . . . . Yuanxin He, Zhenyan Xiong, Bo Wang, Jiao Yuan, Humin Wu, Honghao Xu, Haoren Wang, Xian Shen, Weiming Zhou, and Zhihua Gan Study on Energy-Saving Vacuum Elastic Cold Trap Based on Small Stirling Refrigeration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Haiyue Pei, Jiaxu Xia, Bo Wang, Yujie Lin, Dongli Liu, Qinyu Zhao, and Zhihua Gan Remote Cooling System Study Based on a Two-Stage Cryocooler with a Helium Circulation Loop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Aleksandra Onufrena, Torsten Koettig, Boyan Naydenov, Johan Bremer, Thierry Tirolien, and Marcel ter Brake

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Contents

Very Compact and Mobile Cryostat for Tests of Components at 20–450 K Temperature Range . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Matthias Fischer, Karl-Heinz Kulmer, Stefan Moldenhauer, Rainer Puchleitner, Sergiy Putselyk, Daniel Salleger, and Stefan Zink Analysis of the Application of Uncooled and Cooled Infrared Detection Technology in Intelligent Vehicles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Yubo Wang, Xinqi Fang, Minchen Zhu, Qinyu Zhao, Bo Wang, and Zhihua Gan

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Preliminary Design of the BEPCII-Upgrade Cryogenic System . . . . . . . . . . . . . . Ruixiong Han, Liangrui Sun, Miaofu Xu, Jiehao Zhang, Xiaochen Yang, Minjing Sang, Rui Ye, Zhuo Zhang, Penghui Li, Xiangzhen Zhang, Tongxian Zhao, Zhengze Chang, Mei Li, Yongcheng Jiang, Changcheng Ma, Rui Ge, and Shaopeng Li

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The Design of a High-Pressure Helium Supply System at 20K . . . . . . . . . . . . . . Shiyong Xie, Jiasen Wang, Baoxi Lyu, Jixiang Yan, Dong Xu, and Laifeng Li

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Conceptual Design of Cryogenic System for HFRS of HIAF Project . . . . . . . . . Dongsheng Ni, Zhengrong Ouyang, Lishi Wang, Yue Cheng, Xudong Wang, Li Zhu, Xiaowei Xu, Shaoqi Yang, Xiujuan Xie, Enming Mei, Xiangqi Qin, Wei You, Beimin Wu, Wei Wu, Qinggao Yao, Lina Sheng, and Shaofei Han

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Numerical Simulation of 1.8 K–300 K Sub-Atmospheric Helium Heating Device . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Pengcheng Yang, Qiyong Zhang, Zhigang Zhu, Keping Wu, Yiwen Zong, Zhongyu Zou, and Chengfei Fan Conceptual Design of Helium Recovery Management for S3 FEL . . . . . . . . . . . . Sheng He, Liangbing Hu, Bihui Lai, Xinbo Dong, and Xilong Wang Engineering Design and Development of Cold Box for CRAFT [email protected] Helium Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Shanshan Li, Chuanjia Zhang, Zhigang Zhu, Qiyong Zhang, and Zhu Ping Study on Performance of One High Speed Liquid Hydrogen Submerged Pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Cui Lv, Ziwei Li, Haosu Wang, He Su, Jinzhen Wang, Jin Shang, and Jihao Wu

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Bibliometric Analysis on the Application of Magnetic Bearing in the Field of Cryogenics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sihao Liu, Eric Knauss, Shuyue Zhang, Jin Shang, Jinzhen Wang, He Su, Huaiyu Chen, and Jihao Wu Effect of Tip Clearance on the Performance of Helium Turbine Expander . . . . . Chengfei Fan, Shixiong Chen, Qiyong Zhang, Bao Fu, Shanshan Li, Pengcheng Yang, and Yiwen Zong Performance Analysis of Helium Lubricated Hydrostatic Journal Bearing for Cryogenic Turbine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Shixiong Chen, Qiyong Zhang, Jiefeng Wu, Bao Fu, Shanshan Li, and Chengfei Fan Finite Element Analysis on Scroll of a Scroll Compressor for Electric-Driven Vehicle Air-Conditioner . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Luo Yuchen, Li Zhun, and Li Qiang Thermodynamic Analysis of Single-Stage Mixed-Refrigerant Joule-Thomson Cycle at Cooling Temperature from 100 to 180 K . . . . . . . . . . . Yaxue Wei, Jinxing Wu, Qinglu Song, Dechang Wang, Dandan Sun, and Haocheng Wang A Cryogenic System for Measuring the Thermally Stimulated Depolarization Current . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Jixiang Yan, Rongjin Huang, Peng Jia, Huiming Liu, Yaran Shi, Shiyong Xie, Laifeng Li, and Yuan Zhou Design and Analysis of Vacuum Calibration Apparatus at Cryogenic Temperature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Shixian Wu, Zhengrong Ouyang, Lei Shi, Qiumin Meng, Xin Ai, Xuheng Chen, Dazhi Kuang, Peizhi Ding, and junjie Li

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Cryogenic Heat Transfer and Thermal Insulation Vacuum Break in a Helium Cooled Tube with an Inserted Cavity . . . . . . . . . . . . Nathaniel Garceau, Shiran Bao, and Wei Guo Numerical Study on the Flow and Heat Transfer Characteristics in CFETR Cryostat During a Loss of Vacuum . . . . . . . . . . . . . . . . . . . . . . . . . . . . Jianhong Huang, Jianjian Wei, Jian Ge, Sumei Liu, Yuntao Song, and Tao Jin

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Optimization of Multilayer Insulation and Vapor-Cooled Shield Combination Coupled with a Para-Ortho Hydrogen Converter . . . . . . . . . . . . . . . Chuiju Meng, Xujin Qin, Wenbing Jiang, and Yonghua Huang

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Modeling and Parametric Study of Ice Interface Growth During Microdroplet Impinging on Different Micro-Textures . . . . . . . . . . . . . . . . . . . . . . Xiaoqing Zhou, Guang Yang, Chunyu Li, and Jingyi Wu

435

The Effect of Fill Ratio on the Performance and Flow Regime for Long-Distance Helium Pulsating Heat Pipes . . . . . . . . . . . . . . . . . . . . . . . . . . . Logan Kossel, John Pfotenhauer, and Franklin Miller

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A Novel Cryogenic Loop Heat Pipe Structure and Preliminary Proof of the Concept . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chenyang Zhao, Nanxi Li, Zhenhua Jiang, and Yinong Wu

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Thermal Performance of a 140–200 K Grooved Heat Pipe Under Different Orientations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Nanxi Li, Yongyan Li, Wenyi Zhao, Shiyue Wang, Zhenhua Jiang, and Yinong Wu

456

Investigation on the Thermophysical Process of the Non-vented Storage of Liquid Xenon . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xiuli Wang, Zhuoqun Lei, Yonglin Ju, Yu Chen, and Jianglai Liu

463

Results of a Nitrogen Pulsating Heat Pipe with Subsections in a Series Configuration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zhiyi Jiang, John Pfotenhauer, and Franklin Miller

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Numerical Investigation on the Condensation of Various Refrigerants Outside Horizontal Plain and Low Finned Tubes at Low Temperatures . . . . . . . Shu Li and Yonglin Ju

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Experimental and Theoretical Study of Variable Density Multilayer Insulation (VDMLI) at Different Cold Boundary Temperatures . . . . . . . . . . . . . . Tiantian Xiao, Huiming Liu, Yuchen Zhao, Laifeng Li, and Yuan Zhou

487

Effect of Fill Ratios on the Heat Transfer Performance of Nitrogen Cryogenic Pulsating Heat Pipe . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Yaran Shi, Dong Xu, Bingkun Lyu, Jixiang Yan, Tao Wang, Zhixiong Wu, and Laifeng Li Numerical Study of a Single-Loop Nitrogen Pulsating Heat Pipe . . . . . . . . . . . . Bingkun Lyu, Dong Xu, Wei Wang, Yaran Shi, Chuanjun Huang, Rongjin Huang, and Laifeng Li

494

501

xxx

Contents

Modelling and Experimental Evaluation of the Thermal Budget of the UKRI STFC Daresbury Laboratory Vertical Test Facility . . . . . . . . . . . . . Alastair White, Ayomikun Akintola, Keith Dumbell, Shuisheng He, Sean Hitchen, Conor Jenkins, Bo Liu, Andrew J. May, Shrikant Pattalwar, Mark Pendleton, and Paul A. Smith Study on the Increased Enthalpy of Loss Product on Heat Leak into Cryogenic Vessels . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zheng-qing Li, Sheng-sheng Yang, Xian-hu Han, Yu-hong Cai, Xiao-jin Li, Xiao-xia Li, Min Ma, and Xiao-jun Wang

508

515

Coupling Performance Analysis of Plate-Fin Heat Exchanger Filled with Catalyst for Hydrogen Liquefaction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Pan Xu, Aimin Zhou, Jian Wen, Simin Wang, and Yanzhong Li

522

Study on the Impact of Cold Storage Plate Layout on the Thermal Insulation Performance of VIP Incubator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Lin Yifan, Wang Jiali, Kan Ankang, and Chen Wu

529

Experimental Study on Condensation Characteristics of Volatile Organic Compounds . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hao Xu, Xiafan Xu, Liubiao Chen, Jia Guo, and Junjie Wang

537

A Novel Segmented Non-Uniform Finned Channel of Supercritical LNG Applicable for Printed Circuit Heat Exchanger . . . . . . . . . . . . . . . . . . . . . . . Qingfeng Jiang, Chongyao Pan, Shiqing Wan, Huaibing Li, Qiang Zhu, and Bao Fu A Simulation Study of Enhanced Radiation Cooling on a Radiator with High Emissivity Coating . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Ming Liu, Hongbo Xu, and Nan Peng

544

551

Cryocoolers Investigation on a Free-Piston Stirling Cryocooler at −80 °C for China Space Station . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zhang Yin, Wang Bo, Ni Zhuqing, Luo Gaoqiao, Wang Shuai, Tian Xinghua, Wu Weiwei, Yao Xiaolei, Liu Ting, Zhang Yongqing, Zhu Kuizhang, Fan Yufeng, and Zhang Chi

561

Experiment Study a High-Power Pulse Tube Cryocooler . . . . . . . . . . . . . . . . . . . Haibin Dai and Shaowei Zhu

569

A High Efficiency Stirling/Pulse Tube Refrigerator Working at 15−20 K . . . . . Kongkuai Ying, Zhenhua Jiang, Shaoshuai Liu, and Yinong Wu

576

Contents

A High-Efficiency 80 K Coaxial Pulse Tube Refrigerator with Both Active Phase Shifting and Work Recovery . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hejun Hui, Jiantang Song, Shaoshuai Liu, Lei Ding, Hanzhen Xiang, Zhenhua Jiang, and Yinong Wu Progress in and Outlook for Work Recovery Type Pulse Tube Refrigerator . . . . Shaowei Zhu A 910 mW@15 K Thermal-Coupled Pulse Tube Cryocooler with Active Phase Shifter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Wang Yin, Hejun Hui, Wenting Wu, Jiantang Song, Shaoshuai Liu, Zhenhua Jiang, and Yinong Wu

xxxi

583

590

597

Thermodynamic Performance Analysis of an −180 to −150 °C Refrigeration System with Precooling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Dandan Sun, Haocheng Wang, Qinglu Song, Dechang Wang, and Jinxing Wu

605

Experimental Study of a Single-Stage Coaxial Pulse Tube Cryocooler Aimed at Cooling Two Infrared Detectors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Yinan Han, Ankuo Zhang, Wenhui Yu, and Jing Xie

613

Experimental Research About Displacer Type Micro Pulse Tube Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zhimin Guo, Shaowei Zhu, and John M. Pfotenhauer

621

A Superhigh Performance 40 K Pulse Tube Refrigerator with Miniaturized Active Phase Shifter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Song Jiantang, Hejun Hui, Shaoshuai Liu, Zhenhua Jiang, and Yinong Wu Investigation on Dynamic Pressure Characteristics of a 5–7 K Three-Stage Stirling-Type Pulse Tube Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . Wen Fengshuo, Wu Wenting, Liu Shaoshuai, Song Jiantang, Tang Xiao, Zhu Haifeng, Jiang Zhenhua, and Wu Yinong Calculation Method of the Theoretical Helium Liquefaction Rate Based on the Regenerative Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Qiang Cao, Miaomiao Wang, Tiancheng Xiao, Yuji Chen, Pengcheng Wang, Chaojie Chen, Peng Li, Zhihua Gan, Qinyu Zhao, and Bo Wang Development of a 2 K Joule-Thomson Cryocooler with 4 He . . . . . . . . . . . . . . . . Caiqian Dong, Shaoshuai Liu, Xinquan Sha, Wang Yin, ZhenHua Jiang, and YiNong Wu

628

635

642

649

xxxii

Contents

Characterization of an Embraco Oil-Free Linear Compressor for Driving JT Cryocoolers Working at Liquid Helium Temperature . . . . . . . . . . . . . . . . . . . . Changxu Qiu, Yunwei Shen, Dongli Liu, Lei Liu, Qinyu Zhao, and Zhihua Gan Performance Testing of Helium Valved Linear Compressor System Capable of Providing Compression Ratio Over 200 . . . . . . . . . . . . . . . . . . . . . . . . Xinquan Sha, Lei Ding, Qi Huang, Zicheng Li, Shaoshuai Liu, Zhenhua Jiang, and Yinong Wu Low Thermal Gradient Sorption Compressor Design . . . . . . . . . . . . . . . . . . . . . . A. H. Tolboom, H. J. Holland, C. H. Vermeer, and H. J. M. ter Brake Study of a Two-Stage Pulse Tube Precooling Helium Joule-Thomson Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zhichao Chen, Shaoshuai Liu, Xiaoshan Pan, Zhenhua Jiang, Lei Ding, and Yinong Wu Analysis of Oscillating Flow in Active Piston Phase Shifter Pulse Tube Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zongtao Geng, Chen Zheng, and Zheng Cui Study on Nonlinear Effects of Power Piston Spring Stiffness of Free Piston Stirling Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Kexin Jiao, Mingzhuo Yang, Chunyun Chi, Mingqiang Lin, Ruijie Li, Jian Mou, and Guotong Hong A Pre-cooling Type High-Frequency Pulse Tube Cryocooler Working at Liquid-Helium Temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Biao Yang, Zhaozhao Gao, Xiaotong Xi, Liubiao Chen, and Junjie Wang The Optimization on the Performance for a Fast Cooling Miniature JT Cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xiao Yong Li, Ling Wang, Qiang Long Zhu, Tai He Huang, and Li Huang Research on Mini-Channel Heat Exchanger of Pulse Tube Expander Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Wenhui Cui, Linghui Gong, Qiming Jia, Weiping Zhu, Zhengyu Li, Xiujuan Xie, Meimei Zhang, Ming He, and Han Zhou

656

664

672

679

686

694

701

708

716

Contents

A 150 K Micro Linear Split Stirling Cryocooler for High Operating Temperature Infrared Detectors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Jian Sun, Yong Zeng, Taihe Huang, Li Huang, Zehua Huang, Xiaoyong Li, Qianglong Zhu, Zhiming Zhang, and Yun Wang Calculation and Experimental Study of a 4 K Gas-Coupled Pulse Tube Cryocooler Driven by a Single Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Zhaozhao Gao, Biao Yang, Xuming Liu, Liubiao Chen, and Junjie Wang Start-Up Model Predictions of a Beta-Type Free Piston Stirling Generator . . . . Mingqiang Lin, Chunyun Chi, Kexin Jiao, Guotong Hong, and Jian Mou Simulation and Optimization of the Number of Thermal Buses in ADR Salt Pills . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Ping Liu, Yanan Wang, Jun Shen, Wei Dai, and Ke Li 10W@80K High Frequency Pulse Tube Cryocooler . . . . . . . . . . . . . . . . . . . . . . . Enchun Xing, Qingjun Tang, Hou lei Chen, Yuexue Ma, Tianshi Feng, Yuan li Liu, Nailiang Wang, and Jinghui Cai Theoretical and Experimental Investigation on Resonance Characteristics of a Dual-Coil Linear Compressor for J-T Throttle Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Z. J. Huang, Y. F. Niu, Y. J. Liu, E. C. Xing, Y. L. Liu, C. Zhang, and J. H. Cai Research on Performance of High-Power Linear Compressor for Completely Oil-Free Gifford-McMahon Cryocooler . . . . . . . . . . . . . . . . . . . . Zhijie Huang, Yuefeng Niu, Yanjie Liu, Yuanli Liu, Chen Zhang, Enchun Xing, and Jinghui Cai A 0.79 W/150 K Micro Pulse Tube Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . T. S. Feng, E. C. Xing, M. Gao, Y. T. Zhang, M. L. Liang, H. L. Chen, Y. Q. Xun, and J. T. Liang Experimental Investigation of Controlling Piston Offset in Oil-Free Linear Compressors for J-T Throttling Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . Y. L. Liu, Y. Q. Xun, H. L. Chen, Z. J. Huang, C. Zhang, and J. H. Cai Performance Analysis of the Flexure Spring for the Linear Compressor at Liquid Nitrogen Temperature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chen Zhang, Qingjun Tang, Miguang Zhao, Yuanli Liu, Zhijie Huang, Enchun Xing, Tianshi Feng, and Jinghui Cai

xxxiii

724

732

740

747

755

763

771

780

787

795

xxxiv

Contents

Simulation Study on Operation Characteristics of Cryocooler Regenerator Over Entire Cooling Process . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chen Xiantong, Li Shanshan, and Chen Xi Optimization of Two-Stage High-Efficiency Pulse Tube Cryocooler for Space Application at 20K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Q. L. Zhu, X. Y. Li, T. H. Huang, J. Sun, L. Huang, J. Quan, and J. T. Liang Effect of Mixture Composition on Thermophysical Properties of Refrigerant Mixture for MRJT Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Darshit Parmar and M. D. Atrey Numerical Simulation of Two-Stage Gas Coupled High Frequency Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Wu Xianlin, Liu Sixue, Yang Sujun, Jin Yu, Zhou Qiang, Chen Liubiao, and Zhou Yuan Study on Regenerator Matrix Optimization of Free Piston Stirling Generator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chunyun Chi, Mingqiang Lin, Ruijie Li, Kexin Jiao, Guotong Hong, and Jian Mou Design of High Frequency Coaxial Pulse Tube Cryocooler Working at 200 K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Yanbo Duan, Wei Wang, Bingkun Lv, Jue Wang, Laifeng Li, and Yuan Zhou Research of a High Frequency Reciprocating Piston Gas Bearing . . . . . . . . . . . . Ming-Zhuo Yang, Ming-Qiang Lin, Jian Mou, Ke-Xin Jiao, Chun-Yun Chi, Rui-Jie Li, and Guo-Tong Hong Effect of Flexible Thermal Connections on Temperature Oscillation in the Low-Temperature Range of GM Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . Shanshan Wu, Jue Wang, Hengcheng Zhang, Huiming Liu, Fuzhi Shen, Tiantian Xiao, Chuanjun Huang, Laifeng Li, and Yuan Zhou

802

810

818

825

832

840

848

855

Molecular Dynamics Simulation of the Phase Shift Mechanism on Inertance Pulse Tube Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Longyu Yang, Chen Zheng, and Zheng Cui

862

Study on Decreasing Cool-Down Time of a Rotary Cryocooler for HOT IR Detectors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . T. Xu, C. Zuo, Y. J. Guo, T. H. Huang, W. J. Liu, C. Sun, and L. Huang

870

Contents

xxxv

Cryogenic Fuel and Transportation Numerical Modelling of Transient Chill-Down Operation in Liquid Methane Transfer Line . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Keerthi Raj Kunniyoor and Parthasarathi Ghosh Computations of Capillary-Driven Cryogenic Flows in the Interior Corner with Microstructures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Ran Xu, Huan Jia, Han Chen, Xin Cheng, Mingkun Xiao, Guang Yang, Yonghua Huang, and Jingyi Wu

881

889

Space Cryogenics Effect of Wicking Capability on the Reseal Pressure of Woven Screens for On-Orbit Cryogenic Propellants Management . . . . . . . . . . . . . . . . . . . . . . . . . Chenfangda Shi, Ye Wang, Yilin Lin, Feng Ren, Gang Lei, Guang Yang, and Jingyi Wu Investigation on Bubble Departure Behavior of Liquid Oxygen Under Microgravity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Mingkun Xiao, Guang Yang, Chunyu Li, Yonghua Huang, and Jingyi Wu Thermal Design and Verification of Low-Temperature Storage Device for China Space Station . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Shuai Wang, Fankong Meng, Yufeng Fan, Meng Xiao, Bo Wang, Yin Zhang, Yawei Xu, Huning Yang, Changpeng Yang, Jianyin Miao, and Hongyang Zheng Modeling of Pressure Reducing in a Cryogenic Tank by a Thermodynamic Vent System Considering Flashing . . . . . . . . . . . . . . . . . . . Xujin Qin, Chuiju Meng, and Yonghua Huang Studies on a 60K Cryogenic System Based on Single-Stage Pulse Tube Cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sixue Liu, Yelong Tong, Huizhi Wang, Zhengdong Fu, Qiang Zhou, Lifeng Jiang, Jinyin Huang, Hongxing Zhang, Jianyin Miao, and Liang Zhao Analysis and Optimal Design of Screen Channel Liquid Acquisition Device for Hydrogen . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Jian Li, Yanzhong Li, and Yuan Ma

899

907

914

921

929

935

xxxvi

Contents

Analysis on the Influence of Heat Dissipation of Radiator on China Space Station . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xianlin Wu, Yang Fu, Bangxiang Che, Haiying Han, Yu Jin, Zhenyu Li, Yan Wang, and Jianfeng Cao

942

Thermal-Physical Properties of Cryogenic Fluids A Mesoscale Study of Forced Convection Induced Solid-Air Dendrite Microstructure Evolution Under Liquid Hydrogen Environment . . . . . . . . . . . . . Chaolong Li, Jian Wen, Lei Wang, Yanzhong Li, and Gang Lei Vibrational Spectroscopy of Ethanol Molecules Isolated in a Nitrogen Matrix . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Abdurakhman Aldiyarov, Aliya Tychengulova, Darkhan Yerezhep, and Anel Ismail

951

958

Experimental Study on the Quasi-Static Evaporation Characteristics of Liquid Methane . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Wenxin Zhu, Zhongqi Zuo, and Yonghua Huang

965

Study on Flow of Helium II Through a Porous Element for a Fountain Effect Pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Liguo Wang, Dong Xu, Rongjin Huang, and Laifeng Li

972

Low-Temperature Thermal Conductivity Test of Aerogels Under the Full-Pressure Range . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Xiafan Xu, Jianpeng Zheng, Hao Xu, Liubiao Chen, and Junjie Wang

979

Analysis of the Performance Difference Between4 He and 3 He in Space JT Cryocoolers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Ziyao Liu, Yuexue Ma, Jia Quan, Yanjie Liu, Juan Wang, Jianguo Li, and Jingtao Liang

986

Cryogenic Instrumentation and Control Recent Progress in Instrumentation for Large to Medium Scale Facilities Operating at 4 K Temperature Level . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sergiy Putselyk

997

Architecture Design of Control System for the DALS Test Facility Cryogenic System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1006 Lei Xu, Zheng Sun, Liangbing Hu, Xu Shi, Haining LI, Jichao Dong, and Xilong Wang

Contents

xxxvii

Development of High-Density Signal Feedthroughs for Liquid Argon Calorimeters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1013 Maria Barba, Michel Chalifour, Martin Aleksa, and Johan Bremer Helium Cryogenic Control System Design for CSMC Testing in CRAFT . . . . . 1020 Changheng Xie, Zhiwei Zhou, and Qiang Yu Cryogenics for Superconducting Materials, Devices and Systems Thermal Hydraulic Analysis of Background Magnet of 15T Conductor Performance Test Facility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1031 Libiao Hu, Qiangwang Hao, Yi Shi, Yu Wu, Kaihong Wu, and Chao Dai Conceptual Cooling Design for 14T MRI Superconducting Magnet System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1038 Qiangwang Hao, Libiao Hu, Kaihong Wu, and Yu Wu Thermodynamic Analysis of a Forced-Flow-Cooling System for a CICC Coil in a High Field Superconducting Magnet . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1044 Jingxin Zheng, Junjie Li, and Zhengrong Ouyang Results and Analysis of the First Year of Operation of the UKRI STFC Daresbury Vertical Test Facility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1051 Andrew J. May, Shrikant Pattalwar, David Mason, Keith Middleman, Mark D. Pendleton, Paul A. Smith, Stuart Wilde, Ayomikun Akintola, Alexander R. Bainbridge, Rachael Buckley, Gary Collier, Peter Corlett, Keith Dumbell, Michael Ellis, Mark Hancock, Jane Hathaway, Sean Hitchen, Carl Hodgkinson, Philip Hornickel, Gary Hughes, Conor Jenkins, Geraint Jones, Michael Lowe, Peter McIntosh, George Miller, Jennifer Mutch, Andrew Moss, Adrian Oates, Alan E. Wheelhouse, Alastair A. J. White, and James Wilson Strain Monitoring of ZrW2 O8 Reinforced Epoxy Composites by Using Embedded Fiber Bragg Grating Sensor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1058 Wanyin Zhao, Yuan Zhou, Hongyu Dong, Zhengrong Zhou, Zhicong Miao, Yalin Zhao, Zhixiong Wu, Jijun Xin, Xinran Shan, Chuanjun Huang, and Laifeng Li Cryogenic System Based on a Pulse Tube Cryocooler for Testing Tensile Properties of Micro-Sized Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1065 Haoying Qi, Yemao Han, Hengcheng Zhang, Yuchen Zhao, Fuzhi Shen, Haojian Su, Laifeng Li, and Yuan Zhou

xxxviii

Contents

Analysis of the Shielding Effectiveness of Cryogenic Magnetic Shields Used for Testing Superconducting Single-Flux-Quantum Circuit Chips . . . . . . . 1072 Lingjiao Wei, Guopeng Wang, Jianguo Li, Yuexue Ma, and Guotong Hong Mechanical and Thermal Properties of Cryogenic Materials Mechanical Characterisation of Additively Processed Filled Nylon Powder Composites Under Cryogenic Conditions . . . . . . . . . . . . . . . . . . . . . . . . . 1081 Grant Lumsden, Sarat Singamneni, Bart Ludbrook, Huub Weijers, and Rodney Badcock Improved Thermal Conductivity at Low Temperatures in Epoxy Nanocomposites by Hexagonal Boron Nitride Aerogels . . . . . . . . . . . . . . . . . . . . 1089 Zhicong Miao, Yalin Zhao, Zhengrong Zhou, Haojian su, Mingyue Jiang, Wanyin Zhao, Rongjin Huang, and Laifeng Li Low-Temperature Tensile and Impact Properties of Fe24Mn0.45C High-Manganese Steel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1095 Mingyue Jiang, Chuanjun Huang, Yuguo Chai, Meiyan Liu, Zhicong Miao, Rongjin Huang, and Laifeng Li Thermoelectric Properties of PbSe0.9 Te0.1 at Cryogenic Temperature . . . . . . . . . 1101 Haojian Su, Zhicong Miao, Shanshan Wu, Siyi Zhang, Haoying Qi, Min Zhou, Rongjin Huang, and Laifeng Li Measurement of Apparent Strain of Foil Strain Gauge at Low Temperature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1108 Sara Sato, Minoru Takeda, and Kazuma Maekawa Superconducting Materials and Devices Breakdown Characteristics of Liquid Nitrogen/Tetrafluoromethane/Polypropylene Laminated Paper Insulation System Utilized for Superconducting Energy Pipeline . . . . . . . . . . . . 1119 Zhihao Zhou, Qingquan Qiu, Yuping Teng, Liwei Jing, Naihao Song, Jingye Zhang, Guomin Zhang, and Liye Xiao Preliminary Excitation Test of REBCO External Magnetic Field Coil in Liquid Nitrogen . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1126 Masayoshi Ohya, Shohri Ikuta, Shoichiro Murata, Yugo Yamakawa, Shinsaku Imagawa, Akifumi Iwamoto, and Yasuyuki Shirai

Contents

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Stress Calculation of 50 kJ High Temperature Superconducting Magnet Energy Storage Using FEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1133 Ankit Anand, Abhay Singh Gour, Tripti Sekhar Datta, and Vutukuru Vasudeva Rao Simulation and Analysis of Cryogenic System for FFAG Superconducting Magnets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1140 Zhou Hongji, Fu Wei, Zhang Tianjun, Wang Chuan, Wang Fei, Liu Jingyuan, Zhu Xiaofeng, and Zhang Suping Author Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1149

About the Editors

Prof. Limin Qiu has been the leader of the Cryogenic Group “Cryoboat” at Zhejiang University, China, since 2001 and a teacher there since 1997. He conducts research and teaching on cryogenic engineering, especially in large-scale gas liquefaction and separation, pulse tube cryocooler and its applications, thermoacoustics, cryogenic heat transfer, and high power electronics cooling. He has published over 200 journal papers and 60 international conference papers, including some on ICEC, CEC, ICR and ICC. He is the holder of more than 100 China patents. He worked as a visiting scholar, Humboldt fellow or a guest professor in Osaka City University, Japan; University of Giessen, Germany; National Institute for Fusion Science (NIFS), Japan; University of Wisconsin, USA; University of Twente, the Netherlands. He has received several international and domestic awards including Karl von Linde Award, International Institute of Refrigeration (IIR), 1999; Alexander von Humboldt Fellowship, Alexander von Humboldt Foundation, Germany, 2000; top 100 excellent doctoral theses, Ministry of Education, China, 2000; the second-grade National Award for Technological Invention, the second-grade Award of Science and Technology Progresses, Ministry of Education, China, 2004 and 2013; Award of Science and Technology Progresses for the Youth, Chinese Association of Refrigeration, 2005; the first-grade Award of Science and Technology Progresses, Zhejiang Province, 2006; Natural Science Funds for Distinguished Young Scholar, 2008 and National Award of technology and invention, 2014. He serves as the dean of Qiushi College, Zhejiang University. He is the vice chairman of International Cryogenic Engineering Committee (ICEC) and the editorial board member of Cryogenics, the leading international journal in cryogenic engineering. He was the executive dean, Chu Kochen Honors College, from 2013 to 2017 and the vice dean, Department of Energy Engineering, Zhejiang University, from 2009 to 2013. He is the vice director of Commission of Cryogenic Engineering, Chinese Association of Refrigeration (CAR). He has been serving as the vice president of the Commission A2 (Gas Liquefaction and Separation) of International Institute of Refrigeration from 2007 to 2015 and the board member of International Cryocooler Conference (ICC) from 2010 to 2016. He has been involved with the organization of the International Conference on Cryogenics and Refrigeration (ICCR) since 2003, serving as the co-chairman for ICCR2013, ACASC2015 and ICCR2018 and as the chairman for ICEC28/ICMC2020. Dr. Kai Wang is a research professor in the Institute of Refrigeration and Cryogenics at Zhejiang University. He received his B.Eng. degree in Energy and Environment Systems Engineering and Ph.D. degree in Power Engineering and Engineering Thermophysics from Zhejiang University in 2009 and 2014, respectively. Before joining Zhejiang University, he worked as a postdoctoral researcher in the Clean Energy Processes (CEP) Laboratory at Imperial College London during 2018–2019 and in the

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About the Editors

Energy Research Institute at Nanyang Technological University during 2014–2017. He has been the managing editor of the journal Applied Thermal Engineering since 2019. His research interests focus primarily on high-performance energy technologies, components and systems for liquid-hydrogen production, storage and refuelling, organic Rankine cycles, thermoacoustic power generators/coolers, stirling engines, co/trigeneration systems and solar thermal technologies. He is the recipient of the Sadi Carnot Award from the International Institute of Refrigeration (IIR), one of the IIR Scientific Awards for young researchers working on thermodynamics. To date (January 2023), he has published one book, two chapters, more than 60 peer-reviewed journal papers and owns more than ten patents. Prof. Yanwei Ma is a professor of Institute of Electrical Engineering, Chinese Academy of Sciences (IEE-CAS), Beijing, China. He received his PhD degree from Tsinghua University in 1996 and then worked as a postdoctoral fellow at University of Science and Technology Beijing (USTB). Following the research associate positions at Institute for Materials Research, Tohoku University (Sendai, Japan), National Institute for Materials Science (Tsukuba, Japan) and Universite de Rennes 1 (France), he joined the IEE-CAS in 2004. He is the director of the Superconducting Materials Department at IEE-CAS, a large research group that is mostly devoted to the development of MgB2 and iron-based superconductor wires. He was awarded the “2019 ESAS Award for Excellence in Applied Superconductivity” for outstanding contributions to the development of iron-based wires and tapes. He has published more than 400 refereed SCI journal papers. He has given over 90 invited talks at international conferences. He is a board member of Physica C and Supercond. Sci. Technol., an associate editor of the Superconductivity News Forum, and a board member of International Cryogenic Materials Conference (ICMC).

Plenary

Helium Refrigeration Guy Gistau Baguer(B) Cryoguy, Biviers, France [email protected]

Abstract. It is customary that the Mendelssohn awardee makes a survey about the topic he had been involved in. Such a job has already perfectly been done, for helium refrigeration, by Hans Quack, at the time of his 2008 Mendelssohn Award talk [1]. Therefore, I thought that, instead of repeating same info, I could enlighten some important evolutions that, due to my age, I had the opportunity either to witness or to live, and explain in which context some of them arouse. Then, I will point out some significant moves that have happened since 2008 in the field of helium refrigeration. And finally, I propose a glance on some aspects of future projects. Keywords: helium refrigeration · large power · very large power

1 Some Evolutions in Helium Refrigeration Since 1980 1.1 Automation of Helium Refrigerators In the end of the 70’s, helium refrigerators or liquefiers were “semi-automatic”. That means that, after a rather complicated and tedious cooling down manual procedure was completed, some control loops were activated that were able to keep the system operating, provided no disturbance happened. From time to time, operators were complaining about bad performances, but, for example, after we performed a correct cold box regeneration, the system was nicely operating again. After a number of such uncomfortable situations happened, I decided to “include an operator into the refrigerator”. Of course, at this time, we knew almost nothing about programming. Fortunately, CEA Saclay had started to develop a very smart control system based on an existing small computer (not yet a so-called PC!) and the use of a new programming language: GRAFCET (GRaphe Fonctionnel des Commandes Etapes et Transitions), or in English: Stages and Transitions. The big interest of such an equipment was that we could write, write again, rewrite and modify easily our programs ourselves, by using the very simple GRACET language, without needing a programming specialist. From this time period, I remember having questioned the incredible level of difficulty of automatizing a so common activity, that can be performed by anybody on earth, as driving a car, but that is so complicated to program… This is finally happening, a lot of years later! © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 3–15, 2023. https://doi.org/10.1007/978-981-99-6128-3_1

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To-day, a refrigerator or a liquefier is obviously automatic. However, I should point out that, at the time we started selling such automatic machines, we faced curious behaviours: team leaders were concerned that their competence would be challenged by automation and that their workforce would be reduced. 1.2 The Screw Compressor Since my very first job, in 1965, I have been dealing with (or, more precisely, fighting against) compressors and had troubles with them… That started with oil lubricated reciprocating compressors that were feeding small Joule Thomson hydrogen and helium liquefiers. At this time, we did not really know how to be rid of oil. Later, we used plastic ring reciprocating compressors. No longer oil troubles, but the life time of such rings was between 10000… and 48 h. Could anyone believe that I spent half the week-ends of a calendar year on the compressor of the CELLO refrigerator in DESY, Hamburg? We also used labyrinth compressors that had the inconvenience to be expensive and necessitated heavy concrete blocks to damp vibrations. Information about screw compressors that came from the USA was very exciting, but the oil removal process was rather mysterious: how to get rid of oil so perfectly that nothing gets into the cold box? After visiting BNL, Monsanto, and Balston in the USA, Domnik Hunter in UK, I decided to try by myself. An opportunity appeared when CNRS Grenoble was planning to replace a labyrinth compressor that had recurrent oil leaks along its piston stems. An agreement was reached: Air Liquide designs a compression station fitted with a screw compressor and builds the oil removal system (ORS), CNRS buys the screw compressor and operates the system. At the occasion, when we began running the compressor, we learned about the curious behaviour of a screw compressor according to its operating conditions. To evaluate the oil removal performance, we had to measure the quantity of oil aerosols and vapour that were released by the ORS. My friends form BNL had shown me the equipment they were using: a laser counter that was included in the helium process pipe. This solution avoided the almost impossible problem of getting and transporting a significant sample of aerosol-loaded gas. Oil vapour analysis, for which it is also difficult to get significant samples, was dealt in house at an Air Liquide laboratory. In addition to using the particle counter, we ran one of our turbine static gas bearing system fed by the compressed helium in order to show that the possible remaining oil particles did not endanger its operation. When we were convinced that our system was operating correctly, we connected the compression station to the liquefier and run it for 10000 h. Finally, we were able to offer such a compression solution to customers. An oil lubricated screw compressor followed by its ORS is now the standard solution for most of the helium refrigeration systems (Fig. 1). The ORS is a static equipment that, usually, runs very smoothly. However, if there is some trouble, the consequences are generally heavy: cleaning the internal components of a cold box is very disturbing, both in time and cost. Therefore, the operator must be very vigilant about the operation of the ORS. I strongly advise to fit the ORS with an

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automatic draining system that allows monitoring permanently the first stages coalescer performance.

Fig. 1. A typical Oil Removal System

I also think that there is room for improving the control of the oil flow that is injected into the compressor during the compression phase, according to the operating regimes. 1.3 The Cryogenic Centrifugal Compressor (CCC) Tore Supra (WEST Since 2013) In the early 80’s, EURATOM decided to build the Tore Supra fusion experiment. The requested cryogenic power was 300 W at 1.8 K (15 g/s). At this time, all existing refrigerators operating at 1.8 K were incorporating Roots machines in order to pump on the helium bath. CEA and Air Liquide, who was one of the potential suppliers, thought that, if fusion was really a future solution, larger powers would soon be necessary. The size and the number of Roots machines for the sub atmospheric stream pumping station would then become unrealistic. In order to reduce the size of equipment, one solution was to operate at lower temperatures. In order to increase the reliability, one solution was to use rotating machines operating with non-contact bearings. Such thinking led obviously to cryogenic centrifugal compressors (CCC). After development on Air Liquide’s own funding, centrifugal compressor prototypes fitted with active magnetic bearings were built and tested. In order to reduce the number of cryogenic pumping stages (and therefore, the difficulty for this first trial), it was decided to go to a mixed pumping structure, with two CCC that were incorporated into the Tore Supra refrigerator, followed by a room temperature machine (Fig. 2, left). At this time, the use of screw compressors operating at sub-atmospheric suction pressure had not yet been demonstrated, therefore liquid ring pumps were selected.

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This first refrigerator, incorporating centrifugal cryogenic compressors (Fig. 2, right), proved to be easy to operate. CEBAF After Tore Supra showed that the centrifugal cryogenic compression was a viable and good solution, CEBAF, JLAB, Newport News, VA, USA, decided to go for a full cryogenic compression (Fig. 2, centre) performed by four CCC arranged in pure series, for the 5000 W at 2.0 K refrigerator they needed. At this occasion, we jumped from 15 g/s for Tore Supra up to 240 g/s machines in one step!

Fig. 2. The Tore Supra refrigerator structure, left; the CEBAF refrigerator structure, centre; a Tore Supra cryogenic centrifugal compressor, right.

Starting up the Tore Supra cryogenic compression system was easy, because the system incorporated a volumetric room temperature compressor downstream the centrifugal compressors, but in 1991, at the time of starting up the full cryogenic centrifugal compressor train of CEBAF, I experienced one of the most disturbing situations of my professional life: within seconds, I realised that I did not know how to start up and control such a system! I understood that there is no stable operating point when pumping on a liquid bath at a constant rotational speed (Fig. 3, left) because the flow that feeds the compressor train is generated by its rotational speed variation. The rotational speed control law has to take into account a number of parameters as inlet pressure and temperature, mass flow rate and compression ratio. By the way, a harmonious combination of these parameters allows following a correct path: the red dotted line in the compressor map (Fig. 3, right). Obviously, the situation is much more difficult to manage when a number of compressors arranged in series operate simultaneously (Fig. 4)!

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Fig. 3. No stable operation at constant rotational speed, left; controlled rotational speed, right

Fig. 4. Evolution of bath pressure, mass flow rate and rotational speeds during pump down

1.4 Cycle Efficiency In 1991, when issuing the RFQ for the LEP refrigerators, CERN decided to select the supplier, not only on the capital cost as usually, but on the minimum sum of capital cost (CAPEX) plus operating costs (OPEX), the later, based on a period of time of operation and a cost of energy. This was a real incentive for the potential suppliers to develop efficient cycles. There were two main areas of the cycle to be investigated: precooling and cold end. To get a more efficient precooling, the cycle designer has to make Mr. Carnot happy: increase the number of cooling stages or, in other words, pile-up precooling Brayton cycles. However, according to the cost of extra components, one understands that there would be a limitation of their number. To get a more efficient cold end, all helium should be isentropically expanded (“with external work”) down to the lowest pressure. That means that the conventional JT valve is to be replaced by a turbine. When the turbine discharge pressure is lower than 2.28 bar, the turbine expands into double-phase.

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The interest of such an arrangement is clearly shown on Fig. 5, were one can see that expanding a so-called JT flow of 500 g/s from 10 bar, 6.0 K down to 1.2 bar with a 75% efficiency expander, produces an extra power of 2764 W that is available at 4.41 K. At that time, the turbine makers were generally afraid that the turbine could get in trouble with such operating conditions. Possibilities of “unbalancing”, “cavitation” or “erosion of the wheel” were raised.

Fig. 5. Interest of a turbine expanding in the two-phase region

The first step, to had such a process based on this idea accepted, was to convince my turbine-maker friend Jean-Claude Villard that such a turbine could operate. The second step was to convince CERN that it was a viable solution. For that, we decided to install such a two-phase turbine on a 6-kW refrigerator that was been built at this time. We proposed, firstly to show that the turbine could run as expected, secondly to run it for 1000 h and check its status. The test was fully positive. Presently, the two-phase turbines that operate on CERN LHC refrigerators have been running smoothly for more than 100000 h since 2003. Finally, the CERN incentive allowed to show that a higher efficiency refrigerator, even if it is more complicated, needs a lower cycle mass flow rate that absorbs less power (lower OPEX) and, consequently, all the components have a smaller size (lower CAPEX). The LEP refrigerators that have been built at this time are, still presently, to my knowledge, among the most efficient refrigerators, reaching around 29% compared to Carnot (see Fig. 6). In 1993, at the time of its commissioning, this refrigerator gave the opportunity to show that it was possible to keep a good efficiency, even at turn-down operation, by floating the cycle middle and high pressures (Fig. 7). An even better performance at turn-down would have been possible by using frequency converters instead of the compressor slide valves…

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Fig. 6. Two solutions for the LEP CERN refrigerators

Fig. 7. Power adaptation by floating middle and high pressures (Air Liquide’s designed LEP refrigerator).

1.5 Large Liquefiers When the 7000 L/h Qatar II helium liquefier was built, I had the opportunity to see a liquefier that is 1000 times larger than the ones I was building in 1965! (Fig. 8). 1.6 The Turbo-Brayton Refrigerator This refrigerator operates on a nitrogen cycle; however, I am so admirative of such a system because of its bare simplicity that I decided to add it to my talk. A conventional Brayton cycle would incorporate an oil lubricated screw compressor, an Oil Removal System, and a cold box housing the heat exchanger, the expansion turbine, an adsorber and valves (Fig. 9, left).

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Fig. 8. Myself, inside the cold box of the QATAR II 7000 L/h helium liquefier

Incorporating a centrifugal compressor, has following advantages: as a centrifugal compressor is clean, one can delete the ORS; as it is leak tight, one can delete the cycle adsorber; as it can run at any suction pressure, one can delete the buffer capacity, the load and unload control valves, the cold box isolation valves; as it is fitted with a variable speed drive, one can delete the by-pass and turbine inlet valves. Finally, by arranging the compressor and the turbine on the same shaft, one can recover the energy that is extracted by the turbine, improving strongly the cycle efficiency (Fig. 9, right).

Fig. 9. Simplification of a conventional Brayton cycle (left) leading to the turbo-Brayton structure (right)

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Fig. 10. A turbo-Brayton refrigerator

In addition to such a simplicity, other positive aspects are to be pointed out: • As it is a compact-closed-system skid (Fig. 10) with its cycle gas load already incorporated, the erection on site is simplified and, therefore, performed in a very short period of time. • The compressor rotational speed is the only control parameter of the system. It allows floating the cycle pressures and, consequently, controlling the operating temperature and the absorbed power in a very simple and efficient way. • Finally, due to the active magnetic non-contact bearing technology, the system MTBM is very long.

2 Future 2.1 Very Large Helium Refrigerators and Hydrogen Liquefiers Both scientific and industrial projects need very large helium refrigerators and hydrogen liquefiers. I heard about projects involving helium refrigerators up to 100 kW at 4.5 K or 100 tpd hydrogen liquefiers! The efficiency of such plants will be a very important point. The challenge will be both on gas compression and cycle arrangements. Helium Compression Compression of helium is, presently, mainly performed with oil lubricated screw compressors. Such machines are very reliable but their efficiency is not very good, a huge oil removal system is to be implemented, that could become dangerous in case of incorrect operation and their maximum size makes it necessary to arrange a number of them in parallel to get the required flows. Since a long time, I have been dreaming about centrifugal compressors, but I could never implement one of them in the projects I worked on. At that time, these tentative were made unrealistic due to the too high rotational speeds that would have been necessary, the high number of stages and, consequently, the too high capital cost. However, in a near future, an interesting situation will be experienced at ITER where a centrifugal compressor will circulate 2 kg/s of helium from 16 to 18 bar through the

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thermal shields. However, the shaft seals of this machine, that are not positively leak tight, allow some helium leak that must be recovered, which means that a purification and re-compression unit is to be implemented, making the system more complicated.

Fig. 11. A possible hermetic centrifugal compressor (HOFIM™ compressor system by MAN Energy Solutions Switzerland Ltd)

In order to avoid such leaks, a solution would be to use a hermetic compressor where all rotating equipment is fully enclosed (Fig. 11). To avoid any pollution, the bearing system must be oil-free: active magnetic bearings would be a perfect solution. Possible Cycle Arrangements for Very Large Refrigerators or Liquefiers One way to improve efficiency is to please Mr Carnot by increasing the number of cooling (or expansion) stages. As one is dealing with very high powers, one looks for the highest possible cycle efficiency, therefore, the number of expansion stages is not limited. The temperature differences that are generated by the expanders can be piled up in order that the sum of elementary temperature differences equals 300 K minus the discharge temperature of the coldest turbine. Therefore, the number of expansion stages depends on the cycle expansion ratio and the efficiency of expanders. That leads to a cycle shown in Fig. 12, A, where the warm heat exchanger of each of the conventional Brayton cycle is deleted (“single-HX” Brayton) and the discharge temperature of an expander equals the inlet temperature of the colder one. It happens that the inlet temperatures of the turbines are harmoniously distributed in a geometric progression (Fig. 12, B), which is the optimum arrangement. Same arrangement appears in Fig. 2 of [2]. By the way, such a “single-HX” Brayton structure (Fig. 12, C) is more efficient than the conventional Brayton cycle (Fig. 12, D) because the heat loads that are brought, both by the heat exchanger warm end temperature difference and by the cooling of the gas to be liquefied, are extracted by the turbine at the highest possible temperature: the single-HX Brayton arrangement is the most efficient. Of course, at the highest temperatures, the turbine enthalpy drop can be too high to cope with one-off turbine. Either turbines arranged in pure series and/or operating between high and middle pressures allow solving the problem. Such an arrangement can be used to precool a “JT” cycle with, for example, turbine T8 expanding in supercritical helium and turbine T9 expanding in the two-phase domain, as shown in Fig. 13, A. When the precooling temperature of the JT cycle decreases (T7 discharge temperature), the liquid concentration at the discharge of T9 increases and so does the Carnot efficiency of the cycle.

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Fig. 12. A: piling up single-HX Brayton’s, B: Brayton turbine inlet temperatures distribution, C: “single-HX” Brayton, D: “Full” conventional Brayton.

In order to get orders of magnitude, very simple but not optimised calculation have been performed with REFPROP® and Excel®, assuming: • no pressure drops, • not heat leaks, • smaller heat exchanger temperature differences = 1% of the temperature, with minimum 0.2 K, • expansion turbine efficiencies: 0.75 when higher than 20 K, 0.73 when lower than 20 K, • 95% liquid at discharge of coldest turbine T9, • no recovery of turbine power. According to the above assumptions, a helium refrigerator of 45 kW at 1.2 bar, as shown in Fig. 13, A, would need 4420 kW as an isothermal compression power. Assuming a 55% isothermal efficiency compression as is has been measured on the LEP refrigerator screw compression stations, the Carnot efficiency would be 0.37 (179 W/W), compared to the measured LEP refrigerator efficiency: 0.29 (228 W/W). Assuming a 70% efficiency if running centrifugal compressors would allow reaching 0.48 (140 W/W). When adding another single HX Brayton, between T8 and T9, figures improve slightly, reaching 4340 kW, 175 and 138 W/W. Therefore, it is certainly of a high interest to develop expanders that would allow a high liquid concentration at their discharge side. A similar structure can be used for large helium (Fig. 13, B) or hydrogen liquefiers (Fig. 13, C). As the heat load for cooling the gas to be liquefied, and converting it in case of hydrogen, is very high, it is interesting to use a precooling cycle (or cycles) fitted with single HX Brayton’s, operated with nitrogen, or any other suited gas or gas mixture, down to an optimum temperature. Of course, such cycles should incorporate, as far as possible, centrifugal compressors.

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Fig. 13. Very large: A, helium refrigerator; B, helium liquefier; and C, hydrogen liquefier; with nitrogen precooling (green lines)

As the turbines of such very large systems are powerful, it is a must to plan recovering the energy of the biggest machines (> 100 kW?) in order to, even more, increase the cycle efficiency. 2.2 Cooling Down Large Systems Without LN2 Cooling down very large and heavy systems down to 100 K is generally performed by vaporising and warming up very large quantities of liquid nitrogen. Such a solution, that is presently implemented for LHC, would become unrealistic for larger systems due to liquid nitrogen logistic reasons. In the future FCC, CERN is considering using a dedicated Brayton cycle that would, in addition, have an operational cost that is expected to be lower [3].

3 The Book I Wrote In 2012, I wrote an article about helium refrigeration in the French “Techniques de l’Ingénieur”, but I felt very frustrated to have to stuff so many things into so little space. Soon after, I decided to start the same job, but without any space limitation. It happened to be a very long-term job: eight years… This book is the result of my personal journey into helium refrigeration which is presented in a one-off document, in a logical progression through the main topics of this activity, with emphasis on practical aspects [4]. It gave me the opportunity to understand (at least!) and experience what the French poet Nicolas Boileau (1636 - 1711) wrote such a long time ago (Fig. 14), that is taught to French schoolboys and schoolgirls:

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Fig. 14. Boileau’s rules for well writing

However, I am sorry to disagree with Nicolas on one point: the “words to say it” do not come easily, at least for me! I hope this book, which I used to call my “cryogenic testament”, will facilitate the first steps for newcomers into helium refrigeration and, in a general way, help the cryogenics community. I will continue to keep abreast of developments in helium refrigeration and I encourage new generations to further improve this technology. Acknowledgments. To all people I worked with, who helped me improving, by minute or larger steps, my comprehension of cryogenic systems. To the Air Liquide company where I discovered and loved so much cryogenics and where I had the opportunity to live so many exciting experiences with my old and younger colleagues.

References 1. Quack, H.: Proceedings of the International Cryogenic Engineering Conference 22 (2008). Mendelssohn award talk 2. Ganni, V.: Optimal design and operation of helium refrigeration systems. In: Proceedings of Particle Accelerator Conference 09, Vancouver, BC, Canada, pp. 1931–1935 3. Benedikt, M., et al.: Future circular collider study. Volume 3: the hadron collider (FCC-hh) Conceptual Design Report, preprint edited by CERN accelerator reports, CERN-ACC-2018– 0058, Geneva, December (2018). Published in Eur. Phys. J. ST 4. Gistau Baguer, G.: Cryogenic Helium Refrigeration for Middle and Large Powers. 1srt edition. Springer Nature, Cham, Switzerland (2020). ISBN: 978–3–030–51676–5

Thermodynamics and Heat Transfer in Cryogenics: A Personal View Ray Radebaugh(B) Applied Chemicals and Materials Division, National Institute of Standards and Technology, Boulder, CO, USA [email protected]

Abstract. Cryogenics involves a large difference in temperature, a form of energy potential. To produce and maintain such a large temperature potential requires the efficient use of other energy potentials, such as pressure, magnetic field, electric field, etc. Thermodynamics is a powerful tool that relates these potentials as well as heat and work to show the theoretical limits to what is possible in a thermodynamic system and as a guide to what can be engineered in practice given certain fluid and material properties. The use of Gibbs free energy, exergy, and entropy helps determine limits on system efficiency, but efficiency must be balanced with size in practical cases. A relationship between the two for refrigerators is shown that makes use of known cycle and heat exchanger parameters applicable to a specified configuration (screen, tubes, plates, etc.). This presentation discusses some historical and scientific insights I have observed regarding the usefulness of thermodynamics and heat transfer in chemistry, phase equilibria, superconductivity, dilution refrigerators, and cryocoolers for over 65 years, starting with an air liquefier in high school for a science fair project. Keywords: Cryocoolers · Cryogenics · Efficiency · Exergy · Gibbs free energy · Heat transfer · Thermodynamics

1 Introduction Thermodynamic processes and heat transfer are an integral part of cryogenic refrigeration and liquefaction. The measurement of enthalpy, entropy, and Gibbs free energy for determining spontaneous chemical reactions at room temperature requires specific heat measurements down to at least 5 K for high accuracy. Thermodynamic principles and functions are used extensively for modeling of superconductivity and material properties. Because cryogenics relies inherently on a large temperature difference from ambient, heat transfer becomes an important design consideration of any cryogenics process or equipment. This paper discusses the use of both thermodynamics and heat transfer to find a good balance between efficiency and system size. Several applications to cryogenics are covered using the personal experience of the author. Contribution of NIST, not subject to copyright in the US. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 16–29, 2023. https://doi.org/10.1007/978-981-99-6128-3_2

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2 Thermodynamic Fundamentals 2.1 First and Second Laws of Thermodynamics According to the first law of thermodynamics an energy balance on an open system in thermal contact with one thermal reservoir and in steady state is given by ˙ =W ˙ + m(h Q ˙ e − hi ),

(1)

˙ is heat flow rate from the reservoir to the system, W ˙ is work rate produced where Q (power) by the system, m ˙ is the mass flow rate, and h is the specific enthalpy of the flowing stream at the exit and the inlet. The second law gives ˙ Q + S˙ irr = m(s ˙ e − si ), Tr

(2)

where Tr is the temperature of the thermal reservoir, S˙ irr is the irreversible entropy production rate (always positive), and s is the specific entropy of the flowing fluid. 2.2 Gibbs Free Energy ˙ from Eq. (1) into Eq. (2) yields Substituting Q ˙ = m[(h W ˙ i − Tr si ) − (he − Tr se )] − Tr S˙ irr .

(3)

The process is reversible when S˙ irr = 0. For reversible heat transfer the system is at the temperature Tr . In that case the work produced is the maximum. The combination h−Tr s is the specific Gibbs free energy g, which represents the potential to do maximum work. When g = 0, no work is produced and the inlet and exit states are equal or in equilibrium. The term g can refer to any inlet and outlet material, such as reactants and products in a chemical reaction. A decrease in g means the reaction is spontaneous.

Fig. 1. An open thermodynamic system in thermal contact to multiple temperature sources.

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2.3 Exergy We now consider the case of the more complex system shown in Fig. 1 in which heat is transferred from any number of thermal reservoirs with one of those being at ambient temperature and pressure. The first law for steady state conditions is  ˙0 = ˙j −W ˙ + m(h Q Q ˙ i − he ). (4) j

The second law for steady state is Q ˙j ˙0 Q = + m(s ˙ i − se ) + S˙ irr . T0 Tj

(5)

j

Substituting Eq. (4) into Eq. (5) and rearranging yields [1]   T0 ˙ ˙ = W 1− Qj + m(h ˙ − T0 s)i − m(h ˙ − T0 s)e − T0 S˙ irr . Tj

(6)

j

In a reversible process S˙ irr = 0, so the reversible power produced or required is the ˙ given by exergy E,   T0 ˙ ˙ rev = 1− Qj + m(h ˙ − T0 s)i − m(h ˙ − T0 s)e = E˙ (7) W Tj j

The difference between the reversible power in Eq. (7) and the actual power in Eq. (6) is the lost power or exergy destruction rate, given by ˙ = T0 S˙ irr = E˙ dest , ˙ rev − W W

(8)

which is always positive. An exergy balance for a steady-state system is given by E˙ out = E˙ in − E˙ dest − E˙ unused = ηE˙ in ,

(9)

where η is the exergy efficiency or second law efficiency of the process. Additional information on exergy analysis and entropy minimization is given by Bejan [1]. Exergy flows can be shown with a Grassmann diagram [2]. Figure 2 shows such a diagram that may be typical for a cryocooler. The total loss is often broken into different segments representing various individual losses in different components.

3 Science Fair Project: Air Liquefier In high school I decided to build an air liquefier as a science fair project in my senior year. A two-stage compressor was built using a metal lathe that I had received as a Christmas present from my parents. I decided to use the more efficient Linde dual-pressure method rather than the simpler Hampson process. A ½ hp (400 W) motor and a gear reduction unit were used to drive the compressors at a speed of 120 rpm. The compressor system could produce 20 MPa pressure when closed off.

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Fig. 2. Grassmann diagram that shows exergy flows in a system such as a refrigerator.

The ideal specific liquefaction rate is given by the exergy of liquid air at 100 kPa, which is 734 J/g. With 400 W input power the liquefaction rate is 2.2 L/hr. A secondlaw efficiency of 1% would produce 22 cm3 /hr. The three-channel heat exchanger for the Linde process was made with copper refrigeration tubing, whose diameters and lengths were guesses. Vacuum insulation was beyond my capability, so I used fiberglass insulation from our attic that was packed inside an expanded foam box. A minimum temperature of 115 K was obtained, but 87 K was needed to liquefy air. Though the liquefier was not successful in liquefying air, it received first place in the high school science fair. What I learned from this experience with an air liquefier is that the design and fabrication of efficient heat exchangers is a difficult task and requires a thorough understanding of heat transfer and fluid mechanics.

4 Low Temperature Calorimetry During my junior and senior years at the U. of Mich. (1960–62), Prof. Edgar Westrum asked me to help with the low temperature calorimetry program. The heat capacities of many different compounds were being measured from 5 to 350 K as part of the program to determine thermodynamic functions of many different compounds. A very accurate adiabatic calorimeter had been developed for this effort that could carry out measurements with an uncertainty of less than 0.1% [3] even though the instrumentation used at that time for the measurements are considered antiques by today’s standards. The thermometer was a standard capsule platinum thermometer, whose resistance was read with a bridge circuit using a high precision potentiometer and a mirror galvanometer as the null detector. A focused light beam reflected off the mirror galvanometer to a spot on a scale about 3 m behind the galvanometer. The spot location on the scale was observed by the operator looking through a telescope at a stationary mirror. The long optical path length provided a voltage sensitivity in the thermometer circuit of about 10 nV. The same method was used to measure the heater power. Both measurements were made by the “pilot” in the two-person measurement procedure. The “co-pilot” in the team measured the temperature difference between portions of the adiabatic shield and the calorimeter and manually varied the power to heaters on these different shield portions to maintain the temperature difference to within a few mK. I served as “co-pilot in my junior year but advanced to pilot in my senior year.

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R. Radebaugh

Data analysis consisted of transferring the raw data to computer punch cards. The measured heat capacity data points (per mole) were plotted on a 1-m wide graph and smoothed with a plastic spline held in place with many cast iron “ducks.” Thermodynamic functions were calculated by the following integrals: T1

T2 cp dT + h1 +

h(T ) = h(0) + 0

T1 s(T ) = s(0) + 0

T cp dT + h2 + ....... +

T1

cp h1 dT + + T T1

T2 T1

cp dT Tn

cp h2 dT + + ....... + T T2

T

(10) cp dT , T

Tn

where hi are the heats of transition for first order phase transitions. These thermodynamic functions plus the Gibbs free energy are used in predicting chemical reactions.

5 Thermodynamics of Superconductors My PhD thesis topic under the guidance of Prof. P.H. Keesom of the Physics Department at Purdue Univ. was the Thermodynamic Properties of Vanadium in the Superconducting, Normal, and Mixed States [4, 5]. I measured the specific heat of high-purity, singlecrystal vanadium which has a transition temperature of 5.414 K when extrapolation was made to infinite electron mean free path [4]. Thermodynamics was used to relate the specific heat to magnetic properties. Magnetic measurements were also made on the sample and compared with those derived from specific heat. However, the magnetic measurements were considerably inferior because of hysteresis. 5.1 Specific Heat and the Critical Field Curve For a superconductor the change in Gibbs free energy between the normal state and the superconducting state at zero field is given by G(T ) = Gn (T , H ) − Gs (T , 0) = 21 μ0 Hc2 (T ),

(11)

where H c is the critical field. Gibbs free energies are derived from specific heat measurements, which then determines the critical field curve from Eq. (11). For vanadium μ0 H c (0) = 0.1421 T [4]. Differentiating Eq. (11) gives the entropy difference   d G(T ) dHc (T ) − = Sn (T ) − Ss (T ) = −μ0 Hc (T ) . (12) dT dT According to the third law of thermodynamics S = 0 at T = 0 K. Because H c (0) is finite, dH c (T )/dT = 0 at T = 0 K. The slope of the critical field curve at T c is related to the specific heat jump in zero field from superconducting to normal states by       dHc (T ) 2 C d S(T ) = = μ0 . (13) T Tc dT dT Tc Tc The above analyses show that specific heat curves and thermodynamics completely characterize the critical field curve, including the slopes at T = 0 K and at T = T c .

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5.2 Type II Superconductivity and the Mixed State In addition to the thermodynamic critical field H c , type II superconductors have a lower critical field H c1 and an upper critical field H c2 which separates the superconducting (Meissner) state, the mixed state, and the normal state. Like niobium, vanadium is a type II superconducting element, but one of the purposes of my thesis was to determine if vanadium is an intrinsic type II superconductor in the limit of high purity. Type II superconductivity occurs when the following condition is satisfied: √ 2κ = lim [Hc2 (T )/Hc (T )] > 1, (15) T →Tc

where κ is the Ginzburg-Landau parameter. We found that for vanadium the high √ purity limit becomes 2κ0 = 1.20, which √ shows that vanadium is an intrinsic Type II superconductor [5], as is niobium with 2κ0 = 1.075.

6 Thermodynamics and Heat Transfer in Dilution Refrigerators An important consideration in the search for new refrigeration systems is the second law thermodynamic principle of the cooling power, given by dQ ≤ TdS. The equal sign (maximum cooling) occurs only with a reversible process. With a change of state, the process can be reversible with a steady state refrigeration power of ˙ rev = n˙ T (smd − smo ), Q

(17)

where n˙ is the molar flow rate, smd is the molar entropy in the disorder state, and smo is the molar entropy in the order state. Large entropy differences often occur with changes of state, for example, liquid to gas. A 3 He refrigerator provides good refrigeration power down to about 0.3 K, but for lower temperatures the density in the vapor phase is so low that it becomes very difficult to achieve a reasonable flow rate. 6.1 Phase Separation in 3 He-4 He Solutions In 1956 Walters and Fairbanks found that 3 He-4 He solutions separate into a concentrated phase and a dilute 3 He phase below about 0.8 K [6]. The molar Gibbs free energy 3 of He is equal on the two sides of the phase boundary, which then gave promise to a high refrigeration power because the dilution process would be reversible. In the dilute phase 3 He behaves much like a gas. Measurements [7, 8] showed the 3 He concentration at 0 K to be about 6.4% on the dilute phase curve. Thus, the density remains high as T approaches zero and high diffusion (flow) rates would be possible. Figure 3a shows the phase diagram for 3 He-4 He solutions. Three dilution processes were proposed in 1962 [9]. The third process, suggested by Hall and shown in Fig. 3b, is the process used today. Hall went on to build the first successful dilution refrigerator (DR) in 1966 using the process he proposed. He achieved a temperature of 0.07 K [10].

3 He

22

R. Radebaugh

6.2 Thermodynamic Properties of 3 He-4 He Solutions A few months after results of Hall’s successful dilution refrigerator were published, I finished my PhD at Purdue and went to NBS/Boulder for a postdoc position to develop a model of 3 He-4 He solutions for predicting the thermodynamic properties of the solutions and their use in the design of dilution refrigerators [11]. Edwards, et al. [7] used an ideal Fermi-Dirac gas model with effective mass m* = 2.5m3 to fit their specific heat data of dilute solutions down to 0.1 K, but Anderson’s results [8] showed m* was dependent on concentration. I used a theoretical model of the interaction potential to give m* as a function of 3 He concentration. This weakly interacting Fermi-Dirac model then fit all the specific heat data on dilute solutions. Entropy was calculated using Eq. (10). The refrigeration power is given as  ˙ r (T )/˙n = T s3d (Xd , T ) − s30 (T ) , Q (18) where s3d is the molar entropy of 3 He in 4 He in the dilute phase, X d is the 3 He concentration in the dilute phase along the phase boundary, and s30 is the molar entropy of pure 3 He. Equation (18) is approximated within 1% by 82T 2 J/mol 3 He between 0 K and 40 mK [11]. As 3 He diffuses through the stationary 4 He in the dilute side of the heat exchanger, its concentration decreases as the temperature increases as shown in Fig. 3a to maintain a constant osmotic pressure (more exactly a constant 4 He chemical potential μ4 ). The 3 He specific heat for constant μ is about five times that of pure 3 He, which means that 4 the returning dilute solution in the heat exchanger has considerable excess enthalpy that can be used for precooling electrical leads in the dilution refrigerator.

Fig. 3. (a) Phase diagram of 3 He-4 He solutions. The path followed by 3 He in the dilute side of the heat exchanger between the mixer and still is also shown (constant μ4). (b) Schematic of dilution refrigerator (MC = mixing chamber, HX = heat exchanger).

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6.3 Unique Heat Transfer Issues in a Dilution Refrigerator After completing my work on the properties of 3 He-4 He solutions, a colleague and I designed and built a DR that achieved a low temperature of 10 mK. The DR was patterned after that of Wheatley et al. [12] with continuous and step heat exchangers. The apparatus was used to study heat exchangers and Kapitza thermal boundary resistance [13, 14]. The Kapitza resistance varies as 1/T 3 , which dominates the thermal resistance in heat exchangers below about 100 mK. The step heat exchangers made from copper and filled with 44 µm sintered copper powder provided high surface areas to overcome the Kapitza resistance [12]. Finer powder would increase the area, but the viscosity of the solutions also increases at lower temperatures and would cause viscous heating. At 30 mK the viscosity of pure 3 He is about the same as that of water at 300 K. However, for temperatures below about 100 mK the thermal conductivity of liquid 3 He varies as 1/T. At 20 mK its thermal conductivity is about equal to that of brass. Simply drilling a hole through the sintered powder provides a low flow impedance while the high thermal conductance of the liquid conducts heat from the moving flow channel into the fluid within the sintered powder. In 1972 I was given a sample of 150 nm diameter silver powder by Dr. K. Nakamura of the Ulvac Corp. for DR heat exchanger research. Over the next year we experimented with a continuous coaxial HX in which we sintered the silver powder on the inside and outside wall of the inner CuNi tube. We limited the sintering temperature to 150 °C out of fear of losing surface area. Our test results with the sintered silver powder heat exchanger were disappointing. Performance was slightly worse than a conventional copper powder heat exchanger, although the liquid volume was less [15]. Soon after our failed tests with the silver powder heat exchangers, Frosatti and Thoulouse developed a very successful technique using even smaller 70 nm diameter Japanese silver powder [16, 17]. To achieve good thermal bonding between the powder and a CuNi foil, they first silver plated the foil and then bonded coarser 1 µm French silver to the silver-plated foil. Next, they bonded the 70 nm silver powder on top of the coarser powder using a sintering temperature of 200 °C for 40 min with a hydrogen atmosphere. The foil formed the boundary between the concentrated and dilute solutions. A low temperature of 2 mK was achieved. This type of silver powder heat exchanger is used today in most commercial dilution refrigerators.

7 Cryocooler Thermodynamics and Heat Transfer 7.1 Pulse Tube Cryocoolers In 1984 Mikulin reached 105 K with a pulse tube cryocooler in which they added an orifice inside the pulse tube [18]. Previously the lowest temperature achieved with the basic pulse tube (no orifice) was 124 K [19]. We then quickly put together an apparatus to experiment with this new orifice pulse tube cryocooler. Mikulin’s device was a GiffordMcMahon type of pulse tube, which uses valves, whereas ours was a Stirling type with no valves. We placed the orifice on the other side of the warm heat exchanger to allow the heat exchanger to act as a flow straightener. Our first attempt reached a low temperature

24

R. Radebaugh

of 60 K with a frequency of 9 Hz, average pressure of 1 MPa and a pressure ratio of 1.75 [20]. Both average and peak temperature profiles were measured with a thin foil thermocouple and compared with our model. The minimum gas temperature near the cold end showed the expected undershoot from the cold HX temperature, which provides cooling as the cold gas flows back through the HX. Our initial model of the orifice pulse tube refrigerator (OPTR) used the first law of thermodynamics in which time-averaged enthalpy flow in the pulse tube was calculated [20]. Figure 4a shows the geometry considered, and Fig. 4b shows various energy flows as a function of position. The refrigeration power at the cold end is given by

  ˙ c = H˙ − H˙ + W ˙ c, (19) Q r



 where H˙ is the time averaged enthalpy flow in the pulse tube, H˙ r is the time-averaged ˙ c is the cold end expansion work, which for a enthalpy flow in the regenerator, and W pulse tube is zero. Flow to the right is chosen as positive. For the Stirling or GiffordMcMahon cryocooler the displacer extracts work from the cold end, so the refrigeration  power is still given by Eq. (19) but H˙ becomes zero. The time-averaged enthalpy flow over one cycle is given by

 H˙ =



τ mhdt ˙ =

1 τ

mdh. ˙

(20)

0

In an ideal regenerator with an gas the temperature at any location is constant

ideal  (no temperature oscillation), so H˙ r = 0, but for a real regenerator and/or real gas the enthalpy flow is finite and gives the regenerator loss. No pressure or temperature oscillation occurs after the orifice, which means the time-averaged enthalpy flow after the orifice is zero. No work is extracted either. Thus, the heat rejected from the pulse  ˙ h is exactly H˙ and is easily measured with water cooling of the warm tube warm end Q heat exchanger. The first-law model discussed above provides a good description of the pulse tube cryocooler but not a complete description. Later, we discussed the thermodynamics of oscillating flow using both first and second laws [21]. The second law entropy balance on the cold end gives

  ˙ c /Tc ) + S˙ irr = S˙ − S˙ , (Q (21) r



 where S˙ and S˙ r are the time averaged entropy flows in the pulse tube and the regenerator, respectively. Note that in Fig. 4b the entropy flows are negative, which means the flow is toward the regenerator warm end. ˙ c from Eq. (19) and Eq. (21) gives Combining the first and second laws for Q

  





 ˙ c = H˙ − Tc S˙ − H˙ − Tc S˙ − Tc S˙ irr . (22) W r r The maximum power that can be extracted is for a reversible, isothermal process, in which case



 

 ˙ − G ˙ = P V˙ − P V˙ , ˙ rev = G (23) W r r

Thermodynamics and Heat Transfer in Cryogenics: A Personal View

25

 where P V˙ is known as the acoustic power and equivalent to the Gibbs free energy flow

 ˙ . For the case of a pulse tube, W ˙ rev = 0, so the acoustic power simply passes through G the cold end with no change. For the case

of a Stirling or GM cryocooler where there is a displacer or expander at the cold end, P V˙ = 0 beyond the moving boundary because no flow occurs past the boundary. In that case the reversible work that can be extracted is the acoustic power flow coming in from the regenerator. Thus, the acoustic power (Gibbs free energy) is the potential to do work whenever a moving boundary (displacer) is inserted into the stream to block the flow. That displacer can also be inserted at the pulse tube warm end, as is done with a warm displacer pulse tube.

Fig. 4. (a) Model of the pulse tube cryocooler showing different components and the energy flows. (b) Energy flows vs. position. Positive numbers indicate flow to the right [20, 21].





 Also shown in Fig. 4b is the exergy flow E˙ = H˙ − T0 S˙ in the perfect pulse tube cryocooler. The sudden drop at the cold end is the exergy output for the cold end cooling. The exergy destruction at the orifice is wasted, but if the orifice were replaced with a displacer, the extracted work (exergy) could be fed back and reduce the input ˙ c. power from the compressor to the Carnot value of [(T0 /Tc ) − 1]Q

8 Heat Exchanger Design Thermodynamics gives information about energy flows and the process efficiency. However, it gives little or no information on the system size. The cold finger size usually is dominated by the heat exchanger/regenerator size, so here we explore the relationship between cryocooler performance (efficiency) parameters and heat exchanger dimensions. Cryocooler efficiency is a function of the heat exchanger losses, such as heat exchanger ineffectiveness, pressure drop, conduction loss, and radiation loss (small systems). Important size parameters are configuration (tubes, screen, spheres, etc.), crosssectional area, length, and hydraulic diameter. The calculation of size from a given performance is called a sizing calculation, whereas the calculation of performance from

26

R. Radebaugh

a given size is called a performance calculation. Here we focus on the sizing calculation. Performance and size can be related through only two thermodynamic parameters (specific refrigeration power and HX maximum ineffectiveness) and two heat transfer parameters (the Poiseuille number and the ratio of Colburn j-factor to the friction factor). The approach shown here should be considered as an approximate solution because some simplifications are made, such as using average densities and assuming balanced enthalpy flows in the heat exchanger. The approach also assumes a steady mass flow, but it can be applied approximately to regenerative systems by using the peak flow and pressure drop in the equations. Even though the method provides only approximate results, it gives important insight into scaling, and it is useful as a first iteration for more accurate numerical calculations. These equations would have been very helpful to me for my liquid air science fair project. The method begins with the expression for the pressure drop in a flow channel P =

˙ g )2 L 2fr (m/A , ρDh

(24)

where f r is the Fanning friction factor, Ag is the gas cross-sectional area, L is the length, ρ is the average density (or log-mean for large changes), and Dh is the hydraulic diameter. In most compact HXs flow is in the laminar region in which f r = b/N Re , where b is the Poiseuille number and N Re is the Reynolds number. The second important heat transfer parameter is the ratio of heat transfer to momentum transfer, which is nearly independent of N Re and greatly simplifies the calculations presented here. That ratio, known as the 2/3 Reynolds analogy and given by α = NSt NPr /fr , is approximately 0.05 for a sphere bed, 0.06 for stacked screen, 0.28 for tubes, and 0.34 for parallel plates [22]. Here N St is the Stanton number and N Pr is the Prandlt number. The first essential thermodynamic parameter needed for the calculations is the specific gross refrigeration power ˙ r /m qr = Q ˙ = w + (hT )min ,

(25)

where w is the specific expansion work and (hT )min is the minimum isothermal enthalpy difference between the high- and low-pressure streams over the temperature range of the heat exchanger. The enthalpy term applies to JT cryocoolers. The second important thermodynamic parameter is the maximum HX ineffectiveness λmax = qr /(hh − hc )min ,

(26)

where hh and hc are the specific enthalpies at the hot and cold ends of the HX. There are many steps involved in deriving the equations given below that relate the HX size to cryocooler performance parameters [22, 23]. The gas cross-sectional area for minimum volume can be expressed as  1/2 2/3 NPr Ag = , (27) ˙ hx /Q ˙ r) m ˙ 2αρλmax P0 (P/P0 )(Q ˙ hx is the heat exchanger loss associated with its ineffectiveness. The area here is where Q for one side of a recuperative HX. The performance parameters in this equation are the

Thermodynamics and Heat Transfer in Cryogenics: A Personal View

27

˙ hx /Q ˙ r ). For regenerators the pressure amplitude P1 needs to be losses (P/P0 ) and (Q substituted for the average pressure P0 in these equations. Note that flow rate does not occur inside the brackets in Eq. (27), so Ag should scale as the flow rate. Equation (27) also shows that higher α, ρ, λmax , and P0 all lead to a smaller cross-sectional area. The HX length should be  (Ag /m)(1 ˙ − ng ) kdT L= , (28) ˙ cond /Q ˙ r) ng qr (Q where ng is the channel porosity. For screen or sphere packings the thermal conductivity ˙ is independent of flow must take into account a degradation factor [26]. Because (Ag /m) from Eq. (27) the length is independent of flow. Each side of a recuperative HX may ˙ but one can adjust the conduction have a different length because of a different (Ag /m), loss on each side to provide the same length. The optimum length shown by Eq. (28) can become a problem and require adjustment for the case of very large systems where the volume of the end manifolds start to dominate the volume of the HX. The hydraulic diameter of the HX channel becomes (for Dh independent of Ag )  Dh =

1/2  2bμ(1 − ng ) kdT , ˙ cond /Q ˙ r) ng ρqr P0 (P/P0 )(Q

(29)

where μ is the viscosity. Note that Dh is also independent of flow rate.

9 Minimum Input Power for 80 K and 4 K Cryocoolers 9.1 Scaling Laws for Small Sizes The size calculations discussed above indicate that as the input power for a cryocooler is reduced and the loss fraction is held constant, only the cross-sectional area is reduced. With that scaling law there should be no minimum input power for a given 2nd law efficiency. What that scaling law neglects is radiation and efficiency of the compressor. Power through the HX scales with the cross-sectional area, which for a planar geometry, scales as the width. For that geometry the radiation loss scales with width also. However, for tubular geometry the surface area scales as diameter whereas power scales as diameter squared, so radiation can become the dominate loss at small sizes for tubular geometry. For that case there should be some theoretical minimum input power below which the low temperature cannot be achieved. In 1999 I prepared an efficiency map of cryocoolers available at that time that showed a general trend of Stirling and pulse tube cryocoolers maintaining 2nd law efficiencies of several percent down to 1 W input power [25]. However, more recent data cast doubt on that trend. 9.2 Minimum Input Power—A Challenge A review of some of the smallest rotary Stirling cryocoolers shows that the minimum input power to achieve 80 K is currently about 2 W although there is some uncertainty in

28

R. Radebaugh

these estimates because of the need to extrapolate from actual measurements. Achieving 80 K with a minimum input power may require a unique combination of efficiency (thermodynamics) and size (heat transfer). I offer a challenge to the cryocooler community to see who can be the first to achieve 80 K with less than 1 W of input power. The cooling power can be essentially zero but just enough to cool a thermometer. Applications for such a cooler may be limited at this time, but as detector advances are made, there may be an increasing demand for such cryocooler types. A similar situation occurs for 4 K cryocoolers. The lowest input power currently used for such a temperature is about 250 W [26]. A lower limit may be around 100 W for a cryocooler designed specifically for such conditions. Naturally such low-power cryocoolers will be small, but they may not necessarily be the minimum size.

10 Conclusions The large temperature difference between cryogenic systems and ambient require careful attention to system efficiency and heat flows. The first and second laws of thermodynamics are powerful tools for the design of efficient systems. The parameters associated with reversible work, such as Gibbs free energy and exergy are shown to be useful for setting theoretical efficiency limits. The use of thermodynamics and heat transfer in a wide variety of cryogenic systems were discussed. The systems discussed here are ones the author has been involved with over a period of 65 years, included air liquefaction, phase equilibria, chemical thermodynamics, superconductivity, dilution refrigerators, regenerators, and cryocoolers.

References 1. Bejan, A.: Fundamentals of exergy analysis, entropy generation minimization, and the generation of flow architecture. Int. J. Energy Res. 26, 545–565 (2002) 2. Grassmann, P.: Zur allgemeinen Definition des Wirkungsgrades, Chem.-Ing.-Technik 22(4), 77–80 (1950) 3. Westrum, Jr., E.F.: Application of cryogenic calorimetry to solid-state chemistry. In: Advances in Cryogenic Engineering, vol. 7, pp. 1–15. Plenum Press, New York (1962) 4. Radebaugh, R., Keesom, P.H.: Low-temperature thermodynamic properties of vanadium. I. Superconducting and normal states. Phys. Rev. 149(1), 209–216 (1966) 5. Radebaugh, R., Keesom, P.H.: Low-temperature thermodynamic properties of vanadium. II. Mixed state. Phys. Rev. 149(1), 217–231 (1966) 6. Walter, G.K., Fairbanks, W.M.: Phase separation in He3 -He4 solutions. Phys. Rev. 103(1), 262 (1956) 7. Edwards, D.O., et al.: Solubility of He3 in liquid He4 at 0 K. Phys. Rev. Lett. 15(20), 773–775 (1965) 8. Anderson, A.C., Roach, W.R., Sarwinski, R.E., Wheatley, J.C.: Heat capacity of dilute solutions of liquid He3 in He4 at low temperatures. Phys. Rev. Lett. 16(7), 263–264 (1966) 9. London, H., Clarke, G.R., Mendoza, E.: Osmotic pressure of He3 in liquid He4 , with proposals for a refrigerator to work below 1 K. Phys. Rev. 128(5), 1992–2005 (1962) 10. Hall, H.E., Ford, P.J., Thompson, K.: A helium-3 dilution refrigerator. Cryogenics 6, 80–88 (1966)

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11. Radebaugh, R.: Thermodynamic properties of He3 -He4 solutions with applications to the He3 -He4 dilution refrigerator. NBS Technology Note 362, Washington, DC (1967) 12. Wheatley, J.C., Vilches, O.E., Abel, W.R.: Principles and methods of dilution refrigeration. Physics 4(1), 1–64 (1968) 13. Radebaugh, R., Siegwarth, J.D.: Dilution refrigerator technology. Cryogenics 11, 368–384 (1971) 14. Radebaugh, R., Siegwarth, J.D., Holste, J.C.: Heat transfer between sub-micron silver powder and dilute He3 -He4 solutions. In: Proceedings of the Fifth International Cryogenic Engineering Conference, Japan, pp. 242–245 (1974) 15. Radebaugh, R., Siegwarth, J.D., Oda, Y., Nagano, H.: Experiments with miniature heat exchangers for dilution refrigerators. In: Proceedings of the Fifth International Cryogenic Engineering Conference, Japan, pp. 235–237 (1974) 16. Frossati, G., Thoulouse, D.: Proceedings of the Sixth International Cryogenic Engineering Conference, Grenoble, p. 116 (1976) 17. Frossati, G.: Obtaining ultralow temperatures by dilution of 3 He into 4 He. J. de Physique Colloques 39(C6), 1578–1589 (1978) 18. Mikulin, E.I., Tarasov, A.A., Shkrebyonock, M.P.: Low temperature expansion pulse tubes. In: Advances in Cryogenic Engineering, vol. 29, pp. 629–637, Plenum Press, New York (1984) 19. Gifford, W.E., Longsworth, R.C.: Pulse tube refrigeration. Paper no. 63-WA-290 presented at the Winter Annual Meeting of ASME, Philadelphia (1963); Pulse tube refrigeration progress. In: Advances in Cryogenic Engineering, vol. 10, pp. 69–79, Plenum Press, New York (1965) 20. Radebaugh, R., Zimmerman, J., Smith, D.R., Louie, B.: A comparison of three types of pulse tube refrigerators: new methods for reaching 60K. In: Fast, R.W. (ed.) Advances in Cryogenic Engineering. ACE, vol. 31, pp. 779–789. Springer, Boston, MA (1986). https://doi.org/10. 1007/978-1-4613-2213-9_88 21. Radebaugh, R.: Thermodynamics of regenerative refrigerators. In: Generation of Low Temperature and its Application, Ohtsuka, T. and Ishizaki, Y. (eds.), Chapter 1, pp. 1–20. Shonan Tech Center, Kamakura, Japan (2003) 22. Radebaugh, R., Louie, B.: A simple, first step to the optimization of regenerator geopmetry. In: Proceedings of the Third International Cryocooler Conference, NBS Special Publication 698, Washington, DC, pp. 177–198 (1985) 23. Radebaugh, R.: Microscale heat transfer at low temperatures. In: Kakac, S., et al. (eds.) Microscale Heat Transfer Fundamentals and Applications, pp. 93–123. Kluwer Academic Publishers, New York (2005) 24. Lewis, M.A., Radebaugh, R.: Measurement of heat conduction through bonded re-generator matrix materials. In: Cryocoolers 12, pp. 517–522. Kluwer Academic/Plenum, New York (2003) 25. Radebaugh, R.: Development of the pulse tube refrigerator as an efficient and reliable cryocooler. In: Proceedings of the Institute of Refrigeration (London), vol. 96, pp. 11–29 (1999–2000) 26. Kotsubo, V., et al.: Compact cooling system for superconducting nanowire single photon detectors. IEEE Trans. Appl. Superconductivity 27(4), 9500405 (2017)

Large-Scale Refrigeration and Liquefaction

Design and Preliminary Test of a Helium Cryogenic Refrigerator for the High Energy FRagment Separator of HIAF Shaoqi Yang1 , Wei Pan1 , Rui Xue1,2 , Gang Zhou1 , Dongsheng Ni3 , Xiujuan Xie1(B) , and Wei Wu3 1 Department State Key Laboratory of Technologies in Space Cryogenic Propellants (Technical

Institute of Physics and Chemistry, Chinese Academy of Sciences), Beijing, China {yangshaoqi,panw,zhougang,xiexiujuan}@mail.ipc.ac.cn, [email protected] 2 University of Chinese Academy of Sciences, Beijing, China 3 Institute of Modern Physics, Chinese Academy of Sciences, 509 Nanchang Rd., Lanzhou 730000, China {nidongsheng,wuwei}@impcas.ac.cn

Abstract. As an important part of the High Intensity heavy-ion Accelerator Facility (HIAF), High energy FRagment Separator (HFRS) consists of a two-stage superconducting magnetic system, the pre-separator and the main-separator, which is cooled down by a helium cryogenic refrigerator. In this work, the helium cryogenic refrigerator is designed and optimized to satisfy the cooling capacity larger than 1kW @4.5 K and 10 kW @ (50 ~ 80 K) at the same time. Furthermore, this helium cryogenic refrigerator is integrated and preliminary tested. 300 ~ 4.5 K cool-down processes of cryogenic heat exchangers are presented. The cooling capacity of this refrigerator system is up to 1468 W @4.5 K in the commissioning, which will be improved in the further investigation. Keywords: High energy FRagment Separator (HFRS) · helium cryogenic refrigerator · Preliminary Test · cooling capacity

1 Introduction The new-generation High Intensity heavy-ion Accelerator Facility (HIAF) is being built by the Institute of Modern Physics, Chinese Academy of Sciences (IMP, CAS) [1, 2]. High energy FRagment Separator (HFRS), an in-flight separator at relativistic energy, is an important experimental terminal at the HIAF. It is characterized by high magnetic rigidity, large ion-optical acceptance, and excellent particle identification. Figure 1 shows the layout of the HIAF and the HFRS. The HFRS has a total length of 192 m. There are 11 superconducting dipoles cryostats and 13 superconducting multipole cryostats [3]. It needs transfer lines, valve boxes to distribute the liquid helium. And there are shields in the cryostats, transfer lines and valve boxes. The shields work at 50-75 K. The current leads for dipoles work at 4.5-300 K, and © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 33–40, 2023. https://doi.org/10.1007/978-981-99-6128-3_3

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Fig. 1. Layout of the HIAF [2]

the current leads for multipoles work at 50-300 K. Considering the safety margin and extra equipment to cool in the future, a helium cryogenic refrigerator with the cooling capacity larger than 1 kW @4.5 K and 10 kW @ (50 ~ 75 K) at the same time is needed for HFRS. Helium cryogenic refrigerators have been developed by Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, such as a 2 kW @ 20 K refrigerator [4, 5], a [email protected] K refrigerator [6, 7], a 40 L/h helium liquefier [8]. To meet cooling capacity requirement of the HFRS, a new refrigerator is designed and manufactured. This paper will present the model of the design and the results of the design and test.

2 Theoretical Models 2.1 Process Flow Diagram The process flow diagram of a refrigerator system is shown in Fig. 2. The cooling capacity produced by liquefied nitrogen (LN2), expanders and Joule-Thomson valve can be recovered by the seven heat exchangers. Thus, high pressure helium gas can be cooled from 310 K to 4.5 K step by step (nodes 1 → 21 as shown in Fig. 2). This process supply cooling capacity at 4.5K by Joule-Thomson valve. The outlet of the

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Joule-Thomson valve can be supercritical helium transferred to cryostat or saturated mixer getting into the dewar. Besides, the process supply cooling capacity at 50-75K in the outlet (node 6) of the third heat exchanger (HEX3) and return to the inlet (node 43) of the first stage turbine expander.

Fig. 2. The process flow diagram of the helium cryogenic refrigerator

2.2 Calculation Assumptions In this study, the calculation of the process parameters in the helium cryogenic system is carried out. The calculation assumptions are as follows: 1) System is at steady condition; 2) Isentropic efficiency of expanders are 65%.The efficiency of expanders are selected as to a practical helium refrigerator developed by TIPC; 3) Isentropic efficiency of compressors are 85% which is determined by the empirical experience; 4) The inlet and outlet pressure of the first stage compressor (COMP1) are 105 kPa and 405 kPa; the outlet pressure of the second stage compressor (COMP2) is 1750 kPa; 5) the outlet temperature of compressor (T1) is 310 K; the inlet pressure of the liquid nitrogen is (P40) 104 kPa; here, the numbers behind the character “P” and “T” are the node number shown in Fig. 2; 6) Pressure drops of low and high channels in heat exchangers are 1kPa and 2kPa, respectively. The consumed liquid nitrogen is calculated by converting into consumed electric power in the calculation as follows: ˙ LN Weq_LN = am

(1)

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Comprehensively considering the native market price of liquid nitrogen and electricity, the convert factor, a, is set as 0.48 in our study. 2.3 Calculation Models We use mass conservation and energy conservation to describe the devices. For the compressor, the isentropic compression efficiency and the power are calculated as follows, ηc =

hin −hout_s hin −hout

Wcomp = min (hout − hin )

(2) (3)

where ηc is the isentropic compression efficiency, hin and hout are the specific enthalpy of the inlet and outlet, hout_s is the specific enthalpy of the outlet in a isentropic process, Wcomp is the consumed power. For the turbine expander, the isentropic expansion efficiency is calculated as follows, ηT =

hin −hout hin −hout_s

where ηT is the isentropic expansion efficiency. For Heat exchangers, the energy conservation is calculated as follows,       mh hh_in − hh_out = mc1 hc1_out − hc1_in + mc2 hc2_out − hc2_in + Qleak

(4)

(5)

where mh is the mass flow rate of the hot stream, mc1 and mc2 are the mass flow rate of the cold stream in different pressure, Qleak is the heat leakage. For a tee, the mass conservation and energy conservation is calculated as follows, hin = hout_1 = hout_2

(6)

min = mout_1 + mout_2

(7)

For a mixer, the mass conservation and energy conservation is calculated as follows, min_1 hin_1 + min_2 hin_2 = mout hout

(8)

min_1 + min_2 = mout

(9)

For the load, the energy conservation is calculated as follows, Q = m(hin − hout )

(10)

To obtain a better performance, the optimal target is the ratio of the total cooling capacity to the total consumed power, COP =

Qtotal Wtotal

(11)

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where the Qtotal is the sum of the cooling capacity at 4.5 K and the cooling capacity at 50-75 K. The cooling capacity at 50-75 K must be convert to equal cooling capacity at 4.5 K. The equal factor we used is 0.1. And the W total is the sum of the consumed electric power by the two compressors and the consumed liquid nitrogen. The consumed liquid nitrogen must be convert to equal electric power. The equal factor we used is 0.48, where the unit of the consumed liquid nitrogen is liter per hour and the unit of the electric power is kilowatt. We use the Matlab software accessing the REFPROP dll to calculate the thermodynamic parameters of the helium.

3 Results and Discussion The optimal results of the refrigerator process for the HFRS are got though the calculation. The following Table 1 give the optimal results. The design results show that the design process can supply 1.7 kW @4.5 K and 16.5 kW @ (50 ~ 70 K) at the same time. It meets HFRS requirements. Besides the design process can supply 2.8 kW @4.5 K when the refrigerator works in the 4.5 K refrigeration mode. Also the TS diagram is given as the Fig. 3 shows and the numbers added to the lines have a one-to-one correspondence with the node numbers in the Fig. 2. Table 1. The optimal results of the process. 4.5K refrigeration mode

Application mode

User need

4.5K refrigeration capacity (W)

2823 W

1700 W

1500 W

50-75K refrigeration capacity (W)

0W

16.5 kW

15 kW

Mass flow of liquid nitrogen

191 L/h

191 L/h

-

Power consumption

809 kW

689 kW

-

Mass flow of helium in high pressure

325.6 g/s

325.8 g/s

-

Mass flow of helium in medium pressure

225.6 g/s

135.8 g/s

-

Based on the design result, a refrigerator is manufacture and integrated. Figure 4 shows the compressors, cold box and liquid helium dewar. The rotary screw compressors are all made in China by the Fujian Snowman Co., Ltd. The frequency of the input electricity to the motor of the first stage compressor (COMP1) varies from 25 Hz-50 Hz. The rated power of this compressor is 425 kW/380 V and the size is 5400 × 3000 × 3700. The second stage compressor (COMP2) soft start to realize the overload protection. The rated power of this compressor is 800 kW/10 kV and the size is 6000 × 3000 × 3600. The coldbox is designed and manufactured by the Technical Institute of Physics and Chemistry, CAS, including the key devices such as the turbine expanders and heat

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Fig. 3. The TS diagram of the refrigerator

Fig. 4. The compressors and the coldbox of the refrigerator

exchangers in the coldbox. The size of the coldbox is 3200 × 4500. The dewar is purchased from Wessington ltd. The capacity of the dewar is 2000 L (Table 1). Then we conduct the Preliminary test. The Fig. 5 shows the test results. The left figure shows the cool down process. The legend in the figure is the temperature of the different

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node in the coldbox and the numbers behind the character “T” are the node number shown in Fig. 2. The total time of the cool down is about 60 h. The steady cooling capacity is 1468 W @4.5 K when the refrigerator is operating at 4.5 K refrigeration mode. The right figure shows the heating power and the temperature in the dewar during the test.

Fig. 5. The cooling down process and cooling capacity of the refrigerator in the test

4 Conclusions In this study, a helium cryogenic refrigerator for HFRS is designed and optimized to satisfy the cooling capacity larger than 1.5 kW @4.5 K and 15 kW @ (50 ~ 80 K) at the same time. The design results are 1.7 kW @4.5 K and 16.5 kW @ (50 ~ 80 K) at the same time. It meets the HFRS requirements and can supply extra liquid helium to cool down other equipment. This helium cryogenic refrigerator is integrated and preliminary tested. The cooling capacity of this refrigerator system is up to 1468 W @4.5 K in the commissioning. Next, we will improve the refrigerator in the further investigation. Acknowledgments. The project is supported by the fund of The National Key Research and Development Program of Ministry of Science and Technology of the People’s Republic of China (Grant No. 2020YFB1506201) and the fund of State Key Laboratory of Technologies in Space Cryogenic Propellants (Grant No. SKLTSCP202011).

References 1. Yang, J.C., Xia, J.W., Xiao, G.Q., et al.: High intensity heavy ion accelerator facility (HIAF) in China. Nucl. Instrum. Methods Phys. Res. B 317, 263–265 (2013) 2. Sheng, L.N., Zhang, X.H., Zhang, J.Q., et al.: Ion-optical design of high energy fragment separator (HFRS) at HIAF. Nucl. Instrum. Methods Phys. Res. B 469, 1–9 (2020) 3. Dongsheng, N., et al.: Conceptual design of cryogenic system for HFRS of HIAF project. Presented at ICEC28-ICMC2022, Hangzhou, China (2022)

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4. Zhang, Yu., et al.: Performance analysis of a large-scale helium Brayton cryo-refrigerator with static gas bearing turboexpander. Energy Conv. Manage. 90, 207–217 (2015). https://doi.org/ 10.1016/j.enconman.2014.10.068 5. Lv, C., Qiu, T.N., Wu, J.H., et al.: Modeling and dynamic simulation of a large scale helium refrigerator. Phys. Procedia 67, 135–140 (2015) 6. Xue, R., Yang, S., Xie, X., et al.: Influence of key parameters on the performance of a helium cryogenic system in refrigeration and liquefaction modes. Cryogenics 121, 103386 (2022) 7. Xie, X., Yang, S., Deng, B., et al.: Integration and commissioning of a full localization [email protected] helium refrigerator at TIPC, IOP Conf. Ser.: Mater. Sci. Eng. 502, 012013 (2019) 8. Li, J., Liu, L.Q., Xu, X.D., et al.: Development of a measurement and control system for a 40l/h helium liquefier based on Siemens PLC S7–300. Phys. Procedia 67, 1181–1186 (2015)

Operation of the Upgraded Cryogenics Infrastructure of SM18, CERN Main Facility for Testing Superconducting Magnets, Power Links, and RF Cavities Nicolas Guillotin(B) , Thierry Dupont, Frederic Ferrand, and Antonio Perin TE-CRG-ML, CERN, Meyrin, Switzerland [email protected] https://home.cern/fr/

Abstract. SM18 is CERN’s main cryogenic facility for testing superconducting accelerator magnets, power links, and radiofrequency cavities. Its cryogenics refrigeration capacity and infrastructure have been upgraded to meet the intensive testing requirements and challenges of the High Luminosity Large Hadron Collider (HL-LHC) project. The R&D program of tests for magnets, power links, and RF cavities planned until 2025, and even beyond, requests an increased refrigeration and liquid helium capacity, more flexibility for its distribution and redundancy to fulfil the high availability requirements. In this perspective a new liquefier with a nominal capacity of 35 g/s was installed in 2020 in addition to the existing cryogenic infrastructure. This paper describes the newly installed configuration, its commissioning, the optimization, and performance ramp up of the cryogenics infrastructure of SM18 main test facilities at CERN after a dense operation period of one year. Finally, the SM18 cryogenics operation requirements for the coming years with even wider variety of scheduled tests are detailed. Keywords: SM18 · cryogenics · cold box · compressor · upgrade · ramp up · test benches · liquefaction · capacity · CERN

1 Introduction The SM18 cryogenic test facility at CERN is essential for the R&D program for the High Luminosity Large Hadron Collider (HL-LHC) project [1], the future upgrade of the Large Hadron Collider (LHC). The cryogenics infrastructure of the SM18 has been continuously and significantly developed to adapt to new requirements for radiofrequency (RF) superconducting cavities and cryomodules, for superconducting links and for superconducting horizontal and vertical magnet test benches.

N. Guillotin and T. Dupont—Cryogenics Engineer. F. Ferrand and A. Perin—Section Leader. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 41–47, 2023. https://doi.org/10.1007/978-981-99-6128-3_4

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For the HL-LHC project, the electrical connection between magnets in the LHC tunnel and the power converters in a new transverse tunnel will be supplied by supeconducting links (SClink) [2]. Each 80 m or 120 m SClink for the HL-LHC project will be tested at the SM18 in the coming year. The HL-LHC Inner Triplet string (IT String) [3, 4] is a new test system being installed at the SM18 to validate the collective behaviour of the inner triplet magnets for the future HL-LHC accelerator. The magnet circuits will be powered through a SClink up to the operational current and cooled to 1.9 K. The ramp up of SM18 tests for HL-LHC magnets and RF cavities imposed an upgrade of the cryogenics infrastructure to increase the liquid helium production capacity, to optimize its distribution and to provide redundancy. In this perspective, it was evaluated that an increase of the helium liquefaction capacity of 35 g/s would be necessary [5]. It was therefore decided in 2016 to procure a new helium liquefier [6], with its associated cycle compressor. The new cryogenics infrastructure was commissioned in 2020 with its ancillary equipment and the new liquefier was successfully used as the main source of liquid helium during a full year of operation in 2021.

2 Introduction 2.1 SM18 Test Facility Areas The SM18 hall includes four main test areas: the RF area with 2 bunkers for cryogenics modules and 4 cryostats for superconducting cavities, the horizontal magnets area with 10 test benches equipped with individual cryogenics feed boxes (CFB), the vertical magnets area with 5 cryostats and the area for links and string tests. 2.2 SM18 Cryogenics Infrastructure Before the upgrade, a cold box (CB-A) of 6 kW equivalent at 4.5 K, with liquid nitrogen precooling, supplied liquid helium with a maximum capacity of 25 g/s in pure liquefaction mode to a super insulated 25 000 L helium storage tank. This cold box includes seven turbines and a phase separator. The room temperature high pressure (HP) helium flow, 350 g/s at 19 bar, is provided to the CB-A by three helium screw compressors powered by high voltage motors with a combined power of 1.6 MW. The CB-A was the only liquid helium supply equipment for all the SM18 test benches before the installation of the new liquefier. The new liquefier cold box (CB-B) is equipped with eight turbo-expanders [6] and does not require liquid nitrogen boost. Two 80 K adsorbers in parallel allow for regeneration without interruption of production, as the specification was oriented towards maximum reliability. The high pressure helium is supplied by a screw compressor system with an electrical power of 1.4 MW and provides 300 g/s of helium buffer sets at room temperature, with four tanks of 80 m3 each dedicated to the 6 kW cryogenics station and four tanks of 80 m3 each for the new cryogenics station. The key figures for the cryogenics infrastructure including helium purifiers, balloons, pure helium tanks, cold/warm up units (CWU), warm pumping units (WPU) and main super insulated liquid helium tank are available in [5]. Flow at 16.7 bar and up to 350 g/s at 19 bar. The process compressors are connected to pure helium.

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2.3 Liquid Helium Supply and Return The cryogenics upgrade for the SM18 [5] was based on the use of the new liquefier directly connected to the common super insulated 25 000 L helium tank (see Fig. 1). The liquid helium of the main liquid storage operated at 1 565 mbar is distributed to test benches through transfer lines (TL). The CB-A connections to the test benches by means of the main valve box piping were maintained both for supplies and returns. The CB-A will be directly connected to the IT String test bench through a distribution valve box, while CB-B will be able to supply liquid helium to the other test stands through the common main liquid storage and the main valve box.

Fig. 1. Connections for the helium supplies and returns of the upgraded cryogenics infrastructure.

3 SM18 Test Benches Cryogenics Infrastructure Description 3.1 From Equipment Reception to Performance Tests The new compressor station and CB-B were received on CERN site respectively in April and May 2019. After their installation and connection, a progressive commissioning started with the suppliers for mechanical, electrical, control and tuning aspects. From August to September 2020 the commissioning of the new cold box connected to the compressor station started without connection to the main liquid helium storage in order not to disrupt the tests performed in parallel. Finally, the CB-B was connected to the main liquid helium storage, but disconnected from test benches and CB-A. The operation of the 20 K and 80 K adsorbers regeneration was validated, especially the test of the manual online swap between the two 80 K adsorbers. The difficulties related to the compressor station full mode regulation stability were solved and optimizations for management of the mixed gas returns were applied. These validations allowed for a progressive ramp up of operation of the compressor station up to the nominal level. Then acceptance tests were finalized on September 3rd, 2020 demonstrating a production capacity higher than the specified value of 35 g/s in pure constant level liquefaction mode, leading to handover to operation in September

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and progress towards full operation until end of December 2020. The operation of the tests benches was performed in January 2021, starting with the new cryogenics structure as main liquid helium supply source. 3.2 Liquid Helium Supply and Return Optimization and performance ramp up of the new SM18 cryogenics infrastructure have been performed after internal handover from the project to the operation team. Compressor Station Upgrades. A process overhaul has been applied to improve the compressor start sequences, especially to optimize the valve settings for the control of the HP oil differential pressure and to avoid recurrent stops during the restarts. Initially, about 30% of the compressor restart attempts failed due to such stops before the process upgrade. The low-pressure (LP) regulation loops, the positioners and the actuators have also been largely modified to obtain more stable operation. In terms of process, an automated limitation has been implemented on the LP side to avoid gas loading from the pure helium buffers when the LP is already over its set point. That limitation on the load valve guarantees a smoother operation and reduces the repeated manual tunings. Cold box Upgrades. Several hardware and process upgrades were applied for the adsorbers system. A significant leak in the insulation vacuum appeared on adsorber 1 during its first cool down. The leak was fixed during an intervention by the supplier. On adsorber 2, a leak was also present at cold on the inlet valve, leading to an increase from 7.0 10–6 mbar up to 2.8 10–5 mbar of the vacuum cold box insulation, causing initially some degradations in the performance. A reduced in-line leak is still present after the repair and it does not impact the operation. The management of the cold gas return to the cold box was also totally reviewed to adapt the helium returns as a function of temperatures of the gas from test benches and from the main liquid helium storage. Finally, an important flexibility in terms of operation has been reached by implementing an adaptive liquid production as a function of the pure helium volume available in the buffers and the liquid level in the liquid helium storage. If the volume of pure gaseous helium in the buffers is sufficient and if the level of the liquid helium storage is below 60%, then the cold box operates in full power mode. Otherwise, the cold box operates in economy mode. The implementation of a temperature controller on one of the last turbo-expanders gives the possibility in full power mode to regulate its output temperature as a function of the pressure in the compressor buffer line.

4 Liquefaction Performance with the New Process Compressor Station and Cold Box Compared to Previous Figures In order to compare the performances provided by the two cold boxes CB-A and CBB, it is essential to indicate the main context of operation at the SM18 in 2020, then in 2021.

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4.1 Main Operation Periods Events in 2020 and 2021 at the SM18 Full operation with the CB-A started on January 13th, 2020 with no stop until a COVID19 lockdown period. No SM18 tests were performed from March 17th to May 11th. The operation restarted successfully until end of June, as a clogging was detected at the inlet filter of the first turbine. The nominal operation re-started seven days later. Beginning of September 2020, the CB-A was temporarily disconnected from the main 25 000 L liquid helium tank during five days in order to carry out the performance tests of the new cold box. Then, the CB-A was re-connected to the main liquid helium storage and nominal operation was resumed until the official CERN year end closure in December 2020. The operation with the new cold box started on February 1st, 2021 after the annual SM18 maintenance period. No stop occurred and no difficulties were encountered for the operation of the new liquefier until a major flooding event in the area damaged the main 25 000 L super insulated liquid helium tank. It led to a significant unavailability for all the tests from June 4th to end-July. A 10 000 L helium storage was connected in replacement of the 25 000 L helium storage. The tests restarted until August 9th, as a leak was detected in the common insulation vacuum of the main valve box, due to severe vibrations on the recently renewed return line from the 10 000 L helium storage to the valve box. Repairs were promptly applied in five days and the SM18 cryogenics operation did not stop before the official CERN year end closure in December 2021. 4.2 Liquid Helium Production Capacity A comparison of the estimated liquid helium consumption by the test benches in 2020 and in 2021 is presented in Fig. 2. The main shutdown periods appear for each year. The graphic underlines important weekly variations related to the test benches liquid helium consumption as a function of: the results of the number of tests to be performed, the availability of the magnets, the cavities, the cryomodules and the links. Intraday tunings of the cryogenics parameters are also permanently applied by the operators to optimise the capacities for the test benches, especially to recover normal situation after strong quenches on horizontal test benches or on major liquid helium consumers. The key figures concerning the SM18 liquid helium capacities are summed up in Table 1 for 2020 and 2021. Values from Table 1 report on equivalent average liquid helium consumption by the test benches, while peak values highlight the additional capacity provided by the new liquefier. In order to evaluate with more precision the SM18 performances with the CBA in 2020 and with the new liquefier in 2021, it is useful to focus on two periods with similar continuous operation. During the first chosen period of 34 days, from February 2nd to March 12th (see Fig. 2), the average liquid helium consumption by the test benches was equivalent in 2020 and 2021, respectively 18 g/s with CB-A and 20 g/s with CB-B. During this period, the new cold box CB-B was connected to the main 25 000 L liquid helium storage and optimizations of the cold box parameters were still ongoing. During the second selected period (see Fig. 2) with continuous run of 65 days, from September 7th to December 11th, the average liquid helium consumption by the test benches was 25.9 g/s in 2020 with CB-A. But in 2021, with the new liquefier CB-B

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Fig. 2. SM18 liquid helium consumption in g/s by the test benches. In 2020 with CB-A (green curve), in 2021 with CB-B (blue curve). Table 1. Main figures for the SM18 operation with liquid helium supplied in 2020 and 2021. 2020 with CB-A

2021 with CB-B

Number of days of operation at SM18

274 days

262 days

Average liquid helium consumption by the test benches in g/s

18.0 g/s

19.6 g/s

Peak liquid helium consumption by the test benches in g/s

27 g/s

36 g/s

Total number of tests (*for the HL-LHC project)

50 (*34)

52 (*27)

Liquid helium capacity dedicated to HL-LHC tests in %

67%

59%

it rose up to 34.3 g/s, in spite of a reduced available liquid helium storage as the new liquefier was connected to a 10 000 L liquid helium storage following the flooding incident of June 2021. The global annual availability in 2020 reached 96.8% if the lock down period (55 days) related to COVID-19 restrictions is excluded from the study. In 2021 the global annual availability reached 99.9% if the long stop (57 days) for the main liquid helium tank issue is excluded from the study.

5 Conclusion and Outlook The installation of the new cryogenics station was performed in due time as well as its handover to operation. The specifications have all been respected and the requested functionalities have allowed for a full operation at the SM18 using that station as main liquefier without any perturbations related to that equipment all along 2021. The increased liquefaction capacity at the SM18 offers more margin for the magnet and cavity tests. It was a remarkable challenge for the CERN’s main cryogenic test facility. Several upgrades

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and commissioning of test benches will continue at the SM18 in 2022 and even further. The flexibility of operation with the two cold boxes will allow in 2023 to operate the IT String test bench using the CB-A in parallel of operation of the other SM18 test benches using the CB-B. From the new cryogenics infrastructure commissioning period in 2020 to the final optimizations during the first year of operation in 2021, about half of the new compressor and cold box stations process control has been modified and hardware adjustments were necessary. A very intensive period of optimization of the process in parallel of the dense operation at the SM18 led to a progressive increase of the liquefaction performance in 2021. No major stops of the SM18 liquid helium production are related in 2021 to the new liquefier, demonstrating its reliability and stability. A scenario with both cold boxes liquefying in a common liquid helium storage is considered, requiring specific studies and commissioning to manage the helium storage parameters, complex cold gas returns from test benches and pressures.

References 1. Rossi, L., Apollinari, G., Béjar Alonso, I., Bruning, O., Lamontm, M.: (CERN), HighLuminosity Large Hadron Collider (HL-LHC) Preliminary Design report CERN-2015–05, Geneva (2015) 2. Ballarino, A.: (CERN), Development of superconducting links for the Large Hadron Collider machine, superconductor science and technology, vol. 27, No 04, 044024 (2014) 3. Bajko, M.: (CERN), HL-LHC IT String, HL-LHC collaboration meeting (2016) 4. Bajko, M., Pojer, M.: (CERN), IT String and hardware commissioning, HL-LHC technical report, CERN yellow reports, CERN-2020–010 (2020) 5. Perin, A., Dhalla, F., Gayet, P., Serio, L.: (CERN), Upgrade of the cryogenic infrastructure of SM18, CERN main test facility for superconducting magnets and RF cavities, CEC (2017) 6. Treite, P., Messmer, M., Selva, P., Werz, R.: (Linde Kryotechnik AG), Design and procurement of the 35 g/s helium liquefaction system for the SM-18 facility at CERN, Oral presentation C1Or1A, CEC-ICMC (2021)

The Design Consideration and Optimization for the CEPC Cryogenic System Mei Li, Rui Ge(B) , Shaopeng Li, Ruixiong Han, Miaofu Xu, and Zhengze Chang Institution of High Energy Physics, CAS, Beijing, China [email protected]

Abstract. The Circular Electron Positron Collider (CEPC), with the low Higgs mass of 125 GeV as a Higgs Factory, has the advantage of a higher luminosityto-cost ratio and the potential to be upgraded to a proton-proton collider to reach unprecedented high energies and discover new physics. The Technical Design Report (TDR) of CEPC has begun from 2018. This paper focuses on the design and optimization of the CEPC cryogenic system, which are summarized and progressed as follows: (1) For the CEPC superconducting cavity cryogenic system side, the cooling scheme of the superconducting radio frequency (SRF) cryomodules of both the booster and collider has been modified from parallel to series to improve the cavity performance while reducing cost. The design of both cryomodules has been reviewed and improved based on the growing experience of the IHEP cryogenics group. Prototypes for booster and collider have been built in collaboration with the domestic qualified companies and has been tested in the Platform of Advanced Photon Source (PAPS) infrastructure with the aim of further improvements. (2) The new design of the superconducting (SC) magnet cryogenic system, following the progress of a multiple technology strategy under development for the interaction region (IR) magnets, with normal-conducting sextuples. The choice of using large refrigerators for operational stability and cost was appreciated. All in all, further TDR design is going on. Every member needs to work together to greatly push forward the work and complete the TDR design on time. Keywords: Circular Electron Positron Collider · cryogenic system · Technical Design Rate

1 Introduction The CEPC is a large international scientific facility proposed by the Chinese particle physics community. The CEPC, to be hosted in China in a circular underground tunnel of approximately 100 km in circumference, is a double-ring collider with electrons and positrons, and the detectors are installed at two interaction points (IPs). The primary physics goal is to use the CEPC as a Higgs factory. The CEPC overview is shown in Fig. 1. The CEPC will operate in different modes: H, Z, and W, ttbar, followed by SPPC. Conceptional design rate (CDR) has been finished in 2018 [1, 2], and TDR is expected to be finished in the end of 2022. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 48–55, 2023. https://doi.org/10.1007/978-981-99-6128-3_5

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a) CEPC different working modes

b) CEPC double ring layout

c) CEPC Linac injector of 20GeV Fig. 1. CEPC overview diagram

For the CEPC SRF cryogenic system, there are four cryo-stations with a capacity of 18 [email protected] K for each. The working temperature is 2 K. For the booster ring, there

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are 12 cryomodules for 1.3-GHz 9-cell cavities. And for the collider ring, there are 40 cryomodules for 650-MHz 2-cell cavities. For the magnets side, there are total 4 IR superconducting (SC) magnets working 4K, and 32 sextuple magnets (SM) working in the room temperature. Besides, the detectors adopt a thermosiphon cooling method. The CEPC cryogenic system overview is shown in Fig. 2.

Fig. 2. CEPC cryogenic system overview

2 CEPC Cryogenic System 2.1 CEPC SRF Cryogenic System The CEPC positive and negative electron beam streams are accelerated to 10 GeV after passing through the linear acceleration section, and then accelerated to 120 GeV by 48 1.3-Hz 9-cell superconducting cavities in the booster ring and 120 650-MHz 2-cell superconducting cavities in the collider ring. The cryogenic system process is described as follows: The 3 bara@5 K supercritical helium from the cold box is cooled to 2.2 K by the 2K heat exchanger; it is transported to the throttle valve in front of each module through the multiple transfer line (TL), where the temperature is reduced to 2K after throttling effect through the JT valve; the return helium gas is transferred to the cold end of the heat exchanger to cool the incoming hot flow. Two layers of helium thermal shields are designed at 5–8 K and 40–80 K in the multiple transfer line to minimize heat leakage during the 2.2 K liquid helium transfer. The cooling scheme of the SRF cryomodules of

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both the booster and collider has been modified, from parallel in CDR to series in TDR, to improve the cavity performance while reducing cost. The flow diagrams are shown in Fig. 3 and Fig. 4, respectively.

Fig. 3. Flow chart of SRF cryogenic system CDR Solution

Fig. 4. Flow chart of SRF cryogenic system TDR Solution

The CEPC SRF cryogenic system process calculation is simulated by the EcosimPro software, and the heat leakage of the transfer line is set to 0.15 W/m for the temperature zone @2 K. The process flow calculation diagrams of the CDR and the TDR are shown in Fig. 5. Besides, using the same method, set the heat leakage for 5–8 K and 40– 80 K thermal shields to 0.5 W/m and 2 W/m, respectively. In addition to steady-state calculations, the effects of non-steady-state processes such as warming-up (WU) and cooldown (CD) are usually considered in the design of cryogenic systems. 2.2 CEPC SC Magnets Cryogenic System On both sides of the collision points in the interaction region of CEPC, there are four IR SC magnets cryomodules working at 4 K, and 32 sextuple magnets working at room temperature. The detectors are installed at two IPs. The technical route has two ways: low temperature solenoids, and high temperature solenoid. The following Table 1 gives the estimated heat loads for the SC magnet side.

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Fig. 5. Process flow calculation diagrams of the CDR and the TDR

Table 1. Estimated heat loads for the SC magnets side Name

Unit

No

Heat load for each

Heat load

IR SC sextupole magnet

W

32

10

320

Valve Box of IR SC sextupole magnet

W

32

20

640

Current lead heat load of IR SC sextupole magnet

W

32

——

——

IR SC magnet

W

4

30

120

Valve Box of IR SC magnet

W

4

30

120

Current lead of IR SC magnet

g/s

4

0.5

2

Main distribution valve box

W

4

50

100

Cryogenic transfer-line

W

4000

0.3

1200

Total heat load @4.5K

W

2523.45

Heat load of thermal Shielding 40 ~ 80K

W

256.8

Total equiv. Heat load @4.5K

W

2780.25

[email protected]

[email protected]

Total equiv. Heat load @4.5K with multiplier 1.5 Cooling capacity of [email protected]

4170 W

2

Similarly, the SC magnet process flow calculation is simulated by EcosimPro software, and two schemes are adopted: 3 bar@5K without subcooling and subcooling. Assuming that the heat leakage of supercritical helium flow transfer line is 0.15 W/m for the 4.5 K temperature zone and the heat leakage of the 40-80K thermal shield line is

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2 W/m. Results can be obtained by: The diameter of supply TL is selected as DN25, and the diameter of return TL is DN40. In the case of subcooled supercritical liquid helium, the refrigerator needs to provide 88.5 g/s 3 [email protected] K helium, while in the case of without subcooling method, the refrigerator needs to provide 92.48 g/s 3 bara@5 K helium, as shown in Fig. 6.

Fig. 6. Process flow calculation of SC magnets cryogenic system

Currently, the requirement for IR sextuple magnets has been updated. The magnet bore and pole field are reduced. Conventional sextuple magnet technology can be used (FeCoV pole). Max strength 3886T/m2 is achieved with bore aperture 42 mm, and the max current is less than 200A. Therefore, the cryogenic system designed for the 32 SM magnets will no longer be needed, as shown in Fig. 7. And the estimated heat loads have been modified, as shown in Table 2. As referred to CDR, two individual refrigerators with 1.5 [email protected] K for the detector magnets will be employed. As a result, we will select two individual refrigerators (2 2/2.5 kW @ 4.5K) for the entire CEPC SC magnet cryogenic system when combining the IR SC magnets and detectors cryogenic large refrigerator employment.

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Fig. 7. Flow chart of SC magnets cryogenic system

Table 2. Modified heat loads for the IR SC magnets side in TDR /

/

Name

Unit No

4.5K main loop

Thermal Shielding 40 ~ 80K

Heat load for Heat load each

No

Heat load for Heat load each

IR SC magnet W

4

30

120

/

/

/

Valve Box of W IR SC magnet

4

30

120

/

/

/

Current lead of IR SC magnet

g/s

4

0.8

3.2

/

/

/

Main distribution valve box

W

4

50

100

/

/

/

Cryogenic transfer-line

W

100 0.3

30

100 1.5

150

Total heat load @4.5K

W

/

370

/

150

/

/

(continued)

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Table 2. (continued) /

/

4.5K main loop

Thermal Shielding 40 ~ 80K

Name

Unit No

Heat load for Heat load each

Total heat load @4.5K including current lead

W

/

/

370W + 3.2g/s /

/

150

Total heat load @4.5K

W

/

/

432

/

/

150

Coefficient

/

/

/

1

/

/

0.05

Total heat load @4.5K

W

/

/

432

/

/

7.5

Total equiv. Heat load @4.5K

W

439

Total equiv. W Heat load @4.5K with multiplier 1.5

659

No

Heat load for Heat load each

3 Conclusions Previous CDR and TDR work has provided a solid foundation for future design.For the CEPC cryogenic system TDR design, relevant aspects have been considered, including process scheme, layout, and key equipment design. Further TDR design is going on. The project team members need to work together to greatly push forward the work, and complete the TDR design on time. Acknowledgments. The project is supported by CEPC Scientist Workshop Fund.

References 1. CEPC Study Group. CEPC conceptual design report: accelerator 1. arXiv preprint arXiv:1809. 00285 (2018) 2. CEPC Study Group. CEPC conceptual design report: physics & detector, 2. arXiv preprint arXiv:1811.10545 (2018)

Safety Relief System Design for the DALS Test Facility Cryogenic System Xu Shi1(B) , Zheng Sun1,3 , Bihui Lai2 , Liangbing Hu2 , Guanglong Cui2 , Lei Xu1 , and Xilong Wang1 1 Dalian Institute of Chemical Physics, Chinese Academy of Sciences, Dalian, China

[email protected]

2 Institute of Advanced Science Facilities, Shenzhen, China 3 University of Chinese Academy of Sciences, Beijing, China

Abstract. A new project called Dalian Advanced Light Source (DALS) is proposed by the Dalian Institute of Chemical Physics (DICP), Chinese Academy of Sciences (CAS) and is subject to Chinese central government’s approval. To test SRF cavities and accelerator cryomodules, a series of test benches, including horizontal test bench (HTB), vertical test cryostat (VTC), injector test bench (ITB) and cryogenic test bench (CTB), are under construction in Dalian, China. The dedicated test facility cryoplant (TFCP) will provide cooling power for test benches through the test facility distribution box (TFDB). The cooling capacity of the TFCP is 370W@2 K. Amongst those key issues for the entire system, like reliability, efficiency, process arrangement as well as auxiliary design, top priority has been given to the safety relief system. Such a system shall be able to handle emergency scenarios include but not limited to degradation or sudden loss of vacuum, runaway fill, pipe rupture and rapid energy dissipation in cold helium. In this paper, safety relief solutions for DALS test facility cryogenic system have been introduced and evaluated. Sudden loss of vacuum to air will be viewed as the worst case during the sizing phase, to deter the niobium cavities from danger of over-design pressure, there also equips a decompression system for safety relief. Keywords: DALS Test Facility · Cryogenic System · Safety Relief

1 Introduction In order to produce high repetition rate and high brightness X-Rays through a Free Electron Laser (FEL) facility, Dalian Advanced Light Source (DALS) project is proposed by Dalian Institute of Chemical Physics (DICP), Chinese Academy of Science (CAS). The new linear accelerator will accelerate free electrons to energies of approximately 1.0 GeV in continuous wave (CW) mode. The Superconducting Radio Frequency (SRF) technology will be applied to the linear accelerator [1]. Eight cavities along with one superconducting magnet package are assembled in a cryomodule. There are ten 1.3 GHz cryomodules and two 3.9 GHz cryomodules that will be cooled in helium II bath. The 2 K cryostat will be protected against heat radiation by a nominal thermal intercept and a nominal thermal shield cooled to temperatures from 4.5 K to 8 K and from 40 K to 50 K, respectively [2]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 56–62, 2023. https://doi.org/10.1007/978-981-99-6128-3_6

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Key components of cryomodule mentioned above need to be tested before the final module assembly. The assembled cryomodules need to be qualified before installed in the LINAC. Therefore, a large test facility is constructing in Dalian, China, which includes a horizontal test bench, a vertical test cryostat, a beam test bench, and a cryogenic components test bench. Those test benches will be connected to a dedicated cryogenic infrastructure through a distribution box. Considering the cryogenic installations of DALS test facility have industrial dimensions, risk assessment including operation mishaps and other fault conditions is the most vital aspect that determines reliability, especially loss of beam vacuum to air can’t be ignored right from the beginning [3].

2 Safety Management 2.1 Helium Inventory The DALS test facility cryogenic system consumes substantial amount of helium as shown in Fig. 1. Over 4000 liters of liquid helium will be circulated among refrigerator and the distribution system at nominal operation condition. Half of cold helium is filled in the vertical test cryostat due to its large volume. The cryoplant also needs to serve beam and accelerator cryomodule test stand through TFDB. Approximately 200 liters of LHe will be deposited in the process lines. 50% loss of helium is assumed each year [4]. So 8000 L of liquid helium is the expected annual demand of the whole facility. Since the process is operated in a closed loop, the helium must be stored in warm tank set. Once the helium is contaminated, it will be stored in high pressure cylinder set. The boiling temperature and enthalpy of evaporation of helium is rather low, but the ratio of volume at room temperature to liquid state is extremely high, so it must be thermally isolated. Because even small amount of heat ingress would lead to rapid evaporation and pressure surge. Thus, it’s crucial to guarantee pressure safety of helium vessels.

Fig. 1. Helium inventory summary of DALS test facility cryogenic system

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2.2 Potential Failures Cryogenic system is the most sophisticated and versatile section of the DALS test facility, of which safety is paramount. Accordingly, key priority has been devoted to analyzing potential fault scenarios as depicted in Fig. 2. Potential emergency events include but not limited to: 1) 2) 3) 4) 5) 6) 7) 8) 9)

Sudden loss of insulation vacuum to air Helium flow to the insulation vacuum Sudden loss of beam vacuum to atmosphere Sub-atmospheric helium invade beam vacuum Energy dissipation caused by cavity (simultaneous) quench Pressurized helium fill to sub-atmospheric helium Warm helium flow to cold helium Helium flow to ambient surrounding Other venting cases due to maloperation

Fig. 2. Emergency events of cryomodule

2.3 Heat Flux Ingress Due to Air Condensation Sudden breakdown of vacuum to air usually requires the greatest discharging capability. Considering loss of beam vacuum to in-rushing air is also the severest scenario where air condensation on the inner wall of cavity string, while outer niobium surface exposing to superfluid helium simultaneously. Due to large temperature difference between atmosphere and superfluid helium, low latent heat, thermal conductivity of multiple insulation (MLI) and complex non-steady behavior, it’s inaccurate to estimate the heat input into the plant by theoretical calculation. So, measuring the heat flux by experiments is the most reliable method, which in turn will give accurate instructions for safety design, especially when involving loss of beam vacuum to air. Many researchers had performed heat flux measurement experiments on liquid helium cryostats and cryomodules. In different ways such researchers have sought to size

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an effective and advanced pressure relieving system together with the exploration of theoretical behavior and, importantly, assert measurement of the invading heat flux caused by air condensation. There are some verified parameters on maximum heat flux due to sudden loss of insulation vacuum to air so far, including 0.6 W/cm2 for the superinsulated bath [5] and 3.8 W/cm2 for uninsulated cryostat [5], 6.6 W/cm2 for uninsulated vessel [6]. To be more targeted, researchers in DESY conducted destructive tests to measure the heat loads for cryogenic components of XFEL prototype cryomodule [7]. Owing to the sophisticated structure of DALS test facility cryomodules, heat transfer into process lines and cavities within cryomodule is far different from single lines and cavities when loss of vacuum happens. As for components of cryomodule, 1.5 kW/cm2 for 40/80 K circuits in [7] is adopted by this paper for 50 K lines without MLI, half of which for 40 K lines with MLI from an engineering point of view, 2.3 kW/cm2 for 5 K circuits in [7] is introduced, this paper takes 2.5 kW/cm2 for 4.5/8 K circuits with some margin considerations, 6.5 kW/cm2 for 2 K lines with MLI and 23 kW/cm2 for naked cavity are the same with [7]. While 1 kW/cm2 in [7], 6 kW/cm2 in [5] and 38 kW/cm2 in [5] are employed by 40/50 K transfer lines with MLI, 4.5/8 K transfer lines with MLI and naked cavity in this paper, respectively. Table 1 summarizes the detailed heat flux parameters [5, 7]. Table 1. Heat flux of cryogenic components due to air condensation [5, 7]. Components

Circuit

Heat flux, kW/m2

Cryomodules

50 K naked

1.5 [7]

40 K with MLI

0.75 [7]

4.5/8 K with MLI

2.5 [7]

2 K with MLI

6.5 [7]

naked cavity

23 [7]

40/50 K with MLI

1 [7]

4.5/8 K with MLI

6 [5]

naked cavity

38 [5]

Transfer lines Vertical cryostat

2.4 Safety Relief Approach Safety relief architecture of cryomodules is displayed in Fig. 3. The primary principle is each circuit or vessel that may isolate cryogenic fluids should be equipped with safety valves. Since the pressure of target devices float around atmospheric pressure, a burst disc is not considered because it is likely to be exposed to reverse pressure, which may cause leakage. Given that the pressure of 2 K helium stored in the tank is about 31 mbar, which is lower than 1 bara. If SV1210 is blowdown or failed to reseat, air outside is very possible to invade the sub-atmospheric circuit of Line B to contaminate the helium. In order to

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prevent this situation, a helium buffer sealed by another valve is introduced with SV1210 placed inside. Helium at positive pressure in the header is filled to the buffer to surround the safety valve. Therefore, malfunction of the SV1210 would only lead to the ingress of helium into the pipes. Since the yield strength of niobium cavity is inversely proportional to temperature, the maximum allowable working pressure is 2 bara at room temperature, but 4 bara at cold conditions [8]. The set pressure of SV1210 is 1 brag but the relief capacity in cryogenic temperature is calculated at 4 bara. There is a decompression branch mounted in parallel on a common inlet pipework of the safety valve SV1210, which consists of a digital valve PV1210 actuated by pneumatic controller and a pressure differential switch (PDS). When the pressure reaches the threshold of 1.8 bara, a signal will be conveyed to the control center to move the digital valve to fully open position and drain the boiling helium to the header [9]. Other circuits, like Line A, C&D and E&F, employ safety valves to handle emergency events and their outlet pipework is collected by a manifold to discharge the helium to the outdoors. For safety valve of cryostat vacuum would vent directly into the indoor atmosphere as the possibility of inner pipe rupture is very low.

Fig. 3. Simplified safety scheme of DALS test facility

Based on the heat flux in Table 1, the total heat input Q can be calculated. Set pressure is subject to the protected components. According to Pressure Relief Device Standards recommended by Compressed Gas Association, Inc. The flow capacity of safety valve is largest at the relieving temperature where the following equation is maximum for the specified set pressure. √ v(T )  (1) f (T ) =  dh(T ) v dv p

Safety Relief System Design for the DALS Test

Where v is the specific volume, kg/m3 and v



61



dh(T ) dv p

is the specific heat input, J/kg.

And the required vent flow rate m of the safety valve can be calculated as: Q  m=  ) v dh(T dv

(2)

p

The specific relief system parameters are summarized in Table 2. The inlet nonrecoverable pressure drop should not exceed 3% of the set pressure to avoid chattering [10]. Consequently, safety valves of larger inlet dimensions are selected, especially for Line A. However, keeping the pressure loss below 3% might be progressively difficult at pressure as low as Line B safety valve. Because the required vent mass flow rate of Line B has some margin, along with the introduced decompression system, pressure drop of Line B safety valve up to 0.74 bara is also acceptable. Table 2. Specific parameters of safety relief system Circuit

Operating Temp. K

Set pressure, barg

Relieving Temp. K

Required vent rate kg/s

Installed vent rate, kg/s

Inlet pressure drop, bara

Line A

2.2

16

11.4

0.857

4.7

0.057

Line B safety valve

2

1

5.5

7.2

7.42

0.74

Line B decompression

2

0.5

/

/

/

/

Line C&D

4.5–8

16

11.4

0.403

0.53

0.42

Line E&F

40–50

16

40

0.057

0.238

0.13

All safety valves are flanged connection to guarantee good leak tightness. The inlet of safety valves that connect to warm regions tend to use O-ring sealing, however, safety valves connected to cold volumes will pick PCTFE sealing gasket.

3 Conclusion The safety relief system of DALS test facility cryogenic system has been described. Based on the cryogenic fluid inventory, process parameters and potential failures, the proper and prudent design of helium relief system was carried out. Heat flux due to air condensation adopted by this paper is summarized. Process lines are equipped with safety valves only, while two-stage protection was introduced to prevent RF cavities from overpressure. The civil engineering of DALS test facility is under way. Safety valves had been ordered and will be delivered in August this year, cryogenic system commissioning is expected to start at the beginning of 2023.

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Acknowledgments. This work has been funded by Dalian local government. The authors would like to thank and acknowledge Prof. Guy Gistau Baguer for his constructive comments. We also appreciate colleagues from IASF, IHEP, and SHINE shared their knowledge and offered very valuable references.

References 1. Sun, Z., Huang, L., Shi, X., Wang, X.L.: Conceptual design of DALS test facility cryogenic system. CEC.ICMC 2021 (2021) 2. Theilacker, J., et al.: Guidelines for the design, fabrication, testing, installation and operation of SRF cavities. In: AIP Conference Proceedings, vol. 1218, no. 1, pp. 1223–1230. American Institute of Physics, April 2010 3. Otto, T.: Safety for Particle Accelerators, p. 148. Springer, Cham (2021) 4. Arnold, P., Boros, M., Nilsson, P.: ESS cryogenic controls design. EPJ Techn. Instrum. 8, 8 (2021) 5. Lehmann, W., Zahn, G.: Safety aspects for LHe cryostats and LHe transport containers. In: Proceedings of the 7th International Cryogenic Engineering Conference, London, July 1978 6. Rucinski, R., Bell, D.: D0 Solenoid Upgrade Project: Solenoid Insulatiing Vacuum Vessels; Relief Path Capacity Calculation (No. FERMILAB-D0-EN-345). Fermi National Accelerator Lab. (FNAL), Batavia, IL (United States) (1993) 7. Boeckmann, T., et al.: Experimental tests of fault conditions during the cryogenic operation of a XFEL prototype cryomodule. In: Proceedings of International Cryogenic Engineering Conference, vol. 22, pp. 723–728. HM Chang (2008) 8. Knobloch, J., Anders, W., Xiang, Y.: Cryogenic considerations for CW operation of TESLAtype superconducting cavity modules for the BESSY FEL. In: Proceedings of EPAC (2004) 9. Wiseman, M., Bundy, R., Kelley, J.P., Schneider, W.: CEBAF cryounit loss of vacuum experiment. In: Kelley, J.P. (ed.) Applications of Cryogenic Technology, vol. 10, pp. 287–303. Springer, Boston (1991). https://doi.org/10.1007/978-1-4757-9232-4_24 10. Grohmann, S., Süßer, M.: Conceptual design of pressure relief systems for cryogenic application. In: AIP Conference Proceedings, vol. 1573, p. 1581 (2014) (2015)

Upgrade of the Ex-LEP Helium Refrigerator for HL-LHC at CERN LHC Point 4 Emmanuel Monneret1(B) , Serge Claudet1 , Vanessa Gahier1 , and Robert Herrmann2 1 Technology Department, Cryogenic Group, CERN, Geneva, Switzerland {Emmanuel.Monneret,Serge.Claudet,Vanessa.Gahier}@cern.ch 2 Machinery Department, Linde Kryotechnik AG, Pfungen, Switzerland [email protected]

Abstract. In the context of HL-LHC project, the duty of the existing LHC helium refrigerators will face an increase of beam induced heat loads in the arcs. The cryogenic sectors around the LHC Point 4 (P 4), are considered in priority since the respective refrigerators must cover in addition to the superconducting (sc) magnetic system the sc Radio Frequency (SRF) cavities. Initial studies demonstrated that the upgrade of one of the existing two cryoplants at P 4 would prevent adding a third new refrigerator as foreseen in the baseline scenario. Complementary studies demonstrated that only one helium refrigerator delivered in 1993 by Linde Kryotechnik® needed to be upgraded, representing a particular challenge as it had already been modified twice to match the LEP 200 and the LHC project requirements. This paper will present the upgrade of ex-LEP helium refrigerator, located at LHC P 4 with split cold boxes (surface and underground) emphasizing the challenges to obtain the required additional refrigeration capacity equivalent to 2 [email protected] K with respect to the existing plant capacity of 16.5 [email protected] K. The major action has been to integrate and replace all expansion turbines using new state-of-the-art turbines, while verifying existing equipment manufactured since the nineties. The constraints, singularities of each surface and underground cold boxes and project risks assessment led to different technical solutions for the integration of the new turbines will be described. The performance obtained for the upgraded ex-LEP refrigerator will be presented as well as 18 months of operation facts with cool-down of the corresponding LHC sector and preparation for beams. Keywords: Refrigerator · Helium · Cryogenic

1 Introduction The High-Luminosity Large Hadron Collider (HL-LHC) is a major upgrade of the Large Hadron Collider (LHC) accelerator to increase ultimately the luminosity by a factor 7.5 with respect of the designed luminosity for LHC. The HL-LHC cryogenic configuration comprises (see Fig. 1):

© Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 63–69, 2023. https://doi.org/10.1007/978-981-99-6128-3_7

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• Two new refrigerators (type-G) to be installed respectively nearby the LHC collision Point 1 (ATLAS Experiment) and Point 5 (CMS Experiment) in order to handle the significant increase of dynamic heat loads, mainly due to secondary particles showers; • An upgrade of the ex-LEP refrigerator (type-A) located at LHC point 4, where the two cryogenic plants have to take the SRF modules heat loads in addition to the arcs. This proceeding will focus on the upgrade of the ex-LEP Refrigerator at LHC P 4 for HL-LHC providing the work performed and the performance obtained.

Fig. 1. Location of the ex-LEP refrigerator and the HL-LHC new refrigerators.

2 History and HL-LHC Final Design for LHC P 4 2.1 HL-LHC Design Choice and Refrigerator History from LEP to HL-LHC at LHC Point 4 D. Berkowitz [1] proposed two alternatives design and estimated for the alternative 2 (Fig. 2) the supplementary SRF cooling requirement of the LHC sector 3–4 for HL-LHC at + 2 [email protected] K with respect to the existing refrigerator design for the LHC capacity requirement of 16.5 [email protected] K. To reduce the project complexity and avoid the installation of a supplementary refrigerator, the alternative 2 design was implemented by CERN and the ex-LEP type-A refrigerator was proposed for an upgrade with an additional capacity of 2 [email protected] K to maintain an adequate margin on the refrigerator capacity. On the other hand, the type-B refrigerator was designed for LHC to 18 [email protected] K. It will be able to cover the SRF load (1.16 [email protected] K) with reduced margin and will benefit of a reduction of cooling requirement thanks to the new cooling capacity provided at Point 5 by the new HL-LHC refrigerator (see Fig. 2). 2.2 HL-LHC Capacity Requirements at Point 4 The Table 1 lists the detailed cryogenic capacity requirements for the ex-LEP refrigerator at P 4, considering the history of the ex-LEP refrigerators described by U. Wagner in [2].

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Fig. 2. Studies for possible refrigerators upgrade to comply with HL-LHC heat loads at P 4.

Sulzer-Linde® supplied the original type-A refrigerator in P4 for the Large Electron Positron Collider (LEP). To adapt to higher thermal loads for the LEP and then for the LHC, the type-A refrigerator was upgraded twice and commissioned in 1999 and in 2006. Table 1. Capacity requirements for the Ex-LEP Refrigerator from LEP to HL-LHC Machine

LEP

LEP2

LHC

HL-LHC

Refrigeration Load at 4.5 K [kW]

10.0

16.0

4.02

5.05

Non-isothermal load (4.5 K–300 K) [g/s]

13.0

13.0

27.0

29.8

6.7

6.7

11.7

17.9

Non-isothermal load (4.5 K–20 K) [g/s] Thermal Shield load (50–75 K) [kW] Equivalent capacity at 4.5 K [kW]

213.4

263.7*

31

31

16.5

18.5

*Return at 17 K due to 4.5 K gas helium coming from SRF

The main upgrade challenge for HL-LHC was to increase the ex-LEP refrigerator capacity above its design capacity (+12%) while increasing significantly the nonisothermal cooling between 4.5 K to 17 K (+24%), keeping most of the existing main equipment’s such as the compressor station or the main plate heat exchangers. 2.3 Context, Environment of the Ex-LEP Refrigerator The warm helium compressor station (QSCA) for the ex-LEP refrigerator is located in a surface building (SH4, see Fig. 3), with 2 compression stages, each stage composed of 4 helium-oil screw compressors. It was upgraded in 1999 to cope with the LEP and future LHC operation increasing the compressor unit from 5 to 8 leaving no space available for future upgrade of the compressor station. The ex-LEP refrigerator is split into two cold boxes connected by a vertical cryogenic transfer line (~140 m, see Fig. 3):

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• The surface cold box (QSRA) housing the main exchangers, four turbines and the 80 K and 20 K adsorbers provides helium down to 20 K. An additional box (QSKA), installed for LHC, hosts the nitrogen pre-cooler and the adsorber unit. Both are located in a surface building (SD4); • The underground cold box (QURA) is located in a cavern (US45) housing the cold exchangers, four turbines and the helium phase separator. It provides subcooled helium to the users.

Fig. 3. Configuration of the cryogenic system at P4.

The environment of the surface and underground cold boxes is different and required specific design and faced different challenges during the refrigerator upgrade.

3 The Modifications Preliminary study identified six (6) different options to achieve the capacity requirements defined for the upgrade in Table 1. All options were assessed with respect to several criterion such as technical challenges and engineering, installation risks, or from project risks perspectives with cost and schedule. Among the 6 options, and considering the complexity, CERN selected the one providing the highest performances without any modifications of the warm compressor station (keeping constant the LEP2/LHC compressors maximum flow and pressure). 3.1 Process Study To reach the required capacity, as CERN chose to not modify the compressor station, the upgrade consists to make a new process arrangement using the latest LKT® “TED-type” turbo-expanders. This new type of turbine has higher efficiency and wider operating range. In the surface cold box, the modification of the four turbines arrangement led to a better use of the heat exchange surface available. The first Brayton cycle operates

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between 165 K and 70 K; the second one is mainly used for the thermal shield load between 70 K and 33 K. The underground cold box includes three new “TED” turbines replacing the old “TGL-type” turbines. Only the internal part of the last Joule-Thomson turbine was modified to handle higher flow rate, as not much gain would have justified a change. The upgrade led finally to a significant increase of the cooling capacity of the turbines of the HL-LHC at 148 kW compared to the LHC configuration at 131 kW [2]. 3.2 Thermo-Mechanical and Integration Study Before launching the detailed design and manufacturing phases, the aim of the preliminary study was also to identify any bottleneck either on existing components or the integration of the new ones. Part of the existing components, the main aluminum plate heat exchangers were difficult to replace and even identified as a possible “project killer”. Their design adequacy with the new refrigerator process arrangement was carefully scrutinized, involving the original exchanger manufacturer for the hypothesis and the margin validation. Additional components such as the phase separator and its sub-cooler, piping, valves and the safety valves… were also verified. From integration point of view, 3D models of the two cold boxes (Fig. 4 and Fig. 5) were performed based on the original 2D drawings and successive upgrades documentation. The underground cold box faced the most challenging integration due to space and access limitation.

Fig. 4. 3D Model of the QSRA

Fig. 5. 3D Model of the QURA

Additional works have been performed to improve the operation reliability and availability of the refrigerator. The cooling water panels distribution have been fully modified to accommodate the new powerful turbines. Part of turbines instrumentation have been upgraded (temperature, speed sensors and transmitters). A cryogenic valve was relocated to enable each turbine strings to be isolated for maintenance while keeping the refrigerator in operation. To be complete, the update of the control software and all the documentation were modified.

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3.3 New Development The new TED turbine cartridge is associated to a new design of cold housing. The cold housing is welded to the internal pipework and to the top plate of the cold box. For the QSRA, external and internal access to the cold box was not an issue with sufficient free space to dismount old turbine cold housings and integrate the new ones (Fig. 6). For the QURA the internal space is overcrowded, and the cold box is installed on a steel frame structure 10 m above the slab of the cavern. One solution would have been to dismount part of the internal equipment such as cryogenic valves, the phase separator but with a lot of risks associated to such operation. Multiple scenarios were studied to finally come with an innovative way to update old TGL-type turbines to new TED-type turbines. LKT proposed to keep the original TGL cold housing and to include an adaptor to welcome the new TED turbine cartridges. This solution was supported by additional analysis to demonstrate the feasibility while keeping a similar level of turbine efficiency. It does not require extended work, risks and minimized the time of refrigerator unavailability. The upgrade of the ex-LEP refrigerator has demonstrated the feasibility to upgrade a cold box equipped with old TGL-type turbines using two methods according to the environment with similar performances.

Fig. 6. Difference of integration of new TED-type turbines for QSRA and QURA

4 Performance Test The performance test was managed in 2 phases. The first campaign was performed during summer 2020 to test the new turbine individually at its nominal operating conditions. This test showed that all turbines performed as expected reaching the nominal speed and the correct pressure ratio for the inlet nominal temperature. However, the lack of accuracy on the temperature measurements made difficult the precise evaluation of the turbine efficiency. Nevertheless, the analysis of all process conditions around the turbines allowed concluding with confidence that the turbines operate close to their expected design performances. The new installed “TED” turbine efficiency is anyway estimated to vary from 80% to 84%. The second test campaign, performed after the LHC cool-down in summer 2021, was done with LHC magnets in a cold state. The latter were used to test the upgraded

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refrigerator in a configuration enabling the variation of the non-isothermal load between 4.5 K and 20 K. The thermal shield load was simulated by circulating helium through the LHC thermal shield. This second test campaign was divided in two steps. The first step was to compare the upgraded refrigerator to a defined load scenario already tested early 2019 before the modification. The conclusion of this test was that under the same refrigerator conditions, with a compressor station at fix High Pressure (HP) of 19 bar, the newly upgraded ex-LEP refrigerator could handle + 35% more nonisothermal load between 4.5 K and 20 K. We noted as well, that applying similar loads to the refrigerator, the HP of the high compressor stage was 3 bar lower than before the upgrade, demonstrating a clear cold box efficiency gain estimated to + 7% at full capacity. The second step consisted in simulating the heat loads as close as possible as required for the HL-LHC (Table 1). The upgraded ex-LEP refrigerator equivalent capacity at 4.5 K was evaluated to 18.8 kW, with a cold box efficiency of 54.5%. To note that thermal load was measured between 4.5 K and 20 K instead of 17 K. The test showed an upgraded capacity above the HL-LHC requirements of 18.5 kW confirming that the proposed modification fully complies with the HL-LHC requirements and CERN expectations.

5 Project Status and Conclusion The ex-LEP refrigerator upgrade for HL-LHC, completed during the LHC Long Shutdown 2 in 2019 and 2020, is a success. The performance tests achieved during summer 2021 have confirmed the increase of the ex-LEP refrigerator cooling capacity fully compatible with the HL-LHC requirements. Since end 2020, the upgraded ex-LEP refrigerator is in operation recording already 10’000 h of services. It experienced various LHC operating phases, such as the machine cool-down the LHC, magnets commissioning. Thanks to its higher capacity as well as higher efficiency, operation of both refrigerators located at LHC P 4 providing cooling to sector 3–4 and 4–5 have been optimized to minimize the global energy consumption and are ready for the first beams of the new LHC Run 3. To conclude, this upgrade using more efficient turbines is now a reference for possible need of capacity or increased efficiency. Acknowledgments. Part of the HL-LHC project, Authors would like to thank the colleagues from CERN cryogenic group who contribute to the success of this project as well as Linde Kryotechnik® , industrial partner for their contribution to the project.

References 1. Berkowitz Zamora, D., Claudet, S., Perin, A.: Refrigeration assessment of the existing cryogenic plants for the high luminosity upgrade of the Large Hadron Collider (LHC) – CEC-ICMC (2017) 2. Wagner, U., Claudet, S.: Upgrade of four cryogenic helium refrigerators used in the LEP collider for LHC refrigeration. In: Proceedings of the 21st International Cryogenic Engineering Conference (Prague, Czech Republic) (2006)

Analysis of the Optimal Operating Temperature of Dalian Advanced Light Source Zheng Sun1,2(B) , Xu Shi1 , Lei Xu1 , Liangbin Hu3 , and Xilong Wang1 1 Dalian Institute of Chemical Physics (DICP, CAS), 457 Zhongshan Road, Dalian, China

{zhengsun,xushi,xulei2021,xilongwang}@dicp.ac.cn

2 University of Chinese Academy of Sciences, No. 19(A) Yuquan Road, Shijingshan District,

Beijing, China 3 Institute of Advanced Science Facilities (IASF), Guangming Street, Shenzhen, China

[email protected]

Abstract. Dalian advanced light source (DALS) is a new project proposed by Dalian Institute of Chemical Physics (DICP), Chinese Academy of Science aiming to produce high energy electron beam of 1.3 GeV with a repetition rate up to 1 MHz. It is a continuous wave (CW) extreme ultraviolet (EUV) free electron laser (FEL) based on superconducting radio frequency (SRF) technology. The linac of DALS is composed of ten 1.3 GHz cryomodules and two 3.9 GHz cryomodules. Superconducting cavities in these cryomodules require cryogenic temperature to operate in their superconducting state to keep the low RF losses. The RF loss could be minimized with lower temperature according to BCS theory, but the efficiency of cryogenic system can be severely reduced by lower temperature. Therefore, there is an optimal temperature between the efficiency and the cooling power of the cryoplant, which results in the lowest cost and minimize other operating risks. This paper presents the brief introduction of DALS project, and then the operating cost and capital cost of cryogenic system are analyzed when the operating temperature changes. Some other factors such as microphonics in SRF system, helium properties, the impact of cold compressors, etc. are discussed. Considering the above influences, the recommended operating temperature of the DALS linac is proposed. Keywords: Light source · Cryogenic system · Optimal temperature · superconductivity

1 Introduction Dalian Advanced Light Source (DALS) is a new Free Electron Laser (FEL) project proposed by Dalian Institute of Chemical Physics, Chinese Academy of Science (DICP, CAS) which is under conceptual design, and the cryomodule test facility project of DALS has been designed and is currently under construction at Dalian, China [1]. All cavities in DALS linac need to operate in liquid helium to maintain superconducting state for lowest RF losses, the RF losses of cavities decrease strongly with operating temperature as described by the BCS-theory, but the decreasing temperature also affect © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 70–77, 2023. https://doi.org/10.1007/978-981-99-6128-3_8

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the efficiency, the complexity of the cryoplant, the stable operation of cavities and maybe some other issues. It’s necessary to evaluate the choice of operating temperature not only from an economic point of view but in practical point of view.

2 Heat Load Analysis of DALS Linac 2.1 DALS Cryogenic System The block diagram of DALS cryogenic system is shown in Fig. 1. The whole system has mainly three parts, which are the warm compressor system, cold boxes and cryogenic distribution system, and cryomodules in linac. The cryomodule design is based on TESLA technology and very similar with LCLS II cryomodule design [2]. Superconducting cavities, cryogenic process piping and thermal shielding are assembled in strings with the vacuum barriers at each side of the string. The nominal 2 K superfluid helium in the cavities is produced by throttling the JT valve located inside each cryomodule. The cryogenic system provides 2 K liquid helium to the cavities, there are also 5 K thermal intercept circuit and 40–50 K thermal radiation shield circuit need to be cooled.

Fig. 1. DALS cryogenic block diagram

2.2 Heat Load Sources There are two contributions of cryogenic heat loads in superconducting linac: static and dynamic ones. The static heat load is the heat flux from the environment (300 K) to cryogenic system since they are thermally connected. This kind of heat load always exists even the beam is off. The dynamic heat loads are mainly derived from the RF losses of cavities at a given acceleration field, they are caused by the ohmic dissipation in the surface of SC cavities when subject to RF field, accounting for major part of the total heat loads.

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The RF losses can be calculated as: Pdiss =

(Eacc L)2 (R/Q)Q0

=

(Eacc L)2 Rs (R/Q) G

(1)

  Where Eacc is the accelerating gradient, L is the effective length, QR G is the geometry factor, Q0 is the quality factor depends on Rs , Rs is the RF surface resistance and can be divided into RBCS and Rres . The BCS term RBCS is deduced by BCS theory and depends both on temperature T and frequency f , a simplified form and fit is [3]: RBCS (Ohm) = 2 × 10−4 T1



   f 2 exp − 17.67 1.5 T

(2)

The other term is a temperature independent term denoted the residual resistance Rres . Rres is often thought to be caused by impurities, trapped magnetic flux. For DALS, we plan to use the nitrogen doped cavities with high Q0 that has already been used successfully in the LCLS II project [4]. The nitrogen-doped cavities offer an anti-Q slope which increases the Q0 in medium field region. It has been shown that the nitrogendoping lowers the mean free path of the niobium and thus leads to the very low RBCS [5], the Eq. (2) yields a much larger result so the calculation of Rres is obtained with a program SRIMP [6]. The nitrogen-doped cavities also show lager sensitivity of residual resistance to trapped magnetic flux which need fast cool down through Tc for efficient magnetic flux expulsion, the approximate value of Rres is about 4 n with fast cool down [5]. 2.3 Heat Load Analysis of DALS The heat load analysis is the basis of cost estimation, since the heat loads especially the dynamic heat loads vary with the operating temperature significantly, so the cost changes with temperature. For DALS, the static heat load comes from the cryomodule, feed caps and end caps, the distribution box and transfer lines. The static heat loads are mainly derived from the LCLS II cryomodule acceptance results [4] and the actual values will be obtained through the test facilities of DALS in the future [1]. The cavity specification is the same as LCLS II, the cavities need to achieve quality factor of Q0 ≥ 2.7 × 1010 at 16 MV/m at 2.0 K. The 2 K dynamic heat loads are estimated with Q0 of 2 × 1010 which is a conservative value, so no safety factor is used for 2 K dynamic heat load, and for static heat loads a safety factor of 1.5 is taken, dynamic heat loads in other temperature zones a safety factor of 1.3 is taken. The 2 K heat load is about 2 kW, the heat load for 5 K thermal intercept and 40 K thermal shield is about 1 kW and 10 kW respectively. 2.4 Operating Cost and Capital Cost The operating cost is directly related to cooling capacity and the cryoplant efficiency, the cooling capacity at different operating temperatures can be calculated based on the

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above heat load analysis. The cryoplant efficiency η is related to Carnot efficiency ηcarnot and Carnot efficiency fraction ηr : η = ηcarnot × ηr =

Tc Tc −Th

× ηr =

1 COPinv

(3)

Where Tc is operating temperature, Th is the room temperature (300 K is taken in this paper), ηr varies depending on cryoplant design, COPinv is defined as the wall plug power required in order to adsorb 1 W in Tc . An approximate formula proposed by Green [7] can be used to estimate the ηr : ηr (%) = 15.5 × R(kW )0.23

(4)

Where R is equivalent 4.5 K refrigeration power obtained through Carnot efficiency conversion. The COPinv of other real cryoplants are listed in Table 1. The results are also shown in Fig. 2. It can be found that the real values do not deviate much from the results calculated from formula (4), so we decided to use this formula for calculating COP value. Table. 1. Parameters of large cryoplants worldwide Characteristic

XFEL Jlab CHL-1 Jlab CHL-2 ESS ACCP LHC

Operating T (K)

2

2

2

2

1.8

Cooling power @ operating T (kW) >1.9

4.6

4.8

3

2.4

COPinv @ Op T

–

950 [9]

–

594

900

Equivalent 4.5 K Power (kW)

7.7 [8] 12.7

17.8

9

18

COPinv at 4.5 K

–

240 [10]

251 [11]

231 [12]

–

The average electric price in China is about 0.635 ¥/kWh, and assuming 80% of running time, the operating cost is:   (5) Cop = 5562.6 × R(kW )/η $/year Green [7] also proposed a formula to estimate the capital cost of cryoplant, where R is also equivalent 4.5 K refrigeration power:   (6) Ccap = 2.6 × R(kW )0.63 M $ Figure 3 shows the capital cost based on this formula, and a new formula is fitted based on our test facility and S3 FEL project price data to predict the capital cost:   (7) Ccap = 3.54 × R(kW )0.7 M $ As can be seen in Fig. 3, the latest prices are relatively higher than the calculated value, especially for the cryoplant with larger capacity. The reason for this may be due to inflation, and more specific requirements for large system. The real values of course needed to be determined with vendors.

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2.5 Total Cost Based on above analysis, the total cost is estimated with the sum of 20 years of operation cost and capital cost. As shown in Fig. 4, both operation cost and capital cost have a minimum value with temperature variations. This is because the heat load decreases as the operating temperature decreases, the conversion to equivalent cooling power involves the Carnot efficiency which decreases accordingly. The temperature giving the lowest total cost is about 1.7 K, but the minimum value is broad, operating between 1.5 K and 1.9 K only give 5% additional cost compared to the optimum value. The results here only correspond to DALS linac mainly composed of 1.3 GHz cavities.

Fig. 2. COPinv as a function of 4.5 K equivalent power

Fig. 3. Capital cost as a function of 4.5 K equivalent power

Fig. 4. Total cost as function of operating temperature

3 Other Considerations 3.1 Cold Compressors There are mainly three cooling schemes, first scheme is the pure vacuum pump system only uses sub-atmospheric compressors to compress the helium vapor to atmospheric pressure, but the capacity is limited with about 20000 m3 /h due to prohibitive size of vacuum pumps [13]. For 2 K, 20000 m3 /h corresponds to about 460 W assuming 79% JT valve efficiency, so the scheme is not practical for large cooling capacity. The second cooling scheme is pure cold compressors scheme. Cold compressors have a limited compression ratio of 2.5–3.5, for 1.8 K the overall pressure ratio is 64 so at least 4 or 5 stages are needed, but more stages mean more complex control strategy. Third choice is hybrid cooling scheme, the saturated vapor is compressed in cold compressors to interstage pressure and then final compression is done by the warm sub-atmospheric compressors, the pressure head of volumetric compressor varies with mass flow rate, so the pressure ratio of cold compressors change with the mass flow rate which can provide more flexible control range. Large SRF systems often need several working modes which require cryogenic system capable of working over a wide range of heat loads, so hybrid scheme is chosen for most large 2 K systems such as DESY, ESS and LHC. For now, the lowest temperature of current projects is LHC which have cooling power of 2.4 kW at 1.8 K [14]. So at least 1.8 K is technically feasible.

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3.2 Helium Properties and Microphonics Many helium properties vary greatly with temperature which affects the operation of the cavities, the design of cryomodules, etc. Figure 5 shows that both specific heat and density drops with decreasing temperature, low specific heat means less heat could be stored and higher possibility of vaporization and bubble formation which could cause cavity detuning. Low density means higher volume of pipes needed, but this could be compensated by the lower mass flow rate due to lower temperature. The DALS cryomodule design is mainly based on LCLS II cryomodule which have considered the 1.8 K operating conditions [2]. Figure 6 shows the pressure drop decreases with decreasing temperature since lower mass flow needed, so the LP pipe can remain the same size when operating temperature decreases. Microphonics are mechanical vibration of the cavity that induce a shift of its resonant frequency, this could cause significant degradation in cavity performance. In 2 K regime, some literatures show that the vapor generation through throttling device can cause microphonics, lowering the vapor generation can help mitigate microphonic effects [15]. As Fig. 7 shows, assuming the constant temperature before JT valve, the vapor fraction increases from 18.3% to 25% with bath temperature decreasing from 2 K to 1.6 K, so the lower temperature may be more likely to cause the microphonics.

Fig. 5. Density and specific heat as a function Fig. 6. Pressure drop of helium saturated of helium bath vapor as a function of temperature

Another factor that can lead to bubble formation is the heat load exceeding the critical heat flux. In saturated helium, the limit is given by degree of local subcooling, because the cavities are immersed in liquid helium, so the helium on cavity surface is subcooled, exceeding this heat transfer limit can result in conversion to vapor or He I and cause

Fig. 7. Vapor pressure and vapor fraction through JT valve varying with temperature

Fig. 8. Critical heat flux of saturated helium as a function of temperature

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bubbles and instability, as Fig. 8 shows, the highest critical heat flux is at around 1.85 K. Saturated helium also has the highest thermal conductivity is at around 1.9 K [16].

4 Conclusions The optimal operating temperature is around 1.7 K only considering cost estimation. Operating below 1.8 K is difficult because of the need for more cold compressors and complex control strategies. In views of helium properties and helium bath stability, operating at around 1.8 K have better heat conductivity and the highest critical heat flux. But more vapor fraction through JT valve may lead to microphonics. For now, there are no large SRF systems operating in 1.8 K, so 2 K operation is chosen for the current project design, but we also retain the possibility of operating at 1.8 K. Acknowledgments. The project is supported by the Dalian Municipal Government. The authors would like to express our gratitude to Prof. Guy Gistau Baguer and John Weisend II for their constructive comments and advices. We are also grateful that staffs from IASF, IHEP and SHINE for their very valuable advices and help.

References 1. Sun, Z., Huang, L., Shi, X., et al.: Conceptual design of DALS test facility cryogenic system. CEC.ICMC (2021) 2. Peterson, T., et al.: LCLS-II 1.3 GHz cryomodule design-modified tesla-style cryomodule for CW operation. In: Proceedings of SRF2015, Whistler BC, THPB119 (2015) 3. Padamsee, H.: RF Superconductivity: Science, Technology, and Applications. Wiley, Hoboken (2009) 4. Harms, E.R., et al.: Experience with LCLS-II cryomodule testing at Fermilab. In: Paper THP060, These Proceedings (2019) 5. Gonnella, D.A.: The Fundamental Science of Nitrogen-Doping of Niobium Superconducting Cavities. Cornell University (2016) 6. Halbritter, J.: Fortran-Program for the Computation of the Surface Impedance of Superconductors (No. NP-18355). Kernforschungszentrum, Karlsruhe (West Germany). Institut fuer Experimentelle Kernphysik (1970) 7. Green, M.A.: The cost of helium refrigerators and coolers for superconducting devices as a function of cooling at 4 K. In: AIP Conference Proceedings, vol. 985, no. 1, pp. 872–878. American Institute of Physics, March 2008 8. Blum, L., Wilhelm, H., Petersen, B., Schnautz, T.: Status and commissioning results of the helium refrigerator plant for the European XFEL. In: IOP Conference Series: Materials Science and Engineering, vol. 101, no. 1, p. 012138. IOP Publishing, November 2015 9. Petersen, B.: Some aspects of the layout and optimization for the cryogenic supply of superconducting linacs. Nucl. Instrum. Methods Phys. Res. Sect. A 557(1), 280–286 (2006) 10. Knudsen, P., Ganni, V., Hasan, N., Dixon, K., Norton, R., Creel, J.: Modifications to JLab 12 GeV refrigerator and wide range mix mode performance testing results. In: IOP Conference Series: Materials Science and Engineering, vol. 171, no. 1, p. 012015. IOP Publishing, February 2017 11. Zhang, J., Arnold, P., Kolev, N., Rueegge, A.: Commissioning of the large-scale 2K Helium Refrigeration system at ESS. In: ICEC 2022, April 2022

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12. Gruehagen, H., Wagner, U.: Measured performance of four new 18 kW@ 4.5 K helium refrigerators for the LHC cryogenic system. In: Proceedings of the Twentieth International Cryogenic Engineering Conference (ICEC 2020), pp. 991–994. Elsevier Science, January 2005 13. Lebrun, P., Tavian, L., Claudet, G.: Development of large-Capacity Refrigeration at 1.8 K for the Large Hadron Collider at CERN (No. LHC-Project-Report-6) (1996) 14. Lebrun, P.: Large cryogenic helium refrigeration system for the LHC (No. LHC-ProjectReport-629) (2003) 15. Ravikumar, D.K., Than, Y.R., Longtin, J.P.: Eliminating flow-induced microphonics in a superfluid helium cryogenic system. Cryogenics 104, 102984 (2019) 16. Van Sciver, S.W., Timmerhaus, K.D., Clark, A.F.: Helium Cryogenics, vol. 470. Springer, New York (2012)

Key Design Parameters for Improving Oil Removal in Helium Compressors Stations Krzysztof Brodzinski, Vanessa Gahier(B) , and Udo Wagner Technology Department, CERN, 1211 Geneva 23, Switzerland {Krzysztof.brodzinski,Vanessa.gahier,Udo.wagner}@cern.ch

Abstract. Oil is used in screw compressors as a lubricant, a sealant, and a coolant in Helium compressors stations for refrigeration systems: to avoid any propagation of hydrocarbon contamination in the cold boxes’ systems, a high-performance oil removal system shall be installed downstream of the compression station. However, operational experience on the Large Hadron Collider (LHC) and detectors cryogenic units at CERN, recently demonstrated that a part of the existing oil removal system could be significantly improved. Upgraded new equipment was tested and installed during the last years’ operation and long shutdown of the cryogenic infrastructure. The objective of this paper is to emphasize the key design parameters towards improved performance of the existing oil removal systems for helium compressors stations and provide the latest operational feedback from the LHC cryogenics at CERN. Keywords: Helium compression station · oil removal system

1 Background and Introduction Oil-injected screw-compressors are commonly used for helium refrigerators. This choice is mainly a result of the large displacement capacity, reliability, reduced vibration, and ability to handle helium’s heat of compression [1]. The oil acts as lubricant, sealant and coolant and can be found at levels up to 5% vol. at the outlet of the compressor. To avoid hydrocarbon contamination in cold boxes, an Oil Removal System (ORS) is installed downstream of the last compression stage. Operational experience on the Large Hadron Collider (LHC) and detectors cryogenic units at CERN recently demonstrated that a part of the existing oil removal system could be significantly improved. The objective of this paper is to determine the key design parameters to improve the performance of existing ORS in the helium cryogenics system.

2 LHC System Description The LHC machine is cooled by eight refrigeration plants. Four have been upgraded from the Large Electron-Positron Collider (LEP) cryogenic system and four were designed for the LHC. Both types were provided by two different contractors, resulting in different design. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 78–84, 2023. https://doi.org/10.1007/978-981-99-6128-3_9

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In all installations the design of the ORS is similar. The oil separation is performed into two steps, the bulk separation is achieved in a horizontal vessel, the fine separation is achieved in several stages of coalescing cartridges and oil traces are removed in a charcoal bed (see Fig. 1).

Fig. 1. LHC Oil Removal System illustration.

The different designs vary by dimensions and features of the separator as well as dimensions and type of coalescing cartridges, leading to different performance as explained in this paper.

3 Bulk Separation Quantification - Methodology Two main approaches to design a separation vessel exist: the Souders-Brown equation, or the droplet theory method. The Souders-Brown equation is an empirical approach commonly used in the oil and gas industry. The droplet theory quantification method has been used to assess the steady-state separation performance of LHC’s cryogenic installation separators. The bulk separation can be decomposed in three sections: – The feed pipe and the inlet device where large droplets are separated. – The gravity section where medium size droplets are removed. – The mist removal section where small size droplets are removed. The feed pipe and the inlet device, if present, achieve the first liquid gas separation. The flow pattern in the feed pipe defines the entrainment of liquid in the gas phase and depends on the liquid and the gas superficial velocity. The feed pipe design is crucial: in case of low points, slug flow can occur, and massive re-entrainment can happen in the phase separator leading to lower separation performance. To obtain a stabilized flow pattern, the installation of a straight horizontal pipe of 10 times the pipe diameter is recommended. The liquid entrainment in a horizontal pipe for an annular flow pattern defined by Pan and Hanratty [2] is used for quantification given in Eqs. (1) and (2).    0.5 m 1/(2−m) DU3g ρ0.5 ρ1−m E/EM L ρG G μG = A2 1+m 1 − E/EM σ d32 gρL

(1)

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ρL U2g d32



σ

 0.5 d32 ρ0.5 L g = 0.14 σ0.5

(2)

With increasing velocity, the liquid entrainment increases and the droplet size decreases. A possible operational measure in case of low ORS system performance is a decrease of the volumetric flow by maintaining sufficiently high pressure in the ORS. An inlet device located in the separation vessel allows the separation of gas and liquid and distributes the gas evenly over the separator cross section. When present, the momentum in the inlet device should be limited depending on its complexity to avoid mal-distribution across the separator cross section. In the gas gravity separation section of the vessel, the droplets are eliminated by gravitational force higher than the drag force. In most designs, droplets with a diameter higher than 100 μm are separated in this section. To quantify the gravity separation, the droplet distribution shall be known. In our approach, the upper-limit, log-normal distribution law, given in Eq. (3), was used.   dvμ δd d 2 − ln (3) exp −δ ln fz = √ dm − d dm − dvμ πd(dm − d)

Table 1. Variables used in Eqs. (1) to (3) Unit/Value

Unit/ Value

E/EM

Liquid entrainment ratio –

A2

Empirical constant

9 × 10–8

D

Pipe diameter

m

Ug

Gas superficial velocity

m/s

ρL

Liquid density

kg/m3

ρG

Gas density

kg/m3

σ

Liquid surface tension

N/m

μG

Gas viscosity

Pa s

g

Gravitational constant

m/s2

d32

Sauter diameter

m

m

Settling law exponent

–

d

Droplet diameter

m

δ

Constant

0.72

dvμ

Volume median drop size m

dm

Max. Droplet diameter

m

Mist removal can be performed using and combining different technologies such as vane pack to remove droplets down to 20 μm, knitted mesh for droplets down to 3 μm or fiber bed mesh for droplets down to 0.1 μm. The velocity across the mist elimination device should be sufficiently low to avoid liquid flooding and re-entrainment. Most mist removal technologies use the inertial impaction mechanism for liquid separation by which momentum makes the droplet deviate from the flow streamline to impact the medium, where the droplets coalesce to large droplets that settle by gravity [3]. The inertial capture efficiency is a function of the Stokes number [4] and has been published by Langmuir and Blodgett. Simple overall removal efficiency of knitted mesh removal bed correlation studied by Bradie and Dickinson [5] is used here.

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4 Fine Separation Quantification - Methodology The coalescing cartridge filter uses three mechanisms for liquid separation according to the droplet diameter: inertial impaction (described in the Sect. 3); direct interception: droplet follow the flow streamlines and is separated when it reaches the media and diffusional interception: the very small droplet (28 g/s

(a)

(b)

Fig. 5. Helium mass flow rate versus (a) P1 suction and (b) motor speed with multi-vacuum pump

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4 Commissioning of Circulating Dry Vacuum Pumps with CC in the Superfluid Helium Cryogenic Refrigerator Figure 6 shows the picture of the superfluid helium cryogenic refrigerator including six dry vacuum pumps. The commissioning of circulating dry vacuum pumps with CC has been carried out in 2020. The suction and discharge pressure of pumps from start to stop process is given in Fig. 7. Under condition of 60 Hz frequency, the suction and discharge pressures are 345 and 1150 mbar, respectively. During the operation, the pump group works stably. The maximal cooling capacity of this helium cryogenic refrigerator is up to 510.6 W and maintain stable at 1.967 K temperature for hours, which is measured by thermometer TI4200x and heat load EI4200.

Fig. 6. Picture of superfluid helium cryogenic refrigerator (including the dry vacuum pumps)

Discharge pressure 1150 mbar

suction pressure 345 mbar

mass flow rate 26.8 g/s

(a)

(b)

Fig. 7. (a) The suction and discharge pressure of pumps and (b) cooling capacity

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5 Conclusions and Outlook A set of dry vacuum pumps suitable for superfluid helium cryogenic system are studied and some conclusions can be acquired. 1) The control strategies including one Set-point and Upper & Lower limit, variable frequency regulation and failure protection can effectively adjust for the operation of the circulating dry vacuum pumps. When the suction pressure is 400 mbar the discharge pressure is 1100 mbar, the maximal mass flow rate is up to 39 g/s by multi-vacuum pump (MVP) and variable frequency drive (VFD) of motor. 2) This set of dry vacuum pumps with cold compressor achieves the cooling capacity of 510.6 W @ 1.97 K, which verifies the feasibility of dry vacuum pumps used in the superfluid helium cryogenic refrigerator. 3) This superfluid helium cryogenic refrigerator with dry vacuum pumps is originally self-developed in China. It will be applied in the superconducting cavity vertical test and cryostat test in CiADS of China in the near future. Acknowledgments. The project is supported by the fund of the National Key Research and Development Program (Grant No. 2020YFB1506201) and the Strategic Priority Research Program of Chinese Academy of Sciences (Grant No. XDC10010300) of China. This work is partially supported by Busch AG in Shanghai.

References 1. Van Sciver, S.W.: Helium Cryogenics. Plenum Press, New York (1986) 2. Claudet, G., Aymar, R.: Tore Supra and helium II cooling of large high-field magnets. Adv. Cryo. Eng. 35A, 55–67 (1990) 3. Philippe, L., Laurent, T.: CERN yellow report cooling with superfluid helium, pp. 453–476 (2014). https://doi.org/10.5170/CERN-2014-005.453 4. Lebrun, P., Tavian, L.: Beyond the Large Hadron Collider: a first look at cryogenics for CERN future circular colliders. Phys. Procedia 67, 768–775 (2015) 5. Weise, H.: The TTF/VUV-FEL (FLASH) as the prototype for the European XFEL project. In: Proceedings LINAC, JACOW, pp. 486–490 (2006) 6. Bozhko, Y., Lierl, H., Petersen, B., et al.: Requirements for the cryogenic supply of the European XFEL project at DESY. AIP Conf. Proc. (2006) 7. Arnold, P., et al.: The ESS Test and Instruments Cryoplant – first test results and operation experiences. IOP Conf. Ser. Mater. Sci. Eng. 755, 012095 (2020) 8. Ge, R., Li, S., Han, R., et al.: ADS Injector-I 2 K superfluid helium cryogenic system. Nucl. Sci. Tech. 31, 39 (2020) 9. Xie, X.J., et al.: Integration and commissioning of a full localization [email protected] helium refrigerator at TIPC. IOP Conf. Ser. Mater. Sci. Eng. 502, 012013 (2019) 10. Xue, R., Yang, S., Xie, X., et al.: Influence of key parameters on the performance of a helium cryogenic system in refrigeration and liquefaction modes. Cryogenics 121, 103386 (2022)

Overview of the Maintenance and Consolidation Activities for CERN Cryogenics During LHC Long Shutdown 2 V. Iuga(B) , N. Bonetti, K. Brodzinski, F. Ferrand, C. Fluder, L. Herblin, B. Ivens, S. Junker, A. Perin, O. Pirotte, and M. Pezzetti Technology Department, European Organization for Nuclear Research CERN, Esplanade Des Particules 1 - CH-1211 Geneva 23, Meyrin, Switzerland [email protected]

Abstract. CERN operates and maintains several large cryogenic systems including those serving the Large Hadron Collider (LHC) complex and its associated detectors. After a period of continuous operation of nearly 40’000 h (RUN 2), CERN started in January 2019 the 24 months Long Shutdown 2 (LS2) of the accelerator complex and related experiments. More than 3’000 interventions were performed during the LS2 period on the LHC cryogenics. The main challenges were to apply a relevant maintenance policy and identify the necessary consolidations to ensure for the LHC cryogenic system the requested operational availability of at least 98%. This paper reports on the key activities performed during this major maintenance period, the organisation and quality control applied, difficulties encountered, and the lessons learned from the maintenance & consolidations teams. Additionally, the report on the first operational results obtained after the restart of all LHC cryogenic plants & associated infrastructure for the beginning of RUN 3 will be presented. Keywords: CERN · LHC · Cryogenics · Long Shutdown 2 · Maintenance · Consolidations · Lesson learned

1 Introduction The cryogenic infrastructure of the Large Hadron Collider (LHC) machine is composed of eight cryogenic plants of 18 kW at 4.5 K and eight cryogenic plants of 2.4 kW at 1.8 K units and distribution valve boxes (Fig. 1). Each plant comprises a compressor station and a refrigerator which is providing cooling power for 3.3 km long sectors through cryogenic distribution lines as well as the associated infrastructure to manage an inventory of more than 130 t of helium [1]. In addition, two cryogenic plants are operating at CMS (P5) and ATLAS (P1) detectors facilities. The entire infrastructure has demonstrated a peak availability of more than 98% over the RUN 2 [2]. To perform the necessary technical actions to maintain and improve this high level of performance for the next operation period RUN 3, CERN initiated in early 2019 the LS2 planned for a duration of two years. This was the unique opportunity to © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 92–100, 2023. https://doi.org/10.1007/978-981-99-6128-3_11

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Fig. 1. Cryogenics infrastructure of the LHC accelerator During physics RUN 2, ending in December 2018, most cryogenics equipment had been in continuous operation during 4.5 years with limited maintenance possibilities and had cumulated more than 40’000 running hours.

have the entire LHC machine and cryogenics equipment warmed up for a period of about a year. The objective for cryogenics team was to perform required maintenance program and consolidations of the LHC machine, ATLAS and CMS experiments.

2 Long Shutdown 2 Preparation Period 2.1 LS2 Preparation Steps The preparation phase started one year before the start of LS2, in March 2018. It was defined based on the lessons learned during the previous Long Shutdown 1 (2013–2014), and the yearly technical shutdowns [3]. The activities were performed by several work groups held during operation period including safety related documentation, instrumentation calibration and planning tools. All sections of the cryogenics group contributed to this phase including operation, mechanical, electrical and instrumentation, methodology and logistics, as well as external contractors and suppliers. CERN has continuously improved its maintenance policy and plan since the start of the LHC cryogenics infrastructure operation [4]. All the experience gathered along the years is documented and managed within a Computerized Maintenance Management System (CMMS). As shown in Table 1, an extensive review of most reference documents was carried out during this preparation phase. A significant effort was also put to fully integrate maintenance work and parts tracking within the CMMS to enable detailed quality control during the execution phase. The planning structure and coordination tools were defined to coordinate activities within the LS2 centralized coordination framework for LHC machine, also to allow quick identification of co-activities, critical paths and to precisely plan a large number of interventions on multiples sites in parallel with internal or external resource. Contract technical documentation was also prepared during this phase to specify major overhauling requirements in European industry for rotating machinery including

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Documents

Action

Key figures

Maintenance Plans

Review of content, adequacy, and periodicity

179 plans

Procedures

Technical review and methodology

116 procedures

Checklists

Creation of electronic documents

239 checklists

Material Lists

Review of asset and parts mapping

506 lists

Spare Parts

Implementation of pre-reservation procedure

177 lists/7’263 parts

Plannings

Review of the LS2 maintenance plan

3’000 Work Orders

Contracts

Specification of requirements & work packages

6 major contracts

screw compressors and high voltage motors at contractor premises complying with CERN standards. Regarding on-site maintenance activities, work packages were detailed in the CMMS in coordination with the contractor in charge of execution of the tasks [5]. 2.2 Technical Review Predictive maintenance program is a major part of CERN cryogenics maintenance policy including vibration analysis [6], oil analyses and infra-red camera check of electrical equipment. Measurement results from these campaigns completed during RUN 2 were reviewed and used to plan targeted interventions. The maintenance policy is also based on preventive plan and instrumentation verification campaigns. Technical reviews meetings took place for each type of activity considering analysis of the corrective maintenance performed during RUN2. The main goal was also to define the criticality and to validate activities to be performed during LS2. All technical topics related to maintenance, operation or consolidation projects were then submitted for final approval to the operation and maintenance panel of the cryogenics group. As illustrated by the examples of Table 2, the LS2 maintenance actions were focused on critical assets for safety, process, or reliability. Accessibility and reaction time were also considered especially for assets located underground or in controlled areas. Another important decision was the deployment of an instrumentation calibrator interfaced with CMMS to improve measurement accuracy, efficiency on the field and reporting. 2.3 Consolidation Activities A set of consolidations were implemented during the LS2 on the cryogenic infrastructure to improve the reliability of existing equipment, replace or refurbish obsolete part of the infrastructure or implement new functionalities [7]. Prioritization of these projects was based on incidents that occurred during RUN 2, and cumulated experience on expected device lifetime. As shown in Table 3, several important projects have been addressed

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Table 2. Examples of technical decisions for maintenance during LS2. Domain

Decision

Result

Instrumentation

Sensor’s calibrations reviewed by criticality

20% of 8’669 sensors

Deployment of instrumentation calibrator

2’000 sensors

Vacuum

Review of vacuum criticality & accessibility

28% of 566 devices

Vacuum gauges check and calibration

45% of 411 devices

Mechanics

Replacement of coalescing filters reference

32 coal stages

Conditional maintenance for instrumented filters

47% of 262 assets

Logistics

Water cooling exchangers analysis and cleaning

40% of assets

Anticipated shipping of rotary machines

126 assets

Table 3. Main cryogenics consolidation activities of the LS2. Objective

Project

Volume

Reliability improvement

High stage compressors bearing size increase

11 compressors

Oil pumps with magnetic couplings

12 pumps

Refurbishment of Cold Compressor electronic

4 installations

Switchable water-cooling filters

5 filters

Cryogenic temperature sensors replacement

320 sensors

End of life management

New functionalities

Replacement of ATLAS MR compressor

1 compressor

LHC old generation PLC upgrade

71 PLC

Experiment cryogenics control electric upgrade

10 PLC

Refurbishment of electrical cabinets

24 units

Compressor set cable replacement

3’000 m of cable

Instrumentation cable replacement

800 m of cable

Oil pumps with magnetic couplings

16 pumps

LHC global mass flow measurement at cold

7 flowmeters

during LS2, mainly mechanical or instrumentation changes to leverage reliability, and controls or electricity related projects planned in the obsolescence management program. Consolidation projects require more preparation time compared to standard maintenance preparation as it usually requires engineering studies, on-site pre-qualification tests, external financing, as well as long procurement delivery time.

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3 LS2 Overview and Achievements 3.1 Safety During peak activities of the LS2, more than 100 people were working on-site for cryogenics LS2 activities. We faced minor incidents but no major safety accident, and incidence and frequency rate of our maintenance contractor show better results compared to industrial standards. This was achieved thanks to a careful organization and delegation of work permit validation to local cryogenic control rooms by senior operators, systematic safety visit before start of a new activity with all stakeholders and limitation of the number of activities in parallel. Strong involvement and continuous on-site presence of safety officers to address rapidly safety related issues on the field was also important. 3.2 Maintenance Results During LS2, almost 3’000 preventive and corrective work orders were completed, requiring almost 8’700 activities. The task execution and coordination were reviewed on a weekly basis. The mechanical activities representing almost two thirds of the working hours were identified as the main tasks on the critical path. Electrical/Instrumentation and vacuum activities representing respectively 26% and 9% of the working hours were closely monitored as they were directly related to highly critical device. Tracking tools for actual execution of the work were defined at the very beginning of the shutdown to rapidly identify any discrepancy between the forecast plan and the actual execution. The detailed scheduling of activities, number of work sites in parallel as well as mechanical resource were adapted continuously to comply with the initial target dates. Schedule adherence was better than four weeks until the end of Q1–2020, especially the supply chain management of spare parts and asset logistics never impacted field activities as it had been the case during LS1. Nevertheless, COVID-19 lockdown strongly impacted the initial planning, and afterwards we had to resume activities with new restrictions and limited resource. The progress curve versus initial baseline is shown in Fig. 2.

Fig. 2. Maintenance activities follow-up versus baseline.

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During LS2, quality control was reinforced with respect to LS1. This included the implementation of a dedicated team to perform systematic on-site inspections, especially for critical interventions with predefined control points. This led to significant reduction of rework after interventions compared to LS1. Additionally, LS2 lesson learned showed that the enhanced methodology using CMMS integrated tools such as on-site reporting of Work Orders with electronic tablets in real time, also enabling direct access to work instructions and checklists had an added value to the overall quality of the work. To avoid failures that could have an impact on LHC machine availability, CERN put emphasis on implementation of preventive plans. Figure 3 highlights the ratio between corrective and preventive interventions for the main maintenance tasks. Thus, we anticipate the replacement of the equipment from the early stage of possible failure by using a combination of diagnostic tools and systematic interventions, that may occur. Moreover, inspections and adjustments are preferred to corrective actions.

Fig. 3. Distribution of preventive and corrective interventions by CMMS sub-categories.

One of the maintenance indicators followed over the years for LHC cryogenics is the Corrective over Preventive ratio (C/P). This ratio is based on cost and includes on-site work as well as major overhaul of assets at contractor premises. The C/P show a value of 17.6% over the RUN 2 and LS2 period about 2% above the forecasted ratio calculated in 2017 [8]. 3.3 Consolidations Results All consolidations planned during LS2 have been delivered representing more than 10 critical projects. Even if most of them will show effective results in the coming years compared to the set objectives to maintain or improve cryogenics availability, we can already share some feedbacks. The PLC unit dedicated to the cryoplant control system have reached their programmed end of life. The upgrade campaign was identified critical to maintain the cryogenics availability as any failure would cause important downtime of the cryogenic operation. Their integration was also reviewed to improve electromagnetic compatibility.

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Moreover, every process control system faults identified after restart are systematically investigated to perform the root cause analysis and to identify the optimal consolidation to be applied. Cold Compressors electrical cabinets and its control of the have been consolidated including the process control logic to minimize the recovering time after a nonprogrammed stop of the cold compressor system [9]. Commissioning and operation restart of Cold Compressor systems has been performed with no major issues. Cross impact between Cold Compressor stability and external electrical perturbation remains under close follow-up. For helium compressor station, the recommissioning of older compressor stations after the complete upgrade of the electrical architecture [10], despite the significant technical changes has been performed without major issues. This consolidation is considered as reference for next consolidations to be performed during LS3, since the electrical system will reach their programmed lifetime. High stage compressor models which had the highest failure rate during RUN 2 were consolidated with improved ball bearings type. Complementary to that modification, the installed oil filtration was improved to reduce particles pollution that can damage the bearings. In addition, oil is also filtered using a mobile filtration system to reduce the particle pollution. Finally, oil pumps replaced with magnetic couplings to avoid multiple shaft seal failures encountered in the past are now in nominal operation.

4 Restart of the Cryogenics for LHC Machine After a period of almost two years and a half of stop for maintenance and consolidation activities, the handover to the operation team was completed in October 2020. The cryogenic LS2 maintenance schedule of each LHC sector was performed in close correlation with operational and other departments maintenance constraints. The cumulated number of running hours after LS2 until end of February 2022 is shown in Table 4. Most cryoplants have been in operation for more than 10’000 running hours to provide the cryogenic conditions needed for the prequalification tests of each sector. After the restart, the number of corrective activities was very limited, and it never impacted the LHC accelerator recommissioning plan. Currently, the training phase of the LHC is completed, the cryogenic system operates at nominal power and shall provide reliable operation for the RUN 3.

5 Conclusion and Outlook LS2 represented a challenge that CERN cryogenics achieved successfully based on a detailed preparation and partnership across the cryogenic team and external contractors. The LHC machine is now operating at its nominal capacity and the feedback after the restart is positive with no major issue or concern, thus validating the adopted maintenance strategy. All measures have been taken to ensure the reliable operation for the scheduled running period of 40’000 h that could potentially be extended to consider the evolution of the planned RUN 3.

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Table 4. LHC cryogenics running hours after LS2 until end of February 2022. Restart order

LHC Point

(Cryoplant) Main feed

Avg running hours

2nd

Point 18

(B) Sector 12

13’200 h

5th

Point 2

(A) Sector 23

8’290 h

3rd

Point 4

(A) Sector 34

10’550 h

(B) Sector 45

10’950 h

4th

Point 6

(B) Sector 56

12’390 h

(A) Sector 67

6’840 h

1st

Point 8

(A) Sector 78

9’110 h

(B) Sector 81

10’590 h

–

Point 1

(MR) ATLAS

2’480 h

(SR) ATLAS

13’670 h

–

Point 5

CMS

9’020 h

After the successfully completion of the LS2, the level of expertise increased in terms of maintenance and operation of the LHC machine. Over the years, CERN has developed a series of tools to enhance the availability and maintainability of the cryogenic installations. In addition, CERN is currently working on developing new tools based on artificial intelligence to increase the level of predictability and to identify from the earliest stages a potential threat to the availability of the LHC accelerator.

References 1. Lebrun, Ph.: Cryogenic Systems for Accelerators, in Frontiers of Accelerator Technology, pp. 681–700, Hawaii (1995) 2. Ferlin, G., Bradu, B., Brodzinski, K., Delprat, L., Pezzetti, M.: Cryogenics Experience During Run 2 and impact of LS2 on next run, In: 9th LHC Operations Evian Workshop, pp. 85–90, Evian Les Bains (2019) 3. Serio, L., et al.: CERN experience and strategy for the maintenance of cryogenic plants and distribution systems. In: Advances in Cryogenic Engineering: Cryogenic Engineering Conference, p. 012140, Tucson (2015) 4. Bonetti, N., Ferrand, F., Gayet, P., Knoops, S.: General cryogenic maintenance policy and recent update for CERN assets, In: Cryogenic Operation Workshop, Barcelona (2018) 5. Ferrand, F., Gayet P., Guillotin, N.: Experience from the outsourcing of the cryogenic operation & maintenance at CERN, In: Cryogenic Engineering Conference and International Cryogenic Materials Conference, p. 012084, Connecticut (2019) 6. Pirotte, O.: Vibration monitoring of the rotating machines of the CERN cryogenic systems. In: 7th International Workshop on Cryogenics Operations, Illinois (2016) 7. Delprat, L., Ferrand, F.: LHC cryogenic infrastructure reliability, Towards high availability. In: 6th Accelerator Reliability Workshop, Versailles (2017)

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8. Serio, L., et al.: Outsourcing strategy and tendering methodology for the operation and maintenance of CERN’s cryogenic facilities, In: International Cryogenic Engineering Conference, p. 012115, Madison (2017) 9. Bradu, B., et al.: Process control system evolution for the LHC cold compressors at CERN. In: Cryogenic Engineering Conference and International Cryogenic Materials Conference, p. 012086, Connecticut (2019) 10. Pezzetti, M.: Control of large helium cryogenic systems a case study on CERN LHC. EPJ Tech. Instrum. 8(1), 1–20 (2021)

Commissioning of the Large-Scale 2 K Helium Refrigeration System at ESS J. Zhang1(B) , P. Arnold1 , N. Kolev2 , and A. Rueegge2 1 European Spallation Source (ESS) ERIC, Partikelgatan 2, Lund, Sweden

[email protected] 2 Linde Kryotechnik AG, Daettlikonerstrasse 5, 8422 Pfungen, Switzerland

Abstract. The European Spallation Source (ESS), built in Lund, Sweden, will provide long-pulsed neutron fluxes at very high brightness to the neutron research community. The ESS is driven by a 2.0 GeV proton linear accelerator (LINAC) comprising 43 cryomodules (CMs). The accelerator cryoplant (ACCP) provides the cooling for the CMs and the cryogenic distribution system (CDS) of the LINAC. To maintain the superconducting cavities operating at 2 K, the ACCP can provide a maximum refrigeration capacity of 3.0 kW at 2 K. The ACCP adopts a 3 stages cold compressors string combined with one warm compressor to achieve 31 mbar for the CMs cavity circuit. In a dedicated collaboration between the supplier of the refrigeration plant, Linde Kryotechnik, and ESS, the ACCP has been successfully commissioned and tested in different operating modes. This paper will cover the system design, commissioning experience, performance results and lessons learned. The new developed control logic will be discussed. Keywords: Accelerator Cryoplant (ACCP) · 2 K Helium Refrigeration system · Cold compressors (CCs)

1 Introduction As one of the world-class neutron science facilities, the European Spallation Source (ESS) will provide a 2.0 GeV proton Linac using superconducting RF cavities operating at 2 K, supporting the research range from material science, condensed matter and biomedical studies [1]. The Accelerator Cryoplant (ACCP) will supply super-critical helium to the cryomodules (CMs) in the Linear Accelerator (LINAC) through the Cryogenic Distribution System (CDS). According to the design specification [2], the ACCP has a cooling capacity of ~ 2.5 kW at 2 K, 4.5 K liquefaction rate of 6.8 g/s and 8.5 kW at 40~50 K in stage 1 and ~ 3.0 kW at 2 K, 4.5 K liquefaction rate of 9 g/s and 11.4 kW at 40~50 K in stage 2. In both stages, there are five operation modes, in which the stage 1 nominal design mode (S1ND) will be the most important operation mode in the coming years. In order to maintain the 2 K operation for CMs, three stages of cold compressors (CCs) and one sub-atmospheric warm compressor are adopted. In a collaboration between the supplier, Linde Kryotechnik (LKT), and ESS, the ACCP has been successfully commissioned and tested in different operating modes. The simplified ACCP process flow diagram with a test vessel and heaters are shown in Fig. 1. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 101–108, 2023. https://doi.org/10.1007/978-981-99-6128-3_12

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Fig. 1. The simplified ACCP process flow diagram with a test vessel ((heat exchangers: HX1 ~ HX7; Turbines: T1 ~ T6; different pressure stages: HP, MP, LP and SP; Heaters: Q1, Q2; Thermal shield (TS) [3])

After successful completion of the acceptance test and agreeing on the residual punch points, ACCP has been handed over to ESS in October, 2020. Subsequently, a variety of tests, including CCs off-design tests, ACCP-CDS new modes tests and ACCP envelope capacity tests have been performed, as well as trouble shooting, such as turbines damage and cold compressor speed sensor failure. In the following sections, the commissioning and performance of ACCP and some issues will be discussed.

2 ACCP Commissioning and Performance 2.1 The Commissioning of ACCP After one and a half years’ commissioning and operation, ACCP runs in a very efficient and stable condition. The commissioning experience includes: • Before the cooldown, the impurities of the system are removed by a small flow circulation among ACCP, pure helium tanks and Helium recovery system until the impurity condensation is below 10 ppm. • During the cold box cooldown, the opening of the third turbine string inlet valve is controlled to make sure that the Mach number of turbines T4 and T5 is less than 1 in order to avoid the turbine damage. The adsorbers should start earlier to make sure the impurities adsorbed. • The capacity of the HP compressor is regulated by the slide valve, which will affect the efficiency of HP compressor. Normally, the opening of the HP slide valve should make sure the opening around 20% ~ 30% of the two HP-MP bypass valves in order to keep the MP pressure stable.

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2.2 The Performance of CCs In ACCP, the 3 CCs string combined with a sub-atmospheric pressure (SP) compressor are used to keep all CMs in the LINAC at 2 K. The suction pressure of SP compressor could range from 0.3 bara to 1.05 bara, which will compensate the flow variation for CCs during the pump down or in off-design conditions. The flow scheme for the 2 K part is presented in Fig. 2. The bypass valves CV34998 and CV33801 are used to regulate the CC mass flow (FC34905) and CC1 suction temperature (TC34650). The set point of FC34905 is based upon the SP suction pressure, which means the mass flow set point is regulated by setting the SP suction pressure. The CC1 suction pressure can be set in auto mode. If the CCs haven’t reached the maximum working capacity, the CC1 suction pressure will reach the setpoint.

Fig. 2. The flow scheme of CCs and SP compressor

In S1ND mode, the measured parameters of CCs are shown in table 1. With the CC1 suction pressure of 25.93 mbar and the suction temperature of 4.5 K, the mass flow of CCs is stabilized at 95.3 g/s and the isentropic efficiencies of all the CCs are > 72%. The CCs map is presented in Fig. 3, which shows the pressure ratios of CC2 and CC3 are better than the design (S1D). The performance of CCs met the expectation in S1ND. Table 1. The measured parameters of CCs in S1ND mode

Tinlet , K Pinlet , mbar/bara

CC1

CC2

CC3

4.5

8.9

15.7

25.93 mbar

Toutlet , K

8.9

Poutlet , bara

0.10

0.10 bar 15.7 0.30

0.30 bar 22.7 0.62

mass flow, g/s

95.3

95.3

95.3

Polytropic efficiency, %

79.1%

75.7%

81.5%

Isentropic efficiency, %

73.1%

72.1%

75.5%

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CC1 test

CC2 test

CC3 test

Fig. 3. The CCs map in S1ND

2.3 The Performance of ACCP The parameters comparison among the measurements, ESS requirements and the Design of ACCP in S1ND is listed in Table 2. All the measurements exceed the ESS requirements. The design parameters have a 5% margin than the requirements. Using the exergy efficiency with the definition as same as in [4], the measured WCS and the Cold Box (CBx) exergy efficiencies are 48.5% and 54.0%, respectively. The measured system exergy efficiency is 26.2%, which is 1.4% higher than the Design. The performance of ACCP is better than the expectations. Table 2. The parameters comparison between measurement and design in S1ND Measurements

Requirements

Design*Note

2 K heat load, W

2519

2478

2603

4.5 K coupler cooling, g/s

6.9

6.8

7.14

TS heat load, W

8962

8551

8979

2 K mass flow, g/s

95.3

92.5

97.0

CC1 suction pressure, mbar

25.93

≤ 27.00

26.00

4.5 K equivalent heat load, W

7585

7292

7670

WCS input Power, kW

1900

/

2034

WCS exergy efficiency, %

48.5

/

45.1

CBx exergy efficiency, %

54.0

/

55.0

System exergy efficiency, %

26.2

/

24.8

* Note: The ACCP design parameters have a 5% margin compared with the requirements of ESS

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Based on the test results in different operating conditions, the HP pressure in Gas Management Panel (GMP) and the system efficiency are shown in Fig. 4, in which the HP pressure has a linear correlation with the 4.5 K equivalent heat load. Generally, the system efficiency increases with the 4.5 K equivalent heat load increase. When ACCP operates in the floating mode, i.e. the GMP load/unload valves are closed, LHe Dewar is used to balance the system capacity. The slide valve in HP compressor could control the bypass flow from HP to MP, which affects the WCS efficiency. The HP pressure could be set according to the 4.5 K heat load to fine tune the system stability before operating in the floating mode.

Fig. 4. HP pressure in GMP and system efficiency versus 4.5 K equiv. Heat load for ACCP

3 Issues and Improvements 3.1 Turbines Damage In the third turbine string, there are two serial turbines, Turbine 4 (T4) and Turbine (T5). Both T4 and T5 were damaged during the commissioning. In June 14th , 2021, the T5 expansion wheel and the nozzle were found with damage after the ACCP had a performance degradation for several months, shown in Fig. 5. After LKT’s investigation, the reason for T5 damage is the Mach shock during the cooldown process. To make sure T4/T5 Mach number < 1 during cooldown, a new control logic was implemented in the PLC which controls the opening of T4 inlet valve according to the T4 inlet temperature and T4/T5 pressure ratio.

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T4 occurred a bearing damage and inlet filter deformation after a power outage in Sepember 30th , 2021. With LKT’s inspection, the reason for T4 bearing damage might be T4 incorrectly assembled in LKT workshop and the deformed filter might be caused by very high debris load. Then the spare T4 was installed. The repair was still under warranty and the damaged turbines were swiftly carried out by LKT. Now both T4 and T5 have nominal performance again.

(1) Expansion wheel

(2) Nozzle

Fig. 5. The damaged expansion wheel and nozzle in T5

3.2 CC2 Speed Sensor Failure In October 22nd, 2021, the CC2 speed sensor was broken and sent back to LKT for replacement. During the future operation of LINAC it might happen that we need to replace one CC and keep the rest of the cryogenic system and CMs working at 4.5 K. Therefore, we arranged the test to warm-up the CCs string under the condition of the the CBx in cold. The process includes CCs string warm-up, CC spare replacement, CCs string cool-down and CCs string pump-down, which presented in Fig. 6. It took 24 h to warm up the CCs, 1.2 h to cool down the CCs string and 0.6 h to pump down the CCs string. According to our experience, around 6 h are required to replace the CC spare. Therefore, 32 h in total are needed to replace one CC with CBx cold. In the future, we will further investigate and work on to minimize the downtime caused by one CC replacement after the CDS and CMs are connected with the ACCP.

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Fig. 6. (1) The trends for CCs string warm-up, (2) The trends for CCs cooldown, (3) The trends for CCs pump-down

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4 Conclusion ACCP had the site acceptance test successfully and fulfilled the expectations in July, 2020. Then we had ~ 1.5 years ACCP fine tuning and operation. The ACCP performance are studied, such as CCs off-design and ACCP different heat loads tests. With the issues occurring during the commissioning, the operation process is optimized and more experience has been achieved. After ACCP and CDS connected, ACCP-CDS integrated test will be carried out in Q4, 2022. Acknowledgments. Special thanks to ESS Cryogenics Group and Linde Kryotechnik performing a lot of commissioning and operating work.

References 1. Weisend II, J.G., et al.: Status of the ESS cryogenic system. In: AIP Conference Proceedings, vol. 1573, no. 1, pp. 633–638. American Institute of Physics, January 2014 2. Wang, X., et al.: Specification of the ESS accelerator cryoplant. Phys. Procedia 67, 89–94 (2015) 3. Wang, X.L., Arnold, P., Hees, W., Hildenbeutel, J., Weisend, J.G.: ESS accelerator cryoplant process design. In: IOP Conference Series: Materials Science and Engineering, vol. 101, no. 1, p. 012012. IOP Publishing, November 2015 4. Arnold, P., et al.: Large scale refrigeration plant for ground testing the james webb telescope at NASA johnson space center. In: AIP Conference Proceedings, vol. 1218, no. 1, pp. 1080–1086. American Institute of Physics, April 2010

Numerical Simulation on Flow Resistance of Helium Cryogenic Transfer Lines Sheng Xu1 , Mengjia Chen1 , Lei Zhang2 , Jiuce Sun2 , Yuzhe Lin1 , Zhengrong Ouyang3 , and Shaowei Zhu1(B) 1 School of Mechanical Engineering, Tongji University, 4800 Cao’an Road, Shanghai 201804,

China [email protected] 2 ShanghaiTech University, 393 Middle Huaxia Road, Shanghai 201804, China 3 Shanghai Advanced Research Institute, 99 Haike Road, Shanghai 201804, China

Abstract. The flow characteristics of helium at low temperature are vital for designing a large-scale helium cryogenic system, which is a key component of the scientific apparatus. However, it is difficult to measure the flow resistance characteristics of helium at low temperature experimentally, so the empirical formulas under room temperature are often used. In order to study the applicability of empirical formulas under low temperature conditions, several models of horizontal straight tubes, vertical straight tubes, 90° circular elbows, diffusing tubes and reducing tubes are established by using commercial software FLUENT. Compared with the empirical formula, the flow resistance calculated by the straight-line simulation has a little difference which is less than 10%. The difference of that between the horizontal tubes and the vertical tubes is less than 1%. The simulation results of the flow resistances of the 90°circular elbows are quite different from the empirical formula. The empirical formula of the sudden expansion tube is close to the simulation result of 60-degree diffusing tube. The error between the empirical formula of the reducing tube and the simulation result is small. Keywords: Flow Resistance · Numerical Simulation · Transfer Line

1 Introduction The large-scale helium cryogenic system is a key component of the great scientific engineering, which is an important platform for the development of basic scientific research [1]. The helium cryogenic transmission line is a key component in a large-scale helium cryogenic system, which mainly serves to connect refrigeration equipment and user equipment [2]. The design of the transmission line affects the performance of the helium cryogenic refrigerator directly. At the same time, the transmission lines of the cold compartment cycle and the cold shield cycle will also affect the overall efficiency of the large-scale helium cryogenic system. However, because the flow character of helium at low temperature is difficult to measure, the empirical formula of flow resistance at room temperature is often used © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 109–116, 2023. https://doi.org/10.1007/978-981-99-6128-3_13

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as design parameters in practical design. B Bradu used EcosimPro to simulate LHC beam screen cooling circuits and verified the accuracy of the model experimentally [3]. Rohan Dutta used Aspen Hysys (R) to build a dynamic model of helium liquefier and verified the helium liquefier model using cooling down data from existing helium plants [4]. That two software both employ the empirical formulas. C. Regier established a dynamic model of a liquid helium transfer line at the Canadian Light Source using empirical formulas with auxiliary properties equations, and verified it with experimental data [5]. N. Dittmar studied the influence of the structure on the pressure drop of the helium cryogenic transfer lines, and pointed out that the pressure drop can be reduced by optimizing the geometric parameters [6]. To study the suitability of the empirical formula under low temperature conditions, a helium cryogenic transmission line with an inner diameter of 54.7 mm, which is used for test facility cryogenic system of the Shanghai High repetition rate XFEL and Extreme Light Facility (SHINE), is selected as an example to simulate the flow resistance of helium along the horizontal straight tubes, vertical straight tubes, 90° elbows, diffusing tubes and reducing tubes [7].

2 CFD Models In order to study the flow resistance of helium at low temperature, several models are built using the software ANSYS Fluent. 5 typical working conditions of helium cryogenic refrigerator systems are selected, as shown in Table 1. Among them, case A corresponds to the main circuit cycle, case B and case C correspond to the cold compartment cycle, case D and case E correspond to the cold shield cycle. The properties of helium are calculated by REFPROP. Table 1. Calculation conditions Temperature(K)

Mass flow rate(g/s)

case A

Pressure(Bar) 3

4.5

175

case B

3

4.5

40

case C

3

8

40

case D

18

35

140

case E

17

55

140

The solution method is set to pressure base and steady-state, the turbulence model is set to the Standard k-omega model, and the pressure-velocity coupling uses the SIMPLE method. The inlet boundary condition is set to the mass flow inlet, and the outlet boundary condition is set to the pressure outlet. The simulation of straight line tube uses a two-dimensional axisymmetric model, and the horizontal and vertical lines are simulated by changing the direction of gravity. The Elbow model uses a three-dimensional model. In the calculation of the local resistance

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of the elbow, a 2 m straight tube is added outside the elbow due to the requirement of flow development. The two-dimensional axisymmetric model is built for simulating the diffusing tubes and reducing tubes, which is similar to the elbow model.

3 Results and Analyses In the calculation of empirical formula, the major losses hf are determined by the DarcyWeisbach equation (Eq. 1) while the minor losses hj are obtained by Eq. 2. hf =

fL v 2 D 2g

(1)

2

v hj = ξ 2g

(2)

3.1 Straight Tube Simulation Due to Re > 104 , Mcadams formula [8] for fully rough flow, shown as Eq. 3, is used for comparison. Figure 1 shows the flow resistance changes with the tube length under different working conditions at low temperature. The error decreases with the increase of the tube length. It shows that the empirical formula is accurate for the calculation of long lines. In fact, the length of the tubes used in the helium refrigerator usually is long, so the empirical formula can be perfectly applied to low-temperature pipelines. 160

case B case C

120

Flow Resitance/Pa

Empirical formula Simulation

case A

140

case D

100

case E

80 60 40 20 0

0

1

2

3

Length/m

4

5

6

Fig. 1. Flow resistance changes with line length in cryogenic

The flow resistances along the horizontal straight tube and the vertical straight tube under different working conditions are shown in Fig. 2. It shows that the flow resistances of the horizontal and the vertical straight tube are very close, which indicates that the influence of gravity on the flow resistance of the straight tube is almost negligible. f = −0.184Re−0.2

(3)

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case A case B case C case D case E

Flow resistance/Pa

140 120 100

Horizontal Vertical

80 60 40 20 0

0

1

2

3

Length/m

4

5

6

Fig. 2. Flow resistance changes with line length

3.2 Elbow Model In the calculation of the local resistance of the elbow, since the flow field needs to be fully developed, a sufficient length is added outside. The loss coefficients for 90 ° elbows are calculated by Eq. 4 [9]. The local flow resistance of the elbow is obtained by subtracting the flow resistance of the straight tube in the transition section from the total flow resistance. Therefore, the result obtained by the elbow will inevitably bring in the error of the calculation of the straight tube. However, according to the simulation of the straight tube, the error of the straight tube is small, while the error of the straight tube has less influence on the error in the elbow model. ξ = 0.131 + 0.163(d /D)3.5

(4)

Due to the different ratios of elbow radius and tube diameter in the elbow, there are different local resistance coefficients. Therefore, in the calculation, in order to verify the accuracy of the simulation calculation, the local resistance coefficient of the same transition radius ratio under different working conditions is simulated first. Taking the transition radius ratio as 0.1, the local resistance calculated by the simulation and by the empirical formula is shown in Table 2. With the same transition radius ratio, there is a stable error between the simulation and the empirical formula. Selecting case A as an example, the local resistance of different transition radius ratios is shown in Fig. 3. As the radius of the elbow transition decreases, the flow resistance of the elbow increases, and the error between the empirical formula and the simulation gradually increases. This may be due to the empirical formula derived from previous experiments. Due to the experimental conditions, the fluid used at that time was not cryogenic helium. At the same time, there are many empirical formulas or charts for elbows, and there is a certain gap between the results calculated by different empirical formulas. This paper only compares the much classical one.

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Table 2. Calculation results of flow resistance of elbow Empirical formula(Pa)

Simulation(Pa)

Error

case A

2.77

4.69

69%

case B

0.14

0.22

53%

case C

0.87

1.32

51%

case D

9.7

14.61

51%

case E

16.06

23.95

49%

30 Simulation

Local resistance/Pa

25

Empirical formula

20 15 10 5 0

0

0.2

0.4

0.6

0.8

Transition radius ratio

1

Fig. 3. Flow resistance of the elbow changes with the transition radius

3.3 Diffusing Tube Similar to the calculation of the local resistance of the elbow, in order to fully develop the flow field, straight tubes are added at the inlet and outlet of the diffusing tube. By monitoring the surface total pressure at the inlet and outlet of the model, the total pressure drop of the entire pipeline is obtained. ξ = (A2 /A1 − 1)2

(5)

The local resistance of different working conditions with the connecting tube angle θ of 60 ° and the tube diameter ratio of 1.2 is simulated, and subsequently the simulation results are compared with the empirical formula, as shown in Table 3. Due to θ > 45◦ , the loss coefficients for sudden expansion named as Borda-Carnot equation (Eq. 5) is used [10]. It can be seen from the results that the error between the simulation results of different pipelines and the empirical formula are similar.

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Simulation(Pa)

Error

case A

1.99

1.43

−28.13%

case B

0.10

0.07

−31.46%

case C

0.63

0.43

−30.99%

case D

6.97

4.74

−31.93%

case E

11.54

7.77

−32.69%

In case A, the comparison of simulation and empirical formula for different tube diameter ratios is shown in Fig. 4. In the calculation range, the empirical formula of sudden expansion tube can describe the local resistance coefficients of the diffusing tube with an angle of 60 ° accurately. 12

Local resistance/Pa

10 8 6 4

Simulation

2 0

Empirical Formula 1

1.2

1.4

1.6

Tube diameter ratio

1.8

2

Fig. 4. Flow resistance of diffuse line changes with the tube diameter ratio

3.4 Reducing Tube The comparison of simulation for different tube diameter ratios is shown in Fig. 5. The empirical formula in Ref. [11] is selected. When the tube diameter ratio is 1.2, 1.5 and 1.8 respectively, the local resistance coefficient ξ is 0.1, 0.165 and 0.18. When the tube diameter ratio is small, the simulation results are close to the empirical formula results, but with the increase in the tube diameter ratio, the simulated local resistance will be larger than the empirical formula result. In general, the empirical formula is close to the simulation results.

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50 Simulation

Local resistance/Pa

40

Empirical formula

30 20 10 0

1

1.2

1.4

1.6

1.8

Tube diameter ratio

2

Fig. 5. Flow resistance of shrinking line changes with the tube diameter ratio

4 Conclusion There is little error between the empirical formula result and the simulation result for the straight tube, which is within 10%. Gravity has little effect on the calculation results, that is, the calculation results of the horizontal tube and the vertical tube are almost the same. The elbow calculation results show that the empirical formula of the elbow is suitable for elbows with a larger transition radius. When the radius of the elbow transition is large, there is a relatively stable error in the flow resistance value calculated by the simulation calculation and the empirical formula. The error of reducing and expanding tubes is relatively small. When the angle of the diffusing tube is 60 °, the empirical formula of the sudden expansion tube can better conform to the simulation results. Acknowledgments. The project is supported by National Natural Science Foundation of China (Number 52076151).

References 1. Bonne, F., Alamir, M., Hoa, C., et al.: A Simulink library of cryogenic components to automatically generate control schemes for large Cryorefrigerators. Mater. Sci. Eng. 101, 012171 (2015) 2. Laeger, H; Lebrun, P., Rohner, P.: Long flexible transfer lines for gaseous and liquid-helium. Cryogenics 18(12), 659–662 (1978) 3. Bradu, B., Vinuela, E.B., Gayet, P.: Example of cryogenic process simulation using EcosimPro: LHC beam screen cooling circuits. Cryogenics 53, 45–50 (2013) 4. Dutta, R., Ghosh, P., Chowdhury, K.: Customization and validation of a commercial process simulator for dynamic simulation of helium liquefier. Energy 36, 3204–3214 (2011) 5. Regier, C., Pieper, J., Matias, E.: A dynamic model of a liquid helium transfer line at the Canadian light source. Cryogenics 51, 1–15 (2011)

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6. Dittmar, N., Haberstroh, C., Hesse, U., Krzyzowski, M.: Characterisation and optimisation of flexible transfer lines for liquid helium. Part I: Exp. Results. Cryogenics 75, 6–12 (2016) 7. Jiang, G.Y., et al.:The cryogenic control system of shine. EPJ. Tech. Instrum. 8(11) (2021) 8. MeAdams, W.H.: Heat Transmission. McGraw-Hill, New York (1934) 9. I.E.Idelqik. Hydraulic Resistance (1954) 10. Shao, Y.W.: Fundamentals of Fluid Mechanics. China Architecture & Building Press, Beijing (2015) 11. Chen, Z.R.: Engineering Fluid Dynamics, 3rd edn. Higher Education Press, Beijing (2013)

Comparison and Analysis of Two Hydrogen Liquefaction Processes Based on Helium Expansion Cycle Integrating with Mixed Refrigerant Pre-cooling Yujing Bi and Yonglin Ju(B) Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong, University, Shanghai, China {yujingb,yju}@sjtu.edu.cn

Abstract. To improve the efficiency by taking into account the premise of operation safety, two hydrogen liquefaction processes based on helium refrigeration integrated mixed refrigerant (MR) pre-cooling are proposed. In case 1, a sixcomponent MR is applied to the pre-cooling stage. In case 2, the four-composition of the MR is simplified to reduce the complexity of the proportioning. Meanwhile, a reference case based on the helium expansion cycle integrated with liquid nitrogen (LN2 ) pre-cooling is simulated under the same conditions. Aspen HYSYS is used to simulate the proposed processes, and the genetic algorithm (GA) is selected to optimize the refrigeration cycles. The first and second laws of thermodynamics are used to evaluate the liquefaction process. The output of 5t/d of liquid hydrogen can be achieved after the whole cycle. It can be found that the use of MR can effectively improve efficiency and reduce specific energy consumption (SEC). The simplification of components for MR will lead to a decrease in overall performance, but it will improve safety and reduce operational complexity. The combination of the MR pre-cooling cycle with the helium reverse Brayton cycle can provide guidance for energy saving for small and medium-sized liquefaction plants with safe operation as the primary goal. Keywords: Hydrogen liquefaction · Helium expansion cycle · Mixed refrigerant pre-cooling · Energy saving

1 Introduction In the context of carbon peaks, actively looking for clean energy to replace fossil fuels is the direction of global energy development. Hydrogen has great potential as a future energy carry due to its high calorific value and environmental friendliness. However, the storage and transportation of hydrogen over long distances restrict its feasibility as an energy vector. Hydrogen liquefaction is a promising direction for overcoming these issues, which can provide low-pressure, high-energy-density fuels for a variety of applications [1]. However, the existing hydrogen liquefaction plants, such as the Linde hydrogen liquefaction plant in Ingolstadt [2] and Leuna [3] in Germany, exhibit specific © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 117–124, 2023. https://doi.org/10.1007/978-981-99-6128-3_14

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energy consumption (SEC) in the range of 12.0~15.5 kWh/kg LH2 with the exergy efficiency (EXE) in the range of 19.0%~23.6% [4]. Therefore, high-efficient hydrogen liquefaction processes need to be designed and optimized to improve overall efficiency and reduce capital costs [5]. Hydrogen liquefaction processes can be divided into the Linde-Hampson cycle, Claude cycle, and helium (He)-refrigerated cycle [6]. Most of the actual operating plants are based on the Claude cycle with liquid nitrogen (LN2 ) pre-cooling because the SEC of it is the smallest. Nowadays, the development of reliable and efficient He refrigerators promotes the development of safer hydrogen liquefiers adopting the He-refrigeration cycle. The SEC of hydrogen liquefaction processes with the He expansion cycle is within a reasonable range on the premise of high safety. However, the current research is based on using LN2 as the refrigerant in the pre-cooling stage, which is a major reason for the inefficiency of the process. The characteristics of LN2 as a pre-coolant are not optimal because its isothermal vaporization of it will result in a large temperature difference. In addition, based on the global requirements for pure oxygen, LN2 will not be available as an inexpensive refrigerant for future large-scale hydrogen liquefaction plants [7]. The use of MR can effectively improve efficiency. However, we can find that the MR pre-cooling cycle has not yet been combined with the He expansion cycle to obtain a more efficient and safer hydrogen liquefaction system. In this study, two hydrogen liquefaction processes based on the helium expansion cycle integrating with MR pre-cooling are proposed and simulated in Aspen HYSYS, and the genetic algorithm (GA) is selected to optimize the key parameters. Two cases with different components and ratios of MR are compared and evaluated. Meanwhile, the performance indexes of the two proposed processes are compared with a reference case and other processes. The results show that the integrated process can maintain a reasonable level of energy consumption by considering the premise of process safety, and the use of MR can improve heat transfer performance. The simplification of the components of MR will make the performance worse but simplify the operation.

2 Process Design and Optimization 2.1 Process Description The hydrogen liquefaction process is roughly divided into three temperature stages: pre-cooling, cryogenic cooling, and liquefaction. The pretreated hydrogen is gradually cooled and liquefied through these three stages. Figure 1 shows the schematic diagram of the proposed hydrogen liquefaction process. The cooling capacity is provided by the MR cycle with different components and the He refrigeration cycle. In this work, two MRs with different compositions and ratios are applied to the pre-cooling cycle to compare the effect of different compositions on the performance of the process. The parameters are designed with reference to the actual plant, the feed hydrogen with a flow rate of 5 t/d has been purified and compressed before entering the liquefaction process and will be completely liquefied.

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119

· Case 1: 6 components · Case 2: 4 components

MR cycle

Helium refrigeration cycle

Cryogenic cooling

Pre-cooling

Liquefaction

Pretreated H2 21 bar, 298.15.K

Product LH2 1.5 bar, 21.65 K Ortho-para conversion

Fig. 1. The schematic diagram of the proposed process

The flow chart of the proposed process is shown in Fig. 2. The process consists of three parts: MR pre-cooling cycle, He refrigeration cycle, and hydrogen liquefaction route. The blue line represents the hydrogen route, and the orange and green lines represent the MR pre-cooling and the helium expansion cycle, respectively. The hydrogen is gradually cooled through seven heat exchangers. During this process, four ortho-paraH2 conversions are experienced. After two-stage compression and cooling, the helium enters the heat exchanger to be cooled and then expands twice to provide cooling capacity. MR only provides cold energy to pre-cool hydrogen to 80 K, and because the boiling points of various components are different, it undergoes two gas-liquid separators. C-202

C-201

Helium 219 204 203 202 Cooler-202 Cooler-201

217

218

201

301

321

322

323

309 205 303

Cooler-301 C-301

305

304

308

306

Cooler-302

Mixed refrigerant

319

320

Work flow/Heat flow 318

M-302 314

M-301 V-301

307 302

Hydrogen

216

313 V-302

312

215

316

311

C-302

214

211

209

103

102

210

212

208

207

HEX-4

V-101

E-201 108 109

106 107

105 101

213

317

315 206

E-202

V-303

S-302

S-301 310

HEX-6

HEX-5

112

110 111

104 HEX-1

HEX-2

113

HEX-7 Dewar

HEX-3 Co-1

Co-2

Co-3

Co-4

Fig. 2. Flow chart of the MR pre-cooling hydrogen liquefaction process

2.2 Simulation Conditions and Assumptions Aspen HYSYS is used to simulate and Peng-Robinson (PR) equation is applied to calculate the thermodynamic properties in this work. The following assumptions are set: (1) Streams after compression are cooled to 300 K; (2) The pressure drops in coolers and heat exchangers are ignored; (3) The adiabatic efficiency of the compressors and

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expanders are set at 80%; (4) All the flows in this model are steady-state, and the effects of kinetic energy and potential energy are ignored. The boiling curves of different components are shown in Fig. 3. The working temperatures of heat exchangers in the pre-cooling stage are indicated in the figure. The components of MR are determined by the boiling point distribution of different components under partial pressure, and the boiling points of selected working fluids should be as uniform as possible. The selected components of the MR are listed in Table 1. 101.35 kPa

400

200 kPa

350

Temperature (K)

300

298.15

250

233.15

200 173.15 150 123.15 100

Methane Ethane Propane i-Butane n-Butane i-Pentane n-Pentane Ethylene R14 Nitrogen Hydrogen Neon

80.15 50 0 100

1000

Pressure (kPa)

Fig. 3. Boiling curves for different components Table 1. The components of the mixed refrigerant Components (mol.%)

Case 1

Case 2

CH4

17.86

28.57

C2 H6

7.14

25.71

C3 H8

7.14

n-C4 H10

28.57

R14

23.57

N2

15.71

– 22.86 – 22.86

In order to compare different processes, specific performance parameters need to be defined. SEC, COP, and EXE are commonly used parameters for the performance evaluation of liquefaction systems. The expressions of the three indexes are: SEC =

Wnet · mproduct

=

WC−201 +WC−202 +WC−301 +WC−302 −We−201 −We−202 mLH ˙ 2

COP = EXE =

  mLH ˙ 2 × hFeed −hLH2 Wnet     mLH ˙ 2 hLH2 −h0 −T0 sLH2 −s0 Wnet

Wcooling Wnet

Wmin Wnet

=

=

(1) (2) (3)

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121

2.3 Optimization Method and Results As a nonlinear algorithm, GA is widely used in the liquefaction process optimization and it can be embedded into the simulation process. In this study, the minimum SEC of the whole liquefaction process is taken as the objective function. The penalty function will work when the minimum temperature approach of the heat exchanger is less than 1K, The variation ranges of the design parameters to be optimized remain the same for Case 1 and Case 2. The goal of the optimization is to obtain the optimal ratio of different compositions. The optimization results for the two cases are illustrated in Table 2. The SEC of Case 1 and Case 2 after optimization are 10.2139 and 10.4079 kWh/kg LH2 , respectively. Whether it is before or after optimization, the SEC of Case 1 is smaller than that of Case 2. For Case 1, the SEC of the optimized process is reduced by 1.46% compared to the base case, and for Case 2, it is reduced by 5.33%. It can be seen that optimization is of great significance to the performance improvement of the system. Table 2. The optimization results for Case 1 and Case 2 Parameter

Case 1

Case 2

nCH ˙ 4 (kmol/h)

22.34

57.21

nC2˙ H6 (kmol/h)

12.01

32.56

nC3˙ H8 (kmol/h)

0.7030

nn−C˙ 4 H10 (kmol/h)

37.90

nR14 ˙ (kmol/h)

30.08

nN˙ 2 (kmol/h)

21.38

– 34.79 – 10.40

3 Comparison and Analysis 3.1 Heat Transfer and Exergy Analysis The heat exchanger is the most numerous equipment in the process, and its efficiency has a great influence on the system. The worse the degree of matching between the hot and cold fluid curves, the greater the energy consumption and the exergy loss. Figure 4 and Fig. 5 illustrate the composite curves of the heat exchangers in the pre-cooling stage. The heat transfer effects before and after optimization are also compared. It can be seen from the horizontal comparison that the optimized heat exchange effects of the two cases have been improved. Through longitudinal comparison, it can be found that the temperature difference of the heat exchangers in Case 2 is larger than that in Case 1, which is due to the simplification of the MR compositions. The exergy losses of the main equipment after optimization are shown in Fig. 6. The exergy loss of the optimized Case 1 is 1522.88 kW, which is less than that of Case 2. After optimization, the EXE of Case 1 is 43.45%, which is 1.90% larger than that of Case 2.

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Coolers have the largest proportion of exergy losses because their large temperature difference leads to low efficiency. Exergy losses in compressors and expanders are mainly due to friction and heat losses during their operation. 100

Cold composite Hot composite Delta T

75

HEX-3 50 HEX-1

HEX-2

25

200

HEX-4

75

HEX-3 50 HEX-2

HEX-1

Delta T (K)

200

HEX-4

100

300

Delta T (K)

Temperature (K)

Cold composite Hot composite Delta T

Temperature (K)

300

25

100

100

189.76 389.14

678.56

0 1745.75

1157.80 Heat flow (kW)

189.76 388.42

0 1675.95

675.41 1132.40 Heat flow (kW)

Fig. 4. The composite curves of heat exchangers in the MR pre-cooling stage in Case 1: (a) Based case, (b) Optimized case. 100

HEX-4

Cold composite Hot composite Delta T

75

HEX-3 50 HEX-2

HEX-1 25

100

200

HEX-4

75

HEX-3 50 HEX-2

HEX-1

Delta T (K)

200

100

300

Delta T (K)

Temperature (K)

Cold composite Hot composite Delta T

Temperature (K)

300

25 100

189.76 415.42

761.90 1248.99 Heat flow (kW)

0 1893.75

189.76 386.47

670.33

1130.73 Heat flow (kW)

0 1702.72

Fig. 5. The composite curves of heat exchangers in the MR pre-cooling stage in Case 2: (a) Based case, (b) Optimized case. 1800 1600

Exergy loss (kW)

1400

Compressor Expander

Heat exchanger Valve

Cooler Others

39.48

42.05

268.41

268.41

610.86

622.03

313.78

336.71

257.57

262.09

Optimized case 1

Optimized case 2

1200 1000 800 600 400 200 0

Fig. 6. The exergy losses of the main equipment in optimized Case 1 and Case 2

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Table 3. Comparision of parameters and performance of different systems Processes

Kuzmenko et al. [8]

Ingolstadt [2]

Ref. Case

Case 1

Case 2

Yield (TPD)

5.4

4.32

Flow rate of LN2 (kg/h)

1980

1750

5

5

5

1681

–

–

Flow rate of He (kg/h)

3000

–

2282

2282

2282

SEC (kWh/kg LH2 )

12.7

14.55(13.58)

12.49

10.21

10.41

COP

–

–

0.1036

0.1267

0.1243

EXE (%)

34.6

21

35.53

43.45

42.64

3.2 Comparison Between the Proposed Process and Other Processes The LN2 pre-cooling He expansion refrigeration process under the same conditions except the pre-cooling cycle is modeled as the reference case. Except for the pre-cooling method, other parameters and thermodynamic arrangements are the same as that of Case 1 and Case 2. The detailed parameters and performance comparison with other processes are shown in Table 3, in which the parameters of Ingolstadt are the actual measured values. Energy consumption for LN2 production has been unified to 0.5 kWh/kg for a fair comparison. The horizontal comparison with other processes shows that the performances of the proposed processes are at a better level. In addition, it can be seen that the power consumption of Case 2 is higher than that of Case 1 due to the simplification of the components, but the simplification will reduce the operational complexity.

4 Conclusion In this study, two hydrogen liquefaction processes based on the helium expansion cycle integrating with MR pre-cooling are proposed. Two MR with six and four components for different ratios are used in Case 1 and Case 2. Meanwhile, a reference case with LN2 pre-cooling is established under the same conditions. Aspen HYSYS and GA are used to simulate and optimize the proposed processes. In conclusion, the proposed processes perform well by considering the premise of safe operation due to the physical properties of helium, and the use of six component refrigerants exerts better performance in the target for smaller SEC. Multi-parameter optimization will be the research direction in the future, and global optimization should be conducted to achieve higher efficiency. Acknowledgments. No known competing financial interests or personal relationships exist in the submission of this manuscript.

References 1. Yin, L., Ju, Y.L.: Process optimization and analysis of a novel hydrogen liquefaction cycle. Int. J. Refrig. 110, 219–230 (2020)

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2. Bracha, M., Lorenz, G., Patzelt, A., et al.: Large-scale hydrogen liquefaction in Germany. Int. J. Hydrog. Energy 19(1), 53–59 (1994) 3. Krasae-in, S., Stang, J.H., Neksa, P.: Development of large-scale hydrogen liquefaction processes from 1898 to 2009. Int. J. Hydrog. Energy 35(10), 4524–4533 (2010) 4. Naquash, A., Qyyum, M.A., Min, S., et al.: Carbon-dioxide-precooled hydrogen liquefaction process: an innovative approach for performance enhancement–energy, exergy, and economic perspectives. Energy Convers. Manag. 251, 114947 (2022) 5. Bi, Y.J., Yin, L., He, T.B., et al.: Optimization and analysis of a novel hydrogen liquefaction process for circulating hydrogen refrigeration. Int. J. Hydrog. Energy 47(1), 348–364 (2022) 6. Nandi, T.K., Sarangi, S.: Performance and optimization of hydrogen liquefaction cycles. Int. J. Hydrog. Energy 18(2), 131–139 (1993) 7. Wilhelmsen, Ø., Berstad, D., Aasen, A., et al.: Reducing the exergy destruction in the cryogenic heat exchangers of hydrogen liquefaction processes. Int. J. Hydrog. Energy 43(10), 5033–5047 (2018) 8. Kuz’menko, I.F., Morkovkin, I.M., Gurov, E.I.: Concept of building medium-capacity hydrogen liquefiers with helium refrigeration cycle. Chem. Pet. Eng. 40(1), 94–98 (2004)

Flow Process Analysis of the Cryogenic Distribution System for S3 FEL Project Xinbo Dong1(B) , Zheng Sun2 , Liangbing Hu1 , Shen He1 , Bihui Lai1 , and Xilong Wang2 1 Institute of Advanced Science Facilities, Shenzhen (IASF), Guangming District, Shenzhen,

China [email protected] 2 Dalian Institute of Chemical Physics, Chinese Academy of Sciences (DICP, CAS), Zhongshan Road, Dalian, China

Abstract. Shenzhen Superconducting Soft-X-ray Free Electron Laser (S3 FEL) is located at Shenzhen, China. It is planned to construct a 2.5 GeV CW superconducting radio frequency linear accelerator and four beam lines, aiming at generating X-rays with a high repetition frequency. The accelerator consists of twenty-five 1.3 GHz cryomodules (CM) and two 3.9 GHz cryomodules, of which its SRF cavity need to be operated at 2 K condition. To maintain the 2 K environment, cryogenic distribution system (CDS) is designed to interconnect the CMs and the Cryoplant (CP) and supply cryogens from the CP to the LINAC. This paper presents the preliminary design result and current status of the CDS, which including overall specification and design consideration. In addition, a thermodynamic calculation is performed based on current design result. Those analyses provide sufficient evidence for the performance margin of the CDS design. Keywords: S3 FEL · Superconducting Linear Accelerator · Cryogenic Distribution System · Pressure Drop · Flow Process Analysis

1 Introduction The Shenzhen Superconducting Soft-X-ray Free Electron Laser (S3 FEL) project will be located at Guangming district, Shenzhen city, Guangdong province, China. It is proposed by the Institute of Advanced Science Facilities, Shenzhen (IASF) which is a multi-disciplinary research center based on the integrated particle facilities funded by Shenzhen government. The main purpose of S3 FEL project is to construct a 2.5 GeV CW superconducting radio frequency (SRF) linear accelerator and four experimental beam line, aiming at generating X-rays between 40 eV and 1 keV at a rate up to 1 MHz for scientific research, including drug development, energy science, advanced materials and etc. In order to achieve a high repetition frequency, The S3 FEL Linear Accelerator (LINAC) is based on superconducting radio frequency (SRF) cavity technology which is similar to LCLS-II [1] and SHINE [2] project. According to physic design results, © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 125–132, 2023. https://doi.org/10.1007/978-981-99-6128-3_15

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the electron beam will be accelerated by twenty-five 12-m long 1.3 GHz cryomodules (CMs) and a third-harmonic LINAC comprised of two 3.9 GHz cryomodules (CMs). The LINAC is cooled by the Accelerator Cryoplant (ACCP). To verify SRF technology and support the project commissioning, a test facility will be built served by a Test Facility Cryoplant (TFCP). A prototype accelerator integrated with several CMs will be operated served by a Prototype Accelerator Cryoplant (PACP) before commissioning the main accelerator. All cavities are designed to operate at 2 K. Three independent cryogenic distribution systems (CDS) are needed to supply and return the cryogens from corresponding cryoplant to test benches or to the LINAC via cryogenic pipes. The block diagram of S3 FEL cryogenic system is shown in Fig. 1.

Fig. 1. The block diagram of S3 FEL cryogenic system

This paper mainly focuses on the cryogenic system of the main accelerator of S3 FEL project. The preliminary design results and status are introduced, including CDS specification and cryoplant requirements. In addition, a flow process calculation considering heat load budget, pressure drops and temperature profiles change at nominal operation mode and maximal capacity operation mode are analyzed.

2 Cryogenic System Overall Design 2.1 Overall Specification The S3 FEL cryogenic system consists of three major subsystems: the cryogenic plant (ACCP), the upstream and downstream LINAC and the cryogenic distribution system (CDS). A cryogenic system overview schematic is shown in Fig. 2. The ACCP and the CDS consist of two sets of cold box (CB1 & CB2), two sets of distribution box (DCDB & UCDB), one interconnection distribution box (IB) and other associated auxiliary equipment. The CDS supplies the cryogens from DCDB & UCDB to CM strings (L1, L2 & L3) through transfer lines (TL). At the beginning of the upstream LINAC, two parallel injector units are designed to generate stable electron sources and a sub-distribution box (IVB) is designed to provide them with the helium cooling. The CDS distributes cryogens from the ACCP to the

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LINAC cryomodules via cryogenic transfer lines with feed caps and end caps (FC & EC) connected to each CM string in the LINAC tunnel. The cryogens flow through each LINAC cryogenic string: upstream to FC, L2, L1 and L0 and downstream to FC, L3 and EC. The upstream LINAC has a total 15 CMs while downstream LINAC has a total of 13 CMs. At the end of each string, the flow is rerouted through EC back through the CM string, transfer lines and the distribution boxes to the cold box.

Fig. 2. The cryogenic system overview schematic diagram

The required cryoplant capacity is 4 kW @ 2.0 K, 1.7 kW @5 K, and 16 kW @ 45 K, which will provide a certain margin above the cryogenic heat load [3]. However, to eliminate risk of insufficient cooling capacity in case of cavity performance degradation, it is decided to install two identical cold boxes (CB1 & CB2). When the 2 K heat load exceeds 4 kW, two CBs will be operated simultaneously. The CB1 will supply the upstream LINAC while the CB2 supply the downstream LINAC. 2.2 CM and Transfer Line Based on the mechanical design results, the CM has six main process to transfer cryogens and two auxiliary pipelines, herein named as line A to H. The section view of the CM is shown in Fig. 3

Fig. 3. The 3D model and schematic diagram of the cryomodule

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As described in Sect. 2.1, cryogenic transfer lines with six corresponding pipelines are designed to connect CM string with distribution box via FC & EC. Table 1 shows nominal operating parameters and sizes for these CDS components. Table 1. CDS line size and nominal operating parameters Pipe ID Nominal pipe Nominal pipe size of Nominal pipe size of Temperature Pressure size of CM TL Caps (K) (bar) Line A

DN50

DN 50

DN 50

2.3

3.0

Line B

300 mm

250 mm

250 mm

2

0.031

Line C

DN50

DN 50

DN 50

5

3.2

Line D

DN50

DN 50

DN 50

8

3.0

Line E

DN50

DN 50

DN 50

45

19.2

Line F

DN50

DN 50

DN 50

80

19.0

2.3 CDS Design Consideration The nominal heat load of S3 FEL is 3.55 kW @ 2.0 K, 1.30 kW @ 5 K, and 12.2 kW @ 45 K, including transfer line, distribution boxes, feed & end caps and CMs. The details of the heat load calculation can be found in reference [3]. Based on heat load budget, the nominal mass flow rate of upstream and downstream CM stings are calculated. Besides the nominal operation mode, the CDS shall fulfill the maximum cooling capacity in case of cavity performance degradation. For this case, the heat loads of Line C-F are almost constant comparing with nominal mode. The 2 K total heat load varies with dynamic heat loads determined by Q0 factor of the cavity. Table 2 shows the mass flow rate of nominal operation mode and maximum operation mode. The limitation of ACCPs is also considered during the CDS design process. The primary operating constraints are the pressure drop and temperature change of the supply and return line at the interface with UCDB and DCDB. For Line B, the pressure drop budget is more critical. The inlet pressure shall be not less than 27 mbar considering Table 2. Mass flow rate of each line Pipe ID

Upstream Mass flow (g/s)

Downstream Mass flow (g/s)

Nominal

Maximum

Nominal

Maximum

Line A (inlet)

87

165

81

150

Line B (outlet)

87

165

81

150

Line C& D

27

21

Line E & F

49

35

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the cold compressor limitation while the pressure in the helium bath is 31 mbar. The pressure drop along the whole 2 K return line shall be less than 4 mbar, then a 2 mbar pressure drop budget is given for Line B.

3 Pressure and Temperature Profile Analysis 3.1 Calculation Method To simplify the calculation process, all CMs subcomponents are divided to different calculation nodes and then calculated in sequence. The pressure drop is based on piping model method with flow parameters (density:ρ) updated at each node using HePak properties and then the parameter (velocity:v) is calculated. The input parameters include m (mass flow), D (pipe diameter), L (pipe length), h (height change) and ε (pipe roughness). The pressure (P) Eq. (1) is listed below. ρv2 ρv2 L × + fk + ρgh D 2 2   1 ε 2.51  = −2.0log +  3.7D Re f f

Pj = Pi + f ×

(1)

(2)

where f is friction coefficient calculated from Colebrook Eq. (2) and fk is resistance coefficient for fittings [4]. Initial and final properties of each node are denoted with subscripts i and j. The temperature change is based on the heat load accumulation with length of transfer line and numbers of CMs increase, shown as Eq. (3) Tj = Ti +

Qload m × Cp(Ti ,Pi )

(3)

where Q and Cp represent heat load and specific heat capacity, respectively. The mass flow is constant for Line C-F while Line A and Line B need to consider the mass flow change along the upstream and downstream CM strings. Especially for Line B, the dynamic heat load will evaporate the He-II to saturated helium gas and then the boil off gas flow into B pipe mixing with upstream helium gas. The line B pressure of each CM also affects the saturated temperature of boil off gas at the same time [5]. The mixing temperature is calculated by Eq. (4). hm =

mi hi + mvap hvap   then Tm = T (hm , Pm ) mi + mvap

(4)

where h represents enthalpy, mvap represents mass flow rate of the boil off helium gas.

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3.2 Analysis Results Figure 4, 5, 6, 7 and 8 present the analysis result of pressure drop and temperature profile for Line A to F. The x-axis represents the distance of each LINAC calculation node from the main distribution box along the direction of the accelerator tunnel. It can be concluded from the Fig. 4 and Fig. 5. That the pressure drops and temperature changes are within the budget of process design. In Fig. 4, it has an obvious temperature difference at the end of upstream and downstream CM strings. Especially, the temperature increases fast at IVB section. This is mainly because the mass flow in Line A is almost reduced to zero near the end of CM sting that the remaining helium gas is more vulnerable to the heat load coming from IVB and the transfer lines. Figure 6 presents pressure drop and temperature change profile of line B in maximum flow condition. The pressure drop of upstream and downstream is 68 Pa and 46 Pa respectively. It means that a sufficient margin is reserved for the pressure drop under the maximum flow condition. Figure 7 and Fig. 8 present the accumulated pressure drop and temperature profile for the 5 K circuit (Line C and D) and the 40 K circuit (Line E and F) VS. The distance, respectively. From these two figures, a relatively large pressure difference occurs at X = 0 position can be found between upstream and downstream CM strings. It means that the mass flow value shall be further optimized.

Fig. 4. Pressure drop and temperature change profile for Line A in nominal operation mode

Fig. 5. Pressure drop and temperature change profile for Line B in nominal operation mode

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Fig. 6. Pressure drop and temperature change profile for Line B in maximum flow rate mode

Fig. 7. Line C&D combined pressure and temperature profiles VS LINAC distance

Fig. 8. Line E&F combined pressure and temperature profiles VS LINAC distance

4 Summary The preliminary design and status of the main S3 FEL cryogenic distribution system are introduced in this paper. A thermodynamic calculation considering heat load budget, pressure drops and temperature change at nominal operation mode and max capacity operation mode are analyzed. The results show that the CDS is designed with sufficient margin while considering a variety of CM string, subcomponents based on current 3D layout.

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Acknowledgments. The project is supported by the Shenzhen Local Government. The authors would like to express our gratitude to Prof. Guy Gistau Baguer and John Weisend II for their constructive comments and advices.

References 1. Soyars, W., et al.: Status of the LCLS-II cryogenic distribution system. In: IOP Conference Series: Materials Science and Engineering, vol. 755, no. 1, p. 012059. IOP Publishing (2020) 2. Zhao, Z., Wang, D.: XFEL projects in china. In: Proceedings of the 29th Linear Accelerator Conference (2018) 3. Hu, L.B., et al.: Conceptual design of S3FEL cryogenic system. In: 23rd Cryogenic Engineering Conference and International Cryogenic Materials Conference (2021) 4. Crane Co.: Flow of Fluids Through Valves, Fittings, and Pipe TP-410. Version 2009. Crane Co, USA (2010) 5. Dalesandro, A., Kaluzny, J., Klebaner, A.: Thermodynamic analyses of the LCLS-II cryogenic distribution system. IEEE Trans. Appl. Supercond. 27(4), 1–4 (2016)

Fabrication, Cryogenic Test and Installation of the ESS Cryogenic Moderator System H. Tatsumoto1(B) , Y. Beßler2 , E. Rosenthal2 , P. Arnold1 , M. Kickulies1 , M. Boros1 , A. Horvath1 , M. Segerup1 , P. Tereszkowski1 , and D. Lyngh1 1 European Spallation Source ERIC, Partikelgatan, Lund, Sweden

[email protected] 2 Forschungszentrum Jülich GmbH, Wilhelm-Johnen-Straße, Jülich, Germany

Abstract. The ESS cryogenic moderator system (CMS) has been designed to circulate subcooled liquid hydrogen with a temperature of 17 K and a parahydrogen concentration of more than 99.5% at 1.0 MPa in order to slow down the high energy neutrons by means of two hydrogen moderators. The CMS cold box (CBX) had been fabricated in November 2020 by the ESS in-kind partner, Forschungszentrum Jülich GmbH (FZJ). Subsequently, the CMS CBX cryogenic test had been implemented by using liquid nitrogen until September 2021 at FZJ. The CMS CBX was cooled by a mixture of GN2 and LN2, instead of a helium refrigerator. The performance test for two hydrogen pumps and a measured pressure drop met the design requirement. Furthermore, it was verified that a combination of a heater and a release control valve, which was called a pressure control system, could maintain the CMS pressure at a set point. The CMS installation with the exception of the moderators will be completed at the ESS site by June, 2022. Keywords: Liquid hydrogen · Spallation source · Hydrogen moderator

1 Introduction At the ESS target, high energy spallation neutrons are produced by impinging 5 MW proton beam on the high-Z material, tungsten. The proton beam is pulsed with a repetition of 14 Hz and a pulse length of 2.86 ms. The spallation neutrons are moderated to cold and thermal energies by the moderators. The moderator system consists of a water premoderator, a thermal moderator and two liquid hydrogen cold moderators, all optimized to achieve a high cold neutron brightness [1]. The neutronic performance of the cold moderators degrades rapidly with the decreasing parahydrogen fraction below 99.5%. The nuclear heating is estimated to be 6.7 kW for the 5-MW proton beam power. The ESS cryogenic moderator system (CMS) has been designed to circulate subcooled liquid hydrogen (LH2 ) of a temperature of 17 K and a pressure of 1 MPa to remove the nuclear heating generated in the two hydrogen moderators [2]. There are two pumps, a pressure buffer tank (PCB) with a volume of 0.065 m3 , an ortho-parahydrogen convertor and two heat exchangers in the CMS cold box (CBX). The CMS is cooled by a helium refrigeration plant called the Target Moderator Cryoplant (TMCP) via the plate-fin type © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 133–140, 2023. https://doi.org/10.1007/978-981-99-6128-3_16

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heat exchanger. LH2 is supplied to a distribution box (DB) through a hydrogen transfer line (HTL). The HTL is subsequently split into distribution lines for each moderator. All the hydrogen can be released to the atmosphere through a hydrogen vent line (HVL). Gaseous hydrogen (GH2 ) with a pressure of 2 MPa will be supplied from a GH2 filling station located at the outside of the building via a 100 m-long feed line. The CMS CBX has been fabricated in 2020 by the ESS in-kind partner: Forschungszentrum Jülich GmbH (FZJ) and the cryogenic test has been performed using liquid nitrogen (LN2 ). The whole CMS installation started in January 2022. The first commissioning will start Q3, 2022, using helium instead of hydrogen.

2 Cryogenic Test of the CMS CBX 2.1 Development of a Mixing System Figure 1 shows a set-up of the CMS CBX cryogenic test at the FZJ. A GN2 -LN2 mixing system with design pressure of 0.5 barg was developed to adjust a cooldown temperature and maintain the CMS at a desired temperature, instead of the TMCP. There is a LN2 dewar with a volume of 20 L on a scale. The liquid level was measured by the weight change. The LN2 was supplied via PV-02 before the LN2 run out. The feed of GN2 splits into two lines. One leads to the bottom of the LN2 cryostat to produce cold GN2 . The other is to supply ambient GN2 . The cold GN2 is mixed with the ambient GN2 and adjust the feed temperature at TE-01. 2.2 Cool-Down Test The cooldown operation of the cryogenic test consists of two phases: Phase I and Phase II. In the Phase I, the CMS filled with GN2 at 1 MPa and 296 K was cooled down to 100 K by the mixing system. One of the pumps ran at around 4000 rpm. Additional GN2 was not supplied to maintain the pressure (PT62037) at 1 MPa. Figure 2 shows the Phase I cooldown test result. The feed flow rate of the mixing system is around 10 g/s.

Fig. 1. Overview of the ESS cryogenic moderator system (CMS).

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The feed temperature was decreased at around 0.48 K/min. If temperature differences at the both ends of the heat exchanger (HX-61100) exceeds 35 K, decreasing the feed temperature was tentatively stopped. The CMS circulation flow rate was increased with decrease in temperature. It takes 6.8 h to cool down the CMS to 100 K where the pressure was decreased to 0.4 MPa and the operational mode was moved into the Phase II. The CMS was depressurized and LN2 was directly fed to the CMS at the interface of the supply HTL, meanwhile the mixing system continuously supplies LN2 to the HX-61100. Evaporated GN2 was released via CV-62001.

Fig. 2. Cool-down operation at the cryogenic test.

2.3 Hydrogen Pump Performance Test Two ball bearing type centrifugal pumps are arranged in series. An impeller has a diameter of 95.4 mm. The revolution speed, N , is 1,000 to 14,000 rpm. A discharge flow rate, m, ˙ was measured by an orifice flow meter and a pump head, P, was estimated by a difference between the two pressure transmitters calibrated in advance. Figure 3 shows the pump performances measured under the conditions of (a) liquid nitrogen (LH2 ) at 78.8 K and 0.51 MPa, (b) cold GN2 at 123 K and 0.49 MPa and (c) GN2 at 297 K and 0.5 MPa. The pump head decreases monotonically with increase in flow rate. With increase in pump speed, both P and m ˙ become larger as well. Both of the measured pump performance curves are on the same curve at the same revolution speed. It is known that the performance curve can be arranged using dimensionless expressions of a head coefficient, ψ, discharge coefficient, φ, and the wheel speed, us . −1  (1) ψ = P ρuS2 −1  φ=m ˙ ρus D2

(2)

us = π ND/60

(3)

where ρ is the density and D is the impeller diameter.

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Figure 4 shows the non-dimensional expression of the measured pump characteristics. All the measured pump performances are on the same curve, independent of the temperature and revolution speed. It is verified that the measured performances agree with the design curve within 10%. A fitting curve is newly obtained based on the experimental data in Fig. 4. The pump performances given by the new fitting curve are shown in Fig. 3. They can adequately capture all measured data and further, it is helpful to study a CMS operational procedure.

(a) LH2 at 78.8 K and 0.51 MPa

(b) GN2 at 123 K and 0.49 MPa

(c) GN2 at 297 K and 0.5 MPa

Fig. 3. Measured LH2 pump performance test results.

2.4 Pressure Drop The pressure drops over the CMS CBXP were measured under the condition of (a) LN2 at 79 K and 0.5 MPa and (b) GN2 at 130 K and 0.5 MPa where the bypass valve of CV-62037 was at the position of 100%. Figure 5 shows the measured pressure drops. The flow rates were changed by the pump revolution speed. Tatsumoto et al. [2] have estimated the pressure drop at the nominal condition in the design phase. The pressure drops over the CMS CBX are recalculated according to the current configuration and

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Fig. 4. Dimensionless pump performance curve.

conditions and are also shown in the figure. Fluid properties of parahydrogen at 17 K and 1.0 MPa given by GASPAK [3] have been used. The friction factor of a pipe is calculated using Colebrook equation [4], where a surface roughness of 0.05 mm is used. The measured pressure drops in LN2 and GN2 agree with the calculated values within 15%. It is verified that the CMS CBX has been fabricated as designed.

(a) LH2 at 79 K and 0.5 MPa

(b) GN2 at 130 K and 0.5 MPa

Fig. 5. Measured pressure drops.

2.5 Pressure Control at the Nominal Condition The pressure control buffer (PCB) is prepared to maintain the pressure at a set point. Six temperature sensors are located along the height of the PCB to identify the liquid level, which correspond to 0 (TE-100), 3.14 (TE-102), 10.2 (TE-104), 15.7 (TE-106), 21.2 (TE-108) and 25.9 L (TE-110) as shown in Fig. 6. There are heaters on the PCB and a release valve (CV-62029) on the top of it. The heater produces evaporated gas, increasing pressure. The excess gas is released by CV-62029 in order to adjust the pressure to the

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set point. The released gas is condensed via the HX-61200 and comes back to the main loop. Figure 7 shows a pressure control functional test result of the PCB. LN2 with a temperature of 78.7 K is circulated through the CMS loop by operating two pumps at 3000 rpm and the liquid level of LN2 in the PCB was maintained above the highest location of the level sensors. The total heater power was kept at 700 W and the LN2 temperature was maintained at its saturated temperature of 94 K at 0.5 MPa. The vapor temperature is kept around 102 K by a circulation flow through the CV-62029. The released evaporated GN2 can be condensed and cooled down to 81 K through the HX61200. LN2 is refilled into the PCB and the temperature around the bottom, TE-100 and TE-102, is kept close to the loop temperature. It is verified through the cryogenic test that the CV-62029 PID control and the HX-61200 works well and the pressure can be adjusted to the set point of 0.5 MPa as designed (Fig. 7).

Fig. 6. CMS pressure control system.

(a) CMS CBX

Fig. 7. Pressure control test result.

(b) Top plate of the CMS CBX

Fig. 8. CMS CBX fabrication at FZJ.

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3 CMS Fabrication and Installation Status The CMS CBX fabrication had been completed by the ESS in-kind partner, FZJ, in November, 2020 as shown in Fig. 8. The CMS CBX is separated into two vacuum sections with a diameter of 1.8 m and a total height of 1.6 m, which are suspended from the top flange, to realize a compact cryostat stand and installation height. A hydrogen vent line, to which all the pipes to release hydrogen are connected, is routed from the outlet of the vacuum safety device on the top flange. A Gas Management Panel (GMP) to supply GHe, GH2 and GN2 is mounted on one side of the CMS CBX. An electrical cabinet and a water-cooling system for the hydrogen pumps are temporarily placed close to it for the cryogenic test. The cryogenic test has been completed in September, 2021, as described above. The CMS CBX was delivered to the ESS site and was lifted to the hydrogen room located at the 4th floor of the ESS Target building as shown in Fig. 9. The 38-m long hydrogen transfer lines and distribution valve box had been fabricated around the same time. The installation of the CMS except for the moderators will be completed by the end of May, 2022. As of 1st May, the installation status is shown in Fig. 10. All of the installation work (the CMS CBX, the HTL and the Distribution box) has been completed, with the exception of the hydrogen vent line on the roof top and a hydrogen filling station. As the first step, the commissioning plans to be continuously conducted using helium for 10 months in order to optimize the operational parameters of the CMS and TMCP and accumulate their operational experiences. The first-time hydrogen operation, which will be carried out in the same configuration, cool down the CMS to 17 K at 1.0 MPa and study of the mitigation of the pressure and temperature fluctuation. The moderators will have been installed at the end of 2023. The commissioning of the whole CMS will have been completely finished in 2024.

Fig. 9. CMS CBX delivery

Fig. 10. CMS installation status

4 Conclusions The CMS CBX cryogenic tests have been successfully completed by using LN2 in September, 2021. The developed GN2 -LN2 mixing system worked well to control the CMS temperature to the desired value. It was verified that the pump performance, the pressure drop over the CMS CBX and the pressure control performance met the design

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requirement. The installation of the CMS except for the moderators will be completed at the ESS site by summer 2022 and the commissioning will be conducted using helium as the first step. The moderator will be installed at the end of 2023 and the commissioning of the whole CMS using hydrogen will be completely finished in 2024.

References 1. Garoby, R., Danared, H., et al.: Corrigendum: the European spallation source design. Phys. Scr. 93, 014001 (2018) 2. Tatsumoto, H., et al.: Design status of the ESS cryogenic moderator system. In: IOP Conference Series: Materials Science and Engineering, vol. 755, p. 012101 (2020) 3. GASPAK user’s guide 1998 Cryodata 4. Colebrook, C.F.: J. Institution of Civil Engineers. 11(4), 133–156 (1939)

The Economic Analysis of Helium Liquefaction Plant and Helium Recovery System After Improvement Li Zhun1,2,3(B) , OuYang Zhengrong1,2 , Luo Yuchen3 , Zhang Xuehua4 , and Wang Zezhang3 1 Hefei Institutes of Physical Science, Chinese Academy of Sciences, Hefei, Anhui, China

[email protected]

2 Univercity of Science and Technology of China, Hefei, Anhui, China 3 National Institute of Metrology, Beijing, Chaoyang, China 4 Anhui Vacree Technology Co., Ltd., Hefei, Anhui, China

Abstract. Taking Linde L70 helium liquefaction and recovery plant as an example, this paper introduces the optimization and improvement of the system, which is equipped with a set of external purifier. By comparing the recovery helium with different purity, the liquefaction amount and liquefaction rate of the unit before and after the purifier is installed are analyzed, and the energy conservation and economy of the unit are calculated and compared. Considering the energy saving and economy, it is concluded that the energy saving effect of the system is obvious and the economic benefit is improved, after the external purifier is installed. Keywords: Helium liquefaction · External purifier · Energy saving · Economic

1 Preface In recent years, Asia has been the fastest growing region of helium demand in the world. By 2021, the global annual demand of helium was expected to be 226.6 million cubic meters. According to forecasts from relevant Russian institutions, by 2030, helium demand will reach 220 million to 300 million cubic meters, of which the proportion of the United States will decrease from 50% to 45%, Europe will decrease from 20% to 15%, and Asia-Pacific countries from 25% to 30%. The average annual growth rate (AAGR) of helium demand market in China is higher than 10%.The annual demand is 22 million cubic meters, while the actual supply is only 16 million cubic meters. The difference is 5.4 million cubic meters. It is estimated that China’s annual demand for of helium will increase to 22.4 million cubic meters in 2021, while the supply will maintain at 16.3 million cubic meters, and the gap between supply and demand will further increase to 6.1 million cubic meters [1, 2]. With the rapid development of China’s aviation, aerospace and defense industries, the total domestic demand of helium is growing, with an annual demand of about 22 million cubic meters. According to data, the global helium supply is generally short, with a gap of 50 million cubic meters per year [3]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 141–148, 2023. https://doi.org/10.1007/978-981-99-6128-3_17

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Helium is one of indispensable and rare strategic materials for high-tech industries, but the global distribution of helium resources is extremely uneven. According to the 2021 BLM survey report, the total global helium resources are about 68 billion cubic meters, mainly concentrated. in the United States, Qatar, Algeria, Poland and Russia. Amounting to 88% of global helium. China’s resources account for 2.1% of the global total [4]. In summary, China has a large demand for helium, but at the same time its own helium resources are limited. It is of great significance to research and analyze helium purification and recovery technologies.

2 Helium Purification Process The built-in purifier of the Linde L70 Helium Liquefier is used to remove impurities such as trace water, oxygen, nitrogen and other impurities in pure helium. If a large amount of impurity gas is mixed into the helium gas recovery system, the liquefaction efficiency of the helium liquefier will be reduced, wasting more electrical energy. After sampling tests, the purity of the recovered helium in the No.23th cryogenic building helium recovery system is about 90%–99%. The original helium recovery system is now optimized and improved. An external purifier is installed to form a new helium purification system, as shown in Fig. 1.

Fig. 1. Flow chart of helium liquefaction and recovery system after adding external purifier

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2.1 System Composition The helium purification system is mainly composed of a 30 cubic meter helium air bag, a high-pressure helium compressor, a membrane separation purifier (with its own compressor), a pure helium buffer air bag, and an automatic control unit. The equipment is shown in Fig. 2.

Fig. 2. Helium purification system

2.2 Process Design The system can purify and pressurize helium with a helium content of about 90% to 99% recovered in the No.23th cryogenic building, and fill it into a helium container for storage. According to the scene conditions, the compressor inlet uses two 30 cubic meters of buffer airbags of the original system. The recovered helium gas is directly charged into the buffer airbag. As the amount of gas filled gradually increases, the volume of the airbag gradually increases. When the volume increases to a certain extent, the stroke switch will be triggered, thereby activating the membrane separation and purification device to purify the helium gas. To more than 99%, at the outlet of purified helium gas, the helium purity is monitored in real time by a helium purity meter to meet the purity requirements, as shown in Table 1.The equipment used in this article is the original equipment in Linde 70, which are specialized in measuring helium recovery and purification.

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Heading

Unit/percentage Target value Remark

Processing medium

%

Pure helium

Impurity category Inlet helium purity

Water, oxygen, nitrogen %

90%–99%

Outlet helium purity

%

99.0%

Helium recovery ratio

%

90%

Ratio of above 99% helium to the amount of helium in recovered helium

Purification working pressure MPa

2.00

Booster working pressure

MPa

15.00

Working frequency

hr

4

Work interval between two starts

Helium analyzer accuracy

%

±0.1

Measuring range: 0–100%

2.3 Principle Because cryogenic experiments require frequent insertion and extraction of experimental samples in liquid helium vessel, the evaporated helium gas will be mixed with air to reduce its purity. Therefore, further purification is required to meet the needs for reuse. The helium purification uses membrane separation technology and process, which can purify the recovered helium to more than 99%, and meet the filling requirements after pressurization. As for the helium recovery and helium purification flow, the existing two 30 cubic helium buffer airbags were combined, and the volume of the new buffer airbag triggered the work of the membrane separation and purification equipment. The low-pressure compressor built in the equipment pressurizes the recovered helium to 10–15 bars, and then enters the membrane separation and purification unit for purification operation. The purity of the purified helium after the purification reaches to more than 99%. The purified helium gas is sent to the finished gas buffer tank, and the pressure sensor in the buffer tank is used to trigger the helium high-pressure compressor to be pressurized to a pressure of 150 bars or above to fill the helium gas cylinder. As shown in Fig. 1, after the purity of helium reaches the standard, the cylinder group is filled by pressurization. The helium gas that does not meet the requirements is returned to the new buffer tank through the pipeline for cyclic purification operation until it meets the requirements for use.

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3 Operational Performance Test of Helium Liquefaction and Recovery System After Improvement In actual operation, the performance parameters of liquefaction without external purifier and liquefaction with external purifier can be tested separately. In the improved design of the system, maintenance of the external purifier and other issues are considered. By closing the inlet and outlet solenoid valves and shutoff valves of the external purifier, the process can be switched to the original purification process of the internal purifier. The acceptance performance test table is as follows Table 2. Table 2. Measured performance parameters of helium liquefaction and recovery system after improvement. Test date

With or without external purifier operation

With or without Liquefaction liquid nitrogen output: liter precooling (L)

Liquefaction output: Liter hour (L/h)

Remark (Test duration)

8:00 May 13th to 8:00 May 14th, 2019

NO

YES

284.8

11.9

24 h

8:00 May 16th to 8:00 May 17th, 2019

YES

YES

645.7

26.9

24 h

8:00 Sep 23th NO to 8:00 Sep 24th, 2020

YES

288.9

12.0

24 h

8:00 Sep 27th YES to 8:00 Sep 28th, 2020

YES

644.1

24.8

24 h

8:00 Jun 22th NO to 8:00 Jun 23th, 2021

YES

283.1

11.8

24 h

8:00 Jun 25th YES to 8:00 Jun 26th, 2021

YES

632.6

26.3

24 h

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4 Uncertainty Analysis The resolution displayed by the temperature equipment of the measuring system is 0.01 °C, and the half width √ of the uncertainty interval is 0.005 °C. According to the uniform distribution, k = 3, the standard uncertainty is 0.003°C as shown in formula 1: 0.005 u(td ) = √ ≈ 0.003◦ C 3

(1)

The resolution displayed by the pressure equipment of the measuring system is 0.1 MPa, and the half width √ of the uncertainty interval is 0.05 MPa. According to the uniform distribution, k = 3, the standard uncertainty is 0.029MPa as shown in formula 2: 0.05 u(Pd ) = √ ≈ 0.029MPa 3

(2)

The resolution displayed by the flow rate equipment of the measuring system is 0.1 L, and the half width √ of the uncertainty interval is 0.05 L. According to the uniform distribution, k = 3, the standard uncertainty is 0.029L as shown in formula 3: 0.05 u(Ld ) = √ ≈ 0.029L 3

(3)

The uncertainty of this measurement experiment is combined according to the above uncertainty components, as shown in formula 4:  (4) µ = µ2 (t) + µ2 (p) + µ2 (l) = 0.04

5 Economic Analysis of Helium Liquefaction and Recovery System After Improvement Analysis of the data in Table 1 and Table 2 shows that in the normal operation of the helium liquefaction and recovery unit, the same amount of electricity is consumed, with or without external purification operation, and the production of liquid helium varies widely. Average the two types of date, then formula 1 can be obtained as follows:   Power consumption per unit time   of energy consu min g equipment Power consumption (5) = per liter of liquid helium (Liquid helium output) Table 3 is the rated power of the main energy-consuming equipment of the helium liquefaction and recovery unit. Compared with the operating condition of the external purifier and the operating condition of the external purifier, according to formula (5), it can be obtained that the power consumption per unit liter of helium output is about 8.529 kWh/L, and the operating conditions with an external purifier are about

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Table 3. Performance parameters of main energy consumption equipment of helium liquefaction and recovery system The name of the device

Main helium cycle compressor

Turbine chiller

Air compressor

External purification device

Rated power (kW)

45

3.5

45

8

3.801 kWh/L. The working condition can save 124.4% of electric energy compared with the operating condition of the external purifier, which has a significant energy saving effect. At present, the helium liquefaction and recovery device produce 12,000 L of liquid helium per year, which saves about 56736 kWh of electricity and 42552 yuan of electricity costs annually (Table 4). Table 4. Economic benefits and annual net operating income of electric energy saving of helium liquefaction and recovery system The name of the device

Helium liquefaction output: Liter (L)

Annual power saving (kWh)

Annual operating net income (yuan)

Helium liquefaction and recovery system

12000

56736

42552

1) Consider the characteristics of the running time of the system: 24 h of continuous operation on weekdays. 2) Running time: run irregularly, each running time is 70 h. 3) The electricity price is based on the electricity tariff standard in Beijing, with an average value of 0.75 yuan / kW·h during peak hours.

6 Conclusion (1) After the helium liquefaction and recovery system is added with an external purifier, the operation is in good condition, which can fully meet the demand for liquid helium production and has significant energy saving effects. The annual net profit is 42552 yuan. (2) As the demand for liquid helium increases, the system’s liquid helium production efficiency will be greatly provided, and the energy saving potential is higher.

References 1. Zhang Y., Cai, X., Zhang G., Li, M.: Chinese petroleum enterprises, the development and utilization of helium resources in China is imminent (7), 21–25 (2019)

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2. Tao, X., Li, J., Zhao, L.: Earth science, the current situation of helium resources in China and the discovery of the first super large helium rich reserves: Hotan river gas field 44(3) 2019 3. Editor in chief of the fourth Design Institute of the Ministry of Chemical Industry Cryogenic Handbook (Volume II), Beijing: Chemical Industry Press (7) (1979) 4. Yun, W.: Analysis on competition pattern of domestic helium market and countermeasures for long-term safety supply guarantee. Charming China (11) 2021

Analysis of Large-Scale Energy Storage Technology for Renewable Energy Based on Liquid Hydrogen Ming He1

, Han Zhou1,2 , Cui Lv1 , Wenhui Cui1,2 , Jihao Wu1,2 , Meimei Zhang1 , and Linghui Gong1,2(B)

1 State Key Laboratory of Technologies in Space Cryogenic Propellants,

Technical Institute of Physics and Chemistry, CAS, Beijing, People’s Republic of China {heming,lhgong}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing, People’s Republic of China

Abstract. Hydrogen is a secondary energy that can provide energy without greenhouse effect and pollution, and will play an important role in the future energy system dominated by renewable energy. The core of large-scale development of hydrogen energy industry is to realize low-cost and high-efficiency raw material source, storage and transportation. In this paper, the key technologies for the clean and efficient utilization of liquid hydrogen are reviewed, and the cost factors of hydrogen energy production, storage and transportation are discussed. The energy storage efficiency of compressed air energy storage (25 MPa, 300 K), normal temperature and high pressure hydrogen energy storage (25 MPa, 300 K) and liquid hydrogen energy storage (0.1 MPa, 20 K) are compared and analyzed theoretically. The theoretical calculation shows that the storage energy of liquid hydrogen is 1452 kWh/m3 , it is 3.63 times that of normal temperature and high pressure hydrogen and 27 times that of compressed air. The analysis shows that liquid hydrogen can realize high density, large capacity and long cycle storage of renewable energy, and has high economy. Therefore, the development of key technologies and equipment of liquid hydrogen energy storage will have an important impact on the future energy system. Keywords: Liquid hydrogen · Renewable Energy · Energy storage

1 Introduction Energy is the material basis for the survival and development of human society and occupies an important strategic position in the economy. However, the way of energy utilization based on fossil energy has led to serious environmental pollution and ecological damage. In order to get rid of the dependence on fossil energy, the utilization proportion of renewable energy is increasing. Introducing the energy storage module into the wind and solar power generation system to store the surplus power and release it when necessary, which can realize the stable and reliable power output of large-scale wind and solar power plants to the power grid system and improve the power grid’s ability to accept new energy for power generation [1]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 149–155, 2023. https://doi.org/10.1007/978-981-99-6128-3_18

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For large-scale energy storage technology, the pumped storage power station needs to be built in the process of utilization. The geographical conditions are a great obstacle to the construction of the power station, which requires a lot of water resources and geographical differences. For compressed air energy storage, there are the following problems, such as excessive dependence on large gas storage chambers and required fossil fuels, harsh geographical conditions of large-scale systems and low efficiency of small-scale systems. It can be developed and applied on a large scale only when the terrain conditions and construction period are guaranteed [2]. Hydrogen energy storage technology has the advantages of high energy density, long storage time, low maintenance cost and no pollution in the process.It is considered to be the most potential large-scale energy storage technology [3].

2 Liquid Hydrogen Energy Storage The hydrogen storage system faces three very important challenges: mass hydrogen storage density, volume hydrogen storage density and hydrogen storage cost. In order to meet the practical application of hydrogen energy, the International Energy Agency (IEA) has proposed the goal of mass hydrogen storage density greater than 50 gH2 /kg and volume hydrogen storage density greater than 50 gH2 /L. At the same time, the US Department of energy (DOE) has also formulated the performance requirements and objectives of hydrogen storage containers for hydrogen fuel cell vehicles [4]. The ultimate target mass hydrogen storage density is 65 gH2 /kg, which is higher than the hydrogen storage target of IEA. As one of the storage methods of hydrogen, liquid hydrogen storage has many advantages. 2.1 Properties of Hydrogen The hydrogen can be liquefied by cooling to 20 K, and the density can reach 70.85 kg/m3 (1 bar, 20.37 K). Even if the gaseous hydrogen at room temperature is pressurized, its density is much lower than that of liquid hydrogen (as shown in Fig. 1 (a)). In terms of energy density, even if the pressure of normal temperature and high pressure hydrogen increases to 100 MPa, it is far less than the energy density of liquid hydrogen (as shown in Fig. 1 (b)). As a form of energy storage, the higher the density, the smaller the storage space, the less storage materials required, and the more cost saving. 2.2 Comprehensive Cost At present, the main bottleneck restricting the large-scale utilization of hydrogen energy is still the comprehensive utilization cost of hydrogen. As can be seen from Fig. 2, compared with high-pressure hydrogen storage, when the transportation distance is greater than 500 km, the comprehensive cost of liquid hydrogen storage and transportation is lower. Moreover, with the increase of transportation distance, the comprehensive cost gap becomes larger. The development of liquid hydrogen production, storage, transportation, filling and safety technology has become an indispensable strategic core technology for large-scale utilization of hydrogen energy.

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a Hydrogen density[5]

b

151

Energy density of hydroge

Fig. 1. Comparison between different storage technologies.

Fig. 2. Cost calculation of different transportation modes [6].

2.3 Comparison with Compressed Air Energy Storage Compressed air and compressed hydrogen can be regarded as reversible adiabatic compression process, and the power consumption of the compression process is:    κ−1  κ p2 κ RT1 1− ω= κ−1 p1 Assuming that P2 is 25 MPa, P1 is 0.2 MPa, κ is 1.4, T1 is 300 K, Rair is 0.287 kJ/(kg*K), RH2 is 4.16 kJ/(kg*K) and the compression efficiency is 80%. In addition to the expansion work, the energy storage of compressed hydrogen also includes the chemical energy of hydrogen itself. The energy density of hydrogen is 33.5 kWh/kg. The efficiency of fuel cell is 60%. It is assumed that the potential energy of liquid hydrogen can be regarded as the expansion work of isothermal expansion after constant volume pressurization, and the expansion efficiency is 50%. T2 is 20 K. The expansion work is calculated as follows. ω=RT ln

p1 T1 v2 = RT ln = RT ln v1 p2 T2

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The theoretical storage energy of liquid hydrogen is 1452 kWh/m3 . It is 3.63 times that of normal temperature and high pressure hydrogen and 27 times that of compressed air (Table 1). Table 1. Renewable energy storage efficiency. Energy storage form

Pressure (MPa)

Temperature (K)

Energy (kWh/m3 )

Multiple

Potential energy

Chemical energy

Total energy storage

Compressed air energy storage

25

300

54

0

54

1

Compressed hydrogen energy storage

25

300

50.5

350

400.5

7.4

Liquid hydrogen energy storage

0.1

20

35

1417

1452

27

3 Key Technologies of Liquid Hydrogen Energy Storage 3.1 Hydrogen Production Technology by Electrolytic Water There are mainly three kinds: alkaline water electrolysis hydrogen production technology, proton exchange membrane water electrolysis hydrogen production technology and solid oxide water electrolysis hydrogen production technology. The equipment depreciation and electricity cost of the electrolytic cell account for more than 90% of the cost. However, from the perspective of product and operation, the space for cost reduction in the future lies in reducing the electricity price, increasing the working hours of the electrolyzer, reducing the equipment procurement cost, and diluting the depreciation and other fixed expenditure costs. 3.2 Hydrogen Liquefaction Technology Design and Optimization of Liquefaction Process. The ultimate goal of the research and optimization of liquid hydrogen production process is the cycle efficiency. It can be seen intuitively from Fig. 3 [7] that the efficiency of the theoretical process is low and the unit energy consumption is high. Even for the precooling Claude cycle with the highest efficiency, the theoretical process efficiency is still less than 10% and the unit energy conssumption is higher than 30 kWh/kgLH2 . At present, the relative cycle efficiency of

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the hydrogen liquefaction unit in operation is between 20% and 30%, the unit energy consumption is about 10 ~ 15 kWh/kgLH2 . For example, the hydrogen liquefaction of Ingolstadt and Leuna in the figure is 13.6 kWh/kgLH2 and 11.9 kWh/kgLH2 respectively. The cycle efficiency of the conceptual process is higher than 30%, and the unit energy consumption is less than 10 kWh/kg LH2 .

Fig. 3. Comparison of hydrogen liquefaction process efficiencies by assuming that all processes are with uniform feed pressure equal to that of Ingolstadt plant at 21 bars.

Key components of liquid hydrogen production [8]. Compressor. The compression process of hydrogen and helium is a special process, because their density is very small, and the adiabatic index of helium is 1.7, while that of hydrogen is 1.4, which means that under the same compression conditions, the end temperature of helium is higher than that of hydrogen, and the compression power consumption is higher. Necessary intermediate cooling system is needed in the compression process of helium. Turbine Expander. Turbine expander is the core component of helium liquefier and hydrogen liquefier. Among them, the main component is the bearing part, and a stable, reliable and efficient bearing is the key to the turbine expander. Bearings mainly include oil bearing, static gas bearing and dynamic gas bearing. Dynamic gas bearing is theoretically more efficient than the other two. Unfortunately, expanders based on high-power load (MW range) gas bearings are not available at present. At present, only the selection based on oil bearing is discussed for large-scale hydrogen liquefaction (Figs. 4 and 5). Heat Exchanger. At present, the most commonly used heat exchanger in the lowtemperature section is aluminum plate fin heat exchanger. Different channel design and process design have a great impact on the pressure loss and heat exchange efficiency of the whole heat exchange process. Therefore, the design of heat exchanger is very important in the liquefaction process.

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Fig. 4. Typical bearing systems.

Fig. 5. Aluminum plate fin heat exchanger: 1 – core, 2 – header, 3 – nozzle, 4 – width, 5 – height, 6 – length, 7 – passage outlet, 8 – cover sheet, 9 – parting sheet, 10 – heat transfer fins, 11 – distribution fins, 12 – side bar, and 13 – front bar.

3.3 Hydrogen Fuel Cell Fuel cell has no pollution to the environment and has the advantages of high energy conversion efficiency (up to 60%–80%), zero emission and no noise. However, it has high technical complexity and high battery cost. The application of hydrogen in the field of transportation is mainly proton exchange membrane fuel cell technology, which involves many parts and key materials. In the field of power generation, the application technologies include proton exchange membrane fuel cell and solid oxide fuel cell. Solid oxide fuel cell involves battery chips, stacks and other parts.

4 Conclusion and Prospect Hydrogen storage by electrolyzing water through renewable energy such as scenery can greatly improve the security and stability of power system, and there is almost no pollution emission. It is a form of energy storage with broad application prospects. Hydrogen liquefaction is the key link of liquid hydrogen energy storage. Based on the liquefaction process, optimizing the state points of each link of the process and

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optimizing the structure of key components can further improve the liquefaction rate of practical application, reduce liquefaction power consumption and improve economic benefits. The theoretical minimum demand for hydrogen liquefaction is 3.92 kWh (kg LH2 ). According to the calculation that the efficiency of thermal power conversion equipment is 61%, the liquefaction power consumption can achieve 6.5 kWh (kg H2 ). The improvement of electrolysis equipment and fuel cell technology and the reduction of cost will also be the key to promote the rapid development of liquid hydrogen energy storage. Acknowledgments. This work was supported by the National Natural Science Foundation of China (Grant No. 52106034), the Youth Innovation Promotion Association Innovative, Chinese Academy of Science, and the Talent Cultivation Project of Technical Institute of Physics and Chemistry, Chinese Academy of Sciences (Grant No. 2021-LC).

References 1. Zakeri, B., Syri, S.: Electrical energy storage systems: a comparative life cycle cost analysis. Renew. Sustain. Energy Rev. 42, 569–596 (2015). https://doi.org/10.1016/j.rser.2014.10.011 2. Hadjipaschalis, I., Poullikkas, A., Efthimiou, V.: Overview of current and future energy storage technologies for electric power applications. Renew. Sustain. Energy Rev. 13, 1513–1522 (2009). https://doi.org/10.1016/j.rser.2008.09.028 3. Valera-Medina, A., Xiao, H., Owen-Jones, M., David, W.I.F., Bowen, P.J.: Ammonia for power. Prog. Energy Combust. Sci. 69, 63–102 (2018). https://doi.org/10.1016/j.pecs.2018.07.001 4. He, M., et al.: The design and optimization of a cryogenic compressed hydrogen refueling process. Int. J. Hydrogen Energy (2020). https://doi.org/10.1016/j.ijhydene.2020.11.061 5. Aceves, S.M., et al.: High-density automotive hydrogen storage with cryogenic capable pressure vessels. Int. J. Hydrogen Energy 35, 1219–1226 (2010). https://doi.org/10.1016/j.ijhydene. 2009.11.069 6. Cui, L. et al.: Research progress and energy consumption analysis of hydrogen liquefaction technology. Cryogenics, 11–18 (2019). https://doi.org/10.16711/j.1001-7100.2019.07.003 7. Krasae-in, S., Stang, J.H., Neksa, P.: Development of large-scale hydrogen liquefaction processes from 1898 to 2009. Int. J. Hydrogen Energy 35, 4524–4533 (2010). https://doi.org/10. 1016/j.ijhydene.2010.02.109 8. Stolten, P.D.D., Emonts, D.B.: Hydrogen Science and Engineering : Materials, Processes, Systems and Technology (2016)

Performance Investigation of the Cryogenic Packed Bed Regenerator with Different Parameters for Liquid Air Energy Storage Systems Wei Ji1,3 , Luna Guo1,2 , Xiaoyu Fan1,2 , Jiyun Liu3 , Liubiao Chen1,2 , and Junjie Wang1,2(B) 1 Chinese Academy of Sciences Key Laboratory of Cryogenics,

Technical Institute of Physics and Chemistry, Beijing, People’s Republic of China [email protected] 2 University of Chinese Academy of Sciences, Beijing, People’s Republic of China 3 Institute of Optical Physics and Engineering Technology, Qilu Zhongke, Jinan, People’s Republic of China

Abstract. Liquid air energy storage is a large-scale and long-term energy storage technology for achieving the deep consumption of renewable energy, and it is also an important supporting technology to achieve the carbon emission target around the world. Its key component is the cryogenic regenerator. Considering the safety and cost, the solid phase packed bed with basalt rather than the liquid phase regenerator with propane and methanol was employed. Aimed at improving the cold energy storage efficiency and the energy storage cycle efficiency, the thermocline formed in the cold storage process was theoretically investigated. A shorter thermocline is beneficial to improve the cold storage efficiency. The effect of different regenerator parameters such as porosity, inlet temperature and mass flow of air on the thermocline were analyzed and compared. A one dimensional two-phase porous media model was utilized to establish the theoretical mathematical model. The result shows that for a specific packed bed, a smaller porosity, a higher inlet temperature and a lower mass flow for the regenerator inlet gas can lead to a shorter thermocline and obtain a higher cold storage efficiency. Keywords: liquid air energy storage · regenerator · cold storage efficiency

1 Introduction Liquid air energy storage is a large-scale and long-term energy storage technology which has the advantages of clean, low carbon, safety, long service life and no geographical restrictions [1]. Its key component is the cryogenic regenerator, which can store the high-grade cold energy of liquid air and complete the cold energy transfer between the intermittent energy release process and storage process, increasing the air liquefaction rate and the energy storage efficiency of the system. Cold storage efficiency is the key © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 156–163, 2023. https://doi.org/10.1007/978-981-99-6128-3_19

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factor affecting the “electric to electric” conversion efficiency of liquid air energy storage system, and has become one of the key technical bottlenecks restricting the development of liquid air energy storage technology [2]. Considering the safety and cost problem, the solid phase packed bed regenerator with basalt rather than the liquid phase regenerator with propane and methanol is attracting more attention. Although packed bed has been extensively studied in the field of high temperature heat storage [3–6], it still lacks in-depth exploration in the field of low temperature cold storage in liquid air energy storage system. Jin [7] studied the operation characteristics of the cold storage system in the continuous cold storage/release cycle. Huttermann [8] studied the influence of volume specific heat capacity of cold storage media on effective capacity ratio and exergic efficiency of regenerator. At present, only a few institutions have carried out experimental research on low temperature solid phase cold storage system, and there is a significant deviation between the obtained cold storage efficiency and theoretical design. Chai [9] studied the packed bed regenerator using granite stone and liquid nitrogen as the cold storage medium and heat transfer fluid, respectively. In 2015, Highview company reported the first 350 kW/2.5 MWh liquid air energy storage demonstration project in the world, in which quartzite and air were used as the cold storage medium and heat transfer fluid in the cold storage system, respectively, which made a beneficial exploration for the application of solid phase regenerator in practical projects [10]. Thermocline thickness is an important parameter for the solid phase regenerator. A shorter thermocline is beneficial to improve the cold storage efficiency. However, there are few researches considering the thermocline difference of different parameters of inlet air and solid phase medium. Aimed at improving the cold energy storage efficiency and the energy storage cycle efficiency, the thermocline formed in the cold storage process was theoretically investigated. The effect of different regenerator parameters such as porosity, inlet temperature and mass flow of gas on the dynamic development law with time of the thermocline were analyzed and compared.

2 Simulation Model This study focuses on the evolution of the thermocline in the cold energy storage process of the packed bed, and the physical model is shown in Fig. 1(a). The cold air enters the packed bed from the bottom and transfers the cold energy to the solid particles. Then the reheated air flows out from the top of the packed bed. The static solid phase material realizes the cold energy storage process. In order to simulate the heat transfer process of the packed bed, a two-dimensional axisymmetric calculation domain is adopted. Furthermore, a one dimensional two-phase porous media model is utilized to establish the theoretical mathematical model for the solid phase regenerator. For the solid phase medium: (1 − ε)ρs cp,s

  ∂Ts = ∇ · [(1 − ε)ks ∇Ts ] + qv Tf − Ts ∂t

(2.1)

For heat transfer fluid: ερf cp,f

    ∂Tf + ερf cp,f uf · ∇Tf = ∇ · εkf ∇Tf + qv Ts − Tf ∂t

(2.2)

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where ε is the porosity, ρ (kg/m3) is the density, T (K) is the temperature, t (s) is the time, cp (J/kg·K−1 ) is the specific heat capacity, k (W/m·K−1 ) is the thermal conductivity, u (m/s) is the velocity, s and f represent the solid and fluid, respectively, and qv (W/m3·K−1 ) is the volumetric convective heat transfer coefficient. 6(1 − ε) qsf dp

(2.3)

700 G 0.76 dp0.24 6(1 − ε)

(2.4)

qv = qsf =

where d p (m) is the solid particle diameter, qsf (W/m2 ·K−1 ) is the clearance heat transfer coefficient, G (kg/s·m−2 ) is the mass flow per unit cross-sectional area. The boundary conditions and initial conditions during the cold energy storage process are set as follows. The bottom boundary condition (h = 0 m) is the air mass flow inlet with a constant temperature. The top (h = H m) is the reheated air outlet with a working pressure. The wall is adiabatic. Besides, the initial temperature of the packed bed zone is maintained at ambient temperature. Based on the above model and conditions, COMSOL Multiphysics software is used for simulation and calculation. Triangular and rectangular elements with uneven meshes are used to ensure accuracy. And the grid-independence test is carried out for the same computational domain. The default physics-controlled extra fine mesh (with 25276 elements) can give similar results (the distribution of the temperature field and flow field) to the extremely fine mesh (with 117208 elements). Therefore, the extra fine mesh is chosen for the heat transfer analysis of the packed bed. To verify the accuracy and applicability of the above model, the calculated values are compared with the experimental data in literature [11]. The geometric dimensions, thermal properties of air and solid materials, and boundary conditions are the same as the parameters in the literature. Figure 1(b) shows the comparison results of the axial temperature of the packed bed at different time. The maximum relative error is 5.2%, and the average error is less than 4%. Thus, the packed bed theoretical model is acceptable.

(a)

(b)

Fig. 1. (a) Schematic diagram of the packed bed during the cold energy storage process (b) Comparison of simulation and experimental results

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3 Simulation Result Based on the simulation model above, a basic case was carried out to analyze the dynamic development of the thermocline at different time. Table 1 shows the design parameters of the basic case. The basalt is chosen for the cold storage medium for its low thermal diffusivity and low cost. The nitrogen is used to substitute the air. The temperature zone is set as the same as the cold storage temperature of regenerator in LAES, which is 113.15–303.15 K. The thickness-diameter ratio is set as 4:1. The porosity is measured by experiment, which is 0.4. Table 1. Design parameters of basic case Parameter

Unit

Density

kg/m3

2800

Particle size

mm

10

Heat transfer fluid (nitrogen)

Temperature

K

113.15–303.15

Mass flow

kg/s

0.2

Packed bed

Diameter

m

1

Height

m

4

Solid phase medium (basalt)

Porosity

Value

0.4

At the initial time, the packed bed is at room temperature (303.15 K), and the low temperature nitrogen of 113.15 K enters from the bottom of the packed bed (H = 0 m), and then flows out from the top (H = 4 m), and the cold energy is stored in the cold storage medium. Figure 2 (a) and (b) are temperature contours and axial temperature distributions at different times, respectively. With the continuous transfer of cold energy, a complete thermocline begins to form inside the packed bed. Combined with Fig. 3, it can be seen that the thickness of the thermocline increases with time, from 0.91 m at 0.5 h to 2.24 m at 3.5 h. When the cold storage process reaches 3.5 h, the outlet gas temperature at the top decreases from room temperature due to the continuous extension of the thermocline. Considering the combination of packed bed and liquid air energy storage system, the decrease of the outlet gas temperature will cause the cold loss. Therefore, complete cold storage is not adopted in practical application, and the outlet gas temperature in the cold storage process is limited to 303.15K, corresponding to a cold storage time of 3.5 h.

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(a)

(b)

Fig. 2. The temperature distribution of basic case at different times: temperature contours (a), the axial temperature distribution (b)

Fig. 3. The thickness of the whole thermocline of basic case at different times

4 Parameter Sensitivity Analysis The effect of different regenerator parameters such as porosity, inlet temperature and mass flow of the gas on the dynamic development law with time of the thermocline were analyzed and compared in this section. Figure 4 shows the influence of the porosity on the thermocline thickness. All the thermocline thickness increase with the cold storage time, while a higher porosity can lead to a longer thermocline thickness. When the porosity increase, the axial length of packed bed will be larger and the media per unit length will be cooled faster which can make the thermocline thickness increase. When the cold storage process reaches 2.5 h, the thickness of the thermocline corresponding to the porosity of 0.3, 0.4 and 0.5 is 1.80 m, 1.93 m and 2.36 m, respectively. Figure 5 presents the influence of the nitrogen inlet temperature on the thermocline thickness. The thermocline thickness with different inlet temperatures all increase with the cold storage time, and a higher temperature can lead to a shorter thermocline thickness. A high inlet temperature has a low temperature difference for regenerator and can

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Fig. 4. Influence of the porosity on the thermocline thickness

weaken the heat exchange power between the medium and fluid, which can make the thermocline thickness short. Compared with the thermocline thickness with inlet temperature of 113.15 K, the thickness of the thermocline at 133.15 K and 153.15 K at 3.5 h is 0.1 m and 0.19 m smaller, respectively.

Fig. 5. Influence of the nitrogen inlet temperature on the thermocline thickness

Figure 6 shows the influence of the nitrogen mass flow on the thermocline thickness. The thermocline thickness with different mass flow of nitrogen all increase with the increasing time, and a higher mass flow can lead to a longer thermocline thickness. As the nitrogen mass flow increase, the heat transfer between the medium and fluid is enhanced, which can make the thermocline thickness increase.

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Fig. 6. Influence of the nitrogen mass flow on the thermocline thickness

5 Conclusions For the purpose of improving the cold energy storage efficiency and the energy storage cycle efficiency, the effect of different regenerator parameters on the thermocline were analyzed. The thickness of the thermocline increases with time resulting from the dynamic development of the thermocline, which can maintain the complete length for 3.5 h. For the regenerator with basalt as the medium and nitrogen as the fluid, a smaller porosity, a higher inlet temperature and a lower mass flow are beneficial to the cold storage performance. Acknowledgements. This work was funded by the Youth Science and Technology Innovation Project (CRYOQN202108) from the CAS Key Laboratory of Cryogenics, TIPC.

References 1. Borri, E., Tafone, A., Romagnoli, A., Comodi, G.: A review on liquid air energy storage: history, state of the art and recent developments. Renew. Sustain. Energy Rev. 137, 110572 (2021) 2. Guo, L.N., Gao, Z.Z., Ji, W., Xu, H., Chen, L.B., Wang, J.J.: Thermodynamics and economics of different asymmetric cold energy transfer in a liquid air energy storage system. Energy Technol. 8, 1901487 (2020) 3. Zanganeh, G., Pedretti, A., Zavattoni, S., Barbato, M., Steinfeld, A.: Packed-bed thermal storage for concentrated solar power—pilot-scale demonstration and industrial-scale design. Sol. Energy 86, 3084–3098 (2012) 4. Bindra, H., Bueno, P., Morris, J., Shinnar, R.: Thermal analysis and exergy evaluation of packed bed thermal storage systems. Appl. Therm. Eng. 52, 255–263 (2013) 5. Liu, J., et al.: Experimental study on heat storage and transfer characteristics of supercritical air in a rock bed. Int. J. Heat Mass Tran. 77, 883–890 (2014) 6. Chai, L., Wang, L., Liu, J., Yang, L., Chen, H.S., Tan, C.Q.: Performance study of a packed bed in a closed loop thermal energy storage system. Energy 77, 871–879 (2014) 7. Jin, Y., Wang, L., Yang, C.Y., Song, J., Xu, C.: Cycle performance of a packed bed based cold storage device. Energy Storage Sci. Technol. 6(4), 708–718 (2017)

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8. Hüttermann, L., Span, R.: Influence of the heat capacity of the storage material on the efficiency of thermal regenerators in liquid air energy storage systems. Sol. Energy 174, 236–245 (2019) 9. Chai, L., et al.: Cryogenic energy storage characteristics of a packed bed at different pressures. Appl. Therm. Eng. 63, 439–446 (2014) 10. Morgan, R., Nelmes, S., Gibson, E., Brett, G.: Liquid air energy storage-analysis and first results from a pilot scale demonstration plant. Appl. Energy 137, 845–853 (2015) 11. Meier, A., Winkler, C., Wuillemin, D.: Experiment for modelling high temperature rock bed storage. Solar Energy Mater. 24, 255–264 (1991)

Study on Flow Equalization in Solid Phase Packed Bed Regenerator of Liquid Air Energy Storage Systems Xiaoyu Fan1,2 , Wei Ji1,3(B) , Luna Guo1,2 , Jiyun Liu3 , Liubiao Chen1,2 , and Junjie Wang1,2 1 Technical Institute of Physics and Chemistry, Chinese Academy of Sciences Key Laboratory

of Cryogenics, 29 Zhongguancun East Road, Haidian, Beijing, People’s Republic of China [email protected] 2 University of Chinese Academy of Sciences, No.19(A) Yuquan Road, Shijingshan, Beijing, People’s Republic of China 3 Institute of Optical Physics and Engineering Technology, Qilu Zhongke, Licheng, Jinan, People’s Republic of China

Abstract. As a large-scale energy storage technology, liquid air energy storage (LAES) has many advantages such as large energy capacity, simple process and no geographical restrictions. For the regenerator, solid phase cold storage can effectively reduce the cost and has better safety performance compared with liquid phase cold storage, but a major drawback of solid phase cold storage is the low efficiency caused by uneven flow distribution. In order to solve this shortcoming, several effective modified packed bed structures with flow equalizers were proposed, and the cold storage efficiency of the novel packed bed structures were analyzed. The results show that the novel packed bed structure can effectively improve the flow condition in the packed bed, making the temperature distribution more uniform. The research can greatly reduce the construction cost of LAES system and contribute to the application and promotion of LAES system. Keywords: Liquid air energy storage · Solid phase cold storage · Packed bed · Flow distribution

1 Introduction With the rapid development of renewable energy, the proportion of new energy connected to the grid is increasing. According to the statistical results of REN21, 27.1% of world electricity supply comes from renewable energy [1]. However, renewable energy has the characteristics of fluctuation and intermittence, which poses a threat to the safety and stability of grid [2]. As an effective way to integrate renewable energy into the grid, energy storage has attracted wide attention in recent years and has broad application prospects in the future. Battery energy storage (BES), compressed air energy storage (CAES) and pumped energy storage (PES) are currently mature energy storage technologies, but BES has high investment costs and safety problems, and CAES and PES have strict © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 164–171, 2023. https://doi.org/10.1007/978-981-99-6128-3_20

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geographical constraints [3, 4]. Based on the CAES process, a cryogenic liquefaction process is introduced, and the air is stored in a cryogenic tank in the liquid phase, which is called liquid air energy storage (LAES) [5]. As a large-scale energy storage technology, LAES has advantages of low investment, no geographical constraints, high safety, and has a good promotion prospect [6]. Cold storage unit is a critical component of LAES, and its cold storage efficiency (CSE) has a great impact on the round trip efficiency (RTE) of LAES. Highview Power first built a 350kW/2.5MW pilot system in 2011, in which the CES was 51%, but the RTE of overall system only reached 8% [7]. Based on the experimental results in reference [7], Morgan et al. [8] improved the CES of the cold storage unit from 51% to 91%, and greatly increased the RTE of the system from 8% to 47%. There are two types of cold storage, solid phase cold storage and liquid phase cold storage, in which solid phase cold storage has low investment and no inflammable or explosive risk [9]. Solid phase regenerator stores cold energy in the packed bed with rock or other medium which is conducive to large-scale promotion. However, the RTE of LAES with solid phase cold storage is relatively low, mainly because of the low-grade of cold stored in the packed bed. Lin et al. [10] conducted sensitivity analysis of the solid phase packed bed cold storage in the LAES system. Tafone et al. [11] analyzed the influence of key parameters in the packed bed on RTE of LAES by establishing a steady-state model, and summarized the optimization direction of parameters. Liao et al. [12] studied the dynamic characteristics of the packed bed, and analyzed the effects of air flow rate and working pressure on the heat transfer process. At present, most studies on solid phase cold storage are sensitivity analysis and its integrating with LAES. However, the CES of solid phase cold storage is low, and one of the main reasons is that the temperature distribution in the packed bed is uneven, resulting in the reduction of cold energy grade, which has been seldom studied before. In this paper, the causes of uneven temperature distribution were analyzed, and two methods were proposed to improve the temperature uniformity in the packed bed.

2 The Simulation Model 2.1 Basic Parameters and Boundary Conditions Figure 1 is a schematic diagram of a simulated packed bed which is a two-dimensional axisymmetric model. A one-dimensional two-phase model was used to simulate the packed bed, in which the length of the packed bed was 800 mm and the radius was 100 mm. COMSOL Multiphysics is used as simulation software and the finite element method is adopted. The hybrid mesh of triangle and quadrilateral is adopted, in which the quadrilateral mesh is used for the boundary layer. Skewness is used as the index of the grid test. When the average element quality is greater than 0.8, the grid is available. The change of the result within 5% is adopted as the standard of the grid independence test. In calculation, the time step is set to 60 s. Due to its low thermal conductivity, high heat capacity and low cost of investment, basalt was selected as cold storage medium. The thermophysical properties of basalt are shown in Table 1. In the process of cold storage, cryogenic nitrogen enters from the bottom of the packed bed and conducts convection heat transfer with the basalt in the packed bed. Nitrogen is the heat transfer fluid and

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its thermophysical properties are derived from the properties package in COMSOL. In the process of heat transfer, a temperature gradient from low temperature to normal temperature will be formed in the packed bed, which is called the thermocline, because the cryogenic nitrogen cannot transfer the cold energy totally to basalt per unit volume in unit time. The cold storage process finishes when the outlet temperature of the packed bed begins to drop. The boundary condition of the packed bed are shown in Table 2.

Fig. 1. Schematic diagram of the simulated packed bed. Table 1. Thermophysical properties of basalt. Parameters

Unit

Value

Density

kg/m3

2800

Heat conductivity coefficient

W/(m·K)

2.68

Specific heat capacity

J/(kg·K)

650

Porosity

0.4

Diameter

m

0.005

Table 2. Boundary condition of the packed bed. Parameters

Unit

Value

Ambient temperature

K

298.15

Working pressure

bar

1

Inlet temperature

K

288.15

Outlet temperature

K

138.15

Mass flow rate

kg/s

0.018

Nature convection heat transfer coefficient of the wall (hw )

W/(m2 ·K)

10

Thermal conductivity of insulation layer

W/(m·K)

0.025

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2.2 Assumption For the correct but simple simulation of flow and heat transfer in the packed bed, the following reasonable assumptions were made: a) The inlet cryogenic nitrogen is uniform. b) The physical properties of basalt are constant in the working temperature zone. 2.3 Thermodynamic Model Local thermal nonequilibrium metaphysical coupling was used to consider heat transfer in porous regions where solid and fluid temperatures are unbalanced. This is achieved by coupling the heat equations of the solid and fluid subdomains with transfer terms proportional to the temperature difference between the fluid and the solid. The corresponding heat equations in the solid and fluid subdomains are as follows: θp ρs Cp,s

    ∂Ts + θp ρs Cp,s us · ∇Ts = ∇ · θp ks ∇Ts + qsf Tf − Ts ∂t     ∂Tf 1 − θ p ρf Cp,f  ∂t + 1 − θp ρf Cp,f uf · ∇Tf = ∇ · 1 − θp kf ∇Tf + qsf Ts − Tf

(1) (2)

where θp is the solid volume fraction, ρs and ρf are the solid and fluid densities, Cp,s and Cp,f are the solid and fluid heat capacities at constant pressure, ks and kf are the solid and fluid thermal conductivities, qsf is the interstitial convective heat transfer coefficient, us and uf are the solid and fluid velocity vectors. Radial temperature distribution in the packed bed and cold storage time were selected as evaluation indexes to measure temperature uniformity and CES. The temperature uniformity of a certain layer in the packed bed can be intuitively observed through the radial temperature distribution. In this paper, the radial temperature distribution at the axial height of 0.4 m or 0.8 m is taken as the performance index, because the radial temperature distribution at the axial height of 0.4 m can effectively represent the temperature uniformity in the process of cold storage and the radial temperature distribution at the axial height of 0.8 m can better represent the temperature uniformity at the end of cold storage. The radial maximum temperature difference (TR ) was used as the criterion to measure the uniformity. In addition, under the same thermal insulation conditions, the time of cold storage is a key index to measure temperature uniformity and cold storage efficiency. The longer the cold storage time, the more cold energy is stored in the packed bed, which indicates that the temperature in the packed bed is more uniform and the CSE is higher.

3 Results and Discussion 3.1 Influence of Heat Leakage on Temperature Uniformity of Packed Bed Figure 2 shows the radial temperature distribution at the axial height of 0.4 m in the packed bed under different thermal insulation conditions when the cold storage process is halfway through. With the enhancement of thermal insulation, the radial temperature

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of the packed bed is more uniform and TR is smaller. TR of the packed bed without any thermal insulation measures is the largest, which is 54.85K, while TR of the packed bed with complete thermal insulation is the smallest, which is 35.73K. Therefore, heat leakage is one of the factors of uneven temperature distribution of packed bed, and the thermal insulation performance of packed bed is an important factor of temperature uniformity of packed bed. 280

No insulation measures 10mm thick insulation layer 20mm thick insulation layer 50mm thick insulation layer Fully insulated

Temperature(K)

260

240

220

200

180

-0.10

-0.05

0.00

0.05

0.10

Radial Distance(m)

Fig. 2. Radial temperature distribution under different insulation measures.

3.2 Influence of Guide Plate on Temperature Uniformity of Packed Bed Although thermal insulation enhances the temperature uniformity, there is still a large TR in the packed bed under the condition of fully insulation which is mainly due to the influence of uneven air flow in the packed bed. In order to improve the flow uniformity in the packed bed, the guide plate was used to disturb the flow, so that the temperature distribution in the packed bed is uniform. The following kinds of structural design of guide plates were proposed and their sizes were optimized. a) A circular plate with the diameter of 50 mm. b) A round ball with the diameter of 62 mm. c) A circular plate removes a ring, and its structure is shown in Fig. 3, where D1 = 64 mm, D2 = 120 mm and D3 = 160 mm. In order to verify the performance of the above diversion structures, the radial temperature distribution at the axial height of 0.8 m of the packed bed was calculated along with the cold storage process, as shown in Fig. 4. As shown in Fig. 4 (a), TR of the packed bed without a guide plate is the largest, which is 37.52 K. TR of the packed bed with a circular plate and a round ball are 33.25 K and 29.61 K as shown in Fig. 4 (b) and (c). The packed bed with a circular plate removes a ring as the guide plate has the most uniform radial temperature distribution, and TR is only 3.21 K, as shown in Fig. 4 (d). It shows that the radial temperature uniformity is improved with the use of guide plates. In addition, although the packed bed with a circular plate and a round ball still has a large TR , it can be seen from the figure that this is mainly because of the uneven distribution of temperature on both sides of the packed bed, and the temperature uniformity in the middle has been significantly improved.

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Fig. 3. Schematic diagram of a circular plate removes a ring. 300

300

270

0s 300s 600s 900s 1200s 1500s 1800s 2100s 2400s

240

210

180

Temperature(K)

Temperature(K)

270

0s 300s 600s 900s 1200s 1500s 1800s 2100s 2400s

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150

150

-0.10

-0.05

0.00

0.05

0.10

-0.10

-0.05

Radial Distance(m) (a) 300

0.05

0.10

300

270

0s 300s 600s 900s 1200s 1500s 1800s 2100s 2400s

240

210

180

150

Temperature(K)

270

Temperature(K)

0.00

Radial Distance(m) (b)

0s 300s 600s 900s 1200s 1500s 1800s 2100s 2400s

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-0.10

-0.05

0.00

Radial Distance(m) (c)

0.05

0.10

-0.10

-0.05

0.00

0.05

0.10

Radial Distance(m) (d)

Fig. 4. Radial temperature distribution under different guide plates.

In order to evaluate the CSE of the packed bed with the use of guide plates, the change of outlet temperature of the packed bed under different guide plates is shown in Fig. 5. The packed bed with a circular plate removes a ring has the longest cold storage time, but the improvement of cold storage efficiency is relatively limited. The cold storage time of the packed bed with a circular plate and a round ball is shorter than that of the packed bed without guide plate. Due to the turbulence of the guide plate, the flow velocity increases in some areas of the packed bed, and the thermocline increases, resulting in the decrease of the CSE.

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Without a guide plate Plectane Ball Circular ring

Temperature(K)

270 240 210 180 150 120

0

600

1200

1800

2400

3000

Times(s)

Fig. 5. Outlet temperature changes under different guide plates.

4 Conclusion In this paper, the phenomenon of uneven temperature distribution of solid phase cold storage in LAES was studied, the reasons of uneven temperature distribution were analyzed and two methods were proposed to improve it. The main conclusions are as follows: 1) The uneven temperature distribution in packed bed is due to heat leakage and uneven distribution of flow rate. 2) The effects of different guide plates and insulation measures on temperature uniformity of packed bed were analyzed. Strengthening insulation measures is the effective methods to improve the temperature uniformity in packed bed. However, the use of guide plates has limited effect on improving the efficiency of cold storage. Acknowledgments. This work was funded by the Youth Science and Technology Innovation Project (CRYOQN202108) from the CAS Key Laboratory of Cryogenics, TIPC.

References 1. Renewables 2021 Global Status Report. https://www.ren21.net/gsr-2021/. Accessed 28 May 2022 2. Yao, L.: Challenges and progresses of energy storage technology and its application in power systems. J. Mod. Power Syst. Clean Energy 4(4), 519–528 (2016) 3. Díaz-González, F.: A review of energy storage technologies for wind power applications. Renew. Sust. Energ Rev. 16, 2154–2171 (2012) 4. Ibrahim, H.: Energy storage systems-Characteristics and comparisons. Renew. Sustain. Energy Rev. 12, 1221–1250 (2008) 5. Borri, E.: A review on liquid air energy storage: history, state of the art and recent developments. Renew Sustain. Energy Rev. 137, 110572 (2021) 6. Ji, W.: Thermodynamic analysis of a novel hybrid wind-solar-compressed air energy storage system. Energy Convers. Manage. 142, 176–187 (2017) 7. Morgan, R.: Liquid air energy storage: analysis and first results from a pilot scale demonstration plant. Appl. Energy 137, 845–853 (2015)

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8. Morgan, R.: An analysis of a large-scale liquid air energy storage system. Energy 168(2), 135–144 (2015) 9. Legrand, M.: Integration of liquid air energy storage into the Spanish power grid. Energy 187, 115965 (2019) 10. Lin, X.: Thermodynamic analysis of the cascaded packed bed cryogenic storage based supercritical air energy storage system. Energy Procedia 158, 5079–5085 (2019) 11. Tafone, A.: New parametric performance maps for a novel sizing and selection methodology of a liquid air energy storage system. Appl. Energy 250, 1641–1656 (2019) 12. Liao, Z.: Investigation of a packed bed cold thermal storage in supercritical compressed air energy storage systems. Appl. Energy 269, 115132 (2020)

Preparation and Preliminary Test on the Integrated Test of the Accelerator Cryoplant and the Cryogenic Distribution System at ESS J. Zhang(B) , P. Arnold, J. Fydrych, and J. G. Weisend II European Spallation Source (ESS) ERIC, Partikelgatan 2, Lund, Sweden [email protected]

Abstract. At the European Spallation Source (ESS), a 2.0 GeV proton Linear Accelerator (LINAC) is under construction, which contains 43 cryomodules operating at 2 K. The Accelerator Cryoplant (ACCP) will provide helium cooling to the cryomodules in the LINAC at different temperature levels. The cryogenic distribution system (CDS) connects the cryomodules and ACCP by means of valve boxes and cryogenic multi-transfer lines. The ACCP successfully passed the last acceptance tests in September, 2020. In order to test the performance of CDS, an integrated test of ACCP and CDS without cryomodules connected will be performed in 2022. This paper describes two different series of test cases developed for the ACCPCDS assembly, one to be conducted at 8K feed temperature and over-atmospheric pressure to determine the CDS heat load and the other at 2 K equivalent subatmospheric conditions with the cold compressors (CCs) in operation to check their performance with the CDS attached. The new developed control logic, limitations and the interface parameters will be discussed. Keywords: Accelerator Cryoplant (ACCP) · cryogenic distribution system(CDS) heat load · cold compressors (CCs)

1 Introduction The Accelerator Cryoplant (ACCP) will provide helium cooling to the 43 Cryomodules in the LINAC at different temperature levels, in which the heat loads include 2 K heat load, 4.5 K coupler cooling heat load and Thermal Shield (TS) heat load [1, 2]. The cryogenic distribution system (CDS) connects the Cryomodules and ACCP by means of 43 valve boxes and ~ 400 m long cryogenic multi-transfer lines [3]. The ACCP successfully passed the last acceptance tests in July, 2020. In order to test the performance of CDS and the cold compressors (CCs), an integrated test of ACCP and CDS without Cryomodules connected will be performed in Q3, 2022. For each valve box, the 4.5K circuit and the thermal shield (TS) circuit are short looped within an end cap, shown in Fig. 1. In Fig. 1, the left figure shows the simplified © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 172–178, 2023. https://doi.org/10.1007/978-981-99-6128-3_21

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PI&D of one Valve Box with one End cap and the right picture shows the corresponding structure in the tunnel. The CDS heat load test and the performance of CCs will be performed. ACCP was implemented with two new working modes for the ACCP-CDS integrated tests by the supplier, Linde. In this paper, the new developed control logic, the test method and the latest preparation test results will be presented.

End cap

Valve Box

Fig. 1. One Elliptical valve box with an end cap

2 Test Method In order to avoid the oscillation caused by two phase transition and in order to measure sensible heat, all the tests will be carried out in helium gas region. The main considerations for the ACCP-CDS tests include: • CDS heat load test: CDS heat load will be tested in a series of tests with the mass flow range from 10 g/s to 55 g/s and then the real heat loads will be determined by the linear regression plot. • CCs test: The pump down process of CCs and the operation performance of CCs in different mass flows will be performed, first time with a significant volume. 2.1 The Selection of Parameters The helium properties of enthalpy and the specific heat capacity (Cp) are shown in Fig. 2 [4], which come from Hepak. In order to avoid two phase helium in CDS during the tests, the 4.5 K supply temperature shall be no less than 5.9 K, so the supply temperature for

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CCs test is chosen 5.9 K. In Fig. 2 b), the specific heat capacity (Cp) decreases with the temperature increase, the supply temperature is chosen at 8 K for CDS heat load test so as to get higher temperature difference with the same heat load. The design parameters for CCs test and CDS heat load are listed in Table 1.

a)

b)

Fig. 2. The helium properties, (1) enthalpy vs. temperature (2 K ~ 6 K) and (2) Cp vs. temperature (5.5 K ~ 10 K)

Table 1. The design parameters for ACCP-CDS tests Items

CCs test CDS heat load test

4.5 K supply pressure, bara

3.00

3.00

4.5 K supply temperature, K 5.90

8.00

SP return Pressure, bara

0.027

1.30

SP return temperature, K

4.38

9.5 ~ 13

Total mass flow, g/s

93

10 ~ 50

Non-Isothermal heat load, W 627

627

Total TS heat load, W

5472

5472

TS mass flow, g/s

51

51

2.2 CDS Heat Load Test Acceptance Criteria The heat load is calculated by the difference of enthalpy with the following equation: Qm = m ∗ (Ho − Hi )

(1)

where Qm is the average measured heat load in a period of time, m is the mass flow, Hi is the inlet enthalpy and Ho is the outlet enthalpy. As we all know that the measurements have systematic errors, especially the temperature sensors which are affected strongly by the installation method and installation

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process. Since the enthalpy is mainly dependent on temperature, so the measurement error of temperature sensor is considered. Two assumptions are proposed: • As the specific heat capacity of helium has a little change when T > 8 K, it is supposed to be constant in the temperature range from 8 K to 13 K. • For a given cryogenic transfer line, the real heat load is irrelevant to the mass flow but the work temperature. Therefore, we assume the real heat load is constant in the tests with different working flows while the supply temperature is constant. Then the heat load can be expressed by the following equation: Qr Qm = + Cp ∗ T (2) m m where Qr is the real heat load, Cp is the constant specific heat capacity, T is the systematic error caused by the measurements of inlet temperature T i and outlet temperature T o. The real heat load Qr will be determined by the linear fitting correlation between 1/m and Qm /m based on the Eqs. (1) and (2). T can be gotten by the intercept of the linear fitting correlation. 2.3 CDS Heat Load Test Results in the Test Stand The cryogenic transfer line between the Test & Instrument Cryoplant (TICP) and the cryomodule test stand is 50 m long, which includes 4.5 K supply line, VLP return line, TS supply line and TS return line [4]. In May, 2020, a series of tests were carried out for the 4.5 K CDS circuit with the mass flow ranging from 1.16 g/s to 4.62 g/s and the supply temperature of 4.5 K supply line changing from 6.89 K to 9.53 K [5]. With each mass flow, the average measured heat load for different transfer lines could be achieved based on Eq. (1) with a time period for more than 6 h. Then the real heat loads were correlated according to Eq. (2). The fitting line for 4.5 K supply line real heat load is shown in Fig. 3. In Fig. 3, the 4.5 K supply line real heat load is 5.25 W with a high correlation coefficient of ~ 0.99.

3 New Control Logic Compared with the normal operation modes during ACCP acceptance tests, the ACCPCDS tests are more challenging and need new controllers so as to supply stable gas helium to CDS. The comparison between ACCP acceptance tests and ACCP-CDS tests are listed in Table 2. 3.1 New Controllers Two new working modes for ACCP-CDS tests have been implemented in PLC, i.e. CCs test and CDS heat load test. There are two new controllers, shown in Fig. 4, one new temperature controller is added (TC1) which is deployed on Turbine 4 (T4) inlet valve CV1, one new pressure controller (PC1_2) is added which is deployed on JT turbine (T6) bypass valve CV2 in CDS heat load test and one already existing pressure controller (PC1_1) deployed on sub-cooler inlet valve CV3 in CCs test.

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Qm/m = 5.25*(1/m) - 1.26 R² = 0.99

Fig. 3. The correlation between 1/m and Qm /m for the 4.5 K supply line in the test stand CDS Table 2. The comparison between ACCP acceptance tests and ACCP-CDS tests No

ACCP acceptance tests

ACCP-CDS tests

1#

Helium in liquid and vapor phase

Only gas phase

2#

Temperature stable with 4.5K sub-cooler (600 l) and 20,000 l LHe Dewar

New temperature controller with different principle

3#

JT turbine is in operation

JT turbine doesn’t work in CDS heat load test

4#

Load calibration and stabilization with LHe Dewar

New pressure controllers

5#

Mass flow are decided by heaters

Different mass flows are set in CDS heat load tests by control valves

3.2 New Control Logic Test Results with Heaters In September, 2021, the new control logic was tested with ACCP and the heaters. All the new controllers worked well [7]. In Table. 3, for CDS heat load test, the 4.5 K supply mass flow F1 ranged from 11 g/s to 55 g/s with the reading error less than 5%, and the 4.5 K supply temperature T1 is ~ 8 K with a reading error less than 1.1%. The 4.5 K supply pressure P1 is stabilized at 3 bara with a fluctuation of 0.03 bara. Therefore, ACCP could supply the required stable flow to test CDS heat load. With the 4.5 K supply temperature of 6.2 K, the CCs test was carried out successfully, and the stable operation process is shown in Fig. 5. In Fig. 5, the CCs mass flow is 94.5 g/s with a fluctuation of 0.5 g/s, and the CC1 suction pressure is 26.7 mbar with a fluctuation of 0.3 mbar.

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Fig. 4. The simplified ACCP flow diagram with new controllers (heat exchangers: HX1 ~ HX7; Turbines: T1 ~ T6; different pressure stages: HP, MP, LP and SP [6])

Table. 3 The reading errors of 4.5 K supply mass flow F1 and 4.5 K supply temperature T1 in CDS heat load test mode with ACCP only No

4.5 K supply mass flow F1

4.5 K supply temperature T1

Value, g/s

Reading error, %

Value, K

Reading error, %

1#

55.0

± 0.7%

7.99

± 0.4%

2#

41.1

± 0.2%

8.00

± 0.3%

3#

28.5

± 0.7%

8.01

± 0.5%

4#

20.9

± 5.0%

8.01

± 0.6%

5#

11.2

± 4.0%

8.03

± 1.1%

CCs mass flow: (94.5 ± 0.5) g/s CC1 suction temperature: (4.98 ± 0.07) K

CC1 suction pressure: (26.7 ± 0.3) mbar

Fig. 5. The CCs stable operation in CCs test mode with ACCP only CCs mass flow: (94.5 ± 0.5) g/s

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4 Conclusion and Outlook In ACCP, two new working modes with two new controllers were implemented to test the CDS heat load and the performance of CCs with gas helium. The new control logic was tested successfully with ACCP only in September, 2021, which proved that the CDS heat load test and CCs test were feasible in the expected conditions steadily. In our schedule, ACCP and CDS inter-connection will be proceeded in May, 2022. We will have all the control valves seat leak test and the commissioning of the CDS control system. Then the ACCP-CDS tests will be carried out. Compared with the previous tests with ACCP only, the challenges for ACCP-CDS test include: 1) keeping the cool-down process of the CDS and the valves box evenly; 2) testing the CDS heat load with small mass flow, such as 10 g/s and 3) pumping down the CDS vapor low pressure line steadily without CCs trip.

References 1. Wang, X., Weisend, J., Koettig, T., Hees, W., Darve, C.: ESS accelerator cryogenic plant. HVAC&R Res. 20(3), 296–301 (2014) 2. Fydrych, J., Arnold, P., Hees, W., Tereszkowski, P., Wang, X.L., Weisend, J.G., II.: Cryogenic distribution system for the ESS superconducting proton linac. Phys. Procedia 67, 828–833 (2015) 3. Fydrych, J.: ESS CDS-LTS2 Heat loads calculations, ESS technical document (2017) 4. Zhang, J.: The study of CDS-LTS2 heat loads and 2K performance, ESS technical document (2020) 5. Wang, X. L., Arnold, P., Hees, W., Hildenbeutel, J., Weisend, J.G.: ESS accelerator cryoplant process design. In: IOP Conference Series: Materials Science and Engineering, vol. 101(1), p. 012012. (November 2015) 6. Zhang, J.: Test results for ACCP-CDS test with ACCP only, ESS technical document (2021)

Development of a 10 TPD Hydrogen-Refrigerated Hydrogen Liquefier Jing Li1(B) , Gang Zhou1 , Liqiang Liu1,2 , Kun Yang3 , Lianyou Xiong1,2 , Jihao Wu1,2 , Hailing Qin1 , and Linghui Gong1,2 1 State Key Laboratory of Technologies in Space Cryogenic Propellants, Technical Institute of

Physics and Chemistry, Chinese Academy of Sciences, Beijing 100190, China [email protected] 2 University of Chinese Academy of Science, Beijing 100049, China 3 Beijing Sinoscience Fullcryo Technology Co., Ltd., 1407 Satellite Building, No. 63 Zhichun Road, Beijing 100190, China https://orcid.org/0000-0001-5038-5410

Abstract. In order to develop a 10 TPD (Tons per Day) hydrogen-refrigerated hydrogen liquefier in China, a Process Flow Diagram of this hydrogen liquefier has been proposed. This 10 TPD hydrogen liquefier is based on Claude cycle with a Liquid Nitrogen precooling system. To reach the stream balance and heat balance between hydrogen refrigeration circuit and hydrogen feed gas flow, a thermosiphon hydrogen sub-cooler has been proposed. The hydrogen safety vent lines have been designed according to China national standards for safety. Moreover, the regeneration process of 80 K adsorbers have been introduced. The control logic and the cold box of this liquefier are also presented. Keywords: Liquid Hydrogen · Hydrogen Liquefier · Process Flow Diagram · Control Logic

1 Introduction Hydrogen energy source as a clean and renewable energy resource will be further developed and utilized in the future. Hydrogen energy can balance energy supply and demand, protect the environment and ensure energy security and economic viability [1]. Liquid hydrogen is an important source to get high purity hydrogen and ultra-pure hydrogen. Liquid hydrogen can also be used as refrigerant for high energy physics. In addition, liquid hydrogen has unique characteristics such as lower weight and volume and higher energy content than the gaseous hydrogen. Using a hydrogen liquefier to liquefy hydrogen feed gas is a traditional method to get liquid hydrogen [2]. Most of the existing hydrogen liquefiers in China are small-sized helium-refrigerated hydrogen liquefiers. A 5 TPD (Tons per Day) hydrogen-refrigerated hydrogen liquefier is being established under the support of the fund of the National Key Research and Development Program of Ministry of Science and Technology of the People’s Republic of China. That 5 TPD hydrogen liquefier is designed by Technical Institute of Physics © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 179–186, 2023. https://doi.org/10.1007/978-981-99-6128-3_22

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and Chemistry of CAS and is being manufactured and will be commissioned by Beijing Sinoscience Fullcryo Technology Co., Ltd. Next, a 10 TPD (Tons per Day) hydrogenrefrigerated hydrogen liquefier is under way, which is also designed by Technical Institute of Physics and Chemistry of CAS and is being manufactured and will be commissioned by Beijing Sinoscience Fullcryo Technology Co., Ltd. The main differences between that 5 TPD and 10 TPD hydrogen liquefiers are: First, the different funding agencies, 5 TPD hydrogen liquefier is supported by the National Key Research and Development Program of Ministry of Science and Technology of the People’s Republic of China, 10 TPD hydrogen liquefier is supported by a company. Second, 10 TPD has a larger capacity. The main process flows of these two hydrogen liquefiers are similar. This 10 TPD hydrogen liquefier is based on Claude cycle with a Liquid Nitrogen precooling system and with four hydrogen turbine expanders working in series. A process flow diagram has been presented in this paper. To reach the stream balance and the heat balance between hydrogen refrigeration circuit and hydrogen feed gas flow, a thermosiphon hydrogen sub-cooler has been proposed. The hydrogen safety vent lines have been designed according to China national standards for safety. Control logic and the cold box of this liquefier are also introduced. This paper is organized as follows: After this introduction, Sect. 2 introduces process description of this 10 TPD hydrogen liquefier, including a process flow diagram and a T-S diagram parameters of four hydrogen turbine expanders and four ortho-para convertors are given. The regeneration process of 80 K adsorbers are introduced. Section 3 introduces control logic and control strategy of this 10 TPD hydrogen liquefier. Section 4 introduces cold box. Section 5 makes a conclusion and outlook.

2 Process Description of This 10 TPD Hydrogen Liquefier 2.1 Process Flow Diagram The liquefaction process, which is designed for a capacity of 10 TPD (Tons per Day), is based on a Claude cycle. The necessary refrigeration is provided at three temperature levels using liquid Nitrogen (80 K), two-stage hydrogen turbines arranged in series (80– 30 K) and Joule-Thomson expansion valves (30–20 K). Figure 1 shows the process flow diagram of this 10 TPD hydrogen liquefier. LN2 is fed to the thermosiphon liquid Nitrogen precooler in the cold box at a rate of approx. 1349.64 g/s and a pressure of 3 bar. After heat exchange with the high pressure hydrogen gas, the GN2 is vented to ambient. A summary of the designed technical data of this 10 TPD hydrogen liquefier is given in Table 1. As can be seen from Fig. 1, thermosiphon liquid Nitrogen precooling system has been designed. Thermosiphon structure has been proposed to make use of the density differences of Liquid Nitrogen. Latent heat and sensible heat of Liquid Nitrogen has been used sufficiently. Table 2 shows designed parameters of four hydrogen turbine expanders. Figure 2 shows the T-S diagram of this 10 TPD hydrogen liquefier. From Fig. 2, we can see the parameters of hydrogen refrigeration circuit and hydrogen feed gas flow as well as liquid Nitrogen precooling line.

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Table 1. Designed technical data of this 10 TPD hydrogen liquefier. Item

Parameters

Data

Feed gas specification

Pressure

25 bar

Temperature

< 310 K

Para-content

75%

Pressure

1.5 bar

Temperature

21.48 K

Mass flow rate

150 g/s

LH2 product specification

Para-content

> 95%

LN2 cooling

Mass flow rate

1349.64 g/s

Medium pressure compressor

Pressure ratio

1.05–3.9 bara

Electric power

280.7 kW

High pressure compressor

Pressure ratio

3.9–25 bara

Electric power

3219 kW

Hydrogen liquefier

Capacity

> 10 TPD

Specific energy consumption

10.78 KW·h/Kg LH

Fig. 1. The Process flow diagram of this 10 TPD hydrogen liquefier

To reach stream and heat balance between hydrogen refrigeration circuit and hydrogen feed gas flow, a thermosiphon hydrogen sub-cooler has been proposed. Two-phase

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Inlet temperature (K)

Inlet pressure (bar)

Mass flow rate (g/s)

E11

62.5

24.52

740.0

E12

56.27

18.0

740.0

E21

46.1

12.98

740.0

E22

39.25

8.0

740.0

Fig. 2. The T-S diagram of this 10 TPD hydrogen liquefier

gas-liquid flow after J-T valve CV10 has been subcooled to supercooled liquid hydrogen through heat exchanger HEX-OP4. An isothermal ortho-para conversion has been

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conducted through ortho-para convertor OP4. Then the liquid hydrogen product with Para-hydrogen concentration (>95%) is stored in a vacuum-insulated Dewar D4200. Regeneration gas vent line and Hydrogen gas vent line have been designed separately according to China national standards for safety. Regeneration gas (containing Nitrogen gas) and Hydrogen vent gas have been vented respectively. 2.2 Four Ortho-Para Convertors Total rotational energy of the two spin isomers of molecular hydrogen is not same and the ortho-hydrogen structure is an excited state with a higher energy level than parahydrogen. Thus ortho-hydrogen will spontaneously convert to para-hydrogen to achieve energy balance as the temperature decreases step by step [3]. Four catalytic ortho-para convertors are designed in the process before the hydrogen is liquefied. These four orthopara convertors are applied in different temperatures. Table 3 shows designed parameters of these four ortho-para convertors. As can be seen from process flow diagram of Fig. 1, OP1 and OP4 exchange the orthopara conversion heat at constant temperature with the LN2 and LH2 baths, respectively. OP2 and OP3 are both work adiabatically. Table 3. Designed parameters of four ortho-para convertors. Convertor

Type

Temperature (K)

Para-hydrogen concentration (%)

OP1

Isothermal

78.54

47.91

OP2

Adiabatical

51.7

64.62

OP3

Adiabatical

34.3

89.02

OP4

Isothermal

21.95

96.32

2.3 Regeneration Process of 80 K Adsorbers In the liquefaction flow, the hydrogen feed gas is purified by means of low-temperature adsorption. There are two 80 K adsorbers to purify hydrogen feed gas. These two 80 K adsorbers work in parallel. When one 80 K adsorber works, the other one can be regenerated simultaneously. Figure 3 shows regeneration process flow diagram of one 80 K adsorber. Figure 1 shows, hot Nitrogen gas has been used to regenerate the 80 K adsorbers. High pressure and room temperature H2 feed gas has been used to replace the regenerated 80 K adsorbers. After regeneration and replacement, the 80 K adsorbers have been cooled down to 80 K temperature level before work again.

3 Control Logic and Control Strategy Control logic and control strategies have been proposed to make hydrogen liquefier work efficiently. Control logic of compressor station, turbine strings and liquid hydrogen Dewar have been introduced.

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Working

Dirty

Regeneration

Cool down

Ready

Fig. 3. Regeneration process flow diagram of one 80 K adsorber

Figure 4 shows control logic of compressor station. CH indicates high pressure oilfree reciprocating compressor. CL represents low pressure oil-free reciprocating compressor. This control logic is common and widely used. CV3 and CV4 control high pressure Ph. CV1 controls medium pressure Pm. CV2 and CV21 control low pressure PL. CV2 is the primary control valve and CV21 is the auxiliary control valve.

Fig. 4. Control logic of compressor station

Fig. 5. Control logic of turbine strings and liquid hydrogen Dewar

Figure 5 shows the control logic of turbine strings and liquid hydrogen Dewar. Rotary speed of turbine E11, E12, E21 and E22 is S1, S2, S3 and S4. Outlet temperature of E22 is T2. CV7 and CV11 control the S1, S2, S3 S4 and T2. CV7 is the primary control valve and CV11 is the auxiliary control valve. If S1, S2, S3 and S4 is too high, turn

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down the CV7. Conversely, turn up the CV7. The opening of CV7 should be always conservative. During the cooling down process, if S3 and S4 is too high, CV11 will be put into regulation. To precool the liquid hydrogen dewar D4200, CV9, CV13 and CV14 have been used. From beginning of cooling down, the J-T valve CV9 is used to precool the liquid hydrogen dewar D4200. When the temperature of D4200 T3 reaches to a value, for example, 50 K, J-T valve CV9 is turned off gradually. If T3 is still too high (50 K > T3 > 30 K), CV13 will be turned on. If T3 is a little bit low (30 K > T3 > 23 K), CV14 will be turned on. After precooling of the liquid hydrogen dewar D4200, both CV13 and CV14 will be turned off. The following, J-T valve CV8 will be turned on, hydrogen liquefier will be in the liquefaction mode. Moreover, more than 20 interlocks have been designed for this 10 TPD hydrogen liquefier concerning hydrogen safety, i.e., hydrogen gas components, especially, the oxygen level.

4 Cold Box Different from the vertical cold box of the 4.4 TPD hydrogen liquefaction plant at Ingolstadt in Germany [4], the cold box of this 10 TPD hydrogen-refrigerated hydrogen liquefier has been designed as a horizontal cold box with a vacuum insulated structure. The diameter of this cold box is approximately 4 m, the length of this cold box is approximately 20 m. There are 4 static gas bearing turbines will be mounted vertically on the cold box for convenient installation and maintenance. For convenient maintenance and Ortho-para catalyst loading, maintenance platforms are set both outside and inside of this cold box. Figure 6 shows the 3D design diagram of this cold box.

Vacuum relief valve

Four turbines Cryo-valves

Buffer tank for turbines

Cold box Vacuum pump

Fig. 6. 3D design diagram of the cold box

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5 Conclusion and Outlook A Process Flow Diagram design and a T-S diagram of a 10 TPD hydrogen liquefier has been presented in this paper. The P&ID (Process Instrument Diagram) design has been completed by Technical Institute of Physics and Chemistry of CAS. This 10 TPD hydrogen-refrigerated hydrogen liquefier will be assembled at the end of 2022 and will be commissioned early next year (2023). This 10 TPD hydrogen liquefier will be the second and the biggest large-scale hydrogen-refrigerated hydrogen liquefier in China so far. Acknowledgments. The project is supported by the fund of the State Key Laboratory of Technologies in Space Cryogenic Propellants (Grant No. SKLTSCP202101) and the fund of the National Key Research and Development Program of Ministry of Science and Technology of the People’s Republic of China (Grant No. 2020YFB1506201).

References 1. Won, W., Kwon, H., Han, J.-H., et al.: Design and operation of renewable energy sources based hydrogen supply system: technology integration and optimization. Renewable Energy 103, 226–238 (2017) 2. Li, J., Liu, L.Q., Xiong, L.Y., et al.: Dynamic simulations of medium-sized hydrogen liquefiers based on EcosimPro simulation software. IOP Conf. Series: Mater. Sci. Eng. (755) 012070 (2020) 3. Yujing, B., Yonglin, J.: Design and analysis of an efficient hydrogen liquefaction process based on helium reverse Brayton cycle integrating with steam methane reforming and liquefied natural gas cold energy utilization. Energy (2022). https://doi.org/10.1016/j.energy.2022.124047 4. Bracha, M., Lorenz, G., Patzelt, A., et al.: Large-scale hydrogen liquefaction in Germany. Int. J. Hydrogen Energy 19(1), 53–59 (1994)

Investigation on Performance of Large-Scale Hydrogen Liquefiers Based on Collins and Claude Cycles Rui Xue1,2 , Shaoqi Yang1 , Xiujuan Xie1(B) , Ningning Xie3 , Wei Wu1,2 , Kunyin Li3 , Guoqiang Shen3 , and Linghui Gong1 1 State Key Laboratory of Technologies in Space Cryogenic Propellants, (Technical Institute of

Physics and Chemistry, Chinese Academy of Sciences), Beijing, China {xuerui18,wuwei201}@mails.ucas.ac.cn, {yangshaoqi,xiexiujuan, lhgong}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China 3 Institute of Science and Technology, Three Gorges Corporation, Beijing, China {xie_ningning,yin_likun,Shen_guoqiang}@ctg.com.cn

Abstract. The large-scale hydrogen liquefier is a complex system with high energy consumption and multi-parameter coupling. The design calculation of the hydrogen liquefier is commonly based on steady-state conditions. In this paper, two hydrogen liquefaction systems based on Claude and Collins cycles are simulated and analyzed. The performances of the two systems including total UA of exchangers, specific energy consumption and exergy efficiency are compared, respectively. The results show that hydrogen liquefaction system of Claude cycle has higher energy efficiency and UA than that of Collins cycle. This paper provides references for the selection of process form for hydrogen liquefiers. Keywords: Hydrogen liquefier · Specific energy consumption · Exergy efficiency · UA of exchangers

1 Introduction With the development of China’s economy, the demand for energy is also increasing, while the reserves of oil and coal, as the main energy sources, are decreasing [1]. Thus the transition from fossil fuel to green energy is inevitable. Hydrogen is regarded as the clean energy with the greatest development potential in the 21st century, because of its high heat value of combustion, high energy density, wide range of sources, various forms of utilization, renewable and so on [2]. It is beneficial to solve the energy crisis and environmental pollution. As the world’s largest emitter of greenhouse gases, according to the Paris Agreement, China promises to achieve peak carbon emissions before 2030 [3]. However, the utilization of hydrogen energy needs to solve the problems of hydrogen production, storage, transportation, and application. Among them, the liquefaction process is one of © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 187–193, 2023. https://doi.org/10.1007/978-981-99-6128-3_23

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the most energy-intensive. In recent years, a large number of researchers and research institutions have been working to improve hydrogen liquefaction efficiency. Currently the world’s operating hydrogen liquefaction systems are commonly based on Claude liquefaction cycles with LN2 precooling. Many innovative processes have been proposed. Solar energy, geothermal energy, liquefied nature gas (LNG) cold energy and waste heat have all been used to reduce system energy consumption [4–7]. The specific energy consumption (SEC) of the current hydrogen liquefaction cycle designed in most published papers is mainly in the range of 6 ~ 10 kWh/kg LH2 , and the exergy efficiency (EXE) of the cycle is about 30% ~ 60% [8]. This paper selects Claude and Collins cycles of hydrogen liquefaction to carry on the performance analysis, which can provide reference for the choice of the cycle.

2 Process Flow Diagram of Large-Scale Hydrogen Liquefiers Based on Collins and Claude Cycles Liquid nitrogen tank

( a)

Turbine

LH

6

4

3

Refrigerant

5

7

2 1 8 The raw material of hydrogen

HX1

HX2

HX3

HX4

HX5

HX6

HX7

9

LH2 dewar

Ortho- para hydrogen converter

Liquid nitrogen tank

( b)

Turbine

Refrigerant

LH

4 3 5 2

1 6 The raw material of hydrogen

7 HX1

HX2

HX3

HX4

HX5

HX6

HX7

LH2 dewar

Fig. 1. Hydrogen liquefaction process based on Collins cycle (a) and Claude cycle (b)

Hydrogen liquefaction process based on Claude cycle and Collins cycle is shown in Fig. 1. The raw material of hydrogen at atmosphere temperature and high pressure is gradually liquefied through the heat transfer of the heat exchangers in the cold box. The cooling capacity mainly comes from the liquid nitrogen (LN2 ) precooling and the hydrogen refrigeration (HR) cycle which contains turbines and throttle valve. Furthermore, ortho-para hydrogen conversion through three ortho-para hydrogen converters is required. Because the latent heat of liquid hydrogen (LH2 ) is less than heat of ortho-para hydrogen conversion, the ortho-para hydrogen conversion can occur spontaneously in the long-term storage process of LH2 . Thus a large amount of LH2 will be evaporated if there is much ortho hydrogen in the hydrogen liquefaction.

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The difference between the two processes lies in the arrangement of the turbines. As shown in Fig. 1, the turbines are in series in the Claude cycle and those are in parallel in the Collins cycle.

3 Calculation Model In this paper, Aspen Hysys is selected as the process calculation tool. The efficiency of the compressor and the turbine are calculated as follows: ηCOMP =

WCOMP Hein − Heout , ηEXP = Hcout − Hcin WEXP

(1)

where W COMP is the work required for the isentropic compression, H cin and H cout are the enthalpy of the suction and discharge streams of the compressor, respectively. W TUR is the work produced by the isentropic expansion. H ein and H eout are the enthalpy of the suction and discharge streams of the turbines, respectively. Energy conservation equation in the heat exchangers is expressed., the total heat Qtotal exchanged between cold and heat streams in the heat exchanger is not only equal to the enthalpy difference H lhot between inlet and outlet of every hot stream, but also equal to the enthalpy difference H lcold between inlet and outlet of every cold stream in the heat exchanger. The UA of the heat exchangers is equal to the ratio of Qtotal and LMTD, which can reflect the heat transfer area and volume of heat exchangers. Reducing total UA (UAtotal ) value of seven heat exchangers can speed up cooling process and save some constructive and maintenance costs. Three ortho-para hydrogen converters in this paper are all adiabatic which can be simulated by the model of Conversion reactor in Aspen Hysys. Under the condition of sufficient catalyst, para-hydrogen concentration can at most approach the equilibrium concentration corresponding to the outlet temperature of the converters. Calculation method of para hydrogen equilibrium concentration is as follows [9]: −1  −5.313 eq ) + 0.1 − 2.52 · 10−4 t 3 H2 , para = 0.1 exp( t + 3.71 · 10−3 t 2−2.04·10

−3

− 0.00227

(2)

Here, t = T/T c where T c is the critical temperature of normal hydrogen, T is the temperature of hydrogen.

4 The Assumptions of the Calculation 1) The process is steady-state, ignoring the effects of kinetic and potential energy. 2) The adiabatic efficiency of compressors and turbines are 75%. 3) The pressure drops of hot streams and cold streams in the heat exchanger are 0.01 bar and 0.03 bar, respectively. 4) The temperature, mass flow rate and pressure of raw material hydrogen are 310 K, 59 g/s and 25 bar, respectively.

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5 Results and Discussions To compare two cases, some key nodes are set as the same state parameters. Then the optimizations of the hydrogen liquefaction process in two cases are carried out, respectively. When the parameters of the key nodes are given as Tables 1 ~ 2, reducing the compressor flow rate in the HR cycle can significantly decrease the SEC. However, when the total flow rate of the compressor is too low, the cooling capacity generated by the turbine and throttle valve will be insufficient. It results in “temperature crossing” in Aspen hysys, which violate the second law of thermodynamics. Therefore, under the given calculation conditions, the minimum SEC of two cases can be obtained respectively, and the optimization results of two cycles are compared as follows: Table 1. Key parameters from the hydrogen liquefaction process based on Collins cycle Node

Temperature (K)

1

299

2

310

3

310

4

72

5

Pressure (bar)

Mass flow (g/s)

O-hydrogen (%)

P-hydrogen (%)

1.02

400

75

25

4.80

400

75

25

25.0

400

75

25

24.97

340

75

25

72

24.97

60

75

25

6

45

24.95

50

75

25

7

45

24.95

290

75

75

8

310

25.0

59

75

25

9

21

59

4.6

95.4

1.49

Table 2. Key parameters from the hydrogen liquefaction process based on Claude cycle Node

Temperature (K)

Pressure (bar)

Mass flow (g/s)

O-hydrogen (%)

P-hydrogen (%)

1

299

1.02

346

75

25

2

310

3

310

25.0

4.90

346

75

25

346

75

25

4

72

24.97

288

75

25

5 6

72

24.97

58

75

25

310

25.0

59

75

25

7

21

59

4.6

95.4

1.49

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5.1 Energy Analysis of the Hydrogen Liquefaction Processes Based on Collins and Claude Cycles The total power consumption consists of two parts. One is from the HR cycle. The other is from the LN2 consumption. In the above calculation, LN2 consumption is converted into total power consumption as follows: Converted power from consumed LN2 = convert factor * flow rate of consumed LN2 (3). Considering the native market price of LN2 and electricity, the convert factor is set as 0.48 in this paper. As shown in the Tables 3, SEC of Claude cycle is 14.1% less than the Collins cycle, mainly because the Claude cycle has less power consumption from both HR and LN2 . Table 3. SEC, exergy and UAtotal of the hydrogen liquefaction cycles SEC from LN2 (kWh/kg LH2 )

SEC from HR (kWh/kg LH2 )

Exergy Efficieny (%)

UAtotal (kJ/K·s)

Claude

5.79

9.75

40.3

229

Collins

6.95

11.29

35.21

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5.2 Exergy Analysis of the Hydrogen Liquefaction Processes Based on Collins and Claude Cycles

Fig. 2. Exergy loss on different equipment of Claude cycle (a) and Collins cycle (b).

As Table 3 show, the Claude cycle has higher exergy efficiency than Collins cycle. As Fig. 2 (a) and (b) show, this difference mainly comes from the exergy loss of the heat exchangers. The minor factor for the difference of the cycle exergy efficiency is the exergy loss of the expanders. 5.3 UA of the Hydrogen Liquefaction Processes Based on Collins and Claude Cycles As the Fig. 3 shows, UA of Collins cycle is 7% less than that of Claude cycle which means lower costs for the construction and maintenance of the hydrogen liquefier. The

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main difference results from the UA of HX 4 and HX 6. Commonly, the process design needs to compromise between the two parameters of UAtotal and SEC. 260 240

HX 7 HX 3

HX 6 HX 2

HX 5 HX 1

HX 4

220 200

UA (kJ/K·s)

1 80 1 60 1 40 1 20 1 00 80 60 40 20 0

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Fig. 3. UA of the hydrogen liquefaction processes based on Collins and Claude Cycles

6 Conclusion From above comparison and analysis, the following conclusions are drawn based on the assumption in this paper: 1) The Claude cycle has lower specific power consumption. 2) The Claude cycle has higher exergy efficiency. For both cycles, the main exergy loss comes from the heat exchangers. 3) The Collins cycle has lower UA, which means less cost for construction and maintenance of the hydrogen liquefier and shorter cooling time for the cold box. Acknowledgments. This work was supported by the fund of The National Key Research and Development Program of China (Grant No. 2020YFB1506201) and the project of China Three Gorges Corporation (No.202103391).

References 1. Wu, Z., Zhu, P., Yao, J., et al.: Combined biomass gasification, SOFC, IC engine, and waste heat recovery system for power and heat generation: Energy, exergy, exergoeconomic, environmental (4E) evaluations. Appl. Energy 279, 115794 (2020) 2. Ipriani, G., Dio, V.D., Genduso, F., et al.: Perspective on hydrogen energy carrier and its automotive applications. Inter. J. Hydrogen Energy 39(16), 8482–8494 (2014) 3. Qi, Y., Stern, N., He, J.K., et al.: The policy-driven peak and reduction of China’s carbon emissions. Adv. Climate Change Res. 11(2) (2020) 4. Mehrpooya, M., Sadaghiani, M.S., Hedayat, N.: A novel integrated hydrogen and natural gas liquefaction process using two multistage mixed refrigerant refrigeration systems. Int. J. Energy Res. 44(3), 1636–1653 (2020)

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5. Bg, A., Mm, B., Ma, C., et al.: Hydrogen liquefaction process using solar energy and organic Rankine cycle power system. J. Clean. Prod. 235, 1465–1482 (2019) 6. Seyam, S., Dincer, I., Agelin-Chaab, M.: Analysis of a clean hydrogen liquefaction plant integrated with a geothermal system. J. Cleaner Product. 243, 118562.1–118562.16 (2020) 7. Azizabadi, H.R., Ziabasharhagh, M., Mafi, M.: Introducing a proper hydrogen liquefaction concept for using wasted heat of thermal power plants-case study: Parand gas power plant. Chin. J. Chem. Eng. 12, 10 (2021) 8. Yin, L., Yonglin, J.U.: Review on the design and optimization of hydrogen liquefaction processes. Energy frontier 14(3), 15 (2020) 9. Wilhelmsen, B.D., Aasen, A., et al.: Reducing the exergy destruction in the cryogenic heat exchangers of hydrogen liquefaction processes. Inter. J. Hydrogen Energy, 43(10), 5033–5047 (2018)

Design, Analysis and Optimization of Refrigerant Cycle in Cryo-compressed Hydrogen Storage Process Jingyao Yang1,2 , Haocheng Wang1 , Xueqiang Dong1,2(B) , Yanxing Zhao1 , Maoqiong Gong1,2 , and Jun Shen1,2 1 Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Chinese Academy of Science, Beijing 100190, China [email protected] 2 University of Chinese Academy of Sciences, Beijing 100049, China

Abstract. Large-scale application of hydrogen requires safe, reliable and efficient storage technology. Compared with current hydrogen storage technologies, cryocompressed hydrogen (CcH2 ) has the advantages of high hydrogen storage density, lower energy consumption, no para-to-ortho hydrogen conversion, etc. In this article, several CcH2 processes under different hydrogen storage parameters were simulated and compared, including mixed-refrigerant processes and reverse Brayton processes. The results indicated that the exergy efficiency of mixed-refrigerant processes is significantly higher than that of other processes. The mixed-refrigerant precooling and propane precooling were also compared, and the results showed that the mixed-refrigerant precooling can further reduce power consumption and improve efficiency. The specific power consumption and hydrogen storage density for the dual mixed-refrigerant process at 80 K and 50 MPa are 6.104 kWh/kg and 71.71 kg/m3 , respectively. Under similar hydrogen storage density, the power consumption of dual-MRJT is about 43% lower than that of typical dual-pressure Claude hydrogen liquefaction process (10.85 kWh/kg). Keywords: Cryo-compressed hydrogen · Refrigerant cycle · Dual mixed-refrigerant · Process optimization

1 Introduction Hydrogen (H2 ) is an excellent clean energy carrier with the advantages of extensive sources, high energy density, clean and pollution-free. However, the density of hydrogen is only 0.081 kg/m3 at 300 K and 0.1 MPa, while the volumetric energy density is 1/3000 of gasoline (32.05 MJ/L). Therefore, the development of safe and efficient hydrogen densification technology is the key to achieve large-scale application for hydrogen. Among current hydrogen storage technologies, the compressed gaseous hydrogen storage technology is mature, but the hydrogen storage density is low and the safety requirement is high. Hydrogen liquefaction could reach high density, which is suitable for large-scale storage and transportation, but the energy consumption is high. Metal hydride © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 194–202, 2023. https://doi.org/10.1007/978-981-99-6128-3_24

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hydrogen storage has merits of high volumetric density and low energy consumption, as well as disadvantages of high cost and long storage/release time. As contrast, cryocompressed hydrogen storage (CcH2 ) which combines cooling and pressure boosting has become a potential hydrogen storage technology for its high hydrogen storage density without too high pressure, too low temperature and too much energy consumption. The research on CcH2 was begun by Aceves et al. at Lawrence Livermore National Laboratory [1], who proposed an insulated pressure vessel operated at 20 K and 24 MPa. Based on the aluminum-lined composite-wrapped pressure vessel, they designed and tested three generations of vessels successively. Test results showed that these vessels were not weakened under cryogenic operations [1]. Inspired by these vessels, BMW developed a supercritical CcH2 vehicle storage system with favorable performance and safety [2]. Ahluwalia et al. [3] developed thermodynamic and kinetic models to study the charging/discharging process and storage capacity of CcH2 vessels. The results showed that the theoretical capacity of 71.3 g/L can be realized with an initial temperature below 180 K and a maximum pressure of 35 MPa. The performance of CcH2 system using Gen-3 vessel was also evaluated, which can meet the goals of US Department of Energy (70 g/L) [4]. Petitpas et al. [5] and Moreno-Blanco et al. [6, 7] comprehensively evaluated the factors affecting fill density and system performance of CcH2 vessel. The results showed that the fill density increased monotonically with vessel design pressure. Considering the weight and cost of the system, a design pressure range of 25–35 MPa can be selected. In most research, LH2 is injected into the CcH2 tank by LH2 pump to 35–70 MPa, to achieve low-temperature and high-pressure state by the characteristics of hydrogen itself [8]. However, the energy consumption of LH2 production is relatively high.

Fig. 1. Basic cycle diagram of CcH2

The above works mainly focused on the performance research of CcH2 vessels rather than the overall CcH2 processes. CcH2 generated by injecting LH2 leads to large overall energy consumption. Therefore, an effective and energy-saving CcH2 generation process is required for the popularization and application of CcH2 . In our previous work, a basic CcH2 process was proposed (Fig. 1) [9]. Hydrogen is compressed by multi-stage compressor unit, then cooled by refrigerator and finally sent to storage vessel. As hydrogen stays in gas phase, para-to-ortho conversion can be ignored without catalysts [2]. In this work, the cooling processes for CcH2 will be designed, analyzed, and optimized, to seek the optimal cooling process to realize the efficient storage of hydrogen.

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2 Refrigeration Process Design Our previous work found that the compression power decreases significantly with the increase of compression stages. The power consumption of five-stage compression is 47% of that of single-stage compression, while the power consumption difference between five-stage compression and six-stage compression is less than 3% [9]. Thus, considering the power consumption and system complexity, the process simulation adopts five-stage compression with intermediate cooling. Above the critical temperature, the heat load versus temperature distribution of hydrogen cooling process is almost linear, which is a typically distributed load. In order to reduce cooling energy consumption, it is essential to match the refrigerant and the distributed sensible heat load of hydrogen. Thus, variable-temperature cooling process is preferred, such as the reverse Brayton cycle, mixed-refrigerant J-T cycle, etc. 2.1 Mixed-Refrigerant Process In this work, the mixed-refrigerant process consists of a main mixed-refrigerant J-T (MRJT) cycle, a precooling cycle and a hydrogen path, as shown in Fig. 2. Neonnitrogen-hydrocarbon mixtures are employed as the mixed-refrigerants for the main cycle. According to the target cooling temperature, mixed-refrigerant components are selected according to their effective temperature ranges, such as Ne, N2 , CH4 , C2 H4 , C3 H8 , iC4 H10 . A propane precooling cycle is added to reduce the number of mixedrefrigerant components and the power consumption of the main cycle. Compressed hydrogen is firstly precooled to 263 K by precooling cycle and then cooled to the target temperature by the main MRJT cycle.

Fig. 2. Diagram of mixed-refrigerant process

2.2 N2 /Ne Reverse Brayton Process The N2 /Ne reverse Brayton process in this paper consists of a main N2 /Ne reverse Brayton cycle (RBC), a precooling cycle and a hydrogen path, just as shown in Fig. 3. The main refrigerant will change due to the change of the target cooling temperature. N2 is for the 100 K region while Ne is used in 80 K region. Compared with the mixedrefrigerant process, the N2 /Ne Reverse Brayton process is simpler and the perfusion of the N2 /Ne is more convenient. Similar to the mixed-refrigerant process, Propane is used to precool high-pressure hydrogen and the main refrigerant to 263 K. Then the compressed hydrogen is further cooled to target temperature by the main RBC.

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Fig. 3. Diagram of N2 /Ne reverse Brayton process

2.3 H2 Reverse Brayton Process H2 reverse Brayton process is based on hydrogen expansion refrigeration. High-pressure hydrogen is divided into feed hydrogen and circulating hydrogen. The cooling power generated in the expansion of circulating hydrogen is employed to cool the feed hydrogen. Propane is used as a precooling refrigerant. In order to improve the match of cold and hot streams, two stages of expanders are used, just as shown in Fig. 4. The H2 reverse Brayton process is simpler than the mixed-refrigerant process.

Fig. 4. Diagram of H2 reverse Brayton process

3 Process Calculation In the process simulation of this paper, PR-vdW model is used to calculate the properties of non-hydrogen refrigerants [10]. MBWR model is used for hydrogen. Specific power consumption (SPC) and exergy efficiency (η) of each process are calculated and analyzed. The calculation equations of SPC and η for the whole process are as follows:   (1) SPC = Woverall mH2 η = EQ WC In these equations, W overall (kW) represents the total power consumption of the CcH2 process, including the power consumption of hydrogen compression, precooling and main cooling cycle. mH2 (kg/h) represents the hydrogen inlet mass flow rate. E Q represents the total effective exergy (exergy gained by feed hydrogen). The simulation conditions are following: (1) (2) (3) (4)

Hydrogen is fed at 0.1 MPa and 308.15 K. The ambient temperature is 300 K. The maximum compressor discharge temperature is 385.15 K. The after-cooler outlet temperature is 313.15 K. The single-path pressure drops in heat exchangers are 70 kPa.

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(5) The adiabatic efficiencies of compressors and expanders are 75%. The process overall SPC was taken as the objective function, which was optimized by sequential quadratic programming (SQP) optimization method [11]. SQP method can be applied to optimize problems with or without inequality and equality constraints, which is recognized as one of the most effective optimization methods. Mixed-refrigerant compositions, flow rates and cycle operation pressures are selected as the optimization variables. The minimum temperature difference (T min ) of heat exchangers is taken as constraint. The objective function (2) and constraints (3) are shown in the following equations.  −−→ −−→ SPC = SPC(pyi, pyo, pzi, pzo, my, mz, n1, n2), min f (SPC) = (Wc + Wy + Wz) mH2

Tmin ≥ 3.0K

(2) (3)

where pyi , pyo , pzi and pzo represent the inlet and outlet pressure of precooling compressors and main cooling compressors, respectively. my , mz represent the mass flow rate of precooling refrigerant and main refrigerant, respectively. n1 , n2 represent the compositions of precooling refrigerant and main refrigerant, respectively.

4 Result Discussion 4.1 Comparison of SPC and Exergy Efficiency Under the hydrogen storage temperature of 100 K and 80 K, the variation trend of SPC and exergy efficiency of each process with variable hydrogen storage pressure is shown in Fig. 5. It can be indicated that the increase of hydrogen storage pressure has a great impact on the SPC of H2 reverse Brayton process. Under high hydrogen storage pressures, the SPC of H2 reverse Brayton process will exceed that of N2 /Ne reverse Brayton process. While the hydrogen storage pressure has a little impact on N2 /Ne reverse Brayton process and mixed-refrigerant process. H2 reverse Brayton process could achieve higher exergy efficiency than N2 /Ne reverse Brayton process under lower storage pressure. The power consumption of H2 compression unit increased with higher storage pressure, as shown in Fig. 5(d). However, the hydrogen expanders of H2 reverse Brayton process could lead to higher power consumption and massive exergy losses at higher storage pressure. That maybe the reason for the decrease of the exergy efficiency of H2 reverse Brayton process. The hydrogen storage temperature has a great influence on the SPC of the reverse Brayton process. As a contrast, the mixed-refrigerant process still maintains low energy consumption at the hydrogen storage temperature of 80K. In short, the performance (SPC and exergy efficiency) of mixed-refrigerant process is significantly better than the reverse Brayton process. The parameter ψ (ratio of hydrogen storage density to SPC) was defined, which indicates the hydrogen storage density that can be achieved under unit power consumption. Figure 5(c) shows the trend of parameter ψ for different processes under hydrogen storage conditions. The ψ value increases with the increase of hydrogen storage pressure and decreases with the decrease of hydrogen storage temperature. It can be seen that the mixed-refrigerant process shows obvious advantages, which could achieve relatively high hydrogen storage density under unit power consumption.

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Fig. 5. Variation of SPC (a) and exergy efficiency (b) of each process with hydrogen storage pressure; (c) ψ variation diagram of processes under different hydrogen storage conditions; (d) Power consumption of each unit in CcH2 processes under various storage pressures at 100K

4.2 Comparison of T-Q Distribution of Heat Exchanger The heat exchanger T-Q distribution of each process was analyzed under the typical hydrogen storage condition of 100K and 35MPa, just as shown in Fig. 6. It can be seen that the temperature difference of cold and warm streams in the N2 reverse Brayton process is large in cold region (Fig. 6 (b)). With an additional H2 distributed cooling load, the heat capacity of warm stream is enlarged further, leading to a worse match of cold and warm streams. Multistage expansion could be used to optimize the RBC process structure. For the H2 reverse Brayton process, the temperature difference is smaller in cold region. However, it is very difficult to manufacture expanders under such high expansion ratio. In the mixed-refrigerant process, the mixed-refrigerant and the H2 distributed sensible heat load match well. The temperature difference is relatively small (Fig. 6 (a)). Therefore the heat exchange exergy loss is small and the overall heat exchange efficiency is high. 4.3 Analysis of Precooling Method Based on the above analysis, the mixed-refrigerant process shows obvious advantages in reducing cooling power consumption and heat exchange loss. In this section, propane precooling and mixed-refrigerant precooling for the mixed-refrigerant process were compared. In the dual mixed-refrigerant process (both the main and precooling refrigerants

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Fig. 6. T-Q distribution diagram of heat exchangers under each process at 100K@35MPa

are mixtures), the precooling refrigerant components are determined as C2 H4 , C3 H8 and iC5 H12 . The main refrigerant is the same as that in the mixed-refrigerant process of propane precooling above.

Fig. 7. (a) Comparison of SPC and η of propane precooling and mixed-refrigerant precooling processes; (b)T-Q diagram of heat exchanger under dual mixed-refrigerant at 100 K@35 MPa

The comparison shows that the performance of mixed-refrigerant precooling is better than that of propane precooling. The SPC of mixed-refrigerant precooling is about 9.8% lower than that of propane precooling. The exergy efficiencies of dual mixedrefrigerant processes are relatively high, especially under high hydrogen storage pressures (Fig. 7(a)). Comparing Fig. 7(b) and Fig. 6 (a), it can be found that mixed-refrigerant precooling could achieve a good matching of cold and warm streams’ heat capacities in warm region. Thus, the exergy losses during precooling are reduced. Therefore, in order to improve the efficiency further, mixed-refrigerant precooling can be used in CcH2 processes instead of propane precooling cycles. In addition, dual-MRJT CcH2 process can achieve a hydrogen density similar to that of a typical dual-pressure Claude hydrogen liquefaction process with a 40% reduction in specific power consumption (SPC). For example, CcH2 density is 71.59 kg/m3 at 50 MPa and 80 K with SPC of 6.11 kWh/kg, LH2 density is 70.85 kg/m3 with SPC of

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10.85 kWh/kg [12]. As a result, the combination of CcH2 and high-efficiency mixedrefrigerant refrigeration technology is expected to further reduce the energy consumption of hydrogen storage and improve the economy of hydrogen storage.

5 Conclusion A theoretical analysis on mixed-refrigerant process, N2 /Ne reverse Brayton process and H2 reverse Brayton process for the cryo-compressed hydrogen storage are carried out in this work. Conclusions are listed below: (1) Although the structure of the N2 /Ne and H2 reverse Brayton process might be simpler than the mixed-refrigerant process, the mixed-refrigerant process could achieve good matching of the cold and hot streams, leading to smaller exergy losses. (2) The mixed-refrigerant process shows higher efficiency, which could achieve high hydrogen storage density at relatively low power consumption. (3) The utilization of mixed-refrigerant precooling can further reduce the energy consumption of the mixed-refrigerant process. The dual mixed-refrigerant CcH2 process could achieve similar hydrogen storage density with much lower SPC than the typical dual-pressure Claude hydrogen liquefaction process. Acknowledgments. This work is supported by the National Natural Science Foundation of China under the contract number of 51625603, 52006230, Key deployment projects of Center for Ocean Mega-Science, Chinese Academy of Sciences under the contract number of COMS2020Q1302.

References 1. Aceves, S.M., Berry, G.D., Martinez-Frias, J., et al.: Vehicular storage of hydrogen in insulated pressure vessels. Int. J. Hydrogen Energy 31(15), 2274–2283 (2006) 2. Aceves, S.M., Espinosa-Loza, F., Ledesma-Orozco, E., et al.: High-density automotive hydrogen storage with cryogenic capable pressure vessels. Int. J. Hydrogen Energy 35(3), 1219–1226 (2010) 3. Ahluwalia, R.K., Peng, J.K.: Dynamic of cryogenic hydrogen storage in insulated pressure vessels for automotive applications. Int. J. Hydrogen Energy 33(17), 4622–4633 (2008) 4. Ahluwalia, R.K., Peng, J.K., Hua, T.Q.: Cryo-compressed hydrogen storage. Compendium Hydrogen Energy, 119–145 (2016) 5. Petitpas, G., Moreno-Blanco, J., Espinosa-Loza, F., et al.: Rapid high density cryogenic pressure vessel filling to 345 bar with a liquid hydrogen pump. Int. J. Hydrogen Energy 43(42), 19547–19558 (2018) 6. Moreno-Blanco, J., Petitpas, G., Espinosa-Loza, F., et al.: The fill density of automotive cryo-compressed hydrogen vessels. Int. J. Hydrogen Energy 44(2), 1010–1020 (2019) 7. Moreno-Blanco, J., Petitpas, G., Espinosa-Loza, F., et al.: The storage performance of automotive cryo-compressed hydrogen vessels. Int. J. Hydrogen Energy 44(31), 16841–16851 (2019) 8. Aceves, S.M., Petitpas, G., Espinosa-Loza, F., et al.: Safe, long range, inexpensive and rapidly refuelable hydrogen vehicles with cryogenic pressure vessels. Int. J. Hydrogen Energy 38(5), 2480–2489 (2013)

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9. Zhao, Y., Gong, M., Zhou, Y., et al.: Thermodynamics analysis of hydrogen storage based on compressed gaseous hydrogen, liquid hydrogen and cryo-compressed hydrogen. Int. J. Hydrogen Energy 44(31), 16833–16840 (2019) 10. Kwak, T., Mansoori, G.A.: Van der Waals mixing rules for cubic equations of state. Applications for supercritical fluid extraction modelling. Chem. Eng. Sci. 41, 1303–1309 (1986) 11. Guoguang, M., Chen, Z., Yiran, D., et al.: A study on the use of dual mixed refrigerant in a cascade dual mixed refrigerant cycle. Adv. Mech. Eng. 9(6), 9 (2017) 12. Krasae-in, S., Stang, J.H., Neksa, P.: Development of large-scale hydrogen liquefaction processes from 1898 to 2009. Int. J. Hydrogen Energy 35(10), 4524–4533 (2010)

Study of Pressure and Temperature Fluctuations Applied to the ESS Cryogenic Moderator System H. Tatsumoto1(B) , P. Arnold1 , M. Segerup1 , P. Tereszkowski1 , Y. Beßler2 , and D. Lyngh1 1 European Spallation Source ERIC, Partikelgatan, Lund, Sweden

[email protected] 2 Forschungszentrum Jülich GmbH, Wilhelm-Johnen-Straße, Jülich, Germany

Abstract. The ESS cryogenic moderator system (CMS) will be operated at a pressure of 1.1 MPa and a temperature of 17 K to cool two hydrogen moderators where a nuclear heating is estimated to be 6.7 kW for a 5-MW proton beam operation. A pressure control buffer (PCB) tank with a heater and a control release valve has been designed to mitigate a pressure fluctuation caused by the sudden nuclear heating when the proton beams turn on or off. Excessive gas produced by the heater will be released by the valve and be re-condensed via a small heat exchanger. The pressure fluctuations have been analyzed using a mass and energy conservation. The effects of liquid level, liquid and vapor temperatures in the PCB on the pressure fluctuation have been clarified. The optimal operational conditions have been determined based on the analysis results. Keywords: Liquid hydrogen · Pressure control · Hydrogen moderator

1 Introduction In the beginning, the European Spallation Source ERIC (ESS) will install two hydrogen moderators above the target wheel [1]. The nuclear heating generated at the moderators is estimated to be 6.7 kW for a 5-MW proton beam operation. Figure 1 shows an overview of the ESS cryogenic moderator system (CMS), which consists of two centrifugal pumps in series, a pressure control buffer (PCB) tank with a volume of 0.065 m3 , an orthoparahydrogen convertor, and two kinds of heat exchangers (HX-1 and HX-2). Unlike the J-PARC [2] and SNS [3] where a supercritical hydrogen (1.5 MPa and 18 K) was selected, the CMS has been designed to provide subcooled liquid hydrogen (LH2 ) with a temperature of 17 K, a pressure of 1 MPa and parahydrogen of more than 99.8% to each hydrogen moderator at the flow rate of 0.24 kg/s. The overall CMS volume is 0.414 m3 . The pressure drop is estimated to be 110 kPa at the circulation flow rate of 0.5 kg/s where the two pumps are operated at a speed of 7,500 rpm [4]. Therefore, the pump discharge pressure is determined to 1.1 MPa. When the proton beams turn on or off, liquid the hydrogen temperature at the moderator is rapidly increased by a few Kelvin due to the kW-order nuclear heating. The heat load will be removed through the HX-1 by a helium refrigeration plant, which is called the Target Moderator Cryoplant (TMCP) © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 203–210, 2023. https://doi.org/10.1007/978-981-99-6128-3_25

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[5] and the feed hydrogen temperature should be almost maintained at 17 K. The density change of the liquid hydrogen would bring about a large pressure fluctuation because the CMS forms a closed loop. Therefore, the PCB tank with a control release valve and a heater has been designed to mitigate the pressure fluctuation within ±50 kPa. The heater has a function to aggressively increase the pressure by evaporation of liquid hydrogen. Excessive hydrogen gas will be released by the control valve and be re-condensed via the small heat exchanger (HX-2). In this study, the pressure fluctuations were analyzed for the nominal operational mode and the beam injection one to order. The effects of the liquid level, liquid and vapor temperatures were clarified. The PCB operational parameters were optimized.

2 Pressure Rise Analysis 2.1 Pressure Rise Analytical Model At the J-PARC CMS, Tatsumoto et al. [2] studied a pressure fluctuation caused by a nuclear heating and developed a pressure mitigation system. It was reported that the predicted pressure rises agreed very well with the measured ones for the proton beam powers. The pressure fluctuations applied to the ESS CMS was estimated in the same way as the J-PARC. In the analysis model, the CMS loop is divided into four sections considering the temperature distribution as shown in Fig. 2: (I) the supply line from the HX-1 to the moderators except for the PCB tank (V1 = 0.2091 m3 ), (II) the return line from the moderators to the HX-1 (V2 = 0.1408 m3 ), (III) a liquid phase in the PCB tank (V3 ) and (IV) a vapor phase in the PCB tank (V4 ). The total volume of the PCB tank is 0.065 m3 (= V 3 +V 4 ). The representative temperatures of the supply and return lines should be 17.4 K and 17.6 K, respectively, while the proton beams are off. The temperature only in the section (Il) is increased by 1.56 K at the mass flow rate of 0.5 kg/s after the 5-MW proton beams are being injected. The travel time of the liquid hydrogen from the moderator to the HX-1 is 20.3 s. The temperature in the section (I) is always maintained at 17.4 K under the assumption that the heat load can be completely removed by the HX-1. As the initial condition, the total hydrogen mass was calculated at the pressure of 1.1 MPa. The liquid hydrogen densities in the sections were given by GASPAK [6]. The pressure distribution over the entire CMS loop was neglected in this analysis. A pressure was changed repeatedly under the condition of the beam injection until the total mass is conserved. The representative vapor temperatures in the section (IV) is given in the next section. 2.2 PCB Operational Condition in the Nominal Condition Figure 3 shows an overview of the PCB tank with a volume of 0.065 m3 . The liquid level will be measured by not only a differential pressure transmitter but also five Silicon diode temperature sensors, which are located at 0.00314, 0.0102, 0.0157, 0.0212 and 0.0259 m3 , respectively. There are four heaters wrapped on the PCB tank between each temperature sensor. The release control valve (CV-62029) is located on the top of the

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Fig. 1. Overview of the ESS cryogenic moderator system (CMS).

Fig. 2. Simplified analytical model of the CMS loop.

PCB tank. The released gaseous hydrogen is re-condensed via the HX-2 and is returned to the suction side of the pump through the ortho-para hydrogen convertor. At the nominal condition, the pump discharge pressure will be set to 1.1 MPa. It is important to maintain not only the liquid level but also the vapor temperature in order to avoid any pressure fluctuation. The heater has a function to produce internal pressure by evaporating liquid hydrogen. On the other hand, the excessive gas is released by a PID control of CV-62029 to reduce the pressure to the set point. The CV-62029 position should be controlled in the range of 10 to 20%. At the same time, the subcooled liquid hydrogen with the temperature of 17 K would flow into the PCB tank at the evaporation

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Fig. 3. Overview of the PCB tank.

gas flow rate. Therefore, the liquid level would be kept constant. The average vapor temperature, Tav , is dependent on the evaporated flow rate and the radiation heat load. The upper two thirds of the PCB tank are not insulated with multilayer insulation (MLI). The effects of the unstable vapor temperature, T av , on the pressure change,P B , at the initial vapor temperatures, Tav , of 40, 50, 80 and 100 K are shown in Fig. 4. For higherTav , the vapor temperature fluctuation has a smaller impact on P. It turns out that setting higher Tav would be useful in a stable CMS operation at the nominal operation.

Fig. 4. Effects of Tav on the pressure change at the nominal condition.

The average vapor temperatures,Tav , for the evaporated flow rates, which was used in the pressure rise analysis mentioned in 2.1, were calculated by ANSYS FLUENT in order to estimate the heater power required to evaporate the liquid hydrogen. Figure 5 shows the analysis model, which treats only the vapor phase of the PCB tank where

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the MLI is not wrapped in order to reduce the required heater power. An inlet boundary condition was applied to the liquid surface where the temperature was set to the saturated temperature, Tsat (= 31.9 K), and the pressure was at 1.1 MPa. The evaporation gas flow rate, which was calculated using the latent heat of 1.96 × 105 J/kg and the heat input, was used as an inlet flow condition. A pressure boundary condition was applied to the top of the model. A wall boundary condition was applied to the PCB wall where the effect of a radiation heat transfer was considered. The Realizable κ-ε turbulent model with enhanced wall treatment was used. Polyhedral meshes were applied except for the vicinity of the wall where prism meshes of 10 layers were applied. The properties of parahydrogen given by GASPAK [6] were used and the temperature dependences are considered. Figure 6 shows the simulation results for the heat inputs, Qw . For Qw < 150 W, the average vapor temperature, Tav , is strongly affected by the evaporated flow rate. For Qw > 150 W, it approaches Tsat . The CV-62029 valve positions required for each Qw are also calculated under the condition that the pressure drop through it is 60 kPa. For Qw > 50 W, which corresponds to the evaporation gas flow rate, mev , of 0.33 g/s, the valve position can be maintained at more than 10%. The optimum Tav and liquid level in the PCB tank will be discussed in the next section.

Fig. 5. CFD analysis model.

3 Results and Discussion Figure 7 shows the effects of the return temperature rise T , on the pressure rise P B , and the liquid volume rise of the PCB tank V 3 , under the conditions of V3 = 0.035 m3 , TL = 17.4 K and Tav = 50 K for the two-moderator configuration. The pressure and the liquid level increase almost proportional to T . For the 5-MW proton beam operation where the return temperature rise is 1.65 K, the predicted pressure rise P B , is 75 kPa that is higher than the design criterion of 50 kPa. The effects of V3 , TL and Tav are studied and the PCB operational parameters are optimized to meet the requirement. Figures 8 and 9 show the effects of V3 and TL on P B and V 3 for the 2-MW and 5-MW proton beam operations under the condition of Tav = 80 K that corresponds to Qw

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Fig. 6. Simulation result of the average vapor temperature

of 60 W as shown in Fig. 6. For higher TL , P B and V 3 become smaller. For larger V3 , P B becomes smaller, although V 3 decreases. At TL = Tsat , the required heater power can be minimized and the evaporation of the liquid hydrogen can be produced quickly and effectively. The liquid level was determined to be kept at 0.025 m3 in light of mitigating the pressure change. Accordingly, P B and V 3 can be reduced to 39 kPa and 0.0016 m3 , respectively. Figure 10 shows the effect of Tav on P B and V 3 under the conditions of V3 = ˙ = 0.5 kg/s, which have been determined above. It seems 0.025 m3 , TL = Tsat and rate m that, P B and V 3 are little affected by Tav for Tav > 55 K. P B is lower than the criterion of 50 kPa. For higher Tav , the heater power required to produce the evaporating gas, QW , is decreased as shown in Fig. 6. Accordingly, the average vapor temperature, Tav , is determined to be 80 K where QW = 60 W, mev = 0.31 g/s and CV62029 = 12.3% in order to not only minimize the heater power, QW , and but also maintain the position of CV-62029 in the range of 10 to 15%. The heater power, QL , required to heat up the refilled liquid hydrogen to Tsat is 64.9 W. The heater C that is located on the liquid surface will be used to produce the evaporating gas. The refilled subcooled liquid hydrogen will be heated up by another heater (the heater D), which is located at the bottom. After the 5 MW proton beams are injected, the liquid level is being increased from and the pressure is being increased from 0.025 to 0.0266 m3 at 1.03 mm/s and the pressure is being increased from 1.1 to 1.139 MPa at 1.87 kPa/s during the travel time of 20.3 s. The operational pressure will be maintained at 1.139 MPa during the beam operation. When the proton beams have tripped, the pressure would return to the initial set pressure of 1.1 MPa at 1.87 kPa/s as well as the beam injection. After 20.3 s, the CMS operational condition should have been ready for resuming the proton beam injection. The PCB operational parameters have been determined based on the analytical results and are summarized in Table 1.

Study of Pressure and Temperature Fluctuations Applied

(a) Pressure change

(b) Liquid volume change in the PCB

Fig. 7. Effects of T on PB and V3 .

(a) Pressure rise

(b) Liquid volume change in the PCB

Fig. 8. Effect of V3 and TL on P B and V 3 for a 2-MW proton beam operation

(a) Pressure rise

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(b) Liquid volume change in the PCB

Fig. 9. Effect of V3 and TL on P B and V 3 for a 5-MW proton beam operation.

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(a) Pressure change

(b) Volume change

Fig. 10. Effect of Tav on P B and V 3 . Table 1. PCB optimum operational conditions. QW (W)

QL (W)

mev (g/s)

Tav (K)

V3 (liter)

TL (K)

60

64.9

0.31

80

25

Tsat

4 Conclusions The pressure rise analyses were conducted based on the mass conservation when the proton beams are injected. It turns out that P B is seriously affected by the liquid level, V 3 , the vapor temperature, Tav , and the liquid temperature in the PCB tank, TL . The PCB operational conditions have been optimized to mitigate the pressure rise within ± 50 kPa based on the analytical results.

References 1. Garoby, R., Danared, H., et al.: Phys. Scr. 93, 014001 (2018) 2. Tatsumoto, H., Aso, T., Ohtsu, K., Kawakami, Y.: Development of a partitionable accumulator with pressure tolerance for the cryogenic hydrogen system at J-PARC. Phys. Procedia 67, 123–128 (2015) 3. Spallation Neutron Source Final Safety Assessment Document For Neutron Facilities, Oak Ridge National Laboratory P. O. Box 2008 Oak Ridge, Tennessee 37831–6285 September pp. 3–35 (2011) 4. Arnold, P., Hess, W., Jurns, J., Su, X.T., Wang, X.L., Weisend, II, J.G.: IOP Conf. Ser: Mater. Sci. Eng. 101, 012011 (2015) 5. Tatsumoto, H., et al.: Quack IOP Conf. Ser: Mater. Sci. Eng. 755, 012101 (2020) 6. GASPAK user’s guide (Cryodata) (1998)

Cryogenic Components, Systems, Facilities, and Testing

Multi-field Coupling Rotor Characteristics Analysis for High Speed Turbo Expander Generator Brake Shun Qiu1,2 , Nan Peng1(B) , Liangwei Zheng1,2 , Han Yan1,2 , Changlei Ke1 , Kongrong Li1 , Xiaohua Zhang1 , Bing Dong1 , and Liqiang Liu1,2 1 Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Chinese Academy of Sciences, Beijing, China [email protected], [email protected] 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. At present, the world is in the period of developing large cryogenic system, and the power of turbo expander is increasing continuously. There are several limitations in the traditional braking mode, the direct drive high-speed turbo expander generator braking is becoming the trend. Taking a turbo expander with the power of 30 kW and the speed of 100,000 rpm as an example, the electromagnetic design of high-speed permanent magnet synchronous generator is carried out, and the main structural parameters of the generator are obtained. Furthermore, considering the influence of centrifugal force and interference fit pressure, a twodimensional stress analytical model was established, and the stress distribution of sleeve and permanent magnet are analyzed by theoretical and finite element analysis method. Then the critical speed and modal shape of the rotor were calculated by transfer matrix method, all the critical speed are far away from the rated speed. The results show that the rotor can meet the requirements of the turbo expander generator brake. Keywords: turbo expander · generator brake · stress · rotor dynamics

1 Introduction In recent years, with the rapid development of aerospace and large scientific devices, the scale of low temperature refrigeration/liquefaction systems is increasing, and turbo expander is the main refrigerating component, which is the key to the efficiency of entire system. As the gas expands in the turbo expander, its pressure and temperature decreasing, in this process, the brake is required to absorb power and maintain speed. The conventional braking methods are fan wheel braking and oil braking, there exist several problems such as unable to adjust speed widely and unable to recover power. With the breakthrough of high-speed generator, it is possible to adopt the turbo expander directly coupled to high-speed generator for braking. However, many challenges exist for high-speed generators, such as the electromagnetic design, rotor stress, bearings, © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 213–221, 2023. https://doi.org/10.1007/978-981-99-6128-3_26

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rotor dynamics, thermal and etc. Therefore, the rotor characteristic analysis for highspeed turbo expander generator brake is necessary. C.Zwyssig et al. [1] focuses on the design of a 100 W, 500000 rpm generator for use with a gas turbine, the design selects the suitable machine type and bearing technology, and determines the electromagnetic characteristics. Fang et al. [2] designed the rotor of high-speed high-power permanent magnet synchronous machines, the electromagnetic, thermal and rotor dynamic performances of the designed rotors are analyzed and compared. Zhang et al. [3] analyzed the rotors stress by theoretical and numerical methods, and proposed a new hybrid protective measure. Rodrigo et al. [4] used the transfer matrix method to predict bending critical speeds of hydrogenator shaft lines, and the obtained results are compared with the data from the equipment’s manufactures. This paper focuses on a turbo expander with rated power of 30 kW and rated speed of 100000 rpm, which using high-speed permanent magnet synchronous generator (HSPMSG) for braking. By means of multi-field coupling analysis, the rotor system of high-speed turbo expander generator brake can meet the requirements of mechanical, electromagnetic and dynamic characteristics simultaneously.

2 General Design of HSPMSG The design of HSPMSG can be summarized as the following steps: according to the rated parameters of the generator, select electromagnetic load to obtain the main size of the rotor and stator, and establish a two-dimensional model of the generator by finite element analysis to finish the final design. The active part of an electric generator, which is related the stator bore diameter Dis and the machine active length lef , can be determined by Eq. (1):   Pδ = π 2 /2 · kw · A · B · Dis2 lef · n (1) where Pδ is the air-gap power, kw is the fundamental winding factor, A and B are the electrical loading and magnetic loading, respectively, Dis is the inner diameter of stator, lef is the effective stator length, n is the rated speed. In order to place permanent magnet (PM) in the rotor as much as possible, materials with large residual flux density need to be selected, such as sintered NdFeB or SmCo. In order to conveniently install and adjust PM, a radial tile-shaped rotor magnetic circuit structure is proposed. Designed by Ansoft software, the main parameters of the generator are listed in Table 1. Then, Maxwell2D software was used to simulate the electromagnetic characteristics of the HSPMSG, the magnetic flux distribution, No-load back potential are shown in Fig. 1 (a), (b). The magnetic density of the stator yoke and tooth tip in Fig. 1(a) is up to 1.6T, which is not very high, indicating that the design is reasonable. The root-meansquare (RMS) value of the load voltage in Fig. 1(b) is 427 V, which is close to the rated value. The simulation results show that the designed HSPMSM has good electromagnetic performance and can meet the braking requirements of turbo expander.

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Table 1. Specifications of the HSPMSM. Parameter

Value

Parameter

Value

Rated power (kW)

30

Rated speed (rpm)

100000

Number of poles

2

Number of slots

24

Stator outer diameter (mm)

96

Rotor outer diameter (mm)

35

Air gap length (mm)

2.5

Effective length (mm)

60

Thickness of PM (mm)

6

Pole arc coefficients of PM

0.75

(a)

(b)

Fig. 1. Electromagnetic performance. (a) Magnetic flux distribution. (b) No-load back potential.

3 Rotor Stress Analysis The HSPMSG is coaxially connected to the turbo expander, and the PM cannot bear huge centrifugal force when the rotor rotates at tens of thousands of rpm, thus a retained alloy sleeve or a bound carbon fiber is required to protect the PM. The advantages of sleeve are shielding magnetic field, reducing loss, and its purpose is to produce a certain compressive stress in static state to compensate for the tensile stress in rotational state. According to the rotor mechanical design criteria, the following conditions must be met: (1) The maximum stress of the sleeve and PM must be less than the permissible stress, otherwise it will be damaged; (2) The contact pressure between the shaft, PM and sleeve must be positive, otherwise it will come loose and cannot transfer torque. The three-layer rotor structure of “shaft-permanent magnet-sleeve” is shown in Fig. 2, it was simplified as interference fit between thick-wall cylindrical sleeves, ignoring the axial elongation and assuming axial symmetry, a two-dimensional stress analytical model was established considering interference fit pressure and centrifugal force.

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Fig. 2. Three layer rotor structure of “shaft-permanent magnet-sleeve”.

According to the theory of mechanics of materials, the geometric equation of straindisplacement is shown in Eq. (2). εr =

du u dw , εθ = , εz = dr r dz

(2)

where, εr , εθ , εz is the radical strain, tangential strain and axial strain, u, w is the radical displacement and axial displacement. According to generalized Hooke’s law, the constitutive equation of stress-strain is shown in Eq. (3). ⎧ ⎪ εr = E1 [σr − v(σθ + σz )] ⎪ ⎪ ⎪ ⎪ ⎨ (3) εθ = E1 [σθ − v(σr + σz )] ⎪ ⎪ ⎪ ⎪ ⎪ ⎩ ε = 1 [σ − v(σ + σ )] z θ r E z where, σr , σθ , σz is the radical stress, tangential stress and axial stress, E is the young’s modulus, ν is the Poisson ratio. In rotation state, the rotor is subjected to centrifugal force, and the equilibrium equation of force is shown in Eq. (4). σr − σθ d σr + + ρω2 r = 0 dr r

(4)

Combined Eqs. (2), (3), (4), the displacement equation and stress equations can be obtained as Eq. (5): ⎧ ⎪ E E 3−2v ⎪ 2 2 ⎪ σr = (1+v)(1−v) A − (1+v)r ⎪ 2 B − 8(1−v) ρω r ⎪ ⎨ E E 2v+1 2 2 (5) σθ = (1+v)(1−2v) A − (1+v)r 2 B − 8(1−v) ρω r ⎪ ⎪ ⎪ ⎪ ⎪ ⎩ σz = v(σr + σθ )

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where A, B is the undetermined coefficient, its value depends on the displacement and stress boundary conditions, as shown in Eq. (6): ⎧ ⎪ ⎪ ⎪ σr1 |r=b = σr2 |r=b = −p2 ⎪ ⎪ ⎪ ⎪ ⎪ ⎨ σr2 | r=a = σr3 |r=a = −p1 ⎪ ⎪ ⎪ ⎪ ⎪ ⎪ ⎪ ⎪ ⎩

(6)

σr1 |r=c = 0 u3 |r=0 = 0

where p2 is the contact pressure between sleeve and permanent magnet, and p1 is the pressure between permanent magnet and shaft. As for stress analysis of different rotor components, different strength criteria should be selected. The sleeve is plastic, according to the fourth strength criterion, its equivalent stress (Von Mises stress) should be less than the permissible stress. On the other hand, the permanent magnet is brittle, according to the first strength criterion, its maximum principal stress should be less than the permissible stress. And the strength criterions are shown as Eq. (7): σequi =



0.5 (σr − σθ )2 + (σθ − σz )2 + (σz − σr )2 ≤ [σ ] , σprin = max(σr , σθ , σz ) ≤ [σ ] (7)

Taking the HSPMSG with speed of 100000 rpm and power of 30 kW as an example, the rotor stress is analyzed by theoretical method and FEA method respectively. The outer diameter of the rotating shaft is 8.5 mm, the thickness of the sleeve is 3 mm and permanent magnet is 6 mm.Selecting the static interference between the sleeve, permanent magnet and shaft as 40 μm and 10 μm. The materials of the sleeve, PM and shaft are Inconel 718, NdFeB and 38CrMoAlA respectively, the material properties are shown in Table 2. Table 2. Material property of rotor. material property

Shaft 38CrMoAlA

Permanent Magnet NdFeB

Sleeve Inconel718

Density ρ(kg/m3 )

7850

7400

8190

Elasticity modulus E(GPa)

206

160

205

Poisson ratio ν

0.28

0.24

0.28

Yield stress [σ]MPa

980

1100

1100

Tensile stress [σ]MPa

835

140

735

The Static Structural module in ANSYS Workbench is used to calculate the stress distribution of the rotor, the maximum principal stress of permanent magnet and equivalent stress of sleeve obtained by ANSYS and theoretical calculation are shown in the

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Fig. 3. According to the FEA results, the maximum principal stress of permanent magnet is 127.38 MPa and the maximum equivalent stress of sleeve is 650.63 MPa, both are less than the permissible stress.

Theoratical FEA

stress (MPa)

equivalent stress of sleeve

maximum principal stress of PM

radius (mm)

Fig. 3. Rotor stress distribution.

4 Rotor Dynamic Analysis The rotor is the rotating part of the turbo expander, which has the purpose of fixing the impellers and transmitting power, and can directly determine the performance of the turbo expander. In the turbo expander generator brake system, the impeller is fixed at the ends of the shaft, the rotor is supported by two radial gas bearings and two thrust gas bearings, the stator is located between the two radial bearings. Compared with the traditional fan wheel brake, as the rotor length and mass increased, the position of gas bearings changed, there are significant differences in bearing-rotor dynamics between the two braking methods. In this paper, transfer matrix method is used to calculate the critical speed and modal shape, to verify the feasibility of rotor dynamics in turbo expander generator brake.

(a)

(b)

Fig. 4. The shaft-bearing model. (a) Actual shaft line. (b) Discrete shaft line.

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The stress and deformation of the rotor are represented by the state vector Z, is defined in terms of transversal displacement y, axis slope θ , bending moment M and shear force Q, as shown in Eq. (8).

T Zi = y, θ, M , Q i

(8)

The entire shaft line should be properly discretized in element of shaft, element of mass, and element of guide bearings, as illustrated in Fig. 4. The masses are considered concentrated in points, the shafts are treated as elastic linear segments without mass, combined the mass element and shaft element, the transfer matrix can be related by Eq. (9). ⎡ ⎢1 + ⎢ ⎢ ⎢ Ti = ⎢ ⎢ ⎢ ⎢ ⎣

   2 l2  2 l2 Jp − Jd  mω − kj l + 2EI ω ω 2EI  2    2 l l2 1 + EIl Jp − Jd  2EI mω − kj ω ω EI     2 2 Jp − Jd  l mω − kj 1 ω ω

l3 6EI (1 − γ )

mω2 − kj

0

0

l 3 (1−γ ) 6EI l2 2EI

l 1

⎤ ⎥ ⎥ ⎥ ⎥ ⎥ ⎥ ⎥ ⎥ ⎦

(9)

i

where m is the point mass, ω is the angular frequency of bending vibration,  is the angular speed, kj is the rigidity coefficient of bearing, Jp is the rotor polar moment of inertia, Jd is the rotor transversal moment of inertia, l is the element length, E is the young’s modulus, and I is the moment of inertia, γ is the shear effect coefficient. By covering the entire rotor, from one end to another, one can relate the state vectors of the shaft by the product of all elementary transfer matrices, the total transfer matrix of the system can be related by Eq. (10). Zn+1 = Tn Zn = Tn Tn−1 Zn−1 = Tn Tn−1 · · · T2 T1 Z1 = An Z1

(10)

where An is total transfer matrix, and is of the same order of the elementary transfer matrices, i.e. 4 × 4. By imposing the free-ends conditions, Mn = M0 = 0 and Qn = Q0 = 0, the rotor critical frequency can be calculated by Eq. (11).   a a  (11) (ω2 ) =  31 32  = 0 a41 a42 The length of the rotor is 272.5 mm, the mass is 3.42 kg, the material is 38CrMoAlA, and it is divided into 30 nodes, the critical speed of each order is calculated by transfer matrix method, as shown in Table 3. After calculating the critical speed, the corresponding rotor modes of vibration can be solved by boundary conditions. The vibration mode curves of each order are shown in Fig. 5. It represents oscillation modes, translational mode, first order bending mode and second order bending mode respectively.

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order

First order

Second order

Third order

Fourth order

critical speed(rpm)

20794

28530

146274

392488

second mode

radial relative deflection (y)

radial relative deflection (y)

first mode

length (l/mm)

length (l/mm)

fourth mode

radial relative deflection (y)

radial relative deflection (y)

third mode

length (l/mm)

length (l/mm)

Fig. 5. Modal shape of the rotor.

5 Conclusion In this paper, the rotor of a 30 kW, 100000 rpm turbo expander coupled high-speed generator is analyzed, the rotor can meet the requirements of stress and dynamic characteristics. By interference fit, at rated speed, the maximum principal stress of permanent magnet is less than its allowable stress, the sleeve can effectively protect the permanent magnet from damage in the rotating state, and the simulation results verify the accuracy of the theoretical model. The dynamic analysis of rotor-bearing system using transfer matrix method can quickly calculate its critical speed and modal shape, all the critical speed are far away from the rated speed. Acknowledgments. This work was supported by the fund of The National Key Research and Development Program of China (Grant No. 2020YFB1506201).

References 1. Zwyssig, C., et al.: Design of a 100 W, 500000 rpm permanent-magnet generator for mesoscale gas turbines. In:40th Annual Meeting of the IEEE-Industry-Applications-Society, pp. 253–260. IEEE, Hong Kong, China (2005) 2. Fang, H.Y., et al.: Rotor Design for a High-Speed High-Power Permanent-Magnet Synchronous Machine. In: IEEE Energy Conversion Congress and Exposition (ECCE), pp. 4405–4412. IEEE, Montreal, Canada (2015)

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3. Zhang, F.G., et al.: Rotor retaining sleeve design for a 1.12-MW high-speed PM machine. IEEE Trans. Ind. Appl. 51(5), 3675–3685 (2015) 4. Albuquerque, R., Barbosa, D.L.: Evaluation of bending critical speeds of hydrogenerator shaft lines using the transfer matrix method. Proc. Inst. Mech. Eng. Part C-J. Mech. Eng. Sci. 227(9), 2010–2022 (2013)

Valves for Helium Cryogenics: Short Review and Experience on Installation Sergiy Putselyk(B) MAGNA Energy Storage Systems, Tesma Alle 1, 8261 Sinabelkirchen, Austria [email protected]

Abstract. An operation of cryogenic facilities with the liquid (LHe) or gaseous (GHe) helium is impossible without control devices. The key controlling components are cryogenic valves, which are used to regulate a flow of the cryogenic fluid as well as to liquefy GHe (also named as “JT-valves”). The present paper gives some comments and notes collected from a practical experience on a specification, installation and commissioning of cryogenic valves and Johnston couplings. Author’s experience on valve installation is also presented. A focus will be given to non-standard options, which could be helpful for some specialized applications. Keywords: Cryogenic valves · installation procedure · helium cryogenics

1 Introduction Valves are important components for a reliable operation of cryogenic facilities. Helium cryogenics stays additional requirements, for example, very low heat loads, high leakage tightness, etc., which leads to specialized solutions like valves with long stem length, thin wall thicknesses, bellows for leakage tightness, high manufacturing tolerances. In some cases, very specialized valves are required, for example with double functions of gas flow controlling and safety for helium gas release after quench of superconducting magnet, extra-long stems, etc. In the opening literature sources, specialized valves for space applications [1–3] or ones with cold actuator (also named as “U-boot”) [4–6] are sometimes described. However, descriptions of “standard” and specialized valves with a warm actuator are scare [7–10]. The present paper gives an overview on practical aspects, i.e. from specification till installation and commissioning, which are typically gained from an experience. Non-standard options (or sometimes also named as “specialized valves”) are discussed in detail. In additional, Johnston couplings are shortly considered. In order to fit within available space limitations for papers, discussions on some topics are avoided; so, it is assumed that missing topics have been explained during practical training courses on cryogenic valves and Johnston couplings at several manufacturing firms, e.g. WEKA, Velan, Flowserve, and readers have a practical experience on dis-/assembling of a couple of valves. This work was mainly done during the author work at KIT, Karlruhe, Germany © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 222–232, 2023. https://doi.org/10.1007/978-981-99-6128-3_27

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The second chapter gives an overview on the Johnston couplings, followed by one devoted to a specification of “standard” valves. In the last chapter, specialized valves are discussed.

2 Johnston Couplings Table 1 gives a short listing of specification parameters for Johnston couplings (JC). Table 1. an example of the JC specification. Name

Value

Medium (supply/return)

Helium (LHe, GHe), Nitrogen (LN2, GN2)

Operating pressure (supply/return)

< 25bar(a), or < 40/50bar(a)

Mass flow

-

Thermal anchoring at 40-80K temperature level

Yes/No

Comment

Higher pressures are possible upon request Positioning depends on customer requirements

Leak tightness: - process medium to iso-vacuum - iso-vacuum to atmosphere Flange interface

See note a

Pipe interfaces

See note b

Cryogenic length

875 mm c

Maximal heat load due to thermal conduction over metallic parts

see catalogues, d

Application in radiation areas?

Yes/No

Sealing gas connection for operation below 1 bar(a)?

Yes/No

Special options? - Electrical insulation between male and female parts - hand valve for closing JC - integrated check-valve - cold seal with spring energized v-ring

Yes/No

These options could be very helpful, see text below for further discussion

(continued)

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S. Putselyk Table 1. (continued)

Name

Value

List of spare parts

e.g. rubber o-rings, blind plugs

Special tools

e.g. shett wrench for tightening / removing of the collar nut

Comment

a) Flange interfaces could be very different, and typically presented in specialized drawings, e.g. for WEKA company: with KF-flanges (drawing 911001a), CF-flanges, standard “WEKA” flange (drawing 911121), Helicoflex-sealing (drawing 20020803), subatmospheric operation (drawing 940720), or other adaptor flanges similar to cryogenic valves, (drawing 970320). b) Pipe interfaces are typically kept fixed by each manufacturing firm, though an individual adaptation to customer requests is possible, for example, other diameters and wall thicknesses, flanges (e.g. KF, CF, Helicoflex, etc.). c) Each manufacturing firm has standard lengths for JC (for example, WEKA: 875 mm for Helium and 600 mm for Nitrogen), though individual adaptation to the customer needs is possible. d) The cryogenic heat loads depend on the length, size as well as optional thermal anchoring at the 40-80K temperature level and it is possible to calculate it by the thermal conduction though metallic walls (this value is typically presented in catalogues).

Some special options of JC could be also ordered (only few ones are mentioned), for example: Multi-coaxial JC, see for example Fig. 1, where three flows (two for helium and one for nitrogen) are realized. It is worth to note that this JC is designed for a small temperature difference between helium supply and return flows, because there is no vacuum insulation between these lines. However, it is possible to avoid this limitation by designing JC with an additional insulation vacuum between these lines. Electrical insulation between male and female parts of JC could be realized by a PEUHMW (virgil) with Vins ≤ 2kV. Hand valve for a closing of JC during the demounting of male and female parts to avoid the release of a process gas.

Fig. 1. Multi-coaxial Johnston coupling with three lines. Top: general view, bottom left: cold part, bottom middle: line for middle temperature, bottom right: warm part

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Cold seal with a spring energized V-ring (but also possible to apply a C-ring) to have a better leak tightness between different process lines. It is possible to use either KELF/1.4568 or PCTFE/17–7 PH (or 316 Stainless Steel). Check valves, which are opened by a male part of JC. So, it is also possible to apply a very special JC with an integrated check valve, which is closed during a demounting of the male part and is opened if male and female parts of JC are connected.

3 Standard Valves 3.1 Specification One of the most important documentations related to the cryogenic valves is a specification. Properly defined specification allows avoiding many valve modifications at the later stages, which could lead to a significant cost increase as well as time delays. Below is an example of the valve specification, which was finally supplied by WEKA company. Table 2. Specification example (valve of WEKA company is chosen as an example). Name

Value

Type: regulation On/Off Option (or with helium cold actuator – U-boot)

Yes -

DN/PN

10/25

Medium (He, Ne, N2 …), gaseous or liquid state:

GHe

Comments

Kvs value (Plug number) or Tinlet , 0.12 Toutlet , Pinlet , Poutlet , Mass flow: 0.05 Reference to document: 921104, Rev. 2004/Br “Flow control plugs” Note 1: these plugs/seat numbers are referred to the DN4 and DN6, respectively, however, these plugs/seats must be installed in the DN10 valve Note 2: for the official proposal, tools, which are needed to install/remove/change of seat and plugs, must be also included

Option: changeable seat Plug 36864, Type 6/10–2.5 (installed) Plug 36850, Type 4/10–1.0 (supplied as reserve)

Valve travel total

Standard delivery – 10 mm

15 mm

(continued)

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Name

Value

Comments

Flow characteristics (e.g. linear, equal percentage or special): Note 3: if other characteristics are possible, it must be clearly stated in the official proposal

equal-% linear

Plug 36864, Type 6/10–2.5 Plug 36850, Type 4/10–1.0

Cryogenic length, mm Reference to document: WEKA Kryoventile, FHo/Rev.Nov 2002

600 mm

Thermal anchoring: LN2 temperature and other temperatures

Yes

It is possible to have an additional thermal anchoring at 4 K

Materials: Plug: CuAl10FeNi5 (standard) Stainless Steel polished Seat seals: PTFE (standard) or PCTFE or HD-PE1000 Other: (PI, PEEK, Ni, Satellite 6) Insert: 316 L or Other Body: 316L (standard) or 321 or 304L Upper part, union nuts: Brass (standard) Stainless steel A2 Upper spindle (Stainless steel grade A2/A4) Static O-rings joints: NBR70; FPM75 (standard) Karlez/Chemraz or Helicoflex Security seal for bellows sealed valve: O-ring in NBR70 or FPM75 (standard) or PTFE

Yes Vespel 316L 316L Yes Yes EPDM EPDM

Note: Vespel is needed for radiation stability Note: EPDM is needed for radiation stability

Dimensions of cryogenic connection pipes:

17.2·1.6

L = 47 ÷ 50 mm

Body pattern: E (standard), or Z or D or Y

E

See also drawing 911003 (continued)

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Table 2. (continued) Name

Value

Comments

Pipe connection: Butt weld ends (Standard) KF flange or CF Flange or Other Connection to the vacuum vessel: Welded (standard) O-ring Spindle sealing: Bellows Guard-ring He-guard connection

Yes Yes Bellows No No

Dimensions of the welding flange allows a valve mounting from the top into vacuum vessel Reference to document 910930, Rev. 11.07.2000 HU “Standard dimensions for vacuum welding flange for cryogenic valves DN2…DN150

Connection for a pressure No measurement in volume between valve body and spindle Connection for an installation of safety valve in volume between valve body and spindle

No

Convection brake

No

Actuator: Pneumatic Manual or Electric

Pneumatic

Electrical actuator: Linear actuator or control function Power supply Depending on function: - Limit switch, - Travel switch - Travel feedback potentiometer - Electronically positioner 0/4-20 mA or 0-10V Other equipment (which one?) Manual actuator: Integral Leg design Precision drive 1/100 backlash for control Fit switches: (electromechanical, inductive) Dimension: Material: Integral design Leg design Precision drive (continued)

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Name

Value

Comments

Pneumatic actuator: Pressure (if specially needed) Position (in case of loss of air) Position (normally opened, normally closed) Service or shut-off pressure (bar) Electropneumatic positioner: analog (4-20 mA) Intelligent (4-20 mA) with - Alarm (2-off switches, 1-off alarm, 1-off binary) - Analogues travel feed back as 4-20 mA - HART protocol Intelligent (Profibus PA) with - Alarm (2-off switches, 1-off alarm, 1-off binary) - Analogues travel feed back as 4-20 mA - HART protocol pneumatic positioner signal 0.2–1.0 barg mounting: side (standard) integral smaller size Airset with filter and pressure gages Connection for the air supply Auxiliaries: limit switch: micro or inductive or capacitive 3/2 way solenoid valve Other equipment (if needed)

max. 6bar(g) closed closed up to 6bar(g) No 1/8 NPT-F

Maximal air pressure of 6 bar (g) is available Samson positioner with separated control and reading units will be used The valve stem must have the suitable connecting element/unit to the positioner

Tests: Leak tightness 300 K - To atmosphere - To vacuum - Through the seat 77 K - To atmosphere - To vacuum - Through the seat Functional test Certification Additional tests?

Yes No Standard Yes No

1·10–6 mbar·l/s 1·10–8 mbar·l/s 1·10–4 mbar·l/s CE declaration conformity, Material Certificate, Leak check result certificate, Pressure test certificate, PED conformity certificate

Other actuator: Pneumatic piston, or hydraulic or fluid self-controlled

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The next important valve characteristics is a heat load at 4 K temperature level. At the time being, manufacturing firms optimized the heat load to obtain very low values. It is quite difficult to compare these values proposed by different manufacturing firms due to following reasons: i) main heat load is due to a conduction through the valve body and spindle, while gas conduction or convection between body and spindle is very low and could be practically neglected, ii) these values are not measured with sufficient measurement accuracy at 4 K temperatures to have possibility to compare them from different manufacturing firms. According to the author’s knowledge, two attempts to measure the heat loads on several cryogenic valves manufactured by the WEKA company and installed inside large cryogenic boxes have been performed, but due to measurement uncertainties, only the upper limit was roughly estimated. So, “averaged values” of the heat load for cryogenic valve were measured, and it was confirmed that the specified values presented in catalogues have been guaranteed. In the future, it would be helpful to build a dedicated cryostat for the cryogenic heat load measurements, though measurements will be still a challenge due to small heat loads of the valves in comparison to the “background” heat load of the cryostat. For that reason, it is also possible to define a heat load on 4 K temperature level in the specification; however, it will be a challenging task to prove this value in-situ. Due to an available space limitation in this paper, it is not possible to discuss each option mentioned in Table 2, and for a right valve choice it is necessary to pay a special attention to details. It is also worth to thank cryogenic companies, for example, Flowserve, Welan, Weka, for the excellent developments of cryogenic valves for the helium temperature. Due to many options, it is possible to find a valve configuration practically for any application. 3.2 Installation In many cases, an installation of the cryogenic valves could be a very challenging task. Due to a relatively long valve body and thin walls, it is very easy to bend valves due to improper installation, thermal stresses due to shrinkages, or fluid pressures. In this case a proper valve functioning could not be possible and a quite expensive and timeconsuming re-installation would be required. Therefore, a careful piping design as well as stringent valve installation must be performed. Unfortunately, many manufacturing firms do not provide detailed installation instructions with an exhaustive description of each step and many cryogenic groups have to develop their own “recipes”. Below is one, which is based on the author’s experience, and should not be considered as the only one possible. • Estimation of a pipe displacement: Very careful estimation of the pipe displacement at the cold state must be done either only per hand calculations with analytical formulas or by other programs, for example, ANSYS Mechanical. For such estimations, a valve position at a vacuum flange of the vacuum vessel is assumed to be fixed. It is possible to compare these calculated displacements to the maximum allowed ones, for example, WEKA company document: “20100223 – Allowed Piping Loads – Allowed Displacement”, however, many manufacturing firms provide such documents only for information and no details on a mechanical design of the cryogenic valves is further

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•

•

•

•

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given. Nevertheless, as a rule of thumb for WEKA valves under a careful installation, author used the value of the maximal lateral displacement of 2 mm at cold state for valve length of 600 mm and 3 mm for 800 mm, respectively. Valve dimensions: DN2-DN15 valves are considered as ones with “flexible” body, while DN20 and larger – as ones with “rigid” ones. Practically it means that bodies of valves with DN20 and larger dimensions under the pipe displacements will not be bended. With other words, it is the most probable, that the welding place between the valve body and cryostat flange will be slightly deformed, while for smaller valves, the valve body and the welding place could be bended or deformed at the same time. Options related to a cryostat design: there are two possibilities to mechanically install the valve: a) either a direct welding of valve flange (or sometimes with an additional adaptor flange) to cryostat vacuum vessel, or b) with additional cylinder in between. There is a remarkable difference between these two possibilities. In the first case the valve flange is approximately fixed at the position, because only a welding could be slightly deformed. In the second case there are two weldings and cylinder between them and this “configuration” leads to an additional flexibility of all components (i.e. two weldings, cylinder) and larger allowable tolerances on the valve bending is allowed. In the letter case the valve body is kept approximately straight under additional load due to piping, pressure or thermal shrinkage and a valve bending is “allowed” by two welds and the cylinder. Tools: for the valve installation it is very practical to have two tools: i) a blind flange with a connection for a reforming gas to make the welding, and ii) straightness checking device (“Prüfdorn” according to WEKA notation), which allows to prove the straightness of the valve body without the stem. This device is typically used for valves up to DN15 sizes, because the larger valves have a very rigid body, and it is quite difficult to bend them. Installation procedures: the installation procedure is surely a “cooking recipe” and each manufacturing team has own one. In many cases, the know-how is “kept” by welders, which are “key-persons” in valve installation. At the beginning, valve is spot welded at the top flange (or at cylinder) of vacuum vessel at three points and valve is approximately kept at a required position. After that a piping is very carefully made by bending and adjusting them to the valve flanges within one- or two-millimeters tolerances. If it is necessary, adaptor or transition pipe pieces are prepared and carefully matched to the required positions. The main idea is to adjust the piping to valve connections to keep the residual forces on the valve body to the minimal values. The next step is to spot weld the piping to the valve connections with one or two (in some case three) points with final dimensional check. It is very important not to apply any forces on the valve body, which are sometimes done by welders to mate the connections for an easiness of the sport welding. In many practical cases, there are several other valves or JC present, so the above-mentioned procedure is repeated for all valves or JC. After a visual inspection, the welding of valve body to the pipes starts with larges valves or JC and proceeds to the smaller ones. To keep thermally induced forces due to welding to the minimum, valve bodies are cooling with wet “swam”, and welding is performed in three steps, starting with a very small welding power. However, after the welding there are still residual mechanical deformations on pipes, which also induce forces on the valve bodies. To reduce these forces, the

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spot welds between valve body and top flange (or cylinder) of vacuum vessel are removed and valve is allowed to settle in an equilibrium position, which could be slightly different from the initial one. Next step will be to perform three spot welds again to perform the fixation of the valve in the final positioning, and a final welding is also performed. It is worth particularly to note that in some cases (typically, if work is carefully done), it is not necessary to release three spot weldings between valve and top flange of vacuum vessel for the resettling of the valve in the equilibrium position after the pipe welding. In this case a final welding of the valve flange to the one of vacuum vessel is performed. Final checks: the final valve positioning is checked with the straightness checking device. It is also worth to note that this device is also typically used for a checking of the valve straightness after a delivery from the manufacturing firm or just before the valve installation.

4 Valves with Special Options It is also possible to order the valves with special options. There are quite large numbers of possibilities to make the special valves, but they are typically separated in two groups: • “Combination” of two different valve types, i.e. control and safety ones, e.g. Velan valve at LHC transfer line inside the tunnel. • Some untypical dimensions or designs, for example, valve lengths up to 2 m, or valves with additional separating connections at the cold state, i.e. a very small part of the valve at the 4 K temperature level and bellow between stem and valve body is preinstalled on a cold mass, and after that the second parts of the valve body and stem are connected (such kind of valve was produced by the Weka company). In some cases, in addition to a changeable plug, it is possible to order an option of a changeable seat from “soft” materials, like PTFE, HD-PE or others. It is worth to sincere appreciating the cryogenic firms for the collaborative work on “special valves” because final development costs could be several times the customer paying.

5 Conclusion In the present paper, the Johnston couplings and cryogenic valves operating at 4 K temperature level are discussed. Valves with standard and special options are shortly mentioned. One of possible installation procedures is also presented.

References 1. Frank, D.J., et al.: Testing and application of a motorized valve for the containment of superfluid helium. CEC Proc. 35, 359–367 (1990)

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2. Song, N., et al.: Cryogenic valve pneumatically actuated by use of vapour pressure. Cryogenics 45, p617-619 (2005) 3. Fuchs, F., et al.: Directly operated valve for space cryostats. Cryogenics 27, p15-19 (1978) 4. Suesser, M.: Betriebserfahrung mit Tieftemperatur-Ventilen mit kalten Antrieb, pp. p91-99. Deutsche Kalte Verein, Bremen (2000) 5. Holdener, F.: Kryoventile mit kaltem Antrieb, pp. p79-89. Deutsche Kalte Verein, Bremen (2000) 6. Daido, K., et al.: Characterization of Bonnetless Cryogenic Valves. ICEC Proc. 16, 481–484 (1996) 7. WEKA Seminar for cryogenic valves, KIT, 16 February 2011 or DESY, December (2015) 8. Erni, P., et al.: Cryogenic valves in restricted areas: possible configuration of valve control elements, presentation at ICEC (2016) 9. Boersch, M., et al.: Test procedures and functional verification of cryogenic valves considering operating conditions, Poster presented at CEC (2015) 10. Erni, P., et al.: Testing and qualifying cryogenic safety valves at cold conditions, Poster presented at CEC (2017) 11. Holdener, F., Brütsch, D.: Cryo-components for special applications, Presentation transparencies (2007)

Design, Fabrication and Installation of the Cryogenic Distribution System for Re-Configured FRIB A1900 Fragment Separator Nusair Hasan(B) , Mathew Wright, Venkatarao Ganni, Fabio Casagrande, Shelly Jones, Brandon Laumer, Chinh Nguyen, Adam Fila, and Nathan Joseph Facility for Rare Isotope Beams (FRIB), Michigan State University, 640 S Shaw Lane, East Lansing, MI 48864, USA [email protected]

Abstract. The A1900 fragment separator segment of the continuous wave heavy ion beam linear accelerator at the Facility for Rare Isotope Beams (FRIB) consists of twelve superconducting magnets. In the past, these superconducting magnet cryostats (and associated beamline, cryogenic distribution system) were part of the Coupled Cyclotron Facility (CCF). Reconfiguration of these superconducting magnets for the FRIB fragment separator required re-routing, rebuilding and addition of new cryogenic distribution lines. The re-configured A1900 fragment separator segment cryogenic distribution was designed using the same operational concept used for the FRIB experimental system cryogenic distribution system – which has separate lines for cool-down and 4.5 K operation. This provided flexibility for commissioning, operation, and maintenance of the segment without affecting other loads. These modifications of a legacy system to fit new requirements presented several design challenges which were resolved during the concept design phase. Design, fabrication, and installation of most of the elements of the cryogenic distribution system were carried out in-house at FRIB. This paper presents an overview of the process design, analysis, fabrication, and installation of the re-configured A1900 fragment separator cryogenic distribution system. Keywords: Cryogenic Distribution · Superconducting Magnets · Helium Cryogenics

1 Introduction 1.1 Background The Coupled Cyclotron Facility (CCF) at the National Superconducting Cyclotron Laboratory (NSCL) was shut down last year (2021) after successfully operating its user program for 20 years [1]. Following the completion of the CCF program – the beamlines, instruments, detectors and other components of this facility has been re-configured and integrated to the Facility for Rare Isotope Beams (FRIB) [2]. In the NSCL era, © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 233–239, 2023. https://doi.org/10.1007/978-981-99-6128-3_28

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the cryostats, which were mostly superconducting magnets from the K1200 and K500 cyclotrons, as well as the A1900 fragment separator, S800 spectrograph and other experimental vaults were connected by a cryogenic transfer line network that was fed from the NSCL cryogenic helium refrigerator (colloquially known as the ‘Green Cold Box’). This cryogenic refrigerator along with the main trunk of the distribution transfer line were commissioned in the early 2000’s as part of the NSCL cryogenic system upgrade project [3]. Due to the ‘organic evolution’ of NSCL over the last four decades, the cryostats were connected by a complex distribution system stemming from a main cryogenic distribution line. Physical layout and operational modes of the existing NSCL cryogenic distribution system are discussed in [4]. 1.2 Concept for Re-configuration of Cryogenic Distribution

Fig. 1. Simplified block diagram of the re-configured NSCL cryogenic distribution system (blue text refers to the cryogenic distribution component/interface, red text refers to the cryostat/loads)

Reconfiguration of the A1900 fragment separator for FRIB required re-routing, rebuilding and addition of new cryogenic transfer lines to the existing NSCL cryogenic distribution system. To improve the operability and maintainability of this system, the existing cryostats (on the NSCL cryogenic system) were divided into sub-segments and independent cryogenic distributions connecting to the existing NSCL cryogenic distribution system were designed. A simplified block diagram for the re-configured NSCL cryogenic distribution system is shown in Fig. 1. According to the reconfiguration plan, the NSCL main transfer line (Fig. 1, NSCL TL1) will be connected to a new cryogenic

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distribution segment (Fig. 1, NSCL TL2A), approximately 15.0 m long and branching out in to four sub-segments to support different cryostats.

Fig. 2. (a) Geometrical layout of the re-configured NSCL cryogenic distribution transfer line with associated bayonet boxes, and (b) cross-section of the transfer line.

Geometrical layout of the NSCL cryogenic distribution transfer line (modification) and the sub-segment/expansion interfaces are shown in Fig. 2a. The sub-segments are for the cryostats in A1900 fragment separator, Transfer Hall (T-Hall) beamline, (future) K500 cyclotron, and (future) High Rigidity Spectrometer. The A1900 fragment separator, although supported by the NSCL cryogenic distribution, is an extension of the FRIB target and pre-separator beamline. In addition to these four sub-segments – there is an additional expansion segment for future connection to/from FRIB experimental system (ES) cryogenic distribution. The interface to the existing NSCL cryogenic transfer line (Fig. 1, point A) is welded connection, while that to the sub-segments and the expansion segment are bayonet-style connections. These sub-segments and the expansion interface are consolidated into two bayonet boxes (box 1 and box 2). Altogether, this cryogenic distribution system is designed to support over 55 cryostats, with a total liquid helium inventory of over 10,000 L. Process and component design of this transfer line and the associated sub-segments are discussed below.

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2 Process and Component Design Following the operational concept of the FRIB experimental system cryogenic distribution [5], the new NSCL cryogenic distribution segment and branches are planned to have three different cryogenic circuits enclosed within the vacuum jacketed transfer line – primary (4.5 K) helium supply/return, thermal shield (liquid nitrogen cooled) and helium cool-down lines to cryostats. Each of these lines are 1.0 NPS (approx. 33.4 mm) in diameter, except for the 4.5 K helium return (2.0 NPS/60.0 mm) and vapor nitrogen (VN) return (1.25 NPS/42.1 mm). As mentioned in Sect 1.2, the sub-segments will be connected to the main cryogenic distribution transfer line (NSCL TL2A, Fig. 2a) using bayonet style removable couplings. The cross-section of this transfer line is shown in Fig. 2b. The design capacity for each of these three circuits are presented in Table 1. Design of the cryogenic distribution for each of the sub-segments follow the same concept for the main cryogenic transfer line. However, several additional components, such as – liquid nitrogen thermo-siphon can, cryostat valve/distribution boxes needed to be designed for the sub-segments. Figure 3 shows the simplified schematic for the A1900 fragment separator sub-segment. The other three sub-segments follow a similar concept for the process flow. Design details for these sub-segment components (especially for A1900 fragment separator) are discussed in Sect. 3.

Fig. 3. Simplified schematic diagram of the A1900 fragment separator cryogenic distribution and associated liquid nitrogen thermo-siphon can.

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Table 1. Design capacity for each of the cryogenic circuits of the re-configured NSCL cryogenic distribution transfer line. Distribution Segment

Cryogenic Circuit

Maximum Capacity

Main Transfer Line (NSCL TL2A)

Primary (Helium, 4.5 K)

2000.0 W

Thermal Shield (Nitrogen, 80 K)

250.0 g/s

Cool-down (Helium, 300–80 K)

10.0 g/s

Primary (Helium, 4.5 K)

500.0 W

Thermal Shield (Nitrogen, 80 K)

50.0 g/s

Cool-down (Helium, 300–80 K)

10.0 g/s

Sub-segment (A1900, T-Hall, K500, HRS)

3 A1900 Fragment Separator Cryogenic Distribution System During the design phase of the cryogenic distribution system for the FRIB target vessel and fragment pre-separator, several elements of the cryogenic transfer line (e.g., vacuum break, piping support, piping anchors) were designed from grounds up. These element designs are discussed in [5] and are adopted for the re-configured NSCL cryogenic distribution (including the A1900 fragment separator segment). Some new challenges were also encountered while designing for the legacy cryostats – such as, interfacing with different styles of cryostats (‘continuous filled’ and ‘batch-filled’ types), proper control and distribution of thermal shield (liquid nitrogen) flow due to the inadequacy of system design and instrumentation etc. These design challenges were resolved during the concept design phase and are discussed here. 3.1 A1900 Cryogenic Distribution Layout and Components

Fig. 4. Geometrical layout of the A1900 fragment separator cryogenic distribution and associated components.

The A1900 cryogenic distribution follows the FRIB fragment separator beamline and has the shape of an arc. Geometrical layout of this cryogenic distribution is shown in Fig. 4. This cryogenic transfer line supports twelve cryostats (superconducting magnets)

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at 4.5 K with a liquid nitrogen cooled thermal shield. Major components of this system are shown in Fig. 5. The bayonet box (Fig. 5a) serves as the interface to the NSCL cryogenic distribution. In addition to the bayonet box, there is a LN thermos-siphon can and twelve cryostat valve boxes. These valve boxes are branched off the cryogenic transfer line. The cryogenic transfer line headers are housed in a 10 NPS vacuum jacketed pipe (Fig. 5d). Flexible (expansion) elements are added to the vacuum jacket for proper alignment of the transfer line.

Fig. 5. 3D model of different cryogenic distribution system components. (a) A1900 bayonet box, (b) A1900 LN thermos-siphon can and integrated cryostat valve box, (c) cryostat valve box, and (d) cryogenic transfer line segment.

LN Thermo-Siphon Can In the existing nscl cryogenic distribution system, the liquid nitrogen (ln) cooled thermal shield was set-up to operate in a forced flow (direct cooled) configuration. The ln was directly fed from a storage system, and the return vapor (from cryostats) was partially used in the pre-cooler section of the nscl cryogenic refrigerator. The rest of the (saturated) nitrogen vapor was vented. Proper control and distribution of the thermal shield flow among all the cryostats was very difficult due to the lack of properly placed valves and instrumentation in these legacy cryostats – resulting in two-phase (liquid rich) flow in the shield return piping. To mitigate this issue, a central

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thermo-siphon can (Fig. 5b) feeding all the cryostats is designed for each of the subsegment. For the a1900 fragment separator sub-segment, this thermos-siphon can is mechanically integrated to one of the cryostat valve boxes for space management. Cryostat Valve/Distribution Box Control valves (enclosed in a valve distribution box, for each circuit, Fig. 5c) are added to each of the twelve legacy A1900 cryostats for better controllability and flow isolation (when required). Using bayonet style connections for positive isolation of the cryostats during maintenance is very difficult at this location for the limitation of space. Instead, flange style connections (along with vacuum breaks) are utilized at the cryostats. Retractable metallic expansion joints are used at the cryostat interface to provide room for maintenance.

4 Summary Integration of the NSCL legacy system into the FRIB era system was a complex task due to lack of documentation, unsupported and specialized equipment, organic evolution of NSCL system for the last four decades, etc. However, this also presented a great learning opportunity for the staff and students. The re-configured cryogenic distribution system has been designed, fabricated, installed, and commissioned entirely by the FRIB cryogenics department staff, following an aggressive project schedule. The demolition and reconfiguration of the existing infrastructure (e.g. cryogenic distribution lines, distribution boxes, and support structures) were started in December 2020, and the installation of preliminary systems was started by March 2021, and the commissioning of the reconfigured cryogenic distribution started in December 2021. The system is ready to support the first FRIB experiment in May 2022. Acknowledgments. This material is based upon work supported by the U.S. Department of Energy Office of Science under Cooperative Agreement DE-SC0000661, the State of Michigan and Michigan State University. Michigan State University designs and establishes FRIB as a DOE Office of Science National User Facility in support of the mission of the Office of Nuclear Physics.

References 1. Gade, A., Sherrill, B.M.: NSCL and FRIB at Michigan State University: Nuclear science at the limits of stability. Phys. Scr. 91(5), 053003 (2016) 2. Hausmann, M., et al.: Design of the Advanced Rare Isotope Separator ARIS at FRIB. Nucl. Instrum. Methods Phys. Res., Sect. B 317, 349–353 (2013) 3. McCartney, A., et al.: Cryogenic system upgrade for the National Superconducting Cyclotron Laboratory. AIP Conf. Proc. 613(1), 207–212 (2002) 4. McCartney, A.H., Laumer, H.L., Jones, S.A.: Operational experience of the upgraded cryogenic systems at the NSCL. AIP Conf. Proc. 1218(1), 11–17 (2010) 5. Hasan, N., et al.: Design, fabrication, and installation of the cryogenic distribution system from FRIB target and fragment pre-separator superconducting magnets. In: IOP Conference Series: Materials Science and Engineering, p. 756 (2021)

Sub-Kelvin Cryogenics for Experimental Cosmology Andrew J. May1,2,3(B) 1

2

Department of Physics and Astronomy, University of Manchester, Manchester M13 9PL, UK [email protected] Accelerator Science and Technology Centre, STFC Daresbury Laboratory, Keckwick Lane, Warrington WA4 4AD, UK 3 Cockcroft Institute, Keckwick Lane, Warrington WA4 4AD, UK

Abstract. As the forthcoming generation of Cosmic Microwave Background observatories have moved towards the use of large format detector arrays operating ⎩ κ −1 Ps Ps Ps κ +1

375

(8)

The discrete Reynolds equation can be solved by Newton iteration, and the pressure field distribution in the gas film can be obtained. The dimensionless load can be obtained from the integral form of the following pressure field.      L/R  2π Fx cos ϕ =− d ϕd λ (9) (P − 1) sin ϕ Fy 0 0  (10) W = Fx2 + Fy2 The dimensionless stiffness is calculated by K=

dW dε

(11)

The total dimensionless mass flow is calculated by Qin = k1

n    min i

(12)

i=1

The friction coefficient is calculated by   L/R  2π   1 H ∂P f = + d ϕd λ 6H 2 ∂ϕ 0 0

(13)

In this paper, the performance calculation and analysis of the radial gas bearing in one of the helium turbines of the Comprehensive Research Facility for Fusion Technology (CRAFT) cryogenic system are carried out. The specific bearing structure and input parameters are shown in Table 1. Table 1. Bearing structure and related input parameters Parameters

Value

Unit

Bearing radius (D/2)

16

mm

Bearing length (L)

32

mm

Atmospheric pressure (patm )

1.013 × 105

Pa

Supply pressure (ps )

3pa /6 pa /9 pa

Pa (continued)

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Value

Unit

Supply orifice diameter

0.3

mm

Number of orifices per row

8

\

Helium viscosity (μ)

1.98 × 10–5

Pa•s

Isentropic exponent (κ)

1.67

\

Temperature (T)

293

K

Rotating speed (ω)

100,000

rpm

Fig. 2. Gas film pressure distribution (Ps = 5, ε = 0.5)

3 Results and Discussion Figure 2 shows the gas film pressure distribution cloud diagram of the bearing under the working conditions of supply pressure of 6 bar and eccentricity ratio of 0.5. It can be seen that the bearing has a higher pressure at the orifice, and then gradually decreases to the periphery. Due to the existence of eccentricity ratio, there is also obvious asymmetry in the circumferential distribution of pressure. It is this non-circumferential asymmetry that enables the bearing to have load-carrying capacity. Comparing Fig. 1(a) and Fig. 2(b), it can also be seen that under the same conditions, the high-pressure area in the middle of the double row bearing is significantly larger than that of the single row bearing. Figure 3(a) shows the relationship between the dimensionless load and the eccentricity ratio of the bearing under different gas supply pressures. Overall, the bearing load increases significantly with the increase of eccentricity ratio. When the gas supply is low (Ps = 3), the load difference between the single row bearing and the double row bearing is not obvious. With the increase of the gas supply pressure, the double row bearing shows a more superior load-carrying performance.

Performance Analysis

377

Fig. 3. Bearing performance parameter curves

Figure 3(b) shows the relationship between the dimensionless stiffness and the eccentricity ratio of the bearing. When the eccentricity ratio is less than 0.5, increasing the gas supply pressure and increasing the number of orifices can improve the bearing stiffness. When the eccentricity ratio is greater than 0.5, its increasing effect on stiffness is weakened. It shows that when the eccentricity ratio is high, the dynamic pressure effect of the bearing dominates the bearing stiffness. From the relationship between dimensionless mass flow and eccentricity ratio shown in Fig. 3(c), it can be seen that the mass flow is mainly affected by the gas supply pressure and the number of orifices, while the change in eccentricity ratio has a relatively weak effect. Under the same gas supply pressure, the mass flow of the double row bearing is about twice that of the single row bearing. The friction coefficient characterizes the effect of the gas film on the resistance torque of the rotor, and its relationship with the eccentricity ratio is shown in Fig. 3(d). It can be seen that as the eccentricity ratio increases, the friction coefficient increases rapidly. At low eccentricity ratio, the influence of gas supply pressure and the number of orifices on the friction coefficient is not obvious. At high eccentricity ratio, the increased gas supply pressure and the number of orifices lead to an increase in the friction coefficient.

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4 Conclusion In this paper, the FEM is used to establish a computational model for the helium lubricated hydrostatic bearing used in a CRAFT cryogenic turbine. Through the analysis of the bearing performance, it can be seen that under the same supply pressure, the double row orifices bearing shows certain advantages in load capacity and stiffness than the single row orifices bearing. However, the gas consumption of the double row orifices bearing also increases significantly. Considering that the gas consumption is still within the acceptable range, the cryogenic turbine can be recommended to use double row orifices hydrostatic gas bearings. Acknowledgments. The project is supported by Comprehensive Research Facility for Fusion Technology Program of China under Contract No. 2018-000052-73-01-001228.

References 1. Zhang, X.-B., Chen, J.-Y., Yao, L., Huang, Y.-H., Zhang, X.-J., Qiu, L.-M.: Research and development of large-scale cryogenic air separation in China. J. Zhejiang Univ. Sci. A 15(5), 309–322 (2014) 2. Choi, C.H., et al.: Helium refrigeration system for the KSTAR. Fusion Eng. Des. 81(23–24), 2623–2631 (2006) 3. Gao, Q., Chen, W., Lu, L., Huo, D., Cheng, K.: Aerostatic bearings design and analysis with the application to precision engineering: state-of-the-art and future perspectives. Tribol. Int. 135, 1–17 (2019) 4. Wang, X., Xu, Q., Wang, B., Zhang, L., Yang, H., Peng, Z.: Effect of surface waviness on the static performance of aerostatic journal bearings. Tribol. Int. 103, 394–405 (2016) 5. Faria, M.T.C., Andres, L.S.: On the numerical modeling of high-speed hydrodynamic gas bearings. J. Tribol.-T Asme. 122(1), 124–130 (2000) 6. Zhang, P., Chen, Y., Liu, X.: Relationship between roundness errors of shaft and radial error motions of hydrostatic journal bearings under quasi-static condition. Precis. Eng. 2018(51), 564–576 (2017) 7. Miyatake, M., Yoshimoto, S.: Numerical investigation of static and dynamic characteristics of aerostatic thrust bearings with small feed holes. Tribol. Int. 43(8), 1353–1359 (2010) 8. Du, J., Zhang, G., Liu, T., To, S.: Improvement on load performance of externally pressurized gas journal bearings by opening pressure-equalizing grooves. Tribol. Int. 73, 156–166 (2014)

Finite Element Analysis on Scroll of a Scroll Compressor for Electric-Driven Vehicle Air-Conditioner Luo Yuchen(B) , Li Zhun, and Li Qiang National Institute of Metrology, Chaoyang, Beijing, China [email protected]

Abstract. A three-dimensional model of a new energy vehicle scroll disc is established by Solidworks. Comparing with the deformation and stress of two scrolls before engagement, the deformation and stress distribution of two scroll teeth in meshing state under gas force, temperature, inertia constraints and multi-field coupling loads are analyzed, using Ansys simulation software. The results show that the temperature load field is the most important load field affecting the deformation of the scroll teeth; Under the coupling field, the top deformation of the stationary vortex teeth is the largest, and the maximum stress appears at the root of the moving vortex teeth, and the maximum stress of the coupling field is not the superposition of each load stress. Compared with the single analysis of multi-field loads acting on the two scrolls, the stress and deformation of the assembled scroll increased by 8.63% and 3.63% respectively, while the stress and deformation of the static scroll decreased by 17.2% and 20.38% respectively. Keywords: New Energy Vehicle · Scroll Compressor · Scroll Plate · Finite Element Analysis

1 Introduction At present, the compressors mentioned above are more vortex compressors. At the same time, with the continuous improvement of manufacturing technology, scroll compressors are widely used in the fields of air conditioning and refrigeration, power engineering and transportation [1–3]. Many scholars at home and abroad have made a lot of theoretical achievements in the research of its dynamic characteristics [4–9]. However, in terms of simulation research, most of the current simulation analysis of the scroll compressor of electric vehicles still separate each component of the scroll compressor. For the scroll plate, the deformation under the joint action of coupling after the assembly of dynamic and static scroll bodies under actual working conditions is different from that under separate analysis, At present, some literature have proved through experiments that there will be interference between the scroll disks of the scroll compressor, especially in the case of high temperature and high pressure, there may be serious consequences of gluing [10]. Therefore, it is particularly necessary to simulate the dynamic and static scroll disks as a whole and study their deformation and stress changes under various load fields and coupling states. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 379–386, 2023. https://doi.org/10.1007/978-981-99-6128-3_48

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2 Preconditioning and Parameters The structural parameters of the scroll plate are shown in Table 1. Table 1. Structural parameters of scroll plate Parameter name

sign

Numerical value

Involute type

/

Circular involute high degree curve

Number of vortices

N

1.7

Scroll tooth height

h

19.5 (mm)

Scroll tooth thickness

t

3.26–4.37 (mm)

Exhaust angle degrees

θ

270°

The meshing positions of the dynamic and static scroll plates at the beginning of exhaust are selected for analysis. The operating conditions and relevant load parameters are shown in Table 2. Table 2. Operating conditions and other parameters Parameter name

sign

Numerical value

Suction temperature

T1

16.6 °C

Exhaust temperature

T4

94.24 °C

Suction chamber pressure

Ps

0.3727 MPa

Intermediate working pressure

/

0.4701 MPa

Exhaust chamber pressure

Pd

1.4915 MPa

Speed

n

2500 r/min

Inertial load

Frx

121.75 N

Isentropic index

k

1.15

Compression ratio

/

2.1

Cryogen

/

R134a

The scroll disk of the scroll compressor is modeled by Solid-Works 3D modeling software, and then imported into the corresponding calculation module by using the recognizable interface of ANSYS Workbench. Select the free meshing method in mechanical structure meshing, select solid187 unit to divide the network, and set the minimum unit size of 1mm. The divided entity is shown in Fig. 1. The number of units is 239252 and the number of nodes is 372225.

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Create new materials in ANSYS Workbench engineering database. The material of scroll plate is 4032t6 aluminum alloy, and its material characteristic parameters are shown in Table 3. Table 3. Material parameters Parameter name

Numerical value

Modulus of elasticity

79 GPa

Poisson’s ratio

0.33

Density

2.68 g/cm3

Coefficient of thermal expansion

19.4E-6/°C

Specific heat capacity

864 J/Kg°C

Constraint loading: set the degree of freedom around the end plate of the moving scroll plate along the height direction of the scroll teeth to zero; Set the degree of freedom in the plane direction of the inner wall of the bearing hole on the back of the moving scroll plate to zero; Set the degree of freedom of the top of the bearing hole of the moving scroll disk along the height direction of the scroll teeth to zero; Since the static scroll disk of the scroll compressor is fixed on the frame of the scroll compressor under the constraint of the end face bolts, axial constraints are imposed on the end face of the static scroll disk, and radial and axial constraints are imposed on the bolt holes.

3 Calculation and Solution Type Since the inertial load field has little influence on the scroll plate, the influence of the inertial load field on the scroll plate before and after assembly is no longer listed and analyzed separately in the calculation, but the inertial load is taken into account in the coupling field after assembly. The specific analysis types are shown in Table 4. Table 4. Analysis type Gas load Thermal analysis Thermal stress analysis Gas force analysis Inertial force analysis Coupled field analysis

Temperature load √

Inertial load

√

√

√ √

constraint condition

√ √

√

√

√

√

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3.1 Data Analysis of Scroll Disk Under Thermal Load In the result analysis, in order to better compare the differences between the simulation results by using the two methods of overall simulation analysis and individual part simulation analysis, this paper will additionally introduce the deformation nephogram of the heating load field of the two scroll disks in the separate analysis. The two simulation analysis results of the other load fields are limited by the space requirements, and only the result data will be reflected in Fig. 1.

Fig. 1. Cloud diagram of scroll plate deformation under thermal load during overall analysis

In terms of scroll disk shape variable: as can be seen from Fig. 2, when the overall simulation analysis is adopted for the heating load field of the two scroll disks, the maximum deformation position occurs at the tooth top of the scroll tooth head of the static scroll disk, and the shape variable is about 50.1 μm. The change trend decreases along the direction of the vortex tooth height and from the vortex center to the outside; At the same time, the maximum deformation position of the moving scroll disk also occurs at the tooth top of the scroll tooth head, and the shape variable is about 33.5 μm. The change trend is the same as that of the static scroll.

Fig. 2. Cloud diagram of scroll plate deformation under thermal load when analyzed separately

Figure 2 shows the deformation nephogram of the heating load field when the two scroll disks are simulated and analyzed separately. By comparing Fig. 2 and Fig. 3, it can be seen that compared with the heating load of the two scroll disks analyzed separately, the shape variable of the dynamic scroll disk basically remains unchanged when using

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the overall simulation analysis, while the displacement shape variable of the static scroll disk decreases by 22.1% on average year-on-year. In terms of the stress on the scroll plate: according to the simulation results, the maximum stress occurs at the junction between the inner side of the bearing hole of the moving scroll plate and the end plate of the moving scroll plate during the overall simulation analysis of the heating load of the two scroll plates. The stress is about 176.52 MPa, and the change trend gradually decreases along the outer side of the vortex center and the height of the scroll teeth; At the same time, the maximum stress of the static scroll plate occurs at the tooth root of the scroll tooth head, and the stress is about 120.83MPa, and the change trend is the same as that of the dynamic scroll plate; Compared with the single analysis of the thermal load of the two scroll disks, the average stress of the assembled moving scroll disk is reduced by 30.2% year-on-year, and the average stress of the static scroll disk is reduced by 21.9% year-on-year. The specific changes of the thermal load stress and displacement of the dynamic and static scroll plates by the two analysis methods are shown in Table 5. Table 5. Table of stress and displacement variation of scroll plate under heating load Maximum stress Maximum stress Maximum on moving disc on stationary displacement and plate deformation of moving disc

Maximum displacement and deformation of stationary disk

Separate analysis 252.8 MPa

154.6 MPa

33.5 μm

64.3 μm

Overall analysis

176.5 MPa

154.6 MPa

33.4 μm

50.1 μm

Change ratio

−30.2%

−21.9%

−0.01%

−22.1%

3.2 Data Analysis of Vortex Disk Under Gas Load Stress of scroll plate: according to the simulation results, when the two scroll plates are subjected to gas load for overall simulation analysis, the maximum stress occurs at the tooth root of the scroll tooth head of the moving scroll plate, and the stress is about 44.3 MPa. At the same time, the maximum stress of the static scroll plate also occurs at the tooth root of the scroll tooth head, and the stress is about 38.9 MPa, and the change trend is the same as that of the moving scroll plate; Compared with the single analysis of the gas load on the two scroll disks, the stress on the assembled dynamic scroll disk increased by 6.25%, while the stress on the static scroll disk decreased by 1.19%. Compared with the single analysis of the gas load on the two scroll disks, the deformation displacement of the assembled dynamic scroll disk increased by 4.14%, while the deformation displacement of the static scroll disk decreased by 2.63%. The main reason is that there are multiple exhaust holes in the center of the static scroll plate, which weakens the restraint effect of the end plate on the root of the tooth head, resulting in that the tooth head shape variable of the static scroll plate is slightly larger than that of the dynamic scroll plate. The specific changes of stress and displacement of

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dynamic and static vortex bodies under gas load by the two analysis methods are shown in Table 6. Table 6. Variation of stress and displacement of scroll plate under gas load Maximum stress Maximum stress Maximum on moving disc on stationary displacement and plate deformation of moving disc

Maximum displacement and deformation of stationary disk

Separate analysis 41.73 MPa

39.37 MPa

6.94 μm

7.59 μm

Overall analysis

44.34 MPa

38.99 MPa

7.30 μm

7.57 μm

Change ratio

+6.25%

−1.19%

+4.14%

−2.63%

3.3 Stress and Deformation Analysis of Vortex Disk Under Multi-field Coupling

Fig. 3. Cloud diagram of stress distribution of vortex disk under coupling field during overall analysis

In terms of stress on the scroll plate: according to the results of the overall analysis method in Figure 4, the maximum stress occurs at the tooth root of the scroll tooth head of the moving scroll plate, and the stress is about 176.1 MPa. At the same time, the maximum stress of the static scroll plate also occurs at the tooth root of the scroll tooth head, and the stress is about 130.87 MPa, and the change trend is the same as that of the moving scroll plate; Compared with the single analysis of the combined action of multiple field loads on the two scroll plates, the stress on the moving scroll plate in the overall analysis increased by 8.63%, while the stress on the static scroll plate decreased by 17.2% . The main reason is that the two scroll plates are greatly affected by the gas load field, resulting in reverse changes in the stress distribution and displacement of the two scroll plates in multiple load fields. The specific changes of stress and displacement of

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dynamic and static vortex bodies affected by multiple load fields by the two analysis methods are shown in Table 7. Table 7. Stress and displacement variation of scroll plate under multiple load fields Maximum stress Maximum stress Maximum on moving disc on stationary displacement and plate deformation of moving disc

Maximum displacement and deformation of stationary disk

Separate analysis 162.1 MPa

157.98 MPa

35.8 μm

67.2 μm

Overall analysis

176.1 MPa

130.87 MPa

37.1 μm

53.5 μm

Change ratio

+8.63%

−17.2%

+3.63%

−20.38%

4 Conclusion Through the comparison of the analysis results, it can be seen that the load field that has the greatest influence on the deformation of scroll teeth is the temperature load field; The maximum deformation position appears at the top of the scroll teeth of the static scroll disk, and the corresponding maximum stress is about 176.52 MPa; Under the action of coupling field, the top deformation of static vortex tooth is the largest, and the maximum stress appears at the root of dynamic vortex tooth, and the maximum stress of coupling field is not the superposition of load stress; Compared with the single analysis of the combined action of multiple field loads on the two scroll disks, when the overall analysis method is used to analyze the two scroll disks, the stress on the moving scroll disk increases by 8.63% and the shape variable increases by 3.63%; The stress on the static scroll plate decreased by 17.2%, and the shape variable decreased by 20.38%.

References 1. Peng, B., Zhu, B.: Effect of scroll profile on performance of scroll compressor. Fluid Mach. 44(6), 17–23 (2016) 2. Liu, Z., Zhao, Z., Li, C.: Analysis of crank pin anti rotation mechanism of scroll compressor. J. Lanzhou Univ. Technol. (4), 53–55 (2006) 3. Zhou, Y., Zhang, X., Liu, Z.: Development trend of scroll compressor technology. Refrig. Air Conditioning 17(7), 69–72 (2017) 4. Elson, J., et al.: Scroll technology: an overview of past, present and future developments. In: 19th International Compressor Engineering Conference at Purdue, vol. 1204, pp. 1–8 (2008) 5. Li, F.: Finite element analysis of scroll disk of new oil-free scroll compressor. Lanzhou University of technology (2016) 6. Peng, B., Li, Y., Zhao, S.: Strain analysis and experimental verification of scroll plate of oil-free scroll compressor. Fluid Mach. 46(7), 1–8 (2018) 7. Qiu, H.: improved design of dynamic balance performance of scroll compressor transmission system. Mech. Strength 39(2), 474–478 (2017)

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8. Wang, X.: Study on multi-objective optimization of dynamic balance of transmission system of scroll compressor. Fluid Mach. 44(12), 33–37 (2016) 9. Huang, H., Zhang, C.: Simulation study on dynamic balance of vortex compressor with motion pair clearance. Vib. Shock 35(5), 125–130 (2016) 10. Liu, C., Luo, X.: Modal analysis and harmonic response analysis of connecting rod of six cylinder compressor based on ANSYS. Mech. Des. Manuf. (3), 26–29 (2013)

Thermodynamic Analysis of Single-Stage Mixed-Refrigerant Joule-Thomson Cycle at Cooling Temperature from 100 to 180 K Yaxue Wei1 , Jinxing Wu1 , Qinglu Song1(B) , Dechang Wang1 , Dandan Sun1 , and Haocheng Wang2 1 College of Mechanical and Electrical Engineering, Qingdao University, Qingdao 266071,

China [email protected] 2 Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Beijing 100190, China

Abstract. Rapid developments in clean energy, life science, and aerospace engineering have put forward more requirements for refrigeration technology with temperature below 180 K. Mixed-refrigerant Joule-Thomson (MRJT) refrigeration technology has distinct advantages in this temperature range. In this paper, the thermodynamic analysis of the Linde Hampson cycle and single-stage separation cycle at the refrigeration temperature from 100 to 180 K was carried out, and the energy and exergy models were established. Nitrogen, methane, tetrafluoromethane, ethane, propane, iso-butane, and iso-pentane were employed as the working fluids. The results showed that the performance of Linde-Hampson cycle is close to those of the single-stage separation cycle at the cooling temperature of 180 K. The COP of the single-stage separation cycle is 0.193 with 33.60% exergy efficiency at the temperature of 100 K, and in the Linde-Hampson cycle and single-stage separation cycle, the recuperator exergy loss accounts for 28.45% and 23.45% respectively, exergy losses from recuperator and condenser are the main components in the system. Keywords: Linde-Hampson cycle · Single-stage separation cycle · Thermodynamic analysis

1 Introduction Mixed-refrigerant Joule-Thomson refrigeration cycle (MRJT) has made great progress over the past years. Among the various types of MRJT cycles, the Linde-Hampson cycle and the single-stage separation cycle have advantages over other forms of refrigeration systems, such as: reliable operation, simple system structure, and low manufacturing cost [1], so they are widely used in the cooling temperature range from 100 to 180 K. To improve the performance of the MRJT, many studies have been conducted on the optimization of mixtures and their compositions [2], the thermal cycle of the system [3], and the pinch point of recuperator [4]. Gong et al. [5, 6] studied the thermal performance © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 387–393, 2023. https://doi.org/10.1007/978-981-99-6128-3_49

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of MRJT single-stage cycle and single-stage separation cycle, pointing out that two were close in efficiency under their optimal conditions, with the MRJT single-stage cycle being slightly better. Wang et al. [7] investigated the Joule-Thomson cycle and the Joule-Brayton cycle. They found that under nonideal conditions, Joule-Thomson cycle’s exergy efficiency and volumetric cooling capacity were better than Joule-Brayton cycle between 100 and 180 K, but inferior to Joule-Brayton cycle at 80 K. The primary goal of this study is to analyze and compare the thermodynamic performance of Linde-Hampson cycle and single-stage separation cycle at cooling temperatures from 100 to 180 K, and evaluate temperature matching of hot and cold fluid in recuperators. For the single-stage separation cycle, the cold end temperature’s (T 5 ) permitted range and the effect of the separation fluid’s supercooling on system performance were also investigated.

2 Methodology 2.1 Description of the Linde-Hampson Cycle and the Single-Stage Separation Cycle Figure 1 (a) is the schematic diagram of the Linde-Hampson cycle (Cycle A). The MRJT single-stage separation cycle (Cycle B) is shown in Fig. 1 (b).

Fig. 1. MRJT single-stage cycles. (a) Linde-Hampson cycle. (b) Single-stage separation cycle.

2.2 Refrigerant Screening As shown in Fig. 2, the h peaks of N2 , CH4 , CF4 , C2 H6 , C3 H8 , iC4 H10 and iC5 H12 form a good relay, so the N2 -CF4 -HC mixed-refrigerant is used to achieve a high exergy efficiency.

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Fig. 2. Isothermal throttling effect (h) of the components in MRJT

2.3 Energetic and Exergetic Model Several assumptions and constraints of the model are shown as follows: 1. All systems operate in steady state. 2. The heat loss and pressure drop in heat exchangers and pipelines are ignored. 3. The compressor discharge pressure does not exceed 2.0 MPa, the exhaust temperature does not exceed 393 K, and the pressure ratio is less than 8 [8]. 4. Compressor adiabatic efficiency of 75% [9]. 5. Condenser outlet temperature and ambient temperature of 300 K. In actual projects, the heat exchanger area will increase with the reduction of pinch point temperature difference, so we set a lower limit of 3 K. The parameters setting of the compressor are to ensure that the compressor operates under good working conditions.

3 Results and Discussion 3.1 Energetic and Exergetic Performance Analysis Tables 1 and 2 show the mixture composition and operating pressures for Cycle A and Cycle B from 100 to 180 K, respectively. Figure 3 shows the variation of COP and qv for Cycle A and Cycle B at cooling temperatures from 100 to 180 K. Cycle B has a higher volumetric cooling capacity than Cycle A at 180 K, there is no difference in the COP of the two cycles, and at 160 to 100 K, Cycle B has a higher COP than Cycle A. Figure 4 depicts the exergy loss fraction for each component of the MRJT at 100 to 180 K. Overall, the proportion of exergy loss at the compressor and evaporator in the two cycles does not change much in each temperature zone with different working fluids, while the proportion of exergy loss at the recuperator is relatively large at low temperatures, which is the primary cause of Cycle A and Cycle B’s performance decay in the low temperature zone. At cooling temperatures from 100 to 160 K, Cycle B has better exergy efficiency than Cycle A.

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Table 1. Mixed-refrigerant composition and operating conditions of the Linde-Hampson cycle (Cycle A) Items

180 K

N2 (%, mole)

0

160 K

140 K

120 K

100 K

2.69

8.91

23.23

32.95

CH4 (%, mole)

31.10

40.40

47.48

38.45

27.29

CF4 (%, mole)

21.41

10.10

8.91

11.61

5.66

C2 H6 (%, mole)

19.90

10.44

7.01

7.23

5.26

C3 H8 (%, mole)

9.60

20.20

13.81

6.71

10.37

iC4 H10 (%, mole)

8.27

10.10

2.59

2.39

13.21

iC5 H12 (%, mole)

9.72

6.06

11.29

10.39

5.26

pl (MPa)

0.55

0.50

0.36

0.35

0.50

ph (MPa)

1.84

1.80

2.00

1.57

1.80

Table 2. Mixed-refrigerant composition and operating conditions of the single-stage separation cycle (Cycle B) Items

180 K

N2 (%, mole)

0

160 K

140 K

120 K

100 K

2.56

8.05

25.45

38.18

CH4 (%, mole)

28.89

38.46

48.31

34.55

28.18

CF4 (%, mole)

18.89

15.38

8.05

10.91

9.09

C2 H6 (%, mole)

20.00

12.82

16.10

14.09

8.18

C3 H8 (%, mole)

15.56

16.67

3.22

4.09

6.36

iC4 H10 (%, mole)

10.00

7.69

8.21

1.82

0.91

iC5 H12 (%, mole)

6.67

6.41

8.05

9.09

9.09

pl (MPa)

0.57

0.49

0.45

0.50

0.50

ph (MPa)

1.62

1.80

1.67

1.67

1.61

3.2 MRJT Recuperative Process and Supercooling Analysis The T-Q diagram of Cycle A recuperative process at cooling temperatures (T A6 ) of 180 K and 100 K is shown in Fig. 5. The horizontal coordinate is the dimensionless amount of heat flow. At the cooling temperature of 180 K, the temperature difference (T ) in recuperator is small, and when the cooling temperature is 100 K, T at the hot end of recuperator has a significant tendency to increase, but the temperature matching of the hot and cold fluids in recuperator can be effectively improved by optimizing the mixture composition. The cold end temperature T 5 of HX1 is an important parameter for determining the position of the separation fluid sink into the low pressure return, and the value of T 5 directly determines the size of HX1 and HX2, taking the cooling temperature (T B8 ) of

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Fig. 3. COP and qv of MRJT at 100 to 180 K

Exergy Loss Fraction (%)

100

η Eevap ETV EHX Econd Ecomp

80 60 40 20 0

180

160

140

120

100

Cooling tempertature (K)

Exergy Loss Fraction (%)

100

Cycle B η Eevap ETV Emixer EHX Econd Ecomp

80 60 40 20 0

180

160

140

120

100

Cooling tempertature (K)

Fig. 4. Exergy loss fraction of MRJT at cooling temperatures 100 to 180 K

100 K as an example, the feasibility range of T 5 in Cycle B is shown in Fig. 6, the upper and lower limits of cycle B’s T 5 are analyzed, when the T 5 temperature is too low, the pinch piont appears at the cold end of HX1 and no pinch piont appears at the hot end of HX2; when the T 5 is too high, the pinch piont appears at the hot end of HX2 and no pinch piont appears at the cold end of HX1.

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Fig. 5. T - Q diagrams of recuperators in Cycle A at cooling temperatures (T A6 ) 180 K and 100 K

Fig. 6. Feasibility range of T 5 in Cycle B at cooling temperature (T B8 ) 100 K

4 Summary The thermodynamic comparison of the Linde-Hampson cycle and the single-stage separation cycle from cooling temperature 100 to 180 K was conducted. Following are the conclusions: When the cooling temperature is 180 K, the difference in COP, qv and exergy efficiency between Cycle A and Cycle B is small. As the cooling temperature is lower, the advantage of Cycle B is more obvious. At the cooling temperature of 100 K, the COP of Cycle B is 0.193 and the exergy efficiency is 33.6%, and the good mixture

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composition improves the cycle performance of the system. For the Cycle B, the temperature T 5 has a reasonable range of values, otherwise it will cause temperature cross in the heat exchanger and deteriorate the system efficiency. Acknowledgements. This work was supported by National Natural Science Foundation of China (Grant No. 52106016), Natural Science Foundation of Shandong Province, China (Grant No. ZR2021QE219), and Key R&D Program of Shandong Province, China (Grant No. 2022CXGC020901).

References 1. Rogala, Z.: Application of precooling stage in MR JT cryocoolers. Cryogenics 121, 103395 (2022) 2. Lakshmi Narasimhan, N., Venkatarathnam, G.: Effect of mixture composition and hardware on the performance of a single stage JT refrigerator. Cryogenics 51(8), 446–451 (2011) 3. Geng, H., Cui, X., Weng, J., She, H., Wang, W.: Review of experimental research on Joule– Thomson cryogenic refrigeration system. Appl. Therm. Eng. 157, 113640 (2019) 4. Aspelund, A., Berstad, D.O., Gundersen, T.: An extended pinch analysis and design procedure utilizing pressure based exergy for subambient cooling. Appl. Therm. Eng. 27(16), 2633–2649 (2007) 5. Gong, M., Wu, J., Luo, E., Qi, Y., Zhou, Y.: Study of the single-stage mixed-gases refrigeration cycle for cooling temperature-distributed heat loads. Int. J. Therm. Sci. 43(1), 31–41 (2004) 6. Gong, M., Wu, J., Luo, E.: Performances of the mixed-gases Joule-Thomson refrigeration cycles for cooling fixed-temperature heat loads. Cryogenics 44(12), 847–857 (2004) 7. Wang, H.C., Chen, G.F., Dong, X.Q., Zhao, Y.X., Guo, H., Gong, M.Q.: Performance comparison of single-stage mixed-refrigerant Joule-Thomson cycle and pure-gas reverse Brayton cycle at fixed-temperatures from 80 to 180 K. Int. J. Refrig 80, 77–91 (2017) 8. Sun, S., Guo, H., Gong, M.: Thermodynamic analysis of single-stage compression air-source heat pumps with different recuperation ways for large temperature lift. Int. J. Refrig 108, 91–102 (2019) 9. Guo, H., Gong, M., Qin, X.: Performance analysis of a modified subcritical zeotropic mixture recuperative high-temperature heat pump. Appl. Energy 237, 338–352 (2019)

A Cryogenic System for Measuring the Thermally Stimulated Depolarization Current Jixiang Yan1,2 , Rongjin Huang1,2(B) , Peng Jia1,2 , Huiming Liu1 , Yaran Shi1,2 , Shiyong Xie1,2 , Laifeng Li1,2 , and Yuan Zhou1 1 State Key Laboratory of Technologies in Space Cryogenic Propellants,

Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Beijing 100049, People’s Republic of China [email protected] 2 University of Chinese Academy of Sciences, Beijing 100049, People’s Republic of China

Abstract. The insulating material is a crucial part of superconducting magnets and the dielectric mechanisms of insulating materials are still unclear at cryogenic temperatures. Space charge distribution can partially characterize the dielectric properties of materials. Thermally stimulated currents (TSC) method was adopted to analyze space charge distribution in many cases. In this paper, a cryogenic system for measuring space charge characteristics has been established based on the thermally stimulated currents theory. Besides, a measurement and control system was designed. The whole system contained many good traits. Microscopic parameters of charged particles in various materials can be tested in the platform, especially for the measurement of defect state density distribution of insulating dielectrics. A thermal and reliability experiment of the TSC test platform was carried out. The experiment verified the reliability of the whole experimental system. Keywords: TSC · Space Charge Distribution · Cryogenic Temperature · Insulating Materials

1 Introduction In superconducting magnets, insulating materials such as epoxy often determined the reliability and stability of the entire system. The traditional dielectric insulating behavior at room temperatures had been well explained by the avalanche model and the collisional ionization breakdown model. However, we have experimentally found that epoxy resin has abnormal insulating behavior at low temperatures. Since space charge has an important influence on the electrical behavior of dielectrics and space charge parameters are often measured by using the thermally stimulated current (TSC) method, it is necessary to build an efficient TSC test platform to measure the space charge parameters of insulating materials to establish a TSC database of epoxy resin and predict its breakdown behavior at low temperatures. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 394–401, 2023. https://doi.org/10.1007/978-981-99-6128-3_50

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TSC is a charging and discharging process. Three types of polarization (Fig. 1) can be analyzed from the macroscopic thermally stimulated current spectrum [1]. H. Frei and G. Groetzinger used an external electric field to measure the charge amounts of inner electrets [2]. C. Bucci and R. Fieschi first proposed a complete theory for the dipole polarization process and used the TSC method to study the causes of TSC in ionic crystals [2]. A. Mandowski used the TSC method to numerically study the kinetics of trapping and recombination, predicting the possibility of occurrence of an additional displacement peak in TSC curves [3]. Smaoui, H et al. used the TSC technique to investigate dielectric relaxations and molecular mobility in nanocomposites based on epoxy resin modified by nanoparticles [4]. The TSC method was commonly used to characterize the microscopic parameters of insulating materials, such as trapped charge type, particle activation energy, and dielectric polarization relaxation time [5, 6]. In this study, a cryogenic system with a thermal switch has been established for measuring the thermally stimulated depolarization current. The system design and the cryostat’s performance are investigated and presented in this paper.

Fig. 1. Schematic diagram of TSC.

2 System Design The energy equations of TSC are derived based on the Maxwell model. In the heating process, the heating rate can be represented by β, and the temperature change with time is given by the following expression: T = T0 + βt

(1)

Under the condition of constant temperature relaxation and according to Debye equation, we have: dp(t) dt

t

= − pτ0 e− τ = − τ1 p(t)

(2)

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differentiate Eq. (1) with respect to time and combine it with Eq. (2), take τ = τ0 e kT into consideration, we have:   T H p(T ) = p0 (T0 )exp − τ01β T0 e− kT dT (3) If we only consider the TSC caused by traps, we can get the following expression [7]:

 H − ITSDC (T ) = A exp − kT

1 τ0 β

 exp − T0

T

H  kT

  dT

(4)

where the pre-exponential factor A is a constant, H is the activation energy, k is the Boltzmann constant, T0 is the initial temperature and T is the set temperature, β is the rate of heating, and τ0 is the relaxation time at T0 . According to the TSC method and theoretical model introduced above, the design of the TSC system can be divided into two parts: 1. Design of the cryogenic system; 2. Design of measurement and control system. 2.1 Design of Cryogenic System The cryostat (Fig. 2) should include an experimental bench, vacuum, cable outlets, etc. In this platform, the upper and lower copper electrodes were screwed onto two polytetra fluoroethylene (PTFE) insulating plates. Brass sample stage has a H shape with a certain heat capacity. A vacuum chamber made of 304 stainless steel that can provide a well electromagnetic shield for measurement was designed (Fig. 2(b)). A special thermal switch was designed. This thermal switch (Fig. 3) can realize the on-off state of the thermal and electrical conduction. In the lower part of the vacuum cylinder, a cold chamber was made to temporarily store liquid nitrogen (LN2). The thermal switch is connected to the LN2 through three soft connections made by OFHC (oxygen-free high-purity copper) with a diameter of 8 mm.

Fig. 2. Physical structure of the system.

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Fig. 3. Construction of thermal switch and the sample stage: (1) stainless steel base; (2) PTFE insulation; (3) copper conjunction; (4) M6 threaded rod; (5) O ring dynamic seal; (6) 8 mm copper braids; (7) H-shaped brass sample stage; (8) PTFE bracket.

2.2 Design of Measurement and Control System This article used a Keithley 6517b electrometer for current measurements. The cables used in this paper were triaxial cables with a double-layer shield. A BNC interface was used to establish connections from the vacuum to the atmosphere to minimize the influence of external factors. The test circuit is shown in Fig. 4.

Fig. 4. (1) thermal bridge; (2) sample; (3) electrode; (4) heater and thermometer; (5) experimental bench; (6) stainless steel cylinder; (7) thermal switch; (8) ceramic feedthrough; (9) DC source (DW-P104-2ACF2); (10) mechanical switch; (11) electrometer (Keithley 6517b); (12) copper cable; (13) high voltage divider; (14) oscilloscope; (15) copper braid; (16) cold cavity; (17) LN2 transport tube.

Compared to the liquid nitrogen immersion method (Fig. 5) to perform TSC experiments [8], the sample in this system would not be exposed to refrigerants, which could probably introduce a certain dielectric polarization electric field and increases the error. Townsend discharge would occur at low pressure and limit the polarization voltage of the

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experiment. The dry way of using copper solid to conduct cooling capacity can address this problem.

Fig. 5. Difference between other’s work and our work [8].

Besides, the upper flange of this system can be transferred to an already-built system for a lower temperature TSC test using a GM refrigerator for cooling (Fig. 6). This system was designed with extra cutter flanges for the introduction of other interfaces such as BNC connectors.

Fig. 6. Flange that fits with an established system.

3 Experiment To verify the feasibility of the cryostat, this paper conducted a thermal experiment and a background noise test on the established TSC measurement system. A liquid nitrogen Dewar was used to introduce liquid nitrogen into the cold chamber. After the temperature reaches the target temperature, the thermal switch was lifted to separate the

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thermal bridge from the copper braid. A thermometer (PT100, calibrated by the cryogenic metrology station of the Institute of Physical and Chemical Technology, CAS) with a measurement range of 20 K-350 K and accuracy of 0.1 K was mounted on the surface of the sample (Fig. 7).

Fig. 7. Location of thermometer in the experimental bench.

4 Results and Conclusion The temperature curve of the cooling process is shown in Fig. 8. It can be seen from the figure that the sample temperature has good linearity in the initial heating process because the specific heat of the copper stage does not change much in this temperature region. Also, background current has been tested in shielded and unshielded mode (Fig. 9). It can be seen from the figure that a good shield can greatly reduce the noise from 1E-12 A to 1E-14 A. This system can reach around 145 K, which is sufficient for regular TSC tests. However, the cooling rate needs to be further increased to shorten the experimental time. In future work, this system will be improved and preliminary TSC measurement experiments will be conducted. Furthermore, a lower temperature region TSC test platform will be built to measure the TSC of epoxy resin.

Fig. 8. The temperature of the epoxy sample in the cool-down and heat-up process.

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Fig. 9. Noise current in shielded and unshielded mode.

Acknowledgments. This work was supported by the National Key R&D Program of China (Grant Nos:2017YFE0301504) and the Scientific Instrument Developing Project of Chinese Academy of Sciences (Grant Nos: YJKYYQ20180061).

References 1. Sawa, G., Kawade, M., Lee, D.C., Ieda, M.: Thermally stimulated current from polyethylene in high-temperature region. Jpn. J. Appl. Phys. 13(10), 1547–1553 (1974) 2. Liheng, W.: Thermal Stimulation Theory of Dielectric and Its Application. Science Press, 40–41(1988) 3. Mandowski, A.: Charge carrier trapping in a one-dimensional system. Dielectr. Electr. Insul. IEEE Trans. 8(6), 1007–1010 (2001) 4. Smaoui, H., Mir, L., Guermazi, H., Agnel, S., Toureille, A.: Study of dielectric relaxations in zinc oxide-epoxy resin nanocomposites. J. Alloy. Compd. 477(1–2), 316–321 (2009)

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5. Worsfold, P., Poole, C., Townshend, A., Miró, M.: Encyclopedia of Analytical Science. 3rd edn. Academic Press, 33–42(2019) 6. Glowacki, I., Jung, J., Ulanski, J., Rybak, A.: Conductivity Measurements in Polymer Science: A Comprehensive Reference. 2nd edn. Elsevier, 847–877(2012) 7. Smaoui, H., Arous, M., Guermazi, H., Agnel, S., Toureille, A.: Study of relaxations in epoxy polymer by thermally stimulated depolarization current (TSDC) and dielectric relaxation spectroscopy (DRS). J. Alloy. Compd. 489(2), 429–436 (2010) 8. Jain, D., Chandra, L.S.S., Nath, R., Ganesan, V.: Low temperature thermal windowing (TW) thermally stimulated depolarization current (TSDC) setup. Meas. Sci. Technol. 23(2), 025603 (2012)

Design and Analysis of Vacuum Calibration Apparatus at Cryogenic Temperature Shixian Wu1,2 , Zhengrong Ouyang1 , Lei Shi1 , Qiumin Meng1 , Xin Ai1 , Xuheng Chen1 , Dazhi Kuang1 , Peizhi Ding1 , and junjie Li1(B) 1 Hefei Institutes of Physical Science, Chinese Academy of Sciences, Hefei 230031, China

[email protected] 2 University of Science and Technology of China, Hefei 230026, China

Abstract. A vacuum calibration apparatus is presented which allows the calibration of vacuum in cryogenic temperature (down to ~ 4K) with instrumentation at room temperature. The GM cryocooler was used as the cold source, and the temperature uniformity of the calibration chamber was analyzed and optimized. A structure called Sniffer tube was used to keep the temperature of the measuring pipe above 100K in order to avoid the influence of gas condensation on the measuring pipe, and the influence of the cryosorption on the measuring pipe was analyzed. A method to calculate the vacuum of the calibration chamber is presented when the capacitance diaphragm gauge and hot cathode ionization gauge are used as standard gauges. The results show that the temperature difference between arbitrary points across the calibration chamber is less than 0.25K, and the cryopumping effects of the measuring pipe can be avoid by calibrating different gases, which meets the calibration requirements. Keywords: Cryogenic Vacuum Calibration · Calibration Apparatus · Numerical Analysis · Structure Design

1 Introduction As the temperature moves towards cryogenic temperature, the gas components in the vacuum change with temperatures [1]. Due to the influence of thermal transpiration and cryopumping, the vacuum calibration system at room temperature can no longer respond to the real situation at cryogenic temperature. In fields like space cryoengineering, atomic physics and some cryogenic materials (such as vacuum multilayer insulation), which are located at low temperature and its performance depends on vacuum degree, often cannot be further analyzed because of vacuum physical properties not clear at cryogenic temperature [2, 3]. In the previous experiments, only the vacuum at a specific temperature (such as liquid helium and liquid hydrogen) was measured [4, 5]. However, it is difficult to obtain the change of vacuum and the gas components in the continuous temperature range from room temperature to cryogenic temperature. Due to the gradual increase of the cooling capacity of GM cryocooler since the 1990s [6, 7], it is possible to design a cryogenic vacuum calibration apparatus based on GM refrigerator now. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 402–409, 2023. https://doi.org/10.1007/978-981-99-6128-3_51

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2 Structures and Instruments Consideration 2.1 Temperature Control In order to measure vacuum degree at 4–300K continuous temperature range, a GM Cryocooler with 1st stage capacity of 100K@70W and 2nd stage capacity of [email protected] is selected [8]. The thermal shield is split into two halves. The top of the shield is contacted to the straight pipe of the sniffer tube and can be tightened through the bolts on the side of the shield. Between the 2nd stage of the cryocooler and the calibration chamber, a copper heating block wrapped with a heating wire is placed to control the temperature of the calibration chamber. The power of the heating wire is controlled by the temperature controller to balance the temperature within the range of 4–300K. Three thermometers are needed. One for heating block is to form a temperature control system with a Lakeshore model 336 temperature controller, one mounted on the top of the calibration room is to indicate the temperature difference that meets the requirements, and another to measure the temperature of the top of the thermal shield. The measuring pipe is made of four welded pipes, including upper straight pipe, upper bellows, lower straight pipe, and lower bellows from top to bottom, as shown in Fig. 1. The upper straight pipe is welded with a CF flange, neck pipe on head flange and the upper bellows, respectively. The lower straight pipe is welded with the upper bellows and the lower bellows, respectively. In particular, the lower part of the lower straight pipe cannot contact the lower bellows. This structure is called a Sniffer tube which can keep the temperature of the measuring pipe above 90K [4].

Fig. 1. Section view of cryogenic vacuum calibration apparatus. 1-quadrupole mass spectrometer, 2-vacuum pump system, 3-hot cathode ionization gauge, 4-sniffer tube, 5-vacuum vessel, 6-1st stage of GM cryocooler, 7-capacitance diaphragm vacuum gauge, 8-Six way joint, 9-high vacuum trimmer valve, 10-measuring pipe. 1-thermal shield, 12-calibration chamber, 13-heating block, 14-2nd stage of GM cryocooler.

Due to the extremely thin wall thickness (~0.2 mm), bellows are the primary temperature control structure in the measurement system. The optimal bellows length ratio is proportional to the integral average value as shown in Eq. (1), that is, the length of 300-100K section is 4.8 times that of 100-4K section for 316 stainless steels. The Eq. (1)

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is derived from heat equation and energy conservation equation.  300  100 100 4

λd T λd T

=

L100−300K L4−100K

(1)

where λ is the thermal conductivity as a function of temperature; L is the length of bellows; T is the temperature. 2.2 Vacuum Measurements and Control The vacuum of the apparatus is divided into inner vacuum and outer vacuum. The inner vacuum is the space used to measure and calibrate the vacuum at cryogenic temperature, while the outer vacuum is used to keep the properties of the insulation material well. The Capacitance diaphragm vacuum gauge is a non-pumping gauge with uncertainty less than 0.5% and can be used as a standard vacuum gauge in the vacuum range above 10−3 Pa at room temperature [9]. The vacuum range corrected √ by thermal transpiration effect is about 10−4 Pa according to the formular p1 = p2 T1 /T2 . The hot cathode ionization gauge has higher uncertainty, but larger measurement range (typically 10 ~ 10−12 Pa [9]). For cryogenic vacuum measurement, cathode ionization gauges have a pumping effect. The pumping speed of ionization gauge is between 0.01 and 0.1 L/s [11, 12]. Therefore, the flow conductivity of measuring pipe should be far larger than this pumping speed. That is, the measuring pipe should be as short and wide as possible. The gas component in the vacuum will change at different temperatures due to the cryocondensation effect. A quadrupole mass spectrometer can measure the different gas components in vacuum at different temperatures and pressures. The High vacuum trimmer valve is used to establish the statically vacuum which can control the type and flow rate of gas into the calibration chamber. 2.3 Calibration Chamber The design of the calibration chamber shall first meet the requirements specified in ISO 3567. For spherical calibration chamber, it shall be designed and operated as follows [13]. Frist, the calibration chamber shall have a volume of at least 20 times the total volume of all the gauges and associated pipe work connecting the chamber and the gauges. Secondly, temperature differences between arbitrary points across the calibration chamber shall be less than 1 K. Lastly, the spatial mean temperature of the calibration chamber shall be (296 ± 3)K during calibration, while the mean temperature should not change by more than 1 K. Considering that the temperature uniformity of the calibration chamber needs to be maintained within 1K at cryogenic temperature, copper and aluminum can be used for their high thermal conductivity. But there are fewer gas molecules adsorbed on the copper surface [14] and higher thermal conductivity at cryogenic temperature compared to aluminum. Therefore, copper (RRR ≥ 100) is used as the material of the calibration chamber. The calibration chamber is made in two halves, the upper part of which is connected to the measuring pipe by a hole at the top. The two halves of the calibration chamber are vacuum brazed together with the stainless steel transition part on the top. The roughness of calibration chamber inner wall should be less than 0.04μm to minimize the radiation heat transfer from the measuring pipe [15].

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2.4 Vacuum Vessel In order to minimize the volume of the measuring pipe and the calibration room, the cold head upside is turned down and the measuring pipe directly through the head of the vacuum vessel. Two flanges are mounted on the head to seal the vacuum. In order to facilitate disassembly, the vacuum cylinder is designed into three sections, among which the upper section needs to be supported by screws when the middle section is disassembled and moved down. Then the upper section can be disassembled from the flange welded on the measuring pipe.

3 Numerical Simulation In order to analyze the temperature requirements of the calibration chamber and the measuring pipe, the thermal shield, the measuring pipe, the calibration chamber and the heating block was analyzed numerically. The surfaces out of the thermal shield and in the thermal shield are covered by multilayer insulations. The setting of heat load has a large margin. Due to the large temperature difference between the sniffer tube and the calibration chamber, the inner surfaces of the calibration chamber and the measuring pipe was set to S2S thermal radiation boundary with emissivity equal to 0.1. The thermal conductivities of 316 stainless steel and copper are required from NIST. After grid independence verification, as shown in Fig. 2, the number of grids is set to 75361. The boundary conditions and results are listed in the Table 1. Figure 3 shows the simulation results on heating power of heating wires at different temperatures.

Fig. 2. Grid independence verification

Fig. 3. Heating power for different temperatures

As shown in Table 1 and Fig. 4, the overall temperature difference in the calibration chamber is less than 1K at temperature range of 4-300K. The temperature at the end of the measuring tube is greater than 90K, which meets the requirements.

4 Error Analysis The standard uncertainty of the quantity determined by the calibration, e.g. error of reading, correction factor, sensitivity coefficient, shall be calculated in accordance with ISO/IEC Guide 98–3. The following uncertainty can be significant [13].

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Surfaces

Boundary conditions (group) 1

2

3

4

5

Out of the thermal shield (W/m2 )

1.5

1.5

1.5

1.5

1.5

In the thermal shield (W/m2 )

0.1

0.1

0.1

0.1

0.1

1st stage of cold head (K)

100

118

190

270

300

2nd stage of cold head (K)

4

18.5

80

150

220

Top of the measuring pipe (K)

300

300

300

300

300

Sniffer tube and the calibration room

S2S thermal radiation, Emissivity = 0.1

Heat flow of 2nd stage (K)

1.50

19.7

54.8

102.6

282.0

Maximum temperature difference in calibration chamber (K)

0.109

0.024

0.106

0.22

0.25

Fig. 4. Numerical simulation results of 1st group

Uncertainty of a) base pressure due to inaccuracy of measurement and drift in time, of b) calibration pressure due to non-equal density and velocity distribution of gas molecules at entrance flange of UUC (Unit Under Calibration) and reference gauge, of c) calibration pressure due to drift in time. d) Measurement uncertainty of the reference gauge. e) Uncertainty of indicated reading of the UUC. f) Uncertainty owing to impurities in the calibration gas. g) Repeatability of the measurements. The uncertainty of the instruments can be calculated by the error transfer formula. In addition, the following factors affecting the uncertainty should be considered. The flow conductivity of the measuring pipe is ~ 1L/s. As mentioned above, the pumping speed of cathode ionization gauge is about 0.01 ~ 0.1L/s, so the corresponding uncertainty generated by the pumping speed is 1 ~ 10%. Capacitance diaphragm vacuum gauge has no pumping effect, that is, no such uncertainty. For capacitance diaphragm vacuum gauge is a non-pumping gauge, the measurement results can be corrected directly by thermal transpiration equation. But for hot cathode ionization gauge, a modified equation for cryogenic temperature vacuum measurement is presented by Matsumoto Hisanori [16], which is derived from the linear equation of

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hot cathode ionization gauge based on thermal transpiration equation and law of partial pressure. Pt =



Tt T0

1 2

α K1 K2 ···Kn

·

1 Ie

(2)

where (Tt /T0 )1/2 is the temperature correction factor, and T0 is the gauge temperature, for hot cathode ionization gauge equal to 420K. Tt is the temperature of calibration chamber. Ie is the electron current of the gauge. α/(K1 K2 · · · Kn ) is called the sensitivity correction factor at cryogenic temperature. When a certain temperature is reached, cryocondensation of some gas occurs, which is shown as a reduction of the items in the denominator (K1 K2 · · · Kn ). The sensitivity correction factor can be obtained by means of an experimental comparison with a quadrupole mass spectrometer. Molecule α represents the sum of the partial pressure ion flow of the remaining gas with different coefficients, which is directly measured by the ion current amplifier according to the principle of superposition. Due to the low temperature of the measurement object, cryopumping can affect the measurement. Cryopumping can divided into cryocondensation and cryosorption [1]. Cryocondensation will not occur when the temperature is above the critical temperature of the gas. Some studies have pointed out that in a cryogenic system of 4K, only helium and hydrogen exist [17, 18]. The critical temperatures of helium, hydrogen, and nitrogen are 5.2K, 33.2K, and 126K, respectively. Since the two-stage regenerator of GM refrigerator is a series structure, when the operating temperature of the 2nd stage of the refrigerator rises, the temperature of the 1st stage will also rise. Therefore, the temperature of the Sniffer tube will also rise. Specifically, when the temperature of the calibration chamber is greater than 40K, the temperature of the Sniffer tube will be larger than the critical temperature of nitrogen, so the impact of cryocondensation on the measuring tube can be avoided. When the temperature of the calibration chamber below the critical temperature of a gas, the gauge under different temperatures and pressures cannot be calibrated because the temperature of the calibration chamber corresponds to the vapor pressure of the gas one by one. At this time, gas whose critical temperature is lower than the measured temperature should be selected for calibration. When the temperature of the calibration chamber is lower than 5.2K, only helium can be used as the calibration gas, and at this time, only specific temperature and pressure can be calibrated. For cryosorption, the qualitative analysis can be conducted from the adsorption isotherm data of stainless steel and copper. For a given surface at a given temperature and pressure, there is an upper limit of the amount of gas molecules absorbed by the metal surface. When the temperature or pressure changes, the surface will adsorb or desorb gas until reach to the surface coverage σ2 from a steady state σ1 and then the speed of the pump becomes zero. That means the effect of cryosorption can be avoid through waiting some times for steady state. Finally, according to the error transfer formula, the total uncertainty of the system is 1.8 ~ 10.1% for different pumping effect of hot cathode ionization gauge.

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5 Summary and Discussion In this paper, an apparatus for vacuum calibration at cryogenic temperature was designed. The GM cryocooler and heating block were used to control the temperature of calibration chamber. The length of measuring pipe was minimized by reduce the volume of the calibration chamber as much as possible. The optimal ratio of two bellows on the measuring pipe was calculated, which is equal to 4.8:1 for the length of 100-300K one to the length of 4-100K one. The temperature difference on the calibration chamber was analyzed through numerical analysis for which is less than 1K and temperature of sniffer tube is above 90K. The vacuum pump and trimmer valve were used to control the gas and the pressure in the calibration chamber. The structure of sniffer tube was designed to keep the temperature of measuring pipe above 90K and which is above the critical temperature of the helium and hydrogen to avoid cryocondensation effect, through waiting for a steady state to avoid cryosorption effect. The quadrupole mass spectrometer was used to analyze the gas component at cryogenic temperature. Finally, the thermal transpiration equation was modified to calculate the real pressure on the calibration chamber. After the above analysis, it seems that a cryogenic vacuum calibration apparatus can be made up. But the uncertainty of the apparatus is not good enough yet, especially in the ultra-high vacuum. If the apparatus is built, the more accurate cryogenic temperature pressure with gauges at room temperature can be obtained, which can be used to guide the analysis of physical properties or materials under vacuum cryogenic temperature.

References 1. Welch, K.: Capture Pumping Technology. Elsevier (2001) 2. Cunnington, G.R., Keller, C.W., Bell, G.A.: Interim report, thermal performance of multilayer insulations. NASA-CR 72605, 3–4 (1971) 3. Price, J.W.: Measuring the gas pressure within a high-performance insulation blanket. In: Timmerhaus, K.D. (ed.) Advances in Cryogenic Engineering. Advances in Cryogenic Engineering, vol 13. Springer, Boston, MA (1995). https://doi.org/10.1007/978-1-4757-05164_70 4. Edwards, D., Jr., Limon, P.: Pressure measurements in a cryogenic environment. J. Vac. Sci. Technol. 15(3), 1186–1188 (1978) 5. Wallen, E.: Experimental test of the propagation of a He pressure front in a long, cryogenically cooled tube. J. Vac. Sci. Technol., A: Vac., Surf. Films 15(6), 2949–2958 (1997) 6. Kuriyama, T.: Development of a 5 K GM refrigerator using rare earth compounds as a regenerator matrix. In: Proceedings of International Conference on Cryogenic and Refrigeration (1989) 7. Inaguchi, T, Nagao, M, Naka, K, et al.: Effects of thermal conductance in the cooling stage of a 4K-GM refrigerator on refrigeration capacity. In: Kittel, P. (ed.) Advances in Cryogenic Engineering, vol. 43A, pp. 1807–1814. Springer, Boston, MA, (1998). https://doi.org/10. 1007/978-1-4757-9047-4_228 8. SHI Cryocooler Product Catalogue. https://rowaco.se/app/uploads/2019/06/Cryocooler-Pro duct-Catalogue.pdf. Accessed 10 May 2022 9. Leybold product brochure. https://www.leyboldproducts.com/media/pdf/2b/ff/ae/361001 9902_Sensors_2020_web.pdf. Accessed 29 Dec 2022 10. NIST Cryogenic Material Properties Calculators. https://trc.nist.gov/cryogenics/calculators/ graphcalc.html. Accessed 29 Dec 2022

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11. Berman, A.: Total Pressure Measurements in Vacuum Technology. Academic Press (2014) 12. Peacock, R.N., Peacock, N.T., Hauschulz, D.S.: Comparison of hot cathode and cold cathode ionization gauges. J. Vac. Sci. Technol., A: Vac., Surf. Films 9(3), 1977–1985 (1991) 13. ISO 3567:2011 - Vacuum gauges (2011) 14. Moulard, G., Jenninger, B., et al.: Industrial surfaces behaviour related to the adsorption and desorption of hydrogen at cryogenic temperature. Vacuum 60(1–2), 43–50 (2001) 15. Guobang C, Lihe L.: Technology of Low Temperature Insulation and Heat Transfer. Zhejiang University Press, p. 29 (1989). (in Chinese) 16. Matsumoto, H.: Discussion on some problems of vacuum measurement at low temperature. Vacuum and Cryogenics (01), 39–43 (1985). (translated by Bing Z. in Chinese) 17. Abdel-Samad, S., Abdel-Bary, M., Kilian, K.: Residual gas analysis in the TOF vacuum system. Vacuum 78(1), 83–89 (2005) 18. Chen, S., Wang, R., Li, X.: Experimental investigation and theoretical analysis on measurement of hydrogen adsorption in vacuum system. Int. J. Hydrogen Energy 35(9), 4347–4353 (2010)

Cryogenic Heat Transfer and Thermal Insulation

Vacuum Break in a Helium Cooled Tube with an Inserted Cavity Nathaniel Garceau1,2 , Shiran Bao2 , and Wei Guo1,2(B) 1

Mechanical Engineering Department, FAMU-FSU College of Engineering, Florida State University, Tallahassee, FL 32310, USA [email protected] 2 National High Magnetic Field Laboratory, 1800 East Paul Dirac Drive, Tallahassee, FL 32310, USA Abstract. During a particle-accelerator beamline tube vacuum break, air rushes into the liquid helium cooled tube and condenses on the inner walls. This induces a large heat load onto the liquid helium and causes rapid boiling and potentially a dangerous pressure buildup. To better understand the complex dynamical process, we have conducted systematic experimental studies and modeling using a tube cooled with liquid helium, both normal liquid helium (He I) and superfluid helium (He II). Results indicated a nearly exponential slowing of the gas front. This paper reports the effect of an added cylindrical copper cavity with approximately the same diameter to length ratio as a representative of elliptical beamline cavities. Observations indicated that the gas would propagate past the cavity before fully filling the cavity. This work provides new information toward a full understanding of vacuum break in particle accelerators. Keywords: particle accelerators

1

· liquid helium · vacuum break

Introduction

Modern particle accelerators are long systems composed of interconnected modulates called cryomodules. Superconducting radio-frequency (SRF) cavities immersed in a liquid helium (LHe) bath are housed within these cryomodules. The SRF cavities are connected together to create a long beam tube path which accelerates particles to very high energy [1,2]. This interconnected path creates a large potential risk in the event of vacuum failure into the beam tube. During an accidental rupture, air will rush into the beam tube and then will condense and freeze to the cavity walls. The generated heat is transferred into the liquid helium bath which will cause rapid boiling and potentially dangerous pressure buildup. Due to the catastrophic nature of vacuum failure into a beam tube such as the failure event at CERN in 2008 [3], multiple vacuum failure studies at accelerator labs have been conducted, e.g., [4–7]. These studies showed that gas enters the tube and expands into vacuum at high speeds, but once the gas hits the freezing surfaces of the LHe cooled tube wall, its propagation would slow down radically. c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 413–419, 2023. https://doi.org/10.1007/978-981-99-6128-3_52

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To better understand the complex and dynamical processes occurring during a vacuum break, foundational work has been conducted in our lab over a number of years. Our lab’s pioneering work to understand gas propagation within a beam tube began with Dhuley and Van Sciver who conducted vacuum break experiments by venting nitrogen (N2 ) gas from a buffer tank into a straight LHe cooled vacuum tube [8–10]. They charted the gas propagation in the tube by noting the time of temperature spikes on the surface after breaking the vacuum. Results from their experiments showed a nearly exponential slowing of the propagating gas within the tube. However, their analysis was quantitatively limited and lacked a good explanation of the complex interconnected nature of mass, momentum and heat transfer needed to fully describe the processes. Furthermore, their experimental setup had some design issues which made the analysis in He II baths difficult. Building on Dhuley and Van Sciver’s work, a new experimental system was designed and built so systematic studies in both He I and He II could be conducted. This system was based on a helical tube system which allowed the gas propagation to be observed over a significantly longer length [11,12]. Our first measurements conducted in this new system confirmed the deceleration of the propagating gas front within He I, and it showed He II has a stronger slowing effect than He I. To better understand the complex vacuum break event, a new 1D theoretical model was developed which accounted for the gas dynamics, surface condensation, and heat transfer into He I [12,13]. However, many particle accelerators are cooled by He II due to its efficient heat transfer mechanism known as thermal counterflow. To account for this difference, the model was updated to account for the changes in heat transfer [14]. Both He I and He II simulations were tuned and validated against experimental results. Simulated surface temperature profiles behaved similar to the experimental temperature profiles for their respective run conditions. From these simulations, we were also able to extract useful quantitative behavior of gas dynamics, frost layer growth, heat transfer, and other properties, which are very difficult or impossible to measure within the system experimentally. SRF cavities within particle accelerators have complex shapes which depend on various factors such as the specific particles being accelerated and the desired energy to be achieved. One style of particle accelerator cavity is designed as a series of elliptical cavities [1]. To improve our understanding of the effect of these cavities on the non-uniform slowing of gas propagation during a vacuum failure, a new experiment was designed and built. This paper reports the preliminary observations of a shorter tube system with an inline voluminous cylindrical cavity with similar entrance diameter, outer diameter, and length aspect ratio as a real SRF cavity. Valuable insight was gained from this preliminary experiment about the gas dynamics which will help improve the design of the subsequent experiments. In addition, results suggest that a more comprehensive model to account for the complexities of the cavity geometry will be needed in future.

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Most system components used in our previous He I and He II helical tube experiments were reused for the current experiment [12–14]. However, the helical tube was replaced with a short pipe with a voluminous cavity. The schematic of the system is shown in Fig. 1 (a), and a picture of the assembled system is shown in Fig. 1 (b). In summary, the modified system is composed of a 25.4-mm inner diameter and 1.25-mm thickness tube with a length of 116 cm. The first segment of this tube was a 57 cm stainless steel tube which is vacuum jacketed and contains a heater at the end of the tube. This heater allows for the control of the temperature profile within this section and prevent condensation before the exit of this insulated portion. The exit of the vacuum jacket is connected to the LHe cooled copper pipe via silver brazing. The segment of pipe from the exit of the vacuum jacket to the beginning of the voluminous cavity is approximately 44 cm long. The cavity itself is designed to have the same approximate aspect ratio (i.e., entrance diameter: outer diameter: length) as an SRF cavity [1]. However, we use a cylindrical cavity shape rather than elliptical shape for ease of fabrication and future analysis. Our cylindrical cavity has an entrance diameter of 25 mm, outer diameter of 76 mm, and length of 46 mm. Exit of the cavity is connected to a 25.4 mm inner diameter width and 13 cm in length copper tube.

Fig. 1. (a) Schematic of the new vacuum break system with an inline voluminous cavity; and (b) a picture of the tube and cavity after assembly.

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Experimental Procedure

Our experiments began by charging a 230 L buffer tank with pure nitrogen (99.999% pure) to 100 kPa. This tank allows us to control the gas quality over the course of multiple experiments. Pressure within the buffer tank was measured over the course of an experiment with a factory calibrated TD1000 pressure transducer from Transducers Direct [15]. The copper and stainless steel tube system was completely evacuated before the experiment. The cryostat with the evacuated tube system installed was then precooled with liquid nitrogen (LN2 ). After precooling, the LN2 was completely drained from the cryostat and the system was allowed to warm to at least 90 K to ensure no nitrogen remained within the system. After that, liquid helium (LHe) was slowly transferred into the cryostat. The cryostat bath was filled such that the copper tube was completely covered and the liquid level was at the exit of the vacuum jacket. This liquid level was measured with an American Magnetics Inc. superconducting LHe level sensor. At the start of an experiment, a fast action solenoid valve separating the pressurized nitrogen reservoir and the evacuated tube was opened. Gas flow into the evacuated tube was chocked in a venturi throat at the local speed of sound. Gas propagation within the tube and the cavity was monitored using temperature sensors installed on the outer surface of the copper tube and cavity (see Fig. 1). TM Temperature was measured using seven SD Lake Shore Cernox sensors (thermal response of 15 ms) encapsulated in 2850FT Stycast epoxy following Dhuley’s method [16]. The sensors were approximately located at 14 cm (T1), 24 cm (T2), 34 cm (T3), 44 cm (T4), 46 cm (Tc, mid-point of cavity), 49 cm (T5), and at 59 cm (T6) from the exit of the vacuum jacket (reference origin) as illustrated in TM Fig. 1 (a). Apliezon thermal grease and indium foil were sandwiched between the sensor and the copper wall to reduce contact resistance. All the data was recorded at 4800 Hz Hz using four DT9824 data acquisition modules from Data TM Translation Inc. and National Instruments LabVIEW .

3

Results

Arrival of the propagating gas front at each sensor location is marked by a sharp rise in surface temperature. Time of this rise is called ‘rise time,’ t. To determine this rise time, the temperature data was smoothed using an 80-point moving average to reduce the harmonic and random noise within the data. This is the same data processing procedure used in our prior experiments [12–14]. Figure 2 (a) presents the smoothed experimental data for the He I experiment with 100 kPa buffer tank pressure. Rise time threshold was set at 0.05 K above the bath temperature, and these times are marked by a vertical dashed line in Fig. 2 (a). Note, the rise time of the first sensor (T1) is set to 0 ms. From this figure, two unique results are observed. First, looking at sensors T5 and TC, one can observe that sensor T5 (i.e., the sensor installed just after the cavity) is rising sooner than TC (i.e., the sensor installed on the cavity side

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wall). However, sensor T5 is downstream of TC as illustrated in Fig. 2 (a). This indicates that the gas propagating within the tube passes through the cavity before fully filling the cavity. This implies that in order to build a numerical model of the gas propagation within the tube, a simple 1D model will not be adequate to simulate the observations within the cavity section.

Fig. 2. (a) Smoothed temperature data for preliminary experiment of a tube with a cavity. (b) Average velocity between two sensors graphed at the midpoint between those sensors.

Second unusual observation is that just before the gas enters the cavity, there appears to be an acceleration of the gas front. From our past observations [12– 14], a continuous slowing occurs as the gas propagates down the tube, so the rise times should space out further down the tube. However, one can see in Fig. 2 (a) that the rise times (dashed vertical lines) spacing between the equidistant sensors (T1-T4) does not always increase. This can more clearly be seen if average velocity between the sensors is calculated by dividing the distance between the two sensors over the difference in rise time (v = Δx/Δt). Figure 2 (b) graphs this velocity at the midpoint between the sensors. From this figure, one can clearly see the expected deceleration between the different sequential sensors, but there is an unusual acceleration between sensors T3-T4 just before the cavity. It is not entirely clear why this acceleration occurs. However, some insight might be gained by using our past numerical simulation results for a long He I cooled tube [13]. Figure 3 shows past simulation results for a straight tube cooled by He I with 100 kPa backing tank pressure. Specifically, the figure looks at the entrance region and the buildup occurring during the first few milliseconds. This graph illustrates how the gas starts propagating from the warm insulated tube section then into the cold He I cooled section. As soon as the gas exits the insulated entrance and passes into the LHe cooled section, it starts condensing. Condensation retards the gas propagation. However, for this experiment, the tube length is short so the propagation time scale is less than 25 ms, in contrast to our previous experiments in the helical tube where the propagation time was on

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Fig. 3. Simulated density profiles for a straight tube cooled by He I with 100 kPa backing tank pressure [13].

the order of 1400 ms. From this graph, one can see that the density in the warm section does not saturate until around 25 ms. Therefore, the mass flux entering the cold section continually increases. From our previous experiments [14,17], it is known that higher mass flow rates correspond to higher propagation velocity. Therefore, this transient buildup in the warm section could be the cause of this observed acceleration. This supposition will be examined in our future systematic experiments.

4

Conclusion

Vacuum break in a tube cooled by liquid helium is a complex event which couples various dynamical processes of mass, momentum, and heat transfer. To understand vacuum break in a particle accelerator beam tube, our lab has been conducting a variety of experiments in a simplified versions using tubes cooled by liquid helium. This paper discussed a preliminary experiment for the next stage of development using a pipe with an inserted voluminous cylindrical cavity. This new experiment is designed to shed light on the slowing gas propagation in a single elliptical SRF cavity may have within a particle accelerator beam tube. Results showed two unique features. First, the gas will propagate through the cylindrical cavity before fully filling the cavity. Second, there is an observed acceleration just before the cavity. It is unclear why this acceleration occurs, but it was theorized to be the result of the transient buildup of the gas density in the warm section. Future modeling will need to include a more comprehensive model to simulate the cavity section in the tube. Additionally, a longer segment of pipe before the cavity will need to be added to mitigate the effects of the warm section on the gas propagation.

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Acknowledgments. This work is supported by U.S. Department of Energy under Grant No. DE-SC0020113. The experiment was conducted at the National High Magnetic Field Laboratory, which is supported by National Science Foundation Cooperative Agreement No. DMR-1644779 and the State of Florida.

References 1. Padamsee, H.: CERN Yellow Report CERN-2014-005, pp. 141–169 (2015). https:// doi.org/10.5170/CERN-2014-005.141 2. Pagani, C., Pierini, P.: Cryomodule design, assembly and alignment. In: Proceedings of 12th International Workshop on RF Superconductivity. Ithaca (2005) 3. Lebrun, P., Delucinge, E.: Interim summary report on the analysis of the 19 september 2008 incident at the LHC, Tech. Rep., EDMS 973073, CERN (2008) 4. Wiseman, M., et al.: In Kittel, P.(ed.), Adv. in Cryog. Eng., vol. 39, pp. 997–1003. Springer US, Boston, MA (1994). https://doi.org/10.1007/978-1-4615-2522- 121 5. Seidel, M., Trines, D., Zapfe, K.: Failure analysis of the beam vacuum in the superconducting cavities of the tesla main linear accelerator. Tech. Rep., TESLAReport 2002-06, DESY (2002) 6. Boeckmann, T., et al.: In: 22nd International Cryogenic Engineering Conference and International Cryogenic Materials Conference 2008, pp. 723–728 (2009) 7. Dalesandro, A.A., Theilacker, J., Van Sciver, S.: In AIP Conference Proceeding, vol. 1434 (Spokane, WA, USA, 2012), vol. 1434, pp. 1567–1574 (2012). https:// doi.org/10.1063/1.4707087 8. Dhuley, R., Van Sciver, S.: Int. Heat Mass Tran. 96, 573 (2016). https://doi.org/ 10.1016/j.ijheatmasstransfer.2016.01.077 9. Dhuley, R.C., Van Sciver, S.W.: Int. Heat Mass Tran. 98, 728 (2016). https://doi. org/10.1016/j.ijheatmasstransfer.2016.03.077 10. Dhuley, R.C.: Gas propagation in a liquid helium cooled vacuum tube following a sudden vacuum loss. Dissertation, Florida State University, Tallahassee (2016). Retrieved from FSU, Proquest Diss. Publishing database (UMI No. 10120583) 11. Garceau, N., Bao, S., Guo, W., Van Sciver, S.: Cryogenics 100, 92 (2019). https:// doi.org/10.1016/j.cryogenics.2019.04.012 12. Garceau, N., Bao, S., Guo, W.: Int. Heat Mass Tran. 129, 1144 (2019). https:// doi.org/10.1016/j.ijheatmasstransfer.2018.10.053 13. Bao, S., Garceau, N., Guo, W.: Int. J. Heat Mass Tran. 146, 118883 (2020). https:// doi.org/10.1016/j.ijheatmasstransfer.2019.118883 14. Garceau, N., Bao, S., Guo, W.: Int. Heat Mass Tran. 181, 121885 (2021). https:// doi.org/10.1016/j.ijheatmasstransfer.2021.121885 15. Transducers Direct. TD1000 pressure transducer (2022). https://transducersdirect. com/ 16. Dhuley, R., Van Sciver, S.: Cryogenics 77, 49 (2016). https://doi.org/10.1016/j. cryogenics.2016.05.001 17. Garceau, N., Bao, S., Guo, W.: Effect of mass flow rate on gas propagation after vacuum break in a liquid helium cooled tube. In: IOP Conference Series: Materials Science and Engineering, Volume 755, Advances in Cryogenic Engineering: Proceedings of the Cryogenic Engineering Conference (CEC) 2019, 21–25 July 2019, Hartford, Connecticut, USA, p. 012112. IOP Publishing, Philadephia (2020). https://doi.org/10.1088/1757-899X/755/1/012112

Numerical Study on the Flow and Heat Transfer Characteristics in CFETR Cryostat During a Loss of Vacuum Jianhong Huang1 , Jianjian Wei1 , Jian Ge2 , Sumei Liu2 , Yuntao Song2 , and Tao Jin1(B) 1 Institution of Refrigeration and Cryogenics/Key Laboratory of Refrigeration and Cryogenic

Technology of Zhejiang Province, Zhejiang University, Hangzhou, China [email protected] 2 Institution of Plasma Physics, Chinese Academy of Sciences, Hefei, China

Abstract. The cryostat in Chinese Fusion Engineering Testing Reactor (CFETR) provides a vacuum and low temperature environment to limit the convective heat exchange for superconducting magnets. If the thermal shield is broken due to the rupture of components or welds, the helium gas in cooling tube will flow into the cryostat, which can increase the thermal load of superconducting magnets and even cause damage to the system. A two-dimensional numerical model, using ideal gas model and SST k-ω model, is built to simulate the loss of vacuum accident that helium gas flows into the cryostat under a postulated breach. The velocity at monitoring points, the velocity contours and the average heat flux on the walls of superconducting magnets are obtained under different initial pressure conditions. The results show that a shock wave can be observed near the gas inlet. When the initial pressure in the cryostat ranges from 1000 Pa to 10000 Pa, the velocity at the axis in front of the shock wave is almost unaffected, while the velocity after the shock wave decreases. However, the average heat flux on the walls of superconducting magnets increases rapidly as the pressure increases, which needs to be considered in designing the safety system. Keywords: CFETR Cryostat · Loss of Vacuum Accident · Helium Gas · Numerical Simulation

1 Introduction Nuclear fusion is a promising option for a clean and safe solution for long-term energy needs. The Chinese Fusion Engineering Testing Reactor (CFETR) is a Tokamak-type magnetic confinement nuclear fusion device [1]. To minimize heat loads transferred from outside environment to the superconducting coils operated at 4.5 K, the cryostat surrounding the Tokamak basic machine provides a vacuum environment (10–4 Pa) for the operation of superconducting magnet coils. The thermal shield operating in the temperature range of 80–100 K intercepts thermal radiation from the cryostat inner wall and other warm surfaces [2]. If the cooling tube on thermal shield is broken due to the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 420–426, 2023. https://doi.org/10.1007/978-981-99-6128-3_53

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rupture of components or welds, the helium gas in cooling tube flows into the cryostat and causes a loss of vacuum, which can greatly increase the heat load of superconducting magnets and even cause damage to the system. In recent decades, many small-scale experiments and numerical studies have been implemented on the loss of vacuum accidents. ENEA institutions in Italy [3–5] built a Small Tank for Aerosol Removal and Dust (STARDUST) facility later upgraded to the STARDUST-UPGRADE facility, to comprehensively study the effects of wall temperature, pressurization rate, inlet position on air flow and dust migration in the vacuum vessel under the loss of vacuum condition. Rosa [6] analyzed the ingress of air into the ITER cryostat under a loss of vacuum accident with a postulated breach whose crosssectional area was 0.2 or 1 m2 with ANSYS CFX. Results showed that the maximum pressure in the cryostat was below the design limit pressure of 0.2 MPa. Marco [7] performed a three-dimensional numerical thermo-fluid dynamic analysis on the ITER Cryostat Space Room, then obtained the temperature distribution of the cryostat and the heat load on thermal shield under normal operating conditions. Previous studies about the loss of vacuum accidents in fusion device mainly emphasized on the air flow and dust migration in the vacuum vessel, while few research work done on the fluid flow in the cryostat. The present work is intended to simulate the process that helium flows into vacuum chamber in cryostat under a postulated breach on thermal shield, to analyze the distribution of velocity and temperature fields and the area-weighted average heat flux on walls of the superconducting magnets under different initial pressure conditions. The results are expected to provide a reference for the design of system safety control in the cryostat of CFETR.

2 Simulation Description 2.1 Two-Dimensional Model Simplification Cryostat is one of the most important components in CFETR. When helium leaks from thermal shield into the cryostat vacuum chamber, the main structures involved are thermal shield and superconducting magnets, as shown in Fig. 1. The thermal shield mainly includes vacuum vessel thermal shield (VVTS) and cryostat thermal shield (CTS), surrounding superconducting magnets. The magnet system mainly consists of center solenoid winding (CS), poloidal field winding (PF) and toroidal field winding (TF) [8]. To further simplify mesh construction and facilitate simulation, a two-dimensional axisymmetric model of the simplified structure is built, as shown in Fig. 2. A postulated breach with the cross-section area of 0.02 m2 is located at the center of CTS top surface. Three monitoring points (A, B and C) are set on the axis (0.7 m, 15.2 m and 31.2 m away from the helium breach, respectively) to reveal the evolution of helium jet under different initial pressures in vacuum chamber. When vacuum failure occurs, the helium gas with the pressure of 1.8 MPa and the temperature of 80 K flows into vacuum chamber through the postulated breach. 2.2 Numerical Model The normal operating pressure is as low as 10–4 Pa in the cryostat vacuum chamber. As vacuum failure occurs, the pressure in vacuum chamber rises rapidly. The effect of

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Fig. 1. Simplified cryostat model.

Fig. 2. Boundary conditions of the 2D model.

initial pressure on the tokamaks is not so important for pressures lower than 1000 Pa according to the Reference [9]. The initial pressure in the vacuum chamber is set to 1000 Pa, 5000 Pa, and 10000 Pa to study the viscous flow stage. According to the mass flow rate formula proposed by Xiao et al. [10], when pressure in the vacuum chamber is less than 878.6 kPa, the calculated mass flow rate of helium is 63.35 kg/s. The mass flow rate for the inlet boundary condition is set as follows. m ˙ = 63.35 × (1 − e−t )

(1)

In the two-dimensional axisymmetric model, the pressure-based solver was chosen to solve the compressible Navier-Stokes equation. The SST k-ω model was used to deal with the jet flow. Walls of thermal shield and superconducting magnets were treated as no-slip wall boundaries with the temperature of 80 K and 4.5 K, respectively. The initial temperature in vacuum chamber was assumed to be 42.25 K. The Green-Gauss NodeBased method was selected for the gradient calculation. The PRESTO! scheme was used for the pressure interpolation, and the First Order Upwind discretization scheme was used for calculating the momentum, turbulent kinetic energy, turbulent dissipation rate and energy equations. 2.3 Grid Independence Test and Model Validation The calculated velocity of a monitoring point under three girds of 76910, 136063 and 258516 cells are shown in Fig. 3. The calculation results of 136063 and 258516 cells are very close, which deviate from that of 76910 cells. Considering the accuracy of calculation results and the saving of computing resources, the grid with 136063 cells is selected for calculations with different initial pressures. The numerical model was used to simulate the loss of vacuum accident in the STARDUST [3], and velocity results at the monitoring points are shown in Fig. 4. Overall, the trend of velocity evolution is generally consistent with the experimental results. In terms of maximum velocity, the simulated value is about 449 m/s with the relative error of 9.83%.

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Fig. 3. Mesh sensitivity analysis

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Fig. 4. Model validation

3 Results and Discussions 3.1 Velocity Distribution Inside Vacuum Chamber The velocity distribution contours inside vacuum chamber at 10 s under different initial pressure conditions are basically similar, as shown in Fig. 5. The helium gas flows into the vacuum chamber from the inlet, then hits the bottom wall and spreads along the wall to form recirculation area. Figure 5(d) is a zoomed-in view of the velocity contour near the inlet, indicating the existence of a shock wave. Before the shock wave, the internal energy of the helium gas near inlet is converted into kinetic energy due to the expansion effect and the supersonic flow is realized with the maximum velocity about 900 m/s. After passing through the shock wave, the velocity of helium drops rapidly. When the initial pressure increases from 1000 Pa to 10000 Pa, the velocity in front of the shock wave is almost unaffected, while the velocity after the shock wave decreases due to the increased pressure.

Fig. 5. Velocity contour at 10 s

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3.2 Velocity at the Axis and Monitoring Points The velocity at the axis at 10 s under different initial pressures is shown in Fig. 6 (a). As the pressure in vacuum chamber increases, the position of shock wave shifts slightly towards the helium inlet. The velocity evolution of point A located in front of the shock wave is almost the same, as shown in Fig. 6(b). While velocity evolutions of point B and point C located after the shock wave decrease with the increasing pressure in vacuum chamber, as shown in Fig. 6(c) and (d). The monitoring point B and point C are located in the zone where the jet is perfectly expanded. The initial density of helium in vacuum chamber increases with the initial pressure. A diminution of density difference between the helium in vacuum chamber and the injected helium can cause the helium velocity to decrease, which has been experimentally observed in Reference [11].

Fig. 6. Velocity at axis and monitoring points

3.3 Heat Transfer on the CS Wall Under Different Initial Pressures The area-weighted average heat flux on the walls of superconducting magnet coils under the condition of initial pressure of 10000 Pa is shown in Fig. 7. The helium mainly flows near the CS and forms the recirculation zone at the bottom. The area-weighted average heat flux on the wall of CS is the largest, followed by those of pf-5, pf-4, pf-6 and pf-7. As the initial pressure increases from 1000 Pa to 10000 Pa, the area-weighted average heat flux on the wall of CS both increases, as shown in Fig. 8.

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Referring to the correlation between Nusselt number (Nu) and Reynolds number (Re) for the heat transfer of helium in Reference [12], the Nusselt number in vacuum chamber is correlated with the product of helium density and velocity as follows, Nu ∝ (ρ · U)a

(2)

where a is the coefficient related to a specific condition, ρ is the fluid density, U is the flow velocity. As the pressure in vacuum chamber rises, the average velocity (U ) near CS wall decreases, while the average density (ρ) of helium increases. As shown in Table 1, the product of the average density and velocity of helium increases, which indicates that the Nusselt number rises according to Eq. (2). As the Nusselt number rises, the convective heat transfer coefficient on the CS wall increases and the heat transfer is enhanced. Thus, the helium density plays a dominant role in promoting heat transfer within 10 s under the three initial pressure conditions discussed in this work.

Fig. 7. Heat flux on superconducting magnet walls

Fig. 8. Heat flux on the CS wall

Table 1. Average density and velocity near CS wall at 10 s.   ρ · U (kg/ m2 · s )

Pressure at 0 s (Pa)

Pressure at 10 s (Pa)

ρ(kg/m3 )

U (m/s)

1000

6508

0.037454

164.06

6.08

5000

11616

0.082602

109.10

9.01

10000

17350

0.139572

86.35

12.05

4 Conclusion A simplified two-dimensional model is established for analyzing the loss of vacuum accident that the helium flows into the cryostat of CFETR under a postulated breach in the thermal shield. The flow and heat transfer characteristics of helium under different initial pressures in the vacuum chamber are simulated and analyzed. The main conclusions are summarized as follows.

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(1) When the initial pressure increases from 1000 Pa to 10000 Pa, the velocity in front of the shock wave is almost unaffected, while the velocity after the shock wave decreases. (2) The area-weighted average heat flux on the wall of CS is much higher than that on other walls of superconducting magnet coils since the velocity near the CS wall is higher than those of other superconducting magnet coils. (3) Under the three initial pressure conditions discussed, the heat transfer coefficient on the wall of CS increases with the increasing pressure in the vacuum chamber, and the density dominates the heat transfer. Acknowledgments. This work is financially supported by the National Key Research and Development Program of China (Grant No. 2017YFE0300503) and the State Key Laboratory of Technologies in Space Cryogenic Propellants (SKLTSCP1911).

References 1. Zhuang, G., Li, G.Q., Li, J., et al.: Progress of the CFETR design. Nucl. Fusion 59(11), 112010 (2019) 2. Song, Y., Wu, W., Du, S.: Tokamak Engineering Mechanics. Springer, Berlin (2014) 3. Bellecci, C., Gaudio, P., Lupelli, I., et al.: Loss of vacuum accident (LOVA): comparison of computational fluid dynamics (CFD) flow velocities against experimental data for the model validation. Fusion Eng. Des. 86(4), 330–340 (2011) 4. Porfiri, M.T., Forgione, N., Paci, S., et al.: Dust mobilization experiments in the context of the fusion plants—STARDUST facility. Fusion Eng. Des. 81(8), 1353–1358 (2006) 5. Poggi, L.A., Malizia, A., Ciparisse, J.F., et al.: First experimental campaign to demonstrate STARDUST-upgrade facility diagnostics capability to investigate LOVA conditions. J. Fusion Energy 34(6), 1320–1330 (2015) 6. Lo Frano, R., Aquaro, D., Olivi, N.: Fluid dynamics analysis of loss of vacuum accident of ITER cryostat. Fusion Eng. Des. 109–111, 1302–1307 (2016) 7. Olcese, M., Barbano, F.: Thermo-fluid dynamics analysis of ITER Cryostat Space Room. Fusion Eng. Des. 135, 183–195 (2018) 8. Zheng, J., Liu, X., Song, Y., et al.: Concept design of CFETR superconducting magnet system based on different maintenance ports. Fusion Eng. Des. 88(11), 2960–2966 (2013) 9. Rossi, R., Gaudio, P., Martellucci, L., et al.: Numerical simulations of radioactive dust particle releases during a loss of vacuum accident in a nuclear fusion reactor. Fusion Eng. Des. 163, 112161 (2021) 10. Xiao, J., Travis, J.R., Breitung, W., et al.: Numerical analysis of hydrogen risk mitigation measures for support of ITER licensing. Fusion Eng. Des. 85(2), 205–214 (2010) 11. Ewan, B.C.R., Moodie, K.: Structure and velocity measurements in underexpanded jets. Combust. Sci. Technol. 45(5–6), 275–288 (1986) 12. Liu, Q., Shibahara, M., Fukuda, K.: Transient heat transfer for forced convection flow of helium gas over a horizontal plate. Experimental Heat Transfer 21, 206–219 (2008)

Optimization of Multilayer Insulation and Vapor-Cooled Shield Combination Coupled with a Para-Ortho Hydrogen Converter Chuiju Meng, Xujin Qin, Wenbing Jiang, and Yonghua Huang(B) Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai, China [email protected]

Abstract. Hydrogen molecules have two isomers, namely orthohydrogen and parahydrogen. Due to the difference in rotational energy levels, the conversion process from parahydrogen to orthohydrogen accompanies endothermic effects. A thermodynamic model was established for liquid hydrogen storage in orbit, integrating the multilayer insulation, a vapor-cooled shield, and a para-ortho hydrogen converter. The heat transfer behavior in the vapor venting process accompanying para-ortho hydrogen conversion was studied. The relation between the cooling capacity produced by the para-ortho hydrogen conversion and the position of the vapor-cooled shield inside the multilayer material was discussed. The optimum position of the vapor-cooled shield and the temperature of the para-ortho hydrogen conversion for the integrated system was determined. Keywords: Liquid Hydrogen · Multilayer Insulation · Vapor-cooled Shield · Para-ortho Hydrogen Conversion

1 Introduction Technologies such as multilayer insulation (MLI), vapor-cooled shield (VCS), and thermodynamic vent system (TVS) have been considered effective solutions for liquid hydrogen storage in orbit. Although the boil-off rate of liquid hydrogen has been reduced to 3% per month [1], it still does not meet the requirement for long-term propellant storage missions. Extra measures are needed for further suppression of the heat leakage. Unlike other cryogens, liquid hydrogen has a unique potential to accomplish the purpose by using the endothermic effect of para–ortho hydrogen conversion (POC). The most feasible approach is to fill catalyst in the vapor flow path on the VCS, which does not need major modification of the liquid hydrogen storage system. Nast et al. analyzed the performance of POC for cooling an infrared sensor in spacecraft. The cooling time of the infrared instrument baffle was extended by 40% approximately [2]. Liggett et al. evaluated experimentally the performance of a vapor-cooled shield (VCS) coupled with a POC converter. They found that dual VCS with a converter could reduce the boil-off loss by 26% compared to the configuration with MLI only when the warm boundary remains at 286 K [3]. Bliesner et al. announced that the POC © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 427–434, 2023. https://doi.org/10.1007/978-981-99-6128-3_54

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provided additional 6.5–12.1% cooling power at 90 K, which intended to cool liquid oxygen [4]. Thus it can be seen that POC has promissing application prospects in hydrogen storage at low temperatures, especially for long-term storage mission. However, the coupling mechanism between the VCS and the POC is still ambiguous. In the present study, a one–dimensional model that considers the hydrogen vapor sensible heat in the VCS tube and conversion heat in the POC converter, has been established to investigate and optimize the thermal performance of the VCS/POC combination for cryogenic propellant storage applications. Subsequently, the influences of the VCS position and the warm boundary temperature on the POC cooling effect are discussed as well.

2 Physical Model and Numerical Implementation

Fig. 1. Liquid hydrogen tank’s thermal insulation structure with vapor-cooled shield

The tank uses the same geometry as the MHTB facility [5]. As shown in Fig. 1, the composite insulation structure includes a foam layer, a multilayer insulation (MLI) blanket, a vapor-cooled shield (VCS), and a POC converter. The VCS is embedded and position-adjusted in the MLI blanket. The POC converter is installed at the midpoint of the length of the tube on the VCS. The hydrogen vapor from the liquid hydrogen tank has its enthalpy increased as it flows in the VCS tube but gets cooled when it passes the POC converter. In addition, the POC process results in a change in the isomers’ composition, which is also considered in the model. 2.1 Thermal Model for Composite Insulation with Vapor-Cooled Shield As shown in Fig. 1, the heat transfer is regarded as a one-dimensional heat transfer process from the warm boundary to the cold one. Qtotal is the total heat flux from the warm boundary to VCS. The combination Qsen+conv stands for heat flux leakage taken away by the hydrogen vapor flow in the VCS tube. Qnet is the net heat flux from VCS to the cold boundary, which will be used to calculate the liquid hydrogen evaporation, assuming any parasitic heat leakage is neglected. For the heat transfer process in liquid hydrogen tank’s thermal insulation structure, the energy balance equation is written as, Qtotal =Qsen + conv +Qnet

(1)

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As the VCS is embedded in the MLI blanket, the total heat flux is influenced by the thermal resistance of the MLI on the outside of the VCS, while the net heat flux is affected by MLI thermal resistance on the inner side of the VCS.  Qtotal =(TWBT − TVCS ) RMLI,OUT (2)  Qnet =(TVCS − TCBT ) (Rfoam + RMLI,IN )

(3)

where RMLI, OUT and RMLI, IN are the MLI thermal resistance on outer and inner sides of the VCS, respectively. The “layer by layer” model by Mcintosh [6] is adopted to calculate the thermal resistance of the MLI. The thermal radiation, separator conduction, and residual gas conduction are considered. Rfoam is the foam thermal resistance. T WBT , T VCS, and T CBT are the temperatures of the warm boundary, the VCS, and the cold boundary, respectively. The heat transfer between the VCS and the hydrogen vapor flow in the VCS tube is treated as convection,   (4) Qsen + conv = TVCS − T vapor Rconve As the hydrogen vapor has an apparent temperature gradient along the flow path, the average temperature of the vapor is used in Eq. (4). The T vapor is the average of hydrogen vapor’s inlet temperature and outlet temperature. Rconve is the convective heat transfer resistance,  Rconve = d (λNuA) (5) where d represents the VCS tube’s diameter, and λ is the hydrogen vapor’s thermal conductivity. Nu is the Nusselt number, which takes a fixed value of 3.66 as the hydrogen vapor flows slowly in the tube. The net heat flux into the tank boiling the rest liquid hydrogen, assuming the liquid is saturated. Qnet = mhfg

(6)

where hfg is the latent heat of hydrogen vaporization. Subsequently, the hydrogen vapor enters the VCS tube where it warms up and performances POC partially. Therefore, enthalpies change is expressed by the enthalpy difference between the hydrogen vapor flowing out and into the VCS tube. The energy balance equation in steady state written as Eq. (7) by neglecting the changes of the kinetic energy and potential energy.  o    p p xo,post + Hout (1 − xo,post ) − Hino xo,pre + Hin (1 − xo,pre ) Qsen + conv = mh = Hout (7) p

p

o , H , H o and H are the enthalpies of the orthohydrogen and parahydrogen where Hout out in in at the outlet and inlet of the VCS tube, respectively.

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2.2 Para-Ortho Hydrogen Conversion Model The vapor discharged from the tank consists entirely of parahydrogen. A part of the parahydrogen is converted to orthohydrogen as it flows through the converter where a catalyst is installed. The conversion efficiency is assumed to be 100%. The converted hydrogen is equilibrium hydrogen at the corresponding conversion temperature. After leaving the converter, the isomer concentrations of the converted hydrogen are assumed unchanged until the vapor is exhausted. The isomer concentrations of the equilibrium hydrogen are functions of the temperature only, which complies the Boltzmann distribution law [7], 42θr r 1 + 5 exp(− 6θT r ) + 9 exp(− 20θ 1 − xo,eq NP T ) + 13 exp(− T ) + . . . = = 2θr 12θr 30θr NO xo,eq 3(3 exp(− T ) + 7 exp(− T ) + 11 exp(− T ) + . . .

(8)

Np and No are the molecular numbers of the parahydrogen and orthohydrogen components, respectively. The spinning temperature of hydrogen molecular θ r is 84.837[K] [7]. χ o, eq is the orthohydrogen concentration at an equilibrium state. The conversion heat flux by POC equals the product of the enthalpy difference between the orthohydrogen and the parahydrogen under isothermal conditions and mass flow rate. Qconv = m(xo,post − xo,pre )(ho − hp )

(9)

where x o, post and x o, pre are the concentrations of O-H2 after and before conversion, respectively; m and h are the mass flow rate and the specific enthalpy of the hydrogen vapor, respectively. The thermophysical properties such as specific heat, specific enthalpy, and conversion heat for pure fluids are retrieved from the NIST code REFPROP and relevant literature [8]. The mixture properties are linearly calculated by the isomer concentrations. In the model, thermal resistance parameters (RMLI,OUT , Rfoam , RMLI,IN , Rconve ) are known. And the enthalpies and the concentrations of P-H2 and O-H2 are retrieved from REFPROP according to temperature. Six unknown parameters (Qtotal , Qnet , Qsen+conv , T VCS , T out and m) can be resolved by Eqs. (1)–(7). The pressure difference inside and outside the tank keeps 170 kPa regardless of the liquid hydrogen flushing rate. A transient model is solved through iteration until the temperature reaches steady state. The simulation results for analysis are extracted from the steady state stage.

3 Results and Discussion 3.1 Effect of POC on Temperature Distribution and Heat Flux The temperature profiles in the composite insulation structure were investigated for three different configurations: MLI only, MLI/VCS, and MLI/VCS/POC, as shown in Fig. 2. The x coordinate of 0 mm denotes the foam’s cold surface, and the x coordinate of 71.25 mm represents the MLI’s outmost end (warmest surface). The interface between

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the foam and the MLI is at the x coordinate of 35 mm. The blue vertical dash line represents the position of the VCS. The warm boundary temperature remains 235 K. It can be found that the temperature differences between the three configurations mainly exist in the MLI. The maximum temperature difference occurs at the VCS location. Compared to the MLI/VCS configuration, The addition of POC further enhances the cooling performance of the VCS, in which the temperature dropped from 92.1 K to 85.3 K, by approximately 20.6%. As the thermal resistence of the MLI is much larger than that of the Foam at the on-orbit satage, there is almost no temperature gradient established in the Foam.

Fig. 2. Temperature profile of foam/MLI under different configurations

The significant cooling effect of the VCS would affect the heat flux allocation in the composite insulation structure. As the warm boundary and the cold boundary temperatures remain constant and with the presence of POC, the temperature difference between the VCS and the warm boundary increases, while that between the VCS and the cold boundary decreases. These significant temperature changes resulted in a total heat flux increase and a net heat flux decrease. 3.2 Effect of POC on Net Heat Flux with the Position of VCS’s Changing The net heat flux is the most concerned evaluation parameter, which reflects the thermal protection performance of the whole composite insulation. Figure 3 shows the net heat flux distribution with or without POC for different embed positions of the VCS. If the POC is absent, the midpoint of the MLI thickness is the optimal position to place the VCS, where the net heat flux is minimized [9]. However, with the presence of the POC, the optimum position of the VCS moves innerward from the thickness midpoint of the MLI. The distance between the midpoint and optimal position increases with increasing the warm boundary temperature. The net heat flux increases as the warm boundary temperature rises for all configurations whether the POC is present or not. The hydrogen vapor flow provides greater POC cooling power at higher boundary temperatures. On the other hand, a colder VCS results in a more significant drop in net heat flux. The net heat flux comparison on the optimal configuration with or without POC is shown in Fig. 4. It can be found that the net heat flux decreases by ~10% in the full range

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Fig. 3. Net heat flux comparison with or Fig. 4. Net heat flux comparison for an without POC for different positions of VCS optimum position of VCS with or without POC

of 100 K~305 K. In addition, the porabola shape of the net heat flux curve in Fig. 3 indicates that the value change near the optimum location of the VCS is small. It means an value around the vertex is acceptable rather than the exact optimal value. 3.3 Relationship Between Sensible Heat Flux and Conversion Heat Flux The contribution of POC to the overtall cooling performance of the VCS can be evaluated through the ratio of conversion heat flux to sensible heat flux. The former is obtained by Eq. (9). The latter is calculated by Eq. (4) and Eq. (9). For POC at constant mass flow in the VCS tube, the conversion heat and the orthohydrogen concentration are the determining factors from Eq. (9), which are positively and negatively correlated with the conversion temperature, respectively [10]. Therefore, the conversion heat flux profile is a convex curve with a maximum stationary point. The specific heat is the only decisive factor for the sensible heat of hydrogen vapor, so the sensible heat flux monotonically increases with the increasing VCS temperature. The Ratio of the conversion heat to the sensible heat at different warm boundary temperatures are given in Fig. 5. The ratio profiles are unimodal versus the location of VCS. It is found the the maximum ratios are close to each other in a wide range of the warm boundary temperatures, but the peeks move innerward as the temperature rises. According to Fig. 2, the temperature profile in the MLI blanket in the thickness direction has a linear behavior. Therefore, the position of the VCS determines the temperature of the hydrogen vapor after being heated in the tubes and passing the converter successively at steady state. The ratio of the conversion heat to the sensible heat versus the conversion temperature is presented in Fig. 6. Unlike that in Fig. 5, the ratio peeks appear at the approximately same conversion temperature of 74 K for all the warm boundary temperatures.

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Fig. 6. Ratio of conversion heat to sensible heat versus conversion temperature

4 Conclusions A one-dimensional model for predicting the temperature distribution and heat flux changes of a composite insulation under the influence of para-ortho hydrogen conversion was structured. The impacts of para-ortho conversion cooling effect on the optimum position of vapor-cooled shield and the sensible heat potential have been investigated based on the model. The following conclusions can be drawn: 1) Para-ortho conversion enhances the cooling performance of vapor-cooled shield by converting heat energy to molecular internal energy. Adding a para-ortho converter to the shield reduces the net heat flux through the composite insulation structure. 2) The optimal position of the vapor-cooled shield for minimum net heat flow moves towards the cold boundary due to the para-ortho hydrogen conversion. A reduction in net heat flux by approximately 10% can be achieved by introducing the para-ortho conversion. Acknowledgments. This work is supported by the National Natural Science Foundation of China (No. 51936006) and the Shanghai R&D public service platform project (No. 19DZ2291400).

References 1. McLean, C., Mustafi, S., Walls, L., Pitchford, B., Wollen, M., Schmidt, J.: Simple, robust cryogenic propellant depot for near term applications. In: 2011 Aerospace Conference, pp. 1– 24. , Big Sky, MT, USA (2011) 2. Nast, T.C.: Investigation of a para-ortho hydrogen reactor for application to spacecraft sensor cooling. Lockheed Missiles and Space Co, California (1987) 3. Liggett, M.W.: Space-based LH2 propellant storage system: subscale ground testing results. Cryogenics 33, 438–442 (1993) 4. Bliesner, R.M., Leachman, J.W., Adam, P.M.: Parahydrogen–orthohydrogen conversion for enhanced vapor-cooled shielding of liquid oxygen tanks. J. Thermophys. Heat Transfer 28, 717–723 (2014)

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5. Martin, J.J., Hastings, L.: Large-scale liquid hydrogentesting of a variable density multilayer insulation with a foam substrate. Marshall Space Flight Center (2001) 6. McIntosh, G.E.: Layer by layer MLI calculation using a separated mode equation. In: Kittel, P. (ed.) Advances in Cryogenic Engineering, pp. 1683–1690. Springer, US, Boston, MA (1994) 7. Bliesner, R.M.: Parahydrogen-Orthohydrogen Conversion For Boil-off Reduction From Space Stage Fuel System (2013) 8. Younglove, B.A.: Interactive Fortran program to calculate thermophysical properties of six fluids (1982) 9. Jiang, W.B., Zuo, Z.Q., Huang, Y.H., Wang, B., Sun, P.J., Li, P.: Coupling optimization of composite insulation and vapor-cooled shield for on-orbit cryogenic storage tank. Cryogenics 96, 90–98 (2018) 10. Roy, W., Harkness, W.: Edwards deming: the equilibrium of para and ortho hydrogen. J. Am. Chem. Soc. 54, 2850–2852 (1932)

Modeling and Parametric Study of Ice Interface Growth During Microdroplet Impinging on Different Micro-Textures Xiaoqing Zhou, Guang Yang, Chunyu Li(B)

, and Jingyi Wu

Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, No.800 Dongchuan Road, Shanghai 200240, China [email protected]

Abstract. Owing to supercooled droplets endangering aviation aircraft are microscale, the icing and impingement of microdroplet on microstructure (i.e., microscale-confined condensation) is still vague. The mechanism of ice-water phase front growth on free interface of impinged microdroplets, has gained much attention. This paper aims at retarding freezing on ice phase interface propulsion of microdroplets, influence of spatial morphology of micro-texture on dendritic front growth speed of ice-water interface is revealed. Two types of micro-textures: micro-terraces and micro-columns, are considered. The dynamic method of Coupled Level Set and Volume of Fluid (CLSVOF), and Enthalpy-porosity technique is coupled for determination of the ice-water interface. The reduction influence of breakage on droplet freezing velocity have been uncovered. The possibility of regulating micro-texture arrays to slow freezing of dual micro-droplets was analyzed. Results show that reducing inclination and increasing intervals can reduce freezing rate and droplet contact time, but when inclination reduced to 70°, reducing inclination influences little. Inclination reduction can prompt droplet impingement fracture, and prompting breakage can slow down freezing (48.7% in 8 m/s). The higher the degree of undercooling, the less likely for breakage. Keywords: microdroplet · CLSVOF · ice front growth · breakage

1 Introduction For the challenge of burgeoning anti-icing applications in the space vehicles and airplanes territory, the high-end micro/nano-fabrication of the superhydrophobic textures provide a promising prospect owing to its distinctive quick removal and refresh ability [1]. Microstructures can play a significant role as a passive anti-icing structure: a) Reduce ice adhesion strength down to 1.7 kPa [2]; b) Lower nucleation velocity [3]; c) Lower propagation rate [4]. So far, various microstructures have been adopted [5]. At present, there is no quantitative description of the theory of microstructure morphology properties and ice resistance. According to the latest FAR (Federal Aviation Regulations) Appendix O to Part 25, the magnitudes of supercooled cloud droplets in aviation environment vary from 20 to 2000 μm [6]. Micron droplets and millimeter droplets all must be © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 435–441, 2023. https://doi.org/10.1007/978-981-99-6128-3_55

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considered [7]. Existing study mostly concentrated on millimeter-sized droplets, while in the circumstance that micron topology and droplet own the same scale, the difference of wetting conditions and the interfacial tension of micron droplet cannot be negligible [8]. However, owing to complicated experiment setup, little researches concentrated on microdroplets, so simulation work is necessary. This paper focused on two problems: Quantitative description of the influence of structural characteristics on icing kinetics; Considering impact behavior on microscale. A simulation model based on Coupled level set (LS) and VOF (CLSVOF) method was set up to explore the ice interface growth velocity during microdroplet impinging on microstructures.

2 Materials, Simulation Setup, and Methods Verification 2.1 Theoretical Model This article adopted CLSVOF model to manipulate the situation of multiphase (ice, water, and air) coexistence, which is a surface tracking method commonly employed in multiphase flow simulation suitable for obvious interface flow. Continuum Surface Force (CSF) model adopted by Brackbill et al.[9] was employed herein to calculate the interplay during gas-liquid two phase flow. The solidification/melting model regards the ice area as a porous medium. Internal flow of the phase transition region meets The Darcy’s Law of flow in porous media and conforms to the Carman-Kozeny hypothesis. 2.2 Simulation Setup In the following computational model, micro-pillars and micro-terraces are two types of typical microstructures. Nomenclature of different microstructures can be concluded as D, H, S, and L. D is the top width, H is the height of micropillars, S is the interval between two micropillars, and L is the bottom width (in micropillars, D = L), as shown in Fig. 1. The value of θ (tan−1 ((D − L)/2H)) reflected the inclination of the terraces, and micro-pillar is the micro-terrace which own an inclination of 90°.

Fig. 1. Some quantitative parameters expressed in computational model, θ reflected inclination.

To conduct the numerical analysis, ANSYS-Fluent 19.2 was employed. As for specific model settings, explicit formulation was employed in CLSVOF model. This paper used the CSF model for the surface tension force reckoning and geo-reconstruction

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model for interface capture. Huge difference on heat-transfer property during phase change (liquid water: 0.63 W/(m·K); solid ice: 2.73 W/(m·K)) can influence substrate heat transfer deeply. In our model, UDF was employed to distinguish thermal conductivity before, during and after icing. The model validation was decoupled into dynamic verification and icing verification. The dynamics was validated through the comparison among our early simulation and other literatures’ experiments. Deviation of the relation on time and spreading factor between experiment [10], and simulation is less than 5%. Relative deviation between simulation and experiment is 0.1–0.2 ms, which could mainly be owing to the simplification of air resistance and wall friction. For icing validation, the experiment and simulation results of single droplet ice phase interface evolve own great consistency. In our computational model, microdroplet with diameter (Dd ) of 100 μm impinges into supercooled micro-structured surface, with initial velocity of Vo . Contact angle of superhydrophobic surface is 160°. The droplet will be patched on-top the middle of two micro-terraces’/micro-pillars’ interval. As shown in Fig. 2, grids were locally encrypted. Symmetric structure was adopted. External environment boundary was set as pressureoutlet, and micro-terraces and micro-pillars were set as adiabatic wall boundary. Distance from the circle center to the boundary was higher than 5 Dd in order to avoid boundary restriction.

Fig. 2. Model construction and mesh division of the computational model

2.3 Performance Criterion Two characters representing the dynamic icing kinetics are showing below: 1. The mushy zone area fraction A*: the volume proportion of ice slurry mixture; Mushy zone means zone owning liquid fraction between 0–0.99 [8]. If we keep the droplet area constant, the A* can be defined as the ice accumulation rate. A∗ = Aice /Ad = Aice /π(Dd /2)2 ,

(1)

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herein, Aice is the mushy zone area, Ad is the droplet area, Dd is the droplet initial diameter. Mushy zone possesses irregularity and mobility, as shown in Fig. 3. The mushy zone keeps piling up near the wall, its growth rate can reflect the droplet freezing velocity. 2. The spreading ratio D*: the trend of maximum diffusion diameter with time. D∗ = Dspread /Dd ,

(2)

herein, Dspread is the spreading diameter during droplet diffusion. D* reflects the droplet’s kinetic dynamic behaviors, expressing the maximum limit of kinetic energy converting to surface energy. However, owing to the irregularity and mobility of mushy zone, the variation amplitude of D* may wave a little, and this was demonstrated to be insignificant and have little influence.

Fig. 3. Illustration of ice interface distribution in mushy zone in the whole droplet (left: when impacting on micro-terraces; right: when impacting on micro-pillars).

3 Results and Discussion Results and discussions are carried out from the following three aspects: different microcolumn shape/different impact velocity. 3.1 The Influence of Different Micro-Column Shape For the influence of different micro-column shapes, micropillar and three micro-terraces are considered. One 100 μm droplet impacted on the 263 K supercooled surface with velocity of 5 m/s. Owing to the superhydrophobicity of micro structured surface, droplet falling with a high velocity will rebound back after spread. With the decrease of cylinder’s inclination, the time when droplet reach maximum spreading (i.e., highest point of parabola) would decrease. And the maximum D* would decrease. As shown in the left graph of Fig. 4: Owing to retardation of the slope of micro-terraces, droplet spread completely on micro-pillars at 70 μs, while spreading diameter on micro-terraces is less, and more drops gather. At 120 μs, droplet on the micro-pillars shrink back in a complete form, while on micro-terraces may break up into three pieces. As shown in the right graph of Fig. 4, in the case of D16 H40 S20 L56, droplet would break up into three pieces and finally coalesced.

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Fig. 4. Differences in dynamic behaviors of four microstructures

The shape of microstructures can obviously influence the ice growth velocity. The less the inclination, the slower the freezing rate. However, the manipulation of inclination on slowing down icing is limited. When inclination reduced down to 70°, reducing inclination influences little. The “D16 H40 S20 L36” surface performs worse but the “D16 H40 S20 L46” and “D16 H40 S20 L56” surfaces are very nearly the same. Droplet impinged on the “D16 H40 S20 L56” surface would break up into three pieces and then merged again. The red point represents the time when droplet jumped off the surface, however, droplet on the “D16 H40 S20 L56” surfaces jumped off at 100 μs and then secondary impingement appeared at 140 μs, as shown in Fig. 5.

Fig. 5. Differences in anti-icing performance of four microstructure

3.2 The Influence of Different Impact Velocity These different cases occurred in the surface of D16 H40 S20 L46. There is an apparent coordination between impact velocity and freezing heat transfer, the higher the velocity, the faster the icing process, as shown in Fig. 6. In the case that V = 8 m/s, two breakups were observed at 15 μs and 95 μs. The heat transfer slope slowed down previous to 95 μs, which was weakened by 48.7%. The reasons of this weakening can be concluded as two aspects: a) The whirlpool in the fluid

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domain is strengthened owning to droplet breakup (flow field disturbance at 15 μs is much stronger than at 95 μs, as shown in Fig. 6). b) During breakup, energy is conserved, kinetic energy is converted to surface energy, and flow velocity near boundary decreases and therefore heat transfer is weakened. To sum up, droplets breakup can bring back turbulence gain and kinetics rejection. As long as micro-structure is properly regulated, the weigh between these consequences leads to a final reduction. In addition, contact time varies nonmonotonically with velocity. With the slowing down of the droplet impingement, the contact time is not always lowered.

Fig. 6. Differences in anti-icing performance and flow field disturbance characteristics of various impact velocity.

4 Conclusions This paper concentrated on modeling ice interface growth through parametric study. Some designs guiding efficient anti-icing micro-textures fabrication have been declared: 1. Reducing inclination can prompt fracture of droplet impingement. Prompting breakage can slow freezing (48.7% in 8 m/s, compared to the case without breakage). 2. Reducing each micro-terrace’s inclination (D/L) and increasing each two microterraces’ intervals (S) can reduce freezing rate and droplet contact time. When inclination reaches 70° ((i.e., D16 H40 S20 L46)), reducing inclination influences little. 3. The larger the impingement velocity, the quicker the freezing rate. Droplet impact velocity can influence freezing rate decisively, thus find means to restrain droplets impaction freezing at high velocity is so important. Acknowledgments. This work was supported by Startup Fund for Young Faculty at SJTU (SFYF at SJTU, No. 22X010500267).

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References 1. Zhao, Y., Yang, C.: Retarded condensate freezing propagation on superhydrophobic surfaces patterned with micropillars. Appl. Phys. Lett. 108, 061605 (2016) 2. Feng, S., Delannoy, J., Malod, A., Zheng, H., Quéré, D., Wang, Z.: Tip-induced flipping of droplets on Janus pillars: from local reconfiguration to global transport. Sci. Adv. 6, 4540– 4548 (2020) 3. Alizadeh, A., et al.: Dynamics of ice nucleation on water repellent surfaces. Langmuir. 28, 3180–3186 (2012) 4. Jin, Y., He, Z., Guo, Q., Wang, J.: Control of ice propagation by using polyelectrolyte multilayer coatings. Angew Chem. Int. Ed. 56, 11436–11439 (2017) 5. Chu, F., Wu, X., Wang, L.: Dynamic melting of freezing droplets on ultraslippery superhydrophobic surfaces. ACS Appl. Mater. Inter. 9, 8420–8425 (2017) 6. Peng, Q., et al.: Forced jumping and coalescence-induced sweeping enhanced the dropwise condensation on hierarchically microgrooved superhydrophobic surface. Appl. Phys. Lett. 114, 133106 (2019) 7. Attarzadeh, R., Dolatabadi, A.: Icephobic performance of superhydrophobic coatings: a numerical analysis. Int. J. Heat. Mass. Tran. 136, 1327–1337 (2019) 8. Pan, Y., Shi, K., Duan, X., Greg, F.N.: Experimental investigation of water droplet impact and freezing onmicropatterned stainless steel surfaces with varying wettabilities. Int. J. Heat. Mass. Tran. 129, 953–964 (2019) 9. Brackbill, J.U., Kothe, D.B., Zemach, C.: A continuum method for modeling surface tension. J. Comput. Phys. 100, 335–354 (1992) 10. Zhang, X., Wu, X., Min, J., Liu, X.: Modelling of sessile water droplet shape evolution during freezing with consideration of supercooling effect. Appl. Therm. Eng. 125, 644–651 (2017)

The Effect of Fill Ratio on the Performance and Flow Regime for Long-Distance Helium Pulsating Heat Pipes Logan Kossel(B) , John Pfotenhauer, and Franklin Miller Department of Mechanical Engineering, University of Wisconsin– Madison, Madison, WI, USA [email protected]

Abstract. Helium pulsating heat pipes (PHPs) are an emergent heat transfer technology that can provide efficient, long-distance cooling power for numerous lowtemperature technologies, such as superconducting magnets and space telescopes. An experimental facility is developed to test the performance of a vertically oriented helium PHP with an adiabatic length of 1.25 m. The PHP in these experiments consists of 14 parallel stainless-steel tubes with an inner diameter of 0.5 mm. Each end of the pulsating heat pipe is fixed to a thin copper plate where the condenser (upper) end is cooled by a 4 K cryocooler with a 1.5 W cooling capacity, while a resistive heater provides a prescribed heat load to the evaporator (lower) end. Progressive heat load experiments were performed for fill ratios 37.3% and 70.78%, from which phase and flow regime information at multiple locations is predicted. The maximum performances for each test were 273 kW/m-K at 0.23 W and 136 kW/m-K at 0.43 W for the 37.3% and 70.78% fill ratios, respectively. Keywords: Pulsating heat pipes · Helium · Two-phase flow · Experimental study

1 Introduction Pulsating Heat Pipes (PHPs) are a unique heat pipe variant that have demonstrated numerous advantages over traditional heat pipes, such as decreased weight and better performance. PHPs consist of small-diameter capillary channels assembled into a serpentine shape and form a closed loop, with multiple channels running parallel to each other and turning at each end. Like other heat pipe designs, a PHP has three sections partitioned by heat flow direction: The evaporator, adiabatic, and condenser sections. The evaporator section absorbs heat, and the condenser section expels the heat picked up by the evaporator to a cooling source. In between is the adiabatic section, where the fluid contained in the PHP does not experience heat transfer with the surroundings and travels between the other two sections. Moreover, PHPs using helium as the working fluid have many potential applications by augmenting the cooling distance of cryocoolers without comprising their cooling capacities. One of the most important aspects of a PHP’s design is the inner diameter of the channels. Ideally, the channels should be sized small enough such that the two-phase © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 442–448, 2023. https://doi.org/10.1007/978-981-99-6128-3_56

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working fluid is forced into a plug-slug flow regime, where distinct sections of vapor and liquid form in an alternating order along the length of the channels. This condition is governed by the fluid’s Bond number: The square root of the ratio of the buoyant and surface tension forces, defined in Eq. (1).   Fb (ρl − ρv )gD2 (1) = Bo = Fs σ When the Bond number is less than 2, the surface tension force is strong enough to statically hold the vapor bubbles against the buoyant force [1], creating the alternating train of liquid and vapor sections that defines the plug-slug flow regime. In practice, the phase and flow regime of the working fluid may differ between the three sections and may also depend on numerous parameters such as heat flux, fill ratio, number of turns, and overall length, which makes it difficult to predict a PHP’s behavior for a given set of conditions. Despite pulsating heat pipes being invented more than 30 years ago by Akachi [2], research interest in low-temperature PHPs utilizing helium as a working fluid has picked up only in the last decade. For example, studies from Fonseca, Pfotenhauer, and Miller [3, 4] showed that the effective conductance for 42-turn helium PHPs of different adiabatic lengths (300 mm and 1000 mm) was the same over a range of heat loads at each PHP’s respective optimal fill ratio. The best performance (150 kW/m-K) was observed for the PHP with a 1000 mm adiabatic length at a 58% fill ratio and 0.75 W heat load. This current study of extended-length helium PHPs longer than 1 m is motivated by the supposed length-independence of helium PHPs recognized by Fonseca, Pfotenhauer, and Miller [3, 4]. The prospect of highly effective, long-distance heat transfer via helium pulsating heat pipes is appealing to several space and ground applications such as large liquid fuel storage containers, next-generation infrared space telescopes, and superconducting magnet cooling.

2 Experimental Set-Up An experimental facility was built to accommodate helium pulsating heat pipes with adiabatic lengths up to 1.75 m. This facility is comprised of the PHP test rig contained within a large vacuum chamber with a 2 m diameter and 2.5 m depth. Inside the vacuum chamber, several significant components are fixed to the two stages of the cryocooler: Heat exchangers and a large radiation shield on the first stage and the PHP on the second stage. The purpose of the first stage heat exchangers is to pre-cool room temperature helium before entering the PHP, which considerably reduces the heat load to the second stage when filling. Likewise, the radiation shield aims to re-direct the heat radiating from the vacuum chamber at ambient temperature to the cryocooler’s first stage (which has a much larger cooling power) instead of to the second stage. The shield is made from aluminum sheets and is nearly 2 m long. Additionally, a 20-layer MLI blanket covers the shield to further reduce radiation heat transfer. Results from significant thermal modeling efforts were employed to design these components, with the goal of minimizing the radiation and conduction heat transfer to the cryocooler’s second stage. Figure 1 shows a CAD schematic of these components in relation to each other.

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Fig. 1. The CAD model of the long-distance helium PHP test rig, showing the cryocooler and the experimental components mounted to each stage.

The PHP in this study is built from 0.5 mm ID 304 stainless-steel tubes, with an adiabatic length of 1.25 m, evaporator and condenser lengths of 90 mm, and 14 parallel tubes (or 7 turns). The evaporator and condenser sections are thin copper plates to which the capillary tubing is soldered. Furthermore, the PHP is vertically oriented and bottomheated using a resistive heater. Temperature measurements are made with Lakeshore Cernox CX-1050 sensors (±5 mK) at seven locations on the PHP. Two sensors are fixed to each endplate, averaged to eliminate spatial dependence and used to calculate the conductivity of the PHP. Additionally, three sensors are placed strategically on the center-most neighboring tubes in the adiabatic section. Two of the three adiabatic sensors are placed halfway between the evaporator and condenser sections on adjacent tubes, and the final sensor is located on one of the tubes, a quarter distance between the condenser and evaporator. Moreover, pressures are measured using Omega PX419 transducers (±275 Pa) at both the evaporator and condenser sections of the PHP. Finally, the fill ratio, defined as the liquid volume fraction of the PHP, is determined by an additional pressure measurement of an external buffer tank and precise volume measurements of all plumbing components.

3 Results A progressive heat load experiment was performed for 37.3% and 70.78% fill ratios. Starting from zero, the heat load on the evaporator is incremented by 20 mW, and the measurements are recorded for 15 min per heat load. Finally, the experiment was concluded for the 37.3% fill ratio test when the performance had significantly deteriorated, and for the 70.78% fill ratio when the PHP pressure measurements began to approach the transducers’ maximum allowable pressure (345 kPa). Figure 2 shows the two tests’ averaged effective conductivity and temperature measurements. The effective conductivity is determined using the averaged PHP end temperatures along with the known PHP geometry using the conduction analogy in Eq. (2). Where L is the adiabatic length

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and Ac is the cross-sectional area of PHP. keff =

˙ QL   NAc T evap − T cond

(2)

In Fig. 2, each plot is separated into heat load regions depending on the observed temperature and performance trends. These regions are corroborated by the temperatures vs. time plots in Fig. 3 and the phase diagrams in Fig. 4, and are examined in Sect. 4. Furthermore, since the experimental error is much smaller than the range of measured temperatures for each sensor, the error bars in Fig. 2 represent the standard deviation of the measured temperatures over the test period.

(a)

(b)

Fig. 2. Thermal conductivity and all temperature measurements (evaporator, condenser, and adiabatic sections) versus prescribed evaporator heat load for (a) 37.3% fill ratio and (b) 70.78% fill ratio, averaged over the test period of 15 min.

Instead of averaged values, Fig. 3 shows the raw evaporator and condenser temperature measurements for the experiment duration for each fill ratio. The incrementally increasing heat load is also displayed. These plots are split into the same heat load regions as Fig. 2. Moreover, the averaged evaporator/condenser temperatures and pressures are paired to determine the approximate thermodynamic state of each end of the PHP, visualized using a T-v diagram in Fig. 4. Since temperature and pressure data can only yield specific volume values outside the vapor dome or on the saturated lines, some assumptions are made to determine the two-phase specific volumes for the appropriate evaporator data points. For evaporator states that align with the saturated vapor line, the saturation temperature at the measured pressure is used, and it is assumed that the condenser is filled with saturated liquid. This constrains the specific volume of the remaining fluid to a value inside the vapor dome. A helium equation of state [5] is used to find the specific volume for all other evaporator and condenser points.

4 Discussion Phase and flow regime information can be estimated from the temperature, pressure, and heat load data presented in Figs. 2, 3 and 4. In Fig. 2 (a) and Fig. 3 (a), the 37.3% fill ratio test is split into regions identified by adiabatic and end temperature trends. In

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(a)

(b)

Fig. 3. Evaporator and condenser temperatures and the progressively incrementing heat load, versus time for the entire duration of the (a) 37.3% and (b) 70.78% fill ratio experiments. 37.3% Fill Ratio T-v Diagram

8.0

6.0 5.0

150 kPa 100 kPa

4.0

50 kPa

20 kPa

6.0 5.0

150 kPa 100 kPa

4.0

50 kPa

20 kPa

3.0

3.0 0.05

2.0 0.005

Condenser Evaporator Critical Volume/Pressure

7.0

Temperature [K]

Temperature [K]

7.0

70.78% Fill Ratio T-v Diagram

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Condenser Evaporator Critical Volume/Pressure

0.01

0.1

0.2

0.1

Specific Volume [m3/kg]

(a)

0.5

0.05

1

2.0 0.005

0.01

0.1

0.2

0.1

Specific Volume [m3/kg]

0.5

1

(b)

Fig. 4. Temperature-volume phase diagrams derived from each PHP end’s averaged temperature and pressure data for (a) 37.3% and (b) 70.78% fill ratio.

region I, both adjacent adiabatic tubes measured nearly identical temperatures above the condenser but below the evaporator temperature, and there were minimal temperature oscillations. This region spans heat loads from 0 W to 0.27 W, and displays the most stable evaporator temperatures and the best performance for the 37.3% fill ratio test: 273 kW/m-K at 0.23 W. The fluid appears to be two-phase throughout the PHP in this region, as observed in Fig. 4 (a). The beginning of region II is defined as the heat load at which the evaporator becomes superheated, as evident from the drastic increase in temperature oscillations in Fig. 3 (a) and location on the phase diagram in Fig. 4 (a). In this region, the performance begins to decrease due to the evaporator temperature becoming unstable and increasing more with increasing heat load, which in turn, is the result of the evaporator fluid becoming superheated. An interesting phenomenon occurs that defines region III: the adjacent adiabatic tube temperatures that previously matched begin to diverge at a distinct heat load value, where one tube begins to approach, and eventually match, the evaporator temperature while the other tube continues its previous trend. In this region, the liquid fraction of the fluid in one tube decreases until it becomes superheated, while the adjacent tube remains two-phase. Furthermore, the flow may be transitioning from oscillatory to circulatory

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flow in region III. In regions I and II, the flow is likely oscillatory, with the plugs and slugs oscillating in place and the high performance a result of shuttle heat transfer. Conversely, when the adjacent adiabatic tube temperatures diverge in temperature, as seen in regions IV and V, the flow is likely circulatory, meaning that a given fluid parcel will traverse the entirety of the PHP instead of oscillating in place. In an alternating order, the tubes in the adiabatic section are either warmer or colder, where the colder tubes carry fluid from the condenser to the evaporator, and the warmer tubes carry fluid in the opposite direction, establishing uni-directional flow, which has also been observed in PHP visualization studies [6, 7]. Finally, in region V, a steeper increase in the evaporator temperature is observed, signifying a more aggressive dry-out condition, whereas, in previous regions, the evaporator and matching adiabatic tube were likely drying out intermittently and recovering. The trends of the measured parameters from which flow regime and phase information can be deduced for the 70.78% fill ratio test are markedly different from the low fill ratio test. First, in Fig. 2 (b), the adjacent adiabatic tube measurements match with either the evaporator or condenser temperatures for all heat loads, indicating that the PHP fluid is always uni-directional. Like the 37.3% test analysis, Figs. 2 (b) and 3 (b) are split into heat load regions, albeit for different reasons. Starting from zero heat load, the fluid is two-phase and flowing uni-directionally in region A, where the most stable performance was observed for the 70.78% case. The onset of temperature oscillations in the evaporator section defines region B, despite the fluid remaining two-phase as indicated in Fig. 4 (b). In this region, the evaporator temperature oscillations could result from the Bond number criterion being approached, as opposed to phase change. The fluid in the evaporator and corresponding adiabatic tube is still two-phase, but the surface tension force becomes weaker compared to the buoyancy force on the bubbles, disrupting the plug-slug flow regime. Region C is defined by the point at which the Bond number of the evaporator fluid exceeds 2, confirming that the Bond number criterion is indeed violated. The best performance for the 70.78% fill ratio test occurs in region C, with an effective conductivity of 136 kW/m-K at 0.43 W. The evaporator fluid leaves the vapor dome in region D and becomes supercritical, as displayed in Fig. 4 (b). Despite this phase change, the PHP’s performance is not significantly impacted, a phenomenon that has been observed in previous helium PHP studies [3, 8].

5 Conclusions A long-distance helium pulsating heat pipe with an adiabatic length of 1.25 m was tested at 37.3% and 70.78% fill ratios with a progressively increasing heat load. Temperature and pressure measurements at various locations on the PHP showed contrasting fluid behavior between the high and low fill ratios, which are summarized as follows: 1. At a low fill ratio (37.3%), the PHP performs best with an oscillatory two-phase flow regime. The flow only becomes circulatory after the evaporator fluid is superheated, corresponding to a reduction in performance. 2. At a high fill ratio (70.78%), the PHP performs best with a circulatory two-phase flow regime, with oscillatory flow not observed at any heat load. The onset of temperature oscillations is due to a Bond number violation instead of phase change.

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3. The conditions at high heat load differ for the two fill ratios: Dry-out occurs and limits the performance for the low fill ratio (37.3%), while for the high fill ratio (70.78%), a supercritical operating state was observed where the performance is only slightly impacted. Future experimental research on long-distance helium pulsating heat pipes will consist of testing the complete spectrum of possible fill ratios for the 1.25 m PHP, and further extending the adiabatic length beyond 1.25 m. Acknowledgments. This work was supported by a NASA Space Technology Graduate Research Opportunities Award.

References 1. Khandekar, S.: Thermo-hydrodynamics of closed loop pulsating heat pipes. Dr.-Ing dissertation. University of Stuttgart (2004) 2. Akachi, H.: Structure of a heat pipe. U.S. Patent No. 4921041 (1990) 3. Fonseca, L.: Experimental characterization of cryogenic helium pulsating heat pipes. Ph.D. dissertation, University of Wisconsin – Madison (2016) 4. Fonseca, L., Pfotenhauer, J., Miller, F.: Short communication: thermal performance of a cryogenic helium pulsating heat pipe with three evaporator sections. Int. J. Heat Mass Transf. 123, 655–56 (2018) 5. Ortiz-Vega, D.: A new wide range equation of state for helium-4. Ph.D. dissertation, Texas A&M University (2013) 6. Xue, Z., Qu, W.: Experimental and theoretical research on a ammonia pulsating heat pipe: new full visualization of flow pattern and operating mechanism study. Int. J. Heat Mass Transf. 106, 149–66 (2017) 7. Karthikeyan, V., Khandekar, S., et. al.: Infrared thermography of a pulsating heat pipe: flow regimes and multiple steady states. Appl. Therm. Eng. 62, 470–80 (2014) 8. Li, M., Li, L., Xu, D.: Effect of filling ratio and orientation on the performance of a multiple turns helium pulsating heat pipe. Cryogenics 100, 62–68 (2019)

A Novel Cryogenic Loop Heat Pipe Structure and Preliminary Proof of the Concept Chenyang Zhao1,2 , Nanxi Li1(B) , Zhenhua Jiang1 , and Yinong Wu1(B) 1 Shanghai Institute of Technical Physics Chinese Academy of Sciences, Shanghai, China

{linanxi,wyn}@mail.sitp.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. Many space exploration components require a cryogenic working environment below 200 K. In order to avoid the electromagnetic interference and mechanical vibration of the cryocooler to the detector, an efficient cryogenic heat transfer device is required to realize the transfer from the cryocooler to the payload. Due to their advantages of high heat transfer efficiency, high flexibility, and flexible arrangement, loop heat pipes have become an effective way of cryogenic heat transfer in space. However, researches on cryogenic loop heat pipes (CLHP) are mainly experimental studies in the liquid nitrogen temperature zone, and few studies were carried out for CLHPs in the 150–200 K temperature zone. Therefore, a two-dimensional numerical model was established to calculate the performance of CLHP in different working temperatures. Additionally, an experimental and theoretical research on a 150–200 K CLHP charged with ethylene was carried out in this paper. Tesla valves were used to control the flow direction, the supercritical startup characteristics was experimentally investigated under different heating powers on the secondary evaporator. This work will facilitate the design of cryogenic thermal control for space and ground applications. Keywords: Cryogenic loop heat pipe · Startup characteristics · Thermal performance · Numerical simulation

1 Introduction Loop heat pipe (LHP) is a device that utilizes vapor-liquid two-phase heat exchange and flow of the working medium to achieve efficient heat transfer. The structure is shown in Fig. 1. The capillary wick inside the evaporator provides power for the circulation of the working medium, avoiding the use of moving parts and ensuring the long life and stability of the LHP. Vapor and liquid flow are separated in the thin-walled pipeline, rendering the LHP flexibility and long-distance heat transfer capability. These advantages make the LHP widely used in spacecraft thermal control [1–4]. With the development of aerospace industry, more and more spacecraft components require thermal control systems to create low temperature environment (below 200 K) for deep space exploration, earth observation and space superconductivity applications. Cryocoolers will produce mechanical vibration and electromagnetic interference to the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 449–455, 2023. https://doi.org/10.1007/978-981-99-6128-3_57

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Fig. 1. Schematic diagram of loop heat pipe structure.

detector, so they should be placed far away from the detector. The CLHP has become an effective way for solving the long-distance cryogenic heat transfer problem due to the advantages of high heat transfer efficiency, long heat transfer distance, flexible layout, lack of moving parts, and high reliability. However, the working medium of the CLHP is in the state of supercritical or superheated steam at room temperature, that is, there is no liquid working medium in the capillary wick of the evaporator, and the evaporator cannot be heated to start the CLHP. Therefore, it is necessary to solve the problem of pushing the liquid to the reservoir and evaporator of CLHP before starting. Since Triem T. Hoang et al. [5, 6] proposed the concept and experimental verification of CLHP in 2003, so far, many types of CLHPs have been developed. The development of CLHPs has been comprehensively reviewed, where the CLHPs are divided into five types mainly according to the system structure character [7]. Supercritical start-up of CLHP is achieved by heating the secondary evaporator in series or parallel. These structures inevitably increase the flow resistance of the working conditions or the complexity of the system structure. The Tesla valve was invented by Nikola Tesla in 1920 [8]. The structure is composed of straight and arc channels, with no moving parts, as shown in Fig. 2. Due to its unique structure, the pressure drop of the fluid passing through the Tesla valve in the forward direction is much less than that in the reverse direction. Therefore, Tesla valves can be used to control the direction of fluid flow. Tesla valves work well in a variety of scales, are stable and easy to machine. This gives Tesla valves an advantage in fluid flow design. Scholars have carried out a lot of CFD simulation and experimental research on Tesla valves. In order to obtain better pressure drop performance for the Tesla valve, Zhijiang Jin et al. [9] studied the influence of structural parameters such as hydraulic diameter, valve angle and inner curve radius on the pressure drop under a wide range of inlet velocities. Adrian R. Gamboa et al. [10] used 2D CFD to optimize the shape of the classic T45-R valve [11] in the Reynolds number range of 0-2000. The ratio of reverse to forward flow resistance of the optimized valve (diodicity) is increased by 25%, known as GMF (Gamboa, Morris and Forster) Tesla valve. Scott M. Thompson et al. [12] used various models to numerically simulate the multi-stage GMF Tesla valve and compared the results with experimental data. The results show that the second degree has a power-law relationship with the number of Tesla valves, and the k-kl-ω model

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and the SST k-ω model are closer to the experimental measurements. Therefore, the multi-stage GMF-type Tesla valve and SST k-ω model are adopted in this paper.

Fig. 2. Schematic diagram of a Tesla valve.

The Tesla valve cryogenic loop heat pipe structure is shown in Fig. 3, including the compensation chamber, evaporator, liquid line, vapor line, condenser, secondary heater, inlet Tesla valve and outlet Tesla valve. The specific start-up process is as follows: After the cooling source cools the condenser, and the working medium in the condenser condenses into liquid. When the secondary heater is turned on, the working medium evaporates into superheated vapor. Due to the large reverse flow resistance of the inlet Tesla valve, the vapor can only flow to the condenser and push the liquid flows into the liquid line through the outlet Tesla valve. When the secondary heater is turned of due to the large reverse flow resistance of the outlet Tesla valve, the working medium will be supplemented to the condenser through the inlet Tesla valve. By intermittently turning on and off the secondary heater, the liquid is gradually pushed to the compensation chamber and the evaporator. In this way, liquid phase cryogenic working medium could successfully wet the capillary wick of evaporator. So that the problem of no liquid phase cryogenic working medium wetting in the capillary wick of evaporator is solved.

Fig. 3. Schematic diagram of Tesla valve cryogenic loop heat pipe.

This paper proposes a novel structure to assist the CLHP start-up. Through the design of Tesla valve and secondary heater, the problem of no liquid working medium in the CLHP evaporator was solved, realizing the CLHP start-up. The one-way conduction capability and pressure drop of the Tesla valve were calculated by simulation to verify the feasibility of the novel CLHP.

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2 Description of the System The inlet adopts a 5-stage Tesla valve, and the inner diameter is the same as the vapor line, which is 3 mm. The outlet adopts an 8-stage Tesla valve, and the inner diameter is the same as the liquid line, which is 2 mm. A start-up strategy that starts heating when the temperature drops to 220 K was adopted. The CLHP experiment was carried out in a vacuum tank and the CLHP was covered with multi-layer thermal insulation materials to minimize parasitic heat. Therefore, when simulating and analyzing the Tesla valve components in CLHP, the energy exchange was ignored and only the momentum change was considered. The fluid flow was simulated using the SST k-ω Viscous Model using ANSYS FLU18.1 software. First, the grid independence of the import and export Tesla valves is verified. The grid number of the imported Tesla valve is 620,191, the grid number of the export Tesla valve is 568,804, and the error with the grid number of 150,000 is less than 2%. Then, the fluid working medium is set to ethylene. The inlet boundary is velocity-inlet, which is the flow velocity corresponding to the mass flow corresponding to the power. The outlet boundary is pressure-outlet. The coupled algorithm is used to couple the velocity and pressure fields. When the secondary heating power is 6W, the velocity contours of the vapor passing through the inlet Tesla valve in the forward and reverse direction are shown in Fig. 4 (a) and (b). The velocity contours of the liquid passing through the inlet Tesla valve in the forward and reverse direction are shown in Fig. 5 (a) and (b).

Fig. 4. The velocity contours of the vapor passing through the inlet Tesla valve.

Fig. 5. The velocity contours of the liquid passing through the outlet Tesla valve.

When the secondary heater is turned on, since the flow resistance of the vapor passing through the inlet Tesla valve in the reverse direction is greater than the resistance of the

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liquid passing through the outlet Tesla valve in the forward direction, as shown in Fig. 6 (a). The greater the heating power, the more pronounced the unidirectional characteristic. With a secondary heating power of more than 4 W, the Tesla valve can theoretically achieve unidirectional flow, the liquid working medium will flow out of the condenser through the outlet Tesla valve. When the secondary heater is turned off, since the flow resistance of the vapor passing through the inlet Tesla valve in the forward direction is less than the resistance passing through the outlet Tesla valve in the reverse direction, as shown in Fig. 6 (b), the vapor working medium will flow into the condenser through the inlet Tesla valve. The vapor working medium is condensed into liquid in the condenser. By intermittently turning on and off the secondary heater, after the evaporator capillary wick is wetted, the CLHP can be activated by heating the evaporator.

Fig. 6. Inlet and outlet Tesla valve flow resistance when secondary heater is on and off.

3 Feasibility Analysis In order to verify the feasibility of the CLHP and to obtain the steady state operating performance and capillary force range, an experimental test of a regular CLHP was carried out. The evaporator was pre-cooled, and the structural parameters and experimental results of the CLHP are shown in Table 1 and Fig. 7, respectively. The influence of Tesla valve on the operating pressure drop of the CLHP is shown in Fig. 8. The influence of the inlet and outlet Tesla valves on the pressure drop of the CLHP under the operating conditions of 10–70 W is within 3%. Therefore, theoretically, there is little effect on the heat transfer limit and temperature of the experiment. Limited by the power of the heating plate, the maximum experimental heat load was 70 W, and the CLHP can run in steady state at this time. The experimental result proves that the CLHP can provide at least 19018 Pa capillary force. Therefore, it can be guaranteed that the total pressure drop of the novel CLHP in the 10–60 W operating condition is less than the maximum capillary force.

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component

material

parameter

evaporator

stainless steel 20/24 × 90 mm (Inner Diameter/Outer Diameter × Length)

wick

stainless steel 8 × 75/20 ×8 0 mm (Inner Diameter × Length/Outer Diameter × Length) 1.5 μm (Diameter) 0.46 (Porosity)

compensation chamber stainless steel 20/24 × 60 mm (Inner Diameter/Outer Diameter × Length) 4 × 437 mm (Inner Diameter × Length)

condenser

copper

liquid line

stainless steel 2/3 × 1250 mm (Inner Diameter/Outer Diameter × Length)

vapor line

stainless steel 3/6 × 1200 mm (Inner Diameter/Outer Diameter × Length)

Fig. 7. Cryogenic loop heat pipe experimental results.

4 Summary and Outlook A novel structure to assist CLHP start-up is proposed, which utilizes the outlet and inlet Tesla valves to control the unidirectional flow of fluid. 1. The Tesla valve can achieve unidirectional flow. The greater the heating power, the more pronounced the unidirectional characteristic. It is recommended to use a secondary heating power of 4 W or more. 2. The pressure drop of the Tesla valve accounts for a small amount of the pressure drop of the CLHP, which can theoretically realize the supercritical start of the CLHP. Further experimental verification will be carried out in the future.

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Fig. 8. The influence of Tesla valve on the operating pressure drop of the cryogenic loop heat pipe.

References 1. Yang, T., Zhao, S., Gao, T., et al.: Effects of external heat flux during orbital period of spacecraft on operating characteristics of loop heat pipe. J. Phys. Conf. Ser. 1820, 012089 (2021) 2. Zhang, H.X., Mi, M., Miao, J.Y., et al.: Development and on-orbit operation of loop heat pipes on Chinese circumlunar return and reentry spacecraft. J. Mech. Sci. Technol. 31(6), 2597–2605 (2017) 3. Grob EW, 0148-7191 [R]: SAE Technical Paper (2001) 4. Goncharov, K., Nikitkin, M., Golovin, O., et al.: Loop heat pipes in thermal control systems for “OBZOR” Spacecraft. SAE Technical Papers (1995) 5. Hoang, T.T., O’connell, T.A., Khrustalev, D.K.: Development of a Flexible Advanced Loop Heat Pipe for Across-Gimbal Cryocooling (2003) 6. Hoang, T.T., O’connell, T.A., Ku, J.T., et al.: Design optimization of a hydrogen advanced Loop Heat Pipe for space-based IR sensor and detector cryocooling. In: Heaney, J.B., Burriesci, L.G. (eds.) Cryogenic Optical Systems and Instruments X, pp. 86–96 (2003) 7. Bai, L., Zhang, L., Lin, G., He, J., Wen, D.: Development of cryogenic loop heat pipes: a review and comparative analysis. Appl. Thermal Eng. 89, 180–191 (2015). https://doi.org/10. 1016/j.applthermaleng.2015.06.010 8. Nikola, T.: Valvular conduit (1920) 9. Jin, Z.-J., Gao, Z.-X., Chen, M.-R., et al.: Parametric study on Tesla valve with reverse flow for hydrogen decompression. Int. J. Hydrogen Energy 43(18), 8888–8896 (2018) 10. Gamboa, A.R., Morris, C.J., Forster, F.K.: Improvements in fixed-valve micropump performance through shape optimization of valves. J. Fluids Eng. 127(2), 339–346 (2005) 11. Forster, F.K., Bardell, R.L., Afromowitz, M.A., et al.: Design, fabrication and testing of fixed-valve micro-pumps. Asme-Publications-Fed 234, 39–44 (1995) 12. Thompson, S.M., Jamal, T., Paudel, B.J., et al.: Transitional and turbulent flow modeling in a tesla valve. In: ASME International Mechanical Engineering Congress and Exposition, American Society of Mechanical Engineers (2013)

Thermal Performance of a 140–200 K Grooved Heat Pipe Under Different Orientations Nanxi Li1 , Yongyan Li1,2 , Wenyi Zhao1 , Shiyue Wang1 , Zhenhua Jiang1 , and Yinong Wu1(B) 1 Shanghai Institute of Technical Physics, Chinese Academy of Sciences, Yutian Road,

Shanghai, China [email protected] 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. An omega-shaped grooved heat pipe charged with ethylene was fabricated to meet the heat dissipation requirement for the temperature range of 140–200 K over a long distance. The thermal performance of the heat pipe was experimentally investigated in both horizontal and vertical orientations. Results show that the ethylene heat pipe was able to transfer 20 W of heat over 0.9 m with a temperature difference below 5 K for the entire 140–200 K temperature range when placed horizontally. When positioned vertically with the condensation section above the evaporation section, the heat pipe operated successfully within the temperature range of 160–200 K under the same heat load. However, when the working temperature decreased to 140 K, the temperature difference became larger. Additionally, temperature oscillation was also observed at 140 K. The difference in the thermal performance of the heat pipe under the two operating orientations and the heat transfer anomaly were theoretically analyzed, and suggestions for future improvement of the heat pipe design were also proposed. Keywords: Grooved Heat Pipe · Thermal Performance · Ethylene · Temperature oscillation

1 Introduction Grooved heat pipes are highly efficient and reliable two-phase heat transfer devices that can transfer heat over a long distance with minimal temperature differences, therefore are widely used for space and many terrestrial applications such as the air conditioning systems [1–5]. In particular, optic systems in satellite infrared cameras function the best within the temperature range of 140–200 K. To create this low temperature working environment for the optic systems, cryogenic grooved heat pipes are leveraged to connect the cold source and the camera. Additionally, due to their exceptional heat transfer ability and reliability, cryogenic heat pipes can also be used for terrestrial applications such as cryogenic medical instruments, superconductors, cryocoolers, etc. Heat pipes have high heat transfer efficiency because enclosed the working fluid goes through evaporation on the hot end, called the evaporator, and condensation on the cold © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 456–462, 2023. https://doi.org/10.1007/978-981-99-6128-3_58

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end, called the condenser. The working fluid circulates in the heat pipe under the action of capillary force in the wick structure. There are many wick structures, for example rectangular grooves [6], trapezoidal grooves [7] and triangular grooves [8]. -shaped grooved heat pipes are one of the mostly used types of heat pipes. The unique structure can provide high capillary force and low pressure drop for liquid flow simultaneously. Much work has been done to study the performance of grooved heat pipes. Anand et al. [9] conducted a thorough review on the effect of inclination angle, charge amount, working temperature on the heat transfer capability of grooved heat pipes. However, the effect of these parameters on the heat transfer performance was not clarified. Tang et al. [10] investigated the isothermal performance of a micro-grooved heat pipe charged with ammonia. It was concluded that a lower charging mass would improve the isothermal performance when the heat pipe was placed vertically. However, the isothermal performance of the heat pipe in the horizontal direction was not tested. Yao et al. [11] experimentally investigated the heat transfer performance of a heat pipe under different inclination angles from 0° to 90°and found that a higher inclination angle would enhance the heat transfer performance because gravity assists the backflow of fluid. However, most of the works focus on room temperature heat pipes. The effects of large inclination angle and working temperature on the heat transfer performance of a cryogenic heat pipe were less reported. In order to understand the effect of working temperature and gravity on the performance of a cryogenic grooved heat pipe. An -shaped grooved heat pipe charged with ethylene as working fluid was fabricated to meet the heat dissipation requirement for the temperature range of 140–200 K over a long distance. The thermal performance of the heat pipe was experimentally investigated in both horizontal and vertical orientations. The findings may have guiding significance for the design of cryogenic heat pipe structures and the choosing of filling ratio for terrestrial applications.

2 Experimental Setup 2.1 Experimental Facility Description of the Heat Pipe. The heat pipe was extruded with aluminum, and its dimensions are listed in Table 1. Figure of Merits (FOM) are used to conveniently choose the most working fluid, and the expressions for the liquid and vapor phase FOMSs are shown by Eq. 1 and 2. As illustrated in Fig. 1, although methane has the highest vapor phase FOM, it becomes supercritical state above 190 K. Additionally, ethane and ethylene both have high liquid phase FOM, but the vapor FOM of ethylene is much higher. This shows that within the targeted temperature range, ethylene provides the highest heat transfer capability and efficiency. For this study, a total of 16.76 g of ethylene was charged. This amount ensures that the liquid grooves are filled with liquid and the vapor core is filled with saturated vapor at 170 K, which is the ideal condition for operation. Ml =

σρl hfg μl

(1)

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Mv =

σρv h1.75 fg

(2)

μ0.25 v

where σ is the surface tension, N/m; ρl and ρv are the density of the liquid and vapor respectively, kg/m3 ; hfg is the latent heat, J/kg; and μl and μv are the liquid and vapor viscosity, respectively, Pas.

Table 1. Dimensional Parameters of the heat pipe Parameter

Value

Length

914 mm

Liquid groove diameter

9.5 mm

Slot

0.6(L)*0.3(W)mm

Condenser length

120*40 mm

Liquid groove diameter

1.3 mm

Evaporator Length

450*40 mm

Fig. 1. FOM of different working fluids: Liquid Phase FOM(left) Vapor Phase FOM (right)

Experimental Facility. The experiments were carried out in a vacuum environment. Figure 2 is a schematic of the experimental setup. The experimental facility was comprised of a heating unit, a cryocooler, the tested heat pipe, a data acquisition system, and a vacuum chamber. The heat pipe was wrapped with muti-layer insulation (MLI) material to minimize parasitic heat from the environment. Several Pt1000 thermal resisters were attached on the heat pipe to monitor the temperature of each section. Thin film heaters are used to simulate the heat source. The data acquisition system collected the temperature data every 20 s. The heat pipe was tested both vertically and horizontally under a constant heat load of 20 W at different temperature levels. It should be noted that in both tests the condenser was placed above the evaporator to ensure successful operation.

2.2 Experimental Procedure In each test, the cryocooler was first turned on. When the temperature reached around 200 K, 20 W of heat load was applied to the evaporator. Also, the auxiliary heater wat the cold head of the cryocooler was adjusted until the heat pipe could reach steady state.

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Fig. 2. Schematic of the experimental facility

After the heat pipe was kept at steady state for at least an hour, the power of the auxiliary heater was decreased so that the system temperature could be lowered for the next test case.

3 Results and Discussions 3.1 Experimental Results Horizontal Case. Figure 3 shows the temperature curves of the heat pipe. The heat pipe showed excellent heat transfer performance under all three temperature levels under 20 W. The temperature difference between the condenser and evaporator were all below 1 K. However, the temperature difference increased as the temperature level decreased. Thermal resistance can also be used to evaluate the heat transfer performance. It is defined as the temperature difference divided by the heat load. The highest thermal resistance of the heat pipe was found to be 0.06 K/W at 140 K. Vertical Case. The heat pipe started-up successfully and showed excellent thermal performance from 160 K–200 K, and the temperature difference was also below 1 K. However, the temperature difference became larger below 160 K. At 140 K, the temperature difference reached over 6 K. In addition, large temperature oscillation was found below 160 K. The highest oscillation amplitude occurred in the adiabatic section, which was around 4 K, shown by the black curve in the inset plot. Furthermore, it is also observed that the evaporator and the adiabatic section had the same temperature variation trend which was opposite to that of the condenser. This indicates that when the temperature of the evaporator and the adiabatic section increase, the condenser temperature drops. Oscillation in temperature indicates that the vapor and liquid circulation in the heat pipe becomes unsteady, and the reasons will be discussed later.

3.2 Discussions Effect of Working Temperature on Temperature Difference. The thermal resistance of the heat pipe mainly comes from the conduction of the wick and the liquid film, temperature drop of the vapor, and the two-phase heat transfer of the evaporator and the condenser. Axial conduction through the wick is neglected. Also, thermal resistance of evaporation and condensation were neglected due to their high heat transfer coefficients.

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Fig. 3. Temperature curve of the heat pipe tested under horizontal position (left) and vertical position (right)

The temperature differences are calculated using Eqs. 3, 4, 5 and 6. The temperature difference due to contact resistance of the adhesive was measured using HotDisk. Figure 4 shows the calculated results. It is indicated that as temperature level decreases, the temperature difference of vapor increased drastically. This is mainly due to the decrease of vapor density. The vapor resistance is calculated as follows [12] Tvapor =

RT 2 Pv LPv

(3)

where Tvapor is the temperature difference of vapor, K; R is the molar gas constant, J/(mol·K); T is temperature, K; L is the length of the effective length of the heat pipe, m; Pv is the pressure of vapor, Pa; and Pv is the pressure loss of vapor, which is calculated as follows   ˙ l e + lc 8μv m (4) + la Pv = ρπ rv4 2 where μv is the viscosity of vapor; m ˙ is the mass flow rate, kg/s; ρ is the density of vapor, rv is the radius of vapor core, m, le , l c and l a are the length of the evaporator, the condenser, and the adiabatic section, respectively. The temperature difference across the wick can be calculated as r2 . q Twick = rln( ) k r1

(5)

where q is the heat flux, W/m2 ; r 1 and r 2 are the of the outer and inner radii of the heat pipe, m. k is the thermal conductivity of the wick, W/mK, and is calculated as kwick = (1 − ε)ks + εkl

(6)

where ε is the porosity, and k s and k l are the thermal conductivity of the solid and liquid, respectively, W/mK. Effect of Gravity on the Thermal Performance of the Heat Pipe. Figure 5 illustrates the temperature of the heat pipe in both horizontal and vertical directions. When placed vertically, the temperature difference is always higher. This is due to the liquid pool

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Fig. 4. Temperature differences of each section of the heat pipe under different temperature levels

forming in the evaporator due to gravity. Temperature oscillation was found. This may be caused by two reasons. First, at 140 K, the charging ratio is 100%. However, due to the existence of the liquid pool, a lack of liquid is created in the liquid groove when placed vertically. Therefore, liquid flow becomes discontinuous. In addition, at 140 K, the heat of vaporization decreases, making it easier for bubbles to form. The bubbles forming in the liquid pool rushes out to the condenser due to buoyancy in a periodic manner, creating an oscillation in temperature. At higher temperature ranges, evaporation is still the dominant form of heat transfer, therefore the temperature is steadier.

Fig. 5. Comparison of temperature differences of the heat pipe under different positions

4 Conclusions 1. The temperature difference of the heat pipe increases as the temperature level decreases, and the decrease of the vapor density is the main reason. 2. When the evaporator was placed downwards, a liquid pool was formed due to gravity, causing a lack of return liquid in the liquid groove, which led to the temperature oscillation of the heat pipe. 3. The effect of gravity could be mitigated in space applications, but it should be considered carefully for ground tests and other terrestrial applications. Charging the heat pipe with more working medium can expand its working range, but it comes at a cost of higher storage pressures.

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Acknowledgement. This work was supported by the Shanghai Sailing Program (20YF1455800).

References 1. Oh, H.-U., Shin, S., Beak, C.-W.: Thermal control of spaceborne image sensor using heat pipe cooling system. Aerosp. Sci. Technol. 29, 394–402 (2013) 2. Oh, H.-U., Lee, M.-K., Shin, S., Hong, J.-S.: A numerical and experimental investigation of the thermal control performance of a spaceborne compressor assembly. Cryogenics 51, 477–484 (2011) 3. Misselhorn, J.E., Brennan, P.J., Fleischanman, F.L.: Application of Heat Pipesto a Cryogenic Focal Plane Space Experiment. Presented at the AIM 22nd Thermophysics Conference, Honolulu, HI, U.S.A. (1987) 4. Sherman, A.: Cryogenic and low-temperature heat pipe/cooler studies for spacecraft application 6 (1976) 5. Yau, Y.H., Ahmadzadehtalatapeh, M.: A review on the application of horizontal heat pipe heat exchangers in air conditioning systems in the tropics. Appl. Therm. Eng. 30(2–3), 77–84 (2010). https://doi.org/10.1016/j.applthermaleng.2009.07.011 6. Enke, C.: Transient response of an axially-grooved aluminum-ammonia heat pipe with the presence of non-condensable gas. Appl. Thermal Eng. 10 (2021) 7. Zhang, X., Jiang, D., Wang, H., Liu, X.: Experimental analysis on the evaporator startup behaviors in a trapezoidally grooved heat pipe. Appl. Therm. Eng. 199, 117558 (2021). https://doi.org/10.1016/j.applthermaleng.2021.117558 8. Suman, B., Hoda, N.: Effect of variations in thermophysical properties and design parameters on the performance of a V-shaped micro grooved heat pipe. Int. J. Heat Mass Transf. 48(10), 2090–2101 (2005). https://doi.org/10.1016/j.ijheatmasstransfer.2005.01.007 9. Anand, A.R.: Effect of various parameters on heat transport capability of axially grooved heat pipes. Thermal Sci. Eng. Progress 24, 100890 (2021). https://doi.org/10.1016/j.tsep. 2021.100890 10. Tang, Y., Hu, Z., Qing, J., Xie, Z., Fu, T., Chen, W.: Experimental investigation on isothermal performance of the micro-grooved heat pipe. Exp. Thermal Fluid Sci. 47, 143–149 (2013). https://doi.org/10.1016/j.expthermflusci.2013.01.009 11. Yao, F., Bian, N., Xia, Y., Chen, W., Zhang, R.: Thermal performance of an axially grooved heat pipe subjected to multiple heating sources. Microgravity Sci. Technol. 33(1), 8 (2021). https://doi.org/10.1007/s12217-020-09851-7 12. Reay, D.A., Kew, P.A., McGlen, R.J.: Heat Pipes. Theory, Design and Applications. Elsevier Ltd.

Investigation on the Thermophysical Process of the Non-vented Storage of Liquid Xenon Xiuli Wang1 , Zhuoqun Lei1 , Yonglin Ju1(B) , Yu Chen2 , and Jianglai Liu3,4,5 1 Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai 200240, China [email protected] 2 Department of Energy and Power Engineering, School of Mechanical and Automotive Engineering, Shanghai University of Engineering and Science, Shanghai 201620, China 3 INPAC, School of Physics and Astronomy, MOE Key Laboratory for Particle Physics and Cosmology, Shanghai Key Laboratory for Particle Astrophysics and Cosmology, Shanghai Jiao Tong University, Shanghai 200240, China 4 Tsung-Dao Lee Institute, Shanghai Jiao Tong University, Shanghai 200240, China 5 Shanghai Jiao Tong University Sichuan Research Institute, Chengdu 610213, China

Abstract. The self-pressurization phenomenon of non-vented cryogen storage due to heat loss was investigated in this paper. A dynamic stratification model that incorporated temperature stratification within the fluid was proposed. The heat loss of the high vacuum multi-layer insulated tank was calculated. In particular, the heat loss above the liquidvapor interface was calculated by considering the heat conduction of the neck, the heat convection of the vapor, and the radiation of the multilayer insulation within the vacuum. The temperature distribution and the pressure variation during 30 days of non-vented liquid xenon storage were predicted based on the stratification and homogeneous model.

Keywords: Xenon Storage

1

· Heat Loss · Thermal Stratification

Introduction

In recent years, dark matter searching experiments by large-scale dual-phase xenon Time Projection Chambers (TPCs) are upgrading quickly with the mass of target xenon from 6 tons to 30 tons [1]. Under this circumstance, the xenon handling system needs to be upgraded from gas [2] to a multi-phase handling strategy to achieve high efficiency. A liquid xenon filling, recovery, and storage system (FirstX) is specially designed to handle 30 tons of xenon with five high vacuum multi-layer insulation tanks. The accurate evaluation of the temperature and pressure variations inside the FirstX tank during storage will guide the operation of the system. c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 463–471, 2023. https://doi.org/10.1007/978-981-99-6128-3_59

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The self-pressurization phenomenon due to heat loss during the non-vented storage of the cryogens has been widely investigated. During the storage, the heated liquid adjacent to the wall flows through the boundary layers to the liquid-vapor interface driven by natural convection. Evaporation on the surface leads to self-pressurization, while the heat transfer between the heated liquid surface and the adjacent liquid causes the formation of temperature gradients beneath the interface. Tatom et al. [3], Panzarella and Kassemi [4] reported that the proportion of heat loss that imposed on different parts of fluid has a great influence on the pressure. Seo and Jeong [5] assumed that the heat leak into the liquid volume evenly. As a result, the pressure rise was underestimated. Daigle et al. [6], Liu and Li [7] calculated the heat loss by conducting empirical correlations within the boundary layers. However, the heat loss relied too heavily on qualitative analysis of temperature differences. Bostock [8] reported that the temperature difference between the wall and the liquid nitrogen was 0.01 to 1.0 K magnitude. Incorrect estimation of the temperature difference will result in large errors. Non-equilibrium models have been reported for simulating the storage of LNG [9,10], but the liquid stratification has not been considered. Moreover, different from traditional storage tanks, the FirstX tank has a thick and large neck, through which the heat ingress is considerable. This paper proposed a dynamic stratification model for simulating nonvented cryogen storage with four control volumes: vapor, liquid-vapor interface, stratified liquid, and subcooled liquid. The heat loss of the FirstX were calculated by considering the heat conduction of the thick neck. The temperature distribution and the pressure variation during 30 days of non-vented liquid xenon storage have been presented by using the stratification and the homogeneous model.

2

Description of Model

The schematic diagram of the stratification model is shown in Fig. 1(a). The fluid control volume is separated into four zones: vapor, liquid-vapor interface, stratified, and completely mixed (subcooled) liquid, where the liquid-vapor interface is a thin, massless saturated film corresponding to the vapor pressure. As already mentioned in the Introduction, the heat conduction of the neck should be taken into account. The schematic diagram of heat transfer above the liquidvapor interface is then shown in Fig. 1(b). The stratification model and the heat loss will be demonstrated in detail in the following sections.

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Fig. 1. (a) Schematic diagram of the FirstX tank; (b) heat transfer schematic above the liquid-vapor interface.

2.1

Stratification Model

Mass Conservation. The condensation-evaporation mass flow rate m ˙ v through the liquid-vapor interface is related to the heat exchanger rate through the same liquid-vapor interface, Q˙ li and Q˙ vi , and the latent heat Δh. ˙ ) (Q˙li + Qvi m˙ v = J˙cd = − Δh

(1)

where Q˙ li and Q˙ vi are given by Eq. (6), while the subscript i, l, v represent liquidvapor interface, liquid and vapor, respectively. The length x and the volume Vls of the stratified liquid layer at time t can be written as Eq. (2) using an analytical solution of one-dimensional heat exchanger in a semi-infinite solid. √ x(t) = k at,

J˙cd V˙ ls (t) = x (t)A + ρls

(2)

where a and k are the thermal diffusion number and constant factor, respectively. According to the Schmidt’s experiment [11], k equals 2. Therefore, the mass balance for liquid region can be determined as follows: m ˙ ls = d(ρls Vls ),

m ˙ lm = −(m˙ v + m˙ls )

(3)

while the subscript s, m represent the stratified and the mixed liquid, respectively. Energy Conservation. The heat transfer through the liquid-vapor interface can be calculated with Churchill and Chu correlation which is given as 1

Nu =

0.387Ra 6 hd 2 = (0.825 + 9 8 ) λ 16 ) 27 (1 + ( 0.492 ) Pr

(4)

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where Ra = GrΔP r,

Grls(v) =

gβ(Tls(v) − Ts )d3 v2

(5)

Hence, Q˙ v(l)i = ±0.25hv(l) (Tv(ls) − Ts )πd2

(6)

where d and Ts are the diameter of the inner wall and the temperature of the interface, respectively. In Eq. (6), + is chosen when the energy enters liquidvapor surface, and − otherwise. Then the energy balance of the vapor is given as mv (cv Tv − c∗v Tv∗ ) ˙ v − J˙cd (h∗ − u∗ ) = Q˙ v − Q˙ vs − W (7) qv v dt ˙ v = pV˙ v . Similar, where Tv∗ is the temperature at the previous time, u = cv Tv , W the temperature of the stratified and subcooled liquid can be expressed as Eq. (8) and Eq. (9), respectively. ∗ mlm (cp Tlm − c∗p Tlm ) = Q˙lm − Q˙ls − W˙lm + J˙cd (h∗lq − u∗lm ) + m˙cd (u∗ls − u∗lm ) (8) dt

mls (cp Tls − c∗p Tls∗ ) = Q˙ls − W˙ls (9) dt ˙ for liquid u ≈ h ≈ cp T . Wlm(ls) = pV˙ lm(ls) . The thermal physical properties of cp , cv , P r, λ are updated with NIST Chemistry WebBook [12] corresponding current temperature while the other relevant thermal physical properties are updated by CoolProp [13], the fluid properties database software. As for the equation of state, the vapor pressure is obtained with calculated Tv , mv , Vv using CoolProp and the temperature of the liquid-vapor interface Ts equals the saturation temperature corresponding to the pressure. 2.2

Heat Loss

The convection heat loss within the insulation space is approximately equals zero due to the high vacuum. The radiative heat loss of the MLI and the conductive heat loss of the support are calculated by Fourier’s law of heat conduction. The rate of heat loss that enters the liquid can be calculated as Te − Tl = hAl (Te − Tl ) Q˙ l = λAl δ

(10)

The rate of heat loss entering the subcooled liquid layer is ka times the total rate of heat loss entering the liquid, where ka is the ratio of the bottom head area to the total area in contact with the liquid. Qls = ka Ql = Als /Al Ql

(11)

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λ and δ are the effective mean thermal conductivity and thickness of the MLI, Al is the area of the inner wall that contacts with the liquid, Te is the environment temperature, h is the effective heat convection coefficient of MLI. λ and δ equal to 1e−4 W/(m·K) and 10 mm, respectively. With the definition of the heat exchanger coefficient between the vapor and the inner wall, the thermal balance of an inner wall element dx with the cross-section as Aw becomes dTw d  dTw  − hπd(Te − Tw ) kw A(x) = f mc ˙ p dx dx dx

(12)

where cp is the specific heat capacity of vapor, while the subscript w represents the wall. The gradient f is defined as the heat transfer factor between the ascend gas and the inner vessel that varies from 0 to 1, where 0 corresponds to no heat exchanger and 1 to perfect heat exchanger between the gas the inner vessel. One can obtain the wall temperature distribution of the inner wall by solving Eq. (12) numerically. Then the rate of heat loss that enters vapor Q˙ v and the liquid-vapor interface Q˙ can be written as Q˙v = f mc ˙ p (Tl − Te ),

3

dT  Q˙ = kw A  dx x=0

(13)

Results and Discussion

The developed model was used to simulate 30 days of non-vented liquid xenon storage in the FirstX tank. The FirstX tank is 6 m3 in volume with a cylindrical tank wall, 1.5 m in diameter and 40 mm in thickness, and a large neck, 450 mm in diameter and 16 mm in thickness. In the initial state, the tank was filled with liquid xenon with 34% in volume. The liquid and the vapor were assumed in a saturation state with the pressure equals 0.2 MPa. The heat loss calculation method was validated with Li’s [14] LN2 storage experiment. As shown in Fig. 2(a), the predicted wall temperatures agreed with the experimental data very well. Figure 2(b) displayed the temperature distribution of the liquid xenon storage in the FirstX tank inner wall with gradient f . Due to the thickness of the inner wall and the lower k  , the temperatures were similar with different f , where k  is the sensible heat vs. the latent heat. For xenon, k  equals 0.23 which is only 20% of k  for nitrogen. The total heat loss of liquid xenon storage was calculated as 47 W.

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Fig. 2. (a) Comparison of inner wall temperature of simulation against the experimental data of Li’s [14]; (b) the predicted wall temperature of the FirstX tank with LXe storage.

Stratification model was validated with Schmidt’s [15] LH2 storage experiment. As shown in Fig. 3, the predicted vapor mass and the pressure of the stratification model were much better than that in the homogeneous model. Since the heat losses of the system were constant, the pressure bias between the stratification model and the homogeneous model was mainly due to fluid stratification. In the homogeneous model, the heat transfer coefficient at the interface and the thermal conductivity of the gas and the liquid are assumed to be infinite, and the gas and liquid were in a saturated state. In fact, the temperatures of the fluid near the interface were stratified, and the gas was superheated to the bulk fluid, according to Schmidt’s [15] experiments. In this case, some of the heat used to heat the bulk liquid in the homogeneous model actually heated the gas, causing the pressure in the tank to be higher than that in the homogeneous model.

Fig. 3. Comparison of homogeneous, stratification model against experimental data of Schmidt [15]: (a) vapor mass; (b) pressure.

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After the model validation, the stratification model was used to simulate 30 days of non-vented xenon storage in the FirstX tank. Since the heat inertia of the heavy inner wall is considerable, we calculated the temperature and the pressure considering the temperature change of the inner wall. It is assumed that the rate of wall temperature change was the same as the fluid temperature change during the storage. Results were compared with the stratification and homogeneous model [4]. As shown in Fig. 4, the temperature of the stratified liquid was greater than the vapor and was smaller than the subcooled liquid temperature. With the consideration of stainless steel, the temperature gradients were much smaller.

Fig. 4. The variation of temperature distribution during 30 days of LXe storage without (a) and with (b) considering the heat capacity of the inner wall.

The vapor mass and the pressure during the storage were also investigated. As shown in the Fig. 5, when the heat capacity is not considered, the predicted pressure with the stratification model was 2.25 MPa. By taking into account of the heat capacity of the inner wall, the predicted pressure was 0.558 MPa. The heat capacity of the inner wall greatly reduced the pressure rise by about 75%.

Fig. 5. The variation of vapor mass (a) and pressures (b) during 30 days of LXe storage.

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Conclusion

This paper presented the dynamic model of cryogenic storage that incorporated temperature stratification within the fluid. The heat loss of a cryogenic storage tank above the liquid-vapor interface was calculated by considering the heat conduction of the neck. The temperature distribution and the pressure variations were predicted for 30 days of non-vented liquid xenon storage with the stratification and homogeneous models. It was shown that the stratification phenomenon in the fluid significantly increased the pressure rise. Furthermore, in the simulation scenario described in the paper, the heat capacity of the inner wall significantly reduced the pressure rise by about 75%. Acknowledgments. The authors would like to thank the supports of the PandaX-4T collaboration. This project is supported by grants from the Ministry of Science and Technology of China (No. 2016YFA0400301 and 2016YFA0400302), a Double Topclass grant from Shanghai Jiao Tong University, grants from the Office of Science and Technology, Shanghai Municipal Government (Nos. 11DZ2260700, 16DZ2260200, and 18JC1410200), and the support from the Key Laboratory for Particle Physics, Astrophysics and Cosmology, Ministry of Education.

References 1. Cheng, J., et al.: The China Jinping underground laboratory and its early science. Annu. Rev. Nucl. Part. Sci. 67(1), 231–251 (2017) 2. Zhao, L., et al.: The cryogenics and xenon handling system for the PandaX-4T experiment. J. Instrum. 16(06), T06007 (2021) 3. Tatom, J.W., et al.: Analysis of thermal stratification of liquid hydrogen in rocket propellant tanks. In: Timmerhaus, K.D. (ed.) Advances in Cryogenic Engineering, vol. 9, pp. 265–272. Springer, Boston (1964). https://doi.org/10.1007/978-1-47570525-6 31 4. Panzarella, C.H., Kassemi, M.: On the validity of purely thermodynamic descriptions of two-phase cryogenic fluid storage. J. Fluid Mech. 484, 41–68 (2003) 5. Seo, M., Jeong, S.: Analysis of self-pressurization phenomenon of cryogenic fluid storage tank with thermal diffusion model. Cryogenics 50(9), 549–555 (2010) 6. Daigle, M.J., et al.: Temperature stratification in a cryogenic fuel tank. J. Thermophys. Heat Transfer 27, 116–126 (2013) 7. Liu, Z., Li, Y.: Thermal physical performance in liquid hydrogen tank under constant wall temperature. Renew. Energy 130, 601–612 (2019) 8. Bostock, T.D., Scurlock, R.G.: Low-Loss Storage and Handling of Cryogenic Liquids. ICMS, Springer, Cham (2019). https://doi.org/10.1007/978-3-030-10641-6 9. Felipe, H., Velisa, V.: A realistic vapour phase heat transfer model for the weathering of LNG stored in large tanks. Energy 174, 280–291 (2019) 10. Wang, Z., et al.: Non-equilibrium thermodynamic model for liquefied natural gas storage tanks. Energy 190, 116412 (2020) 11. Schmidt, A.F., et al.: Pressurization and stratification of liquid hydrogen. Jo. Res. Nat. Bureau Stan. Sect. C Eng. Instrum. 65C(2), 81 (1961) 12. Linstrom, P.J., Mallard, W.G. (eds.) NIST Chemistry WebBook, NIST Standard Reference Database Number 69, National Institute of Standards and Technology (2000)

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13. Bell, I.H., et al.: Pure and pseudo-pure fluid thermophysical property evaluation and the open-source thermophysical property library CoolProp. Ind. Eng. Chem. Res. 53(6), 2498–2508 (2014) 14. Li, Y., Wang, R., Wang, C.: Study on effect of liquid level on the heat leak into vertical cryogenic vessels. Cryogenics 50(6), 367–372 (2010) 15. Schmidt, A.F., et al.: An experimental study concerning the pressurization and stratification of liquid hydrogen. In: Timmerhaus, K.D. (ed.) Advances in Cryogenic Engineering, pp. 487–497. Springer, Boston (1960). https://doi.org/10.1007/ 978-1-4757-0537-9 56

Results of a Nitrogen Pulsating Heat Pipe with Subsections in a Series Configuration Zhiyi Jiang(B)

, John Pfotenhauer , and Franklin Miller

Department of Mechanical Engineering, University of Wisconsin - Madison, Madison, USA {zjiang8,fkmiller}@wisc.edu, [email protected]

Abstract. In this paper, the results of a nitrogen-based pulsating heat pipe (PHP) at varied condenser temperatures and fill ratios are compared. The tested PHP consists of two individual subsections, which are connected in a series configuration. Each subsection has 14 stainless-steel capillary tubes which form a total of 28 capillary tubes for the PHP. The PHP was tested at three different initial fill ratios (38%, 50%, and 63%) with the condenser temperature controlled at 79.5 K and 84.5 K. The optimal heat transfer performance of 55,3000 W/m-K was achieved at the initial fill ratio of 50% with the condenser temperature controlled at 84.5 K. At the same condenser temperature, the PHP with a 63% fill ratio achieved the maximum heat transfer capacity of 13.4 W before dry out. The experimental results of the PHP in a series configuration presented in this paper will be compared with subsequent PHP experiments with identical PHP subsections but in varied configurations. Keywords: Pulsating Heat Pipes · Nitrogen · Heat Transfer Performance

1 Introduction The pulsating heat pipe (PHP) has a simple design with a very large heat transfer performance. It was invented by Akachi in the early 90s to assist the heat dissipation of electronic devices [1]. Its heat transfer performance is similar to a conventional heat pipe but without a wick structure, and with capillary sized tubing. The multiple-turn capillary tubing extends back and forth between the evaporator and condenser of the PHP core. The working fluid within the capillary tube of the PHPs is in a two-phase state. A pressure difference induced by the varying saturation conditions in the evaporator and condenser sections drives flow between the two ends. As the fluid flows into the evaporator and absorbs heat, vapor bubbles are generated resulting in a pressure increase. At the other end, vapor bubbles contract, and pressure decreases as heat is released into the condenser. The temperature drop along the capillary tubing of the PHPs is significantly smaller than what would exist in high purity metals with the same length and cross-sectional area in response to the same heat load. As a result, the thermal resistance through the PHP is significantly smaller between the cold end and the heat source [2]. This outstanding heat transfer characteristic enables promising cryogenics applications such as removing heat from superconducting magnets. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 472–478, 2023. https://doi.org/10.1007/978-981-99-6128-3_60

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There has been extensive research on the room temperature pulsating heat pipes since the pulsating heat pipes were invented. The typical working fluids of the room temperature pulsating heat pipes are water, ethanol, acetone, and some conventional refrigerants [3]. Some mixtures of different working fluids have also been found to have excellent heat transfer performance [4, 5]. The hydrodynamic behavior of the vapor bubbles and the liquid slugs can be easily observed through the transparent glass capillary tubing [6, 7]. The effects of the orientation, internal diameter, number of turns, and the length of the adiabatic section on the flow pattern and the performance of the PHPs are thoroughly investigated. Some room temperature PHPs can achieve a heat transfer capacity as high as 2 kW nowadays [8]. There has been an increasing amount of research on cryogenic pulsating heat pipes in the last decade. The main working fluids tested are helium, hydrogen, neon, and nitrogen. Since the driving force of the PHPs heavily relies on the volumetric expansion and contraction of the two-phase fluid and the critical point of the cryogenic working fluids is significantly low, the challenges associated with the visualization of the cryogenic pulsating heat pipes are prohibitive. The flow pattern and motion of cryogenic working fluids inside PHPs are mostly investigated by numerical modeling [9–11]. Additionally, a large number of experimental results are needed in order to better understand the behavior of the cryogenic pulsating heat pipes. The experimental work presented in the following sections on nitrogen-based pulsating heat pipes in a series configuration will form a reference for future work studying the effects of the number of turns and the configurations on the performance of nitrogen pulsating heat pipes.

2 Experimental Setup The schematic of the experimental test rig is presented in Fig. 1. A Sumitomo CH-110 cryocooler is mounted on the top plate of the vacuum dewar. The measured cooling capacity of this cryocooler is around 200 W at 75 K when paired with the Sumitomo F-50 compressor. A copper adapter plate is mounted on the cold finger of the cryocooler. Four 50 W heaters (model KAL-50 with a resistance of 50 Ohms and tolerance of 1%) are mounted on the copper adapter plate. The top of the radiation shield is thermally anchored on the copper adapter plate and the condenser section of the PHP core is thermally anchored on the bottom of the top plate of the radiation shield. The condenser temperature of the PHP thus is controlled by the KAL-50 heaters on the copper adapter plate. The PHP core includes three major parts which are shown in Fig. 1: the condensers, the adiabatic sections, and the evaporators. The condenser and evaporator are rectangular shape copper plates. The distance between the condenser and evaporator plates, the so-called adiabatic section is 1 m long. The capillary tubing meandering between the components of the PHP core is made from 304 stainless steel with an inner diameter of 0.5 mm and an outer diameter of 0.8 mm. The inner diameter of the capillary tubing is smaller than the critical diameter that corresponds to the critical temperature of the nitrogen to guarantee distinct vapor plugs and liquid slugs will be maintained within the

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vapor dome. The critical diameter of the working fluid is expressed as [12]:  σ Dcrital = 2 g(ρliq − ρvap )

(1)

where σ is the surface tension of the fluid. ρ liq and ρ vap are the density of the saturated liquid and vapor, and g is the gravitational acceleration.

Fig. 1. The schematic of the experimental setup.

Lakeshore PT-102 platinum sensors are thermally anchored on the center of each evaporator and condenser plate with Lakeshore VGE-7031 varnish. On the backside of each evaporator plate, KAL-50 resistive heaters are installed to provide the heat load. Each PHP section has a total of 14 capillary tubes, and the two PHP sections are connected in the series configuration as shown in Fig. 2. The gas line is purged with nitrogen gas at least three times in order to remove contamination from the system. Then the buffer tank is filled with nitrogen gas from the gas cylinder. Once the condenser section reaches the targeted temperature, the PHP core is filled with nitrogen gas from the buffer tank to achieve the specified initial fill ratio which is defined as: FR =

Volliq Voltotal

(2)

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where Volliq is the volume of the liquid within the PHP and Voltotal is the total volume of the PHP.

Fig. 2. Condenser plates are connected in the series configuration by the copper capillary tube.

Heat is applied to the evaporator sections once the PHP reaches a steady state at the specified condenser temperature after the filling process. The heat load is stepwise increased from zero to the dry-out condition. The dry-out condition is defined when the temperature of the working fluid in the evaporator transitions out of the vapor dome and the pressure measurement of the fluid stops oscillating. The heat transfer performance of the PHPs degrades significantly when reaching the dry-out condition.

3 Results and Discussion The experiments were conducted at two different condenser temperatures 79.5 K and 84.5 K. The PHP is tested at the initial fill ratios of 38%, 50%, and 63% at each condenser temperature. One of our long-term goals is to characterize the behavior of the PHP from the perspectives of the condenser temperature, the initial fill ratios, and the number of tubes. A varied-configuration PHP is built recently with the same length of the condenser, adiabatic, and evaporator sections. The results presented below and the data collected from the newly built will help describe the characteristics of the thermal performance of the nitrogen-based pulsating heat pipes. At each condenser temperature and fill ratio, the heat load is applied successively until the dry out is reached. The temperature and pressure data were collected for a total of 12 h with a sampling rate of 1000 Hz for each heat load. The time-averaged temperature measurements at the condenser and evaporator sections are used for calculating the effective thermal conductivity: keff =

QL NAc (T e − T c )

(3)

where L is the distance between the center of the evaporator and the center of the condenser, Q is the total heat load on the evaporator plates, N is the total number of

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capillary tubes (28 for the PHP1 and PHP2 combination) and AC is the cross-sectional area of the capillary tubing.

Fig. 3. The effective thermal conductivities versus the heat load on the evaporators at three fill ratios at condenser temperature 79.5 K (top) and 84.5 K (bottom).

The effective thermal conductivities of the PHP at 79.5 K and 84.5 K are shown in Fig. 3. All six effective thermal conductivity plots above present a dome shape behavior. From zero to around 6 W, the effective thermal conductivities grow with the increasing heat load on the evaporators. Recent CFD results suggest that the increased effective conductivity is due to increased velocities of the plug-slug flow, resulting from the increased heat load [13]. From 6 W to dry out, the effective thermal conductivities plateau. It is possible the working fluid is in the annular flow regime. The significant variation among different fill ratios was observed in Fig. 3 between 2 to 8 W and 4 to 8 W for the condenser temperatures of 79.5 K and 84.5 K, presenting different heat transfer characteristics at varied fill ratios. Among the three initial fill ratios, the PHP with a 63% fill ratio shows the largest capacity for transferring heat (10.0 W at 79.5 K and 13.4 W at 84.5 K). PHPs with high fill

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ratios have more liquid within the system that can carry the sensible heat. Additionally, the higher percentage of the fluid within the system will benefit from a greater buoyancy force of the liquid at the high heat load that drives the fluid flows [14]. The optimal effective thermal conductivities were observed at a 50% fill ratio (483000 W/m-K at 79.5 K and 553000 W/m-K at 84.5 K) and at both condenser temperature cases. A 50% fill ratio corresponds to a specific volume at the center of the vapor dome which means there are sufficient bubbles to cause the pressure perturbation and enough liquid to absorb the heat. Overall, PHP tests at the condenser temperature of 84.5 K have a larger capacity for transferring heat before dry-out and a higher mean effective thermal conductivity than at the condenser temperature of 79.5 K. The slope of the saturation line, (dP/dT)sat, appears to play an important role in the performance of pulsating heat pipes. The PHP with a fill ratio of 38% and at the condenser temperature of 79.5 K reached an optimal effective thermal conductivity of 523000 W/m-K. Sagar reported an effective thermal conductivity of around 30000 W/m-K for his single-turn nitrogen-based PHP at a 38% fill ratio with the condenser submerged in boiling LN2 [15]. Although the effective thermal conductivity of Sagar’s single-turn PHP is much smaller than the 28-tube seriesconfiguration PHP, the thermal conductivity per tube is in the same order of magnitude. The single-turn (2 capillary tubes) PHP has a unit effective thermal conductivity of 14955 W/m-K-tube while our 28-tube PHP has a unit effective thermal conductivity of 18665 W/m-K-tube. It suggests that the heat transfer performance of our PHP at a 38% fill ratio is only about 25% higher than Sagar’s single-turn PHP on a per tube basis.

4 Conclusion The heat transfer performance of the nitrogen-based pulsating heat pipes with two PHP subsections connected in series (a total of 28 capillary tubes) is presented. The optimal heat transfer performance of 553,000 W/m-K was achieved at the initial fill ratio of 50% with the condenser temperature controlled at 84.5 K. The maximum capacity of the PHP for transferring heat, 13.4 W, is reached at the initial fill ratio of 63% with the condenser temperature controlled at 84.5 K. The PHP with a higher fill ratio will have sufficient liquid for transferring the sensible heat. The comparison between Sagar’s single-turn PHP and the 28-tube PHP suggests that although the heat transfer performance per turn of the PHPs is similar under similar conditions, the overall performance might vary significantly. The study of varied configuration helium PHPs by Li [16] also suggests that there may be an optimal number of turns that would maximize the heat transfer performance per turn of the PHPs. The goal of future investigations is to explore the optimal configuration of the pulsating heat pipes. The data presented above for the two PHP subsections connected in series will be used as the reference for comparison. Acknowledgments. The project is supported by Sumitomo (SHI) Cryogenics of America, Inc.

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References 1. Akachi, H.: Structure of a heat pipe. U.S. Patent No. 4 921 041, 1 May 1990 2. Natsume, K., et al.: Heat transfer performance of cryogenic oscillating heat pipes for effective cooling of superconducting magnets. Cryogenics 51(6), 309–314 (2011) 3. Wang, X., Jia, L.: Experimental study on heat transfer performance of pulsating heat pipe with refrigerants. J. Therm. Sci. 25(5), 449–453 (2016) 4. Shi, S., Cui, X., Han, H., Weng, J., Li, Z.: A study of the heat transfer performance of a pulsating heat pipe with ethanol-based mixtures. Appl. Therm. Eng. 102, 1219–1227 (2016) 5. Wang, D., Cui, X.: Experiment research on pulsating heat pipe with different mixtures working fluids. In: The 21st International Symposium on Transport Phenomena, Kaohsiung City, Taiwan (2010) 6. Xu, J.L., Li, Y.X., Wong, T.N.: High speed flow visualization of a closed loop pulsating heat pipe. Int. J. Heat Mass Transf. 48(16), 3338–3351 (2005) 7. Khandekar, S., Charoensawan, P., Groll, M., Terdtoon, P.: Closed loop pulsating heat pipes Part B: visualization and semi-empirical modeling. Appl. Therm. Eng. 23(16), 2021–2033 (2003) 8. Czajkowski, C., Nowak, A.I., Błasiak, P., Ochman, A., Pietrowicz, S.: Experimental study on a large scale pulsating heat pipe operating at high heat loads, different adiabatic lengths and various filling ratios of acetone, ethanol, and water. Appl. Therm. Eng. 165, 114534 (2020) 9. Han, DY., Sun, X., Gan, Z.H., Luo, R.Y., Pfotenhauer, J.M., Jiao, B.: Numerical investigation on pulsating heat pipes with nitrogen or hydrogen. In: IOP Conference Series: Materials Science and Engineering, vol. 278. IOP Publishing, Madison (2017) 10. Sagar, K.R., Naik, H.B., Mehta, H.B.: Numerical study of liquid nitrogen based pulsating heat pipe for cooling superconductors. Int. J. Refrig 122, 33–46 (2021) 11. Li, M.N., Wu, Z.X., Xu, D., Li, L.F.: Numerical simulation of 4-turns nitrogen pulsating heat pipe. In: IOP Conference Series: Materials Science and Engineering, vol. 502. IOP Publishing, Oxford (2019) 12. Khandekar, S., Dollinger, N., Groll, M.: Understanding operational regimes of closed loop pulsating heat pipes: an experimental study. Appl. Therm. Eng. 23(6), 707–719 (2003) 13. Xu, C.: Performance characterization of a helium PHP via Fluent. Ph.D. dissertation, Dept. of Mech. Eng. University of Wisconsin-Madison (2022) 14. Li, M., Li, L., Xu, D.: Effect of filling ratio and orientation on the performance of a multiple turns helium pulsating heat pipe. Cryogenics 100, 62–68 (2019) 15. Sagar, K.R.: Experimental investigations on novel condenser based cryogenic pulsating heat pipes. Ph.D. dissertation, Dept. of Mech. Eng. SVNIT, Surat, India (2022) 16. Li, M., Li, L., Xu, D.: Effect of number of turns and configurations on the heat transfer performance of helium cryogenic pulsating heat pipe. Cryogenics 96, 159–165 (2018)

Numerical Investigation on the Condensation of Various Refrigerants Outside Horizontal Plain and Low Finned Tubes at Low Temperatures Shu Li and Yonglin Ju(B) Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai, China {shu1101,yju}@sjtu.edu.cn

Abstract. In the liquefied natural gas (LNG) intermediate fluid vaporizer (IFV), the condensation of refrigerants outside the LNG tubes is the main factor affecting the regasification efficiency. In this paper, the condensation heat transfer characteristics of refrigerants outside horizontal plain and low finned LNG tubes are investigated at low temperatures through numerical simulation. Based on the VOF model and Lee phase change model, the CFD model is validated with the experimental data for the plain and low-finned tubes. For the plain tube, dimethylether (DME) shows the best heat transfer performances in the seven candidate refrigerants. For the low-finned tubes at high heat flux, the retention angle decreases as heat flux increases. The analytical model overpredicts the retention angle at high heat flux and large fin density. The difference in heat flux between flooded and unflooded regions decreases as wall subcooling increases. The tubes with a fin height of 0.3 mm and a fin density between 25–34 fins per inch show high heat transfer performance for the IFV-condenser. In the four heat transfer correlations, the Honda correlation has the highest prediction accuracy under the IFV condition, with a deviation between −10% and 57%. Keywords: Condensation outside Tubes · Low Temperatures · Fin Structures · Heat Transfer Correlations

1 Introduction In the LNG trade, the LNG in the receiving terminals is regasified before the final consumption. Compared with other seawater vaporizers, the intermediate fluid vaporizer (IFV) has the advantages of less icing, compact volume and low seawater requirement, etc. In the heat transfer units of the IFV, the heat transfer coefficient (h) of the condensation outside the LNG tubes in the Condenser is the lowest [1]. Therefore, it is necessary to study and improve the condensation performance of the intermediate fluid. The widely used intermediate fluid in the IFV is propane and the LNG tubes are plain. For the miniaturization of the IFV, the performances of other refrigerants need to be evaluated. High-efficiency finned LNG tubes should also be developed. The low-finned tubes are mainly applied in the condensers in the refrigeration and air-conditioning industry, with the fin height below 1.5 mm. Under the surface tension © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 479–486, 2023. https://doi.org/10.1007/978-981-99-6128-3_61

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effect, the condensate is retained at the tube bottom, known as condensate retention, as shown in Fig. 1. This region is called the flooded region. The heat flux (q) in the flooded region is small since the fin surface is completely flooded by the condensate. The portion of the flooded region is expressed by the condensate retention angle: θ f = θ f1 + θ f2 . The analytical predictions for the retention angle are proposed by Rudy et al. [2] and Honda et al. [3], of which the results are similar. There are also different models for predicting the heat transfer coefficients on the integral finned tubes, in which the heat transfer coefficients in the flooded and unflooded regions are calculated separately [4– 7]. The current researches on condensation outside tubes are mainly concentrated in the refrigeration and air-conditioning field. However, the IFV operating condition is quite different. The outer wall temperature of the LNG tube is low and varies in a wide range. Accordingly, the wall subcooling (T ) is large and the heat flux is high. The saturation temperature (T s ) of the refrigerant is also low, usually below 0 °C. However, there is rare research on refrigerants condensation outside tubes at low saturation temperatures and large heat flux, for either plain or low-finned tubes. In this paper, we studied the performances of different refrigerants in the IFV. The special flow and film distribution characteristics outside the low-finned tubes at large heat flux are investigated. To obtain high-performance low-finned tubes, we also investigated the effects of fin parameters on the flow and condensation heat transfer characteristics under the IFV condition. The accuracy of the heat transfer correlations for the low-finned tube was further evaluated for the future design of the IFV.

Fig. 1. Condensate retention and retention angle

2 Numerical Model 2.1 Physical and Mathematical Model The 3D geometry model and parameters of the low-finned tube are shown in Fig. 2(a). A 2D model is used for the plain tube to simplify the calculation. For the calculation of the multiphase flow, the VOF model is adopted. The Lee model is selected for the ˙ e ) can phase change simulation. The interface liquid-vapor mass transfer rate (m ˙ c and m be obtained by Eq. (1). The mass transfer coefficient (r) in Eq. (1) is set to be 5 e6 s−1 according to the sensitivity verification in Ref. [8]. The interface surface tension (F σ ) is obtained by the continuum surface force (CSF) model and the wall adhesion model, as

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depicted in Eq. (2). The contact angle in the CSF model is set to be 0°.  T −T m ˙ c = rαg ρg sTs g ,Tg < Ts , condensation s ,Tl > Ts , evaporation m ˙ e = rαl ρl TlT−T s Fσ = σ

ρκ∇αl 1 2 (ρl

+ ρg )

=∇·

∇αl σρ∇αl |∇αl | 21 (ρl + ρg )

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(1) (2)

where α is the phase volume fraction, ρ is the density, T is the temperature, σ is the surface tension coefficient, κ is the surface curvature. The subscripts g, l and s represent the gas phase, the liquid phase and the saturated state, respectively.

Fig. 2. (a) 3D geometric model and geometric parameters (b) grid model and mesh refinement

2.2 Numerical Methods The structured hexahedral mesh is generated to discretize the computational domain. In the condensation simulation, the grid size at the mass transfer interface has a great influence on the results. Therefore, the mesh near the tube surface and between the fins is refined. Four mesh schemes are adopted and the condensation heat transfer coefficients for propane at T s = 35 °C are shown in Table 1. Compared with the mesh scheme with a maximum grid size of 9.7 μm, the deviation of that with the size of 13.8 μm is less than 1%. Considering the consumption of computing resources, the scheme with the maximum grid size of 13.8 μm between the fins is selected for the calculation. The grid model and the mesh refinement of the low-finned tube are demonstrated in Fig. 2(b). For the plain tube, the grid height at the mass transfer interface is below 4 μm. The time step in the transient model is set to be 1e−5 s.

3 Results and Discussion 3.1 Model Validation Three cases in the experiment of Ref. [9] are selected for model validation. As shown in Fig. 3, 96.3% of the heat transfer coefficients predicted by the numerical model are within a 10% deviation from the experimental data.

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Grid numbers

52950

99704

118605

216395

Maximum grid size between the fins/μm

46.8

28.6

13.8

9.7

h/kW·m−2 ·K−1

0.555

0.548

0.501

0.506

Deviation

9.68%

8.30%

−0.99%

/

Fig. 3. Comparison between the numerical model and the experimental data

3.2 Evaluation of Different Refrigerants for Plain Tube To improve the performance of the IFV, it is necessary to evaluate the heat transfer efficiency of other refrigerants on the plain tube, as shown in Table 2. First of all, the pressure of the refrigerants in the IFV should be considered. Because of the large size of the IFV, the upper-pressure limit in the IFV is much lower than the condensers in the refrigeration and air-conditioning industry. The saturation pressures (ps ) of R32 and R410A are rather high. The pressure of butane is lower than the ambient pressure. Therefore, R32, R410A and butane are not suitable for the IFV. DME performs the best among the remaining refrigerants. For the IFV system with a vaporization load of 175 t/h, when using DME as the intermediate fluid, the heat transfer area of the IFV-condenser can be reduced by 36.7%, compared with that using propane. Table 2. The heat transfer coefficients for different refrigerants at T s = −7.7 °C, T = 10 K Refrigerant

R32

R410A DME

C3 H6

C3 H8

R134a

I-C4 H10 N-C4 H10

ps /MPa

0.6306 0.6216

0.2018 0.4609 0.3723 0.2195 0.1184

0.0764

h/kW·m−2 ·K−1 3.2985 2.4455

2.3709 2.3214 1.9713 1.8193 1.4922

1.6197

3.3 Condensate Retention Angle In the analytical predictions [2, 3], the retention angle is a function of fluid properties and fin geometry, while is independent of heat flux. The variation of retention angle

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with heat flux outside the low-finned tubes is shown in Fig. 4. At low heat flux, as in the refrigeration and air-conditioning conditions, the retention angle is independent of heat flux, as depicted in Fig. 4(a). However, for the IFV condition in Fig. 4(b), the retention angle decreases as heat flux increases. The condensate film thickness in the unflooded region is ignored in the analytical predictions. At low heat flux, the condensate film in the unflooded region remains thin as heat flux increases. However, at high heat flux, the film is thick and the film thickness increases as heat flux increases. Therefore, the retention angle decreases as heat flux increases. The analytical prediction overpredicts the angle at high heat flux and large fin density.

Fig. 4. Variation of retention angle with heat flux (a) low heat flux (b) high heat flux

3.4 Circumferential Distribution of Heat Flux for the Low-finned Tubes The percentage of the circumferential local heat flux in the total heat flux (PQ ) outside the low-finned tubes is depicted in Fig. 5. In Fig. 5 (a), the heat flux decreases in the circumferential direction at the fin height of 0.3 mm and small wall subcooling. Especially, under the condensate retention effect, the heat flux decreases significantly around the retention angle. However, as the wall subcooling increases, the difference between the flooded and unflooded regions and the retention effect decrease. The curve is quite flat at the subcooling of 80 K. In the unflooded region, the condensate film thickness increases as subcooling increases, while the effective heat transfer area and heat flux decrease. Besides, the condensate film in the flooded region is thinner for the lower fins. Therefore, the heat flux difference between the two regions reduces. When increasing the fin height to 0.6 mm, the heat transfer area in the unflooded region increases and the above difference becomes significant. For higher fin density, it also shows a similar trend, as depicted in Fig. 5(b). In addition, the curves at the fin height of 0.6 mm and 0.9 mm are similar, due to the lower fin efficiency at higher fins. 3.5 Heat Transfer Coefficient for the Low Finned Tubes Effects of Fin Parameters. The outside condensation heat transfer coefficients of the low-finned tubes with different fin parameters are calculated under the IFV condition, as demonstrated in Fig. 6. As fin density increases, the heat transfer coefficient increases as the fin surface area increases. However, the retention angle and the effective heat transfer area decrease as fin density increases. Therefore, there is a peak on the curve,

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corresponding to the optimal fin density. As discussed before, the circumferential distributions of heat flux and retention effect for the fin height of 0.6 mm and 0.9 mm are similar. Therefore, the optimal fin densities for the two heights are both 28fpi. For the fin height of 0.3 mm, the retention effect on heat flux is small. Therefore, as fin density increases, the heat transfer coefficient decreases slower and the optimal fin density increases to 40fpi. By increasing the fin height, the heat transfer coefficient increases with the increase of fin surface area. However, the curves of 0.6 mm and 0.9 mm are close because the fin efficiency of 0.9 mm is low. There is no need to increase the fin height to 0.9 mm. Therefore, the tubes with fin density in the range of 25–34fpi and fin height of 0.6 mm can be considered in the design of the IFV.

Fig. 5. The circumferential distributions of the percentage of the heat flux (a) 23fpi (b) 28fpi

Fig. 6. Effects of fin parameters on the condensation heat transfer for low-finned tubes

Evaluation of Heat Transfer Correlations. Four heat transfer correlations for the condensation outside the integral-finned tubes are evaluated under the IFV conditions, as shown in Fig. 7. The Beatty & Katz model [4] ignores the effect of surface tension and overpredicts the heat transfer. The Webb model [5] used the analytical prediction for retention angle, which overpredicts the angle at large fin density. Therefore, the Webb model overpredicts the heat transfer and the deviation increases as fin density increases. Especially, the deviation is large at the fin height of 0.3 mm since the thick film and low heat transfer in the unflooded region. The Briggs & Rose model [6] used three experimentally fitted constants to simplify the calculation. However, the working fluid and operating conditions of these experiments were quite different from the IFV conditions. Therefore, the error varies in a wide range from −11% to 154% and the model is not suitable for the calculation of the IFV. Under the IFV conditions, the Honda model [7] shows the highest prediction accuracy, with errors ranging from −10% to

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57%. The larger deviation for the fin height of 0.3 mm is also due to the low heat transfer in the unflooded region. Therefore, the Honda model can be considered in the thermal model of the condensation outside low-finned tubes in the IFV.

Fig. 7. Evaluation of heat transfer correlations for finned tubes under the IFV condition

4 Conclusion The main conclusions can be summarized as follows: 1) For the plain tube, DME shows the best heat transfer performance in the seven candidate refrigerants at T s = −7.7 °C. The heat transfer area of the IFV-condenser can be reduced by 36.86% when applying the DME, compared with that using propane. 2) For the low-finned tubes, the retention angle decreases as heat flux increases. The analytical prediction overpredicts the angle at high heat flux and large fin density. 3) The retention effect on heat flux decreases with the increase of subcooling at the lower fin height of 0.3 mm. The circumferential heat flux distributions at the fin height of 0.6 mm and 0.9 mm are similar due to the lower fin efficiency for higher fin. 4) The tubes with a fin height of 0.6 mm and fin density between 25–34 fpi show high heat transfer performance under the IFV condition. The Honda heat transfer correlation has the highest prediction accuracy of the four correlations.

References 1. Pu, L., Qu, Z.G., Bai, Y.H., Qi, D., Song, K., Yi, P.: Thermal performance analysis of intermediate fluid vaporizer for liquefied natural gas. Appl. Therm. Eng. 65(1–2), 564–574 (2014) 2. Rudy, T.M., Webb, R.L.: An analytical model to predict condensate retention on horizontal integral-fin tubes. J. Heat Transfer 107(2), 361–368 (1985) 3. Honda, H., Nozu, S., Mitsumori, K.: Augmentation of condensation on horizontal finned tubes by attaching a porous drainage plate. In: ASME-JSME Thermal Engineering Joint Conference, vol. 3, pp. 289–295. American Society of Mechanical Engineers, New York (1983) 4. Beatty, K.O., Katz, D.L.: Condensation of vapors on outside of finned tubes. Chem. Eng. Prog. 44(1), 55–70 (1948) 5. Webb, R.L., Rudy, T.M., Kedzierski, M.A.: Prediction of the condensation coefficient on horizontal integral-fin tubes. J. Heat Transfer 107(2), 369–376 (1985)

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6. Briggs, A., Rose, J.W.: An evaluation of models for condensation heat transfer on low-finned tubes. J. Enhanced Heat Transfer 6(1), 51–60 (1999) 7. Honda, H., Nozu, S.: A prediction method for heat transfer during film condensation on horizontal low integral-fin tubes. J. Heat Transfer 109(1), 218–225 (1987) 8. Li, S., Ju, Y.L.: Numerical study on the condensation characteristics of various refrigerants outside a horizontal plain tube at low temperatures. Int. J. Therm. Sci. 176, 107508 (2022) 9. Ji, W.T., Chong, G.H., Zhao, C.Y., Zhang, H., Tao, W.Q.: Condensation heat transfer of R134a, R1234ze(E) and R290 on horizontal plain and enhanced titanium tubes. Int. J. Refrig 93, 259–268 (2018)

Experimental and Theoretical Study of Variable Density Multilayer Insulation (VDMLI) at Different Cold Boundary Temperatures Tiantian Xiao1,2 , Huiming Liu1(B) , Yuchen Zhao1,2 , Laifeng Li1 , and Yuan Zhou1 1 State Key Laboratory of Technologies in Space Cryogenic Propellants,

Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Zhongguancun East Road, Beijing 100190, China [email protected] 2 School of Future Technology, University of Chinese Academy of Sciences, Jingjia Road, Beijing 100049, China

Abstract. Cryogenic propellant will play an important role in the future aerospace field because of its superior properties. The zero boil off (ZBO) storage of cryogenic propellant is the key of its long-term on orbit storage. Variable density multilayer insulation material (VDMLI) is one of the effective passive thermal protection technologies for cryogenic propellant storage tanks. The theoretical and experimental study on the thermal performance of VDMLI can provide reliable guarantee for its practical engineering application. A concentric cylindrical calorimeter is designed and built to measure the thermal performance of the VDMLI at low temperature. The effects of cold boundary temperature on the performance of the VDMLI have been analyzed. According to the layer by layer model, the theoretical model of the VDMLI is established to predict the thermal performance of the VDMLI, and it has been verified by experiments. Keywords: VDMLI · cryogenic propellant tank · simulation · variable working conditions

1 Introduction With the continuous development of aerospace technology, cryogenic propellant will play an important role in the future space field and there are many technologies for cryogenic propellant storage [1]. Heat leakage from the external environment will cause the vaporization of cryogenic propellant. Therefore, it is necessary to study the thermal management technology of cryogenic propellant. Multilayer insulation (MLI) or variable density multilayer insulation (VDMLI) is the most effective means of reducing heat load and has been widely used in cryogenic systems [2]. The performance of MLI is affected by many factors, such as properties of multilayer materials, cold and warm boundary temperatures, number of layers, layer density and so on [2–4]. Therefore, the performance of multilayer materials in different conditions must be measured before use. At present, there are two kinds of test platforms © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 487–493, 2023. https://doi.org/10.1007/978-981-99-6128-3_62

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for multilayer insulation materials commonly used at home and abroad. One is the lowtemperature liquid evaporation test platform [5, 6], and the other is the calorimeter test platform with cryocooler as the cold source [7, 8]. NASA first proposed the concept of variable density multilayer insulation material [9]. The principle of VDMLI is to arrange the spacer in a more reasonable place so as to improve the insulation performance of multilayer materials. In order to evaluate the performance of VDMLI from both theoretical and experimental aspects, this paper uses the traditional layer by layer model to predict the performance of VDMLI, and builds a concentric cylindrical calorimeter to test the VDMLI, so as to obtain the heat leakage through the blanket and the temperature distribution inside the materials.

2 Experimental Apparatus In this article, a concentric cylindrical calorimeter is designed and built to measure the thermal performance of the VDMLI. Figure 1 (A) shows the schematic of the calorimeter.

Fig. 1. (A) Schematic of the calorimeter: 1 – 2nd stage cold head, 2 – cold cylinder support rod, 3 – cold cylinder support plate, 4 – cold cylinder, 5 – VDMLI, 6 – upper radiation shield, 7 – lower radiation shield, 8 – vacuum chamber, 9 – lower radiation shield support rod. (B) Schematic of the calorimeter rod calibration: 10 – 2nd stage cold head, 11 – upper radiation shield, 12 – calorimeter rod, 13 – heating resistance, 14 – radiation shield, 15 – thermometer.

The main components and their dimensions and functions of the experimental device are as follows: A two-stage G-M cryocooler is used to maintain the cold boundary temperature. The cold cylinder is made of 204 mm outside diameter, 500 mm long and 2 mm thick aluminium pipe and there is a support plate in the center. The secondary

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cold head of the cryocooler is connected with the support plate through the cold cylinder support rod, in other words, a calibration rod. This rod is 15 mm diameter and 244 mm long. Two copper radiation shields are installed on both ends of the cold cylinder, which are used to eliminate end effects to the measurement. The sample to be tested is wrapped around the cold cylinder. The temperature of the cold cylinder is controlled by the temperature controller. The digital multimeter is used to read the value of each thermometer. A turbomolecular pump is used to create and maintain the vacuum inside the cryostat. The data read by the electronic instrument is recorded and stored by the LabVIEW software on the computer. The reliability of the device is verified according to the heat flux of the VDMLI at 77 K. The heat flux tested by the VDMLI material manufacturer using the lowtemperature liquid evaporation test platform is 2.289 W/m2 , the test data in this article is 1.95 W/m2 , and the error is 14.8%, which is within the scope of engineering application.

3 Test Results and Discussion 3.1 Calorimeter Rod Calibration The measuring principle of the experimental device is to calibrate the inner cylinder support rod and obtain the relationship between the temperature difference on the rod and the heat transfer, Q. Q = ϕ · T

(1)

where Q is the heat transfer from environment, T is the temperature difference on the rod, ϕ is the calibration parameter. Figure 1 (B) shows the schematic of the calorimeter rod calibration. The calorimeter rod is connected with the 2nd stage cold head of the cryocooler, where the temperature is controlled by the temperature controller. The heating resistance is pasted at the bottom of the calorimeter rod, which is powered by DC power supply. At a certain boundary temperature, we can get the heat transfer (Q) through the calorimeter rod by multiplying the output current (I) of the DC power supply and the voltage drop (U) across the heating resistance. At the same time, the temperature difference (T ) on the calorimeter rod is obtained by three thermometers pasted on the rod. In order to reduce the heat transfer between the calorimeter rod and the external environment, the calorimeter rod is covered with MLI, then we install a radiation shield which is also covered with MLI. In this calibration, five different cold boundary temperatures such as 20 K, 40 K, 60 K, 77 K and 100 K are obtained. As an example of the calibration, the calibration curve for the calorimeter rod under 20 K and 77 K is given in the Fig. 2. It can be seen from the calibration curve that under different temperatures, the relationship between the heat transfer (Q) and the temperature difference (T ) is nearly linear. 3.2 Measurement of Different Cold Boundary Temperatures of VDMLI In this experiment, the concentric cylindrical calorimeter introduced in the previous section is used to test the VDMLI at different cold boundary temperatures, and analyze

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Fig. 2. Calibration curve of the calorimeter rod under 20 K and 77 K

the changes of thermal performance of VDMLI caused by the change of cold boundary temperature, including heat flux and internal temperature distribution of materials. The corresponding test results are shown in Table 1 and Fig. 3 (A). Other experimental parameters are: the warm boundary is a vacuum chamber, the temperature is 293 K at room temperature, and the low-temperature vacuum is 10–5 Pa. The tested material is coated as required. By analyzing the test results, we get the following conclusions. With the increase of cold boundary temperature, the heat flux through VDMLI gradually decreases. This is because the difference of cold and warm boundary temperature decreases and the heat transfer decreases, which is consistent with the conclusion of heat transfer law. The internal temperature gradient of VDMLI is larger on the low temperature side and smaller on the high temperature side. Because with the increase of layer number, the radiation heat transfer coefficient and solid thermal conductivity increase. When the heat transfer is certain, the temperature difference between layers will decrease. Table 1. The heat flux of VDMLI at different cold boundary temperatures Cold boundary temperature (K)

Heat flux (W/m2 )

20

2.26

40

2.11

60

1.99

77

1.95

100

1.87

4 Theoretical Model and Verification Fig. 3(B) presents a schematic of the layer by layer model of multilayer insulation. VDMLI material is directly contacts the coldest surface. In this paper, the traditional layer by layer model is used to predict the temperature distribution of the VDMLI material

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Fig. 3. (A) The temperature distribution of VDMLI at different cold boundary temperatures. (B) Schematic of the layer by layer model of MLI.

[10]. The heat flow inside the insulation system includes the thermal radiation qr , the residual gas conduction qg , the solid conduction qs through the multilayer material. At steady state conditions, the total heat flux qtot through the insulation system can be written as, qtot = qr + qg + qs

(2)

The thermal radiation qr can be written as, qr =

σ (TH4 − TC4 ) 1 εH

+

1 εC

−1

= Kr (TH − TC )

(3)

and Stefan-Boltzmann coefficient σ = 5.67 × 10−8 W /(m2 · K 4 ), TH and TC are the temperatures of the warm layer and cold layer surfaces (K), respectively. εH and εC are the emissivities of the warm layer surface and cold layer surface, respectively, both taking 0.04 for aluminized shield. Kr is the radiation heat transfer coefficient Kr =

σ · ε(TH + TC )(TH2 + TC2 ) 2−ε

(4)

The residual gas conduction qg can be written as, qg = C1 · P · α(TH − TC ) = Kg (TH − TC )

(5)

Kg = C1 · P · α

(6)

C1 =

1 R γ +1 ( )2 γ − 1 8π MT

(7)

 where γ = Cp Cv , R is the general gas constant, R = 8.314 J/(mol · K), M is the molecular weight of gas (g/mol), T is the average temperature between warm layer and

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cold layer. For air, C1 = 1.1666. P is the ambient pressure. α is the accommodation coefficient, for air, α = 0.9. The solid conduction qs can be written as, qs = Ks (TH − TC ) Ks =

(8)

C2 · f · k Dx

(9)

and C2 is an empirical constant, taking 0.008 for Dacron net, f is the relative density of the spacer to the solid material (0.02), k is the thermal conductivity of the spacer material (W/(m K)), Dx is the actual thickness of the spacer between reflectors (m). For Dacron net, k is the function of temperature, k = 0.017 + 7 × 10−6 (800 − T ) + 0.0228lnT

(10)

The total thermal resistance between any two radiation shields inside the multilayer insulation material can be written as: Ri =

1 Kr + Kg + Ks

(11)

The temperature of each layer of VDMLI is calculated through an iterative process.

Fig. 4. Comparison between the experimental test results and the model settlement results under 20 K and 77 K

Based on the test results introduced in the previous section, Fig. 4 presents the comparison between the experimental test results and the model settlement results. It can be seen that the calculated value of the temperature distribution of the test material is in good agreement with the experimental value.

5 Conclusions VDMLI material is one of the effective passive thermal protection technologies for cryogenic propellant tanks. The properties of VDMLI are affected by many factors. Before practical application, it is necessary to carry out theoretical and experimental

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study on VDMLI. In this paper, a concentric cylindrical calorimeter is designed to test the thermal performance of VDMLI. The results show that with the increase of cold boundary temperature, the heat flux through VDMLI gradually decreases, and the temperature distribution inside VDMLI is obtained. In addition, this paper uses layer by layer model to simulate VDMLI, and obtains the temperature distribution inside the material. The verification shows that the theoretical results are in good agreement with the experimental results.

References 1. Meyer, M.L., Chato, D.J., Plachta, D.W., et al.: Mastering cryogenic propellants. J. Aerosp. Eng. 26(2), 343–351 (2013) 2. Nast, T.C., Frank, D.J., Feller, J.: Multilayer insulation considerations for large propellant tanks. Cryogenics 64, 105–111 (2014) 3. Fesmire, J.E., Johnson, W.L.: Cylindrical cryogenic calorimeter testing of six types of multilayer insulation systems. Cryogenics 89, 58–75 (2018) 4. Johnson, W.L.: Thermal Performance of Cryogenic Multilayer Insulation at Various Layer Spacings (2010) 5. Fesmire, J.E., Johnson, W.L., Meneghelli, B.J., Coffman, B.E.: Cylindrical boiloff calorimeters for testing of thermal insulation systems. IOP Conf. Ser.: Mater. Sci. Eng. 101, 012056 (2015). https://doi.org/10.1088/1757-899X/101/1/012056 6. Zheng, J., Chen, L., Cui, C., et al.: Experimental study on composite insulation system of spray on foam insulation and variable density multilayer insulation. Appl. Thermal Eng. 130, 161–168 (2018) 7. Celik, D., Hurd, J., et al.: A calorimeter for multilayer insulation (MLI) performance measurements at variable temperature. Cryogenics 55–56, 73–78 (2013) 8. Li, X., Xu, D., Shen, F.Z., Liu, H.M., Li, L.F.: Design of a multi-layer insulation(MLI) measurement system with a G-M cryocooler. IOP Conf. Ser.: Mater. Sci. Eng. 502, 012075 (2019) 9. Hedayat, A., Hastings, L.J., Brown, T.: Analytical modeling of variable density multilayer insulation for cryogenic storage. American Institute of Physics (2002) 10. Wang, B., Huang, Y.H., Li, P., et al.: Optimization of variable density multilayer insulation for cryogenic application and experimental validation. Cryogenics 80, 154–163 (2016)

Effect of Fill Ratios on the Heat Transfer Performance of Nitrogen Cryogenic Pulsating Heat Pipe Yaran Shi1,2 , Dong Xu1(B) , Bingkun Lyu1,2 , Jixiang Yan1,2 , Tao Wang1 , Zhixiong Wu1 , and Laifeng Li1,2(B) 1 State Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Chinese Academy of Sciences, Beijing 100190, China {xudong,laifengli}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100049, China

Abstract. Cryogenic pulsating heat pipe (PHP) is a novel heat transfer unit with unique working mechanism. The advantages of high heat transfer efficiency and strong environmental adaptability make them as desirable means for thermal management in cryogenic applications such as cooling superconductors. In this present paper, a nitrogen-based PHP composed of 4 parallel stainless steel tubes was experimentally studied. The PHP was operated in vertical orientation where the evaporator was at the bottom position, and the lengths of the evaporator, condenser and adiabatic section were 50 mm, 50 mm and 100 mm, respectively. The inner diameter of the pipe was 1mm and the outer diameter was 1.59 mm (1/16 inch). Experiments at three fill rates of 50%, 70% and 90% were performed and the heat transfer performance of nitrogen-based PHP was investigated under various working conditions by changing heat loads and fill ratios. Keywords: Cryogenic Pulsating Heat Pipe · Nitrogen · Fill Ratio

1 Introduction The pulsating heat pipe (PHP) is considered as a novel passive heat transfer device for superconducting magnet cooling applications due to its high heat transfer efficiency and strong adjustability for long-distance cooling. Cryogenic PHPs typically use nitrogen (63.2–126.2 K) [1, 2], neon (24.6–44.5 K) [3, 4], hydrogen (14.0–33.2 K) [5, 6] and helium (2.2–5.2 K) [7–9] as the working fluid, which is selected depending on the critical temperature (Tc) of the superconducting material. With the appearance of high-temperature superconducting magnets with critical temperature higher than 90 K [10], nitrogen-based PHPs have been paid more attention [11–13]. However, the researches on the influence of various parameters on the operating characteristics of the nitrogen PHPs are not comprehensive, and it is difficult to grasp the operation law. Therefore, it is necessary to complete the database of heat transfer performance of nitrogen PHPs under different working conditions. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 494–500, 2023. https://doi.org/10.1007/978-981-99-6128-3_63

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This paper reported the experimental study of a 2-turn nitrogen PHP. The experimental device was described, and experiments under three different filling ratios were carried out to explore the effect of the filling ratios on the heat transfer characteristics of the nitrogen PHP.

2 Experimental Apparatus and Procedure 2.1 Experimental Device Figure 1 shows the schematic of the experimental system and the details of the device. The condenser of the PHP was thermally tightened to the cold head via a conduction cooling plate made of oxygen-free copper. In addition, a heater controlled by a DC power supply provided a steady heat input to the evaporator. A total of eight thermometers (PT100) and two pressure transducers were installed. The data were collected by a Keithley® Model 3706 Multimeter and then output to a computer. 2.2 Cryogenic PHP Prototype The structure of the PHP used in the experiment is shown in Fig. 2. The PHP is composed of 4 parallel stainless steel capillary tubes whose two ends are connected to the filling tube by a T-junction. The outer diameter of the stainless steel tube is 1.59 mm (1/16 inch) and the inner diameter is 1 mm, which is less than the critical diameter of the nitrogen PHP, given by:  D = Dcrit ≤ 2 g(ρlσ−ρv ) (1) where g is the gravitational acceleration, σ, ρl and ρv are the surface tension, liquid density, and vapor density of the working fluid respectively. The lengths of condenser, adiabatic section and evaporator are 50 mm, 100 mm and 50 mm respectively and the tube is welded by tin in the groove of a 100 mm long and 50 mm wide copper plate. 2.3 Experimental Procedure The Filling and Cooling Process It should be noted that during the filling process, the mass of the working fluid introduced into the PHP is assessed by “fill ratio (FR)”, which represents the percentage of the liquid working fluid volume (Vl ) to the total volume of the PHP (VPHP ), namely: FR =

Vl VPHP

× 100%

(2)

Therefore, the mass of the working fluid in the PHP at a specific fill ratio can be calculated, given by: mt = ρl Vl + ρv (VPHP − Vl )

(3)

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Fig. 1. (a) Schematic illustration of the experimental system, (b) the structure of the device: (1) buffer tank; (2) pressure transducer; (3) valve; (4) filling tube; (5) nitrogen gas cylinder; (6) heater; (7) cryogenic pulsating heat pipe; (8) thermal radiation shield; (9) temperature sensor; (10) conduction cooling plate; (11) cold head of cryocooler; (12) cryostat; (13) vacuum pump; (14) single-stage GM cryocooler (Sumitomo, Model: CH-110, 175 W @ 77 K); (15) valve panel.

where ρl and ρv are the densities of the saturated liquid and saturated vapor at 77 K, respectively. Vl and VPHP are the volumes of the liquid nitrogen and the PHP, respectively. This value can be controlled by the pressure change of the buffer tank at the beginning and end of the filling process. Considering nitrogen as an ideal gas, by the law of conservation of mass, there is: P0 VBT Rg Tamb

=

Pf VBT Rg Tamb

+

Pf VFT Rg TFT

+ mt

(4)

where P0 and Pf are the initial and final pressures of the buffer tank during the filling process, respectively. VBT and VFT are the volumes of the buffer tank and filling tube,

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respectively. Rg is the gas constant. mt is the mass of the nitrogen in the PHP. Tamb is the room temperature and TFT is the temperature of the filling tube (average of room temperature and PHP temperature). The Heating Process Heat load with gradual increments was applied and the pressure and temperature changes were recorded during the heating process. The effective thermal conductivity Keff , which is often used to evaluate the thermal performance of the PHP, was determined by: Keff =

Q·L A·T

=

4QL nπ d 2 (Te −Tc )

(5)

where Q is the heat load applied to the evaporator. L is the length of the adiabatic section. n is the number of parallel capillaries. d is the inner diameter of the PHP. Te and Tc are the temperatures of evaporator and condenser, respectively.

Fig. 2. 2-turn Nitrogen PHP prototype

3 Results and Discussion 3.1 Cooling Down Process The typical cooling process of nitrogen PHP is shown in Fig. 3. Firstly, the cryocooler was started and the valve V1 connected to PHP and buffer tank was opened to make the PHP fill with nitrogen gas. The condenser temperature (Tc ) decreased rapidly with the decrease of the temperature of cold head, and it took 1.04 h to drop to 82 K (point A). Subsequently, more and more liquid plugs were generated at the condenser and part of the adiabatic section along with time. Finally, at point B, V1 was closed indicating that the filling process was completed. It was worth noting that the temperature of the condenser was subsequently controlled to increase from 74 K to 77 K. As the temperature of the condenser stabilized at about 77 K, the cooling process was completed, which meant that the condition was available for operating experiments on heat transfer characteristics.

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Fig. 3. The cooling process of the nitrogen PHP (FR = 49.99%).

3.2 Thermal Characteristics Experiments on the heat transfer characteristics of nitrogen pulsating heat pipes under different fill ratios (FR) were carried out. Figure 4 (a) (b) (c) show the changes of temperature and pressure under different heat loads at fill ratio of 49.99%, 70.33%, and 89.93%. With each increase of the heat load, the temperature of the condenser (Tc ) was always constant at 77 K. Meanwhile, the temperature of the evaporator (Te ) and the pressure of PHP (PPHP ) also stabilized at each position after a transient state, and the temperature difference between the evaporator and condenser gradually increased, indicating that the PHP efficiently transferred heat loads from the evaporator to condenser. As the heating power increased to a certain value, the “dryout” can be detected, which was manifested by a rapid increase in evaporator temperature and a rapid decrease in pressure, indicating that PHP can no longer operate as an effective heat transfer element. This value is the ultimate heat load that can be transmitted by PHP, and it increases with the increasement of the filling ratio, which is caused by the reinforcement of the backflow driven force with the increase of the weight of liquid plug. The maximum heat load of 43.941 W can be obtained at the fill ratio of 89.93%. It should be added that once the internal pressure of PHP exceeds the range of the pressure transducer, the value of PPHP will always be displayed as 1000 kPa (for FR = 70.33% and FR = 89.93%, in the condition of a large heat load input). Figure 4 (d) presents the variation of effective thermal conductivity as a function of heat load and fill ratios. It can be seen from the figure that at the filling ratio between 50% and 90%, the effective thermal conductivity increased at first and then decreased with the increasing heat load. With a low heat load, the temperature difference between condenser and evaporator was small, and few vapor bubble was generated at the evaporator, which meant that the thermally driven force was insufficient to generate adequate disturbance. The driven force as well as pulsating velocity of working fluid increased gradually with the increasement of heating power, so the heat transfer efficiency increased. However, with a high heat load, the decrease of heat transfer area of the liquid film caused by the increasement of bubble volume will also led to the deterioration of heat transfer performance. Therefore, there was a peak of thermal conductivity. At the fill ratio of 49.99%, 70.33% and 89.93%, the value was 31229.4 W/(m K), 27264.3 W/(m K), and

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27344.5 W/(m K), respectively. Furthermore, at the fill ratio of 49.99%, the heat transfer performance of PHP was found to be significantly better than that of other fill ratios, reaching the highest value of thermal conductivity of 31229.4 W/(m K) at the heat load of 18.696 W. In addition, in terms of effective thermal conductivity, the values were very similar at the fill ratio of 70.33% and 89.93%. The above results indicated that there existed an optimal fill ratio of nitrogen PHP, which was associated with the heat load input.

Fig. 4. Operation characteristics of the 2-turn nitrogen PHP under different heat loads at different fill ratios: (a) 49.99%, (b) 70.33%, (c) 89.93%; (d) effect of fill ratio and heat load on effective thermal conductivity of the nitrogen PHP.

4 Conclusion The thermal performance of a 2-turn nitrogen PHP at different fill ratios (49.99%, 70.33% and 89.93%) was studied. The experimental device was introduced, and the cooling and heating processes of nitrogen PHP were analyzed. Experimental results showed that the

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effective thermal conductivity showed a trend of first increasing and then decreasing until the dry-out occurred for each fill ratio. In addition, with the increase of the fill ratio, the ultimate power that can be transmitted increased. The maximum effective thermal conductivity of 31229 W/(m K) was observed with a fill ratio of 49.99% at the heat load of 18.696 W.

References 1. Fonseca, L.D., Miller, F., Pfotenhauer, J.: Experimental heat transfer analysis of a cryogenic nitrogen pulsating heat Pipe at various liquid fill ratios. Appl. Therm. Eng. 130, 343–353 (2018) 2. Li, Y., Wang, Q., Chen, S., et al.: Experimental investigation of the characteristics of cryogenic oscillating heat pipe. Int. J. Heat Mass Transf. 79, 713–719 (2014) 3. Liang, Q., Li, Y., Wang, Q.: Effects of filling ratio and condenser temperature on the thermal performance of a neon cryogenic oscillating heat pipe. Cryogenics 89, 102–106 (2018) 4. Liang, Q., Li, Y., Wang, Q.: Study on a neon cryogenic oscillating heat pipe with long heat transport distance. Heat Mass Transf. 54(6), 1721–1727 (2018) 5. Sun, X., Li, S., Jiao, B., et al.: Experimental study on hydrogen pulsating heat pipes under different number of turns. Cryogenics 111, 103174 (2020) 6. Sun, X., Li, S., Jiao, B., et al.: Experimental study on a hydrogen closed-loop pulsating heat pipe with two turns. Cryogenics 97, 63–69 (2019) 7. Xu, D., Li, L., Liu, H.: Experimental investigation on the thermal performance of helium based cryogenic pulsating heat pipe. Exp. Thermal Fluid Sci. 70, 61–68 (2016) 8. Li, M.: Effect of filling ratio and orientation on the performance of a multiple turns helium pulsating heat pipe. Cryogenics 100, 62–68 (2019) 9. Li, M., Li, L., Xu, D.: Effect of number of turns and configurations on the heat transfer performance of helium cryogenic pulsating heat pipe. Cryogenics 96, 159–165 (2018) 10. Iwasa, Y.: Design and operational issues for 77-K superconducting magnets. IEEE Trans. Magn. 24(2), 1211–1214 (1988) 11. Mito, T., Natsume, K., Yanagi, N., et al.: Achievement of high heat removal characteristics of superconducting magnets with imbedded oscillating heat pipes. IEEE Trans. Appl. Supercond. 21(3), 2470–2473 (2011) 12. Bruce, R., Barba, M., Bonelli, A., et al.: Thermal performance of a meter-scale horizontal nitrogen pulsating heat pipe. Cryogenics 93, 66–74 (2018) 13. Sagar, K.R., Desai, A.B., Naik, H.B., et al.: Experimental investigations on two-turn cryogenic pulsating heat pipe with cylindrical shell-type condenser. Appl. Therm. Eng. 196, 117240 (2021)

Numerical Study of a Single-Loop Nitrogen Pulsating Heat Pipe Bingkun Lyu1 , Dong Xu2(B) , Wei Wang1(B) , Yaran Shi2 , Chuanjun Huang2 , Rongjin Huang2 , and Laifeng Li1,2 1 Songshan Lake Materials Laboratory, Dongguan, China

[email protected]

2 State Key Laboratory of Technologies in Space Cryogenic Propellants,

Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Beijing, China [email protected]

Abstract. In this study, a numerical simulation of a two-dimensional single-loop cryogenic pulsating heat pipe (PHP) has been carried out using nitrogen as the working fluid. The Volume of Fluid (VOF) model was employed for the two-phase flow simulation. The initial state and the operating state of the nitrogen PHP were numerically studied. Due to the surface tension effect, an alternation of liquid slugs and vapor plugs distribution was formed in the initial state of the PHP. When a constant heat flux boundary condition was applied to the evaporator section, the circular flow was observed. The relationship between flow and heat transfer of the PHP was analyzed. Keywords: Pulsating Heat Pipe · Nitrogen · Numerical Simulation · Volume of Fluid

1 Introduction Pulsating heat pipe (PHP), invented by Akachi in 1990, is a new type of passive heat transfer device[1]. It consists of a meandering capillary tube that goes back and forth several times between the evaporator section and condenser section enclosing the working fluid. The heat transport is conducted by the pulsation of the vapor plugs and liquid slugs driven by the pressure rise due to phase change when there exists a temperature gradient between both ends of the PHP. High heat transfer efficiency, simple structure, strong adjustability, and high reliability make the PHP more accommodative in cryogenic applications like cooling superconducting magnets. Although the structure of the pulsating heat pipe is very simple, the operation mechanism of the pulsating heat pipe is relatively complicated [2]. However, the theoretical research of cryogenic pulsating heat pipes is still in its infancy and the visualization has not been realized so far. With the development of computer technology, computational fluid dynamics (CFD) numerical simulation is applied to the theoretical analysis of cryogenic PHP. In this paper, a CFD model of the single-loop nitrogen pulsating heat pipe was established. The fluid flow and heat transfer in the initial and operating states were analyzed. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 501–507, 2023. https://doi.org/10.1007/978-981-99-6128-3_64

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2 Numerical Model 2.1 Geometric Model In this work, the model of nitrogen pulsating heat pipe was established and solved by the ANSYS. The modeling software ANSYS DesignModeler was used to draw the geometric model of the two-dimensional (2D) pulsating heat pipe, as shown in Fig. 1(a). The lengths of the adiabatic section, the condenser section, and the evaporator section of the PHP were 100 mm, 50 mm, and 50 mm respectively. The pulsating heat pipe model was placed vertically with the bottom heating mode. The pulsating heat pipe had an inner diameter of 2 mm and used nitrogen as the working medium, which satisfies the critical diameter limit of Eq. (1).  σ D ≤ Dcr = 2 (1) g(ρl − ρv ) where g is the gravitational acceleration, Dcr is the critical diameter of the PHP, D is the inner diameter of the PHP, σ, ρ l , and ρ v are the surface tension, liquid density, and vapor density properties of the fluid, respectively.

Fig. 1. (a) Geometric model and (b) mesh generation of two-dimensional single-loop nitrogen pulsating heat pipe.

The geometric model has meshed in ANSYS Meshing. A structured grid is superior for calculating the surface tension effect accurately. Therefore, the physical model was meshed with a quadrilateral structured grid, as shown in Fig. 1(b). The number of cells is 103740 and the number of nodes is 98800. 2.2 Governing Equation The flow in a cryogenic PHP is a two-phase flow. The VOF model, which is based on the fact that two or more fluids do not permeate each other, was used to simulate the

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two-phase flow. In the calculation cell, the sum of the volume fractions of each phase is uniform. The variables and properties within the cell can be single-phase or two-phase mixtures, depending on the volume fraction value. The VOF method was used to track the surface interface of each cell of the PHP based on the volume fractions of gas phase α v and liquid phase α l , where the subscripts v and l represent gas and liquid phases, respectively. In the control volume: αv + αl = 1 In this model, the gas phase is considered to be a compressible ideal gas and liquid phase is incompressible. The continuity equation of the volume fraction of two-phase flow is given by:   ∂  ˙ vl − m ˙ lv ) (ρl αl ) + ∇ · ρl αl v = Sl + (m ∂t   ∂  ˙ lv − m ˙ vl ) (ρv αv ) + ∇ · ρv αv v = Sv + (m ∂t

(2) the the

(3) (4)



where v is velocity vector, S l and S v are the source terms of liquid and gas phases, ˙ lv are mass transfer from vapor to liquid and mass transfer from respectively, m ˙ vl and m liquid to vapor, respectively. The evaporation and condensation process can be expressed by the Lee model as follows: If T l > T sat , the evaporation process occurs: m ˙ lv = rlv αl ρl

(Tl − Tsat ) Tsat

(5)

If T v < T sat , the condensation process occurs: m ˙ vl = rvl αv ρv

(Tsat − Tv ) Tsat

(6)

where T l and T v are the temperatures of the liquid phase and the gas phase of the working fluid, respectively, T sat is the saturation temperature, r lv and r vl are relaxation factors of mass transfer time. These coefficients affect the accuracy of numerical convergence and interface temperature. A default value of 0.1 is usually set so that the interface temperature remains numerically close to the saturation temperature of the working fluid. The momentum equation is solved in Eq. (7), which depends on the volume fraction of all phases.         ∂    ρ v +∇ · ρ v v = − ∇p + ∇ · μ ∇ v +∇ v T + ρ g +F (7) ∂t 

where ρ is the average density within the control volume, p is pressure, μ is viscosity, g 

is gravitational acceleration, F is the volumetric force calculated by the Laplace equation caused by a curved interface.

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Surface tension is due to the cohesion between molecules in the fluid and the surface forces it generates dominate the small channels. For the modeling of Surface tension, the Continuum Surface Force (CSF) model was adopted. In the VOF model, surface tension is added to the momentum equation as the source term, as shown in the following equation: 

F = σlv

αl ρl kv ∇αv + αv ρv kl ∇αl 1 2 (ρl

(8)

+ ρv )

αl v where σ lv is the surface tension coefficient, k stands for curvature, kl = ∇α , kv = α ∇αl . v The energy equation is:       ∂ (9) + Sh (ρE) + ∇ · v (ρE + p) = ∇ · λeff ∇T + τ · v ∂t where E is the internal energy, λeff is the effective thermal conductivity, and S h is the energy source term caused by phase transition, which is obtained by mass transfer rate ˙ lv = hLH m ˙ vl ). multiplied by latent heat (Sh = −hLH m

2.3 Numerical Simulation Method In order to make the simulation close to the real situation, the NIST Real Gas model was used for the material properties of nitrogen by associating Refprop software developed by the National Institute of Standards Technology (NIST). Since the liquid phase of nitrogen is an incompressible fluid, its physical property parameters change very little with temperature. So, the physical property value at 77.36 K can be used. The physical parameters of nitrogen gas and liquid nitrogen are shown in Table 1. Table 1. Nitrogen property parameter. Properties

Nitrogen gas

Liquid nitrogen

Density (kg/m3 )

Real-gas-law

806.08

Specific heat at constant pressure (J/kg·K)

Real-gas-law

2041.5

Thermal conductivity (W/(m·K))

Real-gas-law

0.14581

Kinematic viscosity coefficient (kg/(m·s))

Real-gas-law

0.00016065

Molar mass (kg/kmol)

28

28

The two-phase flow model VOF equation adopted the explicit form. The dimensionless Coulang number was set to 0.25. The implicit volumetric force equation was used. The saturation temperature of the nitrogen changes with pressure, which is obtained from the Refprop database. The fitting formula is shown in Eq. (11). Tsat (p) = 64.23743 + 1.47536 × 10−4 p − 3.18399 × 10−10 p2 + 4.03529 × 10−16 p3 −2.82499 × 10−22 p4 + 1.08921 × 10−28 p5 − 2.16589 × 10−35 p6 + 1.73337 × 10−42 p7

Other solution settings are listed in Table 2.

(11)

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Table 2. Solution setting of 2D nitrogen pulsating heat pipe by Fluent numerical simulation. Item

Parametric form

Phase

Primary phase-liquid nitrogen and secondary phase-vapor nitrogen

Gravity

Y axis −9.81 m/s2

Surface tension coefficient

0.009133 n/m

Solver setup

Pressure-based

Time setup

Transient

Flow model

Laminar

Operating conditions

Operating temperature 77.36 K Operating pressure 101325 Pa

Pressure-velocity coupling

PISO

Gradient spatial discretization

Least Squares Cell-Based

Pressure spatial discretization

Body Force Weighted

Density spatial discretization

Second-Order Upwind

Momentum spatial discretization

Second-Order Upwind

Volume fraction spatial discretization

Geo-Reconstruct

Energy spatial discretization

Second-Order Upwind

Transient formulation

First Order Implicit

Filling ratio

50%

Time step size

0.0001 s

Maximum iterations

200

3 Results and Discussion 3.1 Initial State In ANSYS Fluent, the temperature of each cell in the initial state was set as 77.36 K. The boundary conditions were set as follows: heat flux of the evaporator section, the condenser section, and the adiabatic section was 0. Figure 2(a) shows the gas-liquid distribution cloud diagram of the initial state of the 2D single-loop nitrogen pulsating heat pipe. It can be seen that there are a random alternate distribution of vapor plugs and liquid slugs in the PHP and the gas-liquid interface is obvious due to surface tension effects. From Fig. 2(b) and (c) we can find that the temperature is kept at about 77 K and the pressure stays around one atmosphere and the gradient can be negligible. This is because the nitrogen PHP is affected by gravity. 3.2 Operating State At the operating state, the boundary conditions were changed as follows: heat load of 5 W was input in the evaporator section; the adiabatic section was set as the adiabatic

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Fig. 2. Initial state contour diagram of two-dimensional single-loop nitrogen pulsating heat pipe: (a) nitrogen gas volume fraction; (b) pressure; (c) temperature.

boundary condition; the condenser section was set at a constant temperature of 75 K. The operating state of the PHP was simulated for 2.8 s, as shown in Fig. 3. As can be seen from Fig. 3(a) that there is a circular flow in the tube. The flow remains unidirectional and the flow type is plug flow. With the increase in running time, the descending tube is mainly the liquid phase while the ascending tube is full of the gas phase. This is because when a PHP operates successfully, vapor plugs generate and grow in the evaporator, simultaneously, they collapse and shrink in the condenser. This volumetric expansion and contraction of the vapor bubbles during phase changes contribute to pressure perturbations, which excites a pulsating or circular flow motion that is able to transfer the heat from the evaporator to the condenser, as shown in Fig. 3(b). At the beginning of the operating state, the temperature of the condenser section remains at 75 K, while the temperature of the evaporator section rises due to the absorption of heat load. Subsequently, as the working fluid in the condenser section enters the left adiabatic section, the temperature of the left adiabatic section soon drops to the same as that of the condenser section. The temperature of the evaporator section continues to rise, and the heat enters the right adiabatic section with the working fluid, causing the temperature to rise in the right adiabatic section. Affected by the flow direction, the left part of the evaporator section is cooled first. The heat load from the right part of the evaporator section will reach the condenser section and be cooled with the flow of the working fluid. Figure 3(c) shows the pressure contour to visualize the pressure difference between the condenser and the evaporator driving liquid nitrogen pulsating/circular flow. It can be found that the pressure in the evaporator is greater than that in the condenser. At the same time, there is an obvious pressure gradient on the gas-liquid interface. The results show that the pressure difference caused by the thermal excitation phase transition is the main driving force for the motion of the nitrogen pulsating heat pipe and thus the heat load can be transferred efficiently.

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Fig. 3. (a) Gas-liquid distribution contour diagram, (b) temperature contour diagram, and (c) pressure contour diagram of working state of two-dimensional single loop nitrogen pulsating heat pipe.

4 Conclusions The initial state and the operating state of the 2D single-loop nitrogen PHP were simulated by using the VOF model in the ANSYS Fluent. In the initial state, the gas-liquid distribution of nitrogen alternates randomly. During the operating state, the flow pattern of the nitrogen PHP is circular flow and the flow type is slug flow. The simulation results are very helpful to understand the flow and heat transfer mechanism of the cryogenic PHP. Acknowledgments. The project is supported by the National Key Research and Development Program of China (Grant No. 2022YFA1603904), the Basic and Applied Basic Research Foundation of Guangdong Province (Grant No. 2020B1515120084), the National Natural Science Foundation of China (Grant No. 52071223) and the Key-Area Research and the Development Program of Guangdong Province (Grant No. 2020B0101340002).

References 1. Akachi, H.: Structure of a heat pipe. U.S. Patent No. 4921041 (1990) 2. Khandekar, S., Pnigrahi, P.K., Lefèvre, F., et al.: Local hydrodynamics of flow in a pulsating heat pipe: a review. Frontiers in Heat Pipes 1, 1–20 (2010)

Modelling and Experimental Evaluation of the Thermal Budget of the UKRI STFC Daresbury Laboratory Vertical Test Facility Alastair White1,2,3 , Ayomikun Akintola1,2 , Keith Dumbell1,2 , Shuisheng He3 , Sean Hitchen1,2 , Conor Jenkins1,2 , Bo Liu3 , Andrew J. May1,2(B) , Shrikant Pattalwar1,2 , Mark Pendleton4 , and Paul A. Smith1,2 1 Accelerator Science and Technology Centre, STFC Daresbury Laboratory, Keckwick Lane,

Warrington WA4 4AD, UK [email protected] 2 Cockcroft Institute, Keckwick Lane, Warrington WA4 4AD, UK 3 Department of Mechanical Engineering, University of Sheffield, Sheffield S1 3JD, UK 4 Technology Department, STFC Daresbury Laboratory, Keckwick Lane, Warrington WA4 4AD, UK

Abstract. A novel vertical test facility (VTF) has now been in operation for over a year at the UKRI STFC Daresbury Laboratory. The facility is capable of testing 3 jacketed superconducting RF cavities in a horizontal configuration in a single cool-down run with superfluid helium at 2 K. This paper reports thermal modelling of the VTF cryostat and comparison with experimentally measured heat loads. Keywords: Vertical test · Superconducting RF cavities · Large-scale test facility · Thermal modelling

1 Introduction The new Superconducting Radio Frequency Laboratory (SuRF Laboratory) Vertical Test Facility (VTF) cryostat has recently completed the commissioning phase at the UKRI STFC Daresbury Laboratory. The VTF cryostat, shown in Figs. 1 and 2 below, supports 2 K RF characterisation of three jacketed SRF cavities in a single cool-down run. As part of the UK’s in-kind contribution to the European Spallation Source (ESS), STFC are testing 88 high-beta cavities. Thermal modelling of the VTF has been under- taken to understand the steady state static heat loads on the 2 K and 80 K stages. Ther- mal analysis comprised hand calculation evaluation of the thermal loads, Finite Element (FE) modelling, and comparison to experimentally-determined values.

2 VTF Cryostat Architecture The unique cryogenic architecture of the VTF tests jacketed cavities in a horizontal configuration as described in Sect. 2.1 and Reference [1]. Compared to a conventional configuration, liquid helium (LHe) consumption is reduced by up to 70%. The VTF © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 508–514, 2023. https://doi.org/10.1007/978-981-99-6128-3_65

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comprises the Cavity Support Insert (CSI, shown in Fig. 1) which holds the three cavities and the fixed helium pipes, and the cryostat vessel (shown in Fig. 2). The cryostat vessel provides an insulating vacuum, a thermal shield at ~80 K, and magnetic shielding to reduce the ambient field experienced by the cavities during test. The CSI is structurally separate to the cryostat vessel and is loaded via crane as shown in Fig. 2.

Fig. 1. Three jacketed cavities wrapped in MLI loaded on CSI (left)

Fig. 2. CSI being loaded into Inner Vacuum Chamber (IVC) (right)

2.1 2 K Circuit In contrast to a conventional vertical test facility which fully immerses bare cavities in a bath of LHe, the Daresbury VTF tests cavities after the helium jackets have been welded on by filling LHe into individual jackets. The 2 K stage of the cryostat is comprised of up to three jacketed cavities, a 2 K header bath above the cavities, and helium connections, fill, and return lines mounted on the CSI. Each component on the 2 K stage is individually wrapped in Multi-Layer Insulation (MLI) to reduce radiative heat load. 2.2 80 K Circuit The vacuum vessel is comprised of an inner and outer chamber (IVC and OVC). The IVC provides an insulating vacuum around the jacketed cavities, with the inner wall of the IVC and thermal shield (wrapped in multilayer insulation, MLI) being cooled by 50 K gaseous helium (GHe). The OVC houses the thermal shield.

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2.3 300 K Components The CSI lid, alongside the OVC outer walls, are at room temperature throughout operation. MLI covered radiation baffles (visible in Fig. 2) are not actively cooled but are at a floating temperature and act to reduce the radiative heat load incident on the 2 K stage from the underside of the lid.

3 Evaluation of Thermal Loads As both the IVC and OVC are held under vacuum < 1e-3 mbar, convective heat loading has been neglected in this analysis. The approaches used for analysis of conductive and radiative loading are described in the following sections. 3.1 Conductive Thermal Loading Heat transfer by conduction to the 2 K circuit takes place through the helium lines, Ultra High Vacuum (UHV) lines and RF cables to each cavity from the lid (300 K). Limited conduction to the 80 K thermal shield takes place through the IVC supporting pillar. Each conductive component was considered individually to determine the heat contribution through each. Hand calculation of the conduction heat loads (Q) were performed to validate the FEA results using Fourier’s law of thermal conduction, of the form Thot ˙ Conduction = A ∫ k(T )dT Q L Tcold

(1)

where A is the cross-sectional area (taken to be constant), L is length, k is thermal conductivity, and T is temperature. Thermal conductivity integrals for 316 stainless steel and OFHC copper (RRR = 100) were taken from NIST [2, 3]. Length values represent the heat transfer path length and not necessarily the total component length. For example, the helium fill line is the length from the 300 K lid to the 2 K top cavity, as the pipe below the top cavity is assumed to contain and be in thermal equilibrium with 2 K LHe. 3.2 Radiative Thermal Loading Radiation transfers heat to the 2 K stage via the lid through two floating MLI-covered baffle plates and from the 80 K thermal shield surrounding the CSI on the remaining three sides. The outer vessel walls radiate heat to the 80 K thermal shield. Radiation heat loads were first found by hand calculation to validate the FEA results considering radiative transfer (Q) between each individual component surface as   σ T14 − T24 ˙ (2) QRadiation1→2 = 1−∈ 1−ε2 1 1 A1 ε1 + A1 F1→2 + A2 ε2 where σ is the Stefan-Boltzmann constant, T is surface temperature, A is surface area, ε is emissivity, and F is the view factor that accounts for the geometry of the system. Emissivities were taken of 0.5 for stainless steel [4] and 0.0017 for MLI [5]. View factors calculated for relevant components were found by taking general forms from the literature [5]. Baffle temperatures were found by balancing radiative heat loads.

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4 Modelling of Thermal Loads Following hand calculations to obtain approximate values, FE simulations were carried out using ANSYS of each conductive and radiative heat load contribution. 4.1 Conductive Thermal Loading The same thermal conductivity values across the temperature domain were used in both the hand calculations and FE simulations. The end temperatures of each component were set as boundary conditions (300 K, 80 K, 2 K as appropriate). The simulated heat flux across one of the UHV pumping lines is shown in Fig. 3 as an example. 4.2 Radiative Thermal Loading In contrast to the hand calculations, the radiative FE simulations were performed on the 2 K stage and 80 K stages as respective wholes, allowing simulation of the radiative interaction between every surface in the VTF. The same temperature and emissivity conditions for each individual surface were used as in the hand calculations. Figure 4 shows the temperatures found in the 2 K loading radiative simulation.

Fig. 3. Heat flux across middle cavity UHV line in conduction simulation (left)

Fig. 4. Temperature distribution on top plate, baffles, header tank, and top cavity in radiative simulation (right)

5 Experimental Measurements LHe transfers into the 2 K circuit are done in this system as a “single-shot” process; transfers of 4 K liquid are carried out each morning to fill the system to the top of the header tank, then a set of pumps is used to bring the partial pressure down to cool the liquid (and hence the cavities) to 2 K, and the day’s RF operations are carried out (for a detailed account, see Ref. [1]). Data was taken from a flow meter (FM2230) downstream of the 2 K pumps which found the helium boil-off rate from the 2 K stage to be 0.15 g/s when a steady-state

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condition had been reached following the fill and pump-down. Taking the latent heat of vapourisation at 2 K to be 93.1 J/mol [6] gave a calculated total heat load on the 2 K stage to be 3.5 W. For the 80 K stage, control valve CV2205 was typically operated at 7.5% opening cooling the thermal shield to ~ 100 K with a helium gas flow rate of 3.4 g/s in steady state. The measured inlet and outlet temperatures of the cooling gas were 50 K and 130 K respectively. Measured outlet pressure was 1 bar absolute with negligible pressure drop through the shield circuit. Using the enthalpy integral of helium gas between 50 K and 130 K (38283 kJ/kg) [7], the total heat load on the thermal shield was determined to be 130 W. It should be noted here that the measured temperature of the shield (~100 K) was different to that assumed during the hand calculations and simulations of 80 K (see following section for discussion of results).

6 Results and Discussion Tables 1 and 2 below present all of the individual and total conductive and radiative heat loads on the 2 K and 80 K circuits determined by hand calculations and FEA, along with the experimentally measured loads for comparison. In both the 2 K and 80 K cases, the analyses found good agreement between hand calculation and FE models, but overestimated heat loads compared to the experimental results. This is largely expected as the theoretical values do not account for thermal contact resistances etc., or, in the case of the 2 K stage, cooling of the pumping lines provided by returning vapour. On the 80 K stage, uncertainty exists in the assumed MLI emissivity value due to variations in manufacturing, assembly, etc. As such, the modelling methods used here provide a conservative guide to cryostat thermal budgeting. Table 1. Heat loads on the 2 K stage

Conduction

Heat load contribution

Hand calculation heat load (W)

FE heat load (W)

Helium bath supports

2.0

2.0

RF Cables Cu core: 0.54 To top cavity 0.44 To middle cavity 0.39 To bottom cavity

0.54 0.44 0.39

RF Cables steel outer: To top cavity To middle cavity To bottom cavity

0.010 0.0090 0.0080

0.010 0.0090 0.0080

Experimental result (W)

(continued)

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Table 1. (continued)

Radiation

TOTAL

Heat load contribution

Hand calculation heat load (W)

FE heat load (W)

UHV Lines: To top cavity To middle cavity To bottom cavity

0.19 0.16 0.13

0.19 0.15 0.12

He liquid column and return line

0.75

1.0

He top fill line

0.18

0.17

He bottom fill line

0.093

0.091

Subtotal

4.9

5.1

Top of helium bath 0.32

0.039

Side of helium bath 0.0051

0.040

Bottom of helium bath

0.0039

0.027

Top section of top cavity

0.0040

0.039

Top section of middle cavity

0.0039

0.037

Top section of bottom cavity

0.0039

0.036

Bottom section of top cavity

0.0039

0.032

Bottom section of middle cavity

0.0039

0.032

Bottom section of bottom cavity

0.004

0.032

Side of cavities

0.24

0.041

Subtotal

0.38

0.35

5.3

5.5

Experimental result (W)

3.5

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Conduction

Radiation

Heat load contribution

Hand calculation heat load (W)

FE heat load (W)

IVC structural support

2.4

2.3

Subtotal

2.4

2.3

Vessel wall to IVC

160

158

Lower baffle to thermal shield

0.81

0.64

Subtotal

161

159

163

161

Total

Experimental result (W)

130

7 Conclusion An analysis of the steady-state static heat loads in the UKRI STFC Daresbury Laboratory Vertical Test Facility has been completed. Very good agreement is shown between hand calculations and FEA, which both provide conservative estimates with respect to experimentally measured values and hence provide a conservative guide to cryostat thermal budgeting and design.

References 1. May, A., et al.: Commissioning of the UKRI STFC daresbury vertical test facility for jacketed SRF cavities. In: 19th International Conference on RF Superconductivity East Lansing, Michigan, USA: SRF2021 (2022) 2. NIST, Material Properties: 316 Stainless. https://trc.nist.gov/cryogenics/materials/316Stainl ess/316Stainless_rev.htm. Accessed 22 Apr 2022 3. NIST, Material Properties: OFHC Copper (UNSC10100/C10200). https://trc.nist.gov/cryoge nics/materials/OFHC%20Copper/OFHC_Copper_rev1.htm. Accessed 22 Apr 2022 4. Nam, H., et al.: Experimental study on the emissivity of stainless steel. In: Proceedings of the Korean Nuclear Society Spring Meeting, Cheju, South Korea (2001) 5. Wilson, M.: Lecture 5: Cryogenics and Practical Matters. CERN, Geneva (2015) 6. Donelly, R.J., Barenghi, C.F.: The observed properties of liquid helium at the saturated vapor pressure. J. Phys. Chem. Ref. Data 27(6), 1217–1274 (1998) 7. NIST, Isobaric Properties for Helium. https://webbook.nist.gov/cgi/fluid.cgi?Action=Load& ID=C7440597&Type=IsoBar&Digits=5&P=1&THigh=130&TLow=50&TInc=1&RefState= DEF&TUnit=K&PUnit=bar&DUnit=kg%2Fm3&HUnit=kJ%2Fkg&WUnit=m%2Fs&Vis Unit=uPa*s&STUnit=N%2Fm. Accessed 22 Apr 2022

Study on the Increased Enthalpy of Loss Product on Heat Leak into Cryogenic Vessels Zheng-qing Li , Sheng-sheng Yang, Xian-hu Han, Yu-hong Cai, Xiao-jin Li, Xiao-xia Li, Min Ma, and Xiao-jun Wang(B) Lanzhou Institute of Physics, Lanzhou 730000, China [email protected]

Abstract. The thermal insulation performance is a very important parameter for cryogenic vessel reflected by heat leak and evaluated by loss product. The standard stipulates that the heat leak in a cryogenic vessel is the latent heat of loss product considered the loss product is saturated as it flows out of the inner vessel. In fact, the gas in cryogenic vessel is unsaturated and its temperature is higher than the saturation temperature during the loss product test of a cryogenic vessel. Therefore a part of the heat leak is absorbed by the loss product because its temperature increases as flowing through the gas of cryogenic vessel. In response to this problem, the heat transfer process was further studied and analyzed during the loss product test, and the energy equations of the gas and liquid were established to research the heat leak in cryogenic vessels, especially for the increased enthalpy of the loss product. The experiment was completed to verify that the heat leak in a cryogenic vessel includes the increased enthalpy of the loss product besides the latent heat in the test. The results indicate that the increased enthalpy is critical for the heat leak in cryogenic vessels during the loss product test, and its ratio is over 26% with respect to the latent heat as the liquid level ratio is less than 85%. Keywords: Enthalpy · Latent heat · Temperature · Loss product · Cryogenic vessel

1 Introduction With the rapid development of the industrial, an increasing number of companies use cryogenic vessels to store and transport various gases because the cryogenic vessel can store more gas in the same volume and its working pressure is less than one tenth of high pressure gas cylinder. Therefore, the cryogenic storage of liquid is efficient and safe [1–3]. The loss product of cryogenic vessels must be regularly tested to evaluate whether the thermal insulation performance meets the requirements of the standard [4, 5]. During the test, if the performance meets the requirements, the cryogenic vessel can continue to be used safely; otherwise, it needs maintenance or warranty in time. To shorten the testing time and save the working medium, the heat leak of cryogenic vessels is analyzed by the convection heat transfer method to create a semi-experiment and semianalytical method that the thermal insulation performance can be evaluated based on the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 515–521, 2023. https://doi.org/10.1007/978-981-99-6128-3_66

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test data at low liquid level, which indicates the heat leak decreases with a decrease in liquid level [6]. In storage and transport process of cryogenic liquids, as all valves of cryogenic vessel are closed, the pressure of cryogenic vessel will continue to increase over time under influence of heat leak that is self-pressurization. Therefore, the selfpressurization process is studied at different liquid levels and heat leak condition to investigate the increased rate of pressure in tested cryogenic vessel [7, 8]. To reduce the heat leak into cryogenic vessel, the thermal insulation materials and structures of cryogenic vessel have been studied to establish the corresponding heat leak calculation models, and their thermal insulation performance are tested under different conditions to select the optimal structure and support the design [9]. Experiments and software were employed to simulate the heat leak at different liquid levels. It was found that the heat leak into cryogenic vessels decreases as the liquid level decreases [10, 11]. When CFD is applied to analyze the process of heat exchange in the cryogenic vessel with and experiments, the studies indicated that the gas is unsaturated and its temperature increases with the height from gas to gas-liquid interface [12]. The liquid temperature near the gas-liquid interface and the inner wall is higher than at any other point, which causes temperature stratification in the liquid because the temperature difference is large between liquid and environment. This means that thermal stratification is more serious if the cryogenic liquid temperature is lower [13]. In these studies, the heat leak of cryogenic vessels was considered to be the latent heat of the loss product, which implied that the gas is saturated and the loss product does not absorb any heat leak as it flows from the gas-liquid interface to the outside of the inner vessel. However, experiments demonstrate that the gas is not saturated and its temperature is higher than the saturation temperature. Therefore, an increased enthalpy is produced as the loss product is not saturated and absorbs heat as it flows from the gasliquid interface to the outside. In this paper, the heat transfer process was further studied and analyzed in the loss product test, and the energy equations of the gas and liquid were established to study the heat leak of cryogenic vessels, proving that the heat leak includes the increased enthalpy and the latent heat of loss produced. The experiment was completed to verify the analysis, and the results indicate that the increased enthalpy is a critical part of the heat leak in cryogenic vessels during the loss product test, especially at low liquid levels.

2 Physics Model Currently, the heat leak of cryogenic vessels is calculated by loss product to evaluate the thermal insulation performance in standards, such as ISO [4]. This test method requires that the pressure of the cryogenic vessel matches the atmospheric pressure firstly. Therefore, the gas valve is fully opened and a large amount of gas is discharged before the test. Then the test system requires 48 h to balance the pressure and stabilize the temperature. Finally, the loss product of the cryogenic vessel is tested over 24 h. The test model is shown in Fig. 1. To perform the analysis discussed in the next section, the following hypotheses are necessary: (1) The liquid is saturated and its temperature is constant because the atmospheric pressure is considered to be constant.

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(2) The heat leak of cryogenic vessels only comes from the environment. (3) The liquid is incompressible.

Fig. 1. Experimental equipment.

2.1 Heat Transfer Analysis As shown in Fig. 1, the liquid and gas in cryogenic vessels are the research objectives. They absorb the heat leak from the environment. During the loss product test, the total mass consists of liquid, gas and loss product; therefore, the mass conservation equation in a cryogenic vessel is expressed as follows: dmlp dmg dml + + =0 dt dt dt

(1)

where ml , mg , and mlp are the masses of liquid, gas, and loss product during the test, respectively. For energy conservation of the test system, the energy equations for liquid and gas are expressed as follows: dmlp d (ρg Vg eg + ρl Vl el ) = QW − hlp dt dt

(2)

where ρ g is the gas density, ρ l is the liquid density, V g is the gas volume, V l is the liquid volume, eg is the special internal energy of the gas, el is the special internal energy of the liquid, QW is the total heat leak into the cryogenic vessel from the environment, and hlp is the special enthalpy of the loss product as it leaves the inner vessel. The first term in Eq. (2) can be expanded as    deg  d ρg Vg d d (ρl Vl ) del (ρg Vg eg + ρl Vl el ) = ρg Vg + eg + (ρl Vl ) + el (3) dt dt dt dt dt During the test, a few of liquid evaporates into the gas to form the loss product. Therefore, the liquid mass decreases, whereas the masses of the gas and loss product increase. These are expressed as follows: d (ρl Vl ) = −M dt

(4)

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d (ρg Vg ) + m ˙ lp = M dt

(5)

where M is the rate of evaporated liquid mass and m ˙ lp = dmlp /dt is the rate of loss product. According to Eq. (5), the evaporated liquid can be divided into two parts: the loss product and the part of increased gas occupying the volume of the evaporated liquid. Therefore, their relationship can be expressed as follows M =m ˙ lp + ρg

M ρl

(6)

The internal energy can be expressed with special heat capacity and temperature as e = cv T

(7)

where cv is the special heat capacity and T is the temperature. Combining Eqs. (4)–(7), Eq. (3) can be rewritten as follows     dTg d dTl (ρg Vg eg + ρl Vl el ) = ρg Vg cvg + ρl Vl cvl + M eg − el − m ˙ lp hlp (8) dt dt dt Meanwhile, the enthalpy can be expressed as h=e+

p ρ

Therefore, based on Eqs. (7)–(9), Eq. (2) is rewritten as follows   dTg dTl ρl ρg Vg cvg + ρl Vl cvl = QW − m ˙ lp hlp − hv + L dt dt ρl − ρg

(9)

(10)

where cvg is the gas specific heat capacity, cvl is the liquid specific heat capacity, T g is the temperature of the gas in the cryogenic vessel, T l is the temperature of the liquid in the cryogenic vessel, hlg is the specific enthalpy of loss product as it leaves the inner vessel, hv is the specific enthalpy of the saturated gas, L is the latent heat, expressed as L = hv – hl , and hl is the specific enthalpy of the saturated liquid. In test process of the loss product, the test system is stable in temperature, and the average temperature of the gas can be considered constant, that is, dT g / dt = 0, while, as a hypothesis, the liquid is saturated and its temperature is also constant, that is, dT l / dt = 0. Based on these considerations, Eq. (10) can be simplified as follows:   ρl QW = m ˙ lp hlp − hv + L (11) ρl − ρg Here, in Eq. (11), the heat leak of the cryogenic vessel can be divided into two parts: latent heat and increased enthalpy of loss product, which are respectively expressed as follows:   ρl QL = ρlp V˙ lp L (12) ρl − ρg

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  QE = ρlp V˙ lp hlp − hv

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(13)

where ρ lp is the density of the loss product, V˙lp is the volume rate of the loss product that can be tested by a mass flow meter, QL is the latent heat of loss product, which is considered to be the total heat leak in a cryogenic vessel according to the standard method, and QE is the increased enthalpy of loss product as it flows from the interface to the outside of the inner vessel, which is ignored in standard method. 2.2 Increased Enthalpy In a loss product test, as the pressure of cryogenic vessel equals to the atmospheric pressure and can be considered constant, the saturated enthalpy of gas hv is constant. As described in Eq. (13), the increased enthalpy is determined by the loss product rate and its temperature when leaving the inner vessel, being hlp =cpg .Tlp . The loss product temperature is considered to be equal to the top gas temperature in the cryogenic vessel as it leaves the inner vessel. If the liquid level is high in a cryogenic vessel, a considerable amount of liquid will be present in cryogenic vessel. Therefore, the respective temperatures of the gas in a cryogenic vessel and loss product will be low during the test process. Otherwise, the respective temperatures of gas in a cryogenic vessel and loss product will be high at a low liquid level during the test process. Meanwhile, the temperature of loss product is not saturated when it leaves the inner vessel Therefore, the increased enthalpy is produced as the loss product flows from the interface to the outside of the inner vessel during the loss product test.

3 Experiment The experimental equipment was built to measure the loss product as Fig. 1. The medium was liquid nitrogen; the volume of cryogenic vessel was 180 L; the material of the inner and outer wall was SA 304; the inner vessel was 450 mm in diameter and 950 mm in length; the elliptical head was130 mm in height. There were multi-layers insulation materials in the vacuum interlayer. The test of loss product was finished as the ISO standard at different liquid levels to investigate the increased enthalpy [11]. The loss products and their temperatures were tested at three different liquid levels. The results are presented in Table 1 and Fig. 2. Table 1. Loss product of tested cryogenic vessel. Le (%) T E (K) V˙lp (L/min)

51.4

70.2

86.1

297.48

295.85

294.68

0.88

0.927

0.967

In Table 1, Le is the liquid level, T E is the environment temperature. As shown in Table 1, the volume rate of loss product increases with an increase in the liquid level.

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Fig. 2. Temperature distribution of gas in tested cryogenic vessel.

This means that the latent heat of loss product QL increases with an increase in the liquid level. In Fig. 2, it indicates the gas temperature increases with a decrease in liquid level ratio. Meanwhile, the temperature of loss product increases with a decrease in liquid level ratio as it leaves the inner vessel, which means that the increased enthalpy also increases with a decrease in liquid level ratio. Based on Table 1 and Fig. 2, the latent heat QL , the increased enthalpy QE , and total heat leak QW are calculated by Eqs. (11)–(13). The results are shown in Table 2. Table 2. Latent heat, increased enthalpy and total heat leak. Le (%)

T lp (K)

hlp (kJ/kg)

QL (W)

QE (W)

QW (W)

86.1

126.08

129.5

4.02

1.05

5.07

70.2

137.21

141.3

3.85

1.24

5.09

51.4

146.09

150.5

3.66

1.35

5.01

As Table 2 shows, the total heat leak QW was noticeably greater than the latent heat of the loss product QL because some of the heat leak was absorbed by the increased enthalpy, as expressed by Eq. (11). However, this part of the heat leak is ignored in the standard method. For example, at a 51.4% liquid level ratio, the total heat leak was 5.01 W, where the latent heat was 3.66 W and the increased enthalpy was 1.35 W. These mean the heat leak is 3.66 W in standard method, but the increased enthalpy of 1.34 W is ignored and the actual heat leak is 5.01 W. The ratio of increased enthalpy to total heat leak was 26.8%. This implies that the latent heat of loss product cannot reflect the actual thermal insulation performance of the tested cryogenic vessel in the standard method.

4 Conclusion Based on the above analysis, experiments, and calculations of the heat leak in cryogenic vessels during the loss product test, the following conclusions can be drawn:

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(1) During the loss product test of a cryogenic vessel, the total heat leak absorbed by the loss product can be divided into two parts: the latent heat and the increased enthalpy. (2) The increased enthalpy of loss product is produced as it flows from the interface of gas and liquid to the outlet of the inner vessel (3) The increased enthalpy of the loss product decreases with an increase in the liquid level because the temperature of the loss product decreases with an increase in the liquid level as the temperature of the gas decreases with an increase in the liquid level.

References 1. Barsi, S., Kassemi, M.: Investigation of tank pressurization and pressure control—part II: numerical modeling. J. Therm. Sci. Eng. Appl. 5, 2–4 (2013) 2. Lisowski, E., Lisowski, F.: Study on thermal insulation of liquefied natural gas cryogenic road tanker. Therm. Sci. 23(Suppl. 4), 1381–1391 (2019) 3. Chen, Q.S., Wegrzyn, J., Prasad, V.: Analysis of temperature and pressure changes in liquefied natural gas (LNG) cryogenic tanks. Cryogenics 44(10), 701–709 (2004) 4. BS ISO 21014, Cryogenic vessels-Cryogenic insulation performance, BSI publications, London, Unite Kingdom (2006) 5. Xue, J.A., Zhou, W.M., Wang, R.S., et al.: GBT 18443.5–2010 Testing method of performance for Vacuum and Insulation Cryogenic Equipment, Part 5: Static Evaporation Rate Measurement, Beijing, China (2010) 6. Zhang, J.F., Wei, G.M., Qu, Z.G.: A time-saving method to evaluate the thermal insulation performance of cryogenic vessels. J. Clean. Prod. 258(9), 1–9 (2020) 7. Majumdar, A., Valenzuela, J., LeClair, A., Moder, J.: Numerical modeling of selfpressurization and pressure control by a thermodynamic vent system in a cryogenic tank. Cryogenics 74, 113–122 (2016) 8. Ludwig, C., Dreyer, M.E., Hopfinger, E.J.: Pressure variations in a cryogenic liquid storage tank subjected to periodic excitations. Int. J. Heat Mass Transf. 66, 223–234 (2013) 9. Fesmire, J.E., Johnsonb, W.L.: Cylindrical cryogenic calorimeter testing of six types of multilayer insulation system. Cryogenics 89, 58–75 (2018) 10. Xu, Y., Fan, X.: Analysis of effect of liquid level on heat loss into cryogenic vessels. Cryogenics 208, 26–30 (2015) 11. Yang, L., Wang, R., Wang, C.: Study on effect of liquid on the heat leak into vertical cryogenic vessels. Cryogenics 50, 367–372 (2010) 12. Li, Y., Wang, R.S.: Numerical simulation and experimental analysis of heat transfer through neck tube into vertical cryogenic vessels. Cryogenics 198, 43–48 (2014) 13. Cheng, X., Li, Y.: Characteristics analysis of cryogenic thermal stratification. Cryogenics 183, 32–36 (2011)

Coupling Performance Analysis of Plate-Fin Heat Exchanger Filled with Catalyst for Hydrogen Liquefaction Pan Xu1 , Aimin Zhou2 , Jian Wen1(B) , Simin Wang3 , and Yanzhong Li1 1 School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an 710049, Shaanxi,

China [email protected] 2 Wuhan 2nd Ship Design and Research Institute, Wuhan 430205, Hubei, China 3 School of Chemical Engineering and Technology, Xi’an Jiaotong University, Xi’an 710049, Shaanxi, China

Abstract. In an effort to promote the design and application of the plate-fin heat exchanger filled with catalyst (PFHEFC) in industrial production, a coupled mathematical model of continuous hydrogen ortho-para catalytic conversion (CHOPCC) is established to analyze its coupling performance. The results show that the j of the catalyst filled channel (CFC) of the PFHEFC is higher 3.3 times than that of the traditional fin channel (TFC), and the thermal enhancement factor (TEF) of the CFC reaches 37.7% of that of the TFC. The outlet para-hydrogen proportion yout p-H2 is approximately positively correlated with mass space velocity vm , and vm ≤ 0.66 kg/(m3 ·s) is the condition for meeting the demand of the PFHEFC at 42– 70 K. Therefore, it shows that the PFHEFC can realize the CHOPCC under the condition of ensuring a certain thermal-hydraulic performance. Meanwhile, for the PFHEFC, the j and TEF of the CFC are respectively 8–10 times and 68% –93% of the TFC under different fin structures. And a large cross-section of the CFC has better coupling performance. Above all, the research results demonstrate the feasibility of the PFHEFC and can facilitate its design. Keywords: Hydrogen liquefaction · CHOPCC · PFHEFC · Coupling performance

1 Introduction Hydrogen energy is a highly concerned energy used for future industrial systems, with liquid hydrogen being the most ideal form of its utilization [1, 2]. However, the high energy consumption and small production capacity of existing hydrogen liquefaction devices limit the promotion and utilization of liquid hydrogen [3]. Therefore, the integrated technology of hydrogen ortho-para converter and low-temperature heat exchanger is one of the optimization schemes mentioned in the many conceptual large-scale hydrogen liquefaction processes. The advantage of this technology is that it allows for the CHOPCC, thereby achieving the goal of optimizing the hydrogen liquefaction process. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 522–528, 2023. https://doi.org/10.1007/978-981-99-6128-3_67

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For this technology, Wilhelmsen et al. [4–6] conducted a performance analysis of the conceptual hydrogen liquefaction process, and a performance evaluation of the main core equipment (PFHEFC). And Donaubauer et al. [7] established also a 1-D model of the PFHEFC to analyze its performance. However, most of the researches on this technology remain in the conceptual analysis stage, lacking the coupling mechanism and coupling performance analysis of the PFHEFC. Therefore, this paper established a coupled mathematical model of the CHOPCC of the PFHEFC to analyze its coupling performance and facilitate its application.

2 Model Development 2.1 Physical Model As shown in Fig. 1, the adopted physical model of the PFHEFC is a two-streams heat transfer structure. Both sides are plain with the same structure, as shown in Table 1. The two sides are respectively normal hydrogen (p = 0.8 MPa) and reaction hydrogen (p = 2 MPa).

Fig. 1. Physical model of the PFHEC

Table 1. Structure parameters of the plate and fin Type

h/mm

s/mm

δ t /mm

δ p /mm

l/m

Value

6.5

2.1

0.3

1.0

2.5

2.2 Mathematical Model As shown in Fig. 2, the hydrogen ortho-para conversion is essentially the spin isomerization reaction of hydrogen molecules caused by temperature changes. The CHOPCC is a complex physical process that involves flow, heat transfer and reaction. In addition to the three basic equations, the core of the mathematical model lies in the mathematical

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description of the hydrogen ortho-para catalytic conversion. Three dynamic models of the hydrogen ortho-para catalytic conversion are proposed in Ref. [4–7] and used in the mathematical model of this paper, as shown in the Table 2. And the packing data of the CFC is shown in Table 3. The mass flow ratio of normal hydrogen and reaction hydrogen are 2.03, and the operating conditions are calculated as shown in Fig. 3. Fluent® is selected as the solver of the numerical solution.

Fig. 2. Hydrogen ortho-para conversion

Table 2. Three dynamic models of the hydrogen ortho-para catalytic conversion Type

Mathematical description n    eq 1−yp−H 2 yp−H 2 r = K ln eq 1−y

Elovich model [4–6]

yp−H 2



p−H 2

eq

First-order model [7]

r = kfirst cH 2 yo−H 2 − yo−H 2

Langmuir-Hinshelwood model [7]

r=



 eq kcH 2 yo−H 2 −yo−H 2  1+k cH 2 yo−H 2

Table 3. Packing data of the CFC Catalyst

ε

d p /mm

ρ s /(kg/m3 )

λs /(W/(m·K))

cp /(J/(kg·K))

Fe2 O3

0.5

0. 59

5240

0.58

700

Fig. 3. Calculation model of the PFHEC

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2.3 Model Verification The verification of thermal-hydraulic behavior in plain is shown in Fig. 4, and the numerical results close to the experimental data [8, 9]. Meanwhile, three dynamic models of the hydrogen ortho-para catalytic conversion are respectively used for calculation and compared with the experimental data [10]. As show in Fig. 5, the Elovich model is the most suitable model with the average relative error of 1.8% with the experimental data.

Fig. 4. Verification of thermal-hydraulic behavior in plain

Fig. 5. Verifications of three dynamic models of the hydrogen ortho-para catalytic conversion

3 Results and Discussion 3.1 Effect of Catalyst The outlet para-hydrogen proportion is 0.8127, which is greater than 95% of the outlet equilibrium para-hydrogen proportion (0.8010). It shows the PFHEC can meet the demand of the CHOPCC. The temperature distribution in the flow direction of the PFHEC and PFHE is respectively shown in Fig. 6. The average temperature differences of the cold and hot sides of the PFHEC and PFHE are respectively 1.06 K and 0.29 K, 0.80 K and 0.83 K. Therefore, the catalyst significantly enhances the heat transfer of the CFC of the PFHEC, resulting in a more uniform temperature cross-section within the CFC. Meanwhile, the j, f and TEF of the CFC of the PFHEC and the TFC of the plate-fin

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heat exchanger (PFHE) are respectively 0.0228, 18.87 and 0.0086, 0.0053, 0.0128 and 0.0227. It shows that the j of the CFC of the PFHEFC is higher 3.3 times than that of the TFC, and the TEF of the CFC reaches 37.7% of that of the TFC. Therefore, the PFHEFC can realize the CHOPCC under the condition of ensuring a certain thermal-hydraulic performance.

(a) PFHEC

(b) PFHE

Fig. 6. Temperature distribution in flow direction of the PFHEC and PFHE

3.2 Effects of Fin Structure Parameters As shown in Fig. 7, the TEF of two sides is negatively correlated with Rehot , l and δ t , and positively correlated with h and s. Meanwhile, for the PFHEFC, the j and TEF of the CFC are respectively 8–10 times and 68%–93% of the TFC under different fin structures.

Fig. 7. Variations of coupling performance with fin structure parameters

The catalyst greatly enhances the heat transfer performance of the CFC and is also the dominant factor of the flow performance of the CFC at the same. Therefore, the thermal-hydraulic performance with Rehot of the CFC tend to be the same under different

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structural parameters. At the same time, a large cross-section of the CFC has better coupling performance. 3.3 Conversion Performance In order to intuitively reflect the normal to secondary conversion performance of the PFHEC, it is found through analyzing numerical data that the outlet para-hydrogen proportion yout p-H2 under different fin structure parameters is approximately positively correlated with mass space velocity vm . And the positive correlation is approximately linear: out yp−H 2 = −0.0693vm + 0.8467

(1)

The fitting degree and average relative error between the linear correlation and numerical results are respectively 0.9874 and 0.27%. Meanwhile, in order to meet the demand of the CHOPCC at 42–70 K, the linear correlation shows the mass space velocity vm should be lower than 0.66 kg/(m3 ·s).

4 Conclusion (1) The Elovich model is the most suitable dynamic model of the CHOPCC with the average relative error of 1.8% with the experimental data [10]. (2) Due to the effect of catalyst, the j of the CFC of the PFHEFC is higher 3.3 times than that of the TFC, and the TEF of the CFC reaches 37.7% of that of the TFC. (3) For the PFHEFC, the j and TEF of the CFC are respectively 8–10 times and 68%– 93% of the TFC under different fin structures. And a large cross-section of the CFC has better coupling performance. (4) The outlet para-hydrogen proportion yout p-H2 is approximately positively correlated with mass space velocity vm , and vm ≤ 0.66 kg/(m3 ·s) is the condition for meeting the demand of the PFHEFC at 42–70 K. Acknowledgments. Thanks for the supports of the National Natural Science Foundation of China (No. 51676146), the Fund Project of State Key Laboratory of Space Cryogenic Propellant Technology of China (No. SKLTSCP 1909) and the Fundamental Research Funds for the Central Universities of China (No. Xzy022022036).

References 1. Chang, H.M., Kim, B.H., Choi, B.: Hydrogen liquefaction process with Brayton refrigeration cycle to utilize the cold energy of LNG. Cryogenics 108, 103093 (2020) 2. Ozcan, H., Dincer, I.: Thermodynamic modeling of a nuclear energy based integrated system for hydrogen production and liquefaction. Comput. Chem. Eng. 90, 234–246 (2016) 3. Krasae-In, S., Stang, J.H., Neksa, P.: Development of large-scale hydrogen liquefaction processes from 1898 to 2009. Int. J. Hydrogen Energy 35(10), 4524–4533 (2010)

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4. Skaugen, G., Berstad, D., Wilhelmsen, O.: Comparing exergy losses and evaluating the potential of catalyst-filled plate-fin and spiral-wound heat exchangers in a large-scale Claude hydrogen liquefaction process. Int. J. Hydrogen Energy 45(11), 6663–6679 (2020) 5. Wilhelmsen, O., Berstad, D., Aasen, A., et al.: Reducing the exergy destruction in the cryogenic heat exchangers of hydrogen liquefaction processes. Int. J. Hydrogen Energy 43(10), 5033–5047 (2018) 6. Hande, R., Wilhelmsen, O.: Minimum entropy generation in a heat exchanger in the cryogenic part of the hydrogen liquefaction process: on the validity of equipartition and disappearance of the highway. Int. J. Hydrogen Energy 44(29), 15045–15055 (2019) 7. Donaubauer, P.J., Cardella, U., Decker, L., et al.: Kinetics and heat exchanger design for catalytic ortho-para hydrogen conversion during liquefaction. Chem. Eng. Technol. 42(3), 669–679 (2019) 8. Zhang, Z.Y., Shi, B.S.: Theory and Device of Cryogenic Technology. Mechanical Engineering Press, Beijing (1987) 9. Kays, W.M., London, A.L.: Compact Heat Exchanger. McGraw-Hill, New York (1964) 10. Hutchinson, H.L.: Analysis of catalytic ortho-parahydrogen reaction mechanisms. University of Colorado (1966)

Study on the Impact of Cold Storage Plate Layout on the Thermal Insulation Performance of VIP Incubator Lin Yifan, Wang Jiali, Kan Ankang, and Chen Wu(B) Merchant Marine College, Shanghai Maritime University, Shanghai 201306, China [email protected]

Abstract. As a kind of efficient cold-chain equipment, the cold storage plate is usually used to maintain the low temperature environment required in the incubator using vacuum insulation panels (VIP). In order to ensure the quality of refrigerated goods, it is necessary to have an appropriate temperature field and holding time in the incubator. In this paper, under the conditions of different ambient temperature, the impact of different layout of cold storage plate on thermal insulation performance of VIP incubator was studied experimentally. By changing the layout mode of 0 °C cold storage plate (bottom, side, parallel, top and bottom mode), the temperature at three different positions in an incubator (box size 500 mm × 390 mm × 350 mm) was tested to investigate the distribution of temperature field and the holding time. It was found from experiments that when the ambient temperature was 33 °C, the holding time at the bottom mode was the longest, up to 31.9 h, but the heat leakage in the top area of the incubator was serious so that it cannot meet the preservation requirements of 2–8 °C. Moreover, the temperature distribution in the incubator was the most uniform when the top and bottom mode was adopted, and the holding time can reach 21.2 h. When the ambient temperature was relatively low, i.e., 22 °C, the holding time at side mode is the longest up to 62.3 h, and the temperature field can meet the requirements; while at the bottom mode, the distribution of temperature field was the most uniform and the holding time can reach 50.1 h. Considering that the holding time can meet the cold-chain transportation requirements in different layouts of the cold storage plate, therefore, when the ambient temperature is high, the top and bottom layout mode for the cold storage plate is recommended, and when the ambient temperature is relatively low such as during the transition season, the bottom mode is recommended for the VIP incubator. Keywords: incubator · cold storage plate · thermal insulation performance · layout mode · temperature field

1 Introduction In recent years, due to the development of China’s economy and the improvement of people’s living standards, the requirements of the cold chain logistics industry for thermal insulation and refrigeration have promoted the development of cold storage technology © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 529–536, 2023. https://doi.org/10.1007/978-981-99-6128-3_68

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[1]. Cold chain logistics requires that products are always in the low-temperature environment required by the products from production, processing, storage and transportation, sales to pre-consumption, so as to ensure that the loss of product quality is minimized. Cold storage transport incubators [2] are attracting more and more researchers’ attention due to their simple equipment, low cost, precise temperature control, and the coexistence of goods of multiple temperatures. Especially in the current epidemic situation, under the trend of sustainable development of drug logistics [4], the use of incubators will become more and more extensive, and the market prospect is good. In the process of cold chain transportation, ensuring the quality of products requires a suitable temperature field inside the incubator. Therefore, in order to improve product quality during transportation, it is not only necessary to pursue the low temperature environment in the incubator, but also consider the temperature field inside the box uniformity. In fact, the temperature distribution inside the incubator is very uneven, but a stable and uniform temperature field is very important for the transportation of products with high temperature requirements [5]. Therefore, many scholars have conducted research on the influencing factors of the temperature field distribution of the incubator, such as the ambient temperature [6], the structural size of the incubator [7], and the thermal insulation material [8], etc. An important factor affecting the temperature field distribution of the incubator is the placement of the refrigerant. The cold storage agent can absorb and store a large amount of cold energy at low temperature, and release a large amount of cold energy at high temperature. It can maintain its own and surrounding low-temperature environment for a long time. It is suitable for application in incubator [9]. However, at present, there is no uniform specification for the placement of the refrigerant in the incubator, so the nonstandard placement of the refrigerant is likely to lead to uneven temperature distribution in the incubator, making some areas unable to meet the insulation requirements [10]. In this experiment, the typical medical incubator in the process of cold chain transportation is taken as the research object, and its suitable refrigeration temperature is 2–8 °C [11]. Based on different ambient temperatures, by changing the placement of the cold storage plate, explore its impact on the temperature field of the incubator.

2 Cold Insulation Experiment At present, vacuum insulation plate (VIP) has been applied to the insulation of goods in cold chain logistics [12]. This experiment used a medical incubator of Pfizer company. The insulation material is VIP. The outer incubator size is 500 × 390 × 350 mm, the inner box size is 400 × 280 × 250 mm, as shown in Fig. 1. The size of the cold storage plate is 245 × 155 × 200 mm. And an Agilent 34972A data collector is needed to collect the experimental data. The specific experimental process is as follows: 1) Set the temperature recording time interval to 20 s. 2) Put two pre-cooled cold storage plates with phase change temperature of 0 °C into the insulation box, and the arrangement is shown in Fig. 2. 3) Fix the temperature sensor at the upper, middle and lower test points on the inner wall of the incubator. Keep the incubator sealed during the experiment.

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Fig. 1. Schematic diagram of Pfizer medical incubator

4) Record the temperature change in the insulation box. When the temperature exceeds the upper limit of the insulation temperature, the experiment ends.

Fig. 2. Schematic diagram of the arrangement of the cold storage plate

3 Experimental Results and Analysis 3.1 Experimental Results When the ambient temperature is about 33 °C, the temperature change curves of the four placement methods and the data of each measuring point are shown in Fig. 3 and Tables 1, 2, 3 and 4: It can be seen that the heat preservation time on the side is the shortest, the heat preservation time on the bottom is the longest and the temperature rise speed is the slowest, and the temperature rise speed on the top and bottom is the fastest. When the ambient temperature is about 22 °C, the temperature change curves of the four placement methods and the data of each measuring point are shown in Fig. 4 and Tables 5, 6, 7 and 8. It can be seen that when the ambient temperature is reduced, the temperature rise speed is greatly reduced, the insulation time is increased. 3.2 Experimental Analysis When the ambient temperature is 33 °C, the temperature of some areas has exceeded the refrigeration temperature range. When placed at the bottom, the average speed of temperature rise is the smallest and the insulation time is the longest, but the temperature in the upper part of the box does not meet the actual requirements. When placed at the top and bottom, the temperature field inside the box is the most uniform, all points can

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Fig. 3. Temperature field change curves of four placement methods at 33 °C

Table 1. Data of each measuring point (33 °C, side) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

24.14

25.8

0.63

Central measuring point

26.88

24.0

0.79

Upper measuring point

29.58

26.3

0.56

Table 2. Data of each measuring point (33 °C, bottom) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

31.81

33.3

0.72

Central measuring point

31.93

31.9

0.75

Upper measuring point

31.99

35.2

0.48

meet the required standards and the temperature difference at each measuring point is the smallest. When the ambient temperature is 22 °C, the temperature rise speed of the side is small, and the insulation aging is the longest. When the bottom is placed, the layered distribution of the temperature field is the most uniform, and the temperature difference

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Table 3. Data of each measuring point (33 °C, top-bottom) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

32.82

32.6

0.76

Central measuring point

32.95

32.8

0.76

Upper measuring point

33.08

21.2

1.18

Table 4. Data of each measuring point (33 °C, parallel) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

33.44

31.9

0.80

Central measuring point

33.47

31.1

0.82

Upper measuring point

33.57

31.9

0.58

Fig. 4. Temperature field change curves of four placement methods at 22 °C

between the measuring points in the vertical direction is the smallest. It can be found that the temperature field inside the incubator is very different under different ambient temperatures.

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Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

21.12

62.3

0.21

Central measuring point

20.76

63.5

0.20

Upper measuring point

17.48

63.5

0.15

Table 6. Data of each measuring point (22 °C, bottom) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

24.64

50.1

0.33

Central measuring point

24.10

51.6

0.31

Upper measuring point

23.30

51.2

0.30

Table 7. Data of each measuring point (22 °C, top-bottom) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

22.50

49.8

0.29

Central measuring point

21.88

51.0

0.27

Upper measuring point

18.60

51.2

0.21

Table 8. Data of each measuring point (22 °C, parallel) Initial temperature (°C)

Holding time (h)

Temperature rise rate (°C/h)

Bottom measuring point

19.60

56.4

0.21

Central measuring point

19.44

57.7

0.20

Upper measuring point

19.41

58.3

0.20

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4 Conclusion Based on the analysis of the above experimental data, the following three conclusions can be drawn: 1) It can be found that the temperature field inside the incubator is very different under different ambient temperatures. When the ambient temperature changes from 33 °C to 22 °C, the top heat leakage is significantly improved, the overall insulation effect is improved by 20 h, and the temperature field distribution uniformity is improved by 2.5 °C. Therefore, in the actual application process, the appropriate ambient temperature can ensure the uniformity of temperature field distribution. 2) When the external temperature is 33 °C, the temperature of the area closest to the refrigerant is lower. The storage agent placed at the bottom has the longest insulation time, which is more than 30 h, but the temperature of the upper part of the box does not meet the actual requirements; The uniformity of temperature field placed at the top and bottom is the best, each point can reach the required standard and the temperature difference of each measuring point is the smallest. If the user does not have high requirements for the uniformity of temperature distribution, the bottom placement method can be used. If the user has high requirements for the uniformity of temperature distribution, the top and bottom placement method can be used. 3) When the external environment is 22 °C, the storage agent placed on the side has the longest heat preservation time; The uniformity of the temperature field is the best when the bottom is placed, and its insulation duration can meet most of the cold chain transportation needs. Because the heat leakage at the top is significantly improved, the top and bottom placement mode does not show advantages at this ambient temperature. Users can choose side placement when seeking maximum insulation duration, and bottom placement when seeking uniformity of temperature distribution.

References 1. Wang, L., Lu, W., Zhai, S.: Simulation and experiment of phase change cold storage incubator. Model. Simul. 10(1), 168–177 (2021) 2. Chen, H., Zhang, J.: Application of cold storage transport incubator in cold chain. Refrigeration Technol. 2010(3), 12–16 (2010) 3. Pambudi, N.A., Sarifudin, A., Gandidi, I.M.: Vaccine cold chain management and cold storage technology to address the challenges of vaccination programs. Energy Rep. 8, 955–972 (2022) 4. Pan, X., Wang, J., Wang, D.: Influence of coolant placement on temperature field in incubator. Packag. Eng. 39(3), 77–82 (2018) 5. Zhu, H., Wang, D., Pan, X.: Research on thermal insulation performance of incubator under different ambient temperature. Packag. Food Mach. 36(4) (2018) 6. Yu, Y., Pan, Z., Lu, L.: Effects of different structural dimensions on the thermal insulation performance of EPP incubators. Struct. Des. Manuf. 39(9), 114–118 (2018) 7. Pan, X., Wang, D., Zhu, H.: Influence of thermal insulation materials on the temperature field in the incubator. Storage Transp. Preserv. 34(8), 115–118 (2018) 8. Sun, J., Xie, J.: Research progress on phase change cold storage materials and their application in cold storage. Food Mach. 37(7), 227–232 (2021)

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9. Xia, C., Yang, Y.: Several issues that should be paid attention to when using incubators to transport refrigerated medicines. Logistics Technol. 42(3), 68–78 (2020) 10. Jiang, H., Zeng, T., Ding, Y.: Selection and analysis of coolant storage in incubators based on cold chain logistics. Packag. Eng. 42(7), 168–174 (2021) 11. Wang, X., Xie, J.: Research progress of cold storage incubator. Food Mach. 35(8), 232–236 (2019) 12. Verma, S., Singh, H.: Vacuum insulation in cold chain equipment: a review. Energy Procedia 161, 232–241 (2019)

Experimental Study on Condensation Characteristics of Volatile Organic Compounds Hao Xu1,2 , Xiafan Xu1,2 , Liubiao Chen1,2,3(B) , Jia Guo1(B) , and Junjie Wang1,2 1 Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Chinese Academy of Sciences, Beijing, China {chenliubiao,guojia}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China 3 Institute of Optical Physics and Engineering Technology, Qilu Zhongke, Licheng District, Jinan, China

Abstract. The volatile organic compounds (VOCs) are important precursors of ozone and PM2.5. VOCs emission reduction plays an important role in carbon emission and environmental protection. For high-value VOCs such as organic solvents, the cryogenic condensation using liquid nitrogen is an effective VOCs recovery method that can satisfy the emission standard with a low refrigeration temperature of −180 °C or even lower. In this paper, in order to explore the VOCs condensation and frosting in the heat exchanger for improving the heat transfer efficiency, a set of visualization experimental equipment for VOCs condensation recovery based on liquid nitrogen cooling was designed, established, and tested. Several windows are set on the heat exchanger shell to observe the process of condensation and even frosting inside the heat exchanger and the distribution of liquid film and frost layer. Experiments and analyses were carried out under different working conditions such as the effect of VOCs concentration, gas flowrate, refrigeration temperature, and other influencing factors. Keywords: Cryogenic Condensation · Liquid Nitrogen · Visualization Experiment · VOCs

1 Introduction VOCs are important precursors for the formation of ozone [1] and PM2.5 [2]. VOCs control plays an important role in promoting carbon emission reduction and environmental protection. For VOCs with high value, such as organic solvents, oil gas, etc., under the condition of concentration higher than 10,000 ppm, the liquid nitrogen (LN2 ) condensation method is a suitable recovery technology [2], which can meet the increasingly stringent VOCs emission standards. This method generally uses LN2 on the tube side to cool the waste gas on the shell side in the shell-and-tube heat exchanger. VOCs are condensed into liquids or frozen into solids on the tube wall, thereby realizing the recovery of VOCs. The heat transfer process of VOCs waste gas plays a key role in the design of the cryogenic heat exchanger in the LN2 condensation recovery VOCs system. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 537–543, 2023. https://doi.org/10.1007/978-981-99-6128-3_69

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At present, some scholars have carried out experimental studies on the visualization of cryogenic condensation with nitrogen [3, 4]. For the condensation process of the mixed gas containing VOCs, the analysis of the condensation temperature is 5–10 °C [5], or investigate the condensation that occurs inside the tube [6] or vertical plate [7], while the cryogenic condensation of the VOCs mixed gas outside the tube is less studied. In this paper, a cryogenic condensation visualization experimental system was designed and developed to analyze the VOCs condensation process. Several windows are set on the heat exchanger shell to observe the condensation process of the waste gas to be treated on the tube wall in real-time. The formation and growth of liquid film or frost layer on the tube wall are observed and recorded. Further research on the LN2 condensation system recovering the exhaust gas of the “MTBE (Methyl tert-butyl ether) + methanol” was carried out. The influence of the gas flowrate and concentration of the exhaust gas and the refrigeration temperature on the system performance was analyzed.

2 Experimental System Figure 1 is a schematic of the developed cryogenic condensation visualization experimental system. It mainly includes a pre-cooling heat exchanger, cryogenic heat exchanger, blower, water bath device, LN2 storage tank, vaporizer, regulating valve, a data acquisition unit, etc. The type of heat exchanger is a shell-and-tube heat exchanger and multiple observation windows are set on the shell to observe the growth of liquid film or frost layer of VOCs gas on the vertical tube wall during the condensation process. The direct measurement data obtained by the visualization experiment are temperature, pressure difference and frost layer thickness. The accuracy of the thermometer is ±0.5 K, and the accuracy of the differential pressure sensor is ±50 Pa. The thickness of the frost layer is compared according to the scale, and the accuracy is ±0.03 mm.

Fig. 1. Schematic of the cryogenic condensation visualization experiment system

The main experiment process is as follows: the blower blows air into the water bath device to generate exhaust gas, which enters the pre-cooling heat exchanger and the cryogenic heat exchanger in turn. And purified gas is returned to the pre-cooling heat exchanger to recover its cold and then is discharged into the environment. The LN2 enters the cryogenic heat exchanger and the pre-cooling heat exchanger in turn from the LN2 storage tank through the regulating valve to control the flow for refrigeration, and the nitrogen discharged from the pre-cooling heat exchanger is rewarmed by the vaporizer. Figure 2 shows the actual experimental device.

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Fig. 2. Photo of the cryogenic condensation visualization experiment system

3 Results and Discussions 3.1 Frost Formation and Growth Figure 3 shows the formation and growth of the frost layer on the vertical tube wall in the observation window during the cryogenic condensation test using air with water vapor as the medium. The frost layer thickness is obtained by using the image software and a unified scale to compare the images in equal proportions after the images are taken by the camera. It can be seen that the vertical tube wall surface is frosted in a short time, the frost layer grows rapidly in the first 60 min, and then the growth trend of the frost layer slows down.

Fig. 3. Diagram of frost formation and growth phenomenon at the tube outer wall in the cryogenic heat exchanger (window A; t means time, minute)

Further referring to Fig. 4, it can be found that during the condensation process, the initial frost layer grows rapidly, and as the thickness increases, the growth of the frost layer begins to gradually slow down, and finally fills the entire tube wall gap. The main reason is that with the increase of the thickness of the frost layer, the thermal

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resistance of the tube wall surface increases due to the growth of the frost layer, so that the condensation heat transfer capacity is weakened, and the growth of the frost layer slows down.

Fig. 4. Diagram of frost thickness at the tube outer wall in the shell-and-tube heat exchanger

Figure 5 shows the pressure difference of the shell-side exhaust gas during the condensation process in the heat exchanger. It can be seen that the shell-side pressure difference of the pre-cooling heat exchanger remains unchanged, while that of the cryogenic heat exchanger gradually increases and the growth trend becomes faster.

Fig. 5. Diagram of pressure difference in the shell side of the heat exchanger

There are two main reasons. On the one hand, it was found that the phenomenon of frosting on different windows was not at the same time during the test. That is to say, the frost layer on the remaining tube bundle wall is still growing as shown in Fig. 6 when the frost layer on the tube bundle wall is full in the Fig. 3, resulting in a rapid increase in the shell side pressure difference. On the other hand, the condensation temperature of the exhaust gas is higher than 0 °C in the pre-cooling heat exchanger and the moisture gathers on the tube wall in the form of droplets or a stream and converges at the bottom of the

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heat exchanger so that it has little impact on the shell side pressure difference. However, the condensation temperature of the exhaust gas of the cryogenic heat exchanger is − 10 °C, the nitrogen temperature is −90 °C, and the tube wall temperature is lower than 0 °C so that the moisture crystallizes rapidly on the tube wall to form a frost layer and begins to grow, reducing the exhaust gas flow area and increasing the flow resistance.

t=187

t=371

Fig. 6. Diagram of frost formation and growth phenomenon at the tube outer wall in the cryogenic heat exchanger (window B)

3.2 Effect of the Inlet Gas Concentration and Refrigeration Temperature Based on the system shown in Fig. 1, another stage heat exchanger is added to realize the cascade utilization of cold. Figure 7 shows the effect of concentration and refrigeration temperature on system performance. It can be seen that the VOCs concentration in the purified gas decreases with the decrease of refrigeration temperature.

Fig. 7. Effect of the inlet gas concentration and refrigeration temperature on the system

The VOCs concentration is 3 ppm when the cooling temperature is −116 °C. The inlet concentration of the three working conditions is at least 10000 ppm and may even be 100000 ppm (the inlet exhaust gas VOCs concentration is obtained by the detection instrument in the petrochemical refinery located in Dongying City, Shandong Province. And the detection instrument shows over range and its maximum is 100000 ppm). The freezing points of MTBE and methanol are around −100 °C under normal pressure, which means that the inlet concentration has little effect on the system outlet concentration after the refrigeration temperature reaches −100 °C.

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3.3 Effect of the Gas Flowrate Figure 8 is the effect of the exhaust gas flowrate on the temperature distribution. It can be seen that the exhaust gas temperature at each stage of the heat exchanger remains basically unchanged with the fluctuating change of the gas flow rate. The main reason is that the system recovers the cold carried by the purified gas.

Fig. 8. Effect of the gas flowrate on the temperature distribution of the system

Figure 9 shows the effect of the gas flowrate on the system LN2 consumption. It can be seen that the system LN2 consumption decreases as the gas flowrate decreases when the refrigeration temperature and inlet VOCs concentration are constant. The recovered condensate decreases accordingly as the gas flowrate decreases with a constant concentration. Therefore, the recovered condensate will carry less cold and the system LN2 consumption will also decrease accordingly.

Fig. 9. Effect of the gas flowrate on the liquid nitrogen consumption of the system

4 Conclusion A cryogenic condensation visualization experiment system was designed and developed, and the cryogenic condensation experiment using air mixed with water vapor was completed. The results show that the formation of the frost layer grows rapidly at early stage,

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and the growth slows down later due to the effect of the frost layer thermal resistance. In addition, the growth of the frost layer causes the shell side pressure difference to increase in the heat exchanger and grows faster. The reason is that there is a time difference in the frost distribution inside the heat exchanger and there are more locations where frost occurs. The experiment of the liquid nitrogen condensation system recovering the exhaust gas of the “MTBE + methanol” was carried out and the influence of refrigeration temperature and flowrate on the system performance was analyzed. The results show that the system outlet VOCs concentration can be as low as 3 ppm when the refrigeration temperature reaches the processing temperature of the medium. The exhaust gas fluctuating flowrate does not affect the temperature distribution of the three-stage heat exchanger. The system LN2 consumption also decreases as the gas flowrate decreases when the gas concentration is constant. Acknowledgments. This research is supported by the Key Research Program of Frontier Sciences, Chinese Academy of Sciences (No. QYZDY-SSW-JSC028), Youth Innovation Promotion Association of the Chinese Academy of Sciences (No. 2019030), and the Foundation of the Director of the Technical Institute of Physics and Chemistry, Chinese Academy of Sciences (No. 210416).

References 1. Zhang, J.F., et. Al.: Ozone pollution: a major health hazard worldwide. Front. Immunol. 10 (2019) 2. Department of Atmospheric Environment, Ministry of Ecology and Environment: A Practical Handbook for Volatile Organic Compounds Control, 1st edn. China Environment Publishing Group, Beijing (2020) 3. Tang, Y., et al.: Experimental study on film condensation characteristics at liquid nitrogen temperatures. Appl. Therm. Eng. 127, 256–265 (2017) 4. Zhu, S.L., et al.: Characteristic analysis of fluctuating liquid film flow behavior and heat transfer in nitrogen condensation. Appl. Therm. Eng. 184, 116249 (2021) 5. Yang, G.C., et al.: Experimental investigation on heat transfer characteristics of two-phase propane flow condensation in shell side of helically baffled shell-and-tube condenser. Int. J. Refrig. 88, 58–66 (2018) 6. Zhuang, X.R., et al.: Experimental investigation on flow condensation heat transfer and pressure drop of R170 in a horizontal tube. Int. J. Refrig. 66, 105–120 (2018) 7. Zhang, L.L., et al.: Experimental and numerical study on filmwise condensation of pure propane and propane/methane mixture. Int. J. Heat Mass Transfer 156, 119744 (2020)

A Novel Segmented Non-Uniform Finned Channel of Supercritical LNG Applicable for Printed Circuit Heat Exchanger Qingfeng Jiang1(B) , Chongyao Pan1 , Shiqing Wan1 , Huaibing Li2 , Qiang Zhu2 , and Bao Fu3 1 School of Energy and Power, Jiangsu University of Science and Technology, Zhenjiang, China

[email protected]

2 Zhangjiagang Furui Special Equipment Co. Ltd., Zhangjiagang, China

{211210801133,209210010}@stu.just.edu.cm 3 Institute of Plasma Physics, Chinese Academy of Sciences, Hefei, China

Abstract. In order to enhance the heat transfer capacity of liquefied natural gas (LNG) in a printed circuit plate heat exchanger (PCHE). In this paper, the PCHE with non-uniform airfoil fin channels is proposed and numerically simulated as a heat transfer channel for supercritical LNG, considering the characteristics of the physical property changes of the heat transfer process of supercritical LNG. The heat transfer and flow characteristics of supercritical LNG in the microchannel are investigated. The results show that the Nusslet number does not fluctuate periodically along the flow direction due to the non-uniform airfoil fin arrangement. Compared with the uniform flow channel, the non-uniform flow channel produces 14.75% smaller for improved entropy generation number. Keywords: Supercritical LNG · Non-uniform · Numerical Simulation · Nusslet Number

1 Introduction The demand for liquefied natural gas (LNG) is increasing rapidly in the energy market due to its clean, environmental and economic advantages. The use of LNG requires its heating and vaporization to natural gas (NG) at supercritical pressure. The vaporization process generates a large number of heat exchange processes, so it is important to study more efficient heat exchangers for the utilization of natural gas. Floating storage and regasification units (FSRU) are an important component distribution of the LNG chain. Due to its offshore characteristics, it requires a relatively small amount of space to transfer a large amount of heat compared to a traditional land-based LNG gasification station. Printed circuit plate heat exchangers (PCHE) have provided extremely high heat transfer efficiency and high-pressure capability. The surface area per volume of an ordinary platefin heat exchanger is approximately 1000 m2 /m3 . Plate and shell heat exchanger is around 300 m2 /m3 . A general shell and tube heat exchanger is less than 100 m2 /m3 . In contrast, © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 544–550, 2023. https://doi.org/10.1007/978-981-99-6128-3_70

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Printed circuit plate heat exchanger can reach more than 2500 m2 /m3 . For the same heat capacity and pressure drop, the PCHE is 1/6 to 1/4 the size of a conventional shell and tube heat exchanger. Therefore, PCHE is particularly suitable as a heat exchanger for the LNG gasification process [1]. However, the heat transfer process of PCHE as a supercritical LNG gasifier involves micro-fluidic flow. Moreover, the characteristics of the supercritical LNG properties change drastically during the heat transfer process. Therefore, it is necessary to propose a better flow channel design form as LNG gasifier. Xu et al. [2] investigated the effect of fin arrangement in the airfoil PCHE channel on flow resistance and heat transfer generation. Simulation studies show that when supercritical CO2 is used as a working fluid, the first consideration is to reduce the flow resistance rather than to increase the heat transfer area. Jin et al. [3] also compared the effects of three fin structures, slit, wavy and airfoil, on the flow heat transfer in the S-CO2 pre-cooler by numerical simulation methods and found that the friction coefficient of the airfoil channel was lower. In this paper, a non-uniform flow channel structure is proposed for PCHE as an LNG gasifier, and the performance comparison with the conventional uniform flow channel is discussed by numerical simulation method.

2 Numerical Models and Methods 2.1 Physical Model This study considers the use of a staggered arrangement between fins in the LNG side channel. The physical geometry of the fins is shown in Fig. 1. The parameters of the airfoil include chord length b, maximum airfoil thickness t, angle of attack α, and curvature f 0 . f 0 represents the maximum distance between the central arced curve and the chord.

Fig. 1. Airfoil fins and LNG side channel geometry

Figure 2(a) shows the flow channel of uniformly arranged fins obtained by the conventional design method, and Fig. 2(b) shows the flow channel of non-uniformly arranged fins designed by the previously proposed design method [4]. The geometric parameters elated to the flow channels given by the two design methods involved in this study are shown in Tables 1 and 2. In order to make the two channel configurations comparable, the fin height was set to 0.00069 m. Also, in order to eliminate the effects of inlet effects and outlet backflow, the inlet and outlet of the heat transfer core were extended by 10% of the actual length. The inlet temperature of LNG was set to 128 K, and the working pressure was 7.3 MPa. The mass flow rate of LNG is 2.48 t/h. The surface area per

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volume of the non-uniform flow channel heat exchanger design is 3260 m2 /m3 and the uniform flow channel surface area per volume is 3760 m2 /m3 . The non-uniform flow channel heat load is 441 kW and the uniform flow channel heat load is 468 kW. The physical properties of LNG based on temperature and pressure were obtained from the NIST Standard Reference Database (REFPROP) version 9.0 to ensure their availability.

Fig. 2. Flow channel structure diagram

Table 1. Uniform flow channel’s parameters Parameter

bi

ti

αi

Px,i

Py,i

f 0,i

Li

Value

0.00549

0.00094

1.2

0.01219

0.00223

0.00041

0.17258

Table 2. Non-uniform flow channel’s parameters Parameter

bi

ti

αi

Px,i

Py,i

f 0,i

Li

First paragraph

0.00834

0.00083

13

0.01878

0.00574

0.00054

0.03653

Second paragraph

0.00433

0.00079

14

0.01907

0.00489

0.00059

0.05288

Third paragraph

0.00629

0.00088

7

0.01663

0.00337

0.00080

0.03962

Fourth paragraph

0.00642

0.00099

0

0.01604

0.00200

0.00042

0.01766

Fifth paragraph

0.00846

0.00099

0

0.01795

0.00201

0.00069

0.02008

Sixth paragraph

0.00408

0.00095

2

0.02687

0.00366

0.00041

0.00581

2.2 Grid Independence and Model Validation Numerical simulations are performed with the aid of ansys fluent software, and discrete equations in second-order upwind scheme are chosen to disperse the convection and

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diffusion terms in the control equations to improve the computational accuracy. The SIMPLE algorithm is also used to couple the pressure and velocity. By solving the control equations in fluid and solid domains, the heat transfer parameters of LNG flow in the flow channel under steady-state conditions are obtained. In order to ensure the accuracy of the calculation results, independent mesh verification of the numerical model as well as experimental verification are also required. The detailed parameters of the five sets of grids are given in Fig. 3(a). It can be seen that when the number of grids reaches 4.2 million, the effect of increasing the number of grids on the calculation results is almost negligible. Therefore, considering the calculation accuracy and calculation time, 4.2 million grids are used for this calculation. In order to verify the reliability of the model used in this paper to calculate the supercritical LNG heat transfer, the trans-critical methane experiment in the literature [5] was chosen as the benchmark experiment. Where the experimental data of Test Case #26 (TC26) was used as the validation data, and the same boundary conditions as the experiment are set in the literature [5]. As shown in Fig. 3(b), comparing the results, it can be found that the prediction results are closer to the experimental values with a maximum error of 9.98%. So, the model used in this paper can be used in the next step of the study.

Fig. 3. Comparison of experimental values and predicted values

3 Results and Discussion 3.1 Local Flow Heat Transfer Characteristics The distribution of the flow and temperature fields calculated by the numerical simulation method for the two design results are shown in the Fig. 4 –Fig. 5. It can be found that the flow velocity increases gradually as the heat transfer proceeds and that larger velocities are generated between the airfoil fins and at the leading edge of the airfoil fins. The uniform design produces greater outlet velocities and outlet temperatures compared to the non-uniform design. This inevitably produces a higher pressure drop. Figure 6(a) shows the temperature and pressure variations along the flow direction of the LNG in both flow channels. It can be found that the LNG temperature rises rapidly in both channels, then the temperature rise rate decreases slowly, and finally

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Fig. 4. Distribution of velocity in heat exchanger channel

Fig. 5. Distribution of temperature in heat exchanger channel

rises rapidly again. This is because the heat transfer process of LNG passes through the pseudo-critical region, where the very high specific heat capacity reduces the rate of LNG temperature rise. And it can be found that the temperature difference of LNG in the two channels in the pre-exchange period is not large, and the pressure drop generated in the non-uniform channel is lower than that in the uniform flow channel at this time. The pressure drop in the non-uniform flow channel is 230 Pa and the pressure drop in the uniform flow channel is 313 Pa. Combining the surface area per volume and heat load of the two designs mentioned in Sect. 2.1, it can be seen that the uniform flow channel design increases the surface area per volume by 19%, but the thermal load only increases by 6% and produces a pressure drop of about 35%. The non-uniform design is therefore superior. The comparisions of the Nu among two channel configurations are shown Fig. 6(b). Apparently, Nu increases with increasing bulk temperature and reaches its highest value near the pseudo-critical region, after which it gradually decreases. And there is a clear difference between the two channel configurations, with significant periodic oscillations in the local Nu variation of the uniform channel due to the periodic arrangement of the airfoil fins. While the non-uniform channels had low Nu in the previous exchange period, high Nu was obtained in the later exchange period. This is due to the relatively dense arrangement of the fins in the later stages of the non-uniform channel. 3.2 Comprehensive Evaluation of Global Flow Heat Transfer Characteristics Figure 7(a) shows the variation of the improved entropy generation number with temperature for the two-channel configuration. It is clear that the improved entropy generation

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Fig. 6. Variation of physical quantities in the flow channel

number for both channels decreases as the heat transfer proceeds. The improved entropy generation number in the flow heat transfer process is higher in uniform flow channels than in non-uniform flow channels, especially in the early stages of heat transfer. And it can be noticed that the non-uniform flow channels do not produce a great improved entropy generation number due to the later arrangement of denser fin structure. In order to evaluate the overall performance of the flow channel, the overall improved entropy generated number and the synergy angle of the LNG in the two structures of the flow channel are given in Fig. 7(b). It can be found that the synergy angle of the non-uniform flow channel is 0.65% bigger than that of the uniform flow channel, but its improved entropy generation number is 14.75% smaller than that of the uniform flow channel. Therefore, the non-uniform flow channel has better overall performance.

Fig. 7. Variation of improved entropy generation number with temperature(a) and improved entropy generation number and synergy angle resulting from a two-channel configuration(b)

4 Conclusion This paper presents and analyses a non-uniform arrangement of supercritical LNG airfoil flow channels that can be used for heat transfer in the LNG gasification process. The conclusions are as follows.

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1. The pressure loss in the non-uniform flow channel is significantly lower compared to the uniform flow channel. Due to the periodic arrangement of the fin structure, the uniform flow channel undergoes periodic oscillations in Nu and reaches a maximum in the pseudo-critical region. In contrast, non-uniform flow channels do not produce periodic variations in Nu and obtain higher Nu near the pseudo-critical region compared to uniform flow channels. 2. The improved entropy generation number of non-uniform flow channels is significantly lower than that of uniform flow channels, especially in the pre-exchange phase, and the denser fin structure arranged at the rear of the non-uniform flow channel does not generate an increased improved entropy generation number. Finally, the overall improved entropy generation number is 14.75% smaller than that of the uniform flow channel and the synergy angle of the non-uniform flow channel is not significantly different from that of the uniform flow channel. Overall, the simulation results show that non-uniform flow channels can be effective in improving the overall performance of PCHEs. The non-uniform flow channel forms can be extended to the design of PCHEs. Acknowledgments. This work is financially supported by the Project funded by China Postdoctoral Science Foundation (No. 2022M712711) and Postgraduate Research & Practice Innovation Program of Jiangsu Province (SJCX23_2198).

References 1. Baek, S., Hwang, G., Jeong, S., et al.: Development of compact heat exchanger for LNG FPSO. In: The Twenty-First International Offshore and Polar Engineering Conference, OnePetro (2011) 2. Xu, X., Ma, T., Li, L., et al.: Optimization of fin arrangement and channel configuration in an airfoil fin PCHE for supercritical CO2 cycle. Appl. Therm. Eng. 70(1), 867–875 (2014) 3. Jin, F., Chen, D., Hu, L., et al.: Thermo-Hydraulic performance of printed circuit heat exchanger as precooler in supercritical CO2 Brayton cycle. Appl. Therm. Eng. 210, 118341 (2022) 4. Jiang, Q., Pan, C., Song, X., et al.: Adaptive design methodology of segmented non-uniform fin arrangements for trans-critical natural gas in the printed circuit heat exchanger. Appl. Therm. Eng. 216, 119011 (2022) 5. Votta, R., Battista, F., Salvatore, V., et al.: Experimental investigation of transcritical methane flow in rocket engine cooling channel. Appli. Thermal Eng., 61–70 (2016)

A Simulation Study of Enhanced Radiation Cooling on a Radiator with High Emissivity Coating Ming Liu1,2

, Hongbo Xu1(B)

, and Nan Peng1

1 Technical Institute of Physics and Chemistry, Chinese Academy of Science, Beijing, China

[email protected], {hbxu,pengnan}@mail.ipc.ac.cn 2 University of Chinese Academy of Science, Beijing, China

Abstract. In the vacuum or natural convection, radiation is one of the main heat transfer methods. The structure and surface emissivity of the radiator have important impacts on radiation performances. A calculation model of radiative heat transfer of the rectangular finned radiator was established, and radiative characteristics under different emissivity and different ratios of fin height to fin spacing (H/s) were analyzed. The results showed that increasing the emissivity greatly enhanced the radiation cooling on the outer surface, while the effect on the fin channels was limited, especially for high and dense fins. It guided the use of the high emissivity coating for reducing the coating cost. The effects of high emissivity coating were further simulated under the vacuum, natural and forced convection. The results showed that at the heating power of 20 W, compared with the radiator coated (ε = 0.93) and uncoated (ε = 0.20), the heat transfer coefficient increased by 95.5%, and the thermal resistance decreased by 39.5% in the vacuum. As for natural convection, the radiation power with coating occupied 39.6% of the total dissipated power and the coating increased the heat transfer coefficient by 26.6% at the heating power of 10 W. Under forced convection, the radiation fraction with coating decreased from 22.6% to 6.4% with the wind speed increased from 0.5 m/s to 4 m/s. In conclusion, for the vacuum, natural convection and low wind speed condition, the high emissivity coating presents a good application prospect in enhancing radiation cooling. Keywords: Radiation cooling · High emissivity coating · Radiator · Simulation

1 Introduction As the size reduces and the power dissipations increase of electronic devices, effective heat dissipation technology is more demanding. Due to the advantages of energy-saving, no noise and simplicity, passive radiation cooling has attracted considerable attention [1, 2]. Although it is usually ignored, radiation plays a significant role in total heat transfer in many cases. In 1988, Buller et al. [3] demonstrated the effects of enhanced radiation cooling by increasing the emissivity of heat sinks. Yu et al. [4] experimentally and numerically studied the radiative performance of a radial heat sink and found that © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 551–558, 2023. https://doi.org/10.1007/978-981-99-6128-3_71

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the maximum radiation fraction reached 27% under natural convection. Zu et al. [5] also investigated the enhanced radiation cooling of a heat sink with the coating (ε = 0.93) in natural convection. SiO2 nanoparticles are selected as the additives for the high emissivity coating which is prepared by the spray approach. And the radiation contribution was 38%. To obtain high emissivity, adding SiO2 particles or carbon materials in coatings is the main method. Mahadik et al. [6] produced double layer Al2 O3 /SiO2 coatings with high emissivity (0.92–0.94) on stainless steel substrate by sol-gel process. Tun-Jen Hsiao et al. [7] prepared MF coatings for enhancing the cooling of LED or CPU and the emissivity of coatings on Al panel was more than 0.9. The temperature of heat sinks decreased obviously by the application of coatings. Overall, the high emissivity coating performs well in terms of improving heat dissipation in natural convection, but there few researches under vacuum and forced convection. In this paper, the effects of enhanced radiation cooling on a typical finned radiator with high emissivity coating were investigated by theoretical and simulative approaches. The thermal performances of the radiator were analyzed in the vacuum, natural convection and forced convection, respectively. The work could provide a reference for evaluating the effects of using the coating under different conditions.

2 Simulation Methods 2.1 The Calculation Model of Radiation Heat Transfer A typical rectangular fin radiator is shown in Fig. 1. Assuming that the surface of the radiator is the gray body wall and the surface temperature is constant, the radiosity of fin channels can be approximately calculated by Ji = εσ Ts4 + (1 − ε)

4 

Ji Fij (i = 1, 2, 3, 4)

(1)

j=1

The total radiation power of the radiator is expressed as Pr = (N − 1)

3  εAi (σ T 4 − Ji ) s

i=1

1−ε

+ εσ Ae (Ts4 − T04 )

Ae = Nt(L + 2H ) + 2HL + 2b(L + W ) + 2LW

(2) (3)

where ε is the surface emissivity of the radiator; F ij is the angle factor of the surface i to the surface j; N is the number of the fins; Ai is the area of the surface i; T 0 is ambient temperature; T s is the surface temperature of the radiator; Ae is the outer surface area other than fin channels. The radiation power can be obtained by these equations. 2.2 Numerical Simulations As shown in Fig. 2, a model of the horizontal finned radiator with a surface heat source was established to simulate the conduction, convection and radiation heat transfer by

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ANSYS Icepak15.0 software. Three temperature monitoring points are selected on the surface of the radiator to judge whether it has reached a stable state. The heat source size located at the bottom center of the radiator is 20 × 20 mm. The size of the radiator is consistent with that in Fig. 1. The length and width of the radiator are 100 mm and 60 mm, respectively. In these simulations, the initial temperatures of the radiator were equal to the ambient temperature of 25 °C and the heating power P was 10 W, 15 W and 20 W, respectively. The surface emissivity of the radiator without the coating was set to 0.20 and the one with the coating was 0.93. According to the grid independence test results, as the number of mesh increasing from 341,848 to 860,200, the deviation of the temperature of point 3 decreased from 0.23 to 0.04 °C. To save computing time, the mesh number 581,960 was selected.

Fig. 1. A typical rectangular fin radiator.

Fig. 2. The simulation model of the radiator.

3 Results 3.1 The Verifications of the Calculation Model and the Numerical Simulation To verify the numerical simulation, the simulation results were compared with the experimental results in the reference [5] under the same heating power of 14 W and the emissivity of 0.934 in Table 1. The relative error of temperatures at point 3 between the experiment and the simulation is 6.4% and it is reasonable. The main reasons for deviations include: (a) errors in numerical simulations; (b) environmental disturbances in the experiment; (c) temperatures measured errors by the thermocouples. Further, the accuracy of the calculation model was verified by compared with the simulation results. According to Table 1, the error in total radiation powers between the simulation and the calculation is 9.5% within the acceptable range. The results indicate the feasibility of theoretical calculation model and simulation method. 3.2 Analysis of Radiation Heat Transfer Properties The radiation performances of the radiator are closely related to the surface temperature, emissivity and structure. As shown in Fig. 3, the total radiation power of the radiator

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was calculated by the Eqs. (1)–(3), which rose with the emissivity under different surface temperatures. The higher the temperature T s , the faster the radiation power grew. The results indicated that increasing the emissivity was extremely effective to enhance the radiation heat transfer. Due to the directivity of radiation, the radiation properties of different surfaces of the radiator are diverse. The radiative heat transfer coefficient of outer surfaces hr, out linearly increased with the emissivity (see Fig. 4) because the surface-to-ambient radiation was only considered. But the heat transfer coefficient of fin channels hr, ch rose slightly because of surface-to-surface radiation of fins. The average radiative heat transfer coefficient hr, avg was between hr, out and hr, ch , hr, out mainly contributed to hr, avg . Moreover, the values of hr, ch under the different ratios of fin height to spacing (H/s) were obtained to analyze the effects of the structure of radiators. As shown in Fig. 5, with the ratio (H/s) increasing, hr, ch decreased rapidly and tended to be consistent in the condition of various emissivity. It suggested that the higher and denser the fins were, the worse the radiative heat transfer performance was, and enhancing radiation by the coating were limited. It is helpful to use coating properly and decrease the cost by partly coating on the outer surfaces with the large angle factor. Table 1. The comparison between the experiment, simulation and the calculation results. Type of results

Temperature at point 3

Total radiation power

Value/°C

Error

Value/W

Error

Experiment

80.9

6.4%

-

-

Simulation

75.7

4.53

9.5%

Calculation

-

Fig. 3. The radiation power of the radiator changed with the emissivity.

-

4.10

Fig. 4. The radiative heat transfer coefficients changed with the emissivity.

Fig. 5. The radiative heat transfer coefficient of fin channels under different ratios.

3.3 Radiation Cooling in Vacuum The heat transfer of the radiator was simulated under different heating powers in the vacuum where radiation is the only way to dissipate heat. As shown in Fig. 6, the temperature of monitoring point 3 increased gradually and reached the stable state and the temperature with coating was much lower than that without coating. And the maximum

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temperature differences between the radiator with coating and without coating were 81.1 °C, 97.7 °C, 109.7 °C under the heating powers of 10 W, 15 W and 20 W, respectively. Meanwhile, specific thermal properties of the radiator were given in Table 2. With the emissivity increased from 0.20 to 0.93, the thermal resistance decreased by 44.9%, 41.9%, 39.5%, and the heat transfer coefficient increased by 116.8%, 103.7%, 95.5% when the corresponding heating powers were 10 W, 15 W and 20 W, respectively. It indicated that the heat dissipation capability of the radiator was greatly improved by the high emissivity coating.

Fig. 6. The temperature curves of point 3 of the radiator with and without coating in vacuum.

Fig. 7. The simulation results of radiator in natural convection. (a) The temperature distribution. (b) The velocity distribution.

Fig. 8. The temperature curves of point 3 of the radiator in natural convection.

Table 2. The thermal properties of the radiator with different heating powers in the vacuum. Heating power /W

Emissivity

Thermal resistance

Heat transfer coefficient

Value/ °C/W

The decreased ratio/%

Value/ W/(m2 ·K)

The increased ratio/%

20.01

−44.9%

1.70

116.8%

−41.9%

2.03

10

0.20 0.93

11.02

15

0.20

16.98

0.93

9.87

0.20

14.91

0.93

9.02

20

3.69 103.7%

4.13 −39.5%

2.32

95.5%

4.54

3.4 Radiation Cooling in Natural Convection Figure 7 showed the temperature and velocity distribution of the central section of the radiator with the heating power of 15 W in natural convection. The temperature of the radiator with coating was significantly lower than that without coating and the air velocity near the radiator with coating was also slower. As shown in Fig. 8, the maximum temperature differences of point 3 between the radiator with coating and

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without coating were 8.8 °C, 11.4 °C, 13.7 °C at the heating power of 10 W, 15 W, 20 W. Table 3 further provided the heat transfer properties of the radiator. The radiation fraction rose from 16.6% to 39.6% due to the coating at the heating power of 10 W. The coating also decreased the thermal resistance by 14.9% and increased the heat transfer coefficient by 26.6%. These demonstrated the significance of radiation and the enhanced heat dissipation performance of the coating in natural convection. 3.5 Radiation Cooling in Forced Convection To evaluate the radiation cooling effect of the high emissivity coating in forced convection, the radiator was simulated at different wind velocities whose direction was parallel to the fin. In Fig. 9, when the heating power was 20 W and the velocity was 0.5 m/s, the surface temperature of the radiator without coating was higher than that with coating, and the temperature difference at the center was about 6.5°C. It showed the coating presented a certain effect of strengthening heat dissipation at low wind speed. Since the velocity distribution was mainly affected by the airflow, the velocity difference between ε = 0.20 and ε = 0.93 was slight. Figure 10 depicted the downward trend of the temperature of point 3 at the heating power of 10 W and 20 W. As the wind velocity increased, the convection heat transfer was enhanced and the temperature difference between the radiator coated and uncoated gradually decreased. Furthermore, the thermal properties of the radiator at different wind speeds were given in Table 4. When the wind velocity rose from 0.5 m/s to 4 m/s, the radiation fraction of the radiator coated decreased from 22.6% to 6.4%, and the decreased ratio of the thermal resistance dropped from 10.3% to 1.9%. At the same time, the increased ratio of the heat transfer coefficient decreased from 17.0% to 4.5%. It indicated that the high emissivity coating could improve the heat dissipation only at the low wind speed and the radiation could be very slight while the velocity is relatively high. Table 3. The heat transfer properties of the radiator in natural convection. Heating power /W

Emissivity

10

0.20 0.93

39.6

4.93

15

0.20

16.1

5.19

0.93

38.3

4.47

0.20

15.9

4.79

0.93

37.6

4.16

20

Radiation fraction/%

Thermal resistance

Heat transfer coefficient

Value/ °C/W

The decreased ratio/%

Value/ W/(m2 ·K)

The increased ratio/%

16.6

5.79

−14.9

6.16

26.6

−13.9

6.86

7.80 25.1

8.58 −13.2

7.42 9.22

24.3

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4 Conclusions In this work, a theoretical model was built to study the radiation properties of a finned radiator. The effects of enhanced radiation cooling on the coated radiator were analyzed numerically in the vacuum, natural convection and forced convection environment, respectively. The calculation results of the model showed that for the radiator with high and dense fins, the enhanced radiation effect of the fin channels was relatively weaker than that of outer surfaces by increasing the emissivity. From the simulations, the enhancement was the most significant in the vacuum and the maximum temperature difference between the radiator uncoated and coated was as high as 109.7 °C. The maximum radiation fraction was 39.6% because of the high emissivity coating which increased the heat transfer coefficient by 26.6% under natural convection. The radiation should not be ignored completely in forced convection. The improvement of the coating was more beneficial at the lower wind velocity.

Fig. 9. The simulation results of radiator in forced convection. (a) The temperature distribution. (b) The velocity distribution.

Fig. 10. The stable temperature of point 3 of the radiator changed with the wind velocity in forced convection.

Table 4. The heat transfer properties of the radiator in forced convection (P = 20 W). Wind velocity/ m/s

Emissivity

0.5

0.20 0.93

2.0 4.0

Radiation fraction/%

Thermal resistance

Heat transfer coefficient

Value/ °C/W

The decreased ratio/%

Value/ W/(m2 ·K)

The increased ratio/%

8.7

3.10

−10.3

12.07

17.0

22.6

2.78

0.93

15.5

1.96

0.20

3.3

1.32

0.93

9.6

1.27

0.20

2.1

0.92

0.93

6.4

0.90

14.12 19.63 −3.5

28.28

6.8

30.21 −1.9

41.99 43.87

4.5

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Acknowledgments. The project is supported by the Project of Stable Support for Youth Team in Basic Research Field, CAS (YSBR-017) and the Fund of Director of Technical Institute of Physics and Chemistry, CAS.

References 1. Hossain, M.M., Gu, M.: Radiative cooling: principles progress potentials. Adv. Sci. 3(7), 1–10 (2016) 2. Zhao, B., Hu, M.K., Ao, X.Z., et al.: Radiative cooling: a review of fundamentals, materials, applications, and prospects. Appl. Energy 236, 489–513 (2019) 3. Buller, L., McNelis, B.: Effects of radiation on enhanced electronic cooling. IEEE Trans. Components Hybrids Manuf. Technol. 11(4), 538–544 (1988) 4. Yu, S.H., Jang, D., Lee, K.S.: Effect of radiation in a radial heat sink under natural convection. Int. J. Heat Mass Transf. 55(1), 505–509 (2012) 5. Zu, H., Dai, W., et al.: Analysis of enhanced heat transfer on a passive heat sink with highemissivity coating. Int. J. Therm. Sci. 166, 106971 (2021) 6. Mahadik, D.B., Gujjar, S., Gouda, G.M., et al.: Double layer SiO2/Al2O3 high emissivity coatings on stainless steel substrates using simple spray deposition system. Appl. Surf. Sci. 299, 6–11 (2014) 7. Hsiao, T.J., Eyassu, T., Henderson, K., et al.: Monolayer graphene dispersion and radiative cooling for high power LED. Nanotechnology 24(39), 395401 (2013)

Cryocoolers

Investigation on a Free-Piston Stirling Cryocooler at −80 °C for China Space Station Zhang Yin1,2 , Wang Bo1,2(B) , Ni Zhuqing1,2 , Luo Gaoqiao1,2 , Wang Shuai3,4 , Tian Xinghua5 , Wu Weiwei1,2 , Yao Xiaolei1,2 , Liu Ting1,2 , Zhang Yongqing1,2 , Zhu Kuizhang1,2 , Fan Yufeng3,4 , and Zhang Chi1,2 1 Laboratory of Cryogenic and Refrigeration Technology, No. 658 Wangjiang Road,

Hefei 230088, Anhui, China 2 16th Institute of China Electronics Technology Group Corporation, No. 658 Wangjiang Road,

Hefei 230088, China [email protected] 3 Beijing Institute of Space System Engineering, Beijing 100094, China 4 Beijing Key Laboratory of Space Thermal Technology, Beijing 100094, China 5 Key Laboratory of Space Utilization, Technology and Engineering Center for Space Utilization, Chinese Academy of Sciences, Beijing 100094, China

Abstract. A free-piston Stirling cooler at −80°C for China Space Station has been successfully developed, which has features of high frequency, short regenerator, and high voltage. The cooler can provide a cooling capacity of 65.0 W at − 87 °C when the input power is 279.6 WAC using a square wave. Considering the vibration output, a sine wave is used. For stable working condition of the refrigerator, the cooler can provide a cooling capacity of 22.5 W at −87 °C when the input power is 99.1 WAC ; For the cooling-down condition of the refrigerator, the cooler can provide a cooling capacity of 43.9 W at −87 °C when the input power is 179.5 WAC . The cooler can withstand the aerospace mechanical and thermal environment tests. Keywords: −80°C · free-piston Stirling cryocooler · China space station · cooling capacity

1 Introduction According to the requirements of space low-temperature storage environment, the development of space cryogenic storage devices has been carried out in the manned aerospace field, such as SOR/F, MELFI, GLACIER, and TWINBIRED low-temperature refrigerators [1–3]. Most cryogenic storage devices utilize free-piston Stirling cryocoolers (46 sets) [4–8]. China Space Station has two Laboratory Modules. Laboratory Module I need a lowtemperature storage device, that is, a refrigerator. The refrigerator has three different temperature storage areas, corresponding to storage areas of −80 °C, −20 °C, and + 4 °C. The −80 °C storage area has a volume of 25 L and is cooled by a Stirling © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 561–568, 2023. https://doi.org/10.1007/978-981-99-6128-3_72

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cryocooler. The main technical indexes of the cryocooler are as follows. ➀ Cooled by liquid of ethylene glycol aqueous solution, the fluid temperature is 23–30.5 °C, and the maximum flow is 70 L/hr; ➁ The cooling capacity is over 18.5 W at −87 °C (input power ≤ 100 WAC ), over 39.5 W at −87 °C (input power ≤ 180 WAC ), and over 65.0 W at − 87 °C (input power ≤ 300 WAC ); ➂ The weight is below 7.0 kg; ➃ Steady-state noise is below 50 dBA; ➄ Meet aerospace mechanical and thermal environment adaptability; ➅ Lifetime should be over 6.67 years [9]. The Stirling cooler is a component, which consists of a free-piston Stirling cryocooler, liquid heat exchanger, thermal insulation component, cold plate coupler, connecting lug, temperature sensors and power supply components, magnetic damping vibration absorber (see Fig. 1). The cryocooler is installed at the back of the + 4°C storage area of the refrigerator (see Fig. 2). The cooler is rigidly connected with the flat plate heat pipe of −80 °C storage area through the cold plate coupler. The waste heat of the cooler is collected by the liquid heat exchanger, then discharged outside of the Space by the liquid circuit. The vibration of the cooler is damped by the magnetic damping vibration absorber.

Fig. 1. Schematic diagram of Stirling cryocooler

Fig. 2. Stirling cryocooler and low-temperature storage device

In this paper, the thermodynamic model of the Stirling cryocooler at −80 °C storage area is established. The free-piston Stirling cryocooler at −80°C is successfully developed, and the cooling performance of the cryocooler is tested. The reliability and lifetime tests are carried out according to the Space’s environmental conditions. The tested results of cooling capacity, steady state noise, vibration output, environmental adaptability and ergonomics meet the Space Station. It was launched together with Laboratory Module I

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in July, 2022, and has supported various scientific experiments and research of the China Space Station.

2 Simulation, Design, and Verification The structural diagram of the free-piston Stirling cooler is shown in Fig. 3. Firstly, the thermodynamic parameters of the cooler are designed by SAGE, and the motor simulation is designed by Maxwell. The design parameters are shown in Table 1. The cooler has a feature of high frequency (80 Hz), short regenerator (37.7 mm), and high voltage (56 V).

Fig. 3. Structure diagram of free-piston Stirling cryocooler

Table 1. Design parameters of −80 °C free-piston Stirling cryocooler Parameters

Calculation results Parameters

Calculation results

Pressure/MPa

2.4

35

Magnet diameter/mm

Frequency/Hz

80

Outer stator diameter/mm 50

Piston diameter/mm

30

Coil voltage/V

56

Displacer diameter/mm

20

Coil diameter/mm

1.25

Regenerator length/mm

37.7

Coil turns

135

Spring stiffness/N·mm−1

12

Motor force/N

98

Cooling capacity/W

85.5

Compressor efficiency/%

85.1

Input power/WAC

292

To obtain the optimal operating parameters of the cooler, the effects of operating frequency ( f ) and inflation pressure (P0 ) on the cooling capacity (Qnet ) are simulated and analyzed, respectively (see Figs. 4 and 5). Figure 4 shows the influence of operating frequency f on P-V work W pv , electric input power W input and cooling capacity Qnet . As seen from the figure, the W pv , W input and cooling capacity Qnet increase with operating frequency f at first, and then decrease. When the operating frequency f is 75 Hz, the cooling capacity reaches the maximum value Qnet of 86.2 W at −80 °C, with W input being 196.2 W, W pv being 118.5 W, and the cooling efficiency COP being 0.29. Considering

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Fig. 4. The influence of frequency f on P-V work, input power and cooling capacity

the emission fundamental frequency must greater than 70 Hz, the operating frequency of the cooler is selected as 80 Hz. Figure 5 shows the effect of inflation pressure P0 on P-V work W pv , electric input work W input and cooling capacity Qnet . As seen from the figure, the W pv , W input and cooling capacity Qnet increase with inflation pressure P0 at first, and then decrease. When the inflation pressure P0 is 2.4 MPa, the P-V work W pv reaches 242 W, the electric input power W input reaches 290 W, the cooling capacity reaches 83.7 W@−80 °C, and the cooling efficiency COP is 0.29. Through simulation results, the inflation pressure is selected as 2.4 MPa, and the cooling capacity and input power of the cooler meet the design requirements.

Fig. 5. The inflation pressure influence of P0 on P-V work, input power and cooling capacity

3 Experimental Research 3.1 Cooling Performance Test The cooling performance has been tested at different fluid temperatures of 26 °C, 28 °C and 30.5 °C, respectively. The test results are shown in Fig. 6. The test results are as follows: 1) When the fluid temperature is 26 °C and the input power is 98 WAC , the cooling capacity is 23.6 W; When the input power is 180 WAC , the cooling capacity is

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45.5 W; When the input power is 279.8 WAC , the cooling capacity is 70.8 W (Square wave); 2) When the fluid temperature is 28°C and the input power is 99.8 WAC , the cooling capacity is 23.1 W; When the input power is 179 WAC , the cooling capacity is 44.79 W; When the input power is 279.9 WAC (Square wave), the cooling capacity is 70.0 W; 3) When the fluid temperature is 30.5 °C and the input power is 99.1 WAC , the cooling capacity is 22.5 W; When the input power is 179.5 WAC , the cooling capacity is 43.96 W; When the input power is 279.6 WAC (Square wave), the cooling capacity is 68.58 W. The test results meet the Space requirement.

Fig. 6. Curves of cooling capacity versus input power under different fluid temperatures (ambient temperature 23 °C)

3.2 Environmental Adaptability Test The cooler experiences vibration and shock tests to simulate the real condition. The cooler was attached to the same chamber of the refrigerator chamber, and this chamber is located on the vibration and shock test platform. The total root mean square acceleration value of the triaxial loading is 9 Grms, and the duration is 90 s. During the mechanical test, acceleration sensors are arranged at the cooler (measuring point 8) and absorber (measuring point 9). The mechanical test site is shown in Fig. 7, and the response results of each measuring point are shown in Table 2. According to the mechanical response results, the maximum root mean square of the cooler is 17.68 Grms (X-direction attach, 297.5 Hz). The largest response is found at the absorber, and the root mean square is 42.52 Grms (Y-direction attach, 190 Hz) because of the collision of the moving parts. The cooling performance of the cooler is tested before and after the mechanical test (see Fig. 8). Before the mechanical test, the cooling capacity is 42.67 W when the input power is 176.4 WAC ; After the mechanical test, the cooling capacity is 42.71 W when the input power is 174 WAC . After the mechanical test, the cooling capacity is hardly reduced, and the leakage rate of the cooler is 1.5 × 10–10 Pa·m3 ·s−1 . The mechanical test results indicate that the cooler can suffer the aerospace mechanical test. The thermal-cycling experiment of the cooler is put in the high-low temperature test chamberjavascript:;. The atmosphere temperature ranges from -5°C to + 40 °C, and the number of cycles is 6.5. The fluid temperature ranges from + 5°C to + 30 °C. In

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Fig. 7. Mechanical test site

Table 2. Acceleration response results of each measuring point Measuring Direction Code Position point

Root mean Maximum Corresponding square/Grms response/g2·Hz−1 frequency/Hz

Direction X 8

9

X

A8X Cryocooler 17.68

6.61

297.5

Y

A8Y

6.54

1.86

197.5

4.95

0.267

495

17.59

6.01

297.5

Z

A8Z

X Y

A9X Vibration A9Y absorber

20.45

Z

A9Z

14.34

2.48

495

X

A8X Cryocooler 12.81

5.79

190

Y

A8Y

16.17

8.68

190

Z

A8Z

4.59

0.29

190

X

A9X Vibration A9Y absorber

11.1

1.99

Y

42.52

67.09

Z

A9Z

15.3

4.93

197.5

X

A8X Cryocooler 11.1

3.37

192.5

Y

A8Y

8.70

3.18

192.5

Z

A8Z

17.13

6.72

230

X

15.41

2.07

192.5

Y

A9X Vibration A9Y absorber

29.76

27.23

195

Z

A9Z

36.88

47.97

232.5

16.3

197.5

Direction Y 8

9

197.5 190

Direction Z 8

9

each cycle, the total maintained time of on-off and continuous operation at high and low extreme temperatures is no less than 4 h, and the cold and hot start are carried out three times, respectively. The curve of the thermal cycling ambient temperature is shown in Fig. 9. The cooling capacity of the cooler is tested before and after the thermal-cycling

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Fig. 8. Comparison of cooling capacity before and after mechanical test

experiment (see Fig. 10). Before the thermal cycling test, the cooling capacity is 44.42 W when the input power is178.4 WAC ; After the thermal cycling test, the cooling capacity is 44.38 W when the input power is 178.1 WAC . After the thermal cycling test, the cooling capacity is hardly reduced. The thermal cycling test results indicate that the thermal design of the cooler meets the requirements. The cryocooler can undertake the thermal-cycling experiment.

Hot start

T/

Two-time Low-temperature Starting ( -87 )

5 min H

g down

0.5

Symbol Description Power on Shut down Start high temperature liquid supply Start low temperature liquid supply 10 min Heat up 2h 2h Maintain Continuous time working

Coolin

Room temperature

2h 2h Continuous Maintain working

15 min

eat up

Th

Tl Cold Start

Two-time Low-temperature Starting (

A total cycle (8.5 h)

Fig. 9. Curve of the thermal-cycling ambient temperature

Fig. 10. Comparison of cooling capacity before and after the thermal-cycling test

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4 Conclusion A free-piston Stirling cooler at −80°C for China Space Station has been successfully developed, which has features of high frequency (80 Hz), short regenerator (37.7 mm), and high voltage (56 V). The cooler can provide a cooling capacity of 65.0 W at −87 °C when the input power is 279.6 WAC using a square wave. Considering the vibration output, a Sine wave is used. For stable working condition of the refrigerator, the cooler can provide a cooling capacity of 22.5 W at −87 °C when the input power is 99.1 WAC ; For the cooling-down condition of the refrigerator, the cooler can provide a cooling capacity of 43.9 W at −87 °C when the input power is 179.5 WAC . The cooler can withstand the aerospace mechanical and thermal environment test.

References 1. Cairelli, J.E., Gaseror, T.: NASA life sciences advanced refrigerator/freezer technology assessment results. J. Aerospace 105(1), 563–573 (1996) 2. Cairelli, J.E.: NASA advanced refrigerator/freezer technology development project overview. In: Ross, R.G. (ed.) Cryocoolers 8, pp. 919–925. Springer, Boston, MA (1995). https://doi. org/10.1007/978-1-4757-9888-3_90 3. Chen, G., Tang, L., Mace, B., et al.: Multi-physics coupling in thermoacoustic devices: a review. Renew. Sustain. Energy Rev. 146,(2021) 4. Petrivelli, A.: The ESA laboratory support equipment for the ISS. ESA Bull. 0376–4265(109), 35–54 (2002) 5. Unger R., Keiter D.: The development of the Cryotel™ family of coolers. In: AIP conference Proceedings. American Institute of Physics, pp. 1404–1411 (2004) 6. Lane N.W, Wood J.G., Unger R.Z.: Free-piston Stirling machine commercialization status at Sunpower. In: International Stirling Engine Conference, pp. 19–21. Rome, Italy (2003) 7. Wood J.: Status of free-piston Stirling technology at Sunpower. In:1st International Energy Conversion Engineering Conference, pp. 234–237. Portsumouth, Virginia (2003) 8. Karandikar, A., Berchowitz, D.: Low cost small cryocoolers for commercial applications. In: Kittel, P. (ed.) Advances in Cryogenic Engineering. A Cryogenic Engineering Conference Publication, vol. 41. Springer, Boston, MA (1996). https://doi.org/10.1007/978-1-4613-03732_196 9. Wang S., Wang B., Xiao M., et al.: Investigation of Low temperature storage device of China Space Station. In: 15th Cryogenic Engineering Conference, pp. 938–947. Shanghai (2021) 10. Wan, F., Tan, Y., Jiang, Z., et al.: Lifetime prediction and reliability estimation methodology for Stirling-type pulse tube refrigerators by gaseous contamination accelerated degradation testing. Cryogenics 88, 116–124 (2017)

Experiment Study a High-Power Pulse Tube Cryocooler Haibin Dai and Shaowei Zhu(B) Tongji University, Shanghai 201804, China [email protected], [email protected]

Abstract. High temperature superconductivity and liquefaction technology develop rapidly in recent years, and the higher requirement for the cryogenic devices supplying more cooling power and higher efficiency is essential. Pulse tube cryocooler is a promising choice due to its advantages of no moving part under low temperature, simpler structure, high reliability and long life. This work introduces the optimizations on a coaxial high-power pulse tube cryocooler by regenerator inner diameter of 100 mm and pulse tube out diameter of 50 mm with 1mm wall thickness. To get the optimized thickness of flow straightener at the cold end, both the cold end and the warm end were tested experimentally by different thickness flow straighteners respectively. Different groups of regenerator screen mesh and two regenerators of 65mm and 80mm were tested, too. A relative Carnot efficiency of 18.7% associated with Cooling power of [email protected] was got. Keywords: High-power Pulse Tube Cryocooler · Regenerator · Flow straightener

1 Introduction With the development of high-temperature superconducting technology, the market demand for large-capacity low-temperature refrigerators in the liquid nitrogen temperature region is increasing [1]. High-temperature superconducting cooling systems usually need to provide hundreds of watts or even several kilowatts of cooling capacity in the liquid nitrogen temperature region [2, 3]. The factors of efficiency, stability, service life, and convenient maintenance are got a lot of attentions. The pulse tube refrigerator is widely used in the aerospace field due to its compact structure, no moving part at the cold end, simple structure, high stability and long service life, which is expected to be the best choice for a cooling system for high-temperature superconducting equipment [4]. Higher cooling power pulse tube refrigerator is important for LNG temperature range [5], too. Pulse tube and regenerator are the key components of a pulse tube refrigerator. The uniformity of the gas in the pulse tube, the layer thickness ratio of different wire mesh screens of the regenerator is the key factors affecting the performance of the high-power pulse tube refrigerator [6–8]. This paper studies the effects of the flow straightener of the pulse tube, ratio of the regenerator screen mesh, length of the regenerator, and the inlet © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 569–575, 2023. https://doi.org/10.1007/978-981-99-6128-3_73

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of the inertance tube for getting higher efficiency for a coaxial pulse tube refrigerator by regenerator inner diameter 100mm and pulse tube out diameter 50mm with 1mm wall thickness.

2 Cold Flow Straightener Effect The basic information of the experimental bench is shown in Table 1. The operating pressure is 3 MPa and the frequency is 56 Hz. Table 1. Experiment system information. Parts

Parameters

Compressor

Amplitude 15 mm, 10 kW

Regenerator

Inner diameter D50 mm, Out diameter D100 mm, length L 80mm

Pulse tube

Out diameter D50 mm, Wall thickness 1 mm, length L 205 mm

Inertance tube

Inner diameter D12 mm, Wall thickness 1 mm, length L 2 m

Buffer

Volume 4.7 L

The cold flow straightener of the pulse tube is important especially for a pulse tube with large diameter. The reason is that the gas flow field in the pulse tube will be very poor and not smooth without any flow straighter or with a too thin layer thickness of mesh screen. In order to study the effect of different parameters on the performance of this refrigerator, and the parameters including the layer thickness of L, mesh number # and wire diameter of d of the cold flow straightener, three groups of cold flow straightener were set up, and the parameters are shown in Table 2. Table 2. Parameters of cold flow straightener. Group

Parameters

1

L5 mm (#100, d 0.1 mm)

2

L3.6 mm (#100, d 0.1 mm) + L1.4 mm (#50, d 0.2 mm)

3

L3.6 mm (#100, d 0.1 mm) + L1.4 mm (#400, d 0.022 mm)

Figure 1 shows the variation of efficiency and cooling capacity with temperature under three groups of cold flow straightener. The definition of the efficiency is relative Carnot efficiency which is real COP over Carnot COP times 100. When pure #100 wire mesh, the hybrid of #100 wire mesh with #400 wire mesh and coarse mesh wire mesh are added respectively, the corresponding no-load temperatures are 54.4K, 57.5K and 59.8K respectively. Likewise, the corresponding highest efficiencies are 12.82%, 13.46%, and 13.22%, respectively. The hybrid screen mesh of flow straightener mesh to efficiency

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Fig 1. a. Cold flow straightener effect on cooling power. b. Cold flow straightener effect on efficiency

is relatively small, and pure #100 filling is a little higher. Subsequently, the effect of flow straightener thickness was done with pure #100 screen mesh, and the cold flow straightener thickness was set to 3 mm, 5 mm, and 8 mm. The total thickness of flow straightener at both ends of the pulse tube is 18 mm, and the length of the pulse tube is not changed in the experiment. The results are shown in Fig. 2. It is shown that the highest efficiency of 17.25% and the cooling capacity of 300W were obtained at 110K when the thickness of the cold flow straightener is 5 mm.

Fig 2. a. Cooling power versus cold flow straighter length. b. Efficiency versus cold flow straighter length

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3 Regenerator Effect 3.1 Regenerator Wire Mesh Effect In order to study the effect of different wire mesh and its layer thickness ratio which is its length over the total length of the regenerator on the performance, different experimental groups filled with different wire meshes with different layer thickness ratios are shown in Table 3. The average pressure is 3 MPa. Table 3. Mesh parameters in regenerator Group

Mesh parameters

1

1/3(#250 d 0.05 mm) + 1/3 (#300 d0.03 mm) + 1/3 (#400 d 0.03 mm)

2

1/2 (#300 d 0.03 mm) + 1/2 (#400 d 0.03 mm)

3

100% #300 d 0.03 mm

It can be seen from Fig. 3 that the efficiency of the refrigerator is the highest with filling method of group 1 with #250, #300 and #400, and a lower no-load temperature can be obtained, too. Specifically, the cooling capacity of 300W was obtained at 110.9K, and the efficiency was 17.25%.

Fig 3. a. Regenerator mesh effect on cooling power. b. Regenerator mesh effect on efficiency

3.2 Influence of Regenerator Length In this experiment, the optimal ratio of mesh obtained in the experiment of 3.1 was used, and the length of the regenerator was changed to 65 mm. The experimental results are shown in Fig. 4.

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Fig 4. Regenerator length effect on cooling power. b. Regenerator length effect on efficiency

With regenerator length of 65 mm, a no-load temperature of 50.6K was obtained at the frequency of 54 Hz, a cooling capacity of 300W was obtained at a temperature of 118.9K, and the relative Carnot efficiency was 17.7%. The efficiency of the refrigerator is no big difference with the regenerator length of 65 mm and 80 mm. For a big pulse tube refrigerator, there is temperature difference along the circumference direction of the regenerator at same position of axial direction, which makes big pulse tube refrigerator hard to get high efficiency. Longer length regenerator has large ratio of radius over length, which may weak the bad influence of temperature difference along circumference direction, but it will increase pressure drop which is not a good thing. So, there should be an optimal length for the regenerator in high power pulse tube refrigerator to get the best performance eventually.

4 Inertance Tube Inlet Size Effect During the experiment, it was found that part of the wire mesh of the hot flow straightener of the pulse tube near the inlet of the inertance tube was broken. The inlet is a small zone with a diameter bigger than inertance tube but smaller than pulse tube. Wire mesh could be broken by the jet flow due to high speed of the gas from the inlet. And the existence of the jet flow would greatly affect the uniformity of the gas inside the pulse tube and worsen the cooling performance. The diameter of the inlet of the inertance tube was increased from 20 mm to 36 mm, and the experimental results were obtained which is shown in Fig. 5. Under the same condition, the cooling power of 300W at temperature of 124.6K was obtained after reaming the inlet of the inertance tube, and the Relative Carnot efficiency of 18.7% was reached, which is 6% higher than the previous one. The efficiency is 18.05% with 202W at 88.4K. It could be deduced from the data lines in Fig. 5 that the cooling power and efficiency would be 169.8W and 16.6% at 80K cold temperature respectively. It shows that the inlet of the inertance tube is also important for efficiency.

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Fig 5. a. Inlet of the inertance tube effect on cooling power. b. Inlet of the inertance tube effect on efficiency

5 Summaries A high-power Stirling type pulse tube refrigerator with hundred watts cooling power was studied experimentally. There is an optimum thickness of 5mm for cold flow straightener to get the highest efficiency. Regenerator with hybrid mesh with different regenerator length and mesh ratio filling were tested. Higher efficiency with a mesh filling of one third of 400#, 300#, and 250# was got. The efficiency with regenerator length of 80 mm is close to the length of 65 mm. With a bigger diameter of inlet of inertance tube, the cooling power of 300W at temperature of 124.6K with relative Carnot efficiency of 18.7% was reached eventually. Acknowledgments. The project is supported by National Natural Science Foundation of China (Number 52076151).

References 1. Hu, J.Y., Zhang, L.M., Zhu, J., et al.: A high-efficiency coaxial pulse tube cryocooler with 500W cooling capacity at 80K. Cryogenics 62, 7–10 (2014) 2. Gong M., Guo, H., Sun, Z., et al.: Advances in mobile natural gas Mini liquefierj. CIESC J. 66, S2 (2015) 3. Hu, J., et al.: Design of a large-capacity multi-piston pulse tube cryocooler. AIP Conf. Proc. 1434(1), 540–546 (2012) 4. Dietrich, M., Yang, L.W., Thummes, G.: High-power Stirling-type pulse tube cryocooler: observation and reduction of regenerator temperature-inhomogeneities. Cryogenics 47(5–6), 306–314 (2007) 5. Zhaorui, G.: Experimentational investigation of high-capacity pulse tube refrigerator at medium and low temperature zone. Tongji University, master theses (2020) 6. So, J.H., Swift, G.W., Backhaus, S.: An internal streaming instability in regenerators. J. Acoust. Soc. Am. 120(4), 1898–1909 (2006)

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7. Fang, K., Qiu, L.M., Liu, D.H., et al.: Effect of non-uniform porosity on temperature inhomogeneity in the regenerator of high power Stirling type pulse tube cryocoolers. In: 24th International Cryogenic Engineering Conference, Fukuoka, Japan, pp. 347–350 (2012) 8. Andersen, S.K., et al.: Numerical study on transverse asymmetry in the temperature profile of a regenerator in a pulse tube cooler. Int. J. Heat Mass Transf. 50(13–14), 275–2804 (2007)

A High Efficiency Stirling/Pulse Tube Refrigerator Working at 15−20 K Kongkuai Ying1,2 , Zhenhua Jiang1,2(B) , Shaoshuai Liu1 , and Yinong Wu1,2 1 Shanghai Institute of Technical Physics Chinese Academy of Sciences, Shanghai, China

[email protected] 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. Stirling/pulse tube refrigerator (SPR) is firstly proposed by Raytheon in 1999. The SPR has significant advantages in terms of high efficiency, long lifetime, and high compactness. The SPR features the ability of inter-stage cooling capacity adjustment by gas distribution. A two-stage gas-coupled SPR driven by a linear compressor working at 40 Hz is developed in our laboratory. The high efficiency SPR can obtain 2.5 W at 70 K plus 0.45 W at 20 K with 217 W electric input power (including displacer input power) or 0.22 W at 15 K plus 1.18 W at 65 K with 200 W compressor input power plus 16 W displacer input electrical power. With an electric input power of 214 W, the refrigerator can provide a cooling power of 0.35 W at 15 K. The no-load temperature is 11.4 K at the second stage while the first stage temperature is 57 K with 180 W compressor input power plus 18 W displacer input electrical power. Keywords: Stirling refrigerator · Pulse tube refrigerator · High efficiency · Working parameters · Phase difference · Experiment

1 Introduction The Stirling/pulse tube refrigerator (SPR) has significant advantages such as high efficiency, long lifetime, and high compactness. The 15–20 K two stages Stirling/pulse tube refrigerator can be used to cool space infrared detectors and optical circuits such as arsenic doped silicon (Si:As) focal planes long-wave infrared detectors and space-borne astronomical receiving systems. The SPR has been researched by Raytheon for more than twenty years, and there are some mature products have been developed [1, 2]. Raytheon’s Stirling/pulse tube cryocoolers have included 55 K/10 K for the low-temperature application, 85 K/35 K for the high-capacity, and 110 K/58 K for the mid-capacity. The main structures of each Stirling/pulse tube refrigerator are identical, the difference is the size of the Stirling regenerator (the displacer), pulse tube, and second stage regenerator. In 2014 Raytheon developed a Compact Inline RSP2 (CI-RSP2) [3] allowing the moving parts for the compressor and the displacer to be consolidated onto a common axis which can markedly decrease the vibration. The CI-RSP2 can obtain 0.48 W at 40 K plus 6 W at 80 K with 171 W input power or 0.25 W at 20 K plus 4.5 W at 70 K with 208 W input power. In 2017 Raytheon introduced the latest performance of the Low-Temperature Raytheon Stirling/Pulse Tube 2stage (LT-RSP2) refrigerator. The maximum capacities at each stage for the LT-RSP2 were approximately 6 W at 55 K and 0.47 W at 10 K [4]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 576–582, 2023. https://doi.org/10.1007/978-981-99-6128-3_74

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Guo et al. [5] from Zhejiang University studied the thermodynamic characteristics of the SPR and explained the effect of the second stage on the first stage and reveal the condition that the SPR can shift in-stage cooling capacity by an ideal thermodynamic model. The study pointed out that SPR’s ability to shift cooling capacity is conditional and proposed a critical value limiting the phase range, where SPR can shift cooling capacity between stages. Based on the above ideal thermodynamic model, Cao et al. [6] introduced five working states of the SPR according to whether the second pulse tube stage exchanger of the hot end of the pulse tube is precooled by the Stirling stage. The theoretical study from the research group of Professor Zhihua Gan (Zhejiang University) is helpful to the design and optimization of the SPR. Over the last several years, the Shanghai Institute of Technical Physics has done a lot of research on the SPR. Our laboratory provides long lifetime, low vibration, and high efficiency cooling for space applications and has abundant experience in designing and fabricating Stirling cryocoolers and pulse tube cryocoolers. Some typical types of Stirling and pulse tube refrigerators have been designed and fabricated such as a single stage pulse tube cryocooler which can provide 3 W at 40 K with an input power of 225 W [7], a pneumatic Stirling cryocooler obtaining 3 W at 80 K with input power of 70 W, a Stirling cryocooler obtaining 2.3 W at 60 K with 67.5 W input power [8], 1.1 W at 35 K with 70 W input power for two-stage Stirling cryocooler. In 2019, a SPR working at 35 K/85 K [9] has been designed and tested in our laboratory. The SPR obtains cooling capacities of 1.06 W at 30 K plus 4.46 W at 80 K with the input electric power of 261.2 W and 1.16 W at 35 K plus 7.25 W at 85 K with the input electric power of 234.6 W (including displacer input power). In 2020, Liu et al. [10, 11] proposed a theoretical model of the SPR and analyzed the thermodynamic characteristics of the SPR theoretically, meanwhile a one-dimensional thermoacousticbased numerical model had been adopted to optimize the structure parameters and the operating parameters. In the following, a SPR designed by our institute is tested, which can obtain a cooling capacity of 2.5 W at 70 K plus 0.45 W at 20 K with 217 W electric input power. The test results of the SPR are compared with reported two-stage refrigerators, and it proves that the SPR has high performance among 15−20 K. Meanwhile, the effects of the displacer phase which are the critical parameters of the SPR on the cooling capacity shift are discussed by experiments.

2 SPR Structure The structure of the SPR includes the main compressor, connect tube, ambient heat exchanger, displacer (filled with 350 # stainless steel screen), expansion chamber, first stage heat exchanger, second stage regenerator, second stage heat exchanger, pulse tube, warm end of pulse tube heat exchanger, inertance tube and gas reservoir. The SRP is tested in a vacuum enclosure and the photo of the cold finger is shown in Fig. 1. The second pulse tube stage is gas-coupled with the first Stirling stage. The second pulse tube is coaxial which makes the SPR more compact. There are two displacement transducers and one pressure sensor to measure the amplitude and phase of the displacer and pressure wave. The temperature at each stage

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Fig. 1. Photo of the SPR without inertance tube and gas reservoir.

heat exchanger, low-temperature inertance tube, and gas reservoir is measured by temperature sensors. The SPR is driven by a main compressor developed by our lab which is the dual opposed piston, flexure bearing with a long lifetime and very low exported vibration. The maximum displacement of the compressor piston is 6mm and the diameter of the compressor piston is 35 mm. The operating frequency of the SPR is 40 Hz. There are also some support systems such as a water-cooling system, vacuum system, power supply, and data acquisition system. The displacer is controlled by an expansion motor, the maximum displacement of the displacer is 4 mm. The SPR’s cold finger is arranged with a gas coupling structure. The ambient heat exchanger is cooled by the water circulating system which is set the temperature to room temperature. The main compressor and displacer are driven by two power supplies. The phase difference between the piston of the main compressor and the displacer is controlled by a signal generator. The inertance tube and the reservoir, used as the phase shifters in the second stage, were cooled by the inter-stage heat exchanger, which improved their phase shifting capability.

3 Experiments The experimental results of the Stirling/pulse tube refrigerator working at 15−20 K are shown in Table 1. The SPR can obtain 0.45 W at 20 K plus 2.5 W at 70 K with 200 W compressor input power plus 17 W displacer input electrical power or 0.22 W at 15 K plus 1.18 W at 65 K with 200 W compressor input power plus 16 W displacer input electrical power. With an electric input power of 200 W (compressor) plus 14 W (displacer) the refrigerator can provide a cooling power of 0.35 W at 15 K. In this condition, the no-load temperature of the first stage is 56 K which is lower than the target first stage temperature. The no-load second temperature is 11.4 K at the second stage while the first stage temperature is 57 K with 180 W compressor input power plus 18 W displacer input electrical power. The displacer input power is a little large because the impedance matching of the displacer is not well designed. The displacer input power can be decreased by adjusting the diameter of the displacer rod and total stiffness of plate spring to achieve pneumatic matching. The cooling powers, temperatures, and input powers of the two-stage cryocoolers reported by SITP, Air Liquide, Technical Institute of Physics and Chemistry, and Raytheon are compared in Table 2. The 15 K pulse tube cooler under development by

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Table 1. Experimental results of the SPR. Compressor Input power (W)

Displacer 1st input power temperature (W) (K)

2st temperature (K)

1st stage cooling capacity (W)

2st stage cooling capacity (W)

Phase difference (°)

200

17

70

20

2.5

0.45

80

200

14

56

15

0

0.35

70

200

16

65

15

1.18

0.22

60

180

18

57

11.4

0

0

75

Air Liquide [12], in collaboration with CEA/SBT and Thales Cryogenics BV, was originally intended to address the He3 J-T precooling requirements of the XMS instrument for the IXO mission (predecessor of ATHENA) can 300 mW at 15 K plus 5 W at 100 K with 300 W electrical input power. The CI–RSP2 [3] developed by can obtain 4.5 W at 70 K plus 0.25 W at 20 K with 208 W input power and Low-temperature RSP2 can obtain 4.5 W at 55 K plus 0.25 W at 10 K with 440 W input power. The gas-coupled High-frequency pulse tube refrigerator (HPTR) [13] developed by the Technical Institute of Physics and Chemistry is driven by a linear compressor using an inertance tube and gas reservoir as two stages phase shifter. Table 2. Several two-stage refrigerators working at 15−20 K. Research institution

Type of cryocooler Cooling power (W)

Heat rejection temperature (K)

Input power (W)

Raytheon

CI–RSP2

4.5 W @ 70 K 0.25 W @ 20 K

300

208

Low-temperature RSP2

4.5 W @ 55 K 0.25 W @ 10 K

300

440

Air Liquide

2-stage 15 K pulse tube cooler

5 W @ 100 K 0.3 W @ 15 K

288

300

Technical Institute of Physics and Chemistry

High-frequency pulse tube refrigerator (HPTR)

0.4 W at 15 K

−

400

SITP

15−20 K SPR

2.5 W @ 70 K 0.45 W @ 20 K

300

217

1.18 W @ 65 K 0.22 W @ 15 K

300

216

Figure 2 illustrates the ability of the SPR to load shift between the stages by adjusting the phase difference between the piston of the main compressor and the displacer. With

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the increase of the phase difference, the cooling capacity of the first stage increases, and the cooling capacity of the second stage decreases with the electrical compressor input power of 200 W. The main reason is that with the change of the displacer phase, the acoustic power distribution in the refrigerator has changed. When the phase difference is 85°, the SPR can obtain 0.4 W at 20 K plus 2.94 W at 70 K. When the phase difference is 90°, the SPR can obtain 0.36 W at 20 K plus 3.05 W at 70 K. By actively controlling the phase difference, the SPR can actively change the in-stage cooling capacity which allows for the ratio of cooling capacities at the two stages to be altered via a simple software command of the phase difference without a significant loss of overall efficiency. However, the SPR’s capability of shifting inter-stage cooling capacity is conditional, which exists a critical value. The cooling capacity of both stages decreases simultaneously as the displacer phase increases when the phase of the displacer exceeds the critical value. As shown in Fig. 2, because the critical value of displacer phase is larger than 90°, the first stage cooling capacity keeps increasing when the displacer phase within 90°. Meanwhile, the displacer input power also changes with an increase in the displacer phase, due to the changes of gas force. In the experiment, when the displacer phase increases from 80° to 90°, the displacer input power increases from 17 W to 21 W. The displacer input power is mainly used to adjust the displacer to reach the proper displacement and phase. Most of the displacer input power is consumed in the mechanical damping and coil resistance of the displacer, and some of the displacer input power contributes to improving cooling efficiency because under certain operating conditions the displacer recovered acoustic work may be greater than the cold end expansion work.

Fig. 2. The load shift of the SPR with different phase differences.

The capability to load shifting for the SPR provides benefits during the design of the system for a typical mission. Normally during the research and development of the refrigerator, the size of the structure and operating parameters are often revised. The cooling capacity of each stage of the actual refrigerator is different from the simulation model. The load shifting capability allows the SPR to change the in-stage cooling capacity to meet the requirement of the overall system.

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4 Conclusions Stirling/pulse tube refrigerator (SPR) has dominant characteristics of high efficiency, long lifetime, and high compactness because of the unique configuration combining the best features of the Stirling refrigerator and pulse tube refrigerator. A SPR has been designed and fabricated in our laboratory, which can obtain a cooling capacity of 2.5 W at 70 K plus 0.45 W at 20 K with 217W electric input power (including displacer input power). With an electric input power of 214W, the refrigerator can provide a cooling power of 0.35 W at 15 K. The experimental results confirm that the 15−20 K SPR has the ability to load shift between their first and second stages by changing the phase difference between the piston of the main compressor and the displacer. When the phase difference increases from 80 degrees to 90 degrees the cooling capacity of the first stage increases from 2.5 W to 3.05 W at 70 K, meanwhile the cooling capacity of the second stage decreases from 0.45 W to 0.36 W at 20 K. The no-load temperature is 11.4 K at the second stage while the first stage temperature is 57 K with 180 W compressor input power plus 18 W displacer input electrical power. Furthermore, the optimization of the displacer’s aerodynamic matching will be carried out in subsequent studies to reduce the input power of the displacer. Acknowledgments. This work is supported by the Hundred Talents Program of the Chinese Academy of Sciences, the National Natural Science Foundation Projects (51806231), and the Strategic Priority Research Program of Chinese Academy of Sciences (XDB35000000).

References 1. Schaefer, B.R., Bellis, L., Ellis, M.J., et al.: Raytheon low temperature RSP2 cryocooler airborne testing. In: AIP Conference Proceedings, vol. 1573, pp. 1318–1324 (2014).https:// doi.org/10.1063/1.4860859 2. Kirkconnell, C.S., Hon, R.C., Kesler, C.H., et al.: Raytheon stirling/pulse tube cryocooler development. In: AIP Conference Proceedings, pp. 909–916 (2008). https://doi.org/10.1063/ 1.2908688 3. Schaefer, B.R., Bellis, L., Ellis, M.J., et al.: Raytheon’s next generation compact inline cryocooler architecture. In: AIP Conference Proceedings, vol. 1573, pp. 371–377 (2014).https:// doi.org/10.1063/1.4860725 4. Conrad, T., Schaefer, B., Bellis, L., et al.: Raytheon long life cryocoolers for future space missions. Cryogenics 88, 44–50 (2017) 5. Guo, Y., Chao, Y., Wang, B., et al.: The thermodynamic characteristics of a Stirling/pulse tube hybrid cryocoole. Cryogenics 96, 133–143 (2018) 6. Chao, Y., Guo, Y., Wang, Y., et al.: Thermodynamic analysis of the working states of the Stirling/pulse tube hybrid cryocooler. Appl. Therm. Eng. 170, 115024 (2020) 7. Zhang, A., Wu, Y., Liu, S., et al.: High-efficiency 3 W/40 K single-stage pulse tube cryocooler for space application. Cryogenics 90, 41–46 (2018) 8. Wu, Y., Liu, E., Jiang, Z., et al.: The integrated cryogenic system for the atmospheric vertical interferometric detector on FY-4 satellite. Tri-Technol. Device Refrig. (TTDR). (2016). https://doi.org/10.1117/12.2223843 9. Liu, B.Q., Jiang, Z.H., Ying, K.K., et al.: A high efficiency Stirling/pulse tube hybrid cryocooler operating at 35 K/85 K. Cryogenics 101, 137–140 (2019)

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10. Liu, B., Jiang, Z., Ying, K., et al.: Theoretical model of a stirling/pulse tube hybrid refrigerator and its verification. Appl Therm. Eng. 189, 116587 (2021) 11. Liu, B., Jiang, Z., Ying, K., et al.: Numerical and experimental study on a Stirling/pulse tube hybrid refrigerator operating around 30 K. Int. J. Refrig. 123, 34–44 (2021) 12. Chassaing, C., Butterworth, J., Aigouy, G., et al.: 15K pulse tube cooler for space missions. Cryocoolers 18, 27–32 13. Wu, X., Chen, L., Liu, X., et al.: An 80 mW/8 K high-frequency pulse tube refrigerator driven by only one linear compressor. Cryogenics 101, 7–11 (2019)

A High-Efficiency 80 K Coaxial Pulse Tube Refrigerator with Both Active Phase Shifting and Work Recovery Hejun Hui1,2 , Jiantang Song1 , Shaoshuai Liu1(B) , Lei Ding1 , Hanzhen Xiang1 , Zhenhua Jiang1,2 , and Yinong Wu1,2(B) 1 Shanghai Institute of Technical Physics, Chinese Academy of Sciences, 500 Yutian Road,

Shanghai 200083, China {liushaoshuai,wyn}@mail.sitp.ac.cn 2 University of Chinese Academy of Sciences, No. 19A Yuquan Road, Beijing 100049, China

Abstract. Pulse tube refrigerator is abundantly used in military and civilian areas owing to the cold head without moving parts. The phase difference between mass flow and pressure wave in the cold finger significantly influences the efficiency of the pulse tube refrigerator. Most pulse tube refrigerators adopt passive phase shifters, such as the inertance tube, double inlet, and passive displacer, which cannot acclimate the phase difference with the variation of cooling temperature. This paper proposes a single-stage pulse tube refrigerator with high-efficiency cooling in the temperature range near 80 K by the active displacer with work recovery. By actively altering the phase difference in an extensive cooling temperature range of 80 K to 120 K, the cold finger of the refrigerator is capable of efficiently providing a considerable cooling capacity. And the active displacer with work recovery can convert the energy dissipated initially as heat into acoustic power. This part of the acoustic power enters the cold finger again to enhance efficiency. The experiment results also show that the refrigerator could provide a cooling capacity of 20 W at 80 K when the input power of the compressor is 256.6 W and 5.1 W input power of the phase shifter, and the phase shifter recovers 17 W of acoustic power concurrently, with the relative Carnot efficiency reaching 21%. Keywords: Pulse tube refrigerator · Active phase shifter · Energy recovery · Cooling performance · Experiment

1 Introduction The coaxial type pulse tube refrigerator, which can be easily coupled with the low temperature heat load, is widely used in various civil and military applications due to the slight vibration output [1]. There is an alternating pressure wave and mass flow in the cold finger. The phase difference distribution between the pressure wave and mass flow in the cold finger significantly influences the refrigeration performance. The pulse tube refrigerator controls the phase distribution within the cold finger by introducing a phase shifter at the hot end of the pulse tube. Many types of phase shifter © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 583–589, 2023. https://doi.org/10.1007/978-981-99-6128-3_75

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structures are proposed to meet the phase shift needs of the refrigerator. The phase shifters commonly used are mainly divided into passive and active types. The passive phase shifters include inertance tube type, double-inlet type, orifice type, piston-type without driving force, etc. The passive phase shifter features the phase shift capability determined by its structural parameters and the operating parameters of the refrigerator. The active phase shifter means that the phase difference can be actively adjusted. The active phase shifter generally used is the piston type phase shifter with motor force, which realizes active phase adjustment by controlling the amplitude and phase of the piston movement. Abolghasemi built a coaxial pulse tube refrigerator with an active displacer and work recovery. The refrigerator with an active displacer is more efficient than using an inertance tube as the phase shifter [2]. When the working condition of the refrigerator changes, the active phase shifter achieves high-efficiency cooling within a wide cooling temperature range and broad cooling capacities by actively adjusting the phase distribution in the cold finger. If there is no work recovery, the acoustic expansion work at the hot end of the pulse tube cannot be recovered. Therefore, the ideal cooling efficiency of the pulse tube refrigerator is not Carnot efficiency but the Kittele efficiency, η = Tc /Th , where Tc is the cooling temperature and Th is the hot end temperature [3]. In 2010, to improve the cooling efficiency, Zhu gives a nodal numerical analysis method for the warm gasdriven displacer and analyzes the effect of the displacer piston motion state on the cooling efficiency [4]. And in 2015, Wang proposed a pulse tube refrigerator with a warm displacer, which improves cooling efficiency by recovering acoustic work at the hot end of the pulse tube [5]. The active displacer with work recovery recovers the acoustic expansion power from the hot end of the pulse tube and transfers the recovered power to the cold finger, which is contributed to improving the cooling efficiency. The active power recovery phase shifter is proposed to combine active phase control and acoustics power recovery advantages. The active displacer with work recovery can actively adjust the phase difference and recover the acoustic work, enabling the pulse tube refrigerator to provide large cooling capacities in broad temperature zones.

2 Structure Design Figure 1 displays the structure of the pulse tube refrigerator with the active displacer. In order to reduce the vibration output, the compressor with a moving coil motor is a double piston opposed structure, there is a piston in each of the two cylinders in the compressor, and the pistons move in opposite states. The diameter of the compressor piston is 38 mm. The cold finger of the refrigerator is the coaxial type, which has a convenient coupling with the load. The active displacer with work recovery is also a dual-piston structure with drive motors. There are three chambers in the phase shifter, including a phase shift chamber and two power recovery chambers. The elaborate phase shifter allows the refrigerator to cool efficiently over a wide temperature range. When the driving force from the motor of the phase shifter is removed, the piston is driven entirely by gas force and can be called the passive displacer. The phase shift capability of the passive displacer is determined by spring stiffness, damping, and kinetic mass, but the passive displacer is still capable of acoustic power recovery. When the

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Fig. 1. The design of the refrigerator

operating conditions of the refrigerator change, there is no motive force to shift the phase distribution within the cold finger actively, which can cause a reduction in cooling efficiency.

3 Numerical Calculation With our many years of research experience in the pulse tube refrigerator field, the onedimensional numerical calculation model has been developed based on the conservation of mass, momentum, and energy equations [6]. Figure 2 and Fig. 3 display the cooling performance variation with phase   difference ˙ (T0 − Tc ) , where Q˙c is and amplitude. The definition for rCOP is rCOP = Q˙c ∗ T0 / W ˙ is the acoustic power, T0 is the ambient temperature, and Tc is the cooling capacity, W the cooling temperature. At cooling temperatures of 80 K, 100 K, and 120 K, the piston motion state significantly impacts the cooling performance. Keeping the amplitude of the compressor piston and the phase shifter piston constant. When the phase difference between the phase shifter piston and compressor piston increased from 40° to 70°, the cooling capacity increased slowly, mainly due to increased acoustic power into the cold finger. However, the cooling efficiency increases first and then less, and all three cooling temperatures have the highest efficiency near 60°, where the highest efficiency reaches 30% when the refrigeration temperature is 100 K. Meanwhile, the calculations show that the acoustic power recovered through the power recovery chamber increases linearly as the phase difference of the phase shifter increases from 40 to 70°. As the amplitude of the phase shifter piston increases from 1.5 mm to 3.5 mm, the cooling capacity gradually increases when the cooling temperature is 100 K and 120 K. Although the input acoustic power keeps increasing, the cooling capacity in 80 K first increases, and then decreases, results in a maximum value occurs. The relative Carnot efficiency increases and decreases in all three temperature regions. The optimum amplitude of the piston at the highest efficiency is 2.7 mm at a cooling temperature of 80 K. However when the cooling temperature is 100 K, the optimum amplitude is 3.1 mm, which means that the optimum amplitude shifts with the change in cooling temperature. With the support of the active displacer with work recovery, the piston movement can be actively adjusted, which allows the refrigerator to achieve high-efficiency cooling in a wide range of temperatures and cooling capacity.

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Fig. 2. Simulation results of the cooling performance vs. the phase difference

Fig. 3. Simulation results of the cooling performance vs. the amplitude of shifter piston

4 Experiment 4.1 Experiment Configuration The prototype we have developed of the refrigerator with an active displacer with work recovery is shown in Fig. 4. The signal generator can alter the input voltages of the compressor and phase shifter, then influence the piston movements of the compressor and phase shifter. A platinum resistance temperature sensor and a ceramic heater are mounted on the cold head. And the cold head is covered with multiple layers of thermal insulation to reduce heat loss. The acoustic power recovered in the experiment can be calculated by measuring the pressure wave and dis-placement of the displacer piston flow in the work-recovery chamber.

Fig. 4. The prototype of refrigerator

4.2 Experimental Results The cool-down curve of the refrigerator with the active displacer with work recovery is shown in Fig. 5, the cold head of the refrigerator is capable of cooling down from room temperature to 40 K in 24 min, and it takes about 27 min to cool down to 39.5 K.

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Fig. 5. Cool-down curve

Keeping the cooling temperature constant at 80 K and exploring the effect of piston motion of the phase shifter, the experimental results are shown in Fig. 5 and Fig. 6. The experimental results show that the amplitude and phase of the phase shifter piston affect the acoustics power conversion efficiency of the compressor and the phase distribution within the cold finger, which in turn has an impact on the refrigeration performance. Figure 6 shows that keeping the input electrical power 300 W and the piston amplitude of phase shifter 2.5 mm constant, the phase difference increases from 45 to 65°, the cooling capacity increases from 19.9 W to 20.9 W, and then decreases, the cooling capacity changes smaller at 50 to 55°. Figure 7 shows that by retaining the phase difference constant at 50°, the cooling capacity increases first to 20.9 W and then decreases to 20.4 W. And the rCOP calculated by acoustic power first increases from 25.5% to 26.2% and then decreases to 25.9%. Figures 6 and 7 show that the simulated calculation results are in trend with the experimental results, which verifies the accuracy of the numerical model and the research analysis.

Fig. 6. Experiment results of the cooling performance vs. the phase difference

Fig. 7. Experiment results of the cooling performance vs. the amplitude of the shifter piston

With the active displacer with work recovery, the refrigerator is capable of adapting to the changing working conditions of the cold head. Figure 7 shows the high cooling efficiency at multiple cooling temperatures and cooling capacities. Keeping the input power 300 W constant, the cooling capacity is 20.9 W, 31.7 W, and 41.8 W when the cold head temperature is 80 K, 100 K, and 120 K, respectively. The motion state of the

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phase shifter is adjusted during the experiments to find the highest efficiency, where efficiency is highest at 100 K. The relative Carnot efficiency is up to 21.7%, which matched the simulation results. The experimental results of the larger cooling capacity at a cooling temperature of 80 K is shown in Fig. 8. When the electrical power into the compressor is 459 W, and the hot end temperature is 300 K, the cooling capacity is 30.3 W (Fig. 9).

Fig. 8. Efficient cooling in multi cooling temperature areas and cooling capacities

Fig. 9. Efficient cooling in multi cooling capacities at 80 K

After analyzing the numerical calculation results and experimental results, the acoustics power is lost more in the connecting tubes. Therefore the structural parameters of the connecting tubes are optimized, the motion state of the phase shifter piston is also adjusted, and a more satisfactory cooling performance is obtained, as Table 1 is shown. The total input electrical power is 261.7 W, including 5.1 W into the phase shifter, and the cooling capacity is 20 W when the cooling temperature is 80 K, with a relative Carnot efficiency of 21%. The measurement results show that the phase shifter simultaneously recovers 17 W of acoustic power. Table 1. The cooling performance after optimizing Parameters

Value

Cold head temperature

80 K

Cooling capacity

20 W

Electrical power into compressor

256.6 W

Electrical power into active work recovery shifter

5.1 W

Reject temperature

300 K

Relative Carnot efficiency (electrical power)

21%

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5 Conclusion Committed improve the cooling efficiency of the coaxial pulse tube refrigerator working at 80 K, we proposed a phase shifter that combines active phase shift and work recovery. The numerical simulations and experiments are conducted and analyzed. The active displacer with work recovery enables the refrigerator to provide highly efficient cooling in multiple temperature areas (from 80 K to 120 K) and numerous cooling capacities (from 20 W to 40 W). The experimental results show that the refrigerator achieved the lowest no-load temperature of 39.5 K. And a relative Carnot efficiency of 21% is obtained at a cooling capacity of 20 W when the cooling temperature is 80 K. Some optimizing work is on the way to achieve a higher cooling efficiency, like optimizing the structural parameters of the phase shifter and improving the acoustic power conversion efficiency of the compressor. Acknowledgments. This work is supported by the Hundred Talents Program of the Chinese Academy of Sciences, the National Natural Science Foundation Projects (51806231), the Strategic Priority Research Program of Chinese Academy of Sciences (XDB35000000).

References 1. Radebaugh, R.: Development of the pulse tube refrigerator as an efficient and reliable cryocooler (2000) 2. Abolghasemi, M.A., Rana, H., Stone, R., et al.: Coaxial Stirling pulse tube cryocooler with active displacer. Cryogenics 111, 7 (2020) 3. Kittel, P., Kashani, A., Lee, J.M., et al.: General pulse tube theory. Cryogenics 36(10), 849–857 (1996) 4. Zhu, S., Nogawa, M.: Pulse tube stirling machine with warm gas-driven displacer. Cryogenics 50(5), 320–330 (2010) 5. Wang, X., Zhang, Y., Li, H., et al.: A high efficiency hybrid stirling-pulse tube cryocooler. AIP Adv. 5(3), 037127 (2015) 6. Liu, S.S., Chen, X., Zhang, A.K., et al.: Investigation of the inertance tube of a pulse tube refrigerator operating at high temperatures. Energy 123, 378–385 (2017)

Progress in and Outlook for Work Recovery Type Pulse Tube Refrigerator Shaowei Zhu1,2(B) 1 Institute of Refrigeration and Cryogenics, School of Mechanical Engineering,

Tongji University, 4800, Cao’an Road, Shanghai 201804, China [email protected] 2 Shanghai Key Lab of Vehicle Aerodynamics and Vehicle Thermal Management Systems, Tongji University, 4800, Cao’an Road, Shanghai 201804, China

Abstract. Pulse tube refrigerator is a cryocooler used in many applications. After the long-time development, the efficiency is still lower than that of Stirling refrigerator. And the performance is much poorer when refrigeration temperature is lower than 20 K. To overcome these problems, some researchers focus on the work recovery type pulse tube refrigerator, and fruitful results were got. Several work recovery type pulse tube refrigerators are discussed. Keywords: Pulse Tube refrigerator · Cryocooler · Stirling refrigerator

1 Introduction When we say pulse tube refrigerator or pulse tube cryocooler, it usually points to double inlet or inertance tube pulse tube refrigerator which are without work recovery, because the expansion power is changed to heat and is discharged to surrounding environment. Because its theoretical efficiency is low, and its phase shifting ability is limited, pulse tube refrigerator can hardly get over 20% efficiency within more than 30 years development. Work recovery type is one of the promising and efficient solutions for pulse tube refrigerator to reach higher efficiency. There are several kinds of work recovery type pulse tube refrigerators, such as double piston, displacer, series and step piston type. It has been gradually getting attention by several research groups, and fruitful results have been got.

2 Work Recovery Type Pulse Tube Refrigerator Figure 1 shows different types of pulse tube refrigerator. For work recovery type pulse tube refrigerator, the expansion power that usually dissipated in other types pulse tube refrigerators can be reused for generating additional cooling power. Pulse tube refrigerator can be considered as a special kind of Stirling refrigerator, the function of the expander in Stirling refrigerator can be used as a referential concept of work recovering in pulse tube refrigerator. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 590–596, 2023. https://doi.org/10.1007/978-981-99-6128-3_76

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Fig. 1. Schematic of Stirling refrigerator and pulse tube refrigerator (PTR) (a) Stirling refrigerator (b) Double piston PTR (c) Basic PTR (d)Orifice PTR (e) Double inlet PTR (f) Displacer PTR (g) Step piston PTR (h) Series PTR

Figure 1a is Stirling refrigerator, it includes compressor, aftercooler, regenerator, cold head exchanger, expander. The expander is rather long for keeping temperature gradient. If the expander is shorter and is put in room temperature, the cylinder part becomes pulse tube, then the refrigerator becomes double piston pulse tube refrigerator [1], as shown in Fig. 1b. The gas column in the long tube and next to the expander is equivalent of the extension of the solid piston, or it could be considered as a gas piston controlled by the solid piston. In addition, Fig. 1b can be seen as a Stirling refrigerator with big dead volume, with advantage of no moving part at low temperature, which is a big merit for manufacture and application. If the expander is removed and the cylinder is covered by a flange, the Stirling refrigerator becomes a basic pulse tube refrigerator, as shown in Fig. 1c. In this case, there is no expander controlling the motion of gas piston in the cylinder, its motion would be very poor. Consequently, the refrigeration efficiency is very low. Adding a capillary tube with buffer on the warm end flange, the basic pulse tube refrigerator becomes an orifice pulse tube refrigerator, as shown in Fig. 1d. The gas flow across the capillary generates irreversible loss, which is from the expansion power of the gas at the cold end of the pulse tube. The cooling power is largely increased. How much the irreversible loss is, how much gross cooling power will be. So, irreversible loss is important here. Adding a bypass between the compressor and the hot end of the pulse tube, it becomes double inlet pulse tube refrigerator, as shown in Fig. 1e. There were two opposite needle-like valves on the bypass line to control DC gas flow at the first experiment of the double inlet pulse tube refrigerator [2]. If the capillary diameter and length are optimized in Fig. 1d, it becomes inertance tube pulse tube refrigerator. Replacing the piston in Fig. 1b by displacer, it becomes displacer pulse tube refrigerator [3], as shown in Fig. 1f. Eliminating the orifice and buffer in the double inlet pulse tube refrigerator with step piston [4] and let the bypass become an inertance tube, it becomes a step piston type pulse tube refrigerator [5], that is shown in Fig. 1g. Using another cold head to replace the expander in Fig. 1b, series pulse tube refrigerator is got [6], as shown in Fig. 1h. The development history of the pulse tube refrigerator is greatly pushed forward by new inventions. The important inventions and associated key technology progresses are

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listed in Fig. 2. The pulse tube refrigerators shown in the center of the list have only one moving part, which are from basic type to series type pulse tube refrigerator. There are two moving parts in double piston type, displacer type and step displacer type pulse tube refrigerator [7] which is a new type of two-stage pulse tube refrigerator. The expansion power of active buffer pulse tube refrigerator is saved in buffers and reused for phase shifting, and this type pulse tube refrigerator was improved from check valve pulse tube refrigerator [1] and was further modified to 4-valve and 5-valve type. The no-load temperature of 60 K, 49 K, 42 K and 4 K was achieved by orifice and double inlet pulse tube refrigerator in the development history of pulse tube refrigerators. With the development of the double inlet pulse tube refrigerator, nodal analysis method, isothermal model and Lagrange method [2, 8] of pulse tube refrigerator were introduced respectively. The double orifice [9] is a method for controlling DC gas flow for getting 4 K. The series pulse tube refrigerator has several modifications including quarter wave type, cascade type, inertance piston type and inertance tube type [10–14]. Thermal acoustic cooler has no moving part, which is another research subfield including cryocooler and room temperature refrigerator.

Fig. 2. Tree of inventions and the associated key technology progresses of pulse tube refrigerators

3 Double Piston and Displacer Type The expander of the double piston pulse tube refrigerator should be controlled by crankshaft [15] or linear motor [16–19]. It is very difficult to get long life with crankshaft. By linear motor, the phase distribution in the cooler could be well controlled, so it is easier to get cooling temperature lower than 20 K. If the expansion power of the pulse tube is less than the loss including gas leakage loss from clearance seal, mechanical loss, iron and eddy current loss of the linear motor, input power is needed. For displacer pulse tube refrigerator, the displacer is driven by gas force (passive displacer) [20]. There are three kinds of displacer, rod less, with rod, and with linear motor (active displacer). The efficiency of 24.2% at 80 K was got by a rod less type displacer pulse tube refrigerator [21]. Displacer with rod was also confirmed to be efficient

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by experiment [22, 23]. Displacer with linear motor is more convenient by handling in experiment, and it is easy to get optimum displacer movement by power source [24]. Theoretically, gas driving displacer can make the same optimum performance as linear motor driving type does, and the circumstance exists at the point where the linear motor current is zero [25]. Due to the recovery of the expansion power, pulse tube refrigerator could work efficiently in near room temperature, and the efficiency of 20.8% was achieved at 170 K [26]. For micro pulse tube refrigerator [27], the expansion power is too small to let inertance tube oscillate sufficiently. The displacer may be an effective method to replace inertance tube. Other problems for micro pulse tube refrigerator include that how small the pulse tube could be, and how to evaluate the heat transfer effect on the pulse tube efficiency. It looks like that the heat transfer phenomenon is quite different from the traditional one and is difficult to find a mathematical law to describe [28]. The complicated heat transfer process makes the traditional 1D simulation difficult to simulate micro pulse tube accurately, and the heat transfer has big effect on the pulse tube efficiency. With more than 20 years development, multi-stage pulse tube refrigerator is still difficult to get high efficiency when temperature is less than 20 K. Step displacer is a new solution for it. The step displacer generates two volumes to recover the acoustic power and shift the phase angle, one for the first stage pulse tube and another for the second stage pulse tube. It can give any phase angle with any pressure ratio, frequency, and charge pressure. With step displacer [29, 30] at 20K, the relative Carnot efficiency of 6.5% was achieved. By double inlet, the relative Carnot efficiency was 3.85%. The efficiency has been increased effectively. Step displacer has two functions, one is for power distribution, another is for phase shifting [31]. A dead volume between displacer and pulse tube is added for separating the two functions. At optimum dead volume, 16.9 K was achieved [32, 33]. As a reversible refrigerator, displacer type can work as different modes of heat pump, heat engine and cold engine [34]. As a heat pump, it is easy to get high temperature difference. As a heat engine, there is no moving part at high temperature, which is a big improvement on Stirling engine. As a cold engine, it can use cold energy such as LNG to generate power.

4 Series Type The concept of series pulse tube refrigerator is quite simple [6]. Two or more pulse tube refrigerators are connected by series. The second pulse tube refrigerator is connected at the hot end of the first pulse tube refrigerator, to use the expansion power to generate additional cooling power. The quarter wave pulse tube refrigerator is a modification of series pulse tube refrigerator, the pulse tube is equivalent to an inertance tube, making big loss and the loss changes to heat for the cold heat exchanger [10]. In cascade pulse tube refrigerator [11, 12], an inertance tube is added between two pulse tube refrigerators. The inertance tube acts as phase shifter for the first pulse tube refrigerator, and transfers the expansion power to the second pulse tube refrigerator. And the efficiency is improved eventually. Furthermore, the inertance tube is replaced by an inertance piston for higher efficiency, in virtue of the loss of the inertance piston is much smaller than the inertance

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tube [13]. The efficiency of 24.5% at 130K was achieved. An interesting modification is the connection of two inertance tube pulse tube refrigerators [14]. The expansion power is partially consumed by the first inertance tube, and the remaining expansion power can be used as the input power of the second stage. Series type is not only efficient, but also has multi-cooling temperatures.

5 Step Piston Type The idea of using inertance tube to recover the expansion power has been in high attention for long time [35]. The step piston pulse tube refrigerator could get the function [5]. The step piston with step cylinder moves back and forth, generating the compression space and expansion space which are in phase with each other. The inertance tube shifts the phase of the pressure in the compression space and expansion space to be different. If the integration of the PV work in compression space is negative, it will be positive in expansion space. The expansion power is transferred from the hot end of the pulse tube through the inertance tube to the expansion space. The key point is how the expansion power is efficiently transferred through the inertance tube. Numerical simulation shows that the power transfer efficiency could be over 80% in some condition [36]. Compared with Stirling refrigerator, the working mechanism of step piston type is different. Its compression space and expansion space are in phase, but the pressures in the two spaces are not in phase. In addition, the step piston pulse tube refrigerator is a reversible refrigerator which has four operation modes as well [37].

6 Conclusion and Future Work Work recovery pulse tube refrigerator has many advantages compared to non-work recover type. High efficiency has been achieved by displacer and series types pulse tube refrigerator. Due to the good phase shifting ability and high theoretical efficiency of work recovery pulse tube cryocooler, it is possible to get 1 kW cooling power at 77 K with COP over 0.1, which is important for the application of high temperature superconductor cable. And getting high efficiency at refrigeration temperature range of 2–20 K is important for the applications such as deep space exploration, SQUID, single photon detector. It is possible to penetrate to vapor cycle field somehow in the future. Acknowledgments. The project is supported by National Natural Science Foundation of China (Number 52076151).

References 1. Mastubara, Y., Miyake, A.: Alternative methods of the orifice pulse tube refrigerator. In: Proceedings of the 5th International Cryocooler Conference, California, pp. 127–135 (1988) 2. Zhu, S.W.: Theory of pulse tube refrigerator and its important improvement: double inlet pulse tube refrigerator. Ph.D. thesis, Xi’an Jiaotong University (1990)

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3. Matsubara, Y.: Future trend of pulse tube cryocooler research. In: Zhang, L., Liang, L., Chen, G. (eds.) Proceedings of the Twentieth International Cryogenic engineering Conference, Beijing, China, 2004, p. 189–196. Elsevier Ltd. (2005) 4. Zhu, S., Peiyi, W., Chen, Z.: Double inlet pulse tube refrigerators: an important improvement. Cryogenics 30, 514–520 (1990) 5. Zhu, S.: Step piston pulse tube refrigerator. Cryogenics 64, 63–69 (2014) 6. Zhu, S., Shin, K., Nogawa, M., Inoue, T.: Pulse tube refrigerator. US6389819 (1999) 7. Zhu, S.: Pulse tube regenerative machine. JP2005–93277 (2005) 8. Zhu, S., Chen, Z.: Enthalpy flow rate of a pulse tube in pulse tube refrigerator. Cryogenics 38, 1213–1216 (1998) 9. Chen, G., et al.: Experimental study on a double-orifice two-stage pulse tube refrigerator. Cryogenics 37, 271–273 (1997) 10. Swift, G.W., Gardner, D.L., Backhaus, S.N.: Quarter-wave pulse tube. Cryogenics 51, 575– 583 (2011) 11. Wang, L.Y., Wu, M., Sun, X., Gan, Z.H.: A cascade pulse tube cooler capable of energy recovery. Appl. Energy 164, 572–578 (2016) 12. Wang, L.: Acoustic- mechanical-electrical coupling and an approach to Carnot efficiency of pulse tube refrigerator. Ph.D thesis, Zejiang University (2016) 13. Xu, J.Y., Hu, J.Y., Hu, J.F., Zhang, L.M., Luo, E.C., Gao, B.: Experimental study of a cascade pulse tube cryocooler with a displacer. Cryogenics 95, 69–75 (2018) 14. Chen Mengjia, X., Sheng, D.H., Shaowei, Z.: Series inertance tube pulse tube refrigerator. Int. J. Refrig 132, 217–221 (2021) 15. Ishizaki, Y., Ishizaki, E.: Experimental performance of modified pulse tube refrigerator below 80 K down to 23 K. In: International Cryocooler Conference, pp. 140–146 (1992) 16. Masuyama, S., Kim, Y.-H., Park, S.-J., Hong, Y.-J., Kim, H.-B., Lee, S.-H.: Experimental research of Stirling type pulse tube refrigerator with an active phase control. Cryogenics 46, 385–390 (2006) 17. Lewis, M.A., Bradley, P.E., Radebaugh, R.: Experiments with linear compressor for phase shifting in pulse tube cryocoolers. AIP Conf. Proc. 1434, 1600–1607 (2012) 18. Yang, S., Wu, Y.N., Zhou, Z.P., Jiang, Z.H., Zhang, A.K.: Simulation and experimental analysis of pulse tube refrigerator with an active phase controller. Cryocooler 19 (2016) 19. Duval, J.M., Charles, I., Coynel, A., Gauthier, A.: Development of 15 K pulse tube cold fingers for space applications at CEA/SBT. Cryocooler 16, 45–50 (2008) 20. Zhu, S.W., Nogawa, M.: Pulse tube Stirling machine with warm gas-driven displacer. Cryogenics 50, 320–330 (2010) 21. Wang, X.T., et al.: A high efficiency hybrid Stirling-pulse tube cryocooler. AIP Adv. 5, 037127 (2015) 22. Shi, Y., Zhu, S.: Experimental investigation of pulse tube refrigerator with displacer. Int. J. Refrig. 76, 1–6 (2017) 23. Wang, M.J., Xu, H.J., Lin, Y.Z., Zhang, Z., Zhu, S.W.: Experimental investigation of pulse tube refrigerator with rod type displacer as phase shifter. Int. J. Refrig. 93, 47–51 (2018) 24. Abolghasemi, M.A., Rana, H., Stone, R., Dadd, M., Bailey, P., Liang, K.: Coaxial Stirling pulse tube cryocooler with active displacer. Cryogenics 111, 103143 (2020) 25. Zhu, S., Guo, Z.: Comparison of solid phase shifters of pulse tube refrigerator. In: Abstracts of CSSJ( Cryogenics and Superconductivity Society of Japan ) Conference, vol. 102, p. 21 (2021) 26. Chen, X., Ling, F.G., Zeng, Y.P., Wu, Y.N.: Investigation of the high efficiency pulse tube refrigerator with acoustic power recovery. Appl. Therm. Eng. 159, 113904 (2019) 27. Guo, Z.M., Pfotenhauer, J.M., Miller, F., Zhu, S.W.: A research of micro-pulse tube cryocooler with displacer phase shifter. Appl. Therm. Eng. 205, 117995 (2022)

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28. Guo, Z.M., Zhu, S.W.: Investigation about basic heat transfer law in the micro pulse tube. Int. J. Refrig 134, 139–145 (2022) 29. Pang, X.: Study on the multi-stage hybrid Stirling-type pulse tube cooler operating at the liquid hydrogen temperature. Ph.D thesis, Technical Institute of Physics and Chemistry, Chinese Academy of Science (2018) 30. Pang, X.M., Wang, X.T., Daia, W., Li, H.B., Wu, Y.N., Luo, E.C.: Theoretical and experimental study of a gas-coupled two-stage pulse tube cooler with stepped warm displacer as the phase shifter. Cryogenics 92, 36–40 (2018) 31. Zhu, S., Lin, Y.: Fundament of input power distribution and phase shifter functions of a step displacer type two-stage pulse tube refrigerator. Int. J. Refrig 113, 31–37 (2020) 32. Lin, Y.: Basic study on step displacer type two-stage pulse tube refrigerator, Ph.D thesis, Tongji University (2020) 33. Lin, Y., Guo, Z., Guo, Z., Zhu, S.: Experimental investigation of the connecting tube effect on a step displacer type two stage pulse tube refrigerator. Appl. Therm. Eng. 173, 115229 (2020) 34. Zhu, S.: Four operating modes of pulse tube heat engine with displacer at room temperature. J. Eng. Thermophys. 37, 232–234 (2016) 35. Swift, G.W., Gardner, D.L., Backhaus, S.: Acoustic recovery of lost power in pulse tube refrigerator. J. Acoust. Soc. Am. 105(2), 711 (1999) 36. Lin, Y., Zhu, S.: Numerical investigation of the new phase shifter for pulse tube refrigeratorinertance tube combining with step-piston. Int. J. Refrig 97, 42–48 (2019) 37. Zhu, S.W.: Four operation modes of a pulse tube machine with a step piston compressor. Energy 100, 190–198 (2016)

A 910 mW@15 K Thermal-Coupled Pulse Tube Cryocooler with Active Phase Shifter Wang Yin1,2 , Hejun Hui1,2 , Wenting Wu1 , Jiantang Song1 , Shaoshuai Liu1(B) , Zhenhua Jiang1,2 , and Yinong Wu1,2(B) 1 Shanghai Institute of Technical Physics, Chinese Academy of Sciences, 500 Yutian Road,

Shanghai 200083, China {liushaoshuai,wyn}@mail.sitp.ac.cn 2 University of Chinese Academy of Sciences, No. 19A Yuquan Road, Beijing 100049, China

Abstract. A two-stage Stirling-type pulse tube cryocooler (SPTC) operating in 15 K is developed in our lab for cooling infrared detectors and per-cooling helium JT cryocooler. The SPTC adopts thermal-coupled structure and each stage adopts a coaxial structure for compactness and convenience. For easy adjustment, the driven compressor of the SPTC is designed as two independent linear compressors. The first stage uses an inertance tube and reservoir as a phase shifter, while the second stage cold finger uses an active warm displacer (AWD) as phase shifter to maximize the cooling performance. Through reasonable design of structure and operating parameters, the SPTC achieves better performance at 15 K. The experiment also investigated the influence of the frequency. With the pre-cooling temperature of 80 K, the second stage cold finger can obtain 470 mW at cooling temperature of 15 K with a total electric input power of 250 W. It can also obtain 910 mW at cooling temperature of 15 K with total input power of 386 W. The cryocooler has obtained a large cooling capacity and high efficiency at 15 K to meet the design and application requirements. Keywords: Pulse tube cryocooler · 15 K · Active phase shifter · Cooling performance · Experiment

1 Introduction Nowadays, the compact cryocoolers working at 15 K are eagerly needed for space applications such as cooling the infrared detectors, low-temperature electronics, and Pre-cooling helium Joule Thomson cryocooler [1]. Among them, the Stirling-type pulse tube cryocooler (SPTC) are widely used because of high reliability, compact structure, and high efficiency [2]. SPTC always adopts a multi-stage structure for temperature around 15 K which causes a decrease in the efficiency [3, 4]. Current research needs to pay attention to improving the performance and efficiency of the SPTC. Recently, a lot of studies on two-stage SPTC have been reported. Based on the basic learning on SPTC, the phase shifter is an extremely important part of SPTC. Various phase shifters have been designed to improve performance such as inertance tube and © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 597–604, 2023. https://doi.org/10.1007/978-981-99-6128-3_77

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reservoir in low temperature, double-inlet, multi-bypass, active displacer, and passive mass-spring oscillator. For example, Pang reported a thermal-coupled two-stage SPTC in 2016 which used an inertance tube as phase shifter. It could end up with a no-load temperature of 15.35 K and obtained 0.73 W at 20 K with input electric power 540 W [5]. Wu reported a gas-coupled two-stage SPTC in 2018 which used a double-inlet, an inertance tube and gas reservoir as phase shifter for each stage. It achieved a no-load temperature of 11.5 K under the input electric power of 200 W, and the cooling capacity of 0.15 W could be obtained at 15 K. Input electric power increased to 400 W, the cooling capacity of 0.35 W could be obtained at 15 K [6]. Zhu compared three modes of phase shifters in 2018. It was concluded that two-stage SPTC obtained a minimum temperature of 18.8 K when an active displacer was used at second stage [7]. Air Liquide Advance Technologies in France designed and manufactured a thermalcoupled two-stage SPTC in 2014 [8]. It used an active displacer as phase shifter in the second stage which has an excellent performance at 15 K. When the total input electric power is 300 W, the cryocooler can obtain 0.416 W at 15 K and 2 W at 80 K which proves the advantages of the active displacer. Since the acoustic power of the hot end of the second stage pulse tube is already very small, the active warm displacer has stronger phase adjustment [9]. Referring to the structure of Air Liquide SPTC, this paper introduces the design and fabrication of a thermal-coupled two-stage SPTC which uses an active displacer without work recovery as the phase shifter at the second stage. It is experimentally confirmed that this structure is beneficial to improve performance and efficiency. The structure of the SPTC, the experimental results and discussions will be presented.

2 Design of the Cryocooler A two-stage SPTC was designed and fabricated by Shanghai Institute of Technology and Physics of Chinese Academy of Sciences. Figure 1 shows the schematic of the SPTC. We use a thermal-coupled two-stage structure and both stages adopt a coaxial structure. It is different from Air Liquide that the driven compressor of the SPTC is designed as two independent linear compressors at each stage. This solution allows us to optimize the operating conditions of each stage individually and to carry out relevant studies on the two stages separately. Phase shifter in the first stage is an inertance tube and reservoir, in the second stage is an active warm displacer (AWD). The regenerator is the most important part of the SPTC, and the regenerator in the low temperature section is particularly important for the multi-stage SPTC. We have carried out a simulation and analysis of the regenerator II in the second-stage pulse tube cryocooler based on the software Regen3.3. We kept the parameters of the cold end constant. Figure 2(a) gives the result of the non-ideal heat transfer loss, axial heat conduction loss, enthalpy flow associated with the enthalpy pressure dependence and non-ideal expansion loss. Figure 2(b) gives the result of the cooling capacity and relative Carnot efficiency of the regenerator. As the length/equivalent radius of the regenerator II increase, the non-ideal heat transfer loss decreases and the enthalpy flow associated with the enthalpy pressure dependence and the PV work by pressure drop at the warm end increase. It can be seen from the results of Fig. 2(b) that the length/equivalent radius of

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Fig. 1. The schematic of the developed two-stage SPTC.

the regenerator II has an optimal value 3.81, which means it is beneficial to performance at 15 K that the regenerator II is shorter and wider.

PV work at cold end:4.65W Volume of the regenerator is constant

2.0

1.5

1.0

0.5

0.0 3.16

3.81

4.55

5.23

6

6.8

cooling capacity at 15 K (W)

Various Losses(W)

2.5

0.040

1.2 non-ideal heat transfer loss axial heat conduction loss enthalpy flow associated with the enthalpy pressure dependence non-ideal expansion loss

cooling capacity at 15 K COP

1.1

0.035

1.0 0.030 0.9 0.025

COP

3.0

0.8 0.020 0.7 0.6

PV work at cold end:4.65W Volume of the regenerator is constant

0.015

0.010

0.5 3.16

length/equivalent radius

(a) various losses in PV work of cold head

3.81

4.55

5.23

6

6.8

length/equivalent radius

(b) performance

Fig. 2. The simulation results of the regenerator II based on Regen.

Design of the other component and optimization of the SPTC were done with onedimensional simulation software. The necessary information of the design parameters of the cold head are shown as Table 1. We used layered filling technology in the regenerator of the second stage. In regenerator I, we filled 350# and 400# stainless steel screens. In regenerator II, we filled a part of 500# stainless steel screens and a part of Er3 Ni.

3 Experimental Set-Up The photos of the cold head and the experimental prototype are shown as Fig. 2. A copper thermal bridge connects the first stage cooler and the middle heat exchanger of the second stage pulse tube. The displacer in the second stage was driven by one moving-magnet linear compressor, which allowed active control of phase between the displacer and the compressor piston (Fig. 3).

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Dimension(mm) Regenerator

F32.9

Pulse tube

F15.5 F22.5

Regenerator I

350# 400#

F22.5

Regenerator II

500#

F21.5

Er3 Ni

F21.5

Pulse tube

F11

Fig. 3. The photos of the cold head and the experimental prototype.

Two LVDTs (linear variable differential transformer) and an oscilloscope are used to measure and read the amplitude and phase of the compressor piston and the displacer piston. In this way, we can measure and control the phase relationship. A dynamic pressure transducer is used for measuring the pressure waves at the connecting tube between compressor and the cold finger of the second stage. Three platinum resistance thermometers with accuracy of 0.1 K are used to measure the temperatures, namely the first stage cold head, the middle heat exchanger, and the outside aluminum shield. The temperature of the 2nd cold head is measured by a Cernox thermometer with 9 mK at 20 K accuracy. Two resistance heaters are used to measure the cooling capacity of each stage. The results of the experiment are as follows.

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4 Experimental Results and Discussions 4.1 Influence of Frequency in Second Stage Operating frequency is an important parameter of SPTC, which will affect the heat transfer efficiency of the regenerator. Increasing the operating frequency can increase the input power density of the compressor, but is limited by the heat capacity of the regenerator materials at extremely low temperatures, and a relatively low operating frequency is often required. In the preliminary experimental study, we found that the frequency has a greater effect on the second stage, which may be the reason for the use of the magnetic regenerator material Er3 Ni. For a more detailed study, we tested the influence of frequency. Two independent linear compressors allow us to adjust the frequency of the second stage individually. The result of the experiment is shown in Fig. 4. The phase difference means the phase difference between the displacement of the compressor piston and AWD piston. In the experimental study, the precooling conditions are consistent which ensure comparability of experimental results. The frequency of the first stage is 50 Hz and the pre-cooling temperature is 80 K. The experimental results in the figures show that the performance of the cryocooler gets better first and then worse with the increase of frequency. And the maximum cooling capacity occurs at 30 Hz as the purple line, and the maximum efficiency occurs at 20 Hz as the red line.

(a) the cooling capacity at 15 K

(b) the performance at 15 K

Fig. 4. The experimental results of the cryocooler

4.2 Cooling Performance The performance of the cryocooler when it gets the maximum efficiency and the maximum cooling capacity are listed respectively in the Table 2. The condition 1 is the maximum efficiency occurs at 20 Hz and the condition 2 is the maximum cooling capacity occurs at 30 Hz. In condition 1, it acquires a cooling power of 0.74 W at 15 K, while the input electrical power of each stage is 134 W and 129 W. The input electrical power of the active warm displacer is 31 W. The maximum relative Carnot efficiency is 4.79%. In condition 2, it acquires a cooling power of 0.91 W at 15 K, while the input electrical

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power of each stage is 190 W and 170 W. The input electrical power of the active warm displacer is 26 W. The total input electrical power is 386 W. This shows that the appropriate reduction of the operating frequency is beneficial to improve the performance of the SPTC which use the magnetic regenerator materials, but the reduction of the frequency reduces the input power density of the compressor which lead to the decrease of the available PV power and the cooling capacity. By selecting the frequency reasonably, the SPTC can obtain a higher cooling capacity while obtaining a good efficiency. Table 2. The cooling performance at 15 K of the SPTC Condition 1

Condition 2

frequency

20 Hz

30 Hz

2nd temperature

15 K

15 K

the middle heat exchanger temperature

80 K

80 K

1st input electrical power 2nd input electrical power

134 W

190 W

129 W

170 W

AWD input electrical power

31 W

26 W

15 K cooling capacity

740 mW

910 mW

rcop

0.0479

0.0448

4.3 Influence of Cooling Temperature in Second Stage We also tested the influence of cooling temperature. The frequency of the second stage is 40Hz and the pre-cooling temperature is 80 K. The result of the experiment is shown in Fig. 5. When the input electrical power of the second stage is 180 W, we can get 1.56 W at 20 K or 3.38 W at 30K. The total input electrical power is about 420 W. It is worth mentioning that when we increase the secondary input power to the compressor maximum load, we can get 1.87 W at 20 K. In this situation, the input electrical power of each stage is 216 W and 243 W. The input electrical power of the active warm displacer is 24 W, and the total input electrical power is 483 W.

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Fig. 5. The influence of cooling temperature of the second stage.

5 Conclusion and Discussion Stirling-type pulse tube cryocooler (SPTC) has great characteristics of low vibration, compact structure, and high efficiency so that can be used for cooling infrared detectors and per-cooling helium JT cryocooler. A thermal-coupled two-stage SPTC has been designed and manufactured in this study. Through simulation calculation by Regen, we designed the regenerator in the low temperature section. Experiment result shows that it can obtain 0.91 W at 15 K with the input electric power of 386 W. We using two compressors to drive each of the two stages allows us to study the operating conditions of the second stage separately. The influence of the frequency is researched. The experimental results show that there exist an optimal frequency 30 Hz to achieve the maximum cooling capacity for 15 K applications. There also exist a frequency 20 Hz to achieve the maximum rCOP 4.79%. Acknowledgments. This work is supported by the Hundred Talents Program of the Chinese Academy of Sciences, the National Natural Science Foundation Projects (51806231), the Strategic Priority Research Program of Chinese Academy of Sciences (XDB35000000).

References 1. Liu, S., Jiang, Z., Ding, L., et al.: Impact of operating parameters on 80K pulse tube cryocoolers for space applications-ScienceDirect. Int. J. Refrig 99, 226–233 (2019) 2. Wang, B., Gan, Z.H.: A critical review of liquid helium temperature high frequency pulse tube cryocoolers for space applications. Prog. Aerosp. Sci. 61, 43–70 (2013)

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3. Zhi, X.Q., Han, L., Dietrich, M., et al.: A three-stage Stirling pulse tube cryocooler reached 4.26 K with He-4 working fluid. Cryogenics 58, 93–96 (2013) 4. Qiu, L.M., Cao, Q., Zhi, X.Q., et al.: A three-stage Stirling pulse tube cryocooler operating below the critical point of helium-4. Cryogenics 51(10), 609–612 (2011) 5. Pang, X., et al.: Development of a two-stage Stirling type pulse tube cooler with precooling inside the pulse tube. Int. Cryocooler Conf. 19, 57–63 (2016) 6. Wu, X., Chen, L., Liu, X., et al.: A 15K two-stage gas-coupled Stirling-type pulse tube cryocooler. IOP Conf. Ser. Mater. Sci. Eng. 502, 01204 (2019) 7. Zhu, H., Jiang, Z., et al.: Comparison of three phase shifters for Stirling-type pulse tube cryocoolers operating below 30K. Int. J. Refrig. 88, 413–419 (2018) 8. Chassaing, C., Butterworth, J., Aigouy, G., et al.: 15K pulse tube cooler for space missions. In: International Cryocooler Conference, vol. 18, pp. 27–32 (2014) 9. Zhu, S., Lin, Y.: Fundament of input power distribution and phase shifter functions of a step displacer type two-stage pulse tube refrigerator. Int. J. Refrig 113, 31–37 (2020)

Thermodynamic Performance Analysis of an −180 to −150 ◦ C Refrigeration System with Precooling Dandan Sun1 , Haocheng Wang2 , Qinglu Song1(B) , Dechang Wang1 , and Jinxing Wu1 1 College of Mechanical and Electrical Engineering, Qingdao University, Qingdao, China

[email protected] 2 Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Chinese Academy of Sciences, Beijing, China

Abstract. The rapid development of science and technology has dramatically increased the requirements for refrigeration technology in the temperature range from −190 ◦ C to −50 ◦ C, such as biomedical science, clean energy and aerospace engineering. Mixed-refrigerant Joule-Thomson (MRJT) refrigeration technology has distinct advantages in this temperature range. The system efficiency deteriorates with the decrease of refrigeration temperature, especially at the temperature below −150 ◦ C. Precooling can effectively reduce the temperature span of the main refrigeration system and improve the system efficiency. This paper presents a low temperature refrigeration cycle coupled with a precooling system to produce the temperature below −150 ◦ C. A mathematical model based on the energy and exergy methods was established to evaluate the system performances at varied operation conditions. The results indicated that the precooling system can significantly improve the refrigeration efficiency. At the precooling temperature of − 40 ◦ C, −50 ◦ C, −60 ◦ C, and the main cycle target temperature of −180 ◦ C to − 150 ◦ C, the composition and working pressure of the R170/R600a binary mixture in the precooling cycle and the N2 /R14/Methane/Ethane quaternary mixture in the main cycle were optimized by genetic algorithm (GA). Under the optimal working conditions, the cooling capacity, power consumption, exergy are analyzed. The results obtained in this paper are of great significance to the performance analysis and optimal design of low-temperature refrigerator. Keywords: Precooling · Mixed-Refrigerant Joule-Thomson · Exergy Analysis · Genetic Algorithm

1 Introduction Many studies have shown that Mixed-Refrigerant Joule-Thomson (MRJT) refrigerators are simple, convenient, efficient, economical and reliable in the temperature range from 80 to 230 K [1–4]. With the development of science and technology, MRJT refrigerators have been applied in various low temperature scenarios. For example, the cryopreservation of vaccines [5], medical cryotherapy technology, high temperature superconductivity technology [6], natural gas liquefaction [7]; and so on. Alexeev and Quack [8] © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 605–612, 2023. https://doi.org/10.1007/978-981-99-6128-3_78

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indicated that the refrigeration capacity of the pre-cooling system will be 1.5–2 times higher than that of the system without precooling. In order to improve the cooling capacity, a large number of scholars take precooling measures [9–12]. Rogala et al. [10] studied the impact of precooling on system performance, and conducted system modeling and component optimization. Of course, the optimization of the components is also very important. Ahmed et al. [13] provided detailed explanations of the current mainstream optimization tools to provide the direction for our optimization. In order to better achieve low-temperature refrigeration, in this paper, genetic algorithm was used to optimize the refrigeration system. At the precooling temperature of −40 ◦ C, −50 ◦ C, −60 ◦ C, and the main cycle target temperature of −180 ◦ C to −150 ◦ C. The composition and working pressure of the R170/R600a binary mixture in the precooling cycle and the N2 /R14/Methane/Ethane quaternary mixture in the main cycle were optimized by GA.

2 Methodology 2.1 System Description The refrigeration cycle consists of two mixed-refrigerant Joule-Thomson (MRJT) cycles, as shown in Fig. 1 (a). The numbers 1,2… 13 are the state points, state points 1 to 6 constitute a precooling cycle, state points 7 to 13 constitute a main cooling cycle. Figure 1 (b) is the pressure-enthalpy diagram (p-h) of the cycle. Two cycles are connected by the cascade heat exchanger.

Fig. 1. (a) Flow chart of the refrigeration cycle. (b) Pressure-enthalpy diagram.

2.2 Refrigerant and Operating Conditions The zeotropic mixture used in the pre-cooling cycle is R600a/R170, and the main cycle is N2 /R14/Ethane/Methane. Thermodynamic properties of pure refrigerants used in this paper are shown in Table 1. The ambient temperature is 25 ◦ C, and the pinch point heat transfer temperature difference in both condensers are 5 ◦ C. The mass flow rate in the main cycle was set as 10 kg h−1 .

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It is assumed that the pressure drop in all heat exchangers is 0, and ignored the heat loss. The relationship between the isentropic efficiency (ηis ) and the pressure ratio (pr ) of both compressors [14] as follows: ηis = 0.65 + 0.015pr − 0.0015pr2

(1)

Table 1. Thermodynamic properties of pure refrigerants. Refrigerant

Chemical name

NBP/◦ C

Freezing point/◦ C

Critical Temperature/◦ C

R600a

Isobutane

−11.75

−159.42

134.66

R170

Ethane

−88.58

−182.78

32.17

N2

Nitrogen

−195.80

−210.00

−146.96

R14

Carbon tetra-fluoride

−128.05

−183.6

−45.5

CH4

Methane

−161.48

−182.46

−82.59

2.3 Refrigerant and Operating Conditions Based on the above assumptions and working conditions, the energy and exergy equations for the overall system is shown as follows: The cooling capacity of the overall system can be obtained by: Qc = m(h13 − h12 )

(2)

The compressor work of the precooling cycle can be obtained by: Wcom. - I = m (h2 − h1 )

(3)

The compressor work of the main cycle can be obtained by: Wcom. = m(h8 − h7 )

(4)

The coefficient of performance (COP) of the system can be obtained by: COP =

Qc Wcom. + Wcom. - I

(5)

In the compressor: Icom. = mT0 (s8 − s7 )

(6)

In the condenser: Icon. = m[h8 − h9 − T0 (s8 − s9 )] − m(1 −

T0 )(h8 − h9 ) Tave + T

(7)

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In the cascade heat exchanger: Icas. = mT0 (s10 − s9 ) + m T0 (s6 − s5 )

(8)

In the expansion valve: Iexp. = mT0 (s12 − s11 )

(9)

In the evaporator: Ieva. = m[T0 (s13 − s12 ) −

T0 (h13 − h12 )] T

(10)

In the recuperator: Irec. = mT0 [(s7 − s13 ) + (s11 − s10 )]

(11)

Irec .I = m T0 [(s4 − s3 ) + (s1 − s6 )]

(12)

In the recuperator -I:

In the expansion valve-I: Iexp. - I = m T0 (s5 − s4 )

(13)

In the condenser-I: Icon. - I = m [h2 − h3 − T0 (s2 − s3 )] − m (1 −

T0 )(h2 − h3 ) Tave + T

(14)

In the compressor-I: Icom. - I = m T0 (s2 − s1 )

(15)

where T 0 is the ambient temperature, T is the evaporation temperature, T ave is the mathematical average of the evaporator’s input and output temperatures, m is the mass flow in the main cycle, m is the mass flow in the precooling cycle and T is the temperature differential between the refrigerant and the hot area, which is considered to be 10 ◦ C. The total exergy can be obtained by: Itotal = Icom. +Icon. +Icas. - I +Iexp. +Ieva. +Icas. - II +Icas. - III +Iexp. - I +Icon. - I +Icom. - I (16)

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Table 2. The parameters of the relevant GA. Parameters

Population size

Generations

Crossover probability

Mutation probability

Value

200

50

0.8

0.01

2.4 Optimization The parameters of the relevant genetic algorithm (GA) are shown in Table 2. The expression of the penalty function is shown below.    p(X )max = COP − σ gHX (X ) + gT (X ) + gpr (X )

(17)

HX

X is the vector of optimization parameters, X = (psus , pdis , n), psus and pdis are the compressor suction and discharge pressures, n is the component of zeotropic mixture. σ is the punishment factor with a value of 10 and the expression of g(X) is shown below. The penalty function gHX (X) is used to ensure that the heat exchanger (HX) pinch point (T min ) is between 3 and 8 ◦ C, gT (X) and gpr (X) are used to limit the compressor discharge temperature to less than 120 ◦ C and compressor pressure ratio is between 2 and 8 ◦ C, respectively. ⎧ ⎨ 3 − Tmin , Tmin < 3 ◦ C gHX (X ) = Tmin − 8, Tmin > 8 ◦ C (18) ⎩ 0, 3 ◦ C ≤ Tmin ≤ 8 ◦ C  0, Td ≤ 120 ◦ C (19) gT (X ) = Td − 120, Td > 120◦ C ⎧ ⎨ 2 − pr, pr < 2 gp (X ) = pr − 8, 8 − pr (20) ⎩ 0, 2 ≤ pr ≤ 8

3 Results and Analysis Figure 2 illustrates the variation of COP at target temperatures from −180 ◦ C to −150 ◦ C. As the pre-cooling can reduce the overall temperature span, reducing the irreversible loss. Obviously, precooling helps to improve the COP of system. But it does not mean that the lower the precooling temperature, the higher overall COP. With the gradual decrease of the precooling temperature, the compressor outlet temperature will gradually increase, resulting in a high pinch point of the heat exchanger, which is not conducive to system operation. The variation of overall COP, the main cycle compressor power, the precooling cycle compressor power, the Qc at different target temperatures is shown in Fig. 3, the different precooling temperatures are presented by (a), (b) and (c). With the reduction of

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Precooling-60 Precooling-50 Precooling-40

0.18 0.17 0.16

COP

0.15 0.14 0.13 0.12 0.11 0.10 0.09 0.08

-180

-175

-170

-165

-160

-155

-150

Target Temperature( )

Fig. 2. The COP of total system at −180 ◦ C to −150 ◦ C

Fig. 3. COP and W at −180 to −150 ◦ C. (a) T pre = −60 ◦ C (b) T pre = −50 ◦ C (c) T pre = − 40 ◦ C.

the precooling temperature, the difference between the discharge pressure of precooling cycle and main cycle decreases, and the Qc increases. The change in the trend of mass fraction, suction pressure and discharge pressure at different target temperatures is shown in Fig. 4, the different precooling temperatures are presented by (a), (b) and (c). As can be seen from these figures, with the decrease of precooling temperature, low boiling point refrigerant gradually increases, high boiling point refrigerant gradually declines. The reduction of high boiling point component can effectively avoid cold blockage. The suction and discharge pressure are not much different between (a), (b) and (c), and the closer the pressure ratio is to the set upper limit, the higher the COP is. Among the components of the main cycle, N2 has the lowest boiling point and Ethane has the highest boiling point, and in order to achieve better temperature relay, R14 is used to connect the temperature zone of methane and ethane. The change in the exergy loss of the system components at different target temperatures is shown in Fig. 5, the different precooling temperatures are presented by (a), (b) and (c). The exergy analysis of each component can provide a theoretical basis for system improvement. With the dreasease of precooling temperature, the compressor exergy loss remains almost constant, the exergy loss of of the system are concentrated in condenser I and recuperator I. Condenser I and recuperator I are two heat exchangers in the precooling system. Due to the temperature mismatch of hot and cold fluids in the heat exchanger, the heat transfer temperature difference is large, resulting in large losses.

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Fig. 4. pr and Mass fraction of the main cycle. (a) T pre = −60 ◦ C. (b) T pre = −50 ◦ C. (c) T pre = −40 ◦ C.

Fig. 5. Exergy loss of system components (a) T pre = −60 ◦ C. (b) T pre = −50 ◦ C. (c) T pre = − 40 ◦ C.

4 Conclusions (1) Take the target temperature −150 ◦ C as an example, at precooling temperatures of −40 ◦ C, −50 ◦ C, and −60 ◦ C, the COP is 0.132, 0.179, 0.182, respectively. Obviously, precooling helps to improve the COP of system. While, with the decrease of the precooling temperature, the compressor outlet temperature will gradually increase, resulting in a high pinch point of the heat exchanger, which will deteriorate the system performance. (2) With the reduction of the precooling temperature, the difference between the discharge pressure of the precooling cycle and the discharge pressure of the main cycle decreases, and the Qc increases. With the decrease of precooling temperature, low boiling point refrigerant gradually increases, high boiling point refrigerant gradually declines. The reduction of high boiling point component can effectively avoid cold blockage. (3) With the precooling temperature decreasing, the compressor exergy loss remains almost constant. The exergy analysis of each component can provide a theoretical basis for system improvement. With the decrease of precooling temperature, the compressor exergy loss remains almost constant, the exergy loss of the system are concentrated in condenser I and recuperator I.

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Acknowledgements. This work was supported by National Natural Science Foundation of China (Grant No. 52106016), Natural Science Foundation of Shandong Province, China (Grant No. ZR2021QE219), and Key R&D Program of Shandong Province, China (Grant No. 2022CXGC020901).

References 1. Pang, W.Q., Liu, J.P.: A strategy to optimize the charge amount of the mixed refrigerant for the Joule-Thomson cooler. Int. J. Refrig. 69, 466–479 (2016) 2. Walimbe, N.S., Narayankhedkar, K.G., Atrey, M.D.: Experimental investigation on mixed refrigerant Joule-Thomson cryocooler with flammable and non-flammable refrigerant mixtures. Cryogenics 50(10), 653–659 (2010) 3. Bychkov, E.G.: An analytical method for ensuring the optimal circulating composition of the multicomponent refrigerant mixture in a steady-state operation mode of the Joule-Thomson refrigerator operating with mixtures. Int. J. Refrig. 130, 356–369 (2021) 4. Rogala, Z.: Composition optimization method for mixed refrigerant MR JT cryocooler. Cryogenics 113, 103–223 (2021) 5. Sun, J., Zhang, M., Gehl, A.: COVID 19 vaccine distribution solution to the last mile challenge: experimental and simulation studies of ultra-low temperature refrigeration system. Int. J. Refrig. 133, 313–325 (2022) 6. Lee, C., Jin, L., Park, C., Jeong, S.: Design of non-flammable mixed refrigerant JouleThomson refrigerator for precooling stage of high temperature superconducting power cable. Cryogenics 81, 14–23 (2017) 7. Chang, H.-M.: A thermodynamic review of cryogenic refrigeration cycles for liquefaction of natural gas. Cryogenics 72, 127–147 (2015) 8. Alexeev, A., Quack, H.: Mixed gas JT cryocooler with precooling stage. Cryocoolers 10 (2009) 9. Wang, H.C., Guo, H., Gong, M.Q.: Development and performance test of a miniature movable mixed-refrigerant liquid nitrogen generator. Cryogenics 96, 1–9 (2018) 10. Rogala, Z.: Application of precooling stage in MR JT cryocoolers. Cryogenics 121, 103–395 (2022) 11. Lee, J., Jeong, S.: Investigation of neon–nitrogen mixed refrigerant Joule-Thomson cryocooler operating below 70 K with precooling at 100 K. Cryogenics 61, 55–62 (2014) 12. Alabdulkarem, A., Mortazavi, A., Radermacher, R., Rogers, P.: Optimization of propane pre-cooled mixed refrigerant LNG plant. Appl. Therm. Eng. 31(6–7), 1091–1098 (2011) 13. Ahmed, R., Mahadzir, S., Biswas, K.: Artificial intelligence techniques in refrigeration system modelling and optimization: a multi-disciplinary review. Sustain. Energy Technol. 47, 101488 (2021) 14. Wang, J.F., Brown, C., Cleland, D.J.: Heat pump heat recovery options for food industry dryers. Int. J. Refrig. 86, 48–55 (2018)

Experimental Study of a Single-Stage Coaxial Pulse Tube Cryocooler Aimed at Cooling Two Infrared Detectors Yinan Han, Ankuo Zhang(B) , Wenhui Yu, and Jing Xie Department of Refrigeration and Cryogenic Engineering, Shanghai Ocean University, 999 Hucheng Ring Road, Shanghai 201306, China [email protected]

Abstract. Space infrared detection is the most widely used platform of pulse tube cryocoolers (PTCs). The development of multi-stage PTC for simultaneous cooling of two or more detectors is discussed. However, its structural characteristics have the problems of large volume, heavy mass and uncompact structure. This paper aims to study the feasibility of multi-temperature cooling of singlestage PTC, to improve the compactness of the cryocooler. In the experiment, multi-temperature cooling of single-stage PTC is achieved by adding intermediate cooling at the cold finger. The large temperature gradient of the cold finger indicates the extra cooling potential. And the effect of the frequency on the temperature of the cold end and inter-stage is discussed. The experiment has also studied two-stage cooling performance under 100 W input power. The results show that the single-stage PTC with multi-stage cooling is feasible. Keywords: Pulse tube cryocooler · Inter-cooling · Compact · Multi-stage

1 Introduction Infrared radiation was first reported in Herschel’s thermometer experiment in the 18th century, and since then infrared detection technology has been developing rapidly [1]. The infrared detector is regarded as the core device of infrared detection technology, transforming the infrared radiation into an electrical signal and obtaining the target image. Generally, infrared detectors normally work in a cryogenic environment. While infrared detectors as the largest application platform of cryocoolers, pulling its development and advancement. The pulse tube cryocooler (PTC) eliminates the cold end moving parts used in the traditional regenerative cryocooler, which has a great application prospect for space infrared detection due to its low vibration, and long operating time [2]. To achieve highquality imaging and reduce the signal-to-noise ratio, the detector is generally required to be cooled to 40 K–90 K, and the optical system needs to be cooled to 120 K–200 K at the same time to avoid the interference of heat leakage. To meet this cooling requirement, the multi-stage PTC is a better choice [3]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 613–620, 2023. https://doi.org/10.1007/978-981-99-6128-3_79

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In recent years, many companies and research institutes have been developing multistage structure PTCs [4–8]. The advent of two-stage PTCs has met the need for simultaneous cooling of two temperature zones for space infrared detection, however, its complex multi-stage structure (Fig. 1) results in large volume and weight, high cost, and inter-stage temperature interaction problems, which increases the instability of the detection system. And it is significantly less competitive in higher temperature zones. Thus, a new method is tried, using one compressor to drive the two cold fingers [9–11]. The development difficulty of this type (Fig. 1) is relatively low, the cost is easy to control, and the influence of inter-stage temperature can be significantly reduced. But there are still problems such as uncompact structure and difficulty in effectively reducing the mass and volume of the whole machine. Compared with the above structure, the single-stage structure is the most compact (Fig. 1), but it is generally only used for cooling at a single temperature, and its multitemperature cooling capacity is ignored [12]. Combined with the characteristics of the large temperature gradient of the regenerator [13], it has a certain multi-temperature refrigeration potential in the higher temperature zone. If it can be better used, it can effectively avoid the above problems. Radebaugh [14] and Zhu [15] confirmed cooling capacity is available in the middle of the regenerator through theoretical analysis and experiments. In this work, based on the coaxial single-stage PTC, the effect of the inter-cooling capacity on the cooling performance is studied to verify its feasibility.

Fig. 1. Schematic of PTC with multi-temperature zones cooling.

2 Development of Multi-stage PTC and Comparison of Cold Finger Structure 2.1 Foreign Research Progress Since the 1990s, various foreign research units have carried out research on multistage pulse tube chillers, including Northrop Grumman Aerospace System (NGAS), Lockheed Martin Advanced Technology Center (LM-ATC) and French Air Liquate company (Air liquid). As shown in Table 1, it can be seen that NGAS has developed multi-stage PTCs that are applied to various temperature zones of low, middle and high temperatures and different cooling capacities. The cold head structure is mainly coaxial type, and

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the weight of the early refrigerator is heavier, and the lightweight research is gradually carried out later. LM-ATC is also the same, and gradually the research direction has shifted to high-frequency micro or small PTCs, which can be used to cool the infrared focal plane working in high temperature zones. Table 1. Development of NGAS’s multi-stage PTC [16–18] Research unit

Time

PTC model

Structure

Cold finger structure

Cooling performance

Mass/kg

NGAS

2003

HCC

Two-stage

Linear

17.5W@85K 2.16W@35K

14.3

2007

HCCQ

Two-stage

coaxial

14.1W@85K 2.1W@35K

14.3

EM

Three-stage

Coaxial (1&2) U-type (3)

0.213W@18K

18.5

ABI

Two-stage

Coaxial &linear

1.9–2.3W@53K 5.1–8.0W@183K

5.5

Two-stage

Coaxial

1.55W@35K

5.5

2009 2012

2.2 Domestic Research Progress Domestic research on multi-stage PTC mainly after 2000, the main research units include the Technical Institute of Physics and Chemistry of the Chinese Academy of Sciences and Shanghai Institute of Technical Physics, and Zhejiang University. Table 2 lists some PTC parameter information. The multi-temperature zone PTC abroad is ahead of the domestic in engineering application, and most of the main research units in China pursue the exploration of lower cooling temperature zone. Considering the compactness of the structure of the PTC, there are few studies on the linear cold finger, and most of them use the coaxial or U-shaped arrangement. However, compared with the single-stage PTC, the multi-stage structure reduces the compactness of the whole machine to a certain extent. In addition, most of the research units adopt two or more phase modulation methods for the multistage PTC, which will increase the complexity of the machine structure to a certain extent.

3 Experiment 3.1 Experiment Setup The schematic diagram and physical diagram of the experimental setup of the PTC are shown in Fig. 2, which consists of four parts: cryocooler module, vacuum module, heat rejection module and test module.

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Research unit

Time

Structure

Cold finger structure

Cooling performance

TIPC CAS

2004

Two-stage

U-type(gas-coupled)

0.41W@35K

2007

Two-stage

Coaxial(1)/U-type(2)(gas-coupled)

0.22W@25K

2015

Two-stage

Coaxial

12.1K(no-load)

2016

Two-stage

Coaxial(thermal-coupled)

0.73W@20K

2008

Two-stage

U-type(thermal-coupled)

1W@35K

2010

Two-stage

U-type(thermal-coupled)

[email protected]

2012

Three-stage

U-type(thermal-coupled)

4.97K(no-load)

Zhejiang University

The linear compressor is driven by an AC power supply. The coaxial arrangement, which places the pulse tube inside the regenerator, makes the structure more compact. In addition, the inertance tube is placed in the gas reservoir to realize the compact design of the PTC. The AC and CHX are designed in a slit configuration to enhance heat transfer. The PT-100 temperature sensors are fixed at the cold end and regenerator (corresponding to the T1 , T2 , T3 and T4 in Fig. 2), and are used to measure the cooling temperature with an accuracy of ±0.1 K. T1 represents the cooling temperature of the cold end, and T3 represents the cooling temperature of the inter-stage. The cooling capacity of each point is determined by measuring the power of the heater, and the heaters are driven by a DC power supply. The cold finger is placed in a vacuum chamber and maintained at a vacuum level of 10–4 Pa, it can reduce heat leakage and cooling loss for high test accuracy. The heat from the hot end and the compressor of the PTC are cooled by a water cooler with a cooling water temperature of 20 °C.

Fig. 2. Schematic diagram and physical diagram of the experimental apparatus (CHX: cold-end heat exchanger, AC: aftercooler).

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3.2 Results and Discussions When the heating load is 0 W, the cooling condition of each point on the cold finger is tested, as shown in Fig. 3. It can be seen that each temperature point has a different degree of decrease with time, and eventually stabilizes at a certain value. The minimum temperature at 1 to 4 increases with the distance from the cold end. When the input power is 100 W, the temperature gradually tends to be stable after 25 min. The minimum T1 of the cold end is 50.4 K, T2 and T4 are respectively 124.7 K and 231.4 K, and the maximum temperature difference between the four points is 18 K. This indicates the cold finger has a great temperature gradient and has a certain cooling potential in the high temperature zone. When the input power increases to 150 W, the temperature of T1 and T2 decreases further, the minimum T1 decreases to 47.3 K, and T2 is 115.2 K. The temperature of T4 has increased, because increasing the input power will bring more heat rejection, and the position of T4 is close to the hot end, which is greatly affected by the hot end. The maximum temperature difference between these four points is larger than that for 100 W input power. In addition, the time taken to cool and stabilize the temperature has been significantly shortened.

Fig. 3. Cooling curves of the cold finger.

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The cooling capacity of the cold end is fixed at 4 W, and there is no heating load in the intermediate. The relationship between the frequency and the cooling temperature is shown in Fig. 4. The temperature of positions 1 and 3 decreases first and then increases with the increase of frequency, the difference is that the lowest temperature of T3 and T4 corresponds to different frequencies. For the cold end temperature, the optimal operating frequency is 49 Hz, and 48 Hz is the optimal frequency relative to the intermediate stage temperature.

Fig. 4. Effect of frequency on temperature.

Figure 5 shows the cooling performance of the cold end and the intermediate when the input power is 100W. Among them, the cooling capacity of the cold end and the intermediate cooling capacity are 1W, 2W, 3W and 4W respectively. When the cold end load is unchanged, increasing the intermediate stage load will inevitably lead to an increase in the intermediate stage temperature, which indirectly indicates the cooling potential of the regenerator. At the same time, the presence of intermediate loads exploits the cooling potential of the regenerator part, so it will cause the cold end temperature to rise.

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Fig. 5. Cooling performance of the PTC under 100 W input power

4 Conclusions Feasible solutions are analyzed for the requirement of multi-temperature cooling for infrared detection. The development and application of traditional multi-stage PTC are reviewed, and the problem of its uncompact structure is pointed out. Due to the cold storage capacity and temperature gradient of the regenerator, the method of simultaneous obtaining cooling at the cold end and the middle of the cold finger is presented. The cooling experiment shows that the regenerator has a certain cooling potential, and the hot end has a great influence on the cooling capacity in the inter-stage. Through the study of the influence of frequency on the temperature of the intermediate and cold end, it is found that the optimal frequency corresponding to the minimum temperature of the two is different. In addition, the performance test at 100 W input power once again shows the feasibility of intermediate cooling, and the influence relationship between intermediate and cold end cooling is also given.

References 1. Herschel, W.: Experiments on the refrangibility of the invisible rays of the Sun. Philos. Trans. Roy. Soc. London 90, 284 (1800) 2. Radebaugh, R.: Pulse tube cryocoolers for cooling infrared sensors. Proc. SPIE Int. Soc. Opt. Eng. 4130, 363–379 (2000)

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3. Zhang, X., Liang, X.Y.: Development of and demands for infrared detectors. Electron. Opt. Control 20(2), 41–45 (2013) 4. Chan, C., Nguyen, T., Jaco, C., Tomlinson, B.J., Davis, T.: High capacity two-stage coaxial pulse tube cooler. Cryocoolers 12, 219–224 (2008) 5. Jiang, Y.Y.: Research on Key Technology of 20K Cryogenic Temperature Two - stage Pulse Tube Cryocooler (2017) 6. Liu, X., Chen, L., Wang, J., et al.: Design of a two-stage gas-coupled high-frequency pulse tube cryocooler working around 4 K. IOP Conf. Ser. Mater. Sci. Eng. 755, 012039 (2020) 7. Fang, K., Nakano, K., Lin, X.G., et al.: Investigation on numerical optimization method for high capacity two-stage 4 K pulse tube cryocooler. IOP Conf. Ser. Mater. Sci. Eng. 502(1), 012040 (2019) 8. Wu, W., Cui, X., Liu, S., et al.: Cooling performance improvement of a two-stage pulse tube cryocooler with er-plated screen as regenerator material. Int. J. Refrig. 131(3) (2021) 9. Ramsey, P.G., Swanson, K.S.: Life testing of the ABI cryocooler: two years complete. Cryogenics 52(4–6), 183–187 (2012) 10. Zhu, H.F., Wu, Y.N., Jiang, Y.Y., et al.: Investigation on impedance character of two cold fingers driven by one compressor. J. Eng. Thermophys. 38(6), 1166–1170 (2017) 11. Liu, S.S, Jiang, Z.H., Zhu, H.F., et al.: Study on performance of an efficiency pulse tube refrigerator operating at two temperature zones. Vac. Cryogenics 025(2), 121–125 (2019) 12. Yang, L.W., Thummes, G.: Single stage high frequency pulse tube cooler with base temperature below 30K. J. Eng. Thermophys. (1), 24–26 (2007) 13. Sun, J.C.: Investigation on regenerator temperature non-uniformity and performance improvement of high power Stirling pulse tube cryocoolers working at liquid nitrogen temperature. Zhejiang University (2013) 14. Radebaugh, R., Marquardt, E.D., Gary, J., O’Gallagher, A.: Regenerator behavior with heat input or removal at intermediate temperatures. In: Ross, R.G. (eds.) Cryocoolers 11, pp. 409– 418. Springer, Boston (2002). https://doi.org/10.1007/0-306-47112-4_52 15. Zhu, S.: 4K pulse tube refrigerator and excess cooling power. Adv. Cryog. Eng. 47 (2002) 16. Jaco, C., Nguyen, T., Tward, E.: High capacity two-stage coaxial pulse tube cooler. Cryocoolers 12(985), 530–537 (2008) 17. Ross, R.G., Johnson, D.L.: NASA advanced cryocooler technology development program. In: Proceedings of SPIE-The International Society for Optical Engineering, vol. 4850, no. 1, p. 1020-8 (2006) 18. Colbert, R., Pruitt, G., Nguyen, T., et al.: ABI cooler system flight module performance. Cryocoolers 15, 49–55 (2010) 19. Yang, L., Thummes, G.J.C.: High frequency two-stage pulse tube cryocooler with base temperature below 20 K. Cryogenics 45(2), 15 (2005) 20. Yang, L., Zhao, M., Liang, J., et al.: Investigation of two-stage high frequency pulse tube cryocoolers. In: International Cryocooler Conference (2007)

Experimental Research About Displacer Type Micro Pulse Tube Cryocooler Zhimin Guo1 , Shaowei Zhu1(B) , and John M. Pfotenhauer2 1 Tongji University, Shanghai 201804, China [email protected], [email protected] 2 University of Wisconsin-Madison, Madison, WI 53706, USA [email protected]

Abstract. Investigations of phase shifters and power recovery mechanisms are of sustainable interest for developing Stirling pulse tube cryocoolers (SPTC) with more compact design and higher efficiency. The warm displacer shows the superiority in phase shifting and work recovery functions, we’ve already explored and quantitively expounded the fundamental working mechanism of warm displacer in the previous work. To examine its advantages over other phase shifters, this work presents the experimental research about a micro pulse tube cryocooler with warm displacer and inertance tube phase shifters. We have developed a 2D model to estimate the displacer type micro SPTC and the numerical calculation reveals good results of the performance of the cooler, thereafter the experiment test set up has been built at the basic of a representative case in the simulation results. Overall, our study concludes that the warm displacer should be raised more concerns in the further research. Keywords: Micro Pulse tube cryocooler · Phase shifter · Displacer

1 Introduction Due to their simple construction and long life owing to a fully passive cold head, pulse tube cryocoolers (PTC) has been used in wide application. Phase shifting mechanism is the main feature of STPCs and is a good guidance to design the phase distribution of the cryocooler in order to get best performance of the system. The best phase relationship is the pressure wave and the mass (or volume) oscillations are in phase with each other at the center of the regenerator [1]. Significant developments of the warm displacer approach [2–4] have demonstrated great improvements, with the cooling power of 10–30 W at 80 K. The highest relative Carnot efficiency of 24.2% was obtained by using a passive warm displacer in the pulse tube cryocooler [5]. Comparing with other alternatives, this phase shifting method seems to be better to miniature pulse tube cryocoolers, since even the inertance tubes are severely limited in their phase shifting capability at small cooling power range[6]. Garaway et al. [7] developed a miniature 150 Hz pulse tube cryocooler. The lowest temperature of 97 K and a cooling power of 530 mW at 120 K were obtained with 5 MPa © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 621–627, 2023. https://doi.org/10.1007/978-981-99-6128-3_80

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operating pressure. Sobol and Grossman developed two micro pulse tube cryocoolers with frequency of 113 Hz, only two minutes was took to cool down from 295 K to 108.7 K and cooling power of 400 mW at 110 K was achieved eventually [8]. A comprehensive investigation about a warm displacer type micro-pulse tube cryocooler was carried out by author’s previous work [9]. Thereafter, we built a displacer type micro pulse tube cryocooler according to the designation from that work, and tested the performance of the micro cryocooler with inertance tube phase shifter and displacer phase shifter.

2 Theoretical Analysis Figure 1 shows the schematic of a displacer type SPTC. Note that the linear motor numbered 10 in the figure is optional according to different modes of displacer. Specifically, the cryocooler works in active type displacer mode with linear motor and passive displacer mode without linear motor. In this experiment designation, a passive displacer type was adopted initially.

1.After cooler 2. Regenerator 3. Cold heat exchanger 4. Pulse tube 5.Displacer front space 6.Displacer piston 7.Displacer back space 8.Displacer rod 9.Displacer spring 10.Linear motor 11.Displacer buffer 12.Compressor compression space

Fig. 1. Schematic of displacer type pulse tube refrigerator

In a warm displacer type PTC, there are two moving parts in the entire system. One is the compressor piston at room temperature, another is a piston or so called warm displacer at the warm end of the pulse tube to recover the expansion acoustic power. Theoretically, to minimize the regenerator loss, any phase angle between the pressure wave and the mass (or volume) flow oscillation can be achieved to meet the optimum performance. As a phase shifter, the warm displacer works on phase shifting by supplying extra space for the gas inertial motion. Thus, the swept volume of the displacer is the crucial parameter once the desired phase relation in the cold head is determined. Once the required phase relationship in the cold head is got for the best performance, the displacer must be driven sufficiently by the gas force on the displacer piston to shift the mass flow

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phase angle. The extra space supplied by the moving displacer can appropriately cancel the phase changing generated by the void volumes of the regenerator and pulse tube. To provide the appropriate phase shifting capacity of the displacer, the spring stiffness can be adjusted according to the physical sizes of the displacer rod and displacer piston and the pressure difference between the warm end and cold end of the regenerator.

3 Inertance Tube Experiment Research 3.1 Experiment Set-Up Figure 2 gives the photo of the cold head. Taking consideration of the convenient assembly of the mesh filling in the micro regenerator, the in-line type structure of the cold tip was chosen. Note that this structure is used for both displacer phase shifter and inertance tube phase shifter, the position of the phase shifters is pointed out in the figure. The copper tube shown here connects the compression space of the displacer and the compressor compression space. The main parameters of this micro cryocooler are given in the Table 1. The materials of pulse tube and regenerator are 304 stainless steel. And the inner diameters of the pulse tube and regenerator are 3 mm and 5 mm, respectively, which are adopted according to the theoretical design, both of them are with wall thickness of 0.25 mm. The mesh filled in the regenerator is Stainless steel screen, and the mesh number is #600 with wire diameter of 0.02 mm. Several inertance tubes were tested to estimate the performance of this micro pulse tube cryocooler. All the tubes have the same inner diameter of 1 mm with different length of 1 m, 1.2 m, 1.4 m, 1.5 m and 1.6 m.

Fig. 2. Experiment system

Fig. 3. Displacer pistons

Figure 3 shows the displacer piston used in the experiment. And the diameters of the displacer piston and the displacer rod are 12 mm and 6 mm respectively. They are manufactured with aluminum and coated with Teflon. Teflon coating is strong enough

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Parts

Parameters

Material

Pulse tube

D3mm, Wall 0.25 mm

Stainless steel

Regenerator

D5mm, Wall 0.25 mm

Stainless steel

Mesh filling

#600, d 0.02 mm

Stainless steel

Inertance tube

D1mm, L(1 m,1.2 m,1.4 m,1.5 m,1.6 m)

Stainless steel

Cold heat exchanger

½ Laser hole drilling + ½ L3mm#150, d 0.06 mm

Copper

for experiment. For longer life time, other material is necessary. Also, flexible spring is needed for much longer life time. The pressures at the compressor compression space and the warm end of the pulse tube were measured. The temperature at the cold end was measured by PT100. 3.2 Experiment Results 3.2.1 Inertance Tube Results Figure 4 shows the cold temperature obtained by several inertance tubes under different frequency. It is found that the optimum length of inertance tube is 1.5 m with the same inner diameter of 1mm, accordingly the lowest cooling temperature of 125 K with frequency of 116 Hz and operating pressure of 3 MPa was got eventually, and the pressure ratio at the compressor was kept around 1.2. In addition, due to the maximum current rating of current source, higher operation frequency was not allowed in this experiment. The performance of this micro cryocooler with inertance tube phase shifter is not good, and the reasons for that could be the following problems. The dead volume between the pulse tube warm end and the inlet of the inertance tube has bad effect, and it is hard to get mostly removed in this small scale due to the assembled method here. The gas movement and heat transfer in a small pulse tube like diameter of 3mm will be strongly affected by the inertia resistance and viscous resistance at the near wall position, the temperature distribution will deviate from the ideal laminarization and the temperature difference along the radial direction is so big that the heat transfer between the gas and wall will go strong. The compressor here is too big for this micro cold head, so that the efficiency of the compressor is very low, and the working frequency is limited while the optimal frequency of the cold head is higher. Figure 5 supply the information of compressor performance under different working conditions. The optimum frequency of the compressor and the cold head seems not matched well, and that may affect the cooling performance so much. 3.2.2 Displacer Results Figure 6 gives the lowest cold temperature of the micro cryocooler with displacer phase shifter under several running frequencies and operating pressures. The cooling performance was very bad in the first trying experiment, and the pressure ratio was low and

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Fig. 4. Cooling temperature distribution with different inertance tubes.

a. Electrical power

b. Current

c. Power factor Fig. 5. Compressor performance in different frequency with several inertance tubes

cannot go up after 1.1. The reason for that could be the matching of the compressor and cold head was not good. The performance was improved a little after a small buffer was added between the compressor and cold finger. And the coldest temperature of 234.4 K was reached with frequency of 100 Hz and operating temperature of 2.2 MPa. But the performance is still too low anyway, which is far away from the design, the reason for

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Fig. 6. Cooling temperature in different frequency with displacer phase shifter

that could be the followings. There could be some mechanical problems about the small displacer piston, and the Teflon coating on the displacer is not well-distributed and the Teflon coating on the rod part was missing. That might make the friction between the piston and wall much stronger, which disturbs the motion of the displacer.

4 Summaries A micro pulse tube cryocooler was manufactured based on previous numerical studying. Experiment with inertance tube was done initially, and the lowest cooling temperature of 125 K at the average pressure of 3.0 MPa and the frequency of 116 Hz was obtained. The test result with displacer is not good which may be from mechanical trouble.

5 Further Research The further experiment research about this micro SPTC with well manufactured micro displacer phase shifters would be very interesting. Acknowledgments. The project is supported by National Natural Science Foundation of China (Number 52076151).

References 1. Radebaugh, R.: Pulse tube cryocoolers for cooling infrared sensors. In: Proceedings of SPIE, The International Society for Optical Engineering, Infrared Technology and Applications XXVI, vol. 4130, pp. 363–379 (2000) 2. Zhu, S.W.: Two-stage pulse tube refrigerator with step displacer as phase shifter. IOP Conf. Ser. Mater. Sci. Eng. 101(1), 012183 (2015) 3. Xu, J.Y., Hu, J.Y., Hu, J.F., et al.: Cascade pulse tube cryocooler using a displacer for efficient work recovery. Cryogenics 86, 112–117 (2017) 4. Shi, Y.Z., Zhu, S.W.: Experimental investigation of pulse tube refrigerator with displacer. Int. J. Refrig 76, 1–6 (2017) 5. Wang, X., et al.: A high efficiency hybrid stirling-pulse tube cryocooler, vol. 5, no. 3. AIP Publishing (2015)

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6. Schunk, L.O., Nellis, G.F., Pfotenhauer, J.M.: Experimental investigation and modeling of inertance tubes. J. Fluids Eng. 127(5), 1029–1037 (2005) 7. Garaway, I., Gan, Z., Bradley, P., Veprik, A., Radebaugh, R.: Development of a miniature 150 Hz pulse tube cryocooler. In: Miller, S.D., Ross, R.G. (eds.) Cryocooler 15, pp. 105–113. Springer, Boston (2009) 8. Sobol, S., Grossman, G.: IOP Conf. Ser.: Mater. Sci. Eng. 278, 012141 (2017) 9. Guo, Z.M., Pfotenhauer, J.M., Miller, F., Zhu, S.W.: A research of micro-pulse tube cryocooler with displacer phase shifter. Appl. Therm. Eng. 205, 117995 (2022)

A Superhigh Performance 40 K Pulse Tube Refrigerator with Miniaturized Active Phase Shifter Song Jiantang1 , Hejun Hui1,2 , Shaoshuai Liu1(B) , Zhenhua Jiang1,2(B) , and Yinong Wu1 1 Shanghai Institute of Technical Physics, Chinese Academy of Sciences, 500 Yutian Road,

Shanghai 200083, China [email protected], {liushaoshuai,jiangzhenhua, wyn}@mail.sitp.ac.cn 2 University of Chinese Academy of Sciences, No. 19A Yuquan Road, Beijing 100049, China

Abstract. Compared to Stirling type refrigerants, Stirling type pulse tube refrigerators have no moving parts at the cold end, making the pulse tube refrigerator have the advantages of long life, high efficiency, and high reliability. The room temperature displacer with active control regulates the mass flow and pressure wave at the cold end of the pulse tube to optimize the performance of the pulse tube refrigerator. Both simulations and experiments have shown that the cooling efficiency of the pulse tube with room temperature displacer phasing is better than that of inertance tube with reservoir phasing in the 40 K temperature region. Adding the structure for power recovery to the active phase shifter can recover the acoustic work at the hot end of the pulse tube and improve the cooling efficiency. The gas load of the work recovery chamber can enhance the resonance efficiency of the displacer piston with micro weight for the same expansion chamber gas load. A cooling capacity of 4 W@ 40 K was obtained at 270 W electrical power with a miniaturized active phase shifter with power recovery. The relative Carnot efficiency is 9.63%. And a cooling capacity of 7 W @ 40 K was obtained at 600 W electrical power, with a relative Carnot efficiency of 7.58%. Keywords: Pulse tube refrigerator · Phase shifter · Active control · Miniaturization · 40 K

1 Introduction With the development of space and long-wave detection technology, to achieve high accuracy and low noise of detectors, long-wave infrared detectors (LWIR) and quantum well detectors require refrigeration in the operating temperature region below 40 K [1]. The pulse tube refrigerator has the advantages of long life, low vibration, high reliability, and high efficiency. In 2018, A 40 K single-stage coaxial pulsed tube refrigerator with inertance tube phasing that can be used to cool QWIP detectors was designed and fabricated by Zhang Anguo et al. of the Chinese Academy of Sciences [2, 3]. A cooling © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 628–634, 2023. https://doi.org/10.1007/978-981-99-6128-3_81

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capacity of 3 W can be obtained at a cold temperature of 40 K, with a Carnot efficiency of 8.4%. In 2021, Liang J.T. et al. designed and fabricated a single-stage coaxial pulse tube refrigerator using inertance tube gas reservoir phasing [4], which achieved a no-load temperature of 23.7 K and could provide 2.1 W cooling capacity at 40 K with 150 W electrical input power at 293 K rejection. The relative Carnot efficiency was up to 8.8% at 40 K. Based on the requirement of 2K × 2K focal plane array LWIR detectors, a pulse tube refrigerator using miniature power recovery active phasing is designed and fabricated by our group. Based on our previous work on a 15 W@170 K pulse tube refrigerator with acoustic power recovery [2, 3]. The structural and phasing parameters of the 40 K pulse tube refrigerator were carefully optimized and designed, and experimental tests were carried out on it. The 4 W@40 K cooling performance was obtained at 270 W input electrical power with a relative Carnot efficiency of 9.63%, which is the highest efficiency for a compact and lightweight work-recovery pulse tube refrigerator working at 4 W@40 K ever reported so far. At the same time, the active power recovery phase shifter is optimized for miniature and low weight, which weighs only 1.3 kg. The experimental verification of the high efficiency and low weight of the active power recovery phase shifter at low temperatures provides a new design direction for the design of single-stage pulse tube refrigerators in the 40 K temperature region.

2 Power Recovery Pulse Tube Refrigerator Test Rig A pulse tube refrigerator using a miniaturized power recovery active phase shifter was designed for 4 W@ 40 K. Figure 1 shows a schematic diagram of the structure of the pulse tube refrigerator. The pulse tube is in coaxial form, and the compressor and active phase shifter are in a split configuration. The parameters of the cold finger are shown in Table 1. The power recovery active phase shifter uses an opposed double piston structure to reduce vibration, and three chambers are formed by pistons with shafts, with the expansion chamber in the middle and the power recovery chamber on both sides. The two power recovery chambers of the power recovery phase shifter, the compressor, and the cold finger inlet are connected by a connecting tube as shown in the figure, and the cold finger outlet and the expansion chamber of the power recovery phase shifter are connected by a connecting tube. Table 1. Structural parameters of the cold finger Parameters

Values

REG length

78 mm

REG diameter

26.5 mm

REG screen mesh

38 mm 350# + 40 mm 400#

PT length

90 mm

PT diameter

12 mm

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REG PT

diameter of piston rod

diameter of piston

HHX

COM CT

PHA

power recovery expansion chamber chamber

(a)

phasor piston

phasor rod

(b)

Fig. 1. Schematic of the structure of (a) the power recovery phase shifter pulse tube refrigerator(PTR: pulse tube refrigerator, COM: compressor, CT: connecting tube, PHA: phase shifter, HHX: hot heat exchanger, PT: pulse tube, CHX: cold heat exchanger, REG: regenerator.) (b) the power recovery phase shifter

3 Simulation Analysis of the Pulse Tube Refrigerator 3.1 Analysis of Power Recovery Active Phasor To decrease the electrical power and increase the motor efficiency, the gas stiffness, the spring stiffness, and the mass of the movable piston are adjusted to make the piston resonate. The piston impedance is expressed as (1) and the electrical power of the phase shifter is expressed as (2).   pc Ac ks pe Ae (1) − + i(ωm − ) + Rm Z = R + iX = v v ω We =

  v2  R(α)2 + Re R2 + X 2 2 (α)

(2)

where α, m, Rm , v, Re , Le , V are the specific thrust, kinetic mass, mechanical damping, piston speed, coil resistance, coil inductance, and voltage. By adjusting the rod diameter, the piston impedance can be adjusted to 0 and the piston is in the state of no power consumption. Figure 2 and Fig. 3 appeal the effect of the rod diameter on the power recovery chamber acoustic power, the expansion chamber acoustic power, gas impedance, and the required motor force when the compressor piston stroke, the piston stroke, and phase of the phase shifter are kept constant. As the rod diameter increases, the area of the work recovery chamber decreases and the recovery acoustic work decreases, while the expansion chamber acoustic work is almost constant. The real and imaginary parts of the gas impedance of the piston decrease as the rod diameter increases and are close to 0 at a rod diameter of 10 mm. The absolute value of the gas force impedance of the piston is smallest at a rod diameter of 10 mm and the required motor force is close to 0 N. As the rod diameter deviates from the optimum value of 10 mm, the absolute value of the gas impedance of the piston increases rapidly, causing the motor force required

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to drive the piston to increase rapidly. The miniature and lightweight phase shifter can only provide a minor motor force. Therefore the diameter of the phase shifter piston rod needs to be optimized to reduce the required motor forces.

30

Acoustic power(W)

25 Th = 300 K Xcom = 3.0 mm Tc = 40 K Xpha = 3.1 mm pm = 3.3 MPa pha = 45 deg f = 50 Hz de = 18 mm

20 15

WPV,e WPV,r

10 5

5

10

15

Rod diameter (mm)

Fig. 2. Variation of acoustic power with rod diameter in the phase shifter

Piston impendance(N/(m/s))

20 10

60

40

0 -10 20

-20 -30 -40 -50 -60

R X Fe

5

0 10

Amplitude of the motro force(N)

BL = 5.3 N/A Rm = 5.5 N/(m/s) Re = 1.3 de = 18 mm Ks = 6200 N/m m = 70 g

30

15

Rod diameter (mm)

Fig. 3. Variation of total impedance and motor force with rod diameter in the phase-shifter

The effect of rod diameter change on the kinetic mass and the required electric work is shown in Fig. 4 when the imaginary part of the total piston impedance X = 0 by keeping the motor parameters constant and adjusting the piston mass. As the rod diameter increases, the movable piston mass increases rapidly, while the electric work first decreases and then increases as the rod diameter increases, and approaches 0 at a rod diameter of 10 mm. For the active phase shifter without the power recovery structure, the piston mass exceeds 200 g to achieve the imaginary part of the piston impedance of 0, which will further increase the weight of the phase shifter. The required electrical power of the phase shifter is close to 0 W.

S. Jiantang et al. piston mass electrical power

Piston mass(g)

200

10 BL = 5.3 N/A Re = 1.3 Ks = 6200 N/m Rm = 5.5 N/(m/s) 0 de = 18 mm X = 0 N/(m/s)

100

We(W)

632

0

5

10

15

-10

Rod diameter (mm)

Fig. 4. Variation of piston mass and shaft work with rod diameter in the phase shifter at the resonance

3.2 Comparison Between Power Recovery and Inertance Tube Based on the optimized parameters of the power recovery phasor, the effect of the phasor impedance on the refrigeration performance is simulated and analyzed. Double-segment inertance tube coupled gas reservoir phasing is the most common phasing method, which is widely used in single-stage pulse tube refrigerators. A gas reservoir with a volume of 250 cc and double-segment inertance tubes with inertance tube diameters of 3 mm and 4.5 mm are used to couple the pulse tube refrigerator, and the refrigeration performance of the cold finger subjected to inertance tube phasing is simulated and compared with the power recovery phasing. When the length of the inertance tube and the stroke and phase of the phasing piston are varied within a certain range, the cold finger impedance of the inertance tube and the power recovery phasor are compared and their effects on the input acoustic work and refrigeration efficiency are shown in Fig. 5 and Fig. 6. The impedance is defined as Z = P/U. When the compressor stroke is fixed, the input acoustic work of the inertance tube is more than that of the active phasing at the same impedance position. The inlet mass flow amplitude of the inlet of the inertance tube phasing is greater than that of the active phasing at the same stroke due to the reduction of the cold finger inlet mass flow amplitude after the addition of the power recovery mass flow. For power recovery active phase shifter, the cooling efficiency shows a peaking trend with the variation of the phasing impedance, and there exists an optimal value that makes the cooling efficiency maximum. Inertance tube phasing also has a similar trend but the maximum value of cooling efficiency does not reach the peak due to the poor inertance tube phasing capability. The input acoustic power of the inertance tube phasing is higher, thus the cooling efficiency of inertance tube phasing is higher at the same impedance position. The maximum cooling efficiency is 1.25%, limited by the phasing capability of the two-segment inertance tube. The structure of the double-inlet inertance tube is applied in the low-temperature pulse tube refrigerator because of the limited phasing ability of inertance tubes. The double-inlet inertance tube phasing achieves a greater range of phase impedance by the mass flow out of the compressor with adjustable orifices. The active phasing has a wider phasing range, with a maximum efficiency of 1.76% at a given stroke and phase. The optimal inertance tube parameters and optimal phasing parameters are shown in Fig. 6.

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Fig. 5. Effect of phase shifter inlet impedance of power recovery piston and inertance tube on the input acoustic power

Fig. 6. Effect of phase shifter inlet impedance of power recovery piston and inertance tube on the cooling efficiency

4 Experiment Result The refrigeration performance was tested by controlling the amplitude and phase of the input voltage with a signal generator and a voltage amplifier. The input power of the compressor and phase regulator is measured using a power meter. The amplitude and phase of the piston displacement of the compressor pistons and phase shifter pistons are measured using LVDT displacement sensors. Temperature sensors are installed at the hot and cold ends of the pulse tube, with the hot end using a PT100 temperature sensor and the cold end using a Cernox 1050 temperature sensor. The temperature of the hot end of the pulse tube is controlled at around 300 K in the experiment. The cold head is also mounted with a heating pad to provide the rated thermal load. The cold head is wrapped with multiple layers. The pressure in the vacuum chamber is maintained at 1e–5 Pa during the test. Based on the optimized phasing parameters, The performance of 40 K under the inertance tube and the active phase shifter was tested. The cooling performance of 4 W@ 40 K can be obtained at an input electrical power of 270 W with an rCOP of 9.63% under the active phase shifter. When the input electrical work is 300 W, a cooling

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capacity of 3.5 W@40 K can be obtained using inertance tube phasing with a rCOP of 7.67%, while 4.67 W@40 K can be obtained using power recovery active phasing with a rCOP of 10.1%. The cooling efficiency of power recovery active phasing is better for the same input power, which is consistent with the previous simulation. At the same time, the required input power for the active phase shifter is less than 1 W. The maximum cooling capacity of 40 K can reach 7 W with a compressor electrical power of 600 W and phase shifter electrical power of 4 W.

5 Conclusion A cooling capacity of 4 W@ 40 K was obtained at 270 W electrical power with a miniaturized active phase shifter with power recovery. The relative Carnot efficiency is 9.63%. The maximum cooling capacity of 40 K can reach 7 W. The power recovery active phase shifter has the advantages of light weight and high efficiency. The power recovery structure contributes to the miniaturized active phase shifter. The diameter of the power recovery rod is optimized for non-power consumption and light mass. Acknowledgments. This work is supported by the Hundred Talents Program of the Chinese Academy of Sciences, the National Natural Science Foundation Projects (51806231), the Strategic Priority Research Program of Chinese Academy of Sciences (XDB35000000).

References 1. Han, Y., Zhang, A.: Cryogenic technology for infrared detection in space. Sci. Rep. 12, 1–15 (2022) 2. Liu, S., Jiang, Z., Zhang, A., Zhu, H., Deng, W., Wu, Y.: High efficiency 35 K single-stage pulse tube cryocooler for a space infrared detector (2018) 3. Zhang, A., Wu, Y., Liu, S., Liu, B., Yang, B.: High-efficiency 3 W/40 K single-stage pulse tube cryocooler for space application. Cryogenics 90, 41–46 (2018) 4. Wang, N., Zhao, M., Chen, H., Wei, L., Liang, J., Cai, J.: 24 K single-stage coaxial pulse tube cryocooler without double-inlet phase shifter (2021) 5. Chen, X., Ling, F., Zeng, Y., Wu, Y.: Investigation of the high efficiency pulse tube refrigerator with acoustic power recovery. Appl. Therm. Eng. 159, 113904 (2019) 6. Deng, W., Liu, S., Chen, X., Ding, L., Jiang, Z.: A work-recovery pulse tube refrigerator for natural gas liquefaction. Cryogenics 111, 103170 (2020)

Investigation on Dynamic Pressure Characteristics of a 5–7 K Three-Stage Stirling-Type Pulse Tube Cryocooler Wen Fengshuo1,2 , Wu Wenting1 , Liu Shaoshuai1(B) , Song Jiantang1 , Tang Xiao1,2 , Zhu Haifeng1 , Jiang Zhenhua1,2 , and Wu Yinong1(B) 1 Shanghai Institute of Technical Physics, Chinese Academy of Sciences, Shanghai, China

{liushaoshuai,liushaoshuai}@mail.sitp.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. The temperature range of 5–7 K is important for advanced high sensitivity space detectors. The multi-stage Stirling-type pulse tube cryocooler and J-T cryocooler were finally used in James Webb Space Telescope to meet the demand of cooling the mid-infrared instrument at 6.2 K. To explore the dynamic cooling characteristics of 5 –7 K, a thermal coupled three-stage Stirling type pulse tube cryocooler has been developed in SITP. The effects of pressure on the impedance characteristics and losses in the regenerator were analyzed with calculation programs in this study. In experimental tests, a no-load temperature of 3.96 K was reached with He-4, and a maximal cooling capacity of 24 mW at 5 K was obtained. The cooling capacity at 6 K and 7 K with three groups of charge pressures in the third stage were compared. A maximal cooling capacity of about 80 mW at 6 K was obtained when the charge pressure is 2.0 MPa. The test results are consistent with the model calculation and prediction, and it is expected to further optimize the cooling performance by adjusting pressure finely. Keywords: Pulse tube cryocooler · Infrared astronomy · 5 –7 K · Working pressure · Cooling capacity

1 Introduction 5–7 K Cryocoolers is important for space applications, such as cooling the doped silicon focal plane arrays to about 6 K during the infrared astronomical observation, and the Terahertz detector usually needs to be cooled to 6 K for a lower noise. Furthermore, these cryocoolers are also suitable for precooling sub-Kelvin crycoolers. There are some methods to reach the cooling target of 6 K in space applications, such as the Turbo-Brayton cryocooler, pulse tube or Stirling coupling with the J-T cryocooler, and multi-stage pulse tube cryocooler, etc. [1]. The liquid helium temperature can be reached with a multi-stage pulse tube cryocooler, which inherits the advantages of compact structure, high reliability and low vibration as single stage ones. The research of multi-stage pulse tube cryocoolers has a history of about 20 years. So far, there are already some multi-stage pulse tube cryocoolers that can work near © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 635–641, 2023. https://doi.org/10.1007/978-981-99-6128-3_82

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the temperature of 6 K, even to liquid helium temperature, which are developed by NGAS [2], Lockheed Martin Advanced Technology Center [3], Zhejiang University [4] and Technical Institute of Physics and Chemistry, Chinese Academy of Sciences [5]. However, at present, the cooling capacity of three-stage pulse tube cryocoolers in this temperature is small and the cooling efficiency is generally low. With He-4 as working gas, the multi-stage pulse tube cryocoolers are generally only obtain a cooling capacity in the order of 30 mW. A three-stage thermal coupled pulse tube cryocooler designed for reaching larger cooling capacity at 5–7 K is introduced in this study. Analyzes and experimentally studies on the influence of working pressure are conducted to promote the cooling performance.

2 Theoretical Analysis and Simulation The oscillating pressure in the pulse tube include the mean pressure and the oscillating part of pressure, as shown in Eq. (1). The mean pressure is related to gas physical properties and the oscillating part is related to the available energy. p(x, t) = p0 + Re[pd (x)eiωt ]

(1)

The mean pressure of the gas has a significant effect on density, and thus higher pressure is conducive to enhancing the inertia effect of the inertance tube. Figure 1 shows the effect of pressure on the impedance of the cold inertance tube and reservoir. As shown in the figure, when the cold inertance tube is cooled to 20–30 K and the mean pressure varies from 0.6 to 1.8 MPa, with consistent acoustic power (1.0 W) at the inlet of the inertance tube, the phase of impedance slightly increases, while the amplitude of the impedance increases obviously with the increase of pressure. The required mass flow is smaller with a higher pressure, which is conducive to reducing the losses of the regenerators. Moreover, comparing the curves of different temperatures, the phases of impedance have small difference, and the impedance is much larger when the temperature is 20 K.

Phase of impedance (°)

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Fig. 1. Effect of pressure on impedance characteristics of cold inertance tube and reservoir

The real gas effect has a significant effect on the performance of the third stage pulse tube cold finger. Two main indicators of the real gas properties, pressure enthalpy

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coefficient and compression factor, are closely related to the mean pressure. The ratio of pressure enthalpy loss in the regenerator to PV work at the cold end is:           Zh Th ∂h ∂h H˙ P (2) = max ρ , ρ ∂P T C ∂P T h ZC TC P V˙ c The enthalpy loss of real gas properties in the lowest temperature part of the regenerator (Reg33) was analyzed with REGEN3.3. When the cooling temperature is 6 K, a reverse enthalpy flow related to pressure is generated when the pressure is lower than 0.8 MPa, and when the cooling temperature is 7 K, the corresponding value is 1.2 MPa. The variation of heat transfer and flow resistance losses in the regenerator with working pressure is shown in Fig. 2. With consistent parameters at the cold end (the cooling temperature is 6 K, the mass flow rate is 2.0 g/s, the phase difference between pressure and mass flow is 20°, and the pressure ratio is 1.15), the proportion of heat transfer and flow resistance losses decreases with the increase of pressure. An excessive flow resistance loss can be seen when the mean pressure is relatively low, under the mean pressure of 0.6 MPa, the proportion of flow resistance loss is about two times that of heat transfer loss. With the same mass flow rate at the cold end, and the frequency maintained at 27 Hz, the flow velocity is much higher as the pressure is smaller. For example, the peak velocity of gas in the Reg33 is 0.94 m/s at 0.6 MPa while it is only 0.44 m/s at 1.8 MPa. REGEN3.3

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Fig. 2. Variation of heat transfer and flow resistance losses of Reg33 with pressure

The cooling capacity of 5–7 K at the cold end of Reg33 was also calculated by REGEN3.3, as shown in Fig. 3. The ordinate represents the ratio of simulated cooling power to PV power at the cold end. In general, the cooling capacity decreases with the increase of pressure. Interestingly, however, there is a cooling capacity peak when the cooling temperature is 6 K or 7 K, and the optimal pressure for 6 K and 7 K are about 0.8 MPa and 1.0 MPa. Which means, with the increase of cooling temperature, the optimal working pressure tends to increase. Considering the factors of the impedance characteristics of the cold inertance tube and the losses in the regenerator, the optimal pressure at 6 K and 7 K should be higher than 0.8 MPa and 1.0 MPa respectively.

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Fig. 3. Variation of the ratio of simulated cooling capacity to acoustic power at cold end of the Reg33 with pressure

3 Experiments and Results 3.1 The Experimental Setup Figure 4 is the photo of the designed 6 K pulse tube cryocooler, a three-stage thermalcoupled scheme has been adopted, and the third stage cold finger is U type structure, while the 1st and 2nd stage cold finger both are coaxial [6]. To well adjust the phase between the pressure wave and mass flow rate, phase shifters of the three stages are inertance tube, active displacer, and cold inertance tube. The working pressure in the third stage pulse tube cold finger is measured in real time by a dynamic pressure sensor installed at the inlet.

Fig. 4. The 6 K pulse tube cryocooler

3.2 Experimental Test and Results The mean pressure in the third stage pulse tube cold finger changes during the cooling down process, and the varies of mean pressure before and after cooling is according to the following formula: 













3 3 P 0 VRegj P 0 Vpt P 0 Vcop P 0 VcT P 0 Vit P 0 Vres P 0 Vtotal P 0 VHEXi + + = + + + + i=1 Z(T )RT HEXi ) j=1 Z(T )RT Regj ) RT0 RT0 RT0 Z(T2 )RT 2 Z(T2 )RT 2 Z(T )RT pt

(3)

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where P 0 and P 0 is the average pressure before and after cooling process, V represents the empty volume of each component, Z is the compressibility factor and T is the temperature of working gas. Assuming that the precooling temperature of the 2nd stage (T2 ) is 20 K, in order to equalize the pressure after cooling to 1 MPa, the charge pressure at room temperature is about 1.68 MPa. When the input electrical powers of the three stages are 230 W, 180 W, and 70 W, the minimum no-load temperature reached 3.96 K. At this time, the precooling temperatures of the 1st stage and 2nd stage were 82.5 K and 18.68 K, a cooling capacity of 24 mW can be obtained at 5 K. The variation of the mean pressure of the third stage pulse tube cold finger is shown in Fig. 5. The 2nd precooling temperature abruptly rises when the driven compressor of the third-stage pulse tube start-up, which cause the fluctuations of average pressure at the 350th minute. And the mean pressure of the third stage pulse tube before startup is 1.68 MPa, and it is reduced to 0.88 MPa after cooling. 1.8

Mean pressure (Inlet of the 3rd-stage PTC)

P0= 1.68 MPa 1.6

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Fig. 5. Variation of the mean pressure during the cooling down process

In a contrast, the cooling capacities of three groups of charge pressures are tested with a consistent input electrical power, which is 200 W, 200 W, and 100 W to the linear compressor of the pulse tube cold fingers. The mean pressures of the third stage pulse tube before startup are 1.2, 1.68, and 2.0 MPa, and the pressure reduced to 0.68, 0.96, and 1.15 MPa after cooling, as shown in Table 1. Table 1. Mean pressures in the 5– 7 K cold finger index

Mean pressure before cooling

Mean pressure after cooling

1

1.20 MPa

0.68 MPa

2

1.68 MPa

0.96 MPa

3

2.0 MPa

1.15 MPa

In the experiment, the operating frequency is adjusted to obtain the optimal cooling performance, and the values of cooling capacity under different pressures are shown in

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Fig. 6. When the charge pressure is 2.0 MPa, the cooling capacity at 6 K and 7 K are relatively larger, which are 79.8 and 145.8 mW, respectively. The total input electrical power is 522 W, where the demand electrical power of the active piston of the 2nd stage pulse tube is 22 W. 160 140

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80 60 40 20 0

1.20/0.68

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Mean pressure before/after cooling (MPa)

Fig. 6. Experimental results of cooling capacities at 6 K and 7 K under three groups of charge pressures

4 Conclusion The influence of working pressure of the 5–7 K pulse tube cold finger is researched. The analysis result shows that from the perspective of the impedance of the cold inertance tube and reservoir, higher pressure is beneficial for cooling performance, while the effect is opposite considering the flow resistance of the regenerator. The relationship between the mean pressure before and after cooling is discussed, and the precooling temperature and the volume of the inertance tube and reservoir can be well designed to optimize the working pressure. Comparing the cooling performance of three groups of working pressures, the experimental result shows that a higher charge pressure has slight advantages for cooling at 6 K and 7 K, and when the charge pressure is 2.0 MPa, 79.8 mW@6 K is obtained within a total electric power of 522 W. Acknowledgments. This work is supported by the Hundred Talents Program of the Chinese Academy of Sciences, the National Natural Science Foundation Projects (51806231), the Strategic Priority Research Program of Chinese Academy of Sciences (XDB35000000).

References 1. Ross, R.G., Johnson, D.L.: NASA’s advanced cryocooler technology development program (ACTDP). In: AIP Conference Proceedings. American Institute of Physics, vol. 823, no. 1, pp. 607–614 (2006) 2. Nguyen, T., Colbert, R., Durand, D., et al.: 10K pulse tube cooler. Cryocooler 14, 27–33 (2007)

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3. Olson, J.R., Moore, M., Champagne, P., et al.: Development of a space-type 4-stage pulse tube cryocooler for very low temperature. In: AIP Conference Proceedings. American Institute of Physics, vol. 823, no. 1, pp. 623–631 (2006) 4. Zhi, X.Q., Lei, H., Dietrich, M., et al.: A three-stage stirling pulse tube cryocooler reached 4.26K with He-4 working fluid. Cryogenics. 58, 93–96 (2013) 5. Chen Liubiao, W., Xianlin, W.J., et al.: Study on a high frequency pulse tube cryocooler capable of achieving temperatures below 4K by helium-4. Cryogenics 94, 103–109 (2018) 6. Fengshuo, W., Shaoshuai, L., Wenting, W., et al.: Comparison of pure stainless steel wire mesh and mixed HoCu2 particle as regenerator material at10–30 K. Chem. Ind. Eng. Progress 41(1), 113–119 (2022)

Calculation Method of the Theoretical Helium Liquefaction Rate Based on the Regenerative Refrigerator Qiang Cao1(B) , Miaomiao Wang1 , Tiancheng Xiao1 , Yuji Chen1 , Pengcheng Wang1 , Chaojie Chen1 , Peng Li1 , Zhihua Gan2,3 , Qinyu Zhao3 , and Bo Wang3 1 Institute of Refrigeration and Cryogenics, Tongji University, Shanghai, China

[email protected]

2 Institute of Refrigeration and Cryogenics, Zhejiang University, Hangzhou, China 3 Cryogenics Center, Zhejiang University City College, Hangzhou, China

Abstract. Helium liquefaction is important for applications of cryogenic engineering, magnetic resonance interference (MRI), helium recovery and purification, and so on. Regenerative refrigerators are generally applied for small-scale applications because of their relatively high refrigeration efficiency, competitive compactness, high reliability, etc. However, the liquefaction efficiency of helium is not high. The main reason is the contradiction between the large sensible heat load and the limited refrigeration efficiency at 4.2 K. There is a useful method of using intermediate refrigeration power, described as “free cooling power”, to enhance the cooling of the sensible heat load. However, the picture of the working mechanism and optimization method of this method is not clear. In this paper, the ideal liquefaction efficiency by application of the temperature-distributed regenerative refrigeration cycle, based on the heat input or removal theory together with a DC flow method, is solved. Numerical simulations are carried out with this method. The working pressure is under a reduced pressure range of 4.0 and 10.0. The hot-end temperature varies between 10 K and 60 K. While the pressure of the helium gas to be liquefied is 1 atm. A theoretical figure of merit (FOM) of 39.2% has been got at a relatively small temperature range of 4.2–30 K. These liquefaction rates are 2–3 times that with only the cold ends. This method should provide a reference for the optimization of small-scale helium liquefaction systems. Keywords: Calculation Method · Helium Liquefaction · Regenerative Refrigerator

1 Introduction Helium liquefaction is important for applications of cryogenic engineering, magnetic resonance interference (MRI), helium recovery and purification, and so on. Improving the liquefaction efficiency is an enduring work. We start with the ideal liquefaction cycle. As we know, the ideal liquefaction cycle adopts two reversible processes: isothermal compression and isentropic expansion. A minimum work of 6848 kJ/kg is required © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 642–648, 2023. https://doi.org/10.1007/978-981-99-6128-3_83

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in these processes. A performance parameter named the figure of merit (shortened as FOMliquefier or FOM) is defined based on a ratio of minimum work to the real liquefaction work. This is obtained based on theoretical calculation with gas properties [1]. The two-stage structure is generally used for helium liquefaction [2]. Extracting the “free cooling power” in the regenerator promotes liquefaction efficiency [2, 3]. Such a “free cooling power” is extracted from the middle of the regenerator, and it imposes a small degradation on the cold-end refrigeration power. The “free cooling power” has been used to cool the sensible heat load [2]. One example shows that 2.04 W of “free cooling power” was extracted from the regenerator, and the liquefaction efficiency increased from 1.4 L/day to 4.8 L/day [4]. However, it is vague to calculate or optimize the “free cooling power” in previous studies. We have put forward a heat input or removal theory [5, 6] to analyze the regenerator working with real gases and also a DC-flow method [7] to transfer the differential heat input, i.e. intermediate refrigeration power. This theory actually provides a foundation for the temperature-distributed regenerative refrigeration cycle, which is different from the traditional high-low-temperature cycle. In this paper, we will further carry out theoretical analyses for improving the liquefaction efficiency with our theory. Numerical cases will be presented to show its effectiveness.

2 Typical Helium Liquefaction Methods with Two Stages The Ericsson cycle is usually applied for ideal regenerative refrigerators, which combines two isobaric processes and two isothermal processes [8]. For a two-stage refrigerator working with an ideal gas, the refrigeration temperature for the two stages should be determined and the refrigeration power should be distributed with an optimization. The heat load of the gas to be liquefied between the ambient temperature (T0 ) and the first-stage temperature (T1 ) should be absorbed by the first stage, and the heat load between the first-stage temperature (T1 ) and the second-stage temperature (T2 ), including the sensible heat load and the latent heat load, should be absorbed by the second stage. T2 is fixed at 4.2 K, while T1 should be optimized. In the calculation, the total power input is fixed at 2000 W. As T1 increases from 10 K, both the refrigeration power in the first stage (Q1 ) and second stage (Q2 ) increase, as shown in Fig. 1. Q1 reaches a maximum value of 117.4 W at 30 K, while Q2 increases monotonously. The maximum liquefaction rate reaches 58.8 L/day when T1 is equal to 35 K, as shown in Fig. 2. The FOM of liquefaction reaches 29.1%. The refrigeration power reaches 117.1 W @35 K and 15.9 [email protected] K in the two stages. If the real gas properties of the working fluid (4 He) are applied in the ideal Ericsson cycle, the average working pressure is fixed at a normal value of 1.5 MPa in the ˙ c ) at each stage is determined with this equation calculation. The refrigeration power (Q [8]: ˙ ˙ c = pVh {(ZT )c − max[(1 − Tβ)ZT ]}. Q (ZT )h

(1)

˙ h is the time-averaged hot-end PV power in the regenerator, (ZT )h is the where pV product of compressibility factor and temperature at the hot end, (ZT )c is the product of the compressibility factor and temperature at the cold end.

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Fig. 1. Calculated refrigeration powers (ideal gas) meeting the liquefaction requirement in the two stages versus the first-stage temperature.

Fig. 2. Liquefaction rate and FOM of liquefaction versus the first-stage temperature, refrigeration powers are calculated with ideal gas properties.

It is clear that the refrigeration power decreases a lot. As shown in Fig. 3. In the calculation, the total power input is also fixed at 2000 W. As T1 increases from 10 K, Q1 reaches a maximum value of 43.2 W at 20 K, while Q2 increases monotonously. The maximum liquefaction rate reaches 20.5 L/day when T1 is equal to 20 K, as shown in Fig. 4. The refrigeration power reaches 3.2 [email protected] K in the second stage in this case. The maximum FOM of liquefaction is only 10.2%, which is 35.0% of that calculated with an ideal gas. A large portion of heat loss of 65.0% occurs because of the real gas properties of 4 He.

3 The Liquefaction Method with the Temperature-Distributed Regenerative Refrigeration Cycle 3.1 Heat Input or Removal Theory We have put forward a heat input or removal theory to analyze the regenerator working with real gases [5], especially near the critical temperature. The heat input is heat added to the regenerator at intermediate temperatures, and the differential heat input is calculated based on gas properties [5]: ˙ 1 d [(1 − T β)ZT ] δ Q/p V˙ h = δT (ZT )h dT

(2)

Examples of 4 He with a pressure ratio between 4.0 and 10.0 are calculated, and a dimensionless parameter named reduced pressure (shortened as Pr ) is defined based on the ratio of the absolute pressure to the critical pressure. The second-stage refrigeration power Q2 is fixed as 1.5 W in the calculation. It shows that there is always a positive differential heat input, i.e. refrigeration power, in the low-temperature range, and it reaches a peak value between 7.3 K and 9.9 K at a reduced pressure range between 4.0 and 10.0. It decreases sharply since the peak, and back to a value that is quite close to zero. When the differential heat input equals zero, the peak value of the integrated heat input, i.e. intermediate refrigeration power, reaches a peak value. The temperature of such a peak value is between 9 K and 15 K at this pressure range. The intermediate

Calculation Method of the Theoretical Helium Liquefaction Rate

Fig. 3. Calculated refrigeration powers (based on real gas properties) meeting the liquefaction requirement in the two stages versus the first-stage temperature.

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Fig. 4. Liquefaction rate and FOM of liquefaction versus the first-stage temperature, refrigeration powers are calculated with real gas properties.

refrigeration power is more than 10 times of the second-stage refrigeration power Q2 , and it decreases slowly when the temperature increases. Fortunately, it keeps positive, which means there is still an intermediate refrigeration power available when T1 gets higher than 60 K. When the reduced pressure is 10, the intermediate refrigeration power equals 7.1 W, 4.7 times of Q2 . 3.2 DC Flow Method for Transferring the Intermediate Refrigeration Power As discussed before, such an intermediate refrigeration power inside the regenerator is difficult to utilize, and radial thermal resistance occurs when it is extracted from outside the regenerator. We have further put forward the internal DC flow method [7], so as to transfer the internal intermediate refrigeration power. The value of the DC flow is determined based on the energy balance, as shown:  Tx ˙  Tx pV h d [(1 − T β)ZT ] δT (3) cp dT = m ˙ DC,(Tc ,Tx ) dT Tc Tc (ZT )h And the DC flow in the whole section should satisfy the above energy balance, so it is determined with this equation: m ˙ DC,min = min(m ˙ DC ), T ∈ (Tc , Th ).

(4)

where Tc is the helium temperature at the cold end, Th is that at the hot end, m ˙ DC is the time-averaged mass DC flow (Figs. 5 and 6). As shown in Fig. 7, the DC flow between the cold end and a variable first-stage temperature T1 is calculated, which shows the value of the DC flow is usually limited to two ends. If the temperature is lower than a certain value, which is between 12.7 K and 17.6 K for a reduced pressure of 4.0–10.0, the DC flow is limited to the cold end; while the temperature gets higher than that value, the DC flow is limited to the hot end. When the pressure gets lower, the value of the DC flow generally gets larger. Here we give one example at a reduced pressure of 6 (1.37 MPa), when Q2 is 1.5 W and the cold-end PV power is 8.97 W, the DC is limited to the cold end when T1 is equal to 15 K. The value reaches 1.28e–4 kg/s. However, the DC value decreases to only 1.72e–5 kg/s, when T1 gets as high as 60 K. The DC is limited to the hot end.

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Fig. 5. Differential heat input versus the temperature for various pressure.

Fig. 6. The intermediate refrigeration power (integrated heat input) over Q2 versus the hot-end temperature.

3.3 Calculation of the Liquefaction Flow Rate The DC flow needs a recuperator to transfer the refrigeration power, stored as the sensible heat capacity, to the gas to be liquefied, as shown in Fig. 8 [9]. Since the recuperator usually works with a high efficiency and can be over 99% if it is well optimized, the heat transfer efficiency from the intermediate refrigeration power to the gas to be liquefied also works with a high efficiency. What’s more, the method of applying the internal DC flow with a recuperator does not generate a radial thermal resistance, and it is adaptive to both a large regenerator and (a) a regenerative refrigerator with a gap between the regenerator and the cylinder. However, there is a challenge for the recuperator. The specific heat capacity of 4 He varies with the pressure, especially at the critical region, the peak value of the specific heat capacity appears closer to the cold end when the pressure gets lower, as shown in Fig. 9. The gas to be liquefied is usually at 1 atm or a bit higher, but the DC flow that is extracted from the regenerator is usually between 1–2 MPa. This means that more refrigeration power is required at a colder temperature, and more DC flow is required to satisfy the refrigeration power demand.

Fig. 7. The DC flow between the cold end and a Fig. 8. Schematic of the liquefaction variable temperature T1 that fulfills the energy balance. process with the DC flow and also a recuperator [9].

The liquefaction rate is determined with this equation:    Tx  Tx ˙2 +m Q ˙ DC,min cp dT ≥ m ˙ liqu q˙ latent + cp dT , Tx ∈ [T2 , T1 ] T2

(5)

T2

where m ˙ liqu is the time-averaged mass DC flow of the liquefaction process, q˙ latent is the latent heat generated in the liquefaction process.

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The calculation of 4 He, in this case, shows that the liquefaction rate keeps constant when the first-stage temperature rises to a turning-point temperature between 23 K and 32 K, corresponding to a reduced pressure of 4 to 10, as shown in Fig. 10. The secondstage refrigeration power Q2 is fixed at 1.5 W. As T1 further increases, the liquefaction rate decreases. As the pressure gets higher, the liquefaction rate gets smaller at that lower temperature range than the turning-point temperature, but it gets larger at a higher temperature than about the turning-point temperature. For example, when the reduced pressure is equal to 8, the liquefaction rate of 1atm 4 He reaches 49.4 L/d when T is 30 K., but it decreases to merely 16.8 L/day when T 1 1 gets 60 K. The required PV power at 30 K is 58.5 W, and the FOM reaches as high as 39.2%. Here an ideal liquefaction cycle between 4.2 K and 30 K is compared.

Fig. 9. The specific heat capacity of 4 He at various pressures.

Fig. 10. Liquefaction rate versus the first-stage temperature T1 for various pressure.

4 Conclusion This paper discusses a theoretical calculation of the liquefaction efficiency in regenerative refrigerators. The liquefaction method is based on our temperature-distributed regenerative refrigeration cycle, and it generates a temperature-distributed refrigeration power related to real gas properties. The liquefaction method with an internal DC flow improves the liquefaction rate greatly. A theoretical FOM of 39.2% has been got at a relatively small temperature range of 4.2–30 K, which is about 4 times the FOM (10.2%) with only the two-stage cold end working with 4 He. This is promising. Acknowledgments. This work is supported by the National Natural Science Foundation of China under the contract number 52176018 and 52006190.

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References 1. Lemmon, E.W., McLinden, M., Huber, M.: NIST standard reference database 23, NIST reference fluid thermodynamic and transport properties -REFPROP. U.S. Deptartment of Commerce (2002) 2. Thummes, G., Wang, C., Heiden, C.: Small scale 4He liquefaction using a two-stage 4 K pulse tube cooler. Cryogenics 38(3), 337–342 (1998) 3. Zhu, S.W., Ichikawa, M., Nogawa, M., Inoue, T.: 4K pulse tube refrigerator and excess cooling power. In: Breon, S., DiPirro, M. (eds.) Advances in Cryogenic Engineering, Vol 47, Pts A and B, pp. 633–640. Amer Inst Physics, Madison, Wisconsin, USA (2002) 4. Wang, C.: Helium liquefaction with a 4 K pulse tube cryocooler. Cryogenics 41(7), 491–496 (2001) 5. Cao, Q.: Attainability of the carnot efficiency with real gases in the regenerator of the refrigeration cycle. Appl. Energy 220(15), 705–712 (2018) 6. Cao, Q., Luan, M.K., Li, P., Wei, L., Wu, Y.: A critical review of real gas effects on the regenerative refrigerators. J. Therm. Sci. 30(3), 782–806 (2021) 7. Cao, Q., Sun, Z., Li, Z.M., Luan, M.K., Tang, X., Li, P., et al.: Reduction of real gas losses with a DC flow in the regenerator of the refrigeration cycle. Appl. Energy 235, 139–46 (2019) 8. Kaushik, S.C., Kumar, S.: Finite time thermodynamic evaluation of irreversible ericsson and Stirling heat engines. Energy Convers. Manage. 42(3), 295–312 (2001) 9. Cao, Q., Luan, M.K., Huo, B.: An efficient liquefaction system with the DC flow method that based on regenerative refrigerators. China, ZL202021794381.2 (2021)

Development of a 2 K Joule-Thomson Cryocooler with 4 He Caiqian Dong1 , Shaoshuai Liu1(B) , Xinquan Sha1 , Wang Yin1,2 , ZhenHua Jiang1,2 , and YiNong Wu1 1 Shanghai Institute of Technical Physics, Chinese Academy of Science, Yutian Road,

Shanghai, China {liushaoshuai,jiangzhenhua,18621939480wyn}@mail.sitp.ac.cn 2 University of Chinese Academy of Sciences, Yuquan Road, Beijing 100049, China

Abstract. The 2 K mechanical Joule-Thomson (JT) throttling refrigeration system is crucial for the scientific observation and quantum communication of space detectors with extremely high sensitivity and resolution, such as X-ray Microcalorimeter Spectrometer and Superconducting Nanowire Single-Photon Detectors. A 2 K-JT refrigeration system is currently under development and the operating mass is 4 He. The JT cycle, precooled by a two-stage thermally coupled pulse tube cryocooler, is driven by a four-stage JT compression system. The precooler can provide 12 K and 85 K temperature zone pre-cooling at the same time. The four-stage direct current (DC) linear compressor set achieves a very low inlet pressure and a greater overall compression ratio of up to 200. The pressure drop in the low-pressure line is a very important parameter in the design of a 2 K-JT refrigeration system. Reducing the pressure drop in the low-pressure flow path of the system can effectively decrease the load on the compressor unit and facilitate the achievement of lower pressure and lower temperature at the evaporator. The design progress of the 2 K-JT refrigeration system is described in terms of integrated consideration of fluid flow and heat transfer. At present, the JT system experimental platform has been built and the corresponding performance tests are being carried out. Keywords: 2 K · (Joule-Thomson) JT · 4 He · low-pressure drop · cryocooler

1 Introduction The development of 2 K-class JT refrigeration technology has benefited from the maturity and application of 4 K-class JT refrigeration technology. Nowadays, helium hybrid JT cryocoolers have become the dominant choice for liquid helium temperature zone space applications. Compared to dewar and other refrigeration methods, the hybrid JT cryocooler has unique advantages such as small size, lightweight, high efficiency, no moving parts at the cold end, high reliability, low vibration, and so on. Research on hybrid JT cryocoolers in the liquid helium temperature region is driven by applications in space astrophysics and quantum mechanics where detectors need to be pre-cooled to © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 649–655, 2023. https://doi.org/10.1007/978-981-99-6128-3_84

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cryogenic temperatures below 2 K [1–3]. The temperature range below 2 K has been successfully obtained internationally using a hybrid refrigeration cycle with helium-3 as the circulating mass in the JT flow path [1, 4–6]. However, the scarcity of 3 He on earth is only 1/106 to 1/107 of helium from the atmosphere and natural gas, and its high price makes its widespread use impractical. With the increasing demand for refrigeration in the 2 K temperature range, it is essential to use the relatively inexpensive 4 He instead of 3 He as the circulating mass in small cryogenic helium hybrid JT flow paths to access temperatures in the 2 K and lower temperature range. National institute of standards and technology has developed a compact, low-power, closed-cycle cooling system. A threestage pulse tube cooler is used as the pre-cooler, providing 2.05 W, 0.08 W, and 0.005 W of cooling at 80 K, 25 K, and 10 K, respectively. When the high pressure is 0.2 MPa and the low-pressure pressure is 1.3 kPa, the mass flow rate is 0.6 mg/s. and can obtain more than 1.2 mW cooling capacity at 2.2 K [7]. Chinese academy of science also reported a hybrid JT cryocooler with a minimum temperature of 2.17 K, 93 K for the primary pre-cooling, and 17.4 K for the secondary pre-cooling, corresponding to a compressor high-pressure pressure of 0.91 MPa and a low-pressure of 4.3 kPa, the compressor power consumption is 37.5 W and the pulse tube power consumption is 298.8 W [8].To obtain a large cooling capacity of 5 mW at 2.2 K, our group is developing a hybrid JT cryocooler with a two-stage pulse tube and a set of four-stage DC compressors.

2 The 2 K Joule-Thomson Cryocooler 2.1 System Structure and Process The 2 K hybrid JT cryocooler schematic is shown in Fig. 1. For pre-cooling, the 2 K-JT system employs a two-stage integrated thermally coupled pulse tube, which is driven by two compressors. When the input electric powers of the first and second stage pulse tube are 212 W and 156 W, the first and second stages obtain 72.35 K and 9.87 K noload refrigeration temperature respectively. With the same input power, the second stage cold head can provide 0.2 W of cooling above 12 K, and the first stage can provide more than 2 W of cooling at 80 K. The four-stage linear DC compressors with different piston diameters are equipped with an arrangement of reed valves and buffer volumes to produce a steady flow of gas, which can provide a suction pressure of 3.8 kPa and pressure ratio below 200 for the JT flow path. The JT flow path is composed of several heat exchange components (three tube-in-tube helical coil counter-flow heat exchangers, two wire mesh-filled pre-cooled heat exchangers, and a pin-ribbed evaporator) with a micro-orifice. To reduce the heat leakage in the 2 K temperature zone, the structure does not use the bypass structure in the design of the structure. 2.2 The Design Value of the JT Cooler To achieve the cooling performance of 5 mW@ 2.2 K with 4 He, the design temperature of the secondary precooling can be 12–15 K. Evaporative pressure needs to reach about 5.3 kPa of suction pressure (2.2 K corresponds to saturation pressure). The design target of the low-pressure drop of heat exchangers is less than 1 kPa to effectively reduce

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Fig. 1. Schematic of the two-stage pulse tube precooling JT hybrid cryocooler

the load of the compressors. The design high pressure is less than 0.8 MPa due to the limitation of the pressure ratio (0.9

Tpre1

80 K

ε1

>0.9

Pl

100 K

>100 K

25

14.56

14.74

50

16.07

18.75

100

16.56

21.2

certain pressure drop. On the premise that the pressure before throttling is not higher than the converting pressure, the greater the resistance through the orifice, the greater the pressure drop generated by throttling, which amplify the integral throttling effect. The amplification of the integral effect is able to improve the performance of the JT cooler. Therefore, increasing the resistance of the throttle element is an effective way to improve the coolers’ performance [6]. On the other hand, the greater the resistance of the throttle, the greater the irreversible losses and the less the flow rate of the cycle. Hence the optimum throttling orifice diameter exists for the JT cooler, as shown in the Fig. 4.

Fig. 4. The cool-down time and flow rate vary with valve’s area

Fig. 5. The comparison cool-down time curves for the single and dual orifice

In order to enlarge the cooling capacity and shorten the cool-down time, it could be achieved by increasing the passageway area of a JT valve. At the same time, the reduction

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in the resistance through the orifice is due to the increasing the area of the throttle orifice, which leads the throttle cooling effect to become worse. When their areas are the same, according to the Fig. 5, the cool-down time of the dual orifice and the single orifice cooler is 8.7s@100K and 11s@100K respectively. The reason is that the former has a smaller aperture and a larger pressure drop, which produces a better throttling effect. 3.3 Finned Tube Exchanger The heat exchanger for the JT cooler generally employ a capillary tube with fins spirally wound on the mandrel, and the channel between the finned tube and the cold finger is used as a return channel, which attain the purpose of cooling the high-pressure fluid. The geometric structure parameters are mainly fin and spiral parameters, as depicted in Fig. 6.

Fig. 6. The schematic diagram for the finned tube. h- Fin height; δ- Fin thickness; t- Fin pitch; d2 - Outer diameter of capillary tube; d1 - Inner diameter of capillary tube

Fig. 7. The contrast between two finned tubes in the imager

The fins enlarge the heat transfer area, which can be expressed in terms of the ribbing coefficient  [7], calculated as:   d22 (d2 +2h)2 − 4 π (d2 + 2h)δ + 2π 4 2h(d2 + δ + h) =1+ (3) = π d2 (δ + t) d2 (δ + t) To further shorten the cool-down time, the comparative experimental study is carried by two fin forms. The finned tube sample morphology is observed under the imaging

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instrument, as shown in Fig. 7 and its specific dimensional parameters are listed in Table 2. The ribbing coefficient  obtained by the parameters of rectangular and circular fins are 3.91 and 1.59 respectively, which means that the heat transfer area is increased by 2.46 times per unit length, compared to circular fins. Table 2. The Geometrical parameters for the different types of finned tube Type

d2 /mm

h/mm

δ /mm

t/mm

rectangle

0.45

0.3

0.08

0.3

circular

0.45

0.2

0.2

0.33

Under the conditions of constant heat exchanger length and spiral parameters, the above-mentioned heat exchangers with different fin forms were used to make the cooler, and the experimental results are shown in Fig. 8. From the results, it can be seen that in case of the double-layer rectangular fins to increase the heat transfer area 2.46 times, the cool-down time of the cooler declines 5.33s@100K to 4.57s@100K.

Fig. 8. Temperature variation curve about different fins

4 Performance Testing and Applications The developed cooler is applied to the 128 × 128 Mid Wave infrared detector, and the diode is arranged near the chip for monitoring the focal plane temperature. Under the ambient temperature conditions, the focal plane temperature of the detector is shown in Fig. 9. From the test curve, the cool-down time of the focal plane is 4.57s@100K, the thermal storage time below 100 K is 74 s. In addition, the imaging effect of the detector is good, and the non-uniformity of the corrected image is less than 6 mV. The actual imaging effect is shown in Fig. 10.

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Fig. 9. The cooling curve for the detector’s focal plane

Fig. 10. The actual output images for the infrared imaging

5 Conclusion This paper introduces a miniature cooler and the optimization study of its key components. The research results include: (1) The volume of the gas cylinder has an important influence on the cool-down time, and it is shortest when a 25 mL high-pressure gas cylinder is used; (2) The optimum throttling orifice size of the throttling element exists for the cool-down time. Besides, for the same area, the dual orifice throttling has a shorter cool-down time than the signal orifice throttling; (3) Optimization of fin form and size parameters, the heat exchanger of the cooler adopts rectangular finned tube which is double winding, and the cool-down time is reduced to 4.57s@100K" 4.57s@100K, which uses the argon as the working fluid; (4) The developed cooler is applied to the 128 × 128 Mid Wave infrared detector, and the imaging effect is good. Besides, the non-uniformity of the corrected image is less than 6 mV.

References 1. Maytal, B., Pfotenhauer, J.: Miniature Joule-Thomson cryocooling: principles and practice. Springer, New York (2012). https://doi.org/10.1007/978-1-4419-8285-8 2. Bonney, G., Stubbs, D.: Design fundamentals of rapid cool-down Joule-Thomson (JT) cryostats and sensors. In: Cryogenic Optical Systems and Instruments VI, vol. 2227, pp. 98–108. International Society for Optics and Photonics, June 1994

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3. Xu, H., Yang, H.: The development of a fast cooling-down conical JT cooler. Chin. Low Temp. Phys. Lett. 37(2), 157–160 (2015) 4. Maytal, B.: Fast cooldown Joule-Thomson cryocooling for infrared detectors. May 19th and 20th, 1998 Technion—Israel Institute of Technology Haifa, Israel, 387 (1998) 5. Ardhapurkar, P., Atrey, M.: Performance optimization of a miniature Joule-Thomson cryocooler using numerical model. Cryogenics 63, 94–101 (2014) 6. Xie, J.: A fast cool-down JT minicryocooler. In: Advances in Cryogenic Engineering, pp. 621– 627. Springer, Boston, MA (1984) 7. Bergman, T., Bergman, T., Incropera, F., Dewitt, D., Lavine, A.: Fundamentals of Heat and Mass Transfer. Wiley (2011)

Research on Mini-Channel Heat Exchanger of Pulse Tube Expander Refrigerator Wenhui Cui1,2 , Linghui Gong1(B) , Qiming Jia1 , Weiping Zhu1 , Zhengyu Li1 , Xiujuan Xie1 , Meimei Zhang1 , Ming He1 , and Han Zhou1,2 1 State Key Laboratory of Technologies in Space Cryogenic Propellants, Technical Institute of

Physics and Chemistry CAS, Beijing, China {cuiwenhui20,zhouhan19}@mails.ucas.ac.cn, {lhgong,jqmipc, zhuweiping,lizhengyu,xiexiujuan,mmzhang,heming}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. The refrigeration system with pulse tube expander is a new type of cryogenic refrigerator that can achieve high reliability in the low-temperature region. The heat exchanger in the pulse tube expander refrigerator needs to simultaneously realize two functions of heat transfer and storage because cold and hot fluids are in different channels, and do not synchronize through the heat exchanger. In this paper, numerical simulation was carried out to evaluate the affecting factors of the performance of CFMCHE, which include the equivalent diameter, wall thickness, tube length, and wall materials. Results show that for the same mass flow rate, increasing the equivalent diameter leads to a significant decrease in pressure drop, and there is an optimal equivalent diameter to maximize the heat transfer effectiveness. And the increase in wall thickness leads to a decrease in heat transfer effectiveness. Moreover, the increase in heat transfer tube length leads to an increase in both effectiveness and pressure drop. In addition, for the same geometric structure, among the materials of stainless steel, aluminum, and copper, the best wall material of CFMCHE is stainless steel. Keywords: Pulse Tube Expander Refrigerator · Mini-channel Heat Exchanger · Numerical Simulation · Thermal Performance

1 Introduction In recent years, research in the field of cryogenic refrigeration has been paid more and more attention domestic and overseas. The development of many new technologies and the implementation of large scientific projects requires a cryogenic environment. However, in the liquid helium temperature area, the development of refrigerators with a cooling capacity in the range of 10–100 W is still blank. The refrigeration system with pulse tube expander based on the reverse Brayton cycle is currently the most potential cryogenic refrigerator to meet this demand. Instead of a piston or turbine, its expansion process is carried out via pulse tubes. Due to the reduction of moving parts, pulse tube expander refrigerators can achieve highly reliable operation in the low-temperature © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 716–723, 2023. https://doi.org/10.1007/978-981-99-6128-3_93

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region [1]. The system diagram of the pulse tube expander refrigerator with MCHE is shown in Fig. 1. The intake and exhaust of the pulse tube are divided into two lines controlled by the cryogenic valves. The performance of the entire refrigerator is not only closely related to the performance of the pulse tube expander but also significantly depends on the performance of its matching heat exchanger. The flow in the hot and cold channels from the matching heat exchanger is intermittent and periodic in the working process. As a result, the traditional partition heat exchanger cannot be used. Compared with the conventional heat exchanger, the mini-channel heat exchanger (MCHE) has the advantages of a compact structure and high heat transfer performance. Besides, its metal wall has the considerable capacity of storing heat. Therefore, it is reasonable to use MCHE to achieve fluid heat exchange in the pulse tube expander refrigerator.

Fig. 1. The system diagram of pulse tube expansion refrigerator with MCHE

The flow and heat transfer characteristics in mini-channels have been studied extensively in the previous researches [2–5]. But there are few researches on the periodic flow and dynamic heat transfer characteristics of two fluids in MCHE. Based on the previous researches, the heat transfer performance of MCHE influenced by different equivalent diameter, tube length, wall thickness, and wall materials has been studied in this paper. And then, the design and optimization of the mini-channel heat exchanger can be achieved through the calculation results of numerical simulation.

2 Physical Model The method of countercurrent heat transfer in the MCHE was adopted. It is composed of several independent neatly arranged rectangular mini-channels which are machined on one metal block. The cross-section of the CFMCHE is shown in Fig. 1, c and h represent the cold and hot fluid channels, which are arranged alternately in layers. Due to its periodic structure, an individual heat exchange unit is used to establish a model for this research. Its geometric model is shown in Fig. 2, ‘a’ is the width of the rectangular mini-channel section, ‘b’ is the length of the rectangular mini-channel section, ‘D’ is the equivalent diameter of the rectangular min-channel, ‘δ’ is the wall thickness, and ‘L’ is the length of the rectangular mini-channel. The geometrical dimensions of all mini-channel heat exchange units used in the calculation are listed in Table 1.

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W. Cui et al. Table 1. Geometric dimensions of heat exchange unit

serial number

material

a (mm)

b (mm)

D (mm)

δ (mm)

L (mm)

1

stainless steel

0.4

240

0.2

2

0.36

2

0.5

5

0.91

3

0.8

8

1.45

4

0.11

11

2

5

0.14

14

2.55

6

0.5

5

0.91

0.2

7

0.6

8

0.8

9

1

10

0.4

220

11

230

12

250

13

260

14

80

15

copper

16

aluminum

Fig. 2. Cross-sectional schematic of a CFMCHE

Fig. 3. A schematic of a heat exchange unit

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3 Numerical Model and Solution 3.1 Model and Grid To study the flow and heat transfer of helium in mini-channels with rectangular crosssection, the CFD fluid analysis software Ansys Fluent was used. And then a 3D numerical simulation model of the heat exchanger unit was built. The grid independence verification was carried out. Figure 3 shows the grid division of the model (Fig. 4).

Fig. 4. The schematic diagram of model meshing

3.2 Calculation Method and Boundary Conditions The equivalent diameter of the fluid channels studied in this paper is in the range of 0.2 mm to 3 mm, according to the research of S.G. Kandlikar et al., its scale belongs to mini-scale [6]. The ideal gas assumption for working fluid helium which in the mini-scale channel are not affected by the rarefaction effect of the gas, the continuum assumption is still satisfied. The viscosity model in the solver uses Laminar, steady-state calculation was used to analyze the influence of mini-channel geometry on heat exchanger performance, and transient calculation was used to analyze the influence of wall material on heat exchanger performance. The boundary conditions are: 1. Inlet boundary: mass flow inlet. According to the working characteristics of the pulse tube expander, in the steady state simulation, the mass flow rate of hot and cold fluids was both set to 0.032 g/s. In the transient simulation, the change of the inlet mass flow rate with time was set by writing UDF. The hot fluid inlet temperature is 300 K and the cold fluid inlet temperature is 77 K. 2. Outlet boundary: pressure outlet. The outlet pressure of hot fluid is 2 MPa, and the outlet pressure of cold fluid is 0.9 MPa, regardless of backflow. 3. Wall boundary: The walls around the heat exchange unit were set as periodic boundary conditions, and the interface walls between the fluid and the solid were set as coupled.

4 Results and Analysis In this paper the heat exchanger effectiveness can be defined as the temperature efficiency which can be expressed as: ε=

Tco − Tci φ = φmax Thi − Tci

(1)

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Tci , Tco , Thi and Tho are the temperature of cold fluid inlet, cold fluid outlet, hot fluid inlet, and hot fluid outlet respectively. The total pressure drop in the heat exchange unit is defined as the sum of the pressure drop of both hot and cold fluids which can be expressed as: Pt = Pc + Ph

(2)

4.1 Effect of Equivalent Diameter Table 2 shows the numerical calculation results of heat exchange units with different equivalent diameters. With the increase of the equivalent diameter, the pressure drops decrease significantly. According to Fig. 5, The equivalent diameter has an optimal value to achieve the maximum heat transfer performance under the same conditions. Table 2. Calculation results of heat exchange units with different equivalent diameters D (mm)

a (mm)

b (mm)

T ho (K)

Ph (Pa)

Tco (K)

Pc ( Pa)

0.36

0.2

2

113.17

16070

273.28

33747

0.91

0.5

5

111.19

400

274.41

877

1.45

0.8

8

112.57

70

274.17

139

2.00

0.11

11

114.57

20

273.40

40

2.55

0.14

14

116.48

10

272.49

15

Fig. 5. Influence of equivalent diameter on heat transfer performance (δ = 0.4 mm, L = 240 mm, stainless steel)

4.2 Effect of Wall Thickness Table 3 shows the numerical calculation results of heat exchange units with different wall thickness. Figure 6 shows the effect of wall thickness on heat exchanger effectiveness. With the increase of wall thickness, the heat transfer effectiveness decreases significantly. The thermal resistance of radial direction will increase and thermal resistance of axial will decrease with the increase of wall thickness. As a result, the loss of axial heat conduction will increase and the heat transfer of radial direction will be weakened. Therefore, the heat transfer effectiveness decreases with the increase of wall thickness.

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Table 3. Calculation results of heat exchange units with different wall thickness D (mm)

λ (mm)

Tho (K)

Tco (K)

0.91

0.2

108.28

275.33

0.91

0.4

111.19

274.41

0.91

0.6

113.11

273.54

0.91

0.8

115.27

195.49

0.91

1.0

117.37

271.36

Fig. 6. Influence of wall thickness on heat transfer performance (D = 0.91 mm, δ = 0.4 mm, stainless steel)

4.3 Effect of Heat Exchanger Tube Length Table 4 shows the numerical calculation results of heat exchange units with different tube length. Figure 7 shows the effect of tube length on effectiveness and pressure drop of the heat exchanger. With the length of heat exchanger increases, the heat transfer of two fluids becomes more adequate. At the same time, the flow resistance will increase. Therefore, the effectiveness and pressure drop both increase with the increase of heat exchange tube length. Table 4. Calculation results of heat exchange units with different tube length D (mm)

L (mm)

T ho (K)

Ph (Pa)

Tco (K)

Pc (Pa)

0.91

220

112.41

380

272.62

804

0.91

230

111.53

400

274.41

840

0.91

240

111.19

420

274.17

877

0.91

250

109.96

430

275.36

912

0.91

260

109.27

450

276.13

948

4.4 Effect of Wall Material Table 5 shows the transient calculation results of the influence of the wall material on the effectiveness of heat exchanger. It can be seen that the heat exchanger made

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Fig. 7. Influence of tube length on heat transfer performance (δ = 0.4 mm, D = 0.91 mm, L = 240 mm, stainless steel)

of stainless steel has the highest effectiveness, followed by copper, and the worst is aluminum. Because the thermal conductivity of stainless steel is much lower than that of copper, the axial heat conduction in stainless steel wall is small and the effectiveness is high. Although the thermal conductivity of aluminum is also lower than that of copper, aluminum has a higher specific heat capacity. During the heat exchange process between the fluids and the aluminum wall, more heat is absorbed by the wall, and less heat is absorbed by the fluid which leads to the lower effectiveness. Table 5. Influence of wall material on heat transfer performance (δ = 0.4 mm, D = 0.91 mm, L = 80 mm) wall material

C p (J/(kg · K) [188.5 K]

λ(W/m · K) [188.5 K]

Tho (K)

Tco (K)

ε

stainless steel

404.9

copper

358.0

12.3

183.97

197.59

0.5203

408.8

190.77

191.08

0.4898

aluminum

814.4

163.6

201.66

165.74

0.4410

5 Conclusion In this paper, helium is used as the flow medium. Based on numerical simulation studies, the performance of rectangular mini-channel heat exchangers was analyzed with different geometries and materials. 1. For the CFMCH made of stainless steel, reducing the wall thickness can reduce the axial heat conduction which can increase the effectiveness of heat exchanger. 2. Under the same geometric parameters, increasing the tube length of the heat exchanger will increase the effectiveness of heat exchanger and pressure drop. 3. With the increase of equivalent diameter, the pressure drop decreases significantly. The equivalent diameter has an optimal value to achieve the maximum effectiveness of the heat exchanger.

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4. For the CFMCH used in pulse tube expander refrigerators, the materials of heat exchanger really matter. According to the research of this paper, the optimal material of the exchanger is stainless steel. Acknowledgments. This work was supported by the fund of The National Key Research and Development Program of China (Grant No. 2020YFB1506201).

References 1. Jia, Q., et al.: Thermodynamic analyses and the experimental validation of the Pulse Tube Expander system. Cryogenics 91, 118–124 (2018) 2. Agostini, B., Bontemps, A., Thonon, B.: Effects of geometrical and thermophysical parameters on heat transfer measurements in small diameter channels. Heat Transf. Eng. 27(1), 14–24 (2006) 3. Liu, D., Garimella, S.V.: Investigation of liquid flow in microchannels. J. Thermophys. Heat Transfer 18(1), 65–72 (2012) 4. Hasan, M.I., Rageb, A.A., Yaghoubi, M., Homayoni, H.: Influence of channel geometry on the performance of a counter flow microchannel heat exchanger. Int. J. Thermal Sci. 48(8), 1607–1618 (2009) 5. Wang, H., Chen, Z., Gao, J.: Influence of geometric parameters on flow and heat transfer performance of micro-channel heat sinks. Appl. Therm. Eng. 107, 870–879 (2016) 6. Kandlikar, S.G., Grande, W.J.: Evolution of Microchannel Flow Passages-Thermohydraulic Performance and Fabrication Technology. Heat Transf. Eng. 24(1), 3–17 (2003)

A 150 K Micro Linear Split Stirling Cryocooler for High Operating Temperature Infrared Detectors Jian Sun, Yong Zeng, Taihe Huang, Li Huang(B) , Zehua Huang, Xiaoyong Li, Qianglong Zhu, Zhiming Zhang, and Yun Wang Wuhan Global Sensor Technology Co. Ltd., Guandong Street, Wuhan, China [email protected]

Abstract. To meet the application requirements of high operating temperature (HOT) infrared detectors, a 150 K micro linear split Stirling cryocooler with small size, light weight and good performance (SWaP3 ) was developed. The cryocooler weigh less than 260 g, and the dimension of the compressor is 33 mm × 67 mm. A set of test platform was built in order to evaluate the performance of the cryocooler. The effects of driving frequency (80 Hz–100 Hz) and compressor piston stroke (2.5 mm–4.7 mm) on motor efficiency, P-V diagram, P-V efficiency, cooling capacity, input power and COP were experimentally investigated. When compressor piston stroke is 3 mm, maximum relative Carnot efficiency of 11.24% and 0.67 W @ 150 K cooling capacity are achieved with 5.8 Wac input power at 88 Hz. When compressor piston stroke is 4.7 mm, maximum cooling capacity of 1.03 W @ 150 K is achieved with 10.9 Wac input power at 90 Hz. The results not only comprehensively evaluate the performance of the cryocooler but also provide directions for further optimization work. Keywords: Linear Stirling Cryocooler · 150 K · Efficiency

1 Introduction With the development of the 3rd generation infrared technology, the operating temperature of infrared detectors has gradually increased, and the HOT (>120 K) Stirling cryocooler has been a research hotspot [1]. High operating temperature increases the Carnot efficiency of the cryocooler and reduces the heat loss of the Dewar, which enables the cryocooler to become smaller, lighter and more efficient (SWaP3 ) [2]. A number of HOT Stirling cryocoolers have developed by international cryocooler manufacturers, such as K590 (RICOR) [3], SX020 (AIM) [4], LC1076 (Cobham) [5], UP8197 (Thales) [6], LF-100 (FLIR) [7], etc. The products weigh between 180 g and 300 g, and relative Carnot efficiency of 4.9%–15.0% can be achieved with a cooling temperature of 120 K–150 K. The reports on China’s micro HOT Stirling cryocooler are scarce. In 2016, the C312 HOT Stirling cryocooler (Kunming Institute of Physics) weighting less than 400 g was reported. A cooling capacity of 0.4 W and relative Carnot efficiency of 8.5% can be achieved with 8 Wac input power at 110 K [8]. In 2018, the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 724–731, 2023. https://doi.org/10.1007/978-981-99-6128-3_94

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SFZ700 HOT Stirling cryocooler (the 16th Research Institute of CETC) weighting less than 400 g was reported. A cooling capacity of 0.8 W and relative Carnot efficiency of 9.0% can be achieved with 15 Wac input power at 110 K [9]. To sum up, developing HOT Stirling cryocoolers with low input power ( 10%) is a challenge. In this paper, a high-efficiency 150 K micro linear Stirling cryocooler was introduced, and the effects of driving frequency and compressor piston stroke on the cooling capacity, input power, P-V diagram and efficiencies of the cryocooler were experimentally studied.

2 The 150 K Micro Linear Split Stirling Cryocooler Figure 1 shows the 150 K micro linear split Stirling cryocooler. The cryocooler weigh less than 260 g, and the dimension of the compressor is 33 mm × 67 mm. The cryocooler consists of a moving-magnet linear compressor, a connecting pipe and an expander. In the linear compressor, some flexure springs with large radial-axial stiffness ratio were used to reduce dry friction between the piston and the cylinder, and the coils were isolated out of the working chamber to reduce gas pollution and enhance the reliability of the cryocooler, and two compressors were placed opposite to reduce vibration. The pneumatic expander adopted the back-pressure chamber structure to reduce the dead volume of the compression chamber and improve the efficiency, and a column spring was used to reduce the radial dimension of the expander.

Fig. 1. The 150 K micro linear split Stirling cryocooler

3 Test Setup Figure 2 shows the schematic diagram of test rig for the linear Stirling cryocooler. The cryocooler was driven by sinusoidal current, and the operating conditions can be changed by adjusting driving frequency and voltage amplitude. The electrical parameters of the cryocooler (such as power factor, input power, etc.) were captured and displayed by the power analyzer. The displacements of the compressor piston and the expander piston were collected in time by two laser displacement sensors, and the gas pressure in the compression chamber was collected in time by a dynamic pressure sensor. The dynamic

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data were transmitted, displayed and stored on the oscilloscope, and the sampling rate was 250 kHz. A diode thermometer was installed on the cold head of the expender to measure the cooling temperature. The cooling capacity was consumed by Joule heat of the electric heater, so cooling capacity was the sum of the electric heating power and the heat leakage of the Dewar.

Fig. 2. Test rig for the linear Stirling cryocooler.

Table 1 shows test conditions of the cryocooler. The first experiment fixed compressor piston stroke at 3 mm, and the driving frequency increased from 80 Hz to 100 Hz. The second experiment fixed drive frequency at 90 Hz, and compressor piston stroke increased from 2.5 mm to 4.7 mm. The cryocooler was filled with helium of 1.9 MPa. Table 1. Test conditions. Parameter (unit)

Test 1

Test 2

Refrigeration temperature (K)

150

150

Frequency (Hz)

80–100

90

Compressor piston stroke (mm)

3

2.5–4.7

4 Results and Discussion 4.1 The Gas Pressure and P-V Diagram The gas pressure in compression chamber of the cryocooler can be expressed as: Pt = P0 + Pm sin(2π ft + θ )

(1)

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Figure 3 shows the effect on the pressure amplitude (Pm ) and the phase different between the gas pressure and compressor piston stroke (θ ). Figure 3(a) shows that the pressure amplitude increases approximately linearly at first and then decreases slowly against frequency. The phase difference decreases linearly with frequency, and the decreasing rate is 0.683°/Hz. When the frequency is 95 Hz, the maximum pressure amplitude is 0.220 MPa, and the phase difference is 32.1°. Figure 3(b) shows that both the pressure amplitude and phase difference increase logarithmically with compressor piston stroke. When compressor piston stroke is 4.7 mm, the maximum pressure amplitude is 0.280 MPa, and the maximum phase difference is 41.2°.

Fig. 3. The pressure amplitude and the phase different

The initial volume of the compression chamber in the half compressor is V0 . After the cryocooler starts, the instantaneous volume can be expressed as: Vt = V0 + A ∗ Pt

(2)

The P-V diagram is an important diagram showing the output characteristics of the compressor. The area enclosed by the curve means the useful work (i.e. P-V power) transmitted by the compressor to the working gas in one cycle, and the slope of the curve is related to the variable index of the thermal process. Figure 4 shows the effect on P-V diagram. Figure 4(a) shows when the frequency increases from 85 Hz to 95 Hz, the P-V power and the variable index are nearly constant with a compressor piston stroke of 3 mm. Figure 4(b) shows when compressor piston stroke increases from 3 mm to 4.7 mm at 90 Hz, the P-V power keeps going up, and the variable index is almost constant. 4.2 Power Factor, Motor Efficiency and the P-V Efficiency Power factor, motor efficiency and the P-V efficiency are the key indexes to evaluate the performance of linear compressors. The power factor is obtained from the power meter. Since iron loss of the motor cannot be measured on the test platform, iron loss will be ignored in motor efficiency calculation, and it can be calculated as follows: ηmotor =

˙ in − I 2 R W ˙ in W

(3)

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Fig. 4. P-V Diagram

The P-V efficiency of this opposed linear Stirling cryocooler can be calculated as follows: ηPV =

2π fXPm Asinθ ˙ in W

(4)

Fig. 5. Power factor, motor efficiency and the P-V efficiency

Figure 5(a) shows that power factor, motor efficiency and the P-V efficiency increase first and then decrease with frequency. When compressor piston stroke is 3 mm, the maximum power factor of 100%, the maximum motor efficiency of 81.4% and the maximum P-V efficiency of 62.0% are achieved at 93 Hz, which is related to the resonance characteristics of linear compressors. Figure 5(b) shows that power factor slowly increases and then decrease against piston stroke, and motor efficiency hardly changes with the stroke, and the P-V efficiency increases parabolic with the stroke. 4.3 Cooling Capacity, Input Power, COP and Relative Carnot Efficiency Figure 6 shows the effect on cooling capacity and input power. Figure 6(a) shows that cooling capacity increases slowly and then decreases rapidly against frequency, which

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is because gas mass flow rate increases with frequency, and the efficiency of the heat exchanger rapidly reduced at high frequencies. The input power fluctuates with frequency, which is related to the resonance characteristics of linear compressors and the variation of the cooling load. Figure 6(b) shows that cooling capacity and input power increase approximately linearly with compressor piston stroke, which is because compressor sweep volume increases with the stroke. When compressor piston stroke is 4.7 mm, the maximum cooling capacity of 1.03 W @ 150 K can be achieved with 10.9 W input power at 90 Hz.

Fig. 6. Cooling capacity and input power

Figure 7 shows the effect on COP and relative Carnot efficiency. Figure 7(a) shows that COP and relative Carnot efficiency increase slowly and then decrease rapidly against frequency. When compressor piston stroke is 3 mm, maximum COP of 11.5%, maximum relative Carnot efficiency of 11.24% and 0.67 W @ 150 K cooling capacity are achieved with 5.8 Wac input power at 88 Hz. Figure 7(b) shows that both COP and relative Carnot efficiency decrease approximately linearly with the compressor piston stroke. When the compressor stroke increases from 2.5 to 4.7 mm at 90 Hz, COP decreases from 11.79% to 9.42%. It can be seen that adjusting the compressor stroke in a fixed frequency can reduce the efficiency of the cryocooler, and changing the compressor stroke and driving frequency at the same time can keep the cryocooler at high efficiency under variable cooling capacity. It is worth noting that the frequency corresponding to the maximum COP (88 Hz) is different from that corresponding to the maximum motor efficiency (93 Hz), which indicates that the cryocooler is currently not a good match. Optimization work of thermodynamic matching of the cryocooler will be carried out soon.

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Fig. 7. COP and relative Carnot efficiency

5 Conclusion A 150 K micro linear split Stirling cryocooler with a weight of less than 260 g for HOT infrared detectors was developed. The effects of driving frequency and compressor piston stroke on motor efficiency, P-V diagram, P-V efficiency, cooling capacity, input power and COP were experimentally investigated. When compressor piston stroke is 3 mm, maximum COP of 11.5%, maximum relative Carnot efficiency of 11.24% and 0.67 W @ 150 K cooling capacity are achieved with 5.8 Wac input power at 88 Hz, and the maximum power factor of 100%, the maximum motor efficiency of 81.4% and the maximum P-V efficiency of 62.0% are achieved at 93 Hz. When compressor piston stroke is 4.7 mm, the maximum cooling capacity of 1.03 W @ 150 K can be achieved with 10.9 W input power at 90 Hz.

References 1. Katz, A., et al.: Development and optimization progress with RICOR cryocoolers for HOT IR detectors. In: Infrared Technology and Applications XLI, vol. 9451, pp. 524–538. SPIE (2015) 2. Chen, X., Sun, H., Nie, X., Gan, Z.: Overview of micro-miniature Stirling cryocoolers for high temperature applications. In: Proceedings of 18th International Cryocooler Conference, pp. 115–120 (2016) 3. Gazit, R., et al.: Low SWaP video core for MWIR imaging. In: Infrared Technology and Applications XLV, vol. 11002, p. 110021W. International Society for Optics and Photonics (2019) 4. Mai, M., Rosenhagen, C., Ruehlich, I.: Development of single piston moving magnet cryocooler SX020. In: Proceedings of 18th International Cryocooler Conference, New York, pp. 65–71 (2014) 5. Squires, M.: Cobham microcooler for high temperature applications. In: Proceedings of 18th International Cryocooler Conference, pp. 73–77 (2014) 6. Willems, D., Arts, R., de Jonge, G., Mullie, J., Benschop, T.: Miniature Stirling cryocoolers at Thales cryogenics: qualification results and integration solutions. In: International Cryocooler Conference, Boulder, pp. 85–93 (2016) 7. Conrad, T., et al.: FLIR FL-100 miniature linear Stirling cryocooler development summary. In: IOP Conference Series: Materials Science and Engineering, vol. 755, no. 1, p. 012045. IOP Publishing (2020)

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8. Xi, Z., et al.: Overview of free piston Stirling cryocoolers for hot detectors. Vac. Cryogenics 24(03), 151–156 (2018). (in Chinese) 9. Hong, Q., et al.: Study on the miniature linear Stirling cryocooler. Cryo. Supercond. 47(02), 28–32 (2019). (in Chinese)

Calculation and Experimental Study of a 4 K Gas-Coupled Pulse Tube Cryocooler Driven by a Single Compressor Zhaozhao Gao1,2 , Biao Yang1,2 , Xuming Liu1,2 , Liubiao Chen1,2,3(B) , and Junjie Wang1,2 1 Chinese Academy of Sciences Key Laboratory of Cryogenics, Technical Institute of Physics

and Chemistry, 29 Zhongguancun East Road, Haidian District, Beijing, People’s Republic of China [email protected] 2 University of Chinese Academy of Sciences, No.19(A) Yuquan Road, Shijingshan District, Beijing, People’s Republic of China 3 Institute of Optical Physics and Engineering Technology, Qilu Zhongke, Licheng District, Jinan, People’s Republic of China

Abstract. High-frequency pulse tube cryocoolers (HPTC) usually adopt the thermal-coupled structure in which multiple cold fingers are cooled in stages to obtain temperatures below 4 K. In order to improve the compactness, a gascoupled HPTC driven by a single compressor capable of achieving liquid helium temperature was developed in this paper. A calculation model was established based on Sage. The calculation results show that a cooling temperature of 3.9 K and 4.5 mW/4.2 K can be obtained with an input power of 380 W when helium-4 is used as the working gas. At present, the lowest temperature of the developed prototype is 4.3 K, and the cooling capacity of 6.0 K is 40.3 mW. Then helium-3 was used to replace helium-4 to further improve the cooling performance, and the improvement of the cooling performance, as well as the changes in the operating parameters and structural parameters, were also calculated. The calculation results show that a cooling temperature of 2.6 K and 15 mW/4.2 K can be obtained with an input power of 380 W when helium-3 was used as the working gas. Keywords: High-frequency pulse tube cryocooler · Gas-coupled · Helium-3 · Liquid helium temperature

1 Introduction The high-frequency pulse tube cryocooler (HPTC) is an advantageous candidate for the small refrigeration device working in liquid helium temperature range due to its small size, low vibration and high reliability [1–3]. At present, the multi-stage thermalcoupled structure is often adopted for HPTC to obtain temperatures below liquid-helium temperature [4–7]. The gas-coupled structure, especially driven by a single compressor, is more compact than the thermal-coupled structure, but an important challenge is that © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 732–739, 2023. https://doi.org/10.1007/978-981-99-6128-3_95

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it is difficult to get down to liquid helium temperature. Changing the working gas is a potential way to obtain lower temperatures. Some theoretical analysis and experimental results show that better refrigeration performance can be obtained by using helium-3 instead of helium-4 as the working gas of HPTC [8–11]. In order to obtain a more compact structure and lower refrigeration temperatures, a gas-coupled HPTC working at the liquid-helium temperature range driven by a single compressor is developed in this paper. The structure of the developed HPTC and the experimental results obtained using helium-4 will be briefly introduced firstly. Then a calculation model was established to study the improvement of refrigeration performance, as well as the changes in the operating parameters and structural parameters when helium-3 was used instead of helium-4 as the working gas.

2 Modeling and Experimental Results of the Cryocooler The schematic of the developed gas-coupled pulse tube cryocooler working in the liquid helium temperature range is shown in Fig. 1(a). Multi-bypass, double-inlet, inertance tube and gas reservoir are applied to joint phasing in this compact structure. The main structural parameters for the numerical calculation of the cryocooler using Sage are summarised in Table 1. The calculation results show that the designed cryocooler can obtain a no-load temperature of 3.9 K, and the cooling capacity is 4.5 mW at 4.2 K with an input power of 380 W when helium-4 is used as the working gas. Based on the calculation results, the developed prototype is shown in Fig. 1(b). Some experimental tests and optimizations have been carried out using helium-4. At present, the lowest temperature that can be obtained is 4.3 K, and the cooling capacity at 6 K is 40.3 mW when the charge pressure of helium-4 is 2.25 MPa, the operating frequency is 25 Hz, and the input power is 475 W [12].

Fig. 1. Schematic (a) and photo (b) of the gas-coupled pulse tube cryocooler.

3 Optimization of Operating and Structural Parameters When Using Helium-3 Instead of Helium-4 as Working Gas Based on the above-mentioned gas-coupled HPTC driven by a single compressor, the change of refrigeration performance when using helium-3 instead of helium-4 as the working gas is calculated. During the calculation, only helium-4 was replaced with

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Parameter

Value

Diameter of Regenerator I, II and III

26.4 mm, 18.4 mm and 12.5 mm

Length of Regenerator I, II and III

35 mm, 50 mm and 45 mm

Specifications for Inertance tube I

2 mm + 3 mm + 5 mm

Specifications for Inertance tube II

1 mm + 2 mm

Volume of gas reservoirs I and II

600 cc and 80 cc

helium-3, and other structural parameters and operating parameters did not change. The calculation results are shown in Fig. 2. Unexpectedly, the refrigeration temperature of each stage of the HPTC increased when it was replaced with helium-3. Specifically, the temperature of the first stage increased from 93.53 K to 107.10 K, the temperature of the second stage increased from 24.85 K to 26.54 K, and the temperature of the third stage increased from 3.28 K to 3.95 K. The above results show that simply changing the working gas can not improve the refrigeration performance of the HPTC. Therefore, the operating parameters of the developed gas-coupled HPTC, such as the charging pressure and operating frequency, and the structural parameters, such as the size of multi-bypass, double-inlet, and inertance tube, are optimized to obtain better refrigeration performance.

Fig. 2. The no-load temperature of the cryocooler with different working gases.

3.1 Variation of Optimal Operating Parameters Corresponding to Different Working Gases Figure 3 shows the relationship between the temperature of the third stage and the charging pressure when the developed HPTC uses helium-3 and helium-4 as the working gas respectively. It can be found that when helium-4 is used as the working gas, the no-load temperature of the cryocooler increases slightly with the increase of the charging pressure, while when helium-3 is applied, the cryocooler can obtain the lowest temperature of 3.1 K at 2.3 MPa, and obtain 15 mW cooling capacity at 4 K. This calculation result

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means that helium-3 can obtain better refrigeration performance than helium-4, and its corresponding charging pressure is higher than that of helium-4.

Fig. 3. The relationship between the cooling temperature and the charging pressure.

The effect of the operating frequency on the no-load temperature of the cryocooler for the two working gases is shown in Fig. 4. It can be seen that when helium-3 is used as the working gas, the no-load temperature of the cryocooler is 2.6 K at the optimum frequency of 34 Hz, while the temperature for helium-4 is 3.2 K at the optimum frequency of 30.5 Hz. It can be concluded that when helium-3 replaces helium-4, reasonably increasing the operating frequency can improve the refrigeration performance.

Fig. 4. The influence of the frequency on the cooling performance under two working gases.

3.2 Variation of Optimal Structural Parameters Corresponding to Different Working Gases For the two working gases, the effect of multi-bypass opening on the no-load temperature of the cold head is shown in Fig. 5. It can be seen that when the working gas is changed from helium-4 to helium-3, the multi-bypass opening corresponding to the lowest temperature is almost unchanged. Figure 6 is the distribution of the gas flow of the two working gases through the multi-bypass. It can be found that the mass flow into the regenerator with the multi-bypass opening corresponding to the lowest temperature

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is also very close, accounting for about 40% of the total mass flow. The phenomenon that the optimal multi-bypass opening does not change with the change of the working gas is very beneficial to the optimization of the HPTC: the multi-bypass is located inside the coaxial HPTC and requires multiple experimental adjustments, based on the above calculation results, helium-4 can be used instead of helium-3 in the experimental optimization process to avoid the waste of expensive helium-3.

Fig. 5. The relationship between refrigeration temperature and multi-bypass opening when different working gases are used.

Fig. 6. Mass flows through multi-bypass when different working gases are used.

The effect of double-inlet opening on the no-load temperature of the cold head is described in Fig. 7. It can be seen that when the working gas is changed from helium-4 to helium-3, the double-inlet opening corresponding to the lowest temperature is also very close, which brings great convenience to the optimization of the double-inlet. After all, double-inlet also requires multiple experimental optimizations, and the use of helium-4 during the experiment can avoid the tedious work in the process of helium-3 charge and recovery, and also reduce the waste of helium-3 during the experiment. Through the above-mentioned process optimization of the cryocooler, the optimal operating parameters and structural parameters are listed in Table 2. And it is verified that helium-3 as the working gas can achieve better refrigeration performance than helium-4 after parameter optimization. In detail, along with the optimal parameters, the performance of the cryocooler with helium-3 used is illustrated in Fig. 8. When the input PV power is 380 W, the cryocooler

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Fig. 7. The influence of the double-inlet opening on the refrigeration performance under two working gases.

Table 2. Optimization results of the cryocooler with two working gases. Parameter

helium-3

helium-4

operating pressure

1.82 MPa

1.60 MPa

operating frequency

34 Hz

30 Hz

multi-bypass opening aperture

1.65 mm

1.60 mm

double-inlet opening aperture

0.75 mm

0.50 mm

can obtain a no-load temperature of 2.6 K. Moreover, with the input of 590 W power, the cold head can provide 25 mW of cooling capacity at 4.1 K.

Fig. 8. Refrigeration performance of the cryocooler with different input PV power when helium-3 is applied.

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4 Conclusion A compact gas-coupled pulse tube cryocooler driven by a single compressor is designed and manufactured. When helium-4 is used as the working gas, the experimental results show that under the input power of 475 W, the prototype achieves a no-load temperature of 4.3 K, and provides a cooling capacity of 40.3 mW at 6.0 K. To further improve the refrigeration performance of the cryocooler, the optimization of operating parameters and structural parameters with applying helium-3 to replace helium-4 is implemented by calculation. The calculation results show that when the working gas is changed from helium-4 to helium-3, in order to obtain better refrigeration performance, the operating parameters such as charging pressure and frequency need to be adjusted to a certain extent, while the structural parameters such as multi-bypass and double-inlet do not need to be adjusted. After optimization, the gas-coupled HPTC can achieve a no-load temperature of 2.6 K and a cooling capacity of 15 mW at 4.2 K with an input power of 380 W. The experimental study of helium-3 as the working gas of the cryocooler will be performed in our follow-up work. Acknowledgments. The project is supported by the National Natural Science Foundation of China (No. 12073058, U1831203), the China National Space Administration (No. D050104, D040305), the Key Research Program of Frontier Sciences, Chinese Academy of Sciences (No. QYZDY-SSW-JSC028), and the Youth Innovation Promotion Association of Chinese Academy of Sciences (No. 2019030).

References 1. Radebaugh, R.: Development of the pulse tube refrigerator as an efficient and reliable cryocooler. In: Proceedings of Institute of Refrigeration, 1999–2000, London (2000) 2. Liu, S.X., Chen, L.B., Wu, X.L., Zhou, Y., Wang, J.J.: 10K high frequency pulse tube cryocooler with precooling. Cryogenics 77, 15–19 (2016) 3. Wu, X.L., Chen, L.B., Liu, X.M., Wang, J., Zhou, Y., Wang, J.J.: An 80 mW/8 K highfrequency pulse tube refrigerator driven by only one linear compressor. Cryogenics 101, 7–11 (2019) 4. Zhi, X.Q., Han, L., Dietrich, M., Gan, Z.H., Qiu, L.M., Thummes, G.: A three-stage Stirling pulse tube cryocooler reached 4.26 K with He-4 working fluid. Cryogenics 5893-6 (2013) 5. Quan, J., et al.: 4 K high frequency pulse tube cryocooler used for terahertz space application. Chin. Sci. Bull. 59(27), 3490–3494 (2014) 6. Dang, H.Z., et al.: Investigations on a 3.3 K four-stage Stirling-type pulse tube cryocooler. Part B Experimental Verifications. Cryogenics 105, 103015 (2020) 7. Chen, L.B., Wu, X.L., Wang, J.J., et al.: Study on a high frequency pulse tube cryocooler capable of achieving temperatures below 4 K by helium-4. Cryogenics 94, 103–109 (2018) 8. Radebaugh, R., Huang, Y., Gallagher, A., Gary, J.: Calculated regenerator performance at 4 K with helium-4 and helium-3. In: Weisend, J.G., et al. (eds.) Advances in cryogenic engineering 52, pp. 225–234. American Institute of Physics, Chattanooga (2008) 9. Nast, T., et al.: Development of remote cooling systems for low-temperature, space-borne systems. Cryocoolers 14, 33–40 (2007) 10. Qiu, L.M., Han, L., Zhi, X.Q., et al.: Investigation on phase shifting for a 4K Stirling pulse tube cryocooler with He-3 as working fluid. Cryogenics 69, 44–49 (2015)

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11. Dang, H.Z., Zha, R., Tan, J., et al.: Investigations on a 3.3 K four-stage Stirling-type pulse tube cryocooler. Part A: Theoretical analyses and modeling. Cryogenics 105, 103014 (2020) 12. Liu, X.M.: Research on two-stage gas-coupled high-frequency pulse tube refrigerator in liquid helium temperature region. University of Chinese Academy of Sciences (2021)

Start-Up Model Predictions of a Beta-Type Free Piston Stirling Generator Mingqiang Lin1 , Chunyun Chi1(B) , Kexin Jiao1,2 , Guotong Hong1,2 , and Jian Mou1 1 Key Laboratory of Space Energy Conversion Technologies, Technical Institute of Physics

and Chemistry, Chinese Academy of Sciences, Beijing, China {mqlin,chichunyun}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. This paper presents the starting characteristic of a beta-type free-piston Stirling generator (FPSG). Firstly, based on the dynamic analysis of the FPSG, the start-up prediction model was established. Then, taking the start-up temperature of the FPSG, the influence of parameters such as heating temperature, charging pressure and displacer spring stiffness on the start-up of the FPSG were studied. As a result, the prediction models showed good agreements in tendency with experimental results. Keywords: Start threshold · Free piston Stirling generator · Dynamic model

1 Introduction Free piston Stirling generator (FPSG) is a reciprocating external heat engine working with the Stirling cycle processes of two isothermal processes connected by two constant volume processes. Since the major advantages of simple structure, high efficiency, high reliability, the FPSG has a broad application prospect, including deep space detectors isotope power, space nuclear power plants, the ground large solar power stations and domestic cogeneration [1]. Compared with traditional kinematic Stirling engines, the piston and displacer in the FPSG are free to oscillate without any mechanical linkages. Therefore, the dynamic instability in the FPSG is more complicated, especially the start-up characteristics [2, 3]. Mou et al. studied the starting characteristics of the FPSG from the perspective of power distribution and proposed three necessary conditions for starting [4]. In the work of Kwankaomeng et al., a mathematical model using Schmidt’s theorem and Newton’s second law of motion was analyzed and a pole placement method to study the start-up and performance of a FPSG was developed. The results shown that the system could work in the frequency range of 6.2–6.4 Hz [5]. As can be seen, many analytical and numerical models have been proposed to analyze stable operation of FPSG, among most of which the moving parts of FPSG were assumed can operate in the steady state or simplified as sinusoidal motion. Unfortunately, the investigation of whether the FPSG can be started is ignored. Simultaneously, it appears that little literature detailed indicators the start-up characteristics of FPSG. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 740–746, 2023. https://doi.org/10.1007/978-981-99-6128-3_96

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Therefore, the main purpose of this study is to predict the start-up of FPSG and find the starting threshold. Taking the start-up temperature of the FPSG as the reference index, an extensive parametric study of the effects of various operating conditions and structural parameters involving heating temperature, charge pressure, spring stiffness of displacer on the start-up of FPSG are performed. And the results will be discussed in the subsequent sections.

2 Derivation of the Numerical Model As seen in Fig. 1, the FPSG is mainly comprised of three spaces (an expansion space, a compression space and a bounce space), two pistons (a piston and a displacer), three heat exchangers (a heater, a regenerator and a cooler) and a linear alternator.

Fig. 1. Schematic diagram of Stirling engine dynamics.

The displacer and piston were each treated as moving rigid bodies that are ‘sprung’ to the engine casing via single mechanical flexure springs. Their governing equations of motion were derived by applying Newton’s second law are extracted as follows: md x¨ d + cd x˙ d + kd xd + (Ap Pc + Ar Po − Ad Pe ) = 0

(1)

mp x¨ p + cp x˙ p + kp xp + Ap (Po − Pc ) = 0

(2)

The main forces acting on the displacer include the compression, expansion and bounce space working fluid pressure forces, denoted by Ad Pe , Ap Pc and Ar Po respectively, and the flexure spring force kd xd , damping force cd x˙ d . Similarly, forces acting on the piston include the compression and bounce space pressure forces, denoted by Ap (Po − Pc ) and the flexure spring force kp xp . The piston is coupled to the generator magnet train that subjects it to the electromagnetic load force, which is coupled into the damping force cp x˙ p . Accordingly, damping force of piston includes mechanical damping and electromagnetic damping force.

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The total mass of the working gas in the FPSG is assumed to be constant and the gas is with a uniform instantaneous pressure over the working space. According to the ideal gas equation and isothermal assumption, the gas pressure equation are described as: P = MR/(Vc /Tk + Vk /Tk + Vr /Tr + Vh /Th + Ve /Th )

(3)

The nonlinear pressure equation is expanded by Taylor expansion, and the secondorder term is ignored. The instantaneous pressure in the working space obtained by linearization is simplified as: P − P0 ≈

∂P ∂P (Ve − Veo ) + (Vc − Vco ) ∂Ve ∂Vc

(4)

Note that two new parameters C 1 , C 2 are defined as the pressure change coefficients relative to the volume change of the expansion space and the compression space: C1 =

∂P MR =− ∂Ve Th



   Vco Vr ln(Th /Tk ) Veo −2 ∂P MR Vco Vr ln(Th /Tk ) Veo −2 + + ; C2 = =− + + Tk Th − Tk Th ∂Vc Tk Tk Th − Tk Th

The equations of motion for the FPSG can be rearranged as:            mp 0 x¨ p Cp 0 x˙ p L L xp 0 + + 1 2 = x¨ d x˙ d L3 L4 xd 0 md 0 Cd 0

(5)

(6)

where: L1 = kp − C2 (Ap − Ar )2 ; L2 = kp − C1 Ad (Ap − Ar ) + C2 (Ap − Ar )(Ad − Ar ) L3 = C2 Ar (Ap − Ar ); L4 = kd − C1 Ad Ar + C2 Ar (Ad − Ar )

3 Results and Discussion In this section, start-up characteristics of the FPSG is analyzed by root locus technique, which can be used to intuitively show the influence of different parameters in the vibration system on the roots of the characteristic equation for the system [6]. The start-up prediction model is validated on the 5 kW FPSG developed by our laboratory. The main operating conditions and design parameters of the 5 kW FPSG are summarized in Table 1. As shown in Fig. 2(a), two pairs of complex conjugates roots for the fourth-order characteristic Eqs.7 are plotted in the complex plane with the real parts along the horizontal axis and the imaginary parts along the vertical axis. According to Lyapunov stability criterion [7], when the real value of the roots develops to be negative, the system motion diminishs with time. In other words, the FPSG cannot be started. In contrast, the positive real value of the roots indicates that the system motion enhanced, which means the FPSG can be started. Therefore, the root with zero real part is defined as the starting critical point of the FPSG, and the corresponding hot-end temperature at this time is defined as the starting temperature.

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Table 1. Generator parameters. Value

Unit

Value

Unit

Heater length

90

mm

Displacer mass

1.56

kg

Regenerator length

55

mm

Piston mass

6.57

kg

Cooler length

60

mm

Displacer spring stiffness

380

kN/m

Displacer area

12867.96

mm2

Piston spring stiffness

180

kN/m

Rod area

490.87

mm2

Displacer damping load

20

Ns/m

12377.09

mm2

Piston damping load

220

Ns/m

Piston area

3.1 Varying Charge Pressure and Hot-End Temperature Figure 2(a) displays the root locus diagram for the hot-end temperature ranging from 50 °C to 300 °C when the charge pressure is 5.5MPa and the cooler temperature is 10 °C. As the hot-end temperature increases, it is observed that the dominant closed-loop poles gradually develops to the right half of the plane with the starting critical point at 172 °C under this working condition. It implies that higher temperature is beneficial to the start of the generator under the same working conditions, which is consistent with the actual experimental phenomenon. Figure 2 (b) shows the root locus predicted by increasing the hot-end temperatures from 50 °C to 300 °C at different charge pressures varied from 4.2 MPa to 5.8 MPa. As seen, the distribution trend of the root locus is basically the same under different charge pressures. The critical point of starting temperature corresponding to the zero point under different charge pressures is studied emphatically.

Fig. 2. Root locus diagram for varying values of the hot-end temperature within the range 50 °C~300 °C for (a) the charge pressures at 5.5MPa and (b) the various charge pressures

As shown in Fig. 3, increasing the charge pressure from 4.2 MPa to 5.8 MPa results in a rapid decrease in the starting temperature of generator. As the amount of gas shuttled for thermal expansion/contraction increases, more heat transfer is required in the corresponding heat exchanger to thermally expand or contract the working fluid, which

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leads to the decrease in hot-end temperature under the same heating power and stroke amplitudes of displacer and piston. For the comparison between the experimental and the numerical results, it can be clearly found that the simulation results are lower than the experimental one, with the maximum deviations at 11%, validating the model to some degree. It can be mainly attribute to the generator dynamics is linearly simplified in the numerical model, and the damping coefficients are all calculated using empirical relations.

Fig. 3. Starting temperatures at different charge pressures

3.2 Varying Springs Stiffness of Displacer As a spring vibrator system, spring stiffness of displacer have a great impact on the stable vibration of the FPSG. Figure 4(a) shows the root locus for varying values of the spring stiffness of displacer when the charge pressure is 5.5 MPa and the heater and cooler temperature are 500 °C and 10 °C. Since the roots are complex conjugate pairs and the left pole disappears quickly as a transient response due to its larger negative real part [6], we only concentrate on the right upper in this analysis. As the spring stiffness of displacer increases, the root locus firstly moves to right half of the plane, then gradually to the left, with a maximum real part, implying the largest allowable power output capability at this point (380 kN/m). In addition, with the further increase of spring stiffness of displacer, the root locus locates more to the left, and eventually all the roots have negative real values. It suggests that larger spring stiffness of displacer results in the suppression of system vibration and eventually leads to the generator stop. Finally, there exists an interval from 120 kN/m to 680 kN/m, only within which can start the generator under the current operating condition. In Fig. 4(b) the root locus of different spring stiffness of displacer are depicted as the hot-end temperature varying from 50 °C to 300 °C. It can be inferred that the distribution trend of the root locus are basically the same under different spring stiffness of displacer. Figure 5 illustrates the simulated and experimental start-up temperature against the spring stiffness of displacer. Within the calculated starting range, as the spring stiffness of the displacer increases, the start-up temperature increases accordingly, with a good

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Fig. 4. Root locus diagram for varying values of springs stiffness of displacer within the range 0 kN/m~1000 kN/m for hot-end temperature at 500 °C (a) and (b) within the range 50 °C~300 °C at different springs stiffness of displacer

match in the tendency. Note that when the spring stiffness of displacer is 100 kN/m and 800 kN/m, the generator fails to be started, which is consistent with the previously predicted results.

Fig. 5. Start-up temperatures versus spring stiffness of displacer

4 Conclusion In this study, a start-up prediction model by drawing the root locus from linear eigenvalue analysis on different parameters for the FPSG was developed and experimentally validated. Through detailed comparison of the experiments and simulation, some important conclusions are briefly summarized as follows: 1.Taking the start-up temperature of the generator as a reference index, within a certain range, increasing charge pressure is beneficial to the start of the generator in the same condition. For the spring stiffness of displacer, there is an optimum point (380

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kN/m) with the largest allowable power output capability and a start interval from 120 kN/m to 680 kN/m in which the generator can be started. 2.The numerical and experimental results are in a good agreement in tendency, which represents that the root locus theory can be an effective tool for predicting the start performance of FPSG at different operating conditions. Acknowledgments. The project is supported by National Key R&D Program of China (Number 2020YFB1901800).

References 1. Laazaar, K., Boutammachte, N.: New approach of decision support method for Stirling engine type choice towards a better exploitation of renewable energies. Energy Convers. Manage. 223, 113326 (2020). https://doi.org/10.1016/j.enconman.2020.113326 2. Chen, P., Yang, P., Liu, L., Liu, Y.: Parametric investigation of the phase characteristics of a beta-type free piston Stirling engine based on a thermodynamic-dynamic coupled model. Energy 219, 119658 (2021). https://doi.org/10.1016/j.energy.2020.119658 3. Lin, M.Q., Mou, J., Chi, C.Y.: A space power system of free piston Stirling generator based on potassium heat pipe. Front. Energy 14(1), 1–10 (2020) 4. Mou, J., Hong, G.: Startup mechanism and power distribution of free piston Stirling engine. Energy 123, 655–663 (2017) 5. Kwankaomeng, S., Silpsakoolsook, B., Savangvong, P.: Investigation on stability and performance of a free-piston Stirling engine. Energy Procedia 52, 598–609 (2014) 6. Sim, K., Kim, D.-J.: Development and performance measurements of a beta-type free-piston Stirling engine along with dynamic model predictions. J. Eng. Gas Turbines Power 139(11), 112806 (2017). https://doi.org/10.1115/1.4036967 7. Tavakolpour-Saleh, A.R., Zare, S.: An averaging-based Lyapunov technique to design thermal oscillators: a case study on free piston Stirling engine. Energy 189, 116127 (2019). https://doi. org/10.1016/j.energy.2019.116127

Simulation and Optimization of the Number of Thermal Buses in ADR Salt Pills Ping Liu1,2 , Yanan Wang1 , Jun Shen1,2 , Wei Dai1,2 , and Ke Li1(B) 1 Key Laboratory of Cryogenic Engineering, Technical Institute of Physics and Chemistry,

CAS, Beijing 100190, China [email protected], {wangyanan,jshen, cryodw}@mail.ipc.ac.cn, [email protected] 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. As an important part of the adiabatic demagnetization refrigeration, the design and structure of salt pills directly affect the performance of the ADR. The primary goal of designing salt pills is to provide the largest cooling capacity within the operating temperature range. For the low-temperature ADR stage, the use of hydrated salt is necessary. Because of the poor thermal conductivity of the hydrated salt, the thermal bus structure is the commonly used method to improve its performance. The determination of the number of thermal buses is a compromise of the effective thermal conductivity and the volume (mass) of the CPA crystal. Increasing the number of thermal buses will effectively increase the effective thermal conductivity of salt pills, but the reduction of paramagnetic salt mass may lead to the reduction of the effective cooling capacity of salt pills. To better guide the design and construction of salt pills, this paper uses COMSOL Multiphysics to simulate the adiabatic and isothermal demagnetization processes of salt pills under different numbers of thermal buses to observe the temperature distribution inside the salt pill, and then determine the optimal number of thermal buses to ensure salt pills can provide effective cooling capacity as much as possible within the working temperature range. Keywords: Adiabatic demagnetization refrigeration · Salt pill · Heat transfer · Numerical simulation · COMSOL Multiphysics

1 Introduction Adiabatic demagnetization refrigeration (ADR) is an important refrigeration technology in space exploration and other scientific research. It has advantages of high refrigeration efficiency, gravity independence, no moving parts, etc. A single-stage ADR consists of a salt pill, a heat switch, a magnet, and a heat sink. A typical thermodynamic cycle includes four processes: isothermal magnetization, adiabatic demagnetization, isothermal demagnetization, and adiabatic magnetization. During the isothermal magnetization, the heat generated by increasing the magnetic field of the salt pill is transferred to the heat sink by turning the heat switch ON; During © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 747–754, 2023. https://doi.org/10.1007/978-981-99-6128-3_97

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the isothermal demagnetization, the salt pill is isolated from the heat sink by turning the heat switch OFF and provides cooling power by decreasing the magnetic field. Several materials have been investigated for ADR, which can be divided into two categories: rare-earth oxides and fluorides, and hydrated paramagnetic salts. Rare-earth oxides and fluorides mainly refer to some oxides or fluorides of Gd, such as GGG (Gd3 Ga5 O12 ) and GLF (GdLiF4 ). GGG is the most commonly used salt in the hightemperature stage of ADR, due to its high thermal conductivity, stable chemical property, and magnetic ordering temperatures down to 380 mK. Hydrated paramagnetic salts such as CPA (CrK(SO4 )2 12H2 O) and FAA (Fe(SO4 )2 NH4 12H2 O), which have 10 mK and 26 mK magnetic ordering temperatures, are commonly used in the low-temperature stage[1]. However, the low thermal conductivity of hydrated paramagnetic salts is the biggest problem that affects the heat transfer performance. The common solution is to grow crystals around bundles of high thermal conductivity metal wires (gold, copper), which are called thermal buses, to increase the contact area between heat transfer structure and hydrated paramagnetic salts, and hence improve the effective thermal conductivity of salt pills [2]. However, a large number of thermal buses will possibly lead to more eddy current heating and reduce the paramagnetic salt mass in a certain volume. At present, there are few published numerical simulation studies on the trade-off between the effective thermal conductivity and the volume (mass) of the CPA crystal, which brings inconvenience to the initial design of salt pills. This research simulates the demagnetization processes (From 1 T/1 K to 0 T/100 mK, covering both the adiabatic process and isothermal process) of the CPA salt pill by using COMSOL Multiphysics. The proportional, integral, and derivative (PID) control is used to demagnetize the salt pill isothermally, and the optimal number of thermal buses is determined by the hold time length of the salt pill (reflecting the effective cooling capacity of the salt pill). In Sect. 2, the model including the geometric setting, heat transfer setting, and magnetic field setting will be introduced. In Sect. 3, the simulation results will be presented and discussed, and the conclusions will be stated in Sect. 4.

2 Simulation Model 2.1 Geometric Setting The structure of the CPA salt pill is shown in Fig. 1 left, which is mainly composed of an upper and lower stainless steel end cap, copper bar, paramagnetic material CPA, and thermal bus structure (copper wires). The dimensions of the model are listed in Table 1. The number of thermal buses n from 30 to 2000 are selected for simulation, respectively. The thermal buses are equally spaced, and the pitch of the wire is between 0.4mm and 2.5 mm. To simplify the calculation in COMSOL Multiphysics, the salt pill is divided into n cylindrical elements equally (neglecting the volume in between these elements) and a two-dimensional axisymmetric model is established to simulate the individual element, as shown in Fig. 1 right. All components use free triangular meshes. The grid independence is tested with a typical heat load, the number of thermal buses n = 300, and t = 1000 s. When the grids vary from a fine grid (the number of mesh is 4043) to a coarser grid (the number of mesh is 3213), the temperature of thermal buses changes from 99.650 mK to 99.651 mK. To

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ensure the accuracy of calculation, this model selects the fine grid for all the simulations, and the meshes are shown in Fig. 1 right.

Fig. 1. The structure of CPA salt pill (Left); Two-dimensional axisymmetric model of simplified CPA salt pill (Right); 1—Copper bar; 2—Lower stainless steel end cap; 3—CPA crystal; 4— Thermal buses (copper wires); 5—Upper stainless steel end cap; I—Heat flow P1 of the CPA salt pill; II—Thermal boundary resistance condition of the CPA salt pill.

Table 1. Geometric dimensions of the CPA salt pill Part

Size (not simplified)

Copper bar

Diameter:11 mm; Height:80 mm

CPA crystal

Diameter:20 mm; Height:65 mm

Copper wires

Diameter:0.25 mm; Height:65 mm

The upper and lower stainless steel end cap

Diameter:20 mm; Height:2 mm

3 Heat Transfer Setting 3.1 Internal Heat Source Setting According to the first law of thermodynamics for the salt pill: du = T ds + μ0 H dM The derivation of the specific total entropy can be defined as the following:     ∂S ∂S dS(T , H ) = dT + dH ∂T H ∂H T

(1)

(2)

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During the process of adiabatic demagnetization, the entropy change of the salt pill dS = 0, so the temperature change of the salt pill is [3]:   T δS dH (3) dT = − cH δH T Since the temperature of the salt pill remains stable during the isothermal demagnetization process, the first term of Eq. (2) is 0. The cooling power of the CPA salt pill during isothermal demagnetization is:   ∂S dH (4) dq = T dS(T , H ) = T ∂H T

4 Boundary Conditions and Initial Conditions During the adiabatic demagnetization process, all boundaries are set as adiabatic except that the r = 0 boundary is axisymmetric. During the isothermal demagnetization process, a heat flow P1 of 1, 3, 5, and 10 µW is set at the lower end of the copper bar (I in Fig. 1 right) to simulate the external heat load of the CPA salt pill, respectively. Since the salt pill model is simplified by n equal cylindrical division, the P1 input value in the COMSOL model is P1 /n µW. Thermal boundary resistance, i.e., the Kapitza thermal resistance, is one of the important factors affecting the heat transfer performance of the salt pill. According to measurement results [4], the thermal boundary resistance between CPA and copper can reach 30 × 10–4 /(AT3 ) (K/W), where A is the contact area. In the COMSOL model, an equivalent thin resistance layer is set between the CPA crystal and copper wires (II in Fig. 1 right) to simulate the thermal boundary resistance. The initial condition of the salt pill is 1 K@1 T.

5 Thermophysical Properties of Salt Pill According to statistical thermodynamics, the entropy of paramagnetic salt pill per mol magnetic ion can be calculated using the following equation [5]: 

      x  x x(2J +1) x(2J + 1) S x sinh = coth − (2J + 1) coth + ln 2 sinh 2 R 2 2 2 (5) where, x = μB gBeff /2k B T, Beff = (B2 + b2 )0.5 , g = 2 is the landé g-factor, μB = 9.27 × 10–24 J/T is the Bohr magneton, and k B = 1.38 × 10–23 J/K is Boltzmann’s constant, J = 3/2 is total angular momentum on each magnetic ion in CPA, b = b0 (1-e−(T /T 0)α ) is the background field generated by each ion’s nearest neighbors with b0 = 0.0841, T 0 = 0.0836, α = 0.7973 [5]. The thermophysical properties including thermal conductivity and specific heat capacity of CPA crystal, copper, and stainless steel are shown in Table 2 below:

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Table 2. Thermophysical properties of the materials in a salt pill Thermophysical properties parameters

Value

Thermal conductivity of CPA (W/cm/K)

0.013T 3 [6]

Special heat capacity of CPA (J/mol/K)

C B /R = (x/2)2 sinh−2 (x/2)-[x(2J + 1)/2]2 sinh−2 [x(2J + 1)/2][7]

Thermal conductivity of the stainless steel (W/m/K)

0.069 @ 1 K 0.305 @ 4 K [8]

Special heat capacity of the stainless steel (J/g/K) 460T + 0.38T 3 @ 1 ~ 10 K [9] 465T + 0.56T −2 @ 0.07 ~ 0.6 K [10] (0.00068T + 0.0000496T 3 )1000/63.5 [11]

Special heat capacity of copper(J/kg/K)

5.1 Magnetic Field Setting The starting magnetic field of adiabatic demagnetization is 1 T, and the demagnetization rate of adiabatic demagnetization is 0.01 T/s. When the temperature of the copper bar T reaches 105 mK during the adiabatic demagnetization process, the COMSOL Multiphysics starts the PID control of the salt pill. PID controller uses a closed-loop feedback mechanism to minimize the control error through proportional, integral, and derivative algorithms. The target temperature of the PID control T set is 100 mK. When the temperature of the copper bar T reaches 100 mK, the heat flow P1 at the lower end of the copper bar changes from 0 µW to P1 , and the salt pill enters the isothermal demagnetization process. To stabilize the temperature of the copper bar T, the PID controller compares the copper bar temperature T with the target temperature T set , and then controls the demagnetization rate Ht, and hence keeps the cooling power equal to the external heat load P1 so that the copper bar temperature T is held constant. The PID control algorithm for Ht is:

t Ht = Kp (T − Tset ) + Ki

(T − Tset )dt + Kd 0

∂ (T − Tset ) ∂t

(6)

K p is the proportional coefficient, which is taken as 10 or 5; K i is the integral coefficient, which is taken as 0.1; K d is the derivative coefficient, which is taken as 0. Usually, PI control can meet the system requirements.

6 Simulation Results 6.1 Temperature Distribution Through numerical calculation, the temperature distribution of the salt pill can be obtained. When the number of thermal buses is relatively small (taking n = 30 as an example), the temperature distribution of the salt pill at the beginning of the isothermal

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Fig. 2. Temperature distribution of the salt pill at the beginning of isothermal demagnetization when the number of thermal buses is n = 30 (Left), n = 1000 (Right)

demagnetization is shown in Fig. 2 Left. The temperature of the CPA crystal main body is 54.25 mK and the temperature of thermal buses is 99.87 mK. When the number of thermal buses is relatively large (taking n = 2000 as an example), the temperature of the CPA crystal main body is 99.58 mK and the temperature of thermal buses is 99.96 mK. It can be concluded from Fig. 2 that the increase in the number of thermal buses will effectively decrease the temperature gradient between CPA crystal and thermal buses. 6.2 Eddy Current Heating Eddy currents are loops of electrical currents induced in a conductor when it is exposed to a time-varying magnetic field due to Faraday’s law of induction. Eddy currents flowing within conductors produce heat which is called eddy current heating. Copper with high electrical conductivity will lead to large eddy current heating when it is placed in a rapidly changing magnetic field. The calculation method of eddy current heating is shown in the COMSOL user manual, which will not be illustrated in detail here [12]. The results show that the maximum eddy current heating is in the order of 10–17 ~ 10–18 µW, which is much less than the cooling power of the salt pill. Therefore, the effect of eddy current heating can be ignored in this model. 6.3 Hold Time During the process of ADR isothermal demagnetization, the temperature of the CPA salt pill remains constant, so the total heat in terms of Joule is expressed as Q = P1 × t (t is the hold time, P1 is the heat load (W)) and should be equal to the effective cooling capacity of the salt pill. For CPA salt pills with the same heat load (W) and different numbers of thermal buses, the longer the hold time means the greater the effective cooling capacity of the salt pill. The relationship between the hold time and the number of thermal buses of the CPA salt pill when the heat load P1 is 1, 3, 5, 10 µW, and T set = 100 mK is shown in Fig. 3.

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Fig. 3. The hold time of CPA salt pill in isothermal demagnetization process with heat load P1 = 1,3,5,10 µW, T set = 100 mK, and the number of thermal buses n varies from 30 to 200,

It can be found from Fig. 3 that for different heat loads, the variation trend of hold time vs. the number of thermal buses is nearly the same. When the number of thermal buses is relatively small, the thermal conductivity of salt pills is poor, resulting in a large temperature gradient between the CPA crystal main body and thermal buses. Taking P1 = 10 µW, n = 30 as an example, the temperature of the CPA crystal main body is 76 mK, which is much less than T set . The hold time is only 912 s, and the effective cooling capacity is 9.12 mJ. With the increase in the number of thermal buses and hence the improvement of the thermal conductivity of the salt pill, the hold time will increase rapidly and reach the maximum value of 1738 s when n = 300, the effective cooling capacity is 17.38 mJ. When n continues to increase, the hold time will slowly decrease due to the mass loss in a certain volume of the salt pill. Taking n = 2000 as an example, the hold time decreases to 1184 s, and the effective cooling capacity is 11.84 mJ. According to Fig. 3, when the heat load P1 is 1, 3, 5,10 µW, the optimal number of thermal buses is always between 200 and 300, and the optimal wire spacing is about 1 mm ~ 1.2 mm, correspondingly.

7 Conclusion The effective cooling capacity of the CPA salt pill is a trade-off between the effective thermal conductivity and the effective volume (mass) of the CPA crystal. The increase in the number of thermal buses (the decrease of the spacing between copper wires in a fixed salt pill volume) will effectively improve the thermal conductivity of the CPA salt pill, and its hold time will increase rapidly towards an optimum value. After that, continuing to increase n will decrease the hold time due to the reduction of effective mass. It is also found that, given the volume of the salt pill, different µW heat load has no obvious effect on the optimal number of thermal buses, which is about 200–300.

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Meanwhile, under the structure and demagnetization rate given in this paper, the eddy current heating is far less than the cooling capacity of the CPA salt pill. An actual CPA salt pill with a diameter of 20 mm, a length of 50 mm, and a thermal bus of 220 has been designed and manufactured. More experimental results and their comparison with the calculation results will be presented in the future. Acknowledgments. The project is supported by National Key Research and Development Plan under Grant No. 2021YFC2203303, the National Natural Science Foundation of China (Grant No. 51925605) and Scientific Instrument Developing Project of the Chinese Academy of Sciences (Grant No. GJJSTD20190001).

References 1. Wikus, P., Canavan, E., Heine, S.T., Matsumoto, K., Numazawa, T.: Magnetocaloric materials and the optimization of cooling power density. Cryogenics 62, 150–162 (2014). https://doi. org/10.1016/j.cryogenics.2014.04.005 2. Bartlett, J., Hardy, G., Hepburn, I.D.: Design and performance of a fast thermal response miniature Chromium Potassium Alum (CPA) salt pill for use in a millikelvin cryocooler. Cryogenics 65, 26–37 (2015). https://doi.org/10.1016/j.cryogenics.2014.11.004 3. Kitanovski, A., Tušek, J., Tomc, U., Plaznik, U., Ožbolt, M., Poredoš, A.: Magnetocaloric Energy Conversion: From Theory to Applications. Springer International Publishing, Cham (2015). https://doi.org/10.1007/978-3-319-08741-2 4. Suomi, M., Anderson, A.C., Holmström, B.: Heat transfer below 0.2°K. Physica 38(1), 67–80 (1968). https://doi.org/10.1016/0031-8914(68)90062-1 5. Peter, J.: Shirron:Applications of the magnetocaloric effect in single-stage, multi-stage and continuous adiabatic demagnetization refrigerators. Cryogenics. 62, 130 (2014) 6. Pobell, F.: Matter and Methods at Low Temperatures. Springer, Heidelberg (2007). https:// doi.org/10.1007/978-3-540-46360-3 7. van Dijk, H., Keesom, W.H.: Measurements of specific heats of iron ammonium alum in high magnetic fields and at liquid helium temperatures. Physica 7(10), 970–984 (1940). https:// doi.org/10.1016/S0031-8914(40)90154-9 8. Ho, C.Y., Chu, T.: Electrical resistivity and thermal conductivity of nine selected AISI stainless steels Report (1977) 9. Du Chatenier, F.J., Boerstoel, B.M., De Nobel, J.: Specific heat capacity of a stainless steel. Physica 31(7), 1061–1062 (1965). https://doi.org/10.1016/0031-8914(65)90148-5 10. Hagmann, C., Richards, P.L.: Specific heat of stainless steel below T = 1 K. Cryogenics 35(5), 345 (1995). https://doi.org/10.1016/0011-2275(95)95355-I 11. Du Chatenier, F.J., De Nobel, J.: Heat capacities of some dilute alloys. Physica 28(2), 181–183 (1962). https://doi.org/10.1016/0031-8914(62)90103-9 12. COMSOL homepage. http://cn.comsol.com/model/eddy-currents-970. Accessed 01 May 2022

10W@80K High Frequency Pulse Tube Cryocooler Enchun Xing1,2 , Qingjun Tang1 , Hou lei Chen1,2(B) , Yuexue Ma1 , Tianshi Feng1,2 , Yuan li Liu1,2 , Nailiang Wang1,2 , and Jinghui Cai1 1 Key Laboratory of Technology on Space Energy Conversion, Technical Institute of Physics

and Chemistry, Chinese Academy of Sciences, Beijing 100190, China [email protected] 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. For the space infrared detectors, there are high requirements for the volume and weight of the pulse tube cryocooler. However, the volume and weight of the pulse tube cryocooler with a cooling capacity of 10 W at the temperature range of 80 K to 120 K are relatively large nowadays. Thus, it is formidable to meet the needs of future space applications. Previous research demonstrates that the lightweight of the pulse tube cryocooler can be achieved by increasing the operating frequency of the cryocooler. Moreover, the efficiency of pulse tube cryocoolers used in the liquid nitrogen temperature can be improved by increasing the charging pressure and applying regenerator packing with a smaller hydraulic diameter. The effects of regenerator length, charge pressure, and regenerator packing on pulse tube cryocoolers are first analyzed in the paper. The cold fingers for pulse tube cryocooler with a length of 35 mm and a cold finger diameter of 20 mm are designed. A lightweight pulse tube cryocooler with a total mass of 2.7 kg and an operating frequency of 126 Hz is obtained by optimizing the factors affecting the efficiency of the cryocooler, such as the phase shifter and the charging pressure. The cooling capacity of 10 W can be obtained at 80 K when the electric power input is 250 W, and the relative Carnot efficiency of the pulse tube cryocooler is 10.9%. Keywords: lightweight · charging pressure · 126 Hz · 10W@80K

1 Introduction Because the cold finger of pulse tube cryocooler has no moving parts, it takes the advantages of compact structure, low mechanical vibration, high reliability and long service life [1]. After years of theoretical and experimental research, great progress has been made in the development of pulse tube cryocooler, and its structural components have been continuously improved, and its application field has gradually become wider. In recent years, with the development of aerospace applications, the structure and the function of infrared detectors have become more complicated, requiring a higher cooling capacity. However, increasing cooling power always increases of the volume and weight of the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 755–762, 2023. https://doi.org/10.1007/978-981-99-6128-3_98

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pulse tube cryocooler, whereas many special aerospace applications have strict size and weight restrictions. As a result, miniaturization of high efficiency pulse tube cryocooler is required urgently in space applications. Previous research shows that increasing the operation frequency of pulse tube cryocooler can be an efficient way to minimize their size and mass because a high operating frequency can increase the energy density [2]. Therefore, the research on 100 Hz pulse tube cryocooler has gradually sprung up at home and abroad. In 2007, NGAS reported the 100 Hz miniaturized coaxial pulse tube cryocooler. The total mass of the cryocooler is 782 g, and the cooling capacity of 1.1 W can be obtained at 77 K [3]. Later, NGAS improved it, and the maximum frequency can reach 144 Hz, When the operating frequency is 100 Hz and the input electric power is 49 W, the minimum temperature of the cryocooler reaches 47 K, and the cooling capacity of 1.3 W can be obtained at 77 K [4]. Technical Institute of Physics and Chemistry (TIPC, CAS) has been working on micro pulse tube cryocooler since 2010. In 2017, a pulse tube cryocooler with a weight of 1.6 kg and an operating frequency of 100 Hz has been developed, the cryocooler can obtain 2.1 W cooling capacity at 80 K under 45 W electric power input [5]. After that, the cooling capacity of 2.2 W is obtained through the optimization of the cold finger [6]. In 2009, Zhejiang University developed a linear pulse tube cryocooler with an operating frequency of 120 Hz and a weight of about 11 kg. When the cryocooler inputs 500 W electric power, the minimum temperature of the cryocooler can reach 49 K, the cooling capacity of 8.0W can be obtained at 78.6 K [7]. TIPC has carried out research on space pulse tube cryocooler for decades and has successfully carried several pulse tube cryocooler on space satellites. This paper will continue to use the mature development technology of space cryocooler in the laboratory, and adopt the method of increasing the frequency to carry out the research on the lightweight and high cooling capacity pulse tube cryocooler. In order to improve the operating frequency of the cryocooler, this paper first analyzes the influence of the length of the regenerator, the filler of the regenerator and the charging pressure of the pulse tube cryocooler, then designs the high-frequency large cooling capacity cooling finger, and optimizes the pulse tube cryocooler through experiments.

2 Design of Cold Finger 2.1 Design of the Length of Regenerator Figure 1 shows a phasor representation of mass conservation within the regenerator of a typical inertance type of pulse tube cryocooler. Because of the volume Vrg of the regenerator, the conservation of mass requires that the flow at the warm end will lead the flow at the cold end, and is given by Eq. (1). ˙ rg PV i2π f PVrg = m˙c + (1) RT RT where the bold variables represent time varying or phasor quantities, the mass flow at the hot end of the regenerator is m˙h , and at the cold end is m˙c . R is the gas constant per unit mass, T is the mean temperature of the regenerator, and P˙ is the rate of change of pressure in the regenerator, P is the dynamic pressure, f is the operating frequency of the cryocooler, i is the imaginary unit. m˙h = m˙c +

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It can be seen from Eq. (1) that when other parameters remain unchanged, the vertical phasor in Fig. 1 increases with the frequency, resulting in an increase in the mass flow at the hot end. In this case the losses of the regenerator become high. Thus, in order to maintain the efficiency of the regenerator, it is necessary to reduce the volume of the regenerator while increasing the operating frequency of the cryocooler, so that the vertical phasor caused by the frequency in Fig. 1 can keep from increasing significantly. It is generally believed that when the pressure in the middle of the regenerator in phase with the mass flow, the average flow in the regenerator is the smallest. At this time, the pressure at the hot end of the regenerator lags behind the mass flow about 30° and at the cold end leads the mass about 30°, the pressure loss of the regenerator is the smallest, and the efficiency of the regenerator is the highest.

Fig. 1. Phasor diagram representing conservation of mass in the regenerator

Fig. 2. Influence of the length of the regenerator on cop

2.2 Selection of Regenerator Matrix Because of the increase of the operating frequency of the cryocooler, the length of the regenerator decreases, the volume specific heat capacity decreases and the heat exchange area of the regenerator decreases. Equation (2) is the thermal penetration depth of helium, Eq. (3) is the viscous penetration depth of helium. Where the k is the thermal conductivity of the gas; ρ is the density; ω is the angular frequency; CP is the specific heat at constant pressure; μ is the viscosity coefficient of gas working medium. With the increase of the operating frequency of the cryocooler, the heat penetration depth of helium decreases. In order to enhance the heat exchange requirements, it is necessary to reduce the hydraulic diameter of the filler of the regenerator and increase the heat transfer area of the regenerator.   (2) δk = 2k ωρCP δμ =

  2μ ωρ

(3)

A large number of experimental studies show that the mixed filling method can improve the performance of the cryocooler. #635 mesh stainless steel wire mesh with the wire diameter of 18um and #500 mesh wire mesh with wire the diameter of 20 um are adopted for the regenerator packing of the 120 Hz high-frequency pulse tube cryocooler.

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2.3 Influence of the Charge Pressure The PV power output by the pulse tube cryocooler can be expressed as: WPV

1 = τ

τ p(t)v(t)Adt

(4)

0

After integration, the PV power output by the compressor is:   pd pxm AsinΦ WPV = Π fpd xm AsinΦ = Π f p

(5)

Equations (4) and (5) are PV power of pulse tube cryocooler [2]. Where the xm is the piston displacement of the compressor, pd is the dynamic pressure amplitude of the compressor, p is the average pressure of the compressor, and the ratio in brackets is the relative pressure, Φ Is the phase difference between piston displacement and pressure, A is the sectional area of the piston, v is the velocity of the piston. It can be seen from the equations that increasing the charging pressure of the pulse tube cryocooler can still keep the PV power output of the cryocooler constant after reducing the displacement of the compressor piston. Generally speaking, the volume and weight of the compressor are positively related to the displacement of the compressor piston. In order to maintain the efficiency of the regenerator. The phase angle of the hot end of the regenerator should remain constant, so the second term in Eq. (1) should remain constant. According to Eq. (1), the following Eq. (6) is derived [7]: ˙ rg pd Vrg PV f = =k RT mRT ˙ p(pd /p)2

(6)

From the Eq. (6) [7], it can be seen that increasing the charging pressure of the pulse tube cryocooler can slow down the decline of the cryocooler performance caused by the increase of the operating frequency of the cryocooler. According to the above analysis, the length of the regenerator is numerically simulated, using regen 3.3 developed by NIST. Figure 2 shows the variation law of the regenerator COP with the regenerator length when operating frequency is 120 Hz, charging pressure is 5 MPa, hot end temperature is 300 K and cold end temperature is 80 K. It can be seen that with the increase of pressure ratio, the length of regenerator corresponding to the highest point of regenerator COP decreases gradually. The pressure ratio provided by the miniaturized high frequency pulse tube cryocooler is about 1.2. Thus, the length of the regenerator corresponding to the highest point of COP is 35 mm.

3 Experimental Apparatus The experimental system mainly includes cooling water circulation system, vacuum system, data acquisition system and cryocooler system. The cooling water circulation system is used to maintain the temperature of the hot end of the cold finger at 300 K. The vacuum system has two main functions: one is to provide a better vacuum environment

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for the cryocooler system before working; the other is to maintain the external vacuum environment of the cold finger. The data acquisition system is mainly used to collect the cold finger temperature and the corresponding operating parameters of the compressor. The pulse tube cryocooler mainly includes linear compressor, coaxial cold finger and phase shifter (inertance tube and gas reservoir). The pulse tube cryocooler is shown in Fig. 3.

Fig. 3. Schematic of the pulse tube cryocooler (1. Laser displacement sensor 2. Observation window 3. Pressure sensor 4. Cold finger 5. Inertance tube 6. Gas reservoir 7. Compressor)

Fig. 4. Influence of the charging pressure on piston displacement

The displacement of the compressor piston is monitored by the laser displacement sensor, and the pressure wave at the outlet of the compressor is monitored by the pressure wave sensor. Considering that the connecting pipe between the cold finger and the compressor is short, it can be approximately considered that the pressure wave at the outlet of the compressor is the pressure wave in the compression chamber of the compressor, and the PV power of the piston calculated by measuring the phase difference between them, which can be used to evaluate the impact of the change of the charging pressure on the pulse tube cryocooler. The heating block is arranged at the cold end to simulate the thermal load. When the heat exchanger at the cold head of the cold finger is in the thermal balance state, the product of heating voltage and current is the cooling capacity at this temperature.

4 Experimental Results and Discussion 4.1 Influence of the Pressure on the Compressor Output and Performance of the Pulse Tube Cryocooler The influence of charging pressure on the output characteristics of the compressor and performance is experimentally studied. Figure 4 shows the piston displacement curve of the compressor under three different charging pressures. With the increase of the charging pressure of the cryocooler, the piston displacement of the compressor under the same input power decreases gradually. When the charging pressure of the cryocooler increases

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from 4 MPa to 6 MPa and the input power is 250 W, the displacement of the compressor piston decreases from 2.96 mm to 2.75 mm. Figure 5 shows the pressure amplitude value of the cryocooler under different input power. With the increase of charging pressure, the pressure amplitude value generated by the compressor under the same input power gradually increases. When the charging pressure of the cryocooler increases from 4 MPa to 6 MP and the input power of the cryocooler is 250 W, the pressure amplitude output by the compressor increases from 319.5 kPa to 447 kPa (Fig. 8).

Fig. 5. Influence of the charging pressure on pressure amplitude

Fig. 7. Influence of charging pressure on cooling capacity

Fig. 6. Influence of the charging pressure on PV power

Fig. 8. Influence of the phase shifter on the cooling capacity

The PV power provide by the compressor is calculated according to the data in Figs. 4 and 5, as shown in Fig. 6. It can be seen from the figure that the PV power provide by the compressor increases with the increase of the input power of the cryocooler, the PV power provide by the compressor increases with the increase of the charging pressure of the cryocooler, and the efficiency of the compressor increases. When the charging pressure of the cryocooler increases from 4 MPa to 6 MPa and the input power of the cryocooler is 250 W, the PV power of the compressor increases from 128 W to 166 W. Figure 7 shows the effect of the charging pressure on the performance of the pulse tube cryocooler when the input power of the cryocooler is 250 W. It can be seen from the figure that with the increase of the charging pressure of the cryocooler, the minimum temperature of the cryocooler increases. When the charging pressure of the cryocooler increases from 4 MPa to 6 MPa, the minimum temperature increases from 53.9 K to

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55.9 K. However, when the working temperature range is higher than 70 K, the performance of the cryocooler increases. When the load of the cryocooler is 10 W, the corresponding temperatures are 88.7 K and 96.4 K with the pressure of 6 MPa and 4 MPa respectively. And when the load is 15 W, the corresponding working temperatures are 104 K and 116 K with the pressure of 6 MPa and 4 MPa respectively. 4.2 Optimizing the Phase Shifter of the Cryocooler The combination of phase shifter case1 adopts copper pipes with inner diameter of 3 mm and 4 mm and the 270cc gas reservoir, in which the length of copper pipes is 0.5 m and 1 m respectively. Compared with case 1, the gas reservoir volume of case 2 is reduced to 100cc. When case 1 is adopted, the optimal operating frequency of the cryocooler is 114 Hz, and when case 2 is adopted, the frequency of the cryocooler is 126 Hz. When the input power of the cryocooler and the charging pressure are 250 W and 6 MPa, the thermal load of the cryocooler is 10 W. The working temperature of the cryocooler is 88.7 K when case 1 is adopted, and 80.1 K when case 2 is adopted. The relative Carnot efficiency is 10.9%.

5 Conclusion In this paper, the effects of regenerator length, charging pressure and regenerator packing on pulse tube cryocooler are analyzed theoretically. Then a pulse tube cryocooler with a cold finger length of 35 mm and a total weight of 2.7 kg is designed. The influence of the charging pressure on the PV power of the compressor and the performance of the pulse tube cryocooler is experimentally studied, and the charging pressure of the pulse tube cryocooler is determined to be 6 MPa. Finally, through the optimization of the phase shifter, an efficient and miniaturized pulse tube cryocooler is obtained. Under the input power of 250 W and the temperature of 80.1K, the cooling capacity of the cryocooler is 10 W, and the relative Carnot efficiency is 10.9%. Acknowledgments. The project is supported by The National Natural Science Foundation of China (No. 52106036), 2009ZYHG0003.

References 1. Radebaugh, R.: The development and application of cryocoolers since 1985. In: Proceedings of ICCR2003, Hangzhou, International Academic Publishers, pp. 858-870 (2003) 2. Radebaugh, R.O., Gallagher, A.: Regenerator operation at very high frequencies for microcryocoolers. AIP Conf. Proc. 823, 1919–1928 (2006) 3. Petach, M., Waterman, M., Tward, E., et al.: Pulse tube microcooler for space applications. Cryocoolers 14, 89–93 (2007) 4. Petach, M., Waterman, M., Pruitt, G., et al.: High frequency coaxial pulse tube microcooler. Cryocoolers 15, 97–103 (2009) 5. Xing, E.C., Chen, H.L., Tang, Q.J., et al.: Investigation on the 1.6 kg miniature coaxial pulse tube cryocooler (in Chinese). Vac Cryog. 23, 217–222 (2017)

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6. Xing, E.C., Chen, H.L., Tang, Q.J., et al.: 2.2W@80K Miniature coaxial pulse tube cryocooler. J. Eng. Thermophys. 39, 484–488 (2018) 7. Gan, Z.H., Wu, Y.Z., Yuan, Y., et al.: Theoretical and experimental study on a 120 Hz single stage pulse tube cryocooler. J. Zhejiang Univ. (Eng. Sci.) 45, 2014–2019 (2011)

Theoretical and Experimental Investigation on Resonance Characteristics of a Dual-Coil Linear Compressor for J-T Throttle Cryocooler Z. J. Huang1,2 , Y. F. Niu1(B) , Y. J. Liu1(B) , E. C. Xing1,2 , Y. L. Liu1,2 , C. Zhang1,2 , and J. H. Cai1 1 Key Laboratory of Technology On Space Energy Conversion, Technical Institute of Physics

and Chemistry, Chinese Academy of Sciences, Beijing 100190, China {niuyuefeng,yjliu}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. In order to explore the relationship among the current, displacement, voltage of the linear compressor and the resonant frequency, a linear compressor was designed for the Joule-Thomson(J-T) throttling refrigerator. The linear compressor utilized the dual-coil to increase thrust. Firstly, the analytical solution of the displacement of the linear compressor under sinusoidal excitation was given by theoretical derivation. Then the changes of the current and displacement phase caused by different frequency excitations were revealed. It indicated that when the current caused the displacement to 90°, the linear compressor reached the resonance state. Besides, in order to verify the correctness of the inference, input same electric power at different frequencies, and test the phase angle of current and displacement. On the other hand, different pressure was charged into the system to test the performance of the compressor with nitrogen. Keywords: Resonance characteristics · Dual-coil linear compressor · J-T throttle cryocoolers

1 Introduction In space exploration, the temperature of the detector is higher than the cosmic background temperature, and the performance will be affected. Therefore, it is indispensable to reduce the temperature of the detector to ensure the accuracy of detection. For deep space detectors such as cosmic background detection, millimeter wave detection, and submillimeter wave detection, the temperature needs to be lowered to below 4 K [1]. The pre-cooled helium J-T throttling refrigerator is widely used in space exploration missions due to high refrigeration efficiency, simple structure and long life at liquid helium temperature. In the superconducting submillimeter wave detector (SMILES) developed in Japan in 2009, a J-T throttling cryocooler pre-cooled by a Stirling cryocooler is used to cool the core components of the detector. The refrigerator achieves a cooling capacity of 20 mW at 4.5 K [2, 3]. Glasgow University reported a small cryocooler based on © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 763–770, 2023. https://doi.org/10.1007/978-981-99-6128-3_99

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Stirling + J-T throttling for space applications in 2017, with a minimum temperature of 4.2 K [4]. Technical Institute of Physics and Chemistry (TIPC, China) successfully developed a J-T throttling refrigerator with two-stage pulse tube cryocooler pre-cooling for space-applied superconducting single-photon detectors in 2018. The lowest temperature of the refrigerator reached 2.6K [5]. As the core component of the J-T throttling refrigerator, the linear compressor is one of the key technologies in the field of space exploration. The laboratory has developed an opposed dual-piston moving coil linear compressor driven by double coils. The mass flow rate of this prototype is 92.76 g/min when the pressure ratio is 3.0. The characteristics of the resonance about the linear compressor are derived in this paper starting from the kinetic equations. The key parameters are measured, analyzed and explored with nitrogen as the working fluid, for instance, pressure ratio, electrical efficiency et al.

2 The Structure and Principle of Linear Compressor Figure 1 is the structure of a double-coil linear compressor. The linear compressor includes two linear motors, a support system and a cylinder-piston system.

Fig. 1. Structure diagram of dual-coil linear compressor (1. Visualization window 2. Exhaust valve 3. Suction valve 4. Flexure spring 5. Outer core 6. Inner core 7. Center frame 8. Cylinder 9. Piston 10. Spindle 11. Permanent magnet 12. Coil).

3 Theoretical Analysis The mover of linear compressor moves under the combined action of five forces which are gas force, electromagnetic force, damping force, spring force and inertia force. Four processes are repeated in the cylinder, namely compression, exhaust, expansion, and suction. The gas force can be obtained from formula 1. Fg (t) = [PA (t) − PB (t)]Ap

(1)

The gas force is periodic, and there are only a finite number of discontinuous points and extreme points of the first type in the same period, which satisfies the Dirichli convergence condition. It can be decomposed into an equivalent gas spring, a gas damper and a static force after expanding it with a Fourier series. Since the double pistons are arranged oppositely at both ends, the static force can be considered to be zero. The gas force can be expressed as Eq. 2. Equation 2 shows that the gas force can be equivalent to

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a spring force and a damping force. Therefore, the mover is under the combined action of damping force, spring force, inertial force and electromagnetic force, which can be regarded as a single-degree-of-freedom [6, 7], damped forced vibration system as shown in Fig. 2. Fg (t) = cg

dx + kg x dt

(2)

Fig. 2. The single-degree-of-freedom and damped forced vibration system

When the electromagnetic force is Fe = F0 sinωt, the dynamic equation can be expressed as Eq. 3. The analytical solution to Eq. 3 can be obtained mathematically as shown in Eq. 4. m¨x + c˙x + kx = F0 sinωt

(3)

x = e−ωn ξ t (C1 cosωd t + C2 sinωd t) + X sin(ωt − ψ)

(4)

Here, ωn is the natural angular frequency; ωd is the damped natural angular frequency; ξ is the damping ratio; cc is the critical damping. X is defined by Eq. 5.  X = A2 + B2 (5) 

 k−mω2 F0

cωF0 ;B = − . Since the initial conditions are (k−mω2 )2 +(cm)2 (k−mω2 )2 +(cm)2 x(0) = 0, x(˙o) = 0, it can be obtained C1 and C2 as shown in Eqs. 6 and 7.

where, A =

C1 = X sinψ

(6)

C2 = X (ωn ξ sinψ − ωcosψ)/ωd

(7)

Finally, Eq. 4 can be written as Eq. 10 from Eqs. 8 and 9. f(t) = e−ωn ξ t (C1 cosωd t + C2 sinωd t)

(8)

g(t) = X sin(ωt − ψ)

(9)

x = f(t) + g(t)

(10)

Figures 3 and 4 are the values of x, f(t) , g(t) calculated using MATLAB. As can be seen from Figs. 3 and 4, f(t) gradually decays with time, and g(t) changes sinusoidally.

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After dozens of cycles, x gradually keeps a steady state and changes sinusoidally. The response of the displacement x excited by the electromagnetic force with the operating frequency as 1/10 of the natural frequency is shown in Fig. 5. Displacement response x exists flutter when the linear compressor just starts to vibrate. But after a few cycles, the displacement response x reaches a steady state, and runs sinusoidally. The response of the displacement x excited by the electromagnetic force with an operating frequency 10 times of the natural frequency is shown in Fig. 6. The displacement response x is relatively obvious at the beginning, but gradually decays to a steady state after several cycles. However, the peak value of the displacement response x will still change according to the natural frequency after reaching the steady state, which is unfavorable for stable operation. Therefore, it is better to vibrate the linear compressor with an electromagnetic force lower than or equal to the natural frequency when starting, which is beneficial to the stable operation of the linear compressor and improve the efficiency and lifespan.

Fig. 3. Vibration response of linear compressor (f n = 50Hz, ω/ωn = 1).

Fig. 5. Vibration response of linear compressor (f n = 50Hz, ω/ωn = 0.1).

Fig. 4. Vibration response of linear compressor (f n = 20Hz, ω/ωn = 1).

Fig. 6. Vibration response of linear compressor (f n = 20Hz, ω/ωn = 10).

From the above analysis, it can be known that f(t) gradually decays with time. Therefore, it can be expressed as g(t) when the linear compressor is running steadily. The angle of the displacement lag sinusoidal excitation is given by Eq. 11. When ω = ωn , or β = 1, ψ = π/2. When the frequency of the electromagnetic force runs at the natural frequency of the linear compressor, the displacement lags the electromagnetic force by 90°. Currently, the current of the linear compressor (current and the electromagnetic force are in the same phase) leads the displacement by 90° and is in the same direction as the speed. ψ = −tan−1

B 2ξβ = tan−1 A 1 − β2

(11)

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4 Experimental Results and Analysis Figure 7 shows the photo of linear compressor. The length of the linear compressor is 540 mm, maximum diameter is 152 mm, Weight is 22.41 kg , maximum power is 500 W.

Fig. 7. The photo of linear compressor.

Fig. 8. Power factor and phase angle with different charging pressure.

4.1 Power Factor and Phase Angle The experimental system was charged with 0.2 MPa nitrogen to verify the results of the theoretical analysis in the previous section. The tests were carried out under the pressure ratio of 1.0, 2.0 and 3.0. The power factor and phase angle are investigated at different operating frequencies, and the results are shown in Fig. 8. As can be seen from Fig. 8, when the pressure ratio is 1.0, the phase angle is 91.6°, the power factor is 0.964, and the power factor reaches the maximum value. The phase angle is 91.3°, the power factor is 0.998 when the pressure ratio is 2.0. The phase angle is 91.7°, the power factor is 0.975 when the pressure ratio is 3.0. The power factor when the phase angle reaches 90° is basically close to the maximum value. It should be noted that due to current distortion, in a few cases, the phase angle is 90°, and in most experiments, the angle is 120°. 4.2 Motor Efficiency and Power Factor The loss of the moving coil linear motor is mainly the Joule heat generated by the coil, which is the main factor affecting the motor efficiency. The formula of the motor efficiency is shown as Eq. 12. ηm =

Pin − I 2 · R Pin

(12)

Under the conditions of fill pressure of 0.3 MPa, 0.5 MPa and 0.7 MPa, the relationship between motor efficiency and power factor with frequency was tested at different pressure ratios. Figure 9 shows the experimental results when the fill pressure is 0.3 MPa at pressure ratios of 2.5 and 3.0, respectively. It should be noted that this prototype will be unstable when it is around 48 Hz, and the data around 48 Hz will fluctuate,the reason for this phenomenon is that the leaf spring reaches the first-order modal resonance point near 48 Hz, and the leaf spring vibrates. It can be seen from Fig. 9 that the motor efficiency and power factor have the same trend with the increase of frequency. The

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maximum value of power factor is 0.992 and 0.887 when the pressure ratio is 2.5 and 3.0, corresponding to 45 Hz and 47 Hz. The motor efficiency arrives at the peak at 47 Hz and 48 Hz, which are 87.7% and 89.3%, respectively. Figure 10 shows the power factor and motor efficiency at different fill pressures when the pressure ratio is 3.0. Likewise, motor efficiency and power factor have similar trends. The power factor reaches the highest point at 47 Hz, 62 Hz and 73 Hz, respectively, and the motor efficiency also reaches the top near these frequencies.

Fig. 9. Motor efficiency and power factor at different pressure ratios with filled to 0.3 MPa.

Fig. 10. Motor efficiency and power factor at different fill pressures with the pressure ratio is 3.0.

Fig. 11. Volumetric efficiency at different pressure ratios with filled to 0.3 MPa.

4.3 Volumetric and Power Efficiency The volumetric efficiency is used to evaluate the degree of deviation from the theoretical capacity of the compressor. Power efficiency evaluates the completeness of the input power of the linear compressor. V 1/2π ·D2 · s · f    k−1 P2 k k m · Rg · T1 ηp = − 1 /Pin k −1 P1 ηv =

(13)

(14)

Figure 11 reveals the volumetric efficiency when the fill pressure is 0.3 MPa at different pressure ratios. It should be noted that this prototype will be unstable and the data will fluctuate when it is around 48 Hz. The reason for this phenomenon is that the flexure spring reaches the first-order modal resonance point and vibrates irregularly around 48 Hz. The volumetric efficiency gradually decreases as the frequency increases, and the higher the pressure ratio, the lower the volumetric efficiency, which is caused by the leakage caused by high pressure ratio. Figure 13 shows a consistent trend of volumetric efficiency with frequency at different inflation pressures when the pressure ratio is 3.0. At the same frequency, there is higher volumetric efficiency at higher inflation pressure (Fig. 12). Figure 13 demonstrates the power efficiency when the fill pressure is 0.3 MPa at different pressure ratios. It can be seen from the figure that as the frequency increases, the electrical efficiency first increases and then decreases. When the pressure ratio is 2.0, the maximum value is obtained at 41 Hz with a value of 78.1%. The maximum value

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Fig. 12. Volumetric efficiency at different fill pressures.

Fig. 13. Power efficiency at different pressure ratios with filled to 0.3 MPa.

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Fig. 14. Power efficiency at different fill pressures with the pressure ratio is 3.0.

is obtained at 47 Hz, and the value is 77.9% when the pressure ratio is 2.5. When the pressure ratio is 3.0, the maximum value is obtained at 49 Hz, and the value is 45.5%. Figure 14 shows that the electrical efficiency first increases and then decreases with frequency under the three charging pressures when the pressure ratio is 3.0. The prototype used in this paper trembles at 63 Hz, but it basically does not affect the data trend.

5 Conclusion 1) The analytical solution of the dynamic equation is given mathematically with the excitation of sinusoidal electromagnetic force is calculated and analyzed. Through calculating and analyzing the analytical solution, it is found that the linear compressor will not run stably after reaching a steady state when it is started at a frequency higher than the natural frequency. Therefore, it is more appropriate to start the experiment at a lower frequency than the natural frequency. 2) Provides help for finding the best operating frequency in the experiment. The frequencies at which the motor efficiency, mass flow per unit power, volumetric efficiency and power efficiency of the linear compressor reach the highest point all fluctuate around the same point under the same operating conditions. 3) The maximum values of power efficiency did not change much with increasing fill pressure. Therefore, it is beneficial to increase the fill pressure in practical to maintain a higher flow rate and high efficiency. Acknowledgments. The project is supported by the Strategic Priority Research Program of Chinese Academy of Sciences. (Grant No. XDA18000000, No. XDA18040000).

References 1. Yuexue, M., Juan, W., Yanjie, L., Jianguo, L., Jingtao, L.: Experimental research on the J-T orifice of a space 4.5 K hybrid J-T cooler. Chin. Sci. Bull. 63(18), 1839–1846 (2018) 2. Otsuka, K., Tsunematsu, S., Okabayashi, A., Narasaki, K., Satoh, R.: Test results after refurbish of cryogenic system for smiles. Cryogenics 50(9), 512–515 (2010)

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3. Shinozaki, K, Sugita, H, et al.: Developments of 1~4 K class space mechanical coolers for new generation satellite missions in JAXA. J. Phys. Chem. B (2008) 4. Gemmell, N.R., et al.: A miniaturized 4 K platform for superconducting infrared photon counting detectors. Supercond. Sci. Technol. 30(11), 11LT01 (2017) 5. Peng, Hu., et al.: Superconducting single-photon detector with a system efficiency of 93% operated in a 2.4 K space-application-compatible cryocooler. Supercond. Sci. Technol. 34(7), 07LT01 (2021) 6. Sun, J., Li, C., Li, J., Cai, J.: Experimental and theoretical investigation of nonlinear dynamic characteristics in an oil-free moving coil linear compressor. IOP Conf. Ser. Mater. Sci. Eng. 758(1), 012086 (2020) 7. Yin, W.: Dynamics analysis of air spring suspension system under forced vibration. China J. Highw. Transp. 19(3), 117–121 (2006)

Research on Performance of High-Power Linear Compressor for Completely Oil-Free Gifford-McMahon Cryocooler Zhijie Huang1,2 , Yuefeng Niu1(B) , Yanjie Liu1(B) , Yuanli Liu1,2 , Chen Zhang1,2 , Enchun Xing1,2 , and Jinghui Cai1 1 Key Laboratory of Technology on Space Energy Conversion, Technical Institute of Physics

and Chemistry, Chines Academy of Sciences, Beijing 100190, China {niuyuefeng,yjliu}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. The demand for completely oil-free Gifford-McMahon (G-M) cryocoolers has been increasing in recent years. Since linear compressors which is completely oil-free and high-efficiency have the good performance and great potential, it is planned to use the linear compressor as the helium compressor in the G-M cryocooler. The compressor in the G-M cryocooler has the characteristics of high power, large flow and high-pressure difference, which are all challenges for linear compressors. A moving coil motor for the linear compressor is designed to meet the requirements. The design of motor stroke is 36 mm to increase the flow rate and improve the volumetric efficiency. The magnetic circuit has also been specially designed to ensure the compressor stroke, and the electromagnetic thrust coefficient has been tested to be 79.85 N/A. The flexure springs with a stroke of 36mm are designed for the motor, but it is found that they are easy to break in the experiment which can lead to problems such as short circuits. The displacement response of the motor at different powers was tested by using a laser sensor, and the optimal resonant frequency was found. The research provides a reference and direction for designing high-power and long-stroke linear compressors. Keywords: Gifford-McMahon cryocooler · linear compressor · oil-free · high power

1 Introduction Currently, the helium compressor that uses oil lubrication is generally equipped with an oil separator and oil adsorber at the exit to separate the lubricating oil in helium. This not only improves the volume and weight of the whole helium compressor, but also restricts the application of the G-M cryocooler. The oil-free characteristics of linear compressors provide a new solution to this problem. Therefore, many research institutions turn their attention to high-power linear helium compressors [1, 2]. In 2005, John A. Corey of the United States took the lead in developing the oil-free linear compressor to apply the G-M cryocooler to the vehicle platform. The compressor © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 771–779, 2023. https://doi.org/10.1007/978-981-99-6128-3_100

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is moving magnetic structure, the maximum displacement is 26 mm, and the input power is 7.5 kW [3]. In 2015, Sumitomo developed a G-M cryocooler driven by an oil-free linear compressor to cool a superconducting single-photon detector. Since the linear compressor doesn’t need oil lubrication, the oil separator and adsorber are canceled in the system, which greatly reduces the volume of the whole system. The volume of original compressor is 70 L, and the volume of linear compressor is 37 L, reducing the height by 25% and the volume by 47% [4]. To improve the efficiency of G-M cryocooler, the cold head need to be provided large helium gas flow rate and high-pressure ratio. Besides, the stroke of the linear motor need to be improved for high efficiency [5]. Our laboratory has developed a high-power linear motor for G-M refrigerator, the maximum stroke is 36 mm.

2 Linear Motor Structure Figure 1 demonstrates the schematic diagram of the linear motor which is moving coil structure. The spindle in the coil drives the pistons on both sides, and the flexure springs are evenly arranged on both sides to support the coil. Figure 2 shows the physical picture of the compressor. The length is 839 mm, the maximum diameter is 173 mm, the mass of the single machine is 43.7 kg, the middle is a water-cooled sleeve, and the both ends are flexure springs support structures.

Fig. 1. Structure drawing of high-power linear motor. (1. Flexure springs 2. Spindle 3. Coil 4. Inner core 5. Permanent magnet 6. Outer core)

Fig. 2. Physical picture of high-power linear motor

2.1 Magnetic Flux Density Optimization The linear motor is designed according to the mature linear motor technology in author’s laboratory. The magnetic field is an important part in the design of linear motor. In order to optimize the design of magnetic circuit, a 2-D model is established in Maxwell. Figure 3 shows the profile of 2-D magnetic circuit, and the key size parameters are given in the figure. Including the length of magnet lm , the width of magnet sm , the length of middle segment lc , the thickness of air gap sd , radius of inner iron rn and the radius of outer iron rw .

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Fig. 3. Magnetic circuit diagram of linear motor

Fig. 4. Magnetic flux density cloud diagram of magnetic circuit

According to Kirchhoff’s voltage law on the coil in the linear motor, the voltage balance equation can be obtained: di + Ke dx Re i + Le dt dt = u

(1)

where Re is equivalent resistance, Le is equivalent inductance, Ke is electromagnetic thrust coefficient. Figure 4 shows the magnetic flux density distribution diagram of the magnetic field when the thickness of air gap is 10 mm. It can be seen from the diagram that the magnetic flux density is relatively uniform at both ends of the air gap, the average value is 0.631 T. Specific values of the linear motor were showed, the effect of the thickness of air gap on the magnetic flux density is simulated with other parameters unchanged. Figure 5 shows the simulation results. When the thickness of air gap increases from 8mm to 12 mm, the magnetic flux density in the air gap decreases from 0.707 T to 0.563 T. When the thickness of air gap is greater than 11 mm, the magnetic flux density is lower than 0.6 T. When the thickness of air gap is 10 mm and 11 mm, the coil can only be wound around 7 layers. Taking all into consideration, the air gap of 10 mm is the best value.

Fig. 5. Magnetic flux density under different thicknesses of air gap

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Fig. 6. The flexure spring of linear motor

Fig. 7. Stiffness curve of flexure spring

2.2 Design of Long-Stroke Flexure Springs The long-stroke flexure spring needs to be designed to meet the 36 mm stroke of linear motor. The flexure spring needs to have a special line to meet the requirement of axial stiffness and radial stiffness. The design of flexure spring lines were designed with Archimedes involute and the specific formulas are as follows:  x = r(cosθ + θ sinθ ) (2) y = r(sinθ − θ cosθ ) where r is the radius of base circle, θ is the involute angle of the circular involute. Figure 6 shows the flexure spring of linear motor. To reduce the increased stress caused by deformation, the diameter of the flexure spring is designed as 160 mm. The axial stiffness of the flexure spring is solved by ANSYS. The central displacement of the flexure spring is 11.98 mm in the axial direction, and the axial stiffness is 2.5 N/mm. Figure 7 illustrates the force of the flexure spring with growing displacement in the test, which can be regarded as a linear relationship when the stroke exceeds 6 mm. After calculation, the actual stiffness of the flexure spring is 3.15 N/mm, which has a difference of 20.63% from the simulated value (Fig. 8). The natural frequency of flexure spring is simulated by ANSYS, as shown in Table 1. The deformation of the flexure spring with modes 1–6 in Fig. 9. It can be seen that the flexure spring will vibrate greatly at some frequencies, so it is necessary to avoid these frequencies during operation. The significant function of flexure spring is to provide radial support for coil and piston components and ensure the positional relationship between the piston and the cylinder. Therefore, the flexure spring is expected to have a larger radial stiffness. With 50 N to the center of the spring in the radial direction, the radial deviation of the center position is solved by ANSYS when its axial displacement is 0, 10 mm, 15 mm and 18 mm respectively. In Fig. 9, the central radial displacement of the flexure spring is different. This may be caused by rotational deformation of the spring. The maximum radial displacement of the flexure spring is shown in Table 2.

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Fig. 8. Modal analysis of flexure spring.

Fig. 9. The central radial displacement of plate spring under the action of 50N radial force under different coaxial displacement.

Table 1. Value of Magnetic Flux Density Mode

1

2

3

4

5

6

Frequency (Hz)

30.283

51.091

51.509

82.503

94.461

94.674

Table 2. Maximum radial displacement of plate spring under different stroke force Displacement (mm)

0

10

15

18

Radial Deformation (mm)

0.0864

0.0867

0.0868

0.0870

3 Experimental Result 3.1 Magnetic Flux Density Testing The assembled magnetic circuit is measured by the Tesla meter. At the same axial position, the probe Measure the value of the magnetic flux density at 30 points along the circumference, and then continue to measure along the axial direction with an interval of 1mm. The magnetic flux density in the air gap is obtained through testing, as shown in Fig. 10.

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3.2 Electromagnetic Thrust Test It is very important for a high-power linear motor to generate enough thrust. The electromagnetic thrust is generated by the energized coil in the magnetic field, which can be calculated by the following formula: F = BIL

(3)

where, B, I, L are magnetic flux density, current and length of conducting wire The electromagnetic thrust is evaluated by electromagnetic thrust coefficient Ke . Ke =

F I

(4)

Fig. 10. Linear Motor Magnetic Flux Density

Fig. 11. The relationship between thrust and current of coil

The value of magnetic flux density in the air gap is shown in Table 3. In the whole air gap, the average value of magnetic flux density is 0.641 T. The error from the calculated value is 1.5%. As shown in Fig. 11, the coil of 1.32 mm wire diameter has higher thrust at the same current. The average electromagnetic thrust coefficient of the coil of 1.32 mm wire diameter is 79.85 N/A. The average electromagnetic thrust coefficient of the coil of 1.8 mm wire diameter is 37.01 N/A. Thinner wires have more thrust. Table 3. Value of Magnetic Flux Density

Magnetic Flux Density ( T )

Average value

Maximum value

Minimum value

0.641

0.755

0.465

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3.3 Stroke and Power Test The long-stroke capability of the linear motor is tested with a coil of 1.32 mm wire diameter. The number of flexure springs installed on both sides of the coil are 38. It can be seen from Fig. 12 that the stroke of the linear motor can reach 30 mm with the input power of 29 W. However, more power is requiring for continuing to increase the stroke. The reason is that the neutral section of the magnetic circuit is set too short.

Fig. 12. The relationship between power and stroke.

Fig. 13. The relationship between power and stroke of kinds of flexure springs

It could be seen from Fig. 13 that fewer flexure springs have a lower power at the same stroke. Therefore, when the number of flexure springs meet the requirements of radial support and axial stiffness, smaller amount is beneficial to reduce power. The reason is that the flexure spring itself has damping properties and consumes power. We use motor efficiency η to evaluate the efficiency of linear motors. The specific formula is as follows η=

P−I 2 R P

(5)

Figure 14 shows the relationship between motor efficiency and stroke of kinds of flexure springs. It can be seen from the Fig. 15 that different groups’ motor efficiency η are all above 0.9. The motor efficiency η is lower when the number of flexure springs is 19, the average value is 0.93. The motor efficiency η is above 0.95 on average for other groups. Figure 15 shows the distribution of the magnetic flux density of the linear motor with ±18 mm (stroke 36 mm) at the middle position. It can be seen from the figure that when the stroke exceeds 15 mm (full stroke 30 mm), the magnetic flux density will increase, and the maximum will reach to 0.46 T. When the energized coil runs in this

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region, a reverse force will be generated, so the power increases. The way to optimize this problem in the future is increase the length of the center segment so that the magneti flux density in the middle of the air gap is close to 0 T.

Fig. 14. The relationship between motor efficiency and stroke of kinds of flexure springs

Fig.15. Magnetic flux density of 18 mm on both sides of the center position

4 Conclusion This paper introduces a linear motor used on the oil-free G-M cryocooler developed by our laboratory independently. The magnetic flux density of air gap is tested by Tesla meter, and the average value is 867.04 mT. The flexure spring used in the linear motor is simulated by ANSYS. The stiffness of the flexure spring is 2.5 N/mm by calculating, which is 20.63% different from the actual measured value. The 1–6 natural frequencies of the flexure spring are obtained, which helps to avoid dangerous frequencies in the experimental test. The coil of 1.32 mm wire diameter has higher electromagnetic thrust coefficient. Under the same stroke, fewer flexure springs have less power consumption. The design stroke of this machine is 36nmm, and linear motors can easily reach 30 mm in the test, continuing to increase the stroke requires multiple times power. Acknowledgments. The project is supported by the Strategic Priority Research Program of Chinese Academy of Sciences. (Grant No. XDA18000000, No. XDA18040000).

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References 1. Corey, J.A., James, E.L., et al.: Development of a linear compressor for Use in G-M cryocoolers. In: Cryocoolers (ed.) vol. 13, pp. 201–208. Springer, US (2004) 2. Hiratsuka, Y.: A Gifford-McMahon Cryocooler below 2K with Helium 4. In: S.D. Miller, R.G. Ross, J.R. Raab (eds.) Cryocoolers 20 3. Maddocks, J.R., Kashani, A., Helvensteijn, B.P.M., et al.: Performance of G-M Cryocooler with oil-free linear compressor. In: AIP Conference Proceedings, pp. 696–674. US (2006) 4. Hiratsuka, Y., Bao, Q., Xu, M.Y.: Performance estimation of an oil-free linear compressor unit for a new compact 2K Gifford-McMahon cryocooler. In: IOP Conference Series: Materials Science and Engineering, vol. 278 (1), pp. 012050–012055 (2017) 5. Sun., J., Li, J.G., et al.: A novel oil-free dual piston compressor driven by a moving coil linear motor with capacity regulation using R134a. Sustainability 13(9) (2021) 6. Kavade, M.V., Patil, C.B.: Optimization of flexure spring using FEA for linear compressor. Int. J. Eng. Sci. 1(12), 37–45 (2012)

A 0.79 W/150 K Micro Pulse Tube Cryocooler T. S. Feng1,2 , E. C. Xing1,2 , M. Gao1,2 , Y. T. Zhang1,2 , M. L. Liang1 , H. L. Chen1,2(B) , Y. Q. Xun1(B) , and J. T. Liang1,2 1 Key Laboratory of Technology On Space Energy Conversion, Technical Institute of Physics

and Chemistry, CAS, Beijing 100190, China {hlchen,yqxun}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. A new single-stage coaxial micro pulse tube cryocooler has been developed by Technical Institute of Physics and Chemistry, CAS. With the characteristics of long-life and low-mass, it could serve for infrared detectors in space and non-space applications. Driven by a high efficiency and lightweight linear compressor, it has a small cold finger with a diameter of 8 mm, and works at 175 Hz. Inertance tube and a reservoir is used as its phase shifter. At a reject temperature of 283 K, it can lift 0.79 W at 150 K with 10 W electrical input. This paper describes the key parameters of components, and presents test data in several working conditions of this pulse tube cryocooler in detail. Keywords: Pulse tube cryocooler · Linear compressor · Coaxial · High frequency

1 Introduction Micro pulse tube cryocoolers focus on providing a required cooling capacity with a tiny size and mass for cold optical equipment operating on the space and ground applications. Their main characters are small size, light weight, fast cooling speed, long lifetime and low vibration. There has been some progress in the area of micro pulse tube cryocooler development. In 2006, Radebaugh [1] calculated the optimized regenerators working at 60 Hz, 400 Hz and 1000 Hz, the solution shows that the COP of micro pulse tube cryocooler working at 400 Hz and 1000 Hz is 78% and 68% of it working at 60 Hz respectively, which reflects that a pulse tube cryocooler can become lighter and smaller without a significant loss of performance. In 2007, M. Petach reported a 100 Hz pulse tube cryocooler [2] they developed with a weight of 782 g, it can obtain 1.1 W@77 K with 45 W PV power consumption, subsequent in 2009, they developed a 144 Hz coaxial pulse tube cryocooler [3], it can cool down to 80 K within about 3.5 min with 325 J heat capacity load. In the same year, S. Vanapalli reported a 120 Hz linear pulse tube cryocooler [4] which takes 5.5 min to cool down from 298 K to 80 K. In 2014, T.C. Nast [5] reported their pulse tube cryocooler weighs 328 g, it can provides 0.85 W @ 150 K with 10 W input power at -15°C reject temperature. In 2019, J.Q. Li developed a 101 Hz pulse tube © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 780–786, 2023. https://doi.org/10.1007/978-981-99-6128-3_101

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cryocooler [6] weighs about 700 g, could obtain 1.1 W@77 K, and takes 6 min cooldown to 77 K. In 2020, T. Feng developed a 120 Hz pulse tube cryocooler weighs 598 g [7],and can obtain 1.2 W@80 K with 45 W electric power input, and it takes about 3 min to cooldown to 80 K with 250 J heat capacity load, in the same year, I.M. McKinley [8] reported their pulse tube cryocooler weighs 477 g, and can provide 0.75 W@80 K with a reject temperature of 220 K at 135 Hz. As increasing the drive frequency of pulse tube cryocoolers to make the micro pulse tube cryocooler smaller and lighter become a common opinion in the development of micro pulse tube cryocooler, researchers are caring about how a small mass and volume a high efficiency cryocooler can reach, and what the drive frequency of it is needed. Our work is to develop such a cryocooler. And in this paper, a novel micro pulse tube cryocooler recently developed by us are introduced, it has a tiny cold finger, and the frequency of 175 Hz, which is rare in pulse tube cryocoolers driven by linear compressor. And the performance testing data of it are present in detail.

2 Exterior and Main Parameters A schematic of this micro pulse tube cryocooler is shown at Fig. 1, it has a spilt configuration, connecting the linear compressor with the cold finger via a connecting tube. A back-to-back moving coil linear compressor is used to drive the cooling element, whose outline is a cylinder with a diameter of 40 mm, a length of 62 mm, and with a weight of 320 g.

Fig. 1. Schematic of this micro pulse tube cryocooler

For the cold finger, it has a coaxial structure with a diameter of 8 mm and a length of 40 mm., the combination of inertance tubes and reservoir is used as its phase shifter of this micro pulse tube cryocooler. The total length of the inertance tube is 1.8 m, including the first tube with a diameter of 1.4 mm and length of 0.3 m, the second tube with a diameter of 2 mm and length of 0.5 m, and the third tube with a diameter of 3 mm and length of 1 m. the volume of the reservoir is 30 cc.

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The filling in the regenerator is mixed matrix including #500 and #635 stainless steel woven net. The working medium is pure helium, and the charge pressure is 4.5 MPa. The main parameter s of this micro pulse tube cryocooler is listed at Table 1. Table 1. Main parameters of this micro pulse tube cryocooler Component

Parameters

Cold Finger

8 mm in diameter and 40 mm in length

Compressor

40 mm in diameter and 62 mm in length

Matrix

#500 and #635 stainless screen

Inertance Tube

A total length of 1.8 m

Reservoir

30 cc in volume

Charge Pressure

4.5 MPa

3 Performance Testing Devices 3.1 Testing Devices The testing devices are shown at Fig. 2. Water-cooling equipment is used to hold the reject temperature at 283 K, and the compressor are working at environmental temperature. The temperature sensor PT100 and the heater bands are installed at the cold tip. The pressure sensor is inserted in the connecting tube to monitor the pressure of the compressor outlet, and windows are set at the end of compressor to use the laser displacement sensor to monitor the displacement of pistons. 3.2 Drive Frequency Figure 3a describes the motor efficiency of the compressor, which is defined as the quantity of the Joule heating (i2 R.) losses in the coils subtracted from the compressor input power divided by the compressor input power [8]. As the cold tip temperature and the reject temperature is hold around 100 K and 283 K respectively, the efficiency is tested at an input power of 10 W and 20 W. The efficiency is decreased when the input power increase from 10 W to 20 W ranging from 155 Hz to 190 Hz, and the optimal frequency of compressor remains at 185 Hz when the input power changed. The highest motor efficiency is 69.2%, and appears at the frequency of 186 Hz. However, the motor efficiency is weakly changed as the frequency change, the efficiency at 175 Hz is only less than it at 185 Hz for about 2%. That means the frequency change did little effect on the efficiency of this compressor, extremely when the input power is low. Figure 3(b) describes the cold tip temperature without cooling load when the input power is 30 W. As this temperature reflects the performance generally, 175 Hz is chosen as the optimal frequency of this micro pulse tube cryocooler.

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Fig. 2. Using water-cooling to hold the reject temperature, and the sensor to monitor the pressure of the compressor outlet and displacement of pistons 0.80

0.74

67.6 67.4

0.72 Reject Temperature: 283 K Charging Pressure: 4.5 MPa 0.70

Temperature (K)

67.2

0.68 0.66 0.64 0.62 0.60

67.0 66.8 66.6 66.4 66.2 66.0

0.58

65.8

0.56

65.6

0.54

65.4

0.52

65.2

0.50

150

155

160

Reject temperature: 283 K Input power: 30 W Charging pressure: 4.5 MPa

67.8

20 W 10 W

0.76

Motor efficiency

68.0

Input power

0.78

165

170

175

Frequency (Hz)

(a)

180

185

190

195

65.0

168

170

172

174

176

178

180

182

Frequency (Hz)

(b)

Fig. 3. (a) The motor has a high efficiency at high frequency, (b) the cold tip reaches a lowest temperature at 175 Hz

3.3 Cooling Performance The cooling performance of this micro pulse tube cryocooler is shown as Fig. 4. Among these two pictures, Fig. 4(a) is the relationship between cold tip temperature and cooling load, drawn depend on the data obtained at tests at the inputs ranging from 10 W to 45 W, and Fig. 4(b) is the relationship between cooling load and compressor input power, drawn with the data estimated depend on the results of tests at the cold tip temperature ranging from 80 K to 150 K. The lowest of this micro pulse tube cryocooler is about 60 K, it can provide near 0.8 W cooling power at 80 K at 45 W compressor input power, and can provide 0.79 W cooling power at 150 K at 10 W compressor input power. The performance and mass of this micro pulse tube cryocooler determined that it suits to work at 80 K to 150 K.

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1.0 0.9

40

Input electric power (W)

Cooling load (W)

0.8 0.7 0.6 0.5 Reject temperature: 283 K Operating frequency: 175 Hz Charging pressure: 4.5 MPa Input electric power

0.4 0.3

10 W 30 W 45 W

0.2 0.1 0.0

60

80

100

120

140

160

35 30 25

Cold tip temperature

20 15 10

180

Reject temperature: 283 K Operating frequency: 175 Hz Charging pressure: 4.5 MPa 0.0

0.5

1.0

1.5

Temperature (K)

2.0

2.5

3.0

80 K 90 K 100 K 120 K 150 K 3.5

4.0

Cooling load (W)

(a)

(b)

Fig. 4. (a) The lowest temperature of this cryocooler is about 60 K, (b) it can obtain 0.79 W at 150 K with 10 W compressor power input 300

None added thermal mass 250 J added thermal mass

280 260

Reject Temperature: 283 K Operating Frequency: 175 Hz Charging Pressure: 4.5 MPa

Temperature (K)

240 220 200 180 160 140 120 100 80 60 40

0

1

2

3

4

5

6

7

8

9

10

Time (min)

Fig. 5. It takes 2 min cooldown to 150 K with 250 J thermal mass added, 4 min to 80 K

3.4 Cooldown Speed Using a specific quality copper cap to simulate the thermal mass added on the cold tip, the cooldown speed of adding 250 J thermal mass and none thermal mass added are tested respectively, the compressor input power during this test process is less than 50 W. The cooldown curve of the micro pulse tube cryocooler is drawn as Fig. 5. It takes 2 min to cooldown to 150 K with 250 J thermal mass added, 3 min to 100 K, 4 min to 80 K and 6 min to the lowest temperature. It takes about half as long as the test 250 J thermal mass added for the test none thermal mass added to cooldown to a specific temperature. 3.5 Pressure at the Compressor Outlet and Pistons Displacement The Fig. 6 described the amplitude of the dynamic pressure and the piston displacement of compressor respectively, as the two pistons of the compressor have almost the same

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moving character at the test, only the data of one has adopted to draw Fig. 6(b). Both of the amplitude of the dynamic pressure and the piston displacement are increased as the input power increased, the amplitude of the dynamic pressure is increased as the cooling load, or put it another way, the cold tip temperature, increased. 1.5

Amplitude of piston displacement (mm)

Amplitude of dynamical pressure (bar)

4.5

4.0

3.5

3.0

Cooling load 0W Reject Temperature: 283 K 0.5 W Operating Frequency: 175 Hz 1W Charging Pressure: 4.5 MPa

2.5

2.0

5

10

15

20

25

30

35

Input electric power (W)

(a)

40

45

50

1.4 1.3 1.2 1.1 1.0 0.9

Cooling load

0.8 0.7 0.6

0W 0.5 W 1W

Reject Temperature: 283 K Operating Frequency: 175 Hz Charging Pressure: 4.5 MPa 5

10

15

20

25

30

35

40

45

50

Input electric power (W)

(b)

Fig. 6. (a) The pressure ratio of warm end is about 1.1 to 1.2, (b) The amplitude of piston displacement is about 0.7 mm to 1.4 mm

As the pressure senser near to the warm end, it can be estimated that the pressure ratio of the warm end is about 1.1 to 1.2 at the input power ranging from 10 W to 45 W, and due to the high drive frequency of 175 Hz the amplitude of piston displacement of compressor is about 0.7 mm to 1.4 mm, which brings the possibility to lighter the compressor.

4 Conclusions A 175 Hz micro pulse tube cryocooler is developed to meet the requirement of infrared detectors in space and non-space applications at the temperature ranging from 80 K to 150 K. It can provide 0.79 W cooling power with an input power of 10 W. Taking about 2 min, it can cooldown to 150 K with 250 J thermal mass added, and 4 min to 80 K. Now the weight of this cryocooler is less than 600 g, not including the dewar, sensor or cooler drive electronics, and to lighter the weight of cryocooler and enhance its performance will be our direction of effort. Acknowledgments. The project is supported by National Natural Science Foundation (Number 52106036), and National Key R&D Program of China under grant 2018YFB0504600, 2018YFB0504603.

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References 1. Radebaugh, R., O’Gallagher, A.: Regenerator operation at very high frequencies for microcryocoolers. In: Proceedings of the Advances in Cryogenic Engineering, pp. 1919–1928. Melville (2006) 2. Petach, M., Waterman, M., Tward, E.: Pulse tube microcooler for space applications. In: Proceedings of the ICC14, pp. 89–93 (2007) 3. Petach, M., Waterman, M., Pruitt, G., Tward, E.: High frequency coaxial pulse tube microcooler. In: Proceedings of the ICC15, pp. 97–103 (2009) 4. Vanapalli, S., Lewis, M., Gan, Z., Radebaugh, R.: 120 Hz pulse tube cryocooler for fast cooldown to 50K. Appl. Phys. Let. 90(7) (2007). https://doi.org/10.1063/1.2643073 5. Nast, T.C., Roth, E., Olson, J.R., Champagne, P., Frank, D.: Qualification of lockheed martin micro pulse tube Cryocooler to TRL6. In: Proceedings of the ICC18, pp. 51–57 (2014) 6. Li, J.Q., et al.: Modelling and experimental study of a 700 g micro coaxial Stirling-type pulse tube cryocooler operating at 100–190 Hz. In: 27th International Cryogenics Engineering Conference and International Cryogenic Materials Conference 2018, Iop Publishing Ltd, (2019) 7. Feng, T., et al.: A 598g Micro Pulse Tube Cryocooler’. In: Proceedings of the ICC21, pp. 157– 161 (2020) 8. McKinley, I.M., Mok, M.A., Rodriguez, J.I.: Characterization testing of space-flight lockheed martin Micro 1-2 cryocooler for the Mapping Imaging Spectrometer for Europa (MISE). In: Proceedings of the ICC21, pp. 135–142 (2020) 9. McKinley, I.M., Hummel, C.D., Johnson, D.L., Rodriguez, J.I.: Characterization testing of Lockheed Martin high-power micro pulse tube cryocooler. In: Proceedings of the CEC2017

Experimental Investigation of Controlling Piston Offset in Oil-Free Linear Compressors for J-T Throttling Refrigerator Y. L. Liu1,2 , Y. Q. Xun1(B) , H. L. Chen1(B) , Z. J. Huang1,2 , C. Zhang1,2 , and J. H. Cai1 1 Key Laboratory of Technology On Space Energy Conversion, Technical Institute of Physics

and Chemistry, Chines Academy of Sciences, Beijing 100190, China {yqxun,hlchen}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. The linear compressor provides the indispensable pressure ratio and mass flow for the Joule-Thomson (J-T) throttling refrigerator. The piston of the single-piston linear compressor will be offset from its initial position at a highpressure ratio, which may result in reduced compressor performance. Three methods of controlling the offset are compared in the paper, which is adopting the symmetrical dual-piston structure, increasing the initial clearance volume, and applying direct current (DC) voltage, respectively. The linear compressor was tested with nitrogen, and the performance of the linear compressor with different adjustment methods was evaluated. The experimental results reveal that the symmetrical dual-piston structure can considerably improve the piston offset and ensure the remarkable performance of the compressor. The method of applying DC voltage can also achieve good performance, but it increases the copper loss of the motor and the complexity of the operation. The research furnishes an effective quantitative analysis for mitigating piston offset and supplies the direction for optimizing compressor performance. Keywords: piston offset · linear compressor · symmetrical dual-piston · J-T throttling refrigerator

1 Introduction The pre-cooled Joule-Thomson (J-T) throttling refrigerator is widely concerned in space cryogenic mechanical refrigeration due to its simple structure, high efficiency, and stable operation. Linear compressors that provide pressure ratio and mass flow for J-T throttling refrigerators have also been studied. Compared with traditional compressors, the linear compressor without a crank connecting rod adopts the flexure spring support and gap sealing technology based on the reed valve. Therefore, the lateral force on the piston is reduced, and the mechanical efficiency is improved. However, the application of gap sealing technology causes gas to leak from the gap between the cylinder and piston. There is a pressure difference between the cylinder and the back pressure chamber when © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 787–794, 2023. https://doi.org/10.1007/978-981-99-6128-3_102

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the average leakage between them is not zero in a cycle [1]. The motion center of the piston is offset from the initial motion position, which is a distinguishing feature of the linear compressor [2]. The offset of the piston motion center will reduce the volumetric efficiency of the compressor at the same stroke. The piston offset will change with the operating conditions, which brings great challenges to controlling the compressor. Therefore, controlling the piston offset is essential to improve compressor performance and control. Some scholars have studied the piston offset. Li et al. [2] analyzed the influence of different operating conditions on the piston offset through simulation and experiment. Fu et al. [3] studied the piston offset of the Stirling compressor with four conditions through experiments and clarified the influencing factors of offset. Ding et al. [4] studied the factors affecting the piston offset of the valved linear compressor. The offset amplitude and start-up characteristics are studied under different conditions. Chen et al. [5] introduced the DC offset control method to solve the moving offset problem based on the multi-loop moving coil linear compressor. Liang et al. [6] reported piston offset in a linear compressor and summed up the methods of controlling the piston offset, including increasing the stiffness ratio appropriately, superimposing a DC voltage, adding ‘stops’ to limit the piston motion, et al. However, there is still insufficient research on the practical effects of using the above methods. It is indispensable to study further and quantify the impact of these methods on compressor performance. The paper compares the effect of three methods on piston offset by adopting the symmetrical dual-piston structure, increasing the initial clearance volume, and applying DC voltage, respectively. The influence of the methods on compressor performance is also analyzed.

2 The Structure and Principle of the Linear Compressor 2.1 Single-Piston Linear Compressor Prototype The experiment employs the single-piston oil-free linear compressor prototype, as shown in Fig. 1. The linear motor adopts a moving coil structure, and it has a higher mechanical efficiency than a rotary motor. The flexure springs are utilized on both sides of the motor to support the mover, which could avoid the dry friction between the piston and cylinder. There is a gap seal between the piston and cylinder, achieving oil-free lubrication. The valve system is a split valve, and both the suction and exhaust valve are reed valves. 2.2 Dual-Piston Linear Compressor Prototype The symmetrical dual-piston linear compressor adopts the same motor to drive two pistons simultaneously, as shown in Fig. 2. The main components are similar to those of the single-piston linear compressor. The working process of the dual-piston linear compressor is slightly more complicated. The detailed process is demonstrated in Table 1, and the a-b-c-d-e-f-a process is carried out in sequence, which is repeated. The electromagnetic force is symmetrically distributed in the single-piston compressor. However, the gas load force in the cylinder is asymmetrically distributed. The gas load force does not match electromagnetic force, which causes the piston to deviate from

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(a) The structure of the linear compressor

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(b) The linear compressor prototype

Fig. 1. Single-piston linear compressor prototype.

(a) The structure of the linear compressor

(b) The linear compressor prototype

Fig. 2. Dual-piston linear compressor prototype.

Table 1. The working process diagram of the dual-piston linear compressor. Process a

Piston movement Compression Compression direction chamber A chamber B A

B

Compression

Expansion

b

A

B

Compression

Suction

c

A

B

Exhaust

Suction

B

Expansion

Compression

d

A

e

A

B

Suction

Compression

F

A

B

Suction

Exhaust

Valve status All closing The suction valve of B opening, others closing The exhaust valve of A and suction valve of B opening, others closing All closing The suction valve of A opening, others closing The suction valve of A and exhaust valve of B opening, others closing

the initial motion position. The dual-piston structure can theoretically realize that the gas load forces in two cylinders are equal in magnitude and opposite in direction. In this way, the gas load force is symmetrical, and the inconsistency is improved.

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3 Experimental Apparatus 3.1 Test Rig A visual real-time test rig of linear compressors is shown in Fig. 3. This system consists of two heat exchangers, linear compressor prototypes, a throttling valve, and other additional components. The compressor is driven by the power supply that provides both AC and DC voltage. The electrical parameters are displayed by the power meter or the three-phase PWM special tester. The laser displacement transducer is adopted on the end cover of the compressor, measuring the piston stroke and offset in real time. The dynamic signals are transmitted to the oscilloscope for display and storage. The pressure, temperature, and mass flow are collected and displayed on the computer.

Fig. 3. A visual real-time test rig of the linear compressor.

3.2 Test Conditions Nitrogen is used to measure the performance of prototypes, and the test conditions are shown in Table 2. The pressure ratio defaults to 2.0 with the variable stroke when the driving frequency is 60.0 Hz. And the piston stroke defaults to 9.0 mm with the variable pressure ratio. Significantly, the throttle valve has been closed to the minimum when the pressure ratio is about 1.7 at the variable pressure ratio. Therefore, the pressure ratio of 1.7, 2.0, 2.5, and 3.0 are selected for analysis. The dimensions of cylinder diameters and the initial clearance lengths (CL) about the single-piston and dual-piston prototypes is to ensure that the exhaust volume of the two prototypes is as consistent as possible.

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Table 2. The test conditions of linear compressors. Parameters (unit)

Value

Fill pressure (MPa)

0.6

Parameters (unit)

Value

Piston stroke (mm)

6.0, 7.0, 8.0, 9.0, 10.0

Ambient temperature (°C) 23.0

Pressure ratio

1.7, 2.0, 2.5, 3.0

Operating frequency (Hz)

60.0

Initial clearance length (mm) 3.73, 5.73, 5.30

Cylinder diameter (mm)

21.0, 26.0 Suction pressure (MPa)

0.25, 0.30, 0.35, 0.40

4 Experimental Results and Discussion 4.1 Piston Offset of the Single-Piston Linear Compressor The offset of the single-piston prototype is about 1.7 mm when the frequency is 60.0 Hz, the pressure ratio is 2.0, and the piston stroke is 9.0 mm. The DC voltage is applied to the prototype with 1.7 mm as the motion center to judge the piston position at the above working point, which means that the piston is offset in the direction of the cylinder or the flexure spring. Control the offset to ±1 mm and ±2 mm, respectively, and the results of compressor performance are shown in Fig. 4. Figure 4(a) manifests the electrical parameters of the compressor, and the variation of compressor efficiency and discharge pressure is demonstrated in Fig. 4(b). The DC voltage decreases as the piston offset increases, from positive voltage to negative voltage input. The experimental voltage and current decrease with the increase of the piston offset, which can be measured with the three-phase PWM special tester. The input power first increases and then decreases with increasing piston offset. The motor efficiency slightly increases with the variable piston offset, and measured power factor increases first then decreases. It can be seen that the volumetric efficiency drops fairly linearly, and the isentropic efficiency first declines and then ascends. The motor, volumetric and isentropic efficiencies of the compressor are defined as Eqs. 1–3. ηmotor = 1 − ηv = ηisen =

˙ isen W ˙ in W

m ˙0 m ˙ theory

=

2 R Irms Pin

4m ˙ 0 Rg T1

π D2 SFP1   P2 k−1 k m ˙ 0 Rg T1 ( ) k − 1 /Pin = k −1 P1

(1) (2) (3)

where, I rms is the effective value of current, R is the resistance, Pin is the input power, m ˙ 0 is the system mass flow, Rg is the gas constant, and k is the adiabatic index. It is worth noting that both the exhaust pressure and the volumetric efficiency decrease with the growing offset. This may be caused by the piston motion center getting farther and farther away from the cylinder. The farther the piston is from the cylinder, the lower the average pressure in the cylinder and the lower the exhaust pressure will be. The effective scavenging volume is lower, which reduces mass flow, and the volumetric

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Fig. 4. Performance of the single-piston linear compressor with different piston offsets.

efficiency drops eventually. Therefore, the piston motion center is biased towards the flexure spring when the offset is 1.7 mm, and the compressor performance is unfavorable. Besides, the results demonstrate that the compressor performance can be effectively improved when the piston offset is zero. Therefore, the following discussion of adding DC voltage to control the offset will be based on this result. 4.2 Piston Offset and Performance of Compressor with Three Methods Figure 5 illustrates the piston offset and mass flow variation with three control methods at different piston strokes and pressure ratios. Adopting the symmetrical dual-piston structure and the DC voltage can significantly improve the piston offset, and the piston offset tends to zero. However, increasing the initial CL only slightly improves piston drift. The factor is that the dual-piston structure can theoretically realize the gas load force symmetry, and the DC voltage can add a push or pull force to move the piston motion center back to its original position. Increasing the initial clearance volume provides a limited improvement in gas fluctuations in the cylinder, thereby relieving piston deflection. The piston offset decreases with the rising piston stroke under the dual-piston structure method and increases the initial CL method (the first two methods). The larger the stroke, the larger the scavenging and the influence of pressure fluctuations is reduced. The mass flow of the system climbs linearly with the growing stroke. And the mass flow of the dual-piston structure is greater than that of other methods. The piston offset rises with the increasing pressure ratio under the first two methods. The greater the gas force with a greater pressure ratio, the increased asymmetry of the system, and the further the piston is pushed away from its initial position. The variation of motor efficiency and power factor with three control methods at different piston strokes and pressure ratios is displayed in Fig. 6. The motor efficiency is basically not affected by piston stroke, pressure ratio, and control method. The power factor is obviously impacted by different methods, among which the dual-piston structure has the largest power factor. The trend of the initial CL to 5.73 mm and 3.73 mm is relatively consistent, which first ascends and then descends with increasing stroke and drops with the growing pressure ratio. The minimum power factor is obtained by applying the DC voltage, which increases with the growth of the stroke and reduces with the rising pressure ratio. This is because the first two methods do not convert the output

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Fig. 5. Piston offset and mass flow with three control methods.

Fig. 6. Motor efficiency and power factor with three control methods.

Fig. 7. Volumetric and isentropic efficiencies with three control methods.

characteristics of the motor, but applying the DC voltage increases the copper loss of the motor and varies the output characteristics. The volumetric and isentropic efficiency with three control methods is described in Fig. 7. The maximum volumetric and isentropic efficiency is acquired by adopting the

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DC voltage at almost any piston stroke and pressure ratio. The efficiency of the dualpiston structure is second only to DC voltage, and the volumetric efficiency is improved by increasing the CL, but the isentropic efficiency is minimal.

5 Conclusion Three methods of controlling piston offset are compared in the paper. Although applying DC voltage can achieve good performance, it needs to test the direction of piston offset in advance, which increases the complexity. What is more, the extra copper loss is added to the motor, which impacts the power factor. The symmetrical dual-piston structure can significantly improve the piston offset and ensure the excellent performance of the linear compressor. Increasing the clearance volume only slightly reduces the piston drift but hurts the overall efficiency of the compressor. The research provides an effective quantitative analysis for ameliorating the piston offset and supplies the direction for optimizing the compressor performance. Acknowledgments. The project is supported by Beijing Municipal Natural Science Foundation (Grant No. 3202033).

References 1. Spoor, P.S., Corey, J.A.: A Novel Method for Controlling Piston Drift in Devices with Clearance Seals. Plenum Press, New York (2004) 2. Li, C.Z., Li, J.G., et al.: Investigation of frequency and piston offset characteristics in an oil-free linear compressor. J. Vibrat. Shock 40(03), 139–146 (2021) 3. Fu, M., Jiang, Z.H., et al.: Analysis on influencing factors of piston drift phenomena in Stirling compressor. Chin. J. Refrig. Technol. 35(03), 24–27 (2015) 4. Ding, L., Zhang, H., et al.: Study on the establishing-process of piston offset in the helium valved linear compressor under different operating parameters. Int. J. Refrig. 133, 80–89 (2022) 5. Chen, H.Y., Liu, X.Y., et al.: Study of the moving offset characteristics of multi-loop moving coil linear compressor. Int. J. Refrig. 133, 289–301 (2022) 6. Liang, K.: A review of linear compressors for refrigeration. Int. J. Refrig. 84, 253–273 (2017)

Performance Analysis of the Flexure Spring for the Linear Compressor at Liquid Nitrogen Temperature Chen Zhang1,2 , Qingjun Tang1(B) , Miguang Zhao1(B) , Yuanli Liu1,2 , Zhijie Huang1,2 , Enchun Xing1,2 , Tianshi Feng1,2 , and Jinghui Cai1 1 Key Laboratory of Space Energy Conversion Technology,

Technical Institute of Physics and Chemistry, CAS, Beijing 100190, China {angqingjun10,mgzhao}@mail.ipc.ac.cn 2 University of Chinese Academy of Science, Beijing 100190, China

Abstract. For the cryogenic requirements of planetary exploration missions in the future, the operating temperature of the space linear compress utilized in space cryocoolers may be lower than 150 K. As an essential component of a linear compressor, the flexure spring has a significant impact on the reliability of the linear compressor. Considering that the performance of flexure springs is not clear at cryogenic temperature, this paper illustrates that the effect of liquid nitrogen temperature on the axial stiffness and equivalent elastic strain of the flexure spring. The axial stiffness and equivalent elastic strain of flexure springs made of 304 stainless steel and SANDVIK 7C27Mo2 are analyzed at ambient and liquid nitrogen temperature. The properties after rewarming have also been determined. The results demonstrate that the axial stiffness of flexure springs increases slightly, and the elastic stress tends to increase at liquid nitrogen temperature. Compared with the characteristics of different temperatures, the performance of SANDVIK 7C27Mo2 is more stable. This research provides significant experimental basis for cryogenic temperature application of flexure springs. Keywords: flexure spring · linear compressor · axial stiffness · strain · liquid nitrogen temperature

1 Introduction Linear compressors have critical applications in special fields such as aerospace which require long life, high reliability and high efficiency. In the future, the operating temperature of the space linear compressor utilized in space cryocoolers needs to be lower than 150 K for planetary exploration missions in a cryogenic space environment [1]. Therefore, it’s necessary to study the linear compressor operating at cryogenic temperature. As a vital component of the linear compressor, the flexure spring is used to maintain the sealing gap between the piston and the cylinder, and provide sufficient stiffness and strength for the reciprocating motion of the piston. The axial stiffness is to support the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 795–801, 2023. https://doi.org/10.1007/978-981-99-6128-3_103

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piston that performs a linear motion. The fatigue strength is to ensure that the stress on flexure springs is less than the fatigue limit of materials [2]. The axial stiffness can be simulated by ANSYS and be measured by the displacement detector and the dynamometer[3]. For estimating the fatigue strength, the stress distribution of flexure springs is usually simulated by ANSYS to search stress concentration points [4]. In the actual stress measurement, the resistance strain gauge method and the stress optical method are mostly used [5]. However, there are few reports on the measurement and analysis of the stiffness, strain, and stress of flexure springs at the liquid nitrogen (LN2 ) temperature. In this paper, the equivalent elastic strain of the flexure spring at LN2 temperature is measured by cryogenic strain gauges, while the reliability of the elastic strain measurement is verified with a cantilever beam with known analytical solutions. Through contrastive analysis of the performance of flexure springs at LN2 and ambient temperature even after rewarming, the characteristics of flexure springs made of different materials at cryogenic temperature and after the treatment of liquid nitrogen are investigated. The primary objective of this paper is to investigate effect of cryogenic temperature on axial stiffness and equivalent elastic stress to further broaden the application range of flexure springs.

2 Finite Element Analysis 2.1 Axial Stiffness Simulation Considering high tensile strength, yield strength and fatigue strength, 304 stainless steel and SANDVIK 7C27Mo2 are chosen as the materials for experimental flexure springs. 304 stainless steel has good processing performance and high toughness, while SANDVIK 7C27Mo2 is a martensitic stainless steel alloyed with molybdenum which has good mechanical properties. The parameters of SANDVIK 7C27Mo2 and 304 stainless steel at room temperature are shown in Table 1. Table 1. Property parameters of 7C27Mo2 and 304 stainless steel at room temperature. Material

Density (g/cm3 )

Young’s Modulus (MPa)

Tensile Strength (MPa)

Yield Strength (MPa)

SANDVIK 7C27Mo2

7.70

210000

1800

1300

304 stainless steel

7.75

193000

550

200

Table 2 is indicative of the simulated and experimental axial-stiffness values of the flexure spring with two different materials under 10N. It can be seen that the increase of the number of meshes has little effect on the simulation stiffness. And the error ratio between simulated and measured values is less than 10%. Therefore, the reliability of the simulation is proved.

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Table 2. Axial stiffness of flexure springs with different materials. Material

Simulated elements

Simulated stiffness (N/mm)

Experimental stiffness (N/mm)

Error value

SANDVIK 7C27Mo2

11403

4.874

4.600

5.96%

45419

4.776

304 stainless steel

3.83%

80170

4.734

2.91%

159355

4.716

2.52%

11403

4.494

45419

4.392

2.40%

80170

4.348

1.38%

159355

4.339

1.17%

4.289

4.78%

2.2 Equivalent Elastic Strain Simulation. The flexure spring is based on the Archimedes profile, and the deformation of the three spring arms is almost the same when the center is stressed (Fig. 1).

Fig. 1. The equivalent elastic strain result.

Fig. 2. Strain gauge sticking points.

Therefore, the interior, middle and external stress concentration points of a single spring arm are selected for strain measurement. Figure 2 demonstrates the strain gauges pasted on the front and back of the three points.

3 Performance Testing 3.1 Experimental Apparatus The experimental apparatus is mainly composed of the stiffness measurement system, the strain gauge measurement system, the liquid nitrogen insulation device, and the flexure spring holder, as shown in Fig. 3. The system device information is in Table 3. In addition, considering the effect of temperature change on resistance, the strain gauge which pasted on the quartz material with a low expansion rate can be used as the temperature compensation sheet of the system.

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Type

Sensitivity

Dynamometer

Mark-10 BG100

0.2%

Strain gauges

KYOWA KFL-5-120-C1

2%

Collector

KEITHLEY 2700

0.008%

Table 4. Comparison of axial stiffness of flexure springs at different temperatures at 5 mm. Material

Ambient Stiffness (N/mm)

LN2 Stiffness (N/mm)

Rewarming Stiffness (N/mm)

SANDVIK 7C27Mo2

4.600

4.609

4.609

304 stainless steel

4.289

4.520

4.529

Fig. 3. Physical map of the experimental system.

3.2 Cantilever Beam Verification Text For verifying the feasibility of the strain gauge measurement at LN2 temperature and obtaining the relationship between the resistance change of the strain gauge and the strain, it’s indispensable to measure and analyze the stainless steel cantilever beam with the known strain analytical solution before the experiment. By testing the cantilever beam, the linear relationship between the resistance change of the strain gauge and the strain is obtained, which can be expressed as: εA = 4.4434Ω + 0.000003

(1)

where εA is the strain, and ΔΩ is the corresponding resistance change of strain gauge. Figure 4 manifests that the cryogenic strain of the cantilever beam is basically consistent with the ambient strain under the same resistance change. Consequently, the results of strain gauge measurements at LN2 temperatures are reliable.

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Fig. 4. Strain measurement data fitting curve at ambient temperature and LN2 temperature.

4 Results and Discussion 4.1 Axial Stiffness at Different Temperatures This part mainly analyzes the change of the axial stiffness of flexure springs of the same material at different temperatures with the increasing displacement. And the axial stiffness of flexure springs of different materials at different temperatures under the same displacement of 5 mm which can ensures that the spring does not break.

Fig. 5. Axial stiffness at different Fig. 6. Axial stiffness at different temperatures and displacements (304 stainless temperatures and displacement (SANDVIK). steel).

From Fig. 5, it is found that the liquid nitrogen processing can increase the axial stiffness of 304 stainless steel. Figure 6 describes that the axial stiffness of SANDVIK at LN2 temperature is not much different from that at room and rewarming conditions, thus it has good stiffness retention and is suitable for cryogenic environments. Table 4 shows that when the displacement of the central position is 5 mm, the axial stiffness of SANDVIK 7C27Mo2 is greater than that of 304 stainless steel at different temperatures.

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4.2 Equivalent Elastic Strain at Different Temperatures When the flexure spring is deformed, the spring arm is simultaneously subjected to stress such as stretching, bending, and torsion. Considering that the size of the strain gauge is extremely small, it mainly receives the strain in the direction of the sense grid length. Therefore, the equivalent strain received by the sheet can be simplified as a straight line treatment without excess strain due to tension and torsion. Considering the error caused by the different pasting positions of the measurement points, quantitative analysis of different materials at the same point is not discussed. Actually, whether the deformation at different temperatures is consistent can be analyzed by equivalent elastic strain. And the elastic stress can also be estimated from the equivalent elastic strain.

Fig. 7. Equivalent elastic strain at different temperature (304 stainless steel).

Fig. 8. Equivalent elastic strain at different temperature (SANDVIK).

In Fig. 7 and 8, under the same temperature conditions, the equivalent elastic strain of flexure springs made of different materials shows an increasing trend with the growing displacement. It is convinced that the equivalent strain of the flexure spring whether 304 stainless steel or SANDVIK at the same point but different temperatures is approximately equal. Therefore, the deformation mechanism at LN2 and room temperature is the same. In contrast, the equivalent elastic strain of different materials at LN2 temperature increases to a certain extent, and the repeatability of the strain experiment is good. However, the linearity becomes worse, which may be caused by the inhomogeneity of temperature in the LN2 environment. On the whole, it can be inferred that SANDVIK 7C27Mo2 has less stress increment and is more suitable for cryogenic operation. Assuming that the Poisson’s ratio of the material is approximately constant, the stiffness and elastic modulus are approximately proportional. The elastic modulus of 304 stainless steel at LN2 temperature is estimated at 203.3 GPa. Compared with its cryogenic measurement value of 201.2 GPa [6], it can be known that the estimated value which is not much different from the literature is credible. The elastic modulus of SANDVIK at LN2 temperature is estimated at 210.4 GPa. Besides, the elastic modulus of 304 stainless steel after rewarming is estimated at 203.8 GPa, and the elastic modulus of SANDVIK after rewarming is also estimated at 210.4 GPa. From the increasing elastic modulus, it can be inferred that the equivalent elastic stress of the flexure spring made of 304 stainless steel has a small increase from room

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temperature to LN2 temperature, even rewarming. In contrast, the equivalent elastic stress of SANDVIK is almost unchanged, indicating that SANDVIK can still maintain characteristics and remain stable at the cryogenic environment and after liquid nitrogen treatment.

5 Conclusion This paper provides an experimental basis for cryogenic temperature application of flexure springs. The axial stiffness and equivalent elastic strain of flexure springs made of 304 stainless steel and SANDVIK 7C27Mo2 at LN2 temperature, ambient temperature and after rewarming are studied. It is found that the deformation mechanism at LN2 and ambient temperature is the same, while the LN2 temperature resulted in a slight increase in axial stiffness and a tendency for equivalent elastic stress to increase. The properties after rewarming have also been determined. The results demonstrate that the axial stiffness of flexure springs increases slightly, the elastic strain keeps unchanged, and the elastic stress also tends to increase at liquid nitrogen temperature. Compared with the characteristics of different temperature, the performance of SANDVIK 7C27Mo2 is more stable and suitable for cryogenic operation of flexure springs. Acknowledgments. The project is supported by The National Natural Science Foundation of China (NO. 52106036) and 2009ZYHG0003.

References 1. Nast, T.C., Helvensteijn, B.P.M., Roth, E., Olson, J.R., Champagne, P., Maddocks, J.R.: Cryocooler with cold compressor for deep space applications. In: Ross, J.R.G. (ed.) Cryocooler 18, pp. 39–44. ICC Press, Miller, S.D. (2014) 2. Lee, C.C., Pan, R.B.: Flexure bearing analysis procedures and design charts. In: Ross, R.G. (ed.) Cryocoolers 9, pp. 413–420. Springer, Boston (1997) 3. Amoedo, S., Thebaud, E., Gschwendtner, M., White, D.: Novel parameter-based flexure bearing design method. Cryogenics 76, 1–9 (2016) 4. Jomde, A., Anderson, A., Bhojwani, V., Kharadi, F., Deshmukh, S.: Parametric analysis of flexure bearing for linear compressor. In: Materials Today: Proceedings, vol. 4, pp. 2478–2486, Materials Today: Proceedings (2017) 5. Khot, M., Gawali, B.: Finite element analysis and optimization of flexure bearing for linear motor compressor. Phys. Procedia 67, 379–385 (2015). https://doi.org/10.1016/j.phpro.2015. 06.044 6. Li, W., Mou, J., Hong, G.T.: Strain gauge for dynamic measurement of piston displacement in stirling engines. J. Vibrat. Measure. Diagn. 37(02), 372–376 (2017)

Simulation Study on Operation Characteristics of Cryocooler Regenerator Over Entire Cooling Process Chen Xiantong, Li Shanshan(B) , and Chen Xi Dalian Minzu University, No. 31 Jinshi Road, Dalian, China [email protected]

Abstract. In order to achieve a faster cooldown time for the cooled devices with rapid cooldown requirements, a larger cooling capacity of the cryocooler over the entire temperature range from ambient down to the cold operating temperature is required. The instantaneous cooling power depends on the phase characteristics at the cold end of the regenerator decided by the coupled phase-shifter and the operation mode over the whole cooling process. The performance of an existing regenerator with different temperatures and phase characteristics at the cold end was simulated based on the software REGEN 3.3. The simulation results can be used to guide the reasonable selection of the phase-shifter and operation mode in order to realize a fast cooldown. Keywords: Cryogenics regenerator · Rapid cooldown · REGEN 3.3 · Operation characteristics

1 Introduction Cryocoolers are widely used to cool the infrared components, cryogenics electronic components, high-temperature superconductors, etc. Some cryocooler applications require faster cooldown times for the cooled masses to enable rapid deployment [1–3]. The cooldown time depends on the instantaneous cooling power over the entire cooling process from the ambient down to the cold operating temperature. The key to achieve a fast cooldown is to increase the cooling capacity over the whole cooling process. The most wildly used cryocooler is the regenerative refrigerator, such as Stirling cryocooler and Stirling-type pulse tube cryocooler. Regenerator is the core segment for the regenerative refrigerator, for the phase characteristics at the cold and hot ends of the regenerator determines the cooling power and the efficiency of the cryocooler. The design principle of the regenerator structure dimensions is to provide the required refrigeration capacity and have a high efficiency at the cold operating temperature. The cooling capacity over the entire cooling process depends on the instantaneous phase characteristics at the cold end of the regenerator which are decided by the coupled phase shifter and the operation mode [4], on which few research has been carried out. This paper presented a detailed numerical simulation on the performance of a regenerator over the whole cooling process based on the software REGEN 3.3. The selection method of the phase shifter and © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 802–809, 2023. https://doi.org/10.1007/978-981-99-6128-3_104

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feasible operation mode to increase the cooling power over the entire cooling process was given based on the simulation results.

2 Parameters Settings of REGEN 3.3 REGEN is a professional software for the design of the regenerator which can efficiently simulate the heat transfer between the gas and the matrix [5]. The simulation study on the performance of a regenerator over entire temperature range was carried out based on the software REGEN 3.3 [6]. The length of the regenerator is 70 mm and the diameter is 34 mm. The matrix is 350# Stainless Steel Screen. The frequency is 65 Hz and the average pressure is 2.8 MPa. We remained the phase characteristics at the inlet of the phase shifter which included the mass flows, the pressure amplitudes, and the phase angles between them unchanged during the entire cooling process. The mass flow at the cold end of the regenerator is inversely proportional to the cold end temperature based on the phasor diagram of the cryocooler [4]. The calculated pressure ratio at the cold end of the regenerator varies from 1.18 to 1.3. The phase angle that the mass flow lags the pressure wave at the cold end of the regenerator is taken as 30°. The calculated cold end temperatures are 240K, 200K, 160K, 120K and 80K respectively. The coefficient of COOLING_MULT [6] in REGEN 3.3 which is used to reduce the regenerator gross refrigeration capacity for the existence of the pulse tube loss is all taken as 0.7.

3 Performance of Regenerator Over Entire Cooling Process Figures 1, 2, 3, 4, 5, 6, 7, 8, 9 and 10 show the changes of the cooling capacity, regenerator efficiency and phase characteristics at the hot end of regenerator over the whole cooling process with different cold end pressure ratios and mass flows.

Fig. 1. Changes of cooling capacity and regenerator efficiency at 80K

Fig. 2. Changes of hot end phase characteristics at 80K

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Fig. 3. Changes of cooling capacity and regenerator efficiency at 120K

Fig. 4. Changes of hot end phase characteristics at 120K

Fig. 5. Changes of cooling capacity and regenerator efficiency at 160K

Fig. 6. Changes of hot end phase characteristics at 160K

Fig. 7. Changes of cooling capacity and regenerator efficiency at 200K

Fig. 8. Changes of hot end phase characteristics at 200K

The efficiency of the regenerator is the ratio of the net refrigeration capacity calculated by the software REGEN 3.3 to the corrected acoustic work at the hot end of the regenerator described in Document [7]. The cold end temperature is 80K in Figs. 1 and 2, 120 K in Figs. 3 and 4, 160 K in Figs. 5 and 6, 200 K in Figs. 7 and 8 and 240 K in Figs. 9 and 10. It can be seen that the efficiency of the regenerator at 80 K decreases and that at the temperature higher than 120 K increases with the decrease of the cold end

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Fig. 9. Changes of cooling capacity and regenerator efficiency at 240K

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Fig. 10. Changes of hot end phase characteristics at 240K

pressure ratio. The larger pressure ratio at the cold end will lead to the larger pressure ratio and mass flow amplitude at the hot end of the regenerator for the temperatures above 120 K, that is, a relatively large compressor displacement is required at the higher temperatures. If a smaller cold end pressure ratio and a larger mass flow amplitude is selected at 80 K, the efficiency at the cold operating temperature will decrease, but the cooling capacity at the higher temperatures will increase for the constant compressor displacement which can reduce the cooldown time of the cooled object.

4 Selection of Phase Shifter and Operation Mode to Achieve Fast Cooldown 4.1 Operation Mode with Constant Phase Characteristics at Inlet of Phase Shifter Based on the research above, different combinations of the mass flow amplitude and pressure ratio at the cold end are selected and the cooling capacities for these combinations are all about 54 W at 80 K. Figures 11 and 12 show the changes of the cooling capacity, regenerator efficiency, hot end pressure ratio and mass flow amplitude over the whole cooling process with unchanged phase characteristics at the inlet of the phase shifter. It is approximately considered that the instantaneous cooling capacity at the cold end temperature of 80–120 K remains constant and equals to that at 80 K to simplify the calculation. The refrigeration powers at the 240 K, 200 K, 160 K, 120 K and 80 K can be used to represent the cooling capacity of the cryocooler over the entire temperature range from ambient down to the cold operating temperature. Then, the cooldown time which refers to the time that per unit mass stainless steel reaches 80 K from 300 K can be calculated [1]. As can be seen from Figs. 11 and 12, when the cold end mass flow amplitude increases from 0.012 to 0.021 kg/s and the cold end pressure ratio decreases from 1.3 to 1.183 at the cold operating temperature of 80 K, the regenerator efficiency at 80 K decreases by 25%. At the same time, the cooling power and regenerator efficiency increases at the temperatures higher than 120 K, which can reduce the cooldown time of the cooled object by approximately 8.4%.

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Fig. 11. Changes of cooling capacity and regenerator efficiency with constant phase characteristics at inlet of phase shifter

Fig. 12. Variation of hot end phase characteristics with constant phase characteristics at inlet of phase shifter

4.2 Operation Mode with Constant Pressure Ratio at Hot End of Regenerator The pressure ratio at the hot end of the regenerator is approximately equal to that at the outlet of the compressor. It is easy to control and adjust the pressure ratio at the hot end of the regenerator over the whole cooling process. The pressure ratio at the hot end remains constant over the whole cooling process and is equal to that at 80 K, then the cooling capacity for the mass flow and pressure ratio combinations in Figs. 11 and 12 is calculated over the entire cooling process. According to Figs. 1, 2, 3, 4, 5, 6, 7, 8, 9 and 10, the ratio of the pressure ratio at the hot to that at the cold end is approximately constant. Therefore, the cold end pressure ratio can be approximately taken as:  = Pr h,80 /(Pr h,T / Pr c,T ) Prc,T

(1)

where Prc,T and Prh,T are the cold and hot end pressure ratio of the regenerator at the cold temperature of T in Fig. 12, and Prh,80 is the hot end pressure ratio at 80 K. It is conservatively considered that the cooling capacity of the regenerator is approximately proportional to the square of the pressure amplitude at the cold end. Then, the refrigeration capacity of the regenerator is approximately equal to  = Qc,T · ( Qc,T

 −1 Prc,T  +1 Prc,T

)2 /(

Prc,T − 1 2 ) Prc,T + 1

(2)

where Qc,T is the cooling power at the cold end temperature of T in Fig. 11. The  are shown in Fig. 13. When the cold end mass flow amplitude calculation results of Qc,T increases from 0.012 to 0.021 kg/s and the pressure ratio decreases from 1.3 to 1.183 at the cold operating temperature of 80K, the cooldown time of the cooled object can be reduced by approximately 33%. It can be seen from Fig. 13 that the hot end mass flow amplitude at the higher temperatures is less than that at 80K, and the collision of compression piston with the cylinder will not happen. 4.3 Operation Mode with Constant Mass Flow at Hot End of Regenerator The displacement of the compressor can be measured by the displacement sensor, and is convenient to adjust and control. It is considered that the displacement of the compressor

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is approximately proportional to the mass flow amplitude at the hot end of the regenerator, which remains unchanged over the whole temperature range and is equal to that at 80 K. Then, the cooling capacity for the mass flow and pressure ratio combinations in Figs. 11 and 12 is calculated over the entire cooling process. The mass flow at the hot end is approximately proportional to the pressure amplitude at the cold end, while the cooling capacity is approximately proportional to the square of the pressure amplitude at the cold end. Therefore, the refrigeration capacity in this operation mode is approximately equal to:  = Qc,T · (mhd ,80 /mhd ,T )2 Qc,T

(3)

where mhd ,T is the hot end mass flow amplitude at the cold end temperature of T in Fig. 12, and mhd ,80 is the hot end mass flow amplitude at 80K. The calculation results  over the entire cooling process are shown in Fig. 14. When the cold end mass of Qc,T flow amplitude increases from 0.012 to 0.021 kg/s and the pressure ratio decreases from 1.3 to 1.183 at the cold operating temperature of 80 K, the cooldown time of the cooled object can be reduced by approximately 36.3%.

Fig. 13. Variation of cooling capacity and hot Fig. 14. Variation of cooling capacity with end mass flow with constant pressure ratio at constant mass flow at hot end of regenerator hot end

4.4 Selection of Phase Shifter and Operation Mode to Achieve Fast Cooldown As can be seen from Figs. 11, 13 and 14, when the cold end mass flow is 0.021 kg/s and the pressure ratio is 1.183 at 80 K, the cooldown time of the cooled object under the operation mode with the constant mass flow at the hot end of the regenerator is reduced by approximately 58.5% compared with that under the operation mode with the constant phase characteristics at the inlet of the phase shifter. Therefore, the operation mode with the constant hot end mass flow (compressor displacement) can be selected to greatly shorten the cooldown time of the cooled object. It can be seen from Sect. 4.1–4.3 that the phase shifter should be selected to provide a large mass flow amplitude under a small pressure ratio in order to decrease the cooldown time. For the pulse tube cryocooler, it is necessary to increase the volume of the coupled

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pulse tube to reduce the pulse tube loss. For Stirling cryocooler, the expansion chamber is located at the cold end and its volume is generally designed based on the phase characteristics at the cold operating temperature. Therefore, a larger mass flow at the higher temperatures may lead to the expansion piston colliding with the cylinder. Compared with the Stirling cryocooler, the pulse tube cryocooler uses the pulse tube to form a gas piston to avoid the possible collision of the expansion piston at the cold end and is more suitable for the applications required faster cooldown times [1]. It can be seen from Figs. 1, 2 and 11 that the efficiency of regenerator with a large mass flow and a small pressure ratio at the cold end is lower than that with a small mass flow and a large pressure ratio at 80K. Therefore, if a dual-driven active displacer is used as the phase shifter for the pulse tube cryocooler, the phase characteristics can be changed at the cold end to achieve a small pressure ratio at the higher temperatures to increase cooling power and a large pressure ratio at the cold operating temperature to improve the efficiency. The pulse tube cryocooler with a dual-driven active displacer as the phase shifter can not only reduce the cooldown time of the cooled object, but also achieve a high efficiency at the cold operating temperature.

5 Conclusion (1) Different phase characteristics at the cold end of the regenerator can be selected to obtain the required cooling power at the cold operating temperature. A smaller mass flow amplitude and a larger pressure ratio at the cold end can obtain a higher regenerator efficiency at the cold operating temperature. (2) For the applications required faster cooldown times, the phase shifter should be selected to provide a large mass flow under a small pressure ratio and the operation mode with constant hot end mass flow (compressor displacement) should be selected. (3) The pulse tube cryocooler with a dual-driven active displacer as the phase shifter can not only reduce the cooldown time of the cooled object, but also achieve a high efficiency at the cold operating temperature. Acknowledgment. The project is supported by National Natural Science Foundation of China (Number 51906032).

References 1. Radebaugh, R., O’Gallagher, A., Lewis, M. A., Bradley, P. E.: Proposed rapid cooldown technique for pulse tube cryocoolers, In: Miller S.D., Ross, Jr. R.G. (eds.) Cryocooler 14, pp. 231–240. ICC, Boulder,CO. (2007) 2. Choi, Y.S., Kim, D.L., Shin, D.W.: Cool-down characteristics of conduction-cooled superconducting magnet by a cryocooler. Physica C 471, 1440–1444 (2011) 3. Choi, Y.S., Kim, D.L., Shin, D.W.: Optimal cool-down time of a 4K superconducting magnet cooled by a two-stage cryocooler. Cryogenics 52, 13–18 (2012) 4. Radebaugh, R.:Thermodynamics of Regenerative Refrigerators, Generation of Low Temperature and It’s Applications, pp.1–20 (2003)

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5. Lewis, M., Kuriyama, T., Xiao, J. H., Radebaugh, R.: Effects of regenerator geometry on pulse tube refrigerator performance, In: Kittel, P.(eds) Advances in Cryogenic Engineering, vol. 43, pp. 1999–2005. Springer, Boston, MA. (1998) 6. Gary, J., O’Gallagher, A., Radebaugh, R., Huang, Y. H., Marquardt, E.: REGEN 3.3: User Manual, NIST ( 2008) 7. Li, S.S., Jiang, Z.H., Fan, W., Wu, Y.N.: Indirect measurement of regenerator phase characteristics for cryogenics refrigerators. Cryogenics 101, 12–21 (2019)

Optimization of Two-Stage High-Efficiency Pulse Tube Cryocooler for Space Application at 20K Q. L. Zhu1 , X. Y. Li1 , T. H. Huang1 , J. Sun1 , L. Huang1,1(B) , J. Quan2 , and J. T. Liang2 1 Wuhan Global Sensor Technology Co., LTD, Guandong Street, Wuhan, China

[email protected] 2 Key Laboratory of Technology on Space Energy Conversion, Technical Institute of Physics

and Chemistry, Chinese Academy of Science, Zhongguancun Street, Beijing, China

Abstract. A high-efficiency pulse tube cryocooler at 20 K is a prerequisite for closed JT coolers to obtain 100 [email protected] K in space. Therefore, it is of strategic importance to develop a two-stage pulse tube cryocooler with high efficiency at 20 K for space applications. Based on this, the low temperature stage compressor, the low temperature stage regenerator structure, the cryogenic inertance tube size and the cryogenic gas reservoir volume were optimized in this paper to improve the performance. Finally, an efficient and compact space pulse tube cryocooler at 20 K was obtained. The cryocooler can obtain a cooling capacity of 670 mW@20 K with a relative Carnot efficiency of 2.54% when the total input power is 370 W, the second stage operating frequency is 27 Hz and the second stage charging pressure is 1.5 MPa. Keywords: Two-stage pulse tube cryocooler · High-efficiency · Space application · 20 K

1 Introduction At present, the best solution to obtain the liquid helium temperature in space is a precooled multi-stage closed JT cooler, in which the pre-cooling stage uses exactly a multistage pulse tube cryocooler, whose performance directly affects the efficiency of the JT cooler. Therefore, it is of great strategic importance to explore the development of a multi-stage pulse tube cryocooler at 20 K. Coulter D R et al. designed a four-stage high-frequency pulse tube cryocooler which was based on the proven three-stage highfrequency pulse tube cryocooler. The cryocooler which can obtain 20 mW@6 K and 150 mW@18 K with the input power of 208 W and operating frequency of 30 Hz, and the lowest no-load temperature can reach 4.85 K [1–3]. At the same time, C. Jaco et al. developed a three-stage high-frequency pulse tube cryocooler with the cooling performance of 400 mW@15 K while the input power of 300 W [4]. In 2016, V. Kotsubo et al. used a three-stage gas-coupled pulse tube cryocooler as the pre-cooling stage of the JT cooler in order to cool a superconducting nano single-photon detector. The prototype © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 810–817, 2023. https://doi.org/10.1007/978-981-99-6128-3_105

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can obtain a cooling capacity of 30 mW@15 K at a total input power of 150 W and an operating frequency of 35 Hz [5]. Duval J.M. et al. developed a two-stage thermalcoupled pulse tube cryocooler, which can obtain a cooling capacity of 300 mW@15 K when the input power of the compressor was 290 W and the pre-cooling temperature was 110 K [6]. Most of the domestic research advances on multi-stage pulse tube cryocoolers were concentrated in Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Shanghai Institute of Technical Physics, Chinese Academy of Sciences and Zhejiang University [7–10].

2 Experimental Setup The structure of the existing two-stage pulse tube cryocooler in our laboratory was thermal-coupled to improve the stability for space applications as shown in Fig. 1. In addition, the compressor of each stage can work at its own optimal frequency, making the whole machine relatively high efficiency. In the figure, “a” indicates the first stage (precooling stage) pulse tube cryocooler, and “b” indicates the second stage (low temperature stage) pulse tube cryocooler. First of all, the pre-cooling stage pulse tube cryocooler selected a single-stage pulse tube cryocooler with a cooling performance of 10 W@80 K, phased only by the inertance and reservoir; secondly, the low temperature stage pulse tube cryocooler selected a phase modulation with low temperature inertance, low temperature reservoir and double-inlet to enhance the cooling performance; finally, according to literature research, the stainless steel mesh cooling efficiency is relatively high above 20 K, therefore, the pre-cooling stage and low temperature stage regenerators were both filled with the SS (stainless steel mesh). The key parameters of the first stage cryocooler were shown in Table 1 and the parameters of the second stage cryocooler were shown in the part 3. Table 1. Key parameters of the first stage cryocooler key parameters

value

The first stage compressor piston diameter/stroke

25 mm/±4.5 mm

The second stage compressor piston diameter/stroke

35 mm/±5.5 mm

The first stage regenerator length/diameter

50 mm/25.4 mm

The first stage pulse tube length

55 mm

3 Optimization Results The prototype was originally optimized to obtain liquid helium temperature, and a cooling capacity of 170 mW@20 K was obtained. It’s clear that the prototype needs to be optimized accordingly to improve the overall efficiency at 20 K. Based on this, we mainly optimized three parts of the cryocooler included the structure of the low temperature stage regenerator, the cryogenic phase shifter and the low temperature stage compressor, and the optimization results are shown as follows.

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Fig. 1. Schematic of the two-stage pulse tube cryocooler

3.1 Optimization of Low Temperature Stage Regenerator Structure The variable cross-section regenerator has a multi-bypass structure at the variable crosssection to enhance the capability of phase shifter, but the amount of working gas entering the cold end is reduced simultaneously, resulting in a relatively small amount of cooling capacity. While the non-variable cross-section regenerator ensure the amount of working gas at the cold end with eliminating the multi-bypass structure, resulting in a higher refrigeration efficiency. For convenience, we will call the two-stage pulse tube cryocooler with the variable cross-section regenerator as cryocooler “A”, and the two-stage pulse tube cryocooler with the non-variable cross-section regenerator as cryocooler “B”. The key parameters of the two cryocoolers cold finger were shown in Table 2. The pre-cooling stage conditions of the two cryocoolers were all kept the same, and only the relevant parameters of the low temperature stage cryocooler were changed during the experiment, it should be noted that the charging pressure of the following two sets of experimental data had been optimized. Here the optimal charging pressure for cryocooler A is 2 MPa, while the optimal charging pressure for cryocooler B is 1 MPa. Table 2. Key parameters of the two cryocoolers cold finger key parameters

A

B

The transition regenerator length/diameter

18 mm/13.7 mm

40 mm/18 mm

The transition regenerator filling

300 & 400#SS

400#SS

The second stage regenerator I length/diameter

67 mm/13.7 mm

40 mm/18 mm

The second stage regenerator II length/diameter

30 mm/9.8 mm

–

The second stage pulse tube Ilength/diameter

65 mm/7 mm

45 mm/8.5 mm

The second stage pulse tube II length

35 mm

–

The multi-bypass orifice diameter

0.52 mm

–

The cryogenic gas reservoir volume

30 cc

230 cc

Figure 2 shows the typical cooling performance curves of two different structure cryocoolers. The cooling slope of cryocooler B in the 20 K temperature region is 23.92

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mW/K, and that of cryocooler A is only 9.7 mW/K, which is less than half of cryocooler B. The reason for the above results is mainly caused by the different structures of the low temperature stage regenerator. For cryocooler A, there is a multi-bypass structure, which leads to the reduction of the cold end mass flow, which directly reduces the amount of working gas at the cold end and greatly reduces the cooling ramp rate. Ultimately, cryocooler B was able to obtain a cooling capacity of 202 mW@20 K at the second stage cold end approximately when the charging pressure was 1 MPa, the operating frequency was 25 Hz. Cryocooler A Cryocooler B

Cooling power/mW

200

150

100

50

0

8

10

12

14

16

18

20

22

24

26

28

Temperature/K Fig. 2. Typical cooling performance curve of two different cryocoolers

3.2 Optimization of Low Temperature Stage Phase Shifter In a general way, the phase shifter capability of cryogenic inertance tubes is stronger than room temperature inertance tubes under small input acoustic power conditions, which enables the cryocooler to obtain better cooling performance. a. Inertance tubes. In order to investigate the capability of cryogenic phase shifter, the effect of four different inertance tubes sizes on the performance of the two-stage pulse tube cryocooler was experimentally investigated firstly. In view of the previous theoretical and experimental studies on the capability of cryogenic phase shifter, it is known that the first inertance tube takes up most of the phasing tasks at low temperature, so the length of the first inertance tube has the greatest influence on the performance of the overall. In the experiment, we kept the length of the second and third inertance tube unchanged, only changed the length of the first inertance tube, its specific dimensions are shown in the Table 3.

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2 mm × 1.0 m + 3 mm × 2 m + 4 mm × 2 m

Case02

2 mm × 1.3 m + 3 mm × 2 m + 4 mm × 2 m

Case03

2 mm × 1.5 m + 3 mm × 2 m + 4 mm × 2 m

Case04

2 mm × 2.0 m + 3 mm × 2 m + 4 mm × 2 m

Figure 3 shows the typical cooling performance curves of the cryocooler under four different combinations of cryogenic inertance tubes. It can be clearly seen that as the length of the first cryogenic inertance tube increases, the no-load temperature of the cryocooler gradually increases, but finally it is all concentrated around 12 K with little change. The reason for the above results is mainly caused by the different operating frequency of the second stage cryocooler. In this condition, the higher the second stage operating frequency, the lower the no-load temperature of the cryocooler. In addition, Case03 showed the best performance in four combinations of cryogenic inertance tubes, and it was able to obtain 313.5 mW@20 K when the input power was 100 W and the charging pressure was 1 MPa.

300

Cooling power/mW

250

Case01 Case02 Case03 Case04

200

150

100

50

0

10

12

14

16

18

20

Temperature/K Fig. 3. Variation curve of typical cooling performance under different cryogenic inertance tube combinations

b. Gas reservoir. The role of the gas reservoir as another important component of the phase shifter has been simplified and neglected gradually. Generally, the gas reservoir is used with

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inertance tubes to provide the desired phase relationship between the pressure wave and the mass flow at the cold end of the pulse tube cryocooler, and the gas reservoir volume will have a certain impact on the performance of the pulse tube cryocooler. Figure 4 shows the typical cooling performance curves of the cryocoolers equipped with two different volumes of gas reservoirs. As can be seen from the graph, the larger the volume of the gas reservoir, the lower the no-load temperature obtained by the cryocooler.

300

Cooling power/mW

250

200

150

100

50

0

10

12

14

16

18

20

Temperature/K Fig. 4. Variation curve of typical cooling power under different gas reservoir volumes

The reason for the above results is that with the increasing of the gas reservoir volume, the internal pressure fluctuation of gas reservoir will gradually weaken, and the range of the impedance phase angle between the pressure wave and the mass flow will gradually increase, both of which will improve the overall performance to a certain extent. What’s more, the slope of the curve shows that increasing the volume of the gas reservoir does not significantly improve the cooling slope of the cryocooler, the difference between the two slope is less than 1 mW/K, which can indicate that the positive effect of increasing gas reservoir volume is attenuated at low temperatures. However, since the no-load temperature of cryocooler B is relatively lower, its cooling capacity at 20 K is relatively higher, the cooling capacity at 20 K was increased from 242.8 mW to 295.6 mW when the volume of gas reservoir was increased from 30 cc to 230 cc. Nevertheless, we chosen the cryocooler A to adapt space applications when the cooling capacity improvement was not significant. 3.3 Optimization of Low Temperature Stage Compressor Previously, the low temperature stage compressor had been using a Leybold linear compressor equipped with column spring (Compressor A), which was found to have a low

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resonance frequency and was not compatible with the optimized cold finger (non-variable cross-section structure), reducing the overall efficiency. Based on this, our laboratory developed a low temperature stage compressor with a relatively higher resonance frequency (Compressor B) to make it more compatible with the existing cold finger. The experimental results of matching the two compressors with the cold finger are shown in the Fig. 5.

Cooling power/mW

200

Compressor A Compressor B

150

100

50

0 10

12

14

16

18

20

Temperature/K Fig. 5. Variation curve of typical cooling capacity under different compressor

The overall cooling performance of the cryocooler with compressor B is significantly better than the cryocooler with compressor A under the same input power. This indicates that compressor B which works at the resonant frequency and its energy conversion efficiency is the highest is better matched with the newly designed low temperature stage cold finger. After the experimental test, the cryocooler with compressor B was able to obtain a cooling capacity of 200 mW@20 K at an input power of 50 W, however, the cryocooler with compressor A could only obtain a cooling capacity of 36 mW@20 K with the same input power, which is a poor coupling match. Through the optimization of the low temperature stage compressor, the overall efficiency at 20 K is significantly improved, and the relative Carnot efficiency of the whole machine can reach 5.6%.

4 Conclusion By optimizing the low temperature stage regenerator structure, the cryogenic inertence tube, the cryogenic gas reservoir and the low temperature stage compressor of the twostage pulse tube cryocooler, the overall cooling capacity at 20 K is significantly improved under the same input power. The experimental results show that a cooling capacity of

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670 mW@20 K and the overall relative Carnot efficiency of 2.54% which meets the demand of 0.5 W@20 K space IR detector can be achieved when the total input power is 370 W, the second stage operating frequency is 27 Hz and the second stage charging pressure is 1.5 MPa. Acknowledgements. This work was supported by National Key R&D Plan under grant NO. 2018YFB0504603 and the National Basic Research Program of China (Grant No.613322).

References 1. Linden smith, C.A., Bowman, R.C., Wade, L.A., et al.: Cryocooler options for NGST and other space applications. Next Generation Space Telescope Science and Technology, vol. 207, p. 106 (2000) 2. Ross, R., Boyle, R.F., Kittel, P.: NASA space cryocooler programs: a 2003 overview. Am. Inst. Phys. (2004) 3. Coulter, D.R., Ross, Jr R.G., Boyle, R.F., et al.: NASA advanced cryocooler technology development program. In: Astronomical Telescopes and Instrumentation. International Society for Optics and Photonics, pp. 1020–1028 (2003) 4. Jaco, C., Nguyen, T., Raab, J.: 10 K pulse tube cooler performance date. In: Cryocoolers, vol. 15, pp. 1–6 (2009) 5. Kotsubo, V., Radebaugh, R., Hendershott, P., et al.: Compact 2.2 K cooling system for superconducting nanowire single photon detectors. IEEE Trans. Appl. Superconduct. 27(4) (2017) 6. Duval, J.M., Charles, I., Coynel, A., et al.L Development of 15 K pulse tube cold fingers for space applications at CEA/SBT. Cryocoolers 16, 45–50 (2011) 7. Wu, X., Chen, L., Liu, X., et al.: An 80 mW/8 Khigh-frequency pulse tube refrigerator driven by only one linear compressor. Cryogenics 101, 7–11 (2019) 8. Dang, H., Zha, R., Tan, J., et al.: Investigations on a 3.3 K four-stage Stirling-type pulse tube cryocooler. part A: theoretical analyses and modeling. Cryogenics 105, 103014 (2020) 9. Dang, H., Zha, R., Tan, J., et al.: Investigations on a 3.3 K four-stage Stirling-type pulse tube cryocooler. part B: experimental verifications. Cryogenics 105, 103015 (2020) 10. Gan, Z., Zhuopei, et al.: Design and preliminary experimental investigation of a 4K Stirlingtype pulse tube cryocooler with precooling. J. Zhejiang Univ. (2009)

Effect of Mixture Composition on Thermophysical Properties of Refrigerant Mixture for MRJT Cryocooler Darshit Parmar(B) and M. D. Atrey Indian Institute of Technology Bombay, Mumbai 400076, Maharashtra, India [email protected]

Abstract. Mixed Refrigerant Joule-Thompson (MRJT) cryocoolers are gaining popularity in low temperature cooling applications due to their lower cost. Mixture of refrigerant used in MRJT cryocoolers dictate the performance of the cryocooler to a great extent. MRJT cryocooler consists of a recuperative heat exchanger where the high and low pressure streams exchange heat for higher efficiency. Thermophysical properties of the mixed refrigerant in the recuperative heat exchanger of an MRJT cryocooler is responsible for a significant contribution of the irreversibility in the cryocooler. In order to minimise these irreversibilities, the present work attempts to analyse the thermophysical properties of the refrigerant mixtures with respect to the mixture constituents and their concentration. Ratio of apparent specific heat capacity for the two streams in the heat exchanger with an objective to minimize the irreversibilities associated with the heat transfer process. The concept of distributed J-T effect and mass bleed is introduced to reduce the irreversibilities in the heat exchanger. Keywords: MRJT cryocooler · Optimisation · Mixture properties

1 Introduction Joule-Thompson (J-T) refrigerators are used for a wide range of cooling applications, however they have not been extensively used for cryogenic temperatures. Main challenges with regards to J-T refrigerators at cryogenic temperatures are their low efficiency and high operating pressures. With introduction of Mixture of Refrigerants as working fluid, both of these issues are taken care of. Such refrigerators are known as Mixed Refrigerant Joule-Thompson (MRJT) refrigerators [1]. For a cooling temperature of up to 80 K, the refrigerant mixture generally consists of low boiling component like Nitrogen, medium boiling components like Methane and Ethane and high boiling components like Propane and Isobutane [2].

2 MRJT Cryocooler There has been a significant increase in research on J-T cryocoolers after the introduction of refrigerant mixtures as working fluid. The study on the MRJT cryocooler involves experimental as well as theoretical investigation. The experimental analysis generally © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 818–824, 2023. https://doi.org/10.1007/978-981-99-6128-3_106

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involves developing MRJT cryocooler for a given cooling temperature. Depending on the refrigerant composition, a particular temperature profile of cold and hot fluid in the heat exchanger can be obtained. G. Venkatarathnam et. al. [3, 4] have studied the occurrence of pinch point in heat condensers and evaporators of zeotropic refrigerant mixtures. The analysis was based on the total heat capacity of the cold and hot streams. A condition based on the rate of change of enthalpy at the ends of the heat exchanger was also represented. Optimisation of MRJT cryocooler was carried out by several researchers [4–7] in the temperature range of 90 K to 120 K. The objective function for the optimisation may be the maximisation of the cooling effect or the exergetic efficiency. Different MRJT cycles were also explored in order to reach lower temperatures with a higher efficiency [6–8]. It was reported that the overall performance of the MRJT cryocooler depends on the temperature profile of the cold and the hot streams in the recuperative heat exchanger. The temperature profile of the cold and hot streams in the heat exchanger in turn depends on the thermophysical properties of the refrigerant mixture. Therefore, in the present analysis, an investigation is carried out to understand the heat exchanger performance with respect to the refrigerant mixture composition. Thermophysical properties such as heat capacity ratio of both the streams are presented for better understanding. The effect of distributed J-T effect and mass bleed effect is also briefly investigated for improving the performance of the MRJT cryocooler. This information can be useful when determining an appropriate initial guess of optimisation of complex MRJT cycles.

3 Theoretical Model In order to understand the effect of individual components on the performance of the MRJT cryocooler, a theoretical model is developed using ASPEN software package [9]. Here, the cycle shown in Fig. 1 is modelled in ASPEN Plus for different evaporator temperature. For the present study, the high and low pressure in the cycle are kept constant at 1.5 MPa and 0.2 MPa respectively. Optimisation of the mixture composition is carried out using exergetic efficiency as the optimisation function. For simplification of the mode, the following assumptions are made. Assumptions – – – –

The compression process is assumed to be isentropic. Inlet to the compression is above dew point temperature. Cooling effect is available at constant evaporator temperature. Minimum Internal Temperature Approach inside the recuperative temperature is taken as 2 K.

For the MRJT cryocooler, exergy gained is the cooling effect (Qe ) available at the evaporator temperature (Te ) while the exergy input is the work of compression (Wc ) and T0 is the ambient temperature. The exergy can be calculated as per Eq. 1,   Qe T0 ηex = −1 (1) Wc Te

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For the present analysis, three different evaporator temperatures i.e. 90 K, 100 K and 120 K are considered. Firstly, the optimisation of mixture composition is carried out in ASPEN Plus for all the three cases. Exergetic efficiency is used as the objective function for the optimisation routine. Following this, a parametric study is carried out to understand the effect of individual components of the refrigeration mixture on the MRJT performance.

4 Results and Discussion Temperature profiles in the heat exchanger of the optimised MRJT cryocoolers for the evaporator temperature of 90 K, 100 K and 120 K are analysed to understand the effect of refrigerant mixture composition on the overall performance. The temperature profiles (T) of the cold and the hot stream vs fraction of heat transferred (q) in the heat exchanger are shown in Fig. 1(a)–(c) for evaporator temperature of 90 K, 100 K and 120 K respectively. Here q equal to 0 corresponds to the cold end of the heat exchanger while q equal to 1 corresponds to the hot end of the heat exchanger. The corresponding optimised mixture composition is also shown in the respective figures. 300

300

250 Temperature (K)

Temperature (K)

Hot stream Cold stream

200

150

250

200

150

100 0.0

Hot stream Cold stream

N2/CH4/C2H6/C3H8/C4H10: 27.3/25.3/11.1/6.0/30.2

90 K

0.2

0.4

0.6

0.8

100 0.0

1.0

0.2

Fraction of heat transferred

0.8

1.0

50 Hot stream Cold stream

Temperature difference (K)

Temperature (K)

0.6

(b) Heat Exchanger temperature profile (Te=100 K)

250

200

150

N2/CH4/C2H6/C3H8/C4H10: 14.9/41.7/12.3/0.1/31.1

120 K

0.0

0.4

Fraction of heat transferred

(a) Heat Exchanger temperature profile (Te=90 K) 300

N2/CH4/C2H6/C3H8/C4H10: 26.5/29.0/12.7/3.6/28.3

100 K

0.2

0.4

0.6

0.8

Fraction of heat transferred

(c) Heat Exchanger temperature profile (Te=120 K)

40

90 K 100 K 120 K

30

20

10 2K

1.0

0 0.0

0.2

0.4

0.6

0.8

1.0

Fraction of heat transferred

(d) Delta T for different evaporator temperature

Fig. 1. Temperature profile inside heat exchanger

The temperature difference between the hot and the cold stream inside the heat exchanger is also compared. The comparison between the temperature differences in

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the heat exchangers for the three cases is also shown in Fig. 1(d). At the warm end of the heat exchanger, the hot stream is in two phase state while the cold stream is still in vapour phase. As a result, the temperature difference between the two streams gradually increases till the point the cold stream condenses. From this point onwards, the temperature difference gradually reduces till the warm end of the heat exchanger. It may also be noted from the figures, that the temperature difference is minimum inside the heat exchanger between 150 K to 200 K for all the cases. Also, the temperature difference at the warm end of the heat exchanger for the evaporator temperature of 90 K is higher compared to 100 K and 120 K. This is mainly due to higher concentration of high boiling components. This also results in having the dew point of the low pressure stream at a higher value of fraction of heat transferred (q). From the figure, it may also be noted that the temperature difference in the heat exchanger in the 125–175 K temperature range is higher for the evaporator temperature of 120 K. This is due to the lower concentration of Nitrogen in the refrigerant mixture. 4.1 Parametric Study Having obtained the insights into the optimised performance of the MRJT cryocooler for different evaporator temperatures, an attempt is made to understand the effect of individual components on the performance of the MRJT cryocooler. The thermophysical property of importance here is the heat capacity ratio between the high and low pressure streams. The optimised mixture composition is tweaked by varying the individual components by ±5% while keeping the proportion of other components constant. The effect of heat capacity on the temperature difference between the hot and cold streams in the heat exchanger is also presented. The temperature difference in the heat exchanger (T) and the heat capacity ratio (Cr ) with respect to fraction of heat transferred (q) in the heat exchanger are presented in Figs. 2(a)–(e) respectively for variations in Nitrogen, Methane, Ethane, Propane and Isobutane respectively. The temperature difference between the two streams in the heat exchanger is shown as solid lines while the heat capacity ratio is shown as dotted lines. Black line indicates the optimised refrigerant mixture, while the red and blue line indicate the refrigerant mixture with higher and lower quantities of the respective component respectively. The results are presented for the evaporator temperature of 100 K. As can be seen from the Fig. 2(a) that if the concentration of Nitrogen is increased, the heat capacity ratio reduces at higher temperature (q ~ 0.3–0.8) but increases at lower temperatures (q ~ 0.2) and vice versa. As the heat capacity ratio reduces at around q = 0.6 in the heat exchanger, the amount of heat required for temperature change of hot stream reduces compared to cold stream resulting in higher temperature change in the temperature of the hot stream. Due to this, the temperature of the hot stream theoretically becomes lower than the cold stream and a temperature cross is observed inside the heat exchanger. However, at q = 0.1–0.3, the heat capacity ratio is higher resulting in a higher change in temperature of the low pressure stream. Therefore, the temperature of the cold stream again becomes lower than that of the hot stream. The condition of MITA being greater than 2 K is satisfied at the ends of the heat exchanger, however, there is violation of that particular constraint inside the heat exchanger. Thus, this particular mixture composition not suitable to reach the desired temperature. However, when the

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nitrogen concentration is reduced, the heat capacity ratio varies in such a way that the overall temperature difference in the heat exchanger increases, reducing the performance of the MRJT cryocooler.

Fig. 2. Temperature difference and Heat capacity ratio vs Fraction of heat transferred (q)

The effect of change in Methane composition on the thermophysical properties as shown in Fig. 2(b) is similar to that of Nitrogen. However, the trend is reversed i.e. with an increase in methane concentration, the heat capacity ratio increases at q = 0.2–0.4 while reduces at q = 0.1–0.2 and vice versa. As a result, there is temperature cross in the heat exchanger when the Methane concentration is reduced while the temperature difference in the heat exchanger increases on increasing the Methane concentration. Comparing the thermophysical properties of the refrigerant mixture by changing Ethane as given in Fig. 2(c), it can be seen that the heat capacity ratio is higher for q = 0.4–0.8 and is lower for q = 0.2–0.4 when the concentration of ethane is increased. It may also be noted that the region where the heat capacity ratio is affected moves slightly to the right side (higher temperature). As a result, the change in heat exchanger temperature difference is seen mostly for q = 0.2–0.8. From Figs. 2(d) and (e), it can be observed that the thermophysical properties of refrigerant mixture does not vary significantly by changing the composition of Propane and Isobutane. Thus, based on the available data, the temperature profile in the heat exchanger can be tweaked as per desire. This can be particularly useful when modelling complex MRJT cycles where if there is temperature cross in the heat exchanger, it

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can be removed by changing the composition of individual components. Such activity is generally carried out to determine the initial guess during optimisation of MRJT cycles. Initial guess for the optimisation can be quite crucial especially where there are large number of independent variables. Better initial guess will not only accelerate the optimisation process but also result in a global optimum rather than a local optimum. 4.2 Improvement of MRJT Cryocooler The irreversibilities in the heat exchanger due to heat transfer across a finite temperature difference deteriorate the MRJT performance significantly. From the previous section, it may be seen that a higher temperature difference is observed near the warm end of the heat exchanger. This higher temperature difference is mainly due to lower heat capacity of the low pressure stream. To reduce the temperature difference in the heat exchanger of the MRJT cryocooler two mechanisms are suggested with the aim of increasing the heat capacity ratio of the low pressure stream. First mechanism is bleeding of mass from high pressure stream to be mixed with low pressure return stream. Second mechanism is the implementation of distributed J-T effect. Here the high pressure stream is partially expanded in the expansion valve and the rest of pressure drop is carried out in the low pressure stream of the heat exchanger. In the mass bleeding mechanism, the heat capacity of the low pressure stream is increased by increasing the mass flow rate in a fraction of the heat exchanger. In the distributed J-T mechanism, excess cooling effect is provided in the low pressure stream of the heat exchanger indirectly increasing the heat capacity of the low pressure stream. Various operating parameters of both the cycles are optimised to obtain the best performance from the modified cycles. The temperature differences in the heat exchanger for the original and two improved cycles are shown in Fig. 2(f). From the figure, it may be noted that the temperature profile for all the three cases are identical with minor variations. However, for the cycle with mass bleed, there is a slight jump in the temperature near q = 0.2. This is due to mixing of the bled stream into the low pressure stream. Here, the overall temperature difference for the cycle with mass bleed is lower near the cold end of the heat exchanger which can improve the performance of the MRJT marginally. However, as a fraction of mass is bled from the high pressure stream, the overall mass flow rate to the evaporator reduces. For the MRJT cycle with distributed J-T effect, the partial expansion of the refrigerant mixture in the heat exchanger may improve the heat exchanger temperature profile, but the cooling effect available in the evaporator reduces as only partial expansion happens in the expansion valve. As a result, all the three cycles i.e. original and two modified ones, show identical performance for the present case (evaporator temperature of 100 K). The modified cycles may not show improvement for the present case, however, they may be useful for lower temperatures (typically less than 80 K).

5 Conclusions A comprehensive analysis of variation of thermophysical properties of the refrigerant mixture with respect to individual components is presented. Based on the study, following conclusions are drawn.

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• The performance of the MRJT cryocooler depends largely on the temperature profiles in the heat exchanger which in turn depends on the heat capacity ratio. • The heat capacity ratio of the two streams depends primarily on low and medium boiling components while does not vary significantly with respect to high boiling components. • The two mechanisms proposed to improve the performance of the MRJT cycle show promise, however, for the temperature of present study, it does not affect the performance significantly for cooling temperatures above 90 K.

References 1. Brodyanskii, V., Gromov, E., Grezin, A., Yagodin, V., Nikol’skii, V., Tashchina, A.: Efficient throttling cryogenic refrigerators which operate on mixtures. Chem. Pet. Eng. 7(12), 1057–1061 (1971) 2. Venkatarathnam, G.: Cryogenic Mixed Refrigerant Processes. Springer, New York (2008) 3. Venkatarathnam, G., Mokashi, G., Murthy, S.: Occurrence of pinch points in condensers and evaporators for zeotropic refrigerant mixtures. Int. J. Refrig. 19, 361–368 (1996) 4. Venkatarathnam, G., Murthy, S.: Effect of mixture composition on the formation of pinch points in condensers and evaporators for zeotropic refrigerant mixtures. Int. J. Refrig. 22, 205–215 (1999) 5. Alexeev, A., Quack, H.: Refrigerant mixture for a mixture-throttling process. US Patent 6513338 B1 (2003) 6. Gong, M., Wu, J., Luo, E.: Performances of the mixed-gases Joule-Thomson refrigeration cycles for cooling fixed-temperature heat loads. Cryogenics 44, 847–857 (2004) 7. Gong, M., Luo, E., Wu, J., Zhou, Y.: On the temperature distribution in the counter flow heat exchanger with multicomponent non-azeotropic mixtures. Cryogenics 42, 795–804 (2002) 8. Asadnia, M., Mehrpooya, M.: A novel hydrogen liquefaction process configuration with combined mixed refrigerant systems. Int. J. Hydrogen Energy 42, 15564–15585 (2017) 9. Aspentech: aspenONE, version 11 (2019)

Numerical Simulation of Two-Stage Gas Coupled High Frequency Cryocooler Wu Xianlin1,2(B)

, Liu Sixue1,2 , Yang Sujun1,2 , Jin Yu1,2 , Zhou Qiang1,2 , Chen Liubiao3 , and Zhou Yuan3

1 Beijing Institute of Spacecraft System Engineering, Beijing, China

[email protected]

2 Beijing Key Laboratory of Space Thermal Control Technology, Beijing, China 3 Technical Institute of Physics and Chemistry, CAS, Beijing, China

Abstract. A two-stage gas-coupled high-frequency pulse tube cryocooler (HFPTC) driven by a linear dual-opposed compressor has been designed. The coaxial structure is used for compactness in both stages. Inertance tubes and double-inlet values are adopted for each stage. A numerical simulation model was developed. The influence of inertance tube and double-inlet on the cooling performance of cryocooler was simulated. The influence of the phase distribution and the character of the gas distribution between two stages were analyzed. It indicates that the second stage’s inertance tube (IT 2) has less influence on the gas distribution in the first stage. However, the first stage’s inertance tube (IT 1) has a significant influence on the second stage. Meanwhile, the value of double-inlet has important influence on the performance of the cryocooler. The double-inlet value of the first stage (DI 1) has significant influence on the output characteristics of the compressor and the gas volume distribution between the two stages. However, the double-inlet value of the second stage (DI 2) has a smaller effect on the output PV characteristics and the phase distribution of the first stage. At the same time, it influences the gas volume distribution and the phase distribution of the second stage. Keywords: Two stage · Gas-coupled · Phase distribution · Gas distribution

1 Introduction HFPTC has the advantages of compactness, small size, low vibration, long life and high reliability due to the absence of moving parts at the cold end. In this way, it has attracted the interest of many researchers. At present, PTC working at around 80 K can achieve a relative Carnot efficiency larger than 20% [1]. It can achieve a cooling capacity of 0.2–1 W at 35 K within 300 W input power [2]. In the lower temperature region, the HFPTC efficiency will be much lower. Multistage structure is adopted to achieve lower temperature [3, 4]. Gas-coupled and thermal-coupled HFPTC are the main coupling structure. Thermal-coupled HFPTCs are easier to adjust the structure of each stage. However, it has drawbacks in terms of compactness and efficiency. Gas-coupled HFPTC can solve the above problem. Due to the complicated gas flow distribution, it is difficult to make each stage work in its optimal range. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 825–831, 2023. https://doi.org/10.1007/978-981-99-6128-3_107

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Many attempts have been made to achieve a lower temperature with gas-coupled HFPTC. Zhu has developed a two-stage gas-coupled SPTC based on a multi-bypass single-stage HFPTC, it has achieved a no-load temperature of 12.1 K with an input electric power of 260 W [5]. Pang has developed a two-stage gas-coupled SPTC with a lowest temperature of 13.6 K and a cooling capacity of 1.06 W at 20 K with 224 W acoustic power [6]. Wu has achieved a lowest temperature of 5.7 K based on a high cooling capacity single-stage HFPTC [7]. In this paper, a two-stage gas-coupled SPTC has been developed based on Sage software. The simulation model of the cryocooler to was established to discuss the influence of inertance tube and double-inlet on the performance of the cryocooler.

2 Design of the Cryocooler A two-stage gas-coupled HFPTC driven by a linear dual-opposed compressor was designed. The schematic of the PTC was shown in Fig. 1. Both of the two stages adopt coaxial structure. It is extremely compact. The profiles of the regenerators and reservoirs in the second stage are generally significantly smaller than those in the first stage. The HFPTC is similar to the single stage PTC in the terms of profiles. The weight of the cold tip of the HFPTC is slightly increased. The pulse tube is placed under the compressor. The drilled heat exchanger is adopted to connect the compressor and regenerator 1.The double-inlet, inertance tube and gas reservoir at ambient temperature were adopted as the phase shifter for the first stage. The cold double-inlet, cold inertance tube and cold gas reservoir were adopted as the phase shifter for the second stage. The inertance tube and reservoir of the second stage were installed on the cold end of the first stage.

Fig. 1. The schematic of two-stage gas-coupled pulse tube cryocooler

The detailed design parameters of the developed cryocooler are summarized in Table 1. A simulation model was developed based on SAGE 10.0. During the simulation, the operating parameters are kept the same. The position stroke is 6.8mm. The frequency is 30 Hz. The charging pressure is 2.3 MPa. The following are the assumptions

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of the model. The infrared radiation among all components was ignored. The thermal resistance between the cold end heat exchanger and the regenerator was ignored. The thermal resistance among inertance tube, reservoir and cold end of the first stage was ignored. The operating parameters were kept the same as the previous experiments. Table 1. Parameters of HFPTC cold tip Parameters

Values (mm)

First section of regenerator I

26 * 25(350# SS) + 26 * 30(400# SS)

Second section of regenerator I

18 * 36(500# SS)

Regenerator II

13 * 44(635# SS + Er3 Ni 70 µm)

Pulse tube I

11.7 * 104

Pulse tube II

5.5 * 44

Gas reservoir

600 cc (I) + 30 cc (II)

Inertance tube I (IT1)

2 mm + 3 mm(L3m) + 4 mm(L4m)

Inertance tube II (IT2)

1 mm + 2 mm(L2m)

3 The Influence of Inertance Tube on the Cryocooler Together with the reservoir, the inertance tube is the main phase shifter component of the PTC. The interaction between the two stages of the phase distribution was calculated. The operating parameters were kept constant except for inertance tube length. Two sets of simulations were performed. Variations in cooling performance between 0.7 m and 1.5 m were calculated for the length of the inertance tube in the second stage, while the length of the first stage was kept at 0.6 m. Variations in cooling performance were calculated between 0.2 m and 1.4 m in the length of the first stage inertance tube, while the length of the second stage was kept constant at 0.9 m. The no-load temperature of

a) The length of IT1

b) The length of the IT2

Fig. 2. The effect of inertance tube length on no-load temperature

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each set for the two stages is shown in Fig. 2. The results indicate that both set simulation working within the optimal range, as both set obtain the lowest temperature in the second stage. The ratio of gas flow into the second stage is shown in Fig. 3. The results indicated that the ratio obviously increases with the increasing the length of the first stage. However, the ratio change is not sensitive to the change of the second stage.

a) The effect of the IT1

b) The effect of the IT2

Fig. 3. The effect of inertance tube length on the gas flow distribution

The effect of the length of inertance tube on the phase angle distribution in the regenerator is shown in Fig. 4. The phase angle at the inlet and outlet of regenerator for each stage was calculated. For the change in the length of the first stage, it is shown that the phase angle decreases significantly in the first stage, while it changes only slightly in the second stage. Regarding the change in the length of the second stage, it is shown that the phase angle does not change much in the first stage, while it decreases significantly in the second phase. The results show that the change of inertance tube of the first stage has a smaller effect on the second stage. However, the effect of the phase angle of the first stage due to the second stage can be neglected.

4 The Influence of Double-Inlet Opening on the Cryocooler The double-inlet value is the other main phase shifter component of the PTC. The influence of opening (equivalent diameter of the hole) of double-inlet value for the two stages was calculated. Two sets of simulations were performed. The influence of opening of double-inlet on no-load temperature is shown in Fig. 5. It indicates that it has different effect on the temperature of the first stage. The influence of opening of the double-inlet on the output PV power of the compressor is shown in Fig. 6. The PV power decreases with the increasing opening of the first stage. The PV power and pressure ratio has no evident change with the increasing opening of the second stage. The influence of opening of the double-inlet on the ratio of the gas flow rate is shown in Fig. 7. With the increasing of double-inlet opening, it shows the completely opposite results. And it has a strong effect on the ratio of the gas flow to the second stage. The effect of the double-inlet opening on the phase angle distribution in the regenerator is shown in Fig. 8. The results show similar conclusions as the inertance tube is varied.

Numerical Simulation of Two-Stage Gas Coupled High Frequency

a) The effect on REG 1 of the IT 1

b) The effect on REG 2 of the IT 1

a) The effect on REG I of the IT 2

b) The effect on REG II of IT 2

Fig. 4. Effect of the length of inertance tube on phase angle distribution in regenerator

a) The effect of the DI 1

b) The effect of the DI 2

Fig. 5. Effect of double-inlet opening of first stage on the temperature of the two stages

a) The effect of the DI 1

b) The effect of the DI 2

Fig. 6. Effect of double-inlet opening of first stage on the PV power

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a) The effect of the DI 1

b) The effect of the DI 2

Fig. 7. Effect of double-inlet opening of second stage on gas distribution along the regenerator

a) The effect on REG I of the DI 1

a) The effect on REG I of the DI 2

b) The effect on REG II of the DI 1

b) The effect on REG II of the DI 2

Fig. 8. Effect of double-inlet opening on phase angle distribution in regenerator

Figure 5, 6, 7 and 8 reveal the no-load temperature, gas volume distribution, phase angle distribution relationship between the two stages. With the increasing of DI 1, the temperature of the first stage decrease. Thus, the temperature of second stage decreases. However, the temperature increases as the gas volume and the phase angle distribution change significantly. With the increasing of DI 2, the temperature of the second stage decreases and then increases as the lower temperature of the first stage and less gas flow rate.

5 Conclusion A two-stage gas-coupled high frequency pulse tube cryocooler has been designed. Simulation models were developed to investigate the effect of inertance tubes and doubleinlet values on the cryocooler performance. The interaction of the phase angle and gas

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flow distribution between the two stages was calculated. The main conclusions were as follows. • The inertance tube of the second stage has little influence on phase angle and gas flow distribution on the first stage. At the same time, the apparent effects of phase angle and gas flow distribution in the second stage are caused by the first stage. • The output PV power of compressor and gas flow distribution were affected by the first stage double-inlet obviously. The double-inlet of the second stage has little influence on the output PV power and phase distribution of the first stage. However, it has an effect on the distribution of gas flow and the phase angle of the second stage. • The relationship between the two gas-coupled stages is complex. We should organized both of the two stage working in proper condition. During the experiment, we usually adopt the control variable method, which does not change the parameters of the two stages at the same time. It has less influence to change the parameters of the second stage. In order to improve the efficiency of experiments, we could design the second stage firstly.

References 1. Xiaotao, W., et al.: A high efficiency hybrid stirling-pulse tube cryocooler. AIP Adv. 5(3) (2015) 2. Xianlin, W., et al.: High efficiency 40 K single-stage Stirling-type pulse tube cryocooler. IOP Conf. Ser. Mater. Sci. Eng. 278 (2017) 3. Xiaoqin, Z., et al.: A three-stage Stirling pulse tube cryocooler reached 4.26 K with He-4 working fluid. Cryogenics 58, 93–6 (2013) 4. Liubiao, C., et al.: Numerical and experimental study on the characteristics of 4 K gas-coupled Stirling-type pulse tube cryocooler. Int. J. Refrig. 88, 204–210 (2018) 5. Xiaoshuang, Z.: Experimental research on a 12.1 K gas-coupled two-stage high frequency pulse tube cryocooler, IOP Conf. Series: Mater. Sci. Eng. 171 (2017) 6. Xiaomin, P., et al.: An integral hybrid two-stage pulse tube cooler with improved efficiency. Cryogenics (2019) 7. Xianlin, W., et al.: An 80 mW /8 K high-frequency pulse tube refrigerator driven by only one linear compressor. Cryogenics 101, 7–11 (2019)

Study on Regenerator Matrix Optimization of Free Piston Stirling Generator Chunyun Chi1 , Mingqiang Lin1(B) , Ruijie Li1,2 , Kexin Jiao1,2 , Guotong Hong1,2 , and Jian Mou1 1 Key Laboratory of Technology on Space Energy Conversion, Technical Institute of Physics

and ChemistryChinese Academy of Sciences, Beijing 100190, China [email protected] 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. Deep space exploration is an important field in scientific research, and free piston Stirling generator has been identified as an excellent candidate for conversion of nuclear energy to electric power. The regenerator is the significant heat exchange component of the generator, and the specification of the matrix is the vitally important factor in determining its performance. With the usage of Sage software, a thermodynamic model is developed in this paper. Then the influence of specification of stainless wire mesh on the generator’s output performance and stable hot end temperature is analyzed, which is verified by experiment. The results show that the model and experiment show the same trend, although there are deviations. When the wire diameter is constant, the efficiency first increases and then decreases with the increase of porosity. When the porosity is constant, the efficiency first increases and then decreases with the increase of the wire diameter. As porosity and wire diameter increase, the stable hot end temperature decreases gradually. This model plays a guiding role in the regenerator matrix optimization of free piston Stirling generator for the future. Keywords: Free piston · Stirling generator · Regenerator · Wire mesh

1 Introduction Deep space exploration is a significant field in scientific research, and the critical issue is the establishment of the energy system that can support long-duration flights [1]. The energy system requires its thermoelectric conversion device has the advantages of high energy-conversion efficiency, high reliability, and long life, free piston Stirling generator is one of the most promising options [2]. As the core heat exchange component of free piston Stirling generator, the regenerator takes the role of alternating heat absorption and release, which plays a crucial role in the performance of the generator [3]. The regenerator performance is determined by many factors, including heat transfer characteristics, flow resistance characteristics, dead volume, and heat loss, etc., which are partly affected by the matrix of the regenerator. The structure and thermophysical properties of the matrix are the main criteria for the © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 832–839, 2023. https://doi.org/10.1007/978-981-99-6128-3_108

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selection of the matrix. The former affects the dead volume and flow resistance of the regenerator, while the latter mainly affects the heat transfer performance [4]. Andersen et al. [4] developed a new structure of regenerator matrix, which can effectively improve the efficiency of the Stirling generator. The results show that the thermal-electric conversion efficiency of the generator is increased from 32.9% to 33.2%, but the output power is reduced by 3%. When the regenerator substrate is completely uniform, the performance can reach 33.0%. Alfarawi et al. [5] researched indicating the power of the free piston Stirling generator filled with three materials such as stainless steel, Monel 400, and ceramic materials. The results show that the lowest indicating power can be obtained when the copper is used as the regenerator filler, while the cooling power is the highest. Garg et al. [6] studied the difference between homogeneous and nonhomogeneous porous media. The experiment result proved that the non-homogeneous porous media could significantly reduce the heat transfer time. Nam and Jeong [7] developed a parallel braided regenerator. It is found that its flow resistance loss was 20%–30% lower than that of the ordinary braided regenerator, while the axial heat conduction loss was larger, thus its performance was poor. Costa et al. [8] established a CFD model of regenerator by using the porous media model. Simulations are carried out for laminar flow with a porosity between 52% and 72% and Reynolds number between 5 and 50. The result shows that when the porosity is 72%, the friction coefficient and pressure drop of the wire mesh are the largest. Previous studies have focused on the material and structure of the filler of the regenerator of the free piston Stirling generator, or the influence of porosity on flow resistance characteristics. However, the influence of the specification of the wire mesh on the generator’s performance has not been studied in depth. In this paper, a thermodynamic model is established by using the Sage software. Then the influence of specification of stainless wire mesh on the generator’s output performance and stable hot end temperature is analyzed, which is verified by experiment.

2 Establishment of Theoretical Model Sage, a commercial software, is a thermodynamic analysis tool that integrates modeling, design, and optimization. With the usage of this software, the model of the free piston Stirling generator is established. The modules of each component of the generator are selected on the software interface, and relationships among the modules, such as mass, energy, and pressure, are connected according to the actual situation. A theoretical model is established with Sage, which is shown in Fig. 1. The output power and thermoelectric conversion efficiency of free piston Stirling generator can be obtained by inputting the structure parameters into the model. Therefore, this model can be used to optimize the wire mesh of free piston Stirling generator. The prototype developed by the Technical Institute of Physics and Chemistry of the Chinese Academy of Sciences is researched in this paper, and the detailed parameters of the generator are listed in Table 1. During the analysis, except for the parameters as variables, other parameters are consistent with the data in Table 1. Inconel 718 is the hot end material for the prototype and the strength of this material decreases significantly when the hot end temperature exceeds 923 K. Therefore, in addition to studying the

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influence of different specifications of wire mesh on the output performance of the generator, its influence on the stable hot end temperature of the generator needs to be considered in this research.

Fig. 1. Sage model diagram of free piston Stirling generator.

Table 1. Parameters of prototype. Variable

Value

Charge pressure

3.5 MPa

Cold end temperature

283 K

Displacer area

706.9 mm2

Rod area

38.5 mm2

Displacer mass

0.09 kg

Piston mass

0.5 kg

Displacer damping

10 N·s/m

Piston damping

10 N·s/m

Displacer spring stiffness

21.6 N/mm

Piston spring stiffness

14.7 N/mm

3 Simulation Results and Analysis The wire diameter and porosity of the wire mesh affect the heat transfer and resistance characteristics of the fluid in the regenerator, thus affecting the performance of the generator. The smaller the wire diameter, the stronger the heat transfer capability, and

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the smaller the porosity, the higher the flow resistance. Therefore, the wire diameter and porosity should be considered comprehensively. In this paper, the wire diameter and porosity are used as variables to study the influence of wire mesh on the performance of the generator. Figure 2 shows the influence of wire diameter and porosity of the wire mesh on the output power of the free piston Stirling generator. As shown in Fig. 2, when the wire diameter is constant, the output power increases first and then decrease with the increase of porosity. The factor which leads to this phenomenon is that the performance of the regenerator is affected by both flow characteristics and heat transfer characteristics. When the wire diameter is constant, the larger the porosity, the smaller the specific surface area, as well as the flow resistance. The decrease of the specific surface area leads to the decrease of free piston Stirling generator output power, while the decrease of flow resistance has a positive effect on the performance of free piston Stirling generator. Therefore, there is an optimal value of porosity for optimal free piston Stirling generator output performance. When the porosity is constant, the output power increases first and then decrease with the increase of wire diameter. The reason for this phenomenon is the same as that mentioned above. Figure 3 demonstrates the influence of wire diameter and porosity of the wire mesh on the hot end temperature of the free piston Stirling generator. It is found that with the increase of porosity and wire diameter, the stable hot end temperature of the generator decreases. For the prototype developed by the laboratory, the hot end temperature cannot exceed 923 K due to the strength limitation of the materials. Wire mesh with greater porosity and smaller wire diameter can be selected to meet the above condition. In consequence, for the optimization of wire mesh, in addition to considering the influence of wire mesh specifications on generator performance, the influence of wire mesh specifications on hot end temperature should be considered. Output power (W) 100 0.80

120

0.72

0.68

20 40

60

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100

140 145

80 60

140 120 100 80

40 20

0.64

1600 1400

700

120

Porosity

Porosity

600 0.80

140

0.76

0.60

Hot end temperature (K)

145

0.72

1600

1200

800 900 1000 1200 1400

1000 900 800 700 600

0.68

0.64

10

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40

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50

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10

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Fig. 2. Influence of wire diameter and porosity Fig. 3. Influence of wire diameter and on the output power. porosity on the hot end temperature.

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4 Experiment Results and Analysis An experimental system is established to test the output performance of the prototype filled with different specifications of the stainless wire mesh, which is shown in Fig. 4. The experimental test system consists of a heating system, cooling system, excitation system, vacuum and charging system, load test system, and data acquisition system. Several heating rods are used to provide a certain heating power by adjusting the regulated power supply. The heat released by the generator is taken away from the cooler by the low temperature circulating water of the water-cooling unit. The excitation system is connected with the load circuit in parallel. When the temperature of the hot end rises to a certain value, the generator is activated by an AC power source, and then disconnects the excitation system. The aim of the vacuum and charging system is to ensure the purity of the helium in the test system. A K-type thermocouple is calibrated to measure the hot end temperature of the generator. When the fluctuation of temperature is less than 0.01 K in 5 s, it can be considered that the system has reached a steady-state. The electrical power produced by the linear alternator is consumed in the load and can be measured by the power meter. The tested data is stored on the computer by the software Labview [1]. In order to verify the correctness of the simulation results, several commonly used wire mesh specifications are selected in this paper, and the details are shown in Table 2. The free piston Stirling generator prototype filled with these specifications of wire mesh are tested respectively and the results are compared with simulations.

Fig. 4. Layout of the FPSE experimental system.

Figure 5 indicates the output power comparison of the generator with different wire mesh specifications. It is found that when the mesh number is 203, the output power of the free piston Stirling generator is the maximum in both simulation and experiment. The output power in simulation is 140.1 W and that in the experiment is 116.5 W. The

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Table 2. Detailed parameters of wire mesh. No

Mesh Number(mesh)

Wire Diameter(µm)

Porosity

Specific Surface Area(m−1 )

1

155

63

0.6981

19171.1

2

180

30

0.8330

22263.2

3

200

30

0.8145

24737.0

4

203

50

0.6861

25108.0

5

250

40

0.6858

31415.9

simulated output power of the generator filled with 155-mesh wire mesh is 125.5 W, which is the worst simulation result among the four kinds of wire mesh. And in the experiment, its output power is also the smallest in the four cases, which is 113 W. This phenomenon means that the experimental results are in great agreement with the simulation results. However, the difference between the four cases in the experiment is not significant, while that of the simulation is more pronounced. This is because the heat involved in the cycle of the generator filled with different wire mesh changes in the experiment while that heat remains constant in the simulation. In the experiment, the higher the stable hot end temperature, the more heat is leaked to the environment. Although the heating power is constant, the actual heat involved in the cycle is reduced, which results in a smaller increase in output power. Whereas the model does not consider this heat loss, the same amount of heat is involved in the cycle for four cases. Therefore, the difference between the four cases in the experiment are less pronounced compared to the simulation. Incidentally, Table 2 lists the five specifications of the wire mesh, while Fig. 4 shows the experiment results for only four specifications of the wire mesh. This is because when testing the generator filled with 254 mesh wire mesh, with the same operating conditions as the generator filled with other specifications of wire mesh, the hot end temperature keeps rising during the experiment, which is about to exceed the critical safe temperature and fails to reach the stable operation state, so the experiment is terminated. Figure 6 displays the hot end temperature of the generator comparison between simulation and experiment under the above wire mesh specifications. According to Fig. 6, the order of the hot end temperature of the generator filled with four specifications of the wire mesh in simulation is as follows: No.1 < No.2 < No.3 < No.4, and that in the experiment is as follows: No.1 < No.2 < No.3 < No.4. The hot end temperature variation trend is consistent between the simulation and the experiment, which indicates the feasibility of the model. With the usage of the model, the wire mesh with appropriate wire diameter and porosity can be selected to optimize the performance of the free piston Stirling generator. Due to the fixed specification of the wire mesh on the market, and the limited time, only the several fixed specifications of wire mesh are researched in this paper. In the future, the fiber cotton will be used as the generator filler instead of the wire mesh. According to the simulation result of the model, the fiber cotton with suitable

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Hot end temperaure (K)

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140

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Experiment Simulation

840

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1

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No.

Fig. 5. Output power comparison of different wire mesh.

1

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Fig. 6. Hot end temperature comparison of different wire mesh.

wire diameter and porosity can be designed to further optimize the performance of the free piston Stirling generator.

5 Conclusion In this paper, a theoretical model is established with the usage of Sage software, which can be used to optimize the stainless wire mesh of the free piston Stirling generator. The influence of stainless wire mesh specification on the generator’s output performance and stable hot end temperature is analyzed and verified experimentally. The results demonstrate that the trend of simulation is consistent with that of the experiment, which implies that the model can be used to select stainless wire mesh with appropriate wire diameter and porosity to optimize the performance of free piston Stirling generator. In the future, this model may be used to further optimize the performance of the free piston Stirling generator filled with fiber cotton. Acknowledgments. This work was supported by the National Key R&D Program of China (2020YFB1901800).

References 1. Chi, C., Mou, J., Lin, M., et al.: CFD simulation and investigation on the operating mechanism of a beta-type free piston Stirling engine. Appl. Therm. Eng. 166, 11471 (2020) 2. Dai, D., Liu, Z., Long, R., et al.: An irreversible Stirling cycle with temperature difference both in non-isothermal and isochoric processes. Energy 186, 115875 (2019) 3. Nielsen, A.S., York, B.T., Macdonald, B.D.: Stirling engine regenerators: how to attain over 95% regenerator effectiveness with sub-regenerators and thermal mass ratios. Appl. Energy 253, 113557 (2019) 4. Andersen, S.K., Carlsen, H., Thomsen, P.G.: Numerical study on optimal Stirling engine regenerator matrix designs taking into account the effects of matrix temperature oscillations. Energy Conver. Manage. 47(8), 894–908 (2006)

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5. Alfarawi, S., Al-Dadah, R., Mahmoud, S.: Potentiality of new miniature-channels Stirling regenerator. Energy Convers. Manage. 133, 264–274 (2017) 6. Garg, S.K., Premachandran, B., Singh, M., et al.: Effect of Porosity of the regenerator on the performance of a miniature Stirling cryocooler. Therm. Sci. Eng. Progress 15, 100442 (2020) 7. Nam, K., Jeong, S.: Development of parallel wire regenerator for cryocoolers. Cryogenics 46(4), 278–287 (2006) 8. Costa, S.C., Barreno, I., Tutar, M., et al.: The thermal non-equilibrium porous media modelling for CFD study of woven wire matrix of a Stirling regenerator. Energy Convers. Manage. 89, 473–483 (2015)

Design of High Frequency Coaxial Pulse Tube Cryocooler Working at 200 K Yanbo Duan1,2 , Wei Wang3(B) , Bingkun Lv1,2 , Jue Wang4 , Laifeng Li1,2 , and Yuan Zhou1,2 1 CAS Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Beijing 100190, China [email protected] 2 Springer University of Chinese Academy of Sciences, Beijing 100049, China 3 Songshan Lake Materials Laboratory, Dongguan 523808, China [email protected] 4 CASIC Space Engineering Development Co. Ltd, Beijing 100854, China

Abstract. With the intensification of global greenhouse gas effect and the development of cryogenic technology, the high frequency pulse tube cryocooler (PTC) with helium gas as operating medium has attracted more attention due to its advantages of simple structure, low vibration, long life and no moving parts at the cold end. However, the application and research of PTC in the high temperature region is less than that in the liquid nitrogen temperature region. In this paper, the operating temperature of the PTC is raised to about 200 K. A theoretical model is established. Secondly, the effects of the operating frequency and charging pressure on the refrigeration performance are studied. Keywords: Pulse tube cryocooler · Sage · High temperature region · Refrigeration performance · Efficiency

1 Introduction Stirling-type pulse tube cryocoolers can convert electrical energy into cooling capacity at low temperatures and play an important role in energy utilization [1, 2]. As a new type of cooling method, PTC has the advantages of low vibration, simple structure, long life and high stability due to no moving parts at the cold end [3, 4], and it has a wide range of cooling temperature areas. On the other hand, PTC uses the inert gas helium as a working medium, which not only protects the ozone layer, but is also a nontoxic and non-hazardous green working medium, promising in the fields of aerospace, high-temperature superconductivity and quantum chip cooling [5, 6]. In space missions, the cooling temperature requirements of infrared detectors and focal plane arrays have increased to more than 150 K. Due to the limitation of volume and weight, small cryogenic PTC are the most reliable as cold sources and have become a research hotspot. In 2014, Air Liquide developed a cooler for cooling infrared detectors for microsatellite missions. The required cooler temperature is from 150 K to © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 840–847, 2023. https://doi.org/10.1007/978-981-99-6128-3_109

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200 K, and the cooling power is between 1 W−3 W. After optimizing the structure of the regenerator and inertance tube, the cooling performance is improved by 8%−25% [7]. Northrop Grumman Aerospace Systems designs and manufactures a refrigerator directly integrated into the SWIR detector, which is small in size and weight and has a performance of 5 W@150 K when the ambient temperature is 300 K [8]. In 2015, Thales Cryogenics launched a small cryocoler designed to cool detectors at high temperature (HOT), with an operating temperature of 120 K−150 K [9]. In addition, in 2020, the Shanghai Institute of Technical Physics optimized parameters such as regenerator, pulse tube, phase angle, etc., and achieved a cooling power of 50 W@170 K and a Carnot efficiency of 15.8% under 228 W input power [10]. In 2021, Wang Bo et al. of Zhejiang University designed a no-load temperature of 99.8 K, when the input power is 500 W, the cooling performance is 181.3 W@233 K, and the Carnot efficiency is 10.4% [11].

Fig.1. Structural drawing of main structure of PTC

2 Numerical Model The schematic diagram of the cryocooler is shown in Fig. 1, consisting of a linear compressor that provides pressure waves, the cold finger unit mainly includes an aftercooler, a regenerator, and a pulse tube. The phase shifter includes an inertance tube and a gas reservoir. The compressor is mainly composed of a linear motor, a flexible spring and two cylinders. The cylinder pistons on both sides are connected to the linear motor mover (permanent magnet) through connecting rods. The compressor is usually constructed with a double motor reversal and a piston displacement phase angle of 180 degrees, which minimises vibration. The inside of the regenerator is filled with porous media, usually stainless steel wire mesh or magnetic filler. The working fluid reciprocates in the regenerator and exchanges heat with the stainless steel wire mesh in the regenerator. In order to reduce irreversible losses and shuttle losses, there is usually an optimal impedance.

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In this paper, Sage software is used for numerical simulation. Sage can model and simulate the various components of the cryocooler, connecting them through mass flow, pressure waves, and energy flow. At the same time, this software also uses a large number of empirical formulas for the calculation of the cryocooler, which has a strong guiding significance for the actual cryocooler. For the working gas, starting from the integral form of the N-S equation, Sage developers made a series of simplifications and assumptions, and finally obtained the one-dimensional differential form continuity equation, momentum equation and energy equation: ∂ρA ∂ρuA + =0 ∂t ∂x

(1)

∂ρuA ∂uρuA ∂P + + A − FA = 0 ∂t ∂X ∂x

(2)

∂A ∂ ∂ρeA +p + (uρeA + uPA + q) − Qw = 0 ∂t ∂t ∂x

(3)

where ρ is gas density, A is flow area, u is mean-flow velocity, and P is pressure. The viscous pressure gradient term F and the heat transfer term Qw can be expressed as:   f K ρu|u|/2 (4) F =− + dh L Qw = Nu (k/dh )Sx (Tw − T )

(5)

where f is friction factor, K is total local loss coefficient, Nu is a Nusselt number, k is gas conductivity, d h is hydraulic diameter, L is length, S x is the wetted perimeter (wetted surface area per unit length) and T w - T is the temperature difference between the negative z surface and section-average. Table 1. Dimensions of the main components. Main components

Value

Regenerator length

30 mm

Inner diameter of regenerator

56 mm

Outer diameter of regenerator

57 mm

Pulse tube length

54.5 mm

Inner diameter of pulse tube

27.5 mm

Stroke of piston

7.5 mm

Diameter of piston

35 mm

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3 Result and Discussion At a defined temperature, there is an optimum length of regenerator that maximises the cooling efficiency. As the length changes, the temperature corresponding to the maximum efficiency also changes. Within a certain range, the optimum length of the regenerator is proportional to the temperature difference between the cold end and the hot end. Since the temperature difference is around 100 K, which is small compared to liquid nitrogen or liquid helium temperature cooler, a shorter regenerator is required to achieve optimum efficiency at 200 K. For the regenerator, in order to obtain a larger cooling capacity, the L/D of the regenerator is relatively small. However, some other problems will arise, such as the large cross-sectional area resulting in uneven temperature, it is necessary to select the appropriate porous fillers to avoid these problems. The dimensions of the main components in the cryocooler are shown in Table 1.

Fig. 2. Influence of inertance tube size on PTC

Fig. 3. Cooling capacity of the PTC versus the reservior volume

3.1 Effects of Phase Shifter on Refrigeration Performance According to our previous experimental experience, the initial filling pressure is set to 3.0 Mpa, and the frequency is in the range of 30−60 Hz. The structure uses a single-segment

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inertance tube and gas reservoir to adjust the phase between mass flow and pressure waves. Firstly, the influence of the length and inner diameter of the inertance tube on the refrigeration performance of the cooler is calculated according to the known parameters. Figure 2(a) shows the effect of the length and inner diameter of the inertance tube on the performance of the PTC at different operating frequencies. The results show that, within a certain range, as the length of the inertance tube increases, there is a maximum value for each type of inertance tube. Figure 2(b) shows the effect of inertance tube length on acoustic power at different inner diameters. It can be seen that at 3.5 m, the inertance tube with an inner diameter of 8 mm consumes the maximum acoustic power, which corresponds to the maximum cooling capacity in Fig. 2(a). That is to say, the inertance tube at this time can adjust the phase of mass flow and pressure wave to the optimum angle. Figure 3 shows the effect of gas reservior volume on cooling power. The results show that at 500–800 cc, the cooling power is maintained at a high level, and the output in this range will lead to a significant attenuation of the cooling power, which is not conducive to improving the cooling performance of the PTC. In the following calculations, the gas reservior volume is 600 cc. 3.2 Effects of Operating Parameters on Refrigerator Performance In addition to the structural parameters of the PTC, the operating parameters also play an important role in the refrigeration performance of the cooler, such as operating frequency, charging pressure and input acoustic power and other parameters. The model is filled with 250# stainless steel wire mesh, and the phase shifter only uses inertance tube and reservoir.

Fig. 4. (a)Effect of frequency on refrigerator efficiency and capacity (b) Effect of charging pressure on refrigerator efficiency and capacity

As shown in Fig. 4(b), the effect of charging pressure on refrigerator performance is presented. As the charging pressure increases, the total amount of working gas inside the cooler increases, and the flow resistance loss in the regenerator also increases, resulting in a gradual increase in the maximum output power of the compressor. According to this mechanism, the changes in Fig. 4(b) can be analyzed. The cooling capacity increases

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linearly with the increase of the charging pressure, mainly because the gas mass increases and the workmanship increases, which improves the cooling capacity of the PTC, but the COP of the cooler reaches a maximum value at 4.0 MPa. Overall, the increase in charging pressure had little effect on COP, and the variation range was small. The displacement of the compressor piston is 6.125 mm, and the charging pressure is 3.0 Mpa. Figure 4(a) is a graph showing the variation of cooling capacity and Carnot efficiency of the PTC at 200 K under different operating frequency. It can be found that when the operating frequency increases, the cooling capacity first increases and then decreases. The cooling reaches its maximum value at 57 Hz, with a maximum of about 167 W. There is an optimal coupling state between the cold finger and the compressor, and the change of frequency has an impact on the impedance characteristic inside the cooler. When the impedance characteristic of the cold finger deviates from the optimal impedance range, the output power of the compressor will decrease. Higher acoustic power can be obtained by increasing the frequency, mainly because the number of work done by the working fluid gas increases with the increase of the reciprocating frequency of the piston when the displacement of the compressor piston is constant. The increase in acoustic power output will naturally lead to higher cooling capacity, but not necessarily the best COP. The maximum COP is obtained when the frequency is 54 Hz, which is about 0.335.

Fig. 5. (a)Available energy loss at different temperatures (b) Available energy loss distribution in regenerator

3.3 Available Energy Loss Analysis The temperature of the cold end of the refrigerator was set to 200 K, and the internal energy loss of the regenerator was studied, as shown in Fig. 5(a). The main energy losses in the regenerator are flow resistance losses (AEfric), gas and solid irreversible heat transfer losses (AEQw) and axial heat transfer losses (AEQx) of the regenerator material. AEfric, AEQw and AEQx all increased as the temperature decreased from 210 K to 180 K. As the regenerator temperature gradient increases, both AEQw and AEQx increase, but AEQw increases more significantly. Among the losses, the flow resistance loss is still the most important part. It can be predicted from Fig. 5(a) that when the temperature is lower and lower, the proportion of AEQw is larger and larger,

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and the axial heat conduction loss also increases slightly, but it is not the main reason for the loss of the regenerator. This also shows that the flow resistance of the working fluid must be optimized in the later stage to reduce the loss of flow resistance in the regenerator. The optimization methods include filling with lower mesh stainless steel wire mesh and optimizing the inflation pressure. The distribution of losses along the regenerator is given in Fig. 5(b), and each histogram represents the total amount of energy lost from the regenerator inlet to the local location. As the temperature decreases, the viscosity of the working gas decreases, the density increases, and the flow resistance decreases. For the cold end, the irreversible heat exchange loss increases significantly, becoming the main part of the available energy loss, especially in the cold end heat exchanger, this phenomenon is more prominent.

4 Conclusion In this paper, the design and performance of the PTC at room temperature are studied. Based on Sage software, the effects of phase shifter, operating parameters and energy loss in regenerator on refrigeration performance were determined. Taking the maximum cooling power as the target, the optimal size of the inertance tube and the reservoir is determined at the cold end temperature of 200 K. The length of the inertance tube is 3.5 m, the inner diameter is 8 mm, and the maximum cooling power is 166.9 W when the volume of the reservoir is 600 cc. The effects of operating parameters such as frequency and charging pressure on refrigeration performance were studied. The maximum COP and cooling power were 0.335 and 220 W when the frequency was 53 Hz and the charging pressure was 4.0 MPa. This work provides a basis for the future application of PTC in room temperature areas such as refrigerators. Acknowledgments. The project is supported by the National Natural Science Foundation of China (Grant No. 52071223), the Guangdong Basic and Applied Basic Research Foundation (Grant No. 2020B1515120084) and the Key-Area Research and Development Program of Guangdong Province under the grant No. 2020B0101340002.

References 1. Hachem, H., Gheith, R., Aloui, F.: Technological challenges and optimization efforts of the Stirling machine: a review. Energy Convers. Manag. 1365–1387 (2018) 2. Wang, K., Sanders, S.R., Dubey, S.: Stirling cycle engines for recovering low and moderate temperature heat: a review. Renew Sustain Energy Rev. 62, 89−108 (2016) 3. Raab, J., Tward, E.: Northrop Grumman aerospace systems cryocooler overview. Cryogenics 50(9), 572–581 (2010) 4. Radebaugh, R.: Pulse tube cryocoolers for cooling infrared sensors. In: Proceedings of the International Symposium on Optical Science and Technology. International Society for Optics and Photonics, pp. 363–379 (2000) 5. Zhang, A., Wu, Y., Liu, S .: High-efficiency 3 W/40 K single-stage pulse tube cryocooler for space application. Cryogenics 90, 41–46 (2018) 6. Ray, R.: Cryocoolers: the state of the art and recent developments. J. Phys. Condensed Matter Inst. Phys. J. 21(16), 1–9 (2009)

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7. Chassaing, C., Butterworth, J., Aigouy, G.: 150K−200K pulse tube cooler for micro satellites. In: ICC18 18th International Cryocooler Conference (2014) 8. Durand, D., Tward, E., Toma, G.: Efficient high capacity space microcooler. In: ICC18 18th International Cryocooler Conference (2014) 9. Arts, R., Martin, J.Y., Willems, D.: Miniature stirling cryocoolers at Thales cryogenics: qualification results and integration solutions. In: SPIE Defense + Security. International Society for Optics and Photonics (2016) 10. Deng, W., Liu, S., Jiang, Z .: Development of a spaceborne pulse tube cooler operating at 170K. Int. J. Refrig. 115, 1−8 (2020) 11. Wang, B., Chao, Y., Zhao, Q.: A high efficiency stirling-type pulse tube refrigerator for cooling above 200K - ScienceDirect. Energy 215, 119120 (2021)

Research of a High Frequency Reciprocating Piston Gas Bearing Ming-Zhuo Yang1,2 , Ming-Qiang Lin1(B) , Jian Mou1 , Ke-Xin Jiao1,2 , Chun-Yun Chi1 , Rui-Jie Li1 , and Guo-Tong Hong1,2(B) 1 Key Laboratory of Space Energy Conversion Technologies, Technical Institute of Physics and

Chemistry CAS, Beijing 100190, China {yangmingzhuo20,linmingqiang15,jiaokexin19, chichunyun16}@mails.ucas.ac.cn, {jmou,rjLi, gthong}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100190, China

Abstract. The gas bearing uses the method of gas lubrication, which can eliminate the friction between the moving piston and cylinder. It is one of the bottle-neck technology of regenerative refrigerator or heat engine to achieve high reliability, high efficiency and long service life. Aerostatic bearing with orifice restrictors has been widely used in Stirling refrigerator, linear compressor and free-piston Stirling engine successfully. For high frequency reciprocating piston, the load capacity of the gas bearing should be strong enough to prevent contact wear between the piston and cylinder wall. Therefore, to study the affecting factors of gas bearing load capacity, a theoretical model was established considering the flow from the gas bearing orifice and the clearance between the piston and cylinder. The correction coefficient of gas bearing load capacity caused by flow dispersion and non-axial flow effect can be obtained by experiments. Based on the experimental correction coefficient, the load capacity of gas bearing system under various conditions can be solved through independent programming in the model. The results show that the gas bearing system is influenced by clearance size, air supply pressure, number and size of gas bearings orifice. In addition, the model has been verified by the experimental results, which provides a guidance for the reciprocating piston gas bearing design of regenerative refrigerator and heat engine. Keywords: Regenerative Refrigerator · Gas Bearing · Load Capacity

1 Introduction The gas bearing technology uses gas as lubricant which has the advantages of no friction, no pollution and strong adaptability [1]. So far, the gas bearing technology has been widely used in rotary machinery and the technology is relative mature. However, there are few researches on gas bearings applied in linear reciprocating machines and the theory is not mature enough [2]. For high frequency reciprocating piston gas bearing system, aerostatic bearings with orifice restrictors are the most widely used [3]. Reducing the friction of moving parts is one important way to improve the reliability and service life © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 848–854, 2023. https://doi.org/10.1007/978-981-99-6128-3_110

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of regenerative refrigerator. Therefore, the technology of high frequency reciprocating piston gas bearing system is the key technology of regenerative refrigerator to achieve high reliability and long service life [4]. For the high frequency reciprocating piston gas bearing system, the system should provide enough load capacity in order to support the piston and eliminate mechanical friction. Therefore, it is necessary to calculate the load capacity of the system at design stage. At present, there are two research methods for analysis and calculation of gas bearing system. The first method is to directly solve the flow field in the clearance between the piston and the cylinder based on Reynolds lubrication equations. The computational fluid dynamics software such as Ansys Fluent is usually used to solve the equations numerically [5]. However, this process costs lots of time and it is difficult to optimize at the design stage. The second method is to simplify the Reynolds equations. The flow process is divided into the throttling flow process and the one-dimensional flow process in the clearance [6]. For the flow process in the clearance, it is divided into multiple rectangular flow channels along the circumferential direction according to the number of gas bearings. Based on the law of conservation of mass, the pressure in the clearance of each channel can be calculated. And then, the load capacity of the gas bearing system can be obtained. It is assumed that the flow in each channel is independent. In fact, when the gas enters the clearance through the gas bearing, the flow diffusion effect will appear. The circumferential pressure difference in the clearance will lead to the non-axial flow effect. The load capacity calculated by the theoretical model will be several times lager than the actual capacity without considering the flow diffusion effect and the non-axial flow effect. According to the researches of Dudgeon and Shires [7, 8], the correction coefficient of these two effects is only related to the geometric parameters of gas bearing system and has nothing to do with the working conditions. Therefore, in this paper, the correction coefficients of these effects were obtained through experiments to calculate the load capacity of the system.

2 Theoretical Model 2.1 Physical Model The structure of gas bearing system shown in Fig. 1 has been widely used in regenerative refrigerator, linear compressor and free-piston Stirling engine. With the reciprocating movement of the piston, some of the gas from the working chamber goes into the gas supply chamber through the unidirectional valve which makes the pressure of the air supply chamber rise. As a result, the gas in the supply chamber enters the clearance between the cylinder wall and piston through the gas bearings orifices in the chamber wall which can achieve the gas lubrication effect. 2.2 Mathematical Model As is shown in Fig. 2, the gas flow process can be divided into two parts, in the gas bearing system. One is the flow through the gas bearing orifice and the other is the flow through the clearance. In order to simplify the study, the clearance is divided into multiple

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Fig. 1. Structure of high reciprocating piston gas bearing system

rectangular flow channels along the circumferential direction according to the number of gas bearings. Therefore, the flow from the clearance can be regarded as one-dimensional axial flow.

Fig. 2. Flow model of gas bearing system

When the piston is subjected to external force, the piston axis will deviate from the cylinder axis. To describe the degree of deviation, the concept of eccentricity is introduced: ε=

δ R

(1)

δ is the axis distance between the piston and the cylinder; R is the radius difference value between the piston and the cylinder. For the flow through the gas bearing orifice, it is assumed that the flow is onedimensional constant entropy flow. According to the isentropic relationship, the mass flow through the gas bearing orifice can be expressed as: m ˙ =

√ CD π d 2 ρs 2RT 4

 γ γ −1

 

Pd Ps

2 γ

−



Pd Ps

 γ +1 γ

0.5 (2)

Pd is the pressure throttled by the gas bearing; Ps is the supply pressure; d is diameter of the orifice; CD is the flow coefficient; According to the research of the Perry et al. [9],

Research of a High Frequency Reciprocating Piston Gas Bearing

CD was expressed in the form of pressure ratio as follows.  2   CD = 0.85 − 0.15 PPds − 0.1 PPds

851

(3)

For the flow through the clearance, it can be assumed to be Couette flow because the clearance is small enough. Therefore, the analytical solution of the flow field in the clearance can be obtained. The mass flow from the clearance can be expressed as:  2 π Dh3 (4) Pd − Pa 2 m ˙ = 24μnRTl D is the diameter of the piston; Pa is the ambient pressure; n is the number of gas bearings in each row of the piston. μ is the viscosity of helium; Based on the law of conservation of mass, the pressure throttled by the gas bearing can be calculated by combining the Eq. (2)–(4). Then the pressure in the clearance can be calculated. Therefore, the force of gas on the piston in each rectangular flow channel and the load capacity of gas bearing system can be obtained as follows:  (5) Wi = Pd (L − 2l)ai + 2l 21 Pd + 21 Pa ai W = Ct

n

Wi θi CW

(6)

i=1

CW is the load diffusion coefficient which can be obtained through the research of Dudgeon and Lowe. CW

 0.21  0.42  0.0505 Pd Ln = 0.89 πdnD πD Pa



πD nd

0.379 

πD nL

0.758

(7)

Ct is the correction coefficient caused by flow dispersion and non-axial flow effect which can be obtained through experiments. Based on the theoretical model above, the load capacity of gas bearing system under various conditions can be solved through independent programming. The correction coefficient Ct is calculated by Eq. (7) with the measured pressure data from the experiments, and then the load capacity calculated by Eq. (6).

3 Experimental Process As is shown in Fig. 3, the gas bearing system has two rows of orifices along axis. In order to apply a uniform load to the piston without generating torque, a new load device has been designed. The piston was attached to the load device and they had one degree of freedom of axial motion. During the experimental process, the load weights were hung on the suspended rod and the 0.2 kg weight was used to pull the piston through a nylon rope which is attached to a fixed pulley. And then, the valve on the gas supply line was opened to supply the helium to the supply chamber. The pressure of the gas supply chamber was obtained through the pressure sensor. The data of the pressure sensor was not recorded until the piston was just moving relative to the cylinder which is regarded as the minimum supply pressure of the system for supporting the load weight. It can be understood that the maximum of load capacity provided by the gas bearing system equals to the gravity of the piston and the weights.

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The sliding of the piston means that the system provides enough load capacity to balance the gravity of the piston and weights at the current gas supply pressure. Otherwise, the maximum static force of the piston is more than 2N, that means the load capacity provided by the system under current pressure is not enough to make the piston float and slide.

Fig. 3. Schematic diagram of experimental device and experimental platform

4 Results and Discussion Figure 4–1 shows the load capacity of the gas bearing system of different orifice size. The experimental results of three kinds of orifices were compared. In the case of 12 orifices on each row and two rows of orifices, the gas bearing system can provide greater capacity with lager size of orifice under the same supply pressure. Figure 4–2, Fig. 4–3 and Fig. 4–4 compare the experimental and theoretical results of different conditions. It can be seen that the calculated results of the corrected theoretical model are in good agreement with the experimental data and have the same variation trend. Therefore, it is reasonable to obtain the correction coefficient caused by flow dispersion and non-axial flow effect by experiments. By comparing theoretical and experimental data, the accuracy of the theoretical model was verified. By comparing Fig. 4–2 and Fig. 3, reducing the number of orifices in each row will reduce the capacity of gas bearing system under the same conditions. Figure 5 shows the load capacity at different clearances calculated by the theoretical model. With the increase of eccentricity, the load capacity provided by the gas bearing system increases. By comparing the results under different clearances, the gas bearing system with smaller clearance can provide lager load capacity under the same conditions.

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Fig. 4. Results of theoretical model and experiments

Fig. 5. Load capacity under different clearances

5 Conclusion In this paper, a theoretical model was established to analyze the load capacity of gas bearing system through independent programming. And then, the experiments were carried on to obtain the correction coefficient. By comparing the experimental and theoretical results, the accuracy of the theoretical model was verified. Based on the results from both experiments and theoretical, the following conclusions can be made: (1) It is reasonable to obtain the correction coefficient caused by flow dispersion and non-axial flow effect by experiments. (2) The gas bearing system can provide lager load capacity with higher supply pressure, lager diameter and more number of orifices of each row under the same conditions.

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(3) Under the same supply pressure, the load capacity provided by the gas bearing system increased with the increase of eccentricity and the decrease of the clearance. Acknowledgments. The project was financially supported by National Key R&D Program of China (No.2020YFB1901800).

References 1. Chen Xuemei, F.: Gas bearing technology and its application. Lubr. Eng. 04, 61–63 (2000) 2. Li Shusen, F.: Application and development trend of gas lubricated bearing technology. Lubr. Eng. 02, 9–10 (1999) 3. Liu Jing, F.: Performance Calculation and Design of Gas Bearings for Free-Piston Stirling Engine. China Ship Research Institute (2004) 4. Zhai WenYing, F.: Overview of gas bearing technology application status in space Stirling cryocooler and convertor. Cryogenics 42(05), 22–27 (2014) 5. Liu XianXian, F.: Research on characteristics of clearance seal coupled gas bearing under oscillating flow. J. Eng. Thermophys. 42(04), 828–832 (2021) 6. Jia ZiLong, F.: Research of design method of gas bearing of free-piston stirling generator. J. Eng. Thermophys. 42(06), 1396–1401 (2021) 7. Dudgeon, E.H.F., Lowe R.G.S.: A Theoretical Analysis of Hydrostatic Gas Journal Bearings. National Research Council, Canada (1965) 8. Shires, G.L.F., Pantall, S.: Aerostatic jacking of a vented aerodynamic journal bearing. Lubric. Wear Convent. 6(6), 87–96 (1964) 9. Perry, J.A.F.: Critical flow through sharp-edged orifices. ASME 71, 757–764 (1949)

Effect of Flexible Thermal Connections on Temperature Oscillation in the Low-Temperature Range of GM Cryocooler Shanshan Wu1,2 , Jue Wang3 , Hengcheng Zhang1(B) , Huiming Liu1 , Fuzhi Shen1 , Tiantian Xiao1,2 , Chuanjun Huang1 , Laifeng Li1,2(B) , and Yuan Zhou1 1 State Key Laboratory of Technologies in Space Cryogenic Propellants, Technical Institute of

Physics and Chemistry, Chinese Academy of Sciences, Beijing 100190, People’s Republic of China {Zhanghengcheng,laifengli}@mail.ipc.ac.cn 2 University of Chinese Academy of Science, Beijing 100049, People’s Republic of China 3 China Aerospace Science and Industry Space Engineering Development Co. Ltd., Beijing 100854, People’s Republic of China

Abstract. Copper flexible connecting structures are commonly used as flexible thermal transfer components at low temperatures. Gifford-McMahon (GM) cryocoolers are widely applied in cryogenic applications due to their wide temperature range coverage and ease of operation. However, the periodic motion of the displacer in a GM cryocooler generates a significant temperature oscillation below 20 K, approximately 300 mK, which is challenging to meet precise measurement requirements. The copper flexible connection has an attenuating effect on temperature oscillations due to the thermal conductivity and interfacial thermal resistance of the flexible connection. Therefore, the theoretical analysis and experiments were conducted to investigate the effect of copper flexible connection on temperature oscillation in the low-temperature range (4−20 K) of GM cryocoolers. Firstly, a COMSOL simulation model was established based on the one-dimensional heat transfer equation to investigate the major factors that reduce temperature oscillations; Secondly, the temperature fluctuations of the secondary cold head of GM chillers in the temperature range of 4−20 K were measured; Finally, the influence of the key parameters of the copper connection on temperature oscillations was experimentally verified on this basis. A base data is provided for the application of the thermal resistance method to suppress the temperature oscillation of the cold head in the lower temperature range. Keywords: Temperature oscillation · GM cryocooler · Copper flexible connection · Low temperature

© Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 855–861, 2023. https://doi.org/10.1007/978-981-99-6128-3_111

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1 Introduction Low-temperature flexible heat transfer connections are typical cold transfer medium, usually used where there is a certain distance between the cold source and the cooling object, or where the cooling object is sensitive to vibration requirements. Flexible heat transfer connections are not only efficient in heat transfer but also sufficiently flexible. At present, copper wire with high thermal conductivity is used to realize thermal connections in most cases. In the field of cryogenic research and engineering applications, accurate and stable cryogenic environment is a guarantee for the reliable progress of experimental research, such as cryogenic superconductivity research, cryogenic material property testing, cryogenic medical research and high-precision temperature calibration [1, 2]. In recent years, cryocoolers have been widely used because of their wide coverage of temperature zones, simple operation and wide application environment. However, whether they are pulse tube cryocoolers, Stirling cryocoolers or Gifford McMahon (GM) cryocoolers, they are all regenerative cryocoolers, and their working processes involve the cyclic alternating flow of the working mass [3]. The working mechanism results in constant cyclical temperature oscillations at the cold head of the cryocooler. The amplitude of the oscillations varies with temperature from 4.2 to 20 K, with the maximum temperature oscillations reaching over 180 mK. The temperature oscillations are transmitted directly to the thermostatic components connected to the cold head, which is unacceptable for most experiments where high temperature accuracy is required. Numerous studies have been carried out to suppress the temperature fluctuation of the cold head. Hasegawa et al. installed G10 material between the second-stage cold head and the copper block of a GM cryocooler to reduce the temperature fluctuation of the cold head from 200 mK to 5 mK at 4.2 K [4]. Dong et al. added polytetrafluoroethylene (PTFE) material between the cold head of the GM cryocooler and copper block, resulting in a copper block temperature fluctuation of less than 4 mK at 20 K [5]. Maezawa et al. suppressed temperature fluctuation transfer by installing a thermal resistance layer, ultimately achieving temperature fluctuations within ±1.6 mK at 4.2 K [6]. However, there is still a lack of research on copper flexible connections as a widely used heat transfer structure in the cryogenic field. In this paper, the effect of copper flexible connections on temperature oscillations in the low temperature range (4−20 K) of the GM cryocooler was investigated through theoretical analysis and experiments. Firstly, the one-dimensional heat transfer equation of the copper flexible connection was analyzed theoretically, on the basis of which a COMSOL simulation model was established. Secondly, the temperature oscillation of the second-stage cold head of the GM cryocooler in the temperature range from 4 to 20 K was measured. Finally, the COMSOL model was used to predict the temperature oscillation of the copper flexible connection. A base data is provided for suppressing the temperature oscillation of the cold head in the cryogenic range.

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2 Model 2.1 Theoretical Model Numerous studies have shown that temperature oscillations exist in cryocoolers at low temperatures, and the temperature oscillations can be suppressed by the thermal resistance method. Illustrations can be performed from the following theoretical models. Assume that the amplitude of the cold head temperature oscillation is T 0 , the average temperature is T m , and the oscillation frequency is f , so the temperature fluctuation function of the cold head with time is simplified as T = T0 sin(2π ft) + Tm

(1)

The thickness of the thermal damping material is assumed to be x, the cross-sectional area is A, the thermal conductivity is λ and the thermal diffusivity is a. The temperature fluctuations in the cold head are attenuated by the thermal damping material. The internal heat flow equation for the thermal damping material is ∂T q = (2) ∂x λ Given that the thermal properties of a thermally damped material are constant, the temperature propagation satisfies the one-dimensional energy equation. ∂ 2T ∂T =a 2 (3) ∂x ∂x Combining Eqs. (1) and (2), the suppressed temperature function can be found as    πf x + Tm1 (4) T (x) = Tx sin 2π ft − a  where Tx = T0 exp(− πaf x), Tm1 = Tm + λq ax . Therefore, the attenuation ratio AR of the thermally damped material is  πf Tx x) (5) = exp(− ARx = T0 a 2.2 Simulation Model Based on the theoretical model, the COMSOL software was applied to build the simulation model. First, as shown in Fig. 1 (a), the copper flexible connection was simplified to a regular cylindrical, suitable material was selected and the thermal physical parameters of the material were set. After setting the initial conditions, the cold end of the cylinder was set as T in Eq. (1), which is a time-dependent sine function. The free end of the cylinder was given a heat flux of 0. Table 1 shows the amplitude of temperature fluctuations at the free end of the copper connection for different grid numbers. The grid number of the model in this paper is 2356, which satisfies the grid independence. The transient study was set up to carry out the calculations. To ensure accurate temperature readings, the time step was set to 0.03125 s. The temperature oscillations of the copper flexible connection at different lengths are shown in Fig. 1(b).

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Table 1. Amplitude of temperature fluctuations calculated with different number of grids Number of grids

156

223

2356

20376

Amplitude of temperature fluctuation of free end/ K

0.000439

0.0085

0.0602

0.0602

Fig. 1. (a) Cu connection model based on COMSOL software; (b) Variation of temperature oscillations of Cu connection at different locations.

3 Experimental Setup On the basis of the above model, temperature oscillations of the GM cryocooler were observed with the device in Fig. 2. As can be seen in Fig. 2(a), the experimental setup consists of the cryocooler, the vacuum chamber, the radiation shield and the copper connection measuring components. The GM cryocooler (Sumitomo RDK-415D) provides 1.5 W of cooling capacity at 4.2 K. The flange of the cryocooler was bolted to the vacuum chamber and sealed with O-rings to maintain the vacuum. The first-stage cold head of the cryocooler was bolted to the radiation shield. The measurement structure of the copper flexible connection at the cold head is shown in Fig. 2(b). The copper flexible connection was cooled by the second-stage cold head. The second-stage cold head was bolted to one end of the copper connection. In order to monitor temperature fluctuations, two rhodium-iron thermometers were placed in the second-stage cold head and at the other end (free end) of the copper connection. A heating resistor (Caddock MP9100–20.0–1%) was attached to the cold head with bolts to control the temperature of cold head. To decrease thermal contact resistance, Apiezon N grease and indium were used between the flange and the radiation shield, and between the joints. Since the test data from the cryocooler manufacturer showed that the main frequency of vibration in the cold head of cryocooler was 1 Hz [7], high frequency acquisition of temperature data is required. In the present experiment, a Keithley 2700 multimeter with kickstart software was adopted for acquisition at frequencies up to 14 Hz. In order to investigate the suppression of temperature fluctuations by copper connection dimensions, temperature measurements were carried out on three sizes of copper flexible connections and their dimensions are shown in Table 2.

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Fig. 2. (a) Experimental setup for measuring the effect of copper connections on temperature fluctuations in the cold head of a GM cryocooler; (b) Partial enlargement of the measured components and the physical picture.

Table 2. Dimensions for three copper connections Sample

Diameter/mm

Length/mm

Number

Cu connection 1

2.4

100

1

Cu connection 2

2.4

300

1

Cu connection 3

2.4

300

3

4 Results and Discussion Figure 3(a) shows the temperature oscillation of the cold head and the three copper connections when the cold head has a cooling capacity of 1.5 W. At this point, the average temperature of the cold head was 4.99 K. It can be seen that the temperature waveform of the cold head resembles a sine curve, with a main frequency of 1 Hz of temperature fluctuations, consistent with the period of the displacer in the cryocooler. The amplitude of temperature fluctuation T 0 of the cold head is approximately 182.5 mK when the heating capacity is 1.5 W. The temperature fluctuations amplitude T1, T2 and T3 of copper connection 1, copper connection 2 and copper connection 3 are 97 mK, 37 mK and 74 mK respectively, with attenuation ratios AR1 , AR2 and AR3 of 0.531, 0.203 and 0.406 respectively. It was found that, in addition to differences in the amplitude of temperature fluctuations, the components differed in their fluctuation phase. The temperature fluctuation of the cold head temperature in the range of 4 to 20 K for different heating applied to the cold head is shown in Fig. 3(b). As can be observed, the temperature fluctuation in the cold head increases firstly and then decreases, with a fluctuation amplitude of up to 244 mK at 9.78 K. The three copper connections exhibited different suppression of temperature fluctuations. Among them, copper connection 1 showed the worst suppression, while copper connection 2, with its longer length of thermal conduction, showed the best suppression. Comparing the experimental results for copper connection 2 and copper connection 3, it is clear that copper connection 3 is less effective in suppressing fluctuations due to its higher number of copper wires, which makes the soldering area less thermally resistant. In combination with the theoretical

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model, it is evident that the AR is not only related to the thermal diffusion coefficient a of the thermally damped material, but also to the length of thermal conduction x. Both smaller a and larger x result in well-suppressed temperature fluctuations at the free end of the copper connection.

Fig. 3 (a) Temperature oscillations of the cold head and the three copper connections when the cold head is heated to 1.5 W; (b) Temperature fluctuations of the cold head and the three copper connections at different temperature.

Fig. 4. Temperature fluctuations of a 2.4 mm diameter copper connection.

Figure 4 shows the range of temperature fluctuations versus the length of the copper connection material in the simulation model built by the COMSOL software. The model provides a prediction of the suppression of temperature fluctuations at different lengths of the copper connection when the temperature of GM cryocooler was 4.99 K. It can be found that for a copper connection with a diameter of 2.4 mm, it takes 1824 mm, 192 mm and 96 mm to achieve an AR of 0, 0.25 and 0.5 respectively.

5 Conclusion In the present paper, the effect of copper flexible connections on the suppression of temperature fluctuations in the 4 to 20 K temperature range of GM cryocooler was investigated. Theoretical models were combined with experimental results to analyze and predict the suppression of temperature fluctuations by copper connections. The main conclusions are as follows.

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1. Theoretical analysis revealed that the main factors influencing the attenuation ratio (AR) are the thermal diffusion coefficient a and the length of thermal conduction x. Smaller a and larger x both provide better attenuation of temperature fluctuations. 2. The temperature fluctuation of the GM cryocooler increases and then decreases from 4 to 20 K, reaching a maximum amplitude of 244 mK at 9.78 K. 3. The smaller the thermal cross-sectional area and the longer the thermal conduction length of the copper connection, the better the suppression of temperature fluctuations. The simulation model predicted that at the temperature of the cold head of 4.99 K, a single 2.4 mm diameter copper connection would completely suppress the temperature fluctuations in the copper connection at a length of 1824 mm. Acknowledgments. The project was supported by the National Natural Science Foundation of China (Grant No.: 52007186 & 51877209), Research fund of State Key Laboratory of Technologies in Space Cryogenic Propellants (Fund No.: SKLTSCP202001), the Key Research Program of the Chinese Academy of Sciences, Grant No. ZDRW-CN-2021–4-1.

References 1. Ohtani, Y., et al.: Sub-cooled nitrogen cryostat for 66 kV/750A superconducting fault current limiter magnet. In: AIP Conference Proceedings, vol. 710, pp. 867-876 (2004) 2. Brian, I., et al.: A cryostat and temperature control system optimized for measuring relaxations of glass-forming liquids. Rev. Sci. Instrum. 79, 045105 (2008) 3. Gifford, W.E.: The Gifford-McMahon cycle. Adv. Cryog. Eng. 11, 152–159 (1966) 4. Yasuhiro, H., Daiki, N., Masayuki, M., Hiroya, Y., Takashi, K.: High-precision temperature control and stabilization using a cryocooler. Rev. Sci. Instrum. 81, 094901 (2010) 5. Dong, B., et al.: A new cryostat for precise temperature control. In: AIP Conference Proceedings, vol. 1552, pp. 825–829 (2013) 6. Maezawa, H., et al.: Stability of a quasi-optical superconducting NbTiN hot-electron bolometer mixer at 1.5 THz frequency band. IEEE Trans. Appl. Superconduct. 21(3), 640–644 (2011) 7. Takayuki, T., et al.: Vibration analysis of cryocoolers. Cryogenics 44(5), 309–317 (2004)

Molecular Dynamics Simulation of the Phase Shift Mechanism on Inertance Pulse Tube Refrigerator Longyu Yang1,2 , Chen Zheng2 , and Zheng Cui1,2(B) 1 Institute of Thermal Science and Technology, Shandong University, Jinan, China

[email protected], [email protected]

2 Thermal Science Research Center, Shandong Institute of Advanced Technology, Jinan, China

[email protected]

Abstract. At present, it has become a hot spot to study the thermodynamic mechanism of the pulse tube from the perspective of microscopic particles, but the explanation of the phase shift mechanism on inertance pulse tube is not clear enough. In this paper, the inertance pulse tube and the basic pulse tube model were constructed by the molecular dynamics simulation method. A reflect wall moving at a sinusoidal velocity was used as the driving piston. The instantaneous temperature, pressure, mass flow and particle velocity of the pulse tube were statistically analyzed. The simulation results show that, due to the existence of the inertance tube and the gas reservoir, the inertance pulse tube used the inertia effect of the oscillating airflow to adjust the phase difference between the pressure wave and the velocity wave, making it smaller than that in the basic pulse tube and the cooling efficiency of the pulse tube refrigerator was improved. At the same time, the amplitude of the temperature of the inertance pulse tube was lower. The results of the current work were helpful to deeply understand the phase shift mechanism on inertance pulse tube from the microscopic perspective. Keywords: Molecular Dynamics Simulation · Pulse Tube Refrigerator · Inertance Tube · Phase Shift Mechanism

1 Introduction The pulse tube refrigerator has no moving parts at low temperature, so it has a simple structure and reliable operation [1–3]. Due to the lack of phase modulation components at the hot end of the basic pulse tube refrigerator, the cooling efficiency is low [4]. The pulse tube refrigerator with various phase modulation mechanisms at the hot end have been widely studied and applied. At present, the most common phase-modulated pulse tube refrigerator is the inertance pulse tube refrigerator, while the methods used to study the inertance pulse tube refrigerator are mainly experiment, theoretical analysis and numerical simulation. Liu et al. [5] found that the volume of the gas reservoir and the surface roughness of the inertance © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 862–869, 2023. https://doi.org/10.1007/978-981-99-6128-3_112

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tube have a certain influence on the cooling efficiency of the refrigerator by experiment. Thermoacoustic theory and circuit analogy models are commonly used research methods for theoretical analysis. Luo et al. [6] presented a laminar-flow thermoacoustic transmission-line model and a turbulent-flow thermoacoustic model to theoretically analyze the phase modulation mechanism of the inertance tube. The results show that the turbulent-flow thermoacoustic model is more accurate. Roach and Kashani [7] studied the advantages obtained by replacing the orifice of a small-hole pulse tube refrigerator with an inertance tube by building a circuit analogy model. A lot of scholars have also carried out numerical simulation research. Based on the Computational fluid dynamics (CFD), Rout et al. [8] modeled an inertance pulse tube refrigerator, and the dimensions of pulse tube and regenerator were optimized by Response Surface Methodology and Non-Sorted Genetic Algorithm II. Wang et al. [9] used the SAGE software to optimize the regenerator length, outer diameter and frequency of an inertance pulse tube refrigerator with an operating temperature of 60 K, and analyzed the effect of the regenerator length on the cooling efficiency. Radebaugh et al. [10] analyzed a transmission line model of the inertance tube with REGEN3.2 software, calculated the phase modulation performance of the inertance tube and compared it with the experimental results, and found that under certain conditions, the two can be well matched. Liu et al. [11] used DeltaEC to propose a coupling model of an inertance pulse tube refrigerator and studied the characteristics of phase shift and energy flow. The results show that the phase shift performance of the refrigerator decreases with the increase of the number of coiling turns. The various numerical simulation software mentioned above all study the performance of the inertance pulse tube refrigerator from a macro perspective, and cannot accurately describe the internal state and properties of the working substance inside the refrigerator, and there is a certain difference with the actual working substance. In contrast, the molecular dynamics simulation method provides a powerful tool for simulating micro-dynamic behavior oriented to the atomic level. At present, the molecular dynamics simulation method [12, 13] has been gradually used to study the alternating oscillation process of the working substance inside the pulse tube refrigerator from a microscopic perspective, but the explanation of the phase shift mechanism on inertance pulse tube is not clear enough. Therefore, in this paper, the inertance pulse tube and the basic pulse tube model were constructed by the molecular dynamics simulation method. The instantaneous temperature, pressure, mass flow and particle velocity of the pulse tube were statistically analyzed. The phase modulation mechanism of the inertance pulse tube refrigerator is investigated from the atomic level by using the molecular dynamics simulation method.

2 Simulation System and Method 2.1 Simulation Model The two types of the pulse tube model are shown in Fig. 1. Since the temperature control command is used to equate to the pre-cooling of the regenerator, the regenerator is removed from the model and the simulation time is shortened. The initial simulation model of the inertance pulse tube refrigerator consists of a piston cylinder, a pulse tube,

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an inertance tube and a gas reservoir. While the pulse tube refrigerator only consists of a piston cylinder and a pulse tube. The model size is scaled according to the real refrigerator, and the parameters are shown in Table 1. The initial filling pressure and temperature of He are 5 bar and 287 K respectively, which are the same as the actual situation. The period of the piston is 2000 ps, although the average speed of the piston is higher than that of the actual piston in this case, it is reasonable considering that there is no friction loss in the simulation [13]. The fully reflective wall is set in the x and z directions. The periodic boundary condition is set in the y direction to simulate infinite volume. The red wall on the left side of the model is the compression piston surface in Fig. 1. The relationship between the position of the compression piston surface and the simulation time is expressed by the following equation. X = A[1 − cos ω(t − t0 )dt]

(1)

where X is the position of the compression piston surface; A is the stroke of the piston, which is 143.03 nm; ω is the angular velocity; t is the simulation time; t 0 is the initial time; dt is the time step of the simulation, which is 0.0008 ps. 2.2 Simulation Method In this paper, the interaction between the atom in the simulation is described by the standard 12–6 Lennard-Jones (L-J) potential energy function and the parameters are shown in Table 2. The simulation process is mainly divided into two steps. The NVT ensemble is used to carry out the equilibrium dynamics simulation process in the first step, running for 80 ps. At this stage, the global temperature is controlled to 287 K. Next, the NVE ensemble is used to simulate the actual alternating oscillation process of the gas working substance, running for 20000 ps, which is ten piston periods. The software LAMMPS is used to complete the described simulation process, and output the pressure, temperature, mass flow and velocity parameters [14].

Fig. 1. The simulation models of the pulse tube refrigerator.

3 Results and Discussion In order to obtain the variation of the temperature with time inside the pulse tube, a micro grid is divided every 100 nm along the x direction of the pulse tube, and then the instantaneous temperature of each grid is calculated. Figure 2 shows the temperature

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Table 1. The size parameter settings of the model. part

x/nm

y/nm

z/nm

Piston Cylinder

143.03

0.57212

100

Pulse Tube

858.18

0.57212

100

Inertance Tube

429.09

0.57212

20

Gas Reservoir

429.09

0.57212

100

Table 2. The values of the energy parameter and the size parameter. pair

ε/eV

σ /nm

He-He

0.000607098

2.103

Fe-Fe

0.5261

2.3005

He-Fe

0.0178726

2.20175

variation with time of the cold and hot end of the two types of the pulse tube. It can be seen that the temperature curves are similar to the sinusoidal function curve, and the period is similar to the piston period. Besides, the temperature amplitude of the cold and the hot end of the inertance pulse tube is smaller than that of the basic pulse tube.

Fig. 2. The temperature at the cold and hot end of pulse tube. (a) The basic pulse tube; (b) The inertance pulse tube.

The pressure at the cold and the hot end of the two types of the pulse tube in the range of 14,000 ps to 20,000 ps are shown in Fig. 3. The pressure curves are similar to the sinusoidal function curve and the period is similar to the piston period. For the inertance pulse tube, the peak pressure at the cold and hot end is higher than that of the basic pulse tube. The average peak pressure at the cold and hot end of the former are 7.82 bar and 8.31 bar respectively, while that of the latter are 7.59 bar and 8.28 bar

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respectively. The reason is that after the inertance tube and gas reservoir are added, more He enters the pulse tube to participate in cooling, increasing the cooling performance [15]. In addition, the amplitude of the pressure in the pulse tube decreases due to the buffering effect of the inertance tube and the gas reservoir. Figure 4 shows the curve of mass flow with time for each section in the pulse tube. In this paper, the number of He atoms passing through a section at 1 ps is taken as the mass flow of the section and take the positive direction of x axis as positive. The mass flow curve of each section in the pulse tube is also similar to the sinusoidal function curve and the period is similar to the piston period. For two types of the pulse tube, the peak value of the mass flow in the section near the middle of the pulse tube is the largest, while the trough value of the mass flow in the section near the cold end is the largest. The mass flow amplitude is the smallest at the section near the hot end. In addition, the mass flow curve of the section near the cold end in the two types of the pulse tube are basically consistent. While the peak of the curve of the section near the hot end in the

Fig. 3. The pressure at the cold and hot end of pulse tube. (a) The basic pulse tube; (b) The inertance pulse tube.

Fig. 4. The curve of mass flow with time for each section in the pulse tube. (a) The basic pulse tube; (b) The inertance pulse tube.

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inertance pulse tube is smaller than that in the basic pulse tube. This is due to the fact that some He atoms in the inertance tube and the gas reservoir keep going in and out of the pulse tube during the operation of the pulse tube, which leads to the reduction of the peak of the mass flow near the hot end. While the inertance tube and the gas reservoir have less effect on the mass flow near the cold end. Figure 5 shows the phase difference of the pressure and velocity wave at the cold and hot end of the two types of the pulse tube. The pressure wave leads the velocity wave at the cold end, while the velocity wave leads the pressure wave at the hot end. The phase difference at the cold end of the basic pulse tube is close to 90°, which is basically consistent with the enthalpy flow phase modulation theory [16]. According to the enthalpy flow phase modulation theory, the phase difference between the pressure wave and the velocity wave in the basic pulse tube is close to 90°, and it has almost no cooling capacity. After the inertance tube and gas reservoir are added, the phase difference at the cold end becomes close to 72°. In addition, the difference in phase difference at the hot end of the two types of the pulse tube is also close to 18°, which shows that the addition of the inertance tube and gas reservoir does not change the difference in phase difference at the cold and hot end of the two types of the pulse tube. The results are consistent with the experimental results in the Reference [17].

Fig. 5. The phase difference of the pressure and velocity wave. (a) The cold end of the basic pulse tube; (b) The hot end of the basic pulse tube; (c) The cold end of the inertance pulse tube; (d) The hot end of the inertance pulse tube.

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4 Conclusion In this paper, the basic pulse tube and inertance pulse tube models are constructed by the molecular dynamics simulation method. The phase shift mechanism of the inertance pulse tube is analyzed from the microscopic perspective by analyzing the parameters of temperature, pressure, mass flow and velocity wave in the pulse tube. The research results are as follows. The temperature, pressure and mass flow trends of the cold end and hot end of the two types of the pulse tube refrigerator obtained by the molecular dynamics simulation method are consistent with the conclusions of the enthalpy flow model, so it is feasible to study the pulse tube refrigerator by the molecular dynamics simulation method. The research results of the phase difference between the pressure wave and the velocity wave at the cold and hot end of the pulse tube show that the phase difference at the cold end is close to 90° in the basic pulse tube, while it is close to 72° in the inertance pulse tube. In addition, the difference in phase difference at the hot end is also close to 18°. It can be seen that the phase shift capability of the inertance tube and the gas reservoir is verified by using the molecular dynamics simulation method. Acknowledgments. The project is supported by the National Science Foundation of Shandong Province, China (No. ZR2021QE182), the Postdoctoral Science Foundation of China (No. 2022M711955) and the Taishan Scholar Project (TSQN202103142).

References 1. Gifford, W.E., et al.: Surface heat pumping. In: Advances in Cryogenic, vol. 11, pp. 171−179 (1966). https://doi.org/10.1007/978-1-4757-0522-5_18 2. Nast, T., et al.: Development of a 4.5 K pulse tube cryocooler for superconducting electronics. In: AIP Conference Proceedings, vol. 985, pp. 881–886 (2008) 3. Che, Y., et al.: Simulation on alternating oscillation flow in microchannel pulse tube coupled with active piston using non-equilibrium molecular dynamics. Chem. Phys. Lett. 759, 137965 (2020) 4. Longsworth, R.C. An experimental investigation of pulse tube refrigeration heat pumping rates. In: Advances in Cryogenic Engineering. Springer US, pp. 608–618 (1967). https://doi. org/10.1007/978-1-4757-0489-1_63 5. Liu, S., et al.: Investigation on phase shifter of a 10 W/70 K inertance pulse tube refrigerator. Int. J. Refrig. 74, 450–457 (2017) 6. Luo, E., et al.: Inertance tube models and their experimental verification. In: AIP Conference Proceedings, vol. 710, pp. 1485−1492 (2004) 7. Roach, P.R., Kashani, A.: Pulse tube coolers with an inertance tube: theory, modeling and practice. Adv. Cryog. Eng. 43, 1895–1902 (1998) 8. Rout, S.K., et al.: Multi-objective parametric optimization of inertance type pulse tube refrigerator using response surface methodology and non-dominated sorting genetic algorithm. Cryogenics 62, 71–83 (2014) 9. Wang, N., et al.: A high efficiency coaxial pulse tube cryocooler operating at 60 K. Cryogenics 93, 48–50 (2018) 10. Radebaugh, R., et al.: Inertance tube optimization for pulse tube refrigerators. In: AIP Conference Proceedings, vol. 823, p. 59 (2006)

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11. Liu, S., et al.: Impact of coiled type inertance tube on performance of pulse tube refrigerator. Appl. Therm. Eng. 107, 63–69 (2016) 12. Qi, Y., et al.: Study on micro thermodynamic process of gas flow in pulse tube by unequilibrium molecular dynamics simulations. Int. J. Heat Mass Transf. 137, 669–676 (2019) 13. Lu, Y., et al.: Study on micro-liquid droplets behavior at the expander of the pulse tube cryocooler under 5 K based on molecular dynamics. J. Mol. Liq. 346, 118222 (2022) 14. Plimpton, S.: Fast parallel algorithms for short-range molecular dynamics. J. Comput. Phys. 117(1), 1–19 (1995) 15. Jiang, Y., et al.: Phase shift mechanism of pulse tube cooler based on fluid network theory. J. Nanjing Univ. Aeronaut. Astronaut. 37(4), 452–456 (2005) 16. Storch, P.J., Radebaugh, R.: Development and experimental test of an analytical model of the orifice pulse tube refrigerator. In: Fast, R.W. (eds.) Advances in Cryogenic Engineering. A Cryogenic Engineering Conference Publication, vol. 33, pp. 851–859. Springer, Boston, MA (1988).https://doi.org/10.1007/978-1-4613-9874-5_103 17. Liu, S., et al.: Investigation of the inertance tube of a pulse tube refrigerator operating at high temperatures. Energy 123, 378–385 (2017)

Study on Decreasing Cool-Down Time of a Rotary Cryocooler for HOT IR Detectors T. Xu(B) , C. Zuo, Y. J. Guo, T. H. Huang, W. J. Liu, C. Sun, and L. Huang Wuhan Globe Sensor Technology Co, Ltd., Wuhan, China [email protected]

Abstract. The CDT(cool-down-time)of the cryocooler is a key performance for IR detectors, especially for the specific application. In this paper, the model predicting the CDT of the rotary cryocooler was established, and the effects of the key parameters including filling pressure, frequency and the length of cold finger on the CDT of the cryocooler were investigated by simulation. The results show that shortening cold finger is the best way to decrease the CDT compared with optimizing other parameters. Finally, the optimization was made for a rotary stirling cryocooler for HOT IR detectors, the CDT of the cryocooler was predicted decreasing from 167.4 s to 105.6 s at 150 K@20 °C, and the power consumption at 70 °C of ambinent temperature roughly doubled simultaneously. Keywords: Rotary Stirling Cryocooler · Cool-down Time · Model · Cold Finger

1 Introduction Historically, most of the IR applications worked at 70 K−85 K with a limited cooling capacity for cryocooler. Today, as the development of Chip Technology, there is a change of requirements for cryocooler. For example, the operating temperature increased to 130 K−150 K for HOT application [1, 2], the CDT of detectors are limited for the special application. On one hand, it is necessary to update the cryocooler owing to transformation of the requirements. On the other hand, it is difficult for cryocooler manufacturer to update the products in due course owing to fixed development cycle. So, it is necessary for a cryocooler to address a different requirements of applications by some uncomplicated optimization [3, 4]. In order to obtain the direction of optimization quickly, it is important to research the effect of some key parameters on the performance of cryocooler. In this paper, research on decreasing cool-down time (CDT) of a rotary cryocooler under special application has been done. First of all, the model predicting the CDT of the rotary cryocooler was established, and the effects of the key parameters including filling pressure, frequency and the length of cold finger on the CDT of the cryocooler were investigated by simulation. Finally, the optimization was made for a rotary cryocooler for HOT IR detectors.

© Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 870–877, 2023. https://doi.org/10.1007/978-981-99-6128-3_113

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2 Theoretical Model Thermal mass of the load (F(t)) is proportional to the temperature of the cold end (T(t)), the effective mass of heat conduction(mi ) and specific heat capacity(ci ): F(t) =

c 

mi T(t)

(1)

i

The cooling capacity of cryocooler is proportional to the temperature of the cold end (T(t)): Q = aT(t) + b

(2)

The energy balance equation of the cooling process has following relationship: cm

dT + aT(t) + b = 0 dt

So, the CDT of cryocooler can be expressed as: c i mi · ln(1 + CDT = t(Tcold ) − t(Thot ) = a

(3)

(Thot −Tcold )·a Q(Tcold ) )

(4)

 Thus it can be seen that major effect factors include the coefficient of thermal mass ( ci mi ), the coefficient of performance (a) and the cooling capacity at target temperature (Q(Tcold ). For a cryoccoler, it is convenient to optimize the cryocooler by changing filling pressure, frequency and the length of cold finger.  The model of dewar is built by finite element, and the ci mi was calculated. Thermodynamic model of the cryocooler is built by Sage stirling model, the coefficient of performance (a) and cooling capacity at target temperature (Q(Tcold ) can be obtained, and the effects of the key parameters including filling pressure, frequency, and the length of cold finger on the CDT of the cryocooler were investigated.

3 Simulation Results and Analysis The simulation was based on a rotary cryocooler for HOT IR detector. The basic parameters of the cryocooler are as following: Based on the literature and practical experience, it is necessary to estimate the environmental adaptability of the cryocooler in the process of optimizing the CDT. According to the national military standard [5], the cryocooler need to keep working normally at 71 °C and −45 °C environmental temperature. However, the performance of the cryocooler will change with environmental temperature, which will lead to problems such as noise or temperature fluctuation [6]. Thus, the performance at different environmental temperature were estimated and the following requirements were considered: 1) The operating frequency at 200 mW@150 K@−45 °C should be higher than 15 Hz in order to avoid temperature fluctuation;

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2) The power consumption and the load torque should be as small as possible in order to avoid mechanical vibration or noise,especailly for ambinent temperature of 70 °C. For a rotary cryocooler for HOT IR, the effects of the key parameters including filling pressure, frequency and the length of cold finger on the CDT of the cryocooler were investigated as follows. In the simulation, the different performance index were caculated. At 20 °C of ambinent temperature, the performance index including CDT, peak PV power, peak load torque and so on were caculated. At 70 °C of ambinent temperature, the peak PV power was caculated in order to assess the nosie owing to increased peak PV power. At −45 °C of ambinent temperature, the operating frequency at 200 mW@150 K was caculated to assess the temperature fluctuation. 3.1 Changing Filling Pressure Figure 1,2 show the cooling capacity at 20 °C/−45 °C ambient temperature and the load torque with different filling pressure. The results show that the coefficient of performance (a) and cooling capacity at target temperature increased with filling pressure increasing. According to model formula, the CDT of cryocooler at 3MPa will decrease from 167.4 s to 114.3 s. Besides, the operating frequency at 200 mW@150 K@−45 °C decreased from 13.7 Hz to 10.8 Hz which will lead to temperature fluctuation and the load torque increased from 0.0052 N·m to 0.00945 N·m, which will induce more noise.

Fig. 1. Cooling capacity with different filling pressure

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Fig. 2. Cooling capacity at –45°C and load torque with different filling pressure

3.2 Changing Operating Frequency Figure 3,4 show the cooling capacity at 20 °Cambient temperature, PV power and load torque with different operating frequency. The results show that the coefficient of performance (a) and cooling capacity at target temperature increased with operating frequency increasing. According to model formula, the operating frequency need to increase from 75 Hz to110 Hz in order to achieve 110s of CDT. Besides, the peak PV power will increase 77% from 3.6 W to 6.4 W.

Fig. 3. Cooling capacity with different operating frequency

3.3 Changing the Length of Cold Finger Figure 5,6 show the cooling capacity at 20 °C/−45 °C ambient temperature and load torque with different length of cold finger. The results show that the coefficient of performance (a) and cooling capacity at target temperature increased with the length of cold finger shortening. According to model formula, the CDT of cryocooler decreased from 167.4 s to 118.5 s when the length of cold finger was reduced from 35 mm to 15 mm. The operating frequency at 200 mW@150@−45 °C keep stable, but the load torque increase from 0.0052 N·m to 0.0121 N·m,which will lead to side effect such as vibration and noise.

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Fig. 4. PV power and load torque with different operating frequency

Fig. 5. Cooling capacity with different length of cold finger

Fig. 6. Cooling capacity at –45 °C and load torque with different length of cold finger

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4 Optimization Results According to above analysis, increasing filling pressure, increasing operating frequency and shortening length of cold finger all will decrease the CDT of cryocooler, but all lead to side effect such as power consumption increasing or load toque increasing simultaneously. The simulation results show that shortening the length of cold finger is of greatest effect on CDT with minimal side effect. For example, increasing frequency will significantly increased load torque and increasing the filling pressure will lead to more power comsuption increasing. By comprehensive consideration, the optimization scheme was determined to decreasing filling pressure and shortening the length of cold finger. Figure 7,8 show the cooling capacity at 20 °C ambient temperature,PV power and motor torque before and after optimization. By the way, The CDT of the cryocooler declined from 167.4 s to 105.6 s at 150 K@20 °C, the operating frequency at 200 mW@150 K@−45 °C increased from 13.7 Hz to 16.7 Hz which will avoid temperature fluctuation, and the power consumption roughly doubled at 70 °C of ambinent temperature simultaneously. Table 1 show the basic parameters of the cryocooler. Table2 show the detail before and after optimization.

Fig. 7. Cooling capacity before and after optimization

Fig. 8. Cooling capacity at −45 °C and load torque before and after optimization

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Temperature of the cold end(K)

Filling pressure(MPa)

Operating frequency(Hz)

Length of the cold finger(mm)

CDT(s)

150

2

75

35

167.4

Table 2. Optimization results Ambient temperature

Performance

Before optimization(35 After optimization(15 mm + 2.0 MPa) mm + 1.6 MPa)

20 °C

CDT/s

167.4

Peak PV power/W

3.6

4.66

Peak load torque/N·m

0.00763

0.0104

Operating frequency/Hz

20

Operating PV power/W

0.766

1.569

70 °C

Operating PV power/W

1.075

2.332

−45 °C

Operating frequency/Hz

13.7

105.6

25.8

16.7

5 Conclusion The theoretical model predicting the CDT of the rotary cryocooler was established, and simulation results show that increasing filling pressure, increasing operating frequency and shortening the length of cold finger will decrease the CDT of cryocooler, and the power consumption will increase or operating frequency at lower ambient temperature will decrease simultaneously. By shortening the length of cold finger and decreasing filling pressure for a rotary cryocooler for HOT IR detectors, the CDT of the cryocooler was predicted decreasing from 167.4 s to 105.6 s at 150 K@23 °C, and the power consumption at 70 °C of ambinent tempearture roughly doubled simultaneously.

References 1. Filis, A., Haim, Z.B., Pundak, N., Broyde, R.: Micro miniature rotary sitirling cryocooler for compact, lightweight and low power thermal imaging systems, SPIE. In: Defense, Security, and Sensing, vol. 7298, pp. 729818:1–9 (2009) 2. www.ricor.com/product/#integral-rotary

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3. Katz, A., Haim, Z.B., Riabzev S., et al.: Development and optimization progress with RICOR cryocoolers for HOT IR detectors. In: Proceedings of SPIE, vol. 9821, pp. 9821, pp. 9210N:1– 13 (2016) 4. Vasse, C., Raynal, G., Martin, J.Y., Abousleiman, V., Benschop, T.: Applicable range and performance prediction model for Thales rotary coolers. In: Infrared Technology and Applications XLVII, vol. 11741, pp. 73−82. SPIE (2021) 5. General specification for Stirling cryocooler 6. Filis, A., Haim, Z.B., et al.: RICOR’s rotary cryocoolers development and optimization for HOT IR detectors.In: Proceedings of SPIE (2012)

Cryogenic Fuel and Transportation

Numerical Modelling of Transient Chill-Down Operation in Liquid Methane Transfer Line Keerthi Raj Kunniyoor(B) and Parthasarathi Ghosh Cryogenic Engineering Centre, Indian Institute of Technology, Kharagpur 721302, India [email protected]

Abstract. An accurate chill-down model is essential for the efficient design of cool-down operation in the cryogenic transfer line, which is crucial for optimized propellant usage in space flights. Heat transfer constitutive relations play a vital role in the accuracy of the thermo-hydraulic code used for this purpose. Hence, the recently proposed correlations in the literature for various two-phase boiling regimes and their transition criteria are investigated for their applicability to LCH4 chill-down using a non-homogeneous two-phase model. Built-in correlations in SINDA/FLUINT (S/F), correlation sets based on LN2 and LH2 cryogens, correlations reported based on LOX and LAr data, and correlations employed in previous chill-down models are studied. Though available correlation could predict rewetting temperature within 11% MAE, higher deviations are noted in the film boiling regime, which persists for the significant duration of the LCH4 chill-down period. Nevertheless, S/F built-in correlations and the chill-down correlation set based on LN2 and LH2 cryogens gave a rough estimate for the wall thermal transients and a reasonably good prediction for the chill-down duration. Keywords: Chill-down · heat transfer correlations · LCH4 · numerical modelling · transfer line

1 Introduction LOX/LCH4 propulsion technology has been considered as a promising alternative for the traditional cryogenic (LOX/LH2 ) and semi-cryogenic (LOX/kerosene) propulsion systems used in space launch vehicles due to being relatively less expensive, ease of handling, non-toxic, and ease of storing. Moreover, the possibility of synthesizing methane in space enhances the importance of liquid methane propulsion systems for future longrange space missions. At present, when global space agencies are actively involved in building reliable and cost-effective LOX/LCH4 propulsion technology to power future launches, it is also important to store, handle and transport LCH4 efficiently [1]. The transfer lines used for cryogenic fluid management in space industries and liquefaction plants must be cooled to liquid temperature before the transfer process can be initiated. This is to avoid any phase change of the incoming liquid and the possibility of two-phase flow at the pipe outlet, which may result in combustion instabilities in space launch vehicles [2]. Usually, a part of the fluid to be transported is sacrificed to achieve © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 881–888, 2023. https://doi.org/10.1007/978-981-99-6128-3_114

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this by directly passing it through the feed line and allowing it to boil by taking heat from the channel wall. This transient stage of fluid transfer, during which the pipeline that is initially at ambient condition is brought to liquid temperature, characterizes phase change flow and conjugate heat transfer phenomenon [3]. The development of accurate thermo-hydraulic models that can capture the complex heat and mass transfer phenomenon during the chill-down phase change flow is critical in saving valuable propellants, reducing non-productive time, and ensuring safe operation during chill-down [2]. Accurate constitutive relations to capture the wall-fluid thermal interaction play an essential role in the success of these numerical models. Assessment studies have reported that the traditional heat transfer correlations would result in huge prediction errors due to the inherent property difference of cryogenic fluid from room temperature fluids and the nature of quenching operation, which is different from steadyheated tube experiments from which available correlations are developed [4]. This paved the way for recent chill-down correlation development that has improved the model predictions [2, 5–7]. However, most of these studies are performed using LH2 and LN2 , though some data are also available for LOX and LAr chill-down. Recently, LCH4 chill-down data are also made available for the first time by the research group [8]. Also, the correlations proposed in [2] using LN2 and LH2 data are reported to give higher MAE (mean absolute error) when applied for LCH4 chill-down heat transfer prediction. Hence, the present work aims to model the LCH4 chill-down operation using a non-homogeneous two-phase model in SINDA/FLUINT. The built-in heat transfer correlation set in S/F is initially tested, followed by a thorough investigation of alternative heat transfer correlations in the literature for each boiling regime and its transition criteria for LCH4 chill-down modeling. Finally, the correlation set that gives the best prediction is used in the chill-down model for wall temperature prediction.

2 Numerical Model The numerical model developed using Thermal desktop, FloCAD, and SINDA/FLUINT (S/F) software package is shown in Fig. 1(a & b) [9]. Here, the transfer line wall material is discretized into wall nodes interconnected by wall conductors. Whereas the fluid domain is discretized using lumps and paths. The thermal interaction between the wall and fluid is established using a wall-fluid heat transfer tie which estimates the quench heat flux based on boiling regime-specific heat transfer correlations. The boiling characteristics during chill-down depend on several parameters related to feed line, flow, and ambient conditions, such as pipe geometry, orientation, flow inlet conditions, mass flow rate, etc. A non-homogeneous two-phase slip flow model is used to simulate the phase change flow during chill-down. It solves the mixture mass Eq. (1) mixture energy Eq. (2) and phase momentum equations Eq. (3) along with constitutive relations for wall-fluid friction (Kf ), change in momentum due to change in density (Rl ), interphase drag (FD), momentum transfer due to phase generation (FG), virtual mass (AM), etc. Equation (3) is written for the liquid flow rate. The corresponding equation for vapor flow rate can be obtained by simply interchanging the subscripts 1 (for liquid) and 2 (for vapor) in Eq. (3). For brevity, the momentum equation is not written separately for the vapor phase and can be easily obtained by changing the subscripts. Hence, a separate correlation for slip ratio

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is not needed as in the two-phase drift flux model. The solution of governing equations gives the velocity of each phase. Further, the continuity and energy equations are solved at the fluid lumps, where the scalar properties, such as pressure, temperature, quality, etc., are defined. The phase momentum equations are solved at paths where the mass flow rate and flow velocity are defined (i.e., a staggered grid approach). Additionally, the energy conservation Eq. (4) on the wall node is also solved. Parasitic heat in-leak to the transfer line from its surrounding is also considered in the model (last term in Eq. (4)). For vacuum-insulated lines, this last term accounts for the radiation and gas conduction heat in-leak. dm ˙ up − m ˙ down =m dt

(1)

dU ˙ ˙ up hup − m ˙ down hdown + Q =m dt

(2)

Acj ˙ jpath(1) dm = Lj path(1) [αjpath(1) {(Pup − Pdown ) − (ρgsinθ L)jpath(1) } dt path(1) exp ˙ jpath(1) |m ˙ j2path(1) }] ˙ jpath(1) | jpath(1) + Rl m + αj 1 {−Kfjpath(1) m path(1) ˙ jpath(1) − αj 1 FDjpath(2) m ˙ jpath(2) + αj 1 FDjpath(1) m path(1) path(2) 1 1 ˙ jpath(1) ) + ( αj ˙ jpath(2) )] +jpath(1) [( αj FGjpath(1) m FGjpath(2) m path(1)

˙j dm

path(2)

(3)

˙j dm

+αjpath(2) AMjpath(2) path(2) − αjpath(1) AMjpath(1) path(1) dt dt     dTi   KAc mCp i = (Ti±1 − Ti ) + hk Asin Tjlump − Ti + qpara dt L i±1,i

(4)

Fig. 1a. Numerical model isometric view (1) inlet, (2) fluid lump, (3) heat transfer tie, (4) wall node, (5) heat in-leak, (6) fluid path, (7) temperature sensor, (8) outlet

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Fig. 1b. A schematic drawing of transfer line (front view)

3 Results and Discussion 3.1 LCH4 Chill-Down Prediction Using SINDA/FLUINT built-in Correlations The LCH4 chill-down experimental data used in this work are taken from [8]. It is performed using a 4.11 m long, 2.21 cm inner diameter, 1.65 mm thick vacuum insulated horizontal stainless steel transfer line with outer wall temperature measurements available at two stations, S2 and S4. LCH4 is supplied to the transfer line at 448.2 kPa and 131K inlet flow condition with an inlet subcooling degree of 2.4 K. The average inlet mass flux is 150 kg/m2 s. The numerical model consists only of the feed line, and other flow components in the experimental setup used in [8] were not modeled. Based on a grid-independent study, 30 pipe divisions are used in the numerical model. Further, mass flow inlet and pressure outlet boundary conditions are used with an initial condition set based on the experimental conditions at time zero. SINDA/FLUINT built-in heat transfer correlations, shown in table 1, are used initially to understand their suitability for LCH4 chill-down modeling. The algorithm used for boiling regime selection is discussed in Sect. 3.6.7 of [9] and hence is not presented here for brevity. Table 1. SINDA/FLUINT built-in heat transfer correlations Boiling regime

Correlation [9]

Film boiling (hfb )

Bromley correlation for low-quality forced convection film boiling, Max (Groeneveld, Dittus-Boelter) for the high-quality film boiling

Rewetting temperature Minimum of departure from film boiling (Ramilson-Leinhard) and (Trewet ) Leidenfrost temperatures (Zuber correlation) Transition boiling (htb )

Interpolation based on Ramilison-Leinhard scaling laws

Critical heat flux (qchf )

Modified Zuber-Griffith correlation with sub-cooling correction by Gambill multiplied with a scaling factor of π/24

Nucleate boiling (hnb ) Chen correlation Single phase (hsp )

Dittus-Boelter correlation

Based on the heat transfer coefficient versus equilibrium quality profile characteristics discussed in [10, 11], chill-down is predominantly in an inverted annular film boiling

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regime. Inverted slug flow is also noted downstream. Due to higher heat transfer during the transition and nucleate boiling, associated boiling regimes, such as annular and bubbly flow, exists for a short period towards the end of the chill-down. Further, flow rate fluctuations arising from phase change are also noted in the simulation as flow proceeds to downstream nodes. Regarding the accuracy of numerical prediction, Fig. 2(a) compares the wall temperature from S/F at stations S2 and S4 with the corresponding experimental results. Further, the boiling regime prediction from S/F at station 2 is also plotted alongside its wall temperature transient in Fig. 2 (b). At both stations, S/F predicts the film boiling to initiate in the first few seconds of the chill-down operation, whereas experimental data shows a slow decrease in the wall temperature initially, which may be due to the single-phase vapor flow through the downstream location of the pipe. Later, when the wall temperature drops below 250K, the slope of the experimental wall temperature transient shows a sudden increase, which can be attributed to the initiation of film boiling at these downstream locations. Further, from experimental results at S2 and S4, a higher cooling effect is noted at the pipe upstream during film boiling, as indicated by a higher wall temperature vs. time curve slope. On comparing these observations with S/F predictions during film boiling, it is noted that the initial film boiling heat transfer predictions are in the range of the experiment; however, as wall temperature approaches the rewetting point, heat transfer is under-predicted. The rewetting temperature prediction (177K) at station 2 is lower than the experimental data (188K), whereas it is over-predicted (177K) at station 4 (163K in the experiment). Further, the transition and nucleate boiling heat transfer predictions are comparable with the experiment at station 4, whereas it is under-predicted at station 2.

Fig. 2. (a) SINDA/FLUINT validation, (b) boiling regime at station 2. In Fig. 2. (b), right vertical axis, 0 = single phase, 1 = subcooled nucleate boiling, 2 = saturated nucleate boiling, 3 = transition boiling, 4 = film boiling, 13 = transition boiling with partial wall dry-out, 14 = film boiling with partial wall dry-out [9]

3.2 Investigation of Heat Transfer Correlations for LCH4 Chill-Down Modelling The experimental wall temperature transients obtained from [8] are converted to heat flux based on an inverse heat conduction problem. Further, pressure and mass flux data

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Correlation

MAE

PDI 50

PDI 75

Comments

Jin et al. [6]

210.3

26.7

58.9

Based on LN2 , LH2 , and LAr chill-down

Darr et al. [14]

536.2

50

53.5

Based on LH2 in vertical line

Darr et al. [2]

153.2

29.4

53.5

LN2 , LH2 chill-down (various orientations)

Jin et al. [15]

197.2

11.6

51.7

LN2 chill-down of horizontal line

Luchinsky et al. [16]

599.4

36.6

50

Modified Bromley correlation

Jin et al. [17]

349.9

53.5

54.4

LN2 chill-down of horizontal line

Darr et al. [5]

162.2

30.3

50.8

LN2 chill-down under various orientations

Hu et al. [18]

2713

0

0

LN2 chill-down of rectangular channel with saw teeth inner wall surface

Chi et al. [19]

740.8

18.7

26.7

LH2 chill-down, Dittus-Boelter form

needed for correlation assessment are taken from [8], whereas the enthalpy and quality are calculated using the first law of thermodynamics. Table 2 shows the summary of the film boiling correlation assessment. It is noted that all the recent correlations gave a higher mean absolute error. Minimum MAE is noted for Darr et al. [2], 153.2%, with 30.3% and 50.8% data within ±50% and ±75% of experimental results, respectively. Further, in Table 2, PDI is used as an abbreviation for the percentage of data inside ±X % of the experimental results. For instance, the percentage of data inside ±50% of the experimental results is written in short as PDI 50. Regarding the rewetting temperature prediction, the best results are obtained using Darr et al. [12] (11.8% MAE) and Jin et al. [13] (12.7% MAE). The performances of other rewetting correlations are summarized in Table 3. Transition and nucleate boiling regimes characterize higher heat transfer coefficients and thus exist only for a very short period. Nucleate boiling correlations reported in [5] and [2] showed an MAE of 125.97% and 73.63%, respectively, whereas qchf correlations from [2] and [6] are found to give 47.5% and 2428.3% MAE respectively. 3.3 Transfer Line Wall Temperature Prediction Using Final Correlation Set The final correlation set consists of hfb, hnb, and qchf correlations proposed by Darr et al. [2] and Trewet from Darr et al. [12]. Here, the transition boiling regime is modeled by interpolating the minimum and critical heat flux. From Fig. 3, which shows the wall temperature prediction comparison with the experimental data, it is noted that the film boiling heat transfer is mostly under-predicted, whereas the rewetting temperature is over-predicted (11.8% MAE). Further, the slope of wall thermal transients post-rewetting point is comparable with that of the experiment, especially at station 4.

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Table 3. Chill-down rewetting temperature correlations assessment for LCH4 Correlation

MAE

Comments

Zhang et al. [7]

25.0

Based on outlet constrained LN2 feed line chill-down

Jin et al. [6]

4732

LN2 , LH2 , and LAr chill-down in horizontal feed line

Jin et al. [13]

12.7

Based on LAr chill-down of horizontal feed line

Darr et al. [2]

27.7

LN2 , LH2 chill-down under various orientations

Darr et al. [12]

11.8

LN2 chill-down under various orientations

Hu et al. [18]

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203.6

Berenson correlation with Iioeje correction

Fig. 3. Prediction from final correlation set

4 Conclusion Numerical modeling of feed line chill-down operation in the LCH4 transfer line is investigated using a non-homogeneous two-phase model with different heat transfer constitutive relations from literature. SINDA/FLUINT built-in correlations, as well as a correlation set consisting of hfb, qchf, and hnb from [2], Trewet from [12], gave a reasonable estimate for wall thermal transient, through high MAE is noted during film boiling, which persists for the significant duration of the LCH4 chill-down operation. Further, both predicted the time taken for chill-down operation with reasonable accuracy. Available chill-down correlations need modification to improve LCH4 model predictions, which further calls for additional experimental and numerical studies on this subject.

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References 1. Neill, T., Judd, D., Veith, E., Rousar, D.: Practical uses of liquid methane in rocket engine applications. Acta Astronaut. 65(5–6), 696–705 (2009) 2. Darr, S.R., et al.:Two-phase pipe quenching correlations for liquid nitrogen and liquid hydrogen. J. Heat Transf. 141(4), 042901 (2019) 3. Kunniyoor, K.R., Govind, R., Venkateswaran, K.S., Ghosh, P.: Liquid hydrogen pipeline chill-down: mathematical modelling and investigation. Cryogenics 118, 103324 (2021) 4. Hartwig, J., Darr, S., Asencio, A.: Assessment of existing two phase heat transfer coefficient and critical heat flux correlations for cryogenic flow boiling in pipe quenching experiments. Int. J. Heat Mass Transf. 93, 441–463 (2016) 5. Darr, S.R., et al.: An experimental study on terrestrial cryogenic tube chilldown II. effect of flow direction with respect to gravity and new correlation set. Int. J. Heat Mass Transf. 103, 1243–1260 (2016) 6. Jin, L., Lee, J., Jeong, S.: Investigation on heat transfer in line chill-down process with various cryogenic fluids. Int. J. Heat Mass Transf. 150, 119204 (2020) 7. Zhang, J., Wang, K., Chen, L.: Experimental study on liquid oxygen chill-down in a horizontal exit-contracted pipe. Cryogenics 120, 103387 (2021) 8. Hartwig, J., Meyerhofer, P., Stiegemeier, B., Morehead, R.: Liquid methane and liquid oxygen horizontal chilldown experiments of a 2.54 and 11.43 cm transfer line. Appl. Therm. Eng. 205, 118042 (2022) 9. Cullimore, B., Ring, S., Johnson, D.: User’s Manual, SINDA/FLUINT, thermal/fluid network analyzer for CRThermal desktop version 6.2. Cullimore and Ring Technologies, Inc. (CRTech) (2020) 10. Kunniyoor, K.R., Ghosh, P.: Development of transient flow film boiling heat transfer correlations for energy efficient cryogenic fluid management during feed line quenching operation. Int. J. Heat Mass Transf. 204, 123806 (2023) 11. Kunniyoor, K.R., Ghosh, P.: Investigation of quench flow boiling heat transfer correlations for liquid oxygen feed line chill-down with outlet contraction. Cryogenics 128, 103593 (2022) 12. Darr, S.R., Dong, J., Glikin, N., Hartwig, J.W., Chung, J.N.: Rewet temperature correlations for liquid nitrogen boiling pipe flows across varying flow conditions and orientations. J. Therm. Sci. Eng. Appl. 11(5), 051008 (2019) 13. Jin, L., Cho, H., Jeong, S.: Experimental investigation on line chill-down process by liquid argon. Cryogenics 97, 31–39 (2019) 14. Darr, S.R., Hartwig, J.W.: Two-phase convection heat transfer correlations for liquid hydrogen pipe chilldown. Cryogenics 105, 102999 (2020) 15. Jin, L., Cho, H., Lee, C., Jeong, S.: Experimental research and numerical simulation on cryogenic line chill-down process. Cryogenics 89, 42–52 (2018) 16. Luchinsky, D.G., Khasin, M., Timucin, D., Sass, J., Brown, B.: Inferential framework for two-fluid model of cryogenic chilldown. Int. J. Heat Mass Transf. 114, 796–808 (2017) 17. Lingxue, J., Park, C., Jeong, S.: Analysis of heat transfer characteristics for Liquid nitrogen line chill-down process. In: Asian Conference on Thermal Sciences 2017. The Korean Society of Mechanical Engineers (2017) 18. Hu, H., Wijeratne, T.K., Chung, J.N.: Two-phase flow and heat transfer during chilldown of a simulated flexible metal hose using liquid nitrogen. J. Low Temp. Phys. 174(5), 247–268 (2014) 19. Chi, J.W.H. Cooldown Temperatures and Cooldown Time during Mist Flow. In: Timmerhaus, K.D. (eds.) Advances in Cryogenic Engineering. Advances in Cryogenic Engineering, vol 10, pp. 330−340. Springer, Boston, MA (1965). https://doi.org/10.1007/978-1-4684-3108-7_40

Computations of Capillary-Driven Cryogenic Flows in the Interior Corner with Microstructures Ran Xu, Huan Jia, Han Chen, Xin Cheng, Mingkun Xiao, Guang Yang(B) Yonghua Huang, and Jingyi Wu

,

Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai 200240, China [email protected]

Abstract. The capillary-driven flow in the interior corners is crucial for cryogenic propellant management under microgravity. However, interior corner flow is difficult for cryogenic fluids due to low surface tension. In this study, the capillary-driven flow of liquid nitrogen in an interior corner with microstructure was simulated based on the Volume-of-Fluid method. Flow patterns of the interior corner with and without microchannels were comprehensively studied. The effects of microstructures’ width and depth were studied to reveal their impacts on flow behavior. The results show that the rational design of the microstructures effectively promotes the mass flow rate compared with the pure interior corner. This study provides insights into the flow physics of the interior corner with microstructures and a framework to optimally design related devices. Keywords: Capillary-driven flow · Cryogenic fluid · Microstructure

1 Introduction The on-orbit propellant storage technology is significant to the rapid development of space science. Due to lacking gravity in space, the gas-liquid interface is uncertain. Thus, propellant management devices (PMDs) are widely used in the field of propellant storage in space due to the requirement of transferring pure liquid to the engine. One of PMDs is the deflector in the surface tension tank. In microgravity conditions, the liquid is dominated by surface tension which is created by molecule cohesion [1]. In most cases, the molecule cohesion of the solid is much greater than that of the liquid, so the liquid will spread out along the wall in the corner, and it is called interior corner flow. Since the Concus-finn condition was proposed in 1969 [2], many studies on internal corner flow have been carried out. However, due to the high cost and difficulties of conducting microgravity experiments, numerical simulations are widely employed. Li et al. [3] analyzed the capillary flow in the fan-shaped asymmetric interior corner by numerical simulation. Gurumurthy et al. [4] investigated the spontaneous rise of liquid in an array of rectangular corners under gravity with the Volume-of-Fluid method. It’s © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 889–895, 2023. https://doi.org/10.1007/978-981-99-6128-3_115

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revealed that the rivulets at long times obey the one-third power law in time, with a weak dependence on the geometry. McCraney et al. [5] simulated the free surface with an OpenFOAM solver. Their predictions of drain transients were validated by the space experiments. In recent years, cryogenic fluids have been employed as rocket propellants due to their high specific impulse. However, the feature of low surface tension leads to difficulty in interior flowing driven by surface tension. Inspired by microstructure showing great influence on the flow patterns, a composite interior corner with microstructures was designed in this study. The Volume-of-Fluid method was employed to numerically investigate the spontaneous rise of liquid nitrogen in the interior corner with microstructure under microgravity. Flow patterns of an interior corner with and without microchannels were comprehensively studied. The microchannel structures were optimized by different widths and depths to reveal their impacts on flow behavior. We hope that this approach gives guidance in designing propellant management devices.

2 Method 2.1 Physical Model The flow characteristics of an interior corner with microchannels (ICM) were numerically studied. The physical model and the relevant dimensions are schematically shown in Fig. 1. Rectangle microchannels were added to the side wall of triangle prism. The pure interior corner (PIC) without microchannels was used as the controlled group as well to investigate the performance advancement of the interior corner with microchannels. The widths and depths of microchannels were investigated in this study.

Fig. 1. (a) The isometric view and (b) the top view of the physical model for ICM and PIC. The blue part is liquid and the grey part is vapor. The values of L, H, D, w0 , x, 2α are 35 mm, 10 mm, 8 mm, 0.6 mm, 0.4 mm, 45°, respectively.

Liquid nitrogen was initially filled with a fixed length at the bottom of the triangular prism. Liquid and vapor nitrogen were used in this study. The temperature of working fluids was 73 K higher than the triple point of nitrogen (63.18 K). Thus, the liquid and vapor nitrogen coexisted in the interior corner. The physical properties of working fluids are shown in Table 1. The static contact angle is considered on account of the assumption that the walls are ideally smooth.

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Table 1. Physical properties of the working fluids. Temperature (K) Density (kg/m3 ) Viscosity (µPa·s) Surface tension coefficient (mN/m)

Fluid

Nitrogen (liquid) 73 Nitrogen (vapor)

824.85

191.20

2.84

5.12

9.8

2.2 Governing Equations and Boundary Conditions In this study, the flow was regarded as laminar, isothermal, and incompressible. The effect of viscous dissipation and gravity was neglected for the simulations. The Volume-ofFluid (VOF) method was used for numerical simulation, which could be solved through the finite volume method [6]. The continuity, Navier–Stokes, and phase equations for the 3D incompressible flow are given as Eqs. 1–3. ∇ ·U =0

(1)

∂ρU + ∇ · (ρUU ) = −∇P + Fs + ∇ · [μ(∇U + ∇U T )] ∂t

(2)

∂αl + ∇ · (U αl ) = 0 ∂t

(3)

where U is the velocity vector, P is the pressure field, α l is the phase fraction of liquid. F s is surface tension. ρ, μ are the density, and viscosity of mixed fluids, respectively, which can be calculated as eqs. 4 and 5. ρ = αl ρl + (1 − αl )ρv

(4)

μ = αl μl + (1 − αl )μv

(5)

Boundary conditions were set as follows. The side Walls 1–3 and the undersurface were no-slip walls. The top surface was at atmospheric pressure. The contact angles of wall 1, wall 2, and undersurface were set to 0° due to the property of the cryogenic fluid. [7] Wall 3 was set to 90°, which was assumed to not affect interior corner flow. Due to the symmetric geometry, the half of model is numerically calculated. The pressure-velocity coupled equations resulting from the discretization of (1), (2) were solved by employing the pressure implicit splitting operator algorithm (PISO). The iterative convergence for each time step was clarified when the relative variation of any variable is less than 10–6 . 2.3 Mesh Independent and Validation A grid independence simulation was conducted to examine the reliance of numerical accuracy on grid cells. The test was performed during liquid nitrogen flow through the interior corner with microchannel at different times. Fig. 2a demonstrates that the flow

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Fig. 2. (a) Variation of flow length with total grid numbers during 0.1 s using liquid nitrogen flow. The dimensions of these microchannels are d = 0.2 mm, w = 0.4 mm. (b) Comparison of the simulated flow length results with those of benchmark experiments.

length at the moment of 0.1s is almost invariable for each case as the grid number increases. The accuracy of the numerical model was examined by comparing the simulated results with those obtained through experiments. Fig. 2b exhibits the validation of the flow length for different viscosity values of polydimethylsiloxane (PDMS) in the triangle prism [8]. For high viscosity, the simulated flow lengths from the present model are in accordance with the experimental data and analytical solution of the same geometry. Although this divergence for low viscosity is notably larger than that for high viscosity, the average difference between the experiment and simulation is less than 20%. Thus, the average difference of the present numerical approach can satisfactorily simulate interior corner flow.

3 Result and Discussion 3.1 The Flow Patterns The instantaneous snapshots of capillary rise are observed with h = 0.2 mm and w = 0.5 mm microchannels as shown in Fig. 3a. The initially flat surface forms a meniscus and creates a local underpressure region, which leads to capillary rise. The meniscus close to the microchannel and interior corner leads to the rise of liquid level, while the capillary rise can only be seen in the interior corner of PIC (Fig. 3b). Capillary rise of the microchannel away from the interior corner is faster than that close to the interior corner at the beginning. The liquid column extends in the initial stages due to the inertia force [10], which decays as affected by the viscous friction. Subsequently, the capillary rise in interior corner surpasses and drives the flow in the microchannel close to the interior corner. Although the flow length of I0 in PIC is longer than that of M0 in ICM (Fig. 4a), the microchannel plays an important role in improving the mass flow rate. As seen in Fig. 4b, the mass flow rate raises dramatically at the beginning because the capillary force dominates the flow. [9] As the fluid flows continuously, the viscous friction resistance

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also increases. With time elapsing, the viscous friction exceeds the capillary force as the dominant driving force. Thus, the flow slows down and the mass flow rate declines. Thus, the increment effect of microchannel on mass flow rate gradually recedes.

Fig. 3. The instantaneous snapshots of capillary rise (a) with h = 0.2 mm and w = 0.5 mm microchannels in ICM, (b) in PIC. The blue part is liquid and the grey part is vapor. The color of the liquid close to wall 1 is deepened.

Fig. 4. Transient characteristics comparison between ICM and PIC. (a) Flow length of different corners, (b) Mass flow rate. I0 means the interior corner of PIC. M0 and M1–M3 are located in the interior corner and microchannels with d = 0.2 mm and w = 0.5 mm in ICM. The mass flow rate is selected in the initial cross section.

3.2 The Influence of Width With the depth fixed at 0.2 mm, the mass flow rate at the selected moment increases slightly with the increase of width as shown in Fig. 5a. This phenomenon results from the increase of the fluid volume in each microchannel. Especially, the liquid is redistributed between the internal corner and microchannels, which is called the liquid distribution behavior. The flow length with width at the moment of 0.05s is illustrated in Fig. 5b. With the width increasing, the liquid tends to distribute to the microchannel away from the internal corner (M3), while the flow length in the microchannels close to the internal corner (M1, M2) is shorter. However, the mechanism of the liquid distribution behavior needs to be further explored.

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Fig. 5. Variation of (a) the mass flow rate with width at the moment of 0.02–0.08 s, (b) flow length with width in different channels. d is equal to 0.2 mm.

3.3 The Influence of Depth The variation of the transient mass flow rate and flow length with depth is presented in Fig. 6 at different time steps. It is found that the mass flow rate increases with the depth. Compared with Fig. 6a, the increase of per unit depth could significantly promote the mass flow rate, which displays steeper lines in Fig. 6b. In other words, depth has a more pronounced effect on mass flow rate than width. Besides, as the depth increases, the flow length in all microchannels (M1−M3) shows an upward trend while the flow length in M0 decreases. This is because the flow increment resulting from the depth change is also influenced by the distribution behavior, displaying the phenomenon that the fluid that should flow into the interior corner is transported to the microchannels.

Fig. 6. Variation of (a) the mass flow rate with depth at the moment of 0.02–0.08 s, (b) flow length with depth in different channels. w is equal to 0.4 mm.

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4 Conclusion In this study, the capillary-driven flow of liquid nitrogen in an interior corner with microstructures was simulated using the VOF method. The accuracy of the model was verified by comparing the simulated results with experimental data. We studied the flow behaviors of the interior corner with and without microchannels, and found that the microchannels play an important role in improving the mass flow rate. The variation trends of mass flow rate and flow length with the width and depth of microchannels were obtained. The results show that the flow distribution behavior is different for various widths and depths, which is affected by the distance between the interior corner and microchannels. Through fully resolved numerical modeling, this study provides insights into the flow physics of the interior corner with microstructures, which is promising to be applied to the design of propellant management devices. Acknowledgments. The project is supported by the National Natural Science Foundation of China (Nos. 51936006 and 52276013).

References 1. Chen, Y., Collicott, S.H.: Experimental study on capillary flow in a vane-wall gap geometry. AIAA J. 43(11), 2395−2403.(2005) 2. Li, J., Chen, X., Huang, Y.: The review of the interior corner flow research in microgravity. Procedia Eng. 31, 331−336 (2012) 3. Yong-Qiang, L., Wen-Hui, C., Ling, L.: Numerical simulation of capillary flow in fan-shaped asymmetric interior corner under microgravity. Microgravity Sci. Technol. 29, 65−79 (2017) 4. Gurumurthy, V.T., Roisman, I.V., Tropea, C., et al.: Spontaneous rise in open rectangular channels under gravity. J. Colloid Interface Sci. 527, 151–158 (2018) 5. McCraney, J., Weislogel, M., Steen, P.: OpenFOAM simulations of late stage container draining in microgravity. Fluids 5(4), 207 (2020) 6. Gurumurthy, V.T., Rettenmaier, D., Roisman, I.V.: Computations of spontaneous rise of a rivulet in a corner of a vertical square capillary. Colloids Surf. Physicochemical Eng. Aspects 544, 118−126 (2018) 7. Wang, Z., Yang, G., Wang, Y., et al.: A three-dimensional flow model of screen channel liquid acquisition devices for propellant management in microgravity. Npj Microgravity 8(1), 28 (2022) 8. Weislogel, M.M., Lichter, S.: Capillary flow in an interior corner. J. Fluid Mech. 373, 349–378 (1998) 9. Dreyer, M., Delgado, A., Path, H.J.: Capillary rise of liquid between parallel plates under microgravity. J. Colloid Interface Sci. 163(1), 158−168 (1994) 10. Quéré, D., Raphaël, É., Ollitrault, J.: Rebounds in a capillary tube. Langmuir 15, 3679–3682 (1999)

Space Cryogenics

Effect of Wicking Capability on the Reseal Pressure of Woven Screens for On-Orbit Cryogenic Propellants Management Chenfangda Shi1 , Ye Wang1 , Yilin Lin1 , Feng Ren2 , Gang Lei3 , Guang Yang1(B) and Jingyi Wu1

,

1 Institute of Refrigeration and Cryogenics,

Shanghai Jiao Tong University, Shanghai 200240, China [email protected] 2 Aerospace System Engineering Shanghai, Shanghai 201109, China 3 State Key Laboratory of Technologies in Space Cryogenic Propellants, Beijing 100028, China

Abstract. In the microgravity environment, the instability of the gas-liquid distribution in the storage tank of propellants is the key issue that the power system of the spacecraft should address. Passive phase separation based on porous materials, is a promising technique to guarantee vapor-free propellant delivery. Porous woven screens have been widely utilized due to their physical durability. The reseal pressure of the screen describes the ability to recover the phase separation capability after the bubble breakthrough failure. In this study, experiments are carried out to measure the reseal pressure and the bubble point pressure of various mesh specification. Different working fluids, including water, kerosene and HFE 7500, are considered. A dimensionless relaxation pressure is proposed to describe the self-healing ability of the screen, which is a function of the bubble point pressure and the reseal pressure. The results show that the relaxation pressure is inversely proportional to the wicking parameter, which is described by the surface tension, contact angle and viscosity of the fluids. Better wicking capability accelerates the rewetting and the resealing process of the opening pores after the bubble breakthrough, which further improves the self-healing ability of the porous screen. Keywords: Cryogenic Propellant · Porous Screen · Phase Separation · Reseal Pressure · Wicking Capability

1 Introduction With the development of aerospace industry, deep-space exploration missions require spacecraft to adapt to longer working hours, greater distances and larger loads. Thus, cryogenic propellants with higher specify impulse and lower cost have increasingly become the priority. For long-term missions, it is impractical to carry enough propellant from the ground. As early as in the 1960s, researchers first proposed the concept of on-orbit refueling [1]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 899–906, 2023. https://doi.org/10.1007/978-981-99-6128-3_116

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During the working process, the storage tank is in microgravity environment where the gas and liquid phases mix together without clear interphase. Effective phase separation method is required to ensure that pure liquid fuel is delivered to the engine. Considering the microgravity environment, on-orbit gas/liquid separation methods based on surface tension have been proposed. Liquid acquisition devices (LADs) can be mainly categorized to vane type, sponge type and screen-channel type [2]. Benefiting from the capillary effect, the screen-channel LADs overcome the disadvantages of low surface tension of the cryogenic propellants and do not consume extra energy. The applicability of screen channel LADs to microgravity environment and cryogenic fluid has been verified by experiments [3]. Most of the studies on liquid acquisition devices focus on the flow-through-screen pressure loss and bubble point pressure[4], as a larger bubble point pressure directly determine a better critical phase separation ability of the porous screen, also called mesh. In recent years, several studies have been conducted to gain a better understanding of the phase separation performance of screen channel LADs. Hartwig et al. [5–7] experimentally studied the liquid acquisition performance of screen channel LADs for cryogenic fluids such as liquid hydrogen and liquid oxygen, and proposed performance optimization methods. Kudlac [8] analyzed the operating performance of different screen structures by considering the experimental data on bubble point pressure and pressure loss prediction equations. Jurns et al. [9, 10] investigated the effect of microporous crosssectional shape on the bubble point pressure of the screen. Subsequently, Meserole et al. [11, 12] studied the effect of pressurized gas on separation performance by experimental and theoretical analysis. Reseal pressure is another important parameter, which characterizes the self-healing ability of the screen, and smaller reseal pressure indicates a better self-healing ability. Hartwig and Kamotani [13] obtained an empirical model of reseal pressure through a large number of experimental data. They pointed out that there are seven parameters affect the reseal pressure, including surface tension, contact angle, screen reseal diameter, liquid temperature, degree of subcooling, and pressurant gas type and temperature. In this study, we conduct more experiments to study the effects of the fluid property and mesh specification on the reseal pressure. Moreover, the reseal pressure is analyzed in a dimensionless form, to better understanding its physical insights. Relaxation pressure is proposed, which reflects the difference between the bubble point pressure and the reseal pressure.

2 Experimental Setup The experimental devices to measure the reseal pressure consists of a transparent pipe, a data acquisition system, a gas cylinder, a gas flow controller and the screen components (Fig. 1). The transparent pipe consists of the upper and lower parts. The screen is assembled between the two sections by compression. The two probes of the differential pressure sensor are arranged above and below the screen, respectively. The lower part formed a completely closed volume by the chamber wall and the sealing liquid inside the screens. The upper part was open to ambient environment. A differential pressure transducer (DPT) was connected to the closed volume on one side and to the ambient

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environment on the other side. The differential pressure was recorded at 3 Hz using the data acquisition system. A high-speed camera was mounted to record the bubble breakthrough processes.

Fig. 1. Configuration (left) and photographs (right) of the experimental devices [14].

Prior to each test, the screen samples were fully cleaned. After mounting the screen sample on the test chamber, the fluid was filled into the upper part with a fill height of 15 mm to maintain a totally closed volume underneath. Then, the closed volume was pressurized by injecting nitrogen at a uniform and continuous flow rate of 3 sccm, which was optimized to minimize the influence of the pressurization rate and to acquire a quasisteady condition [15]. The pressurizing was stopped when there were bubbles popping out from the screen, in which case the screen failed completely. Besides, the camera can record the whole process of the screen from bubble breakthrough to self-healing. These processes can also be observed by the variations of the DPT signal. Figure 2 shows the pressure change process under a typical working condition (with water and oxidation mesh), where the bubble point pressure (pbp ) and the reseal pressure (pres ) are marked in the figure.

Fig. 2. Variation of pressure difference over time [14]

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In this study, a relaxation pressure (Prel ∗ ) is defined, which is dimensionless and is written as follow. ∗ Prel = Prel /Pbp

(1)

where Prel is the difference between bubble point pressure and reseal pressure. Prel = Pbp − Pres

(2)

Five specifications of the porous screens were considered in the experiments, including 40 × 430, 100 × 850, 130 × 1100, 203 × 1600, and 325 × 2300, all of which were fabricated using Dutch Twill weaving. Here, M × N denotes M warp wires and N shute wires per square inch on the mesh. The effect of modified screens (original, oxide and fluoride) on the self-healing process was also investigated. After different surface treatments, the screens were proved to change their wettability but the pore structures remain unchanged [14]. The contact angles between different fluids and the three different stainless steel sheets (SSS) are shown in Table 1. Table 1. Contact angle of different fluids on various surfaces Fluids

SSS

Oxide@SSS

Fluoride@SSS

HFE 7500

≈0°

≈0°

≈0°

Water

73°

48°

105°

Kerosene

20°

≈0°

53°

3 Results and Discussion 3.1 Effect of Contact Angle In order to explore the influence of the contact angle on the relaxation pressure independently, the same fluid (water) and the same density screen components 325 × 2300 with different surface treatments was used in experiments. Repeat the experiment at least three times for each data point. Figure 3a indicates that the bubble point pressure decreases with the increase of contact angle, while the reseal pressure is less sensitive to the changes of contact angle. Besides, as shown in Fig. 3b, that a larger contact angle leads to a higher relaxation pressure. Smaller value of the contact angle indicates better hydrophilicity of the screen [16], which results in a quicker process for the screen to return to the sealing state. Therefore, the corresponding relaxation pressure is smaller with better hydrophilicity and enhancing the resealing performance. As a result, the smaller contact angle results in a better resealing performance. Improving the wettability of the screen is an effective means to enhance its self-healing ability.

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Fig. 3. Effect of contact angle on (a) bubble point pressure and reseal pressure and (b) relaxation pressure

3.2 Effect of Fluid Properties In order to comprehensively consider the effects of fluids and screens, a wicking parameter were defined, which related to surface tension, contact angle and viscosity [17]. α = (σ cosθ )/μ

(3)

As shown in Fig. 4, the relaxation pressure is inversely proportional to the wicking parameter as defined in Eq. (3). When the wicking parameter of the fluid is larger, it means that the capillary action is more beneficial to its performance and results in a better self-healing performance. This finding can be applied to different combinations of fluids and meshes as discussed in this study. The fitting curve function between the capillary parameter and the relaxation pressure is: ∗ = 2.19/ (α + 2.65) Prel

(4)

3.3 Effect of Mesh Specification The influence of mesh specification on the self-healing performance is discussed in this section with water as the test fluid. It can be seen from Fig. 5a that, within the error range, both the bubble point pressure and reseal pressure increase with the increase of mesh density, also with, the decrease of the pore diameter. Besides, the relaxation pressure of these five screen densities (40 × 430, 100 × 850, 130 × 1100, 203 × 1600, 325 × 2300) fluctuate in the range of 0.1 to 0.6, as shown in Fig. 5b. The increase of bubble point pressure makes the screen’s phase separation performance more applicable. Meanwhile, the increase of relaxation pressure indicates that the screen’s self-healing performance becomes worse during the dynamic operation.

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Fig. 4. Effect of wicking parameter on relaxation pressure. Mesh specification: 325 × 2300 [14]

Fig. 5. Effect of mesh density on (a) bubble point pressure and reseal pressure and (b) relaxation pressure

4 Conclusion In this paper, experiments were conducted to investigate the effects of wicking parameter and mesh type on resealing performance. A relaxation pressure is defined as the difference between the reseal pressure and the bubble point pressure divided by the bubble point pressure to describe the self-healing performance of the screen. The experiment results show that: (1) Better surface wettability strengthens the capillary flow of fluid in the porous screen. Therefore, the bubble point pressure increases and the relaxation pressure decreases with the decreasing contact angle. (2) For the same density of the porous screen, the relaxation pressure is inversely correlated with the wicking parameter (α = (σcosθ)/μ) that mainly influenced by fluid properties. For different fluids, when the wicking parameter is larger, the self-healing performance of screen is better.

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(3) Reducing the mesh density can achieve better self-healing performance, but at the same time, it may have a greater impact on other performance like bubble point pressure. Therefore, in actual situations, the choices of the mesh density needs to be considered comprehensively. It should be noted that, as a preliminary work, the experiments in this study are conducted with room temperature fluids. For cryogenic flow in practical applications, more effects should be considered, such as temperature gradient, thermal expansion, heat transfer, etc. Furthermore, more experiments with cryogenic fluids will be conducted in the future to further improve the present results. Acknowledgments. The project is supported by National Natural Science Foundation of China (Nos. 52276013 and 51936006).

References 1. Street, D., Wilhite, A.: A scalable orbital propellant depot design. Georgia Institute of Technology, April 2006 2. Tegart, J.R., Aydelott, J.C.: Effect of vibration on retention characteristics of screen acquisition systems. J. Spacecr. Rocket. 16(5), 319–325 (1979) 3. Dominick, S.: Orbital test results of a vaned liquid acquisition device. In: The 30th Joint Propulsion Conference, vol. 6, pp. 27–29 (1994) 4. Wang, Z., et al.: A three-dimensional flow model of screen channel liquid acquisition devices for propellant management in microgravity. npj Microgravity 8, 28 (2022) 5. Hartwig, J., Chato, D., McQuillen, J.: Screen channel LAD bubble point tests in liquid hydrogen. Int. J. Hydrogen Ener. 39(2), 853–861 (2014). USA 6. Hartwig, J., Adin Mann, J., Darr, S.R.: Parametric analysis of the liquid hydrogen and nitrogen bubble point pressure for cryogenic liquid acquisition devices. Cryogenics 63, 25–36 (2014). Guildford 7. Hartwig, J.W., et al.: A steady state pressure drop model for screen channel liquid acquisition devices. Cryogenics 64, 260–271 (2014). Guildford 8. Kudlac, M.: Screen channel liquid acquisition devices for liquid oxygen. In: Proceedings of the 42nd Joint Propulsion Conference and Exhibit, Sacramento, USA (2006) 9. Jurns, J., Mcquillen, J.: Liquid acquisition device testing with subcooled liquid oxygen. In: Proceedings of the 44th Joint Propulsion Conference and Exhibit, Hartford, USA (2008) 10. Jurns, J., Mcquillen, J.B., Gaby, J.D., et al.: Bubble Point Measurements with Liquid Methane of a Screen Channel Capillary Liquid Acquisition Device. Cleveland, USA (2008) 11. Meserole, J.S., Jones, O.S.: Pressurant effects on cryogenic liquid acquisition devices. J. Spacecr. Rocket. 30(2), 236–243 (1993) 12. Savas, A.J., Hartwig, J.W., Moder, J.P.: Thermal analysis of a cryogenic liquid acquisition device under autogenous and non-condensable pressurization schemes. Int. J. Heat Mass Transf. 74, 403–413 (2014) 13. Hartwig, J., Yasuhiro, K.: The static reseal pressure model for cryogenic screen channel liquid acquisition devices. Int. J. Heat Mass Transf. 99, 31–43 (2016) 14. Wang, Y., Wang, Z., Cheng, X., Yang, G., Wu, J.: Pressure-driven phase separation based on modified porous mesh for liquid management in microgravity. Langmuir 38(9), 2919–2927 (2022)

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15. Conrath, M., Dreyer, M.: Gas breakthrough at a porous screen. Int. J. Multiphase Flow 42, 29–41 (2012) 16. Li, Y., He, S., Xu, Z., et al.: Investigation on the intrinsic wetting thresholds of liquids by measuring the interaction forces of self-assembled monolayers. Nano Res. 15, 4344–4349 (2022) 17. Wang, Y., et al.: Flow physics of wicking into woven screens with hybrid micro-/nanoporous structures. Langmuir 37(7), 2289–2297 (2021)

Investigation on Bubble Departure Behavior of Liquid Oxygen Under Microgravity Mingkun Xiao, Guang Yang, Chunyu Li, Yonghua Huang, and Jingyi Wu(B) Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai 200240, China [email protected] Abstract. The investigations on the characteristics of single bubble growth and detachment are of great importance to reveal the mechanism of the flow and heat transfer of liquid oxygen under microgravity. In this paper, the coupled level set and volume of fluid (CLSVoF) method is employed in OpenFOAM to capture gas-liquid interface. The results have shown that the degree of wall superheat has great effects on the bubble growth rate, while that of liquid subcooling is only consequential under microgravity. The bubble departure diameter and frequency of the oxygen are proportional to g −1/3 and g 1/2 , respectively. The degrees of wall superheat and liquid subcooling both have uniform impact on the bubble departure behaviors in different gravity levels, whereas that of liquid subcooling is only influential under microgravity on bubble departure diameter. Keywords: Liquid oxygen fluid dynamics

1

· Microgravity · Bubble · Computational

Introduction

Liquid oxygen is significant for cryogenic propellants with large specific impulse in deep space exploration. Due to the heat flux in space, it is vital to study the growth and departure of isolated bubble boiling of liquid oxygen in microgravity for storing and using cryogenic propellants. Investigations on dynamic parameters of the bubble are undertaken, such as bubble growth rate, bubble departure diameter and frequency [1–6]. Bubble growth rate denotes the change of bubble diameter over the time during bubble growth [7]. Bubble departure diameter is the equivalent diameter of the bubble after it detaches from the heating surface. Van Stralen and Zijl [8] and Gorenflo et al. [9] pointed out the relationship between the bubble departure diameter Dd and the gravitational acceleration, namely Dd ∝ g −1/3 . As for bubble detachment frequency, Cole [3] and Ivey [4] proposed the relation of f Ddn ∝ g 1/2 to reflect the impact of gravitational acceleration on bubble detachment frequency. Basically, most of these correlations focus on normal temperature fluids. As there are few studies about cryogenic fluids [10,11] on pool boiling in isolated bubble c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 907–913, 2023. https://doi.org/10.1007/978-981-99-6128-3_117

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regime, investigations are deficient in liquid oxygen under microgravity, which further illustrates the difficulties of liquid oxygen experiments under microgravity. Hence, numerical simulation has ascended as a crucial approach to undertake the research on the growth and detachment behaviors of single bubble pool boiling of liquid oxygen under microgravity. In this paper, CLSVoF method [12] has been employed to ensure the conservation of the mass and inhibit the diffusion of the interface. This method has been assessed in the previous work and found to be accurate for cryogenic two-phase flows [13]. This paper focuses on the growth and departure behaviors of liquid oxygen on single bubble nucleate pool boiling, and concentrates on the bubble growth rate, bubble departure diameter and frequency. Numerical models are constructed from OpenFOAM [14].

2

Modeling Two-Phase Flow

CLSVoF method regards the two fluids as the single fluid and describes them with phase fraction and signed distance function. The VoF phase equation is solved initially: ∂α m ˙ + ∇ · (Uα) + ∇ · [Uc α(1 − α)] = (1) ∂t ρl where U is velocity, t time, Uc = min (cα |U|, |U|max ) ∇α/|∇α| compression velocity and cα compression factor. The phase fraction α is between 0 and 1, which in the present work represents the gas and liquid phases, respectively, and m ˙ is the net mass flux rate represented by the Tanasawa model [15] ⎧ m ˙ =m ˙ c−m ˙e ⎪ ⎪   ⎪  ⎪ ⎨ ρg hv (T −Tsat ) 2γe M m ˙ e = 2−γe 2πR T > Tsat 3/2 Tsat   ⎪  ⎪ ⎪ ρg hv (Tsat −T ) ⎪ 2γc M ⎩m ˙ c = 2−γ T ≤ Tsat 3/2 2πR c

(2)

Tsat

where γe and γc are the evaporation and condensation coefficients, respectively, M molecular weight, R universal gas constant, T temperature, Tsat saturation temperature, ρ density, and hv latent heat. The subscript l and g represents liquid and gas, respectively. After advection of α, level set equation is employed to solve φ: φ0 = (2α − 1) · Γ (3) Γ = 0.75Δx ∂φ

∂τ = S (φ0 ) (1 − |∇φ|) φ(x, 0) = φ0 (x)

(4)

where φ is level set function, Δx mesh step size, τ = 0.1Δx artificial time step, S signed function. Eq. (3) is the initialization of φ, and Eq. (4) the solution of φ. Subsequently, Delta function is calculated by φ:

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0 δφ = 1  2ε

 1 + cos

πφ ε



|φ| > ε |φ|  ε

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(5)

where ε = 2Δx is interface thickness parameter. Delta function is employed for the calculation of surface tension force: ˆ φ (ˆ Fσ = [σκφ ∇φ + σT (∇T − n nφ · ∇T )) ∇φ] δφ

(6)

where Fσ is surface force, σ surface tension coefficient, σT surface tension coefˆ φ = ∇φ/|∇φ| interface unit ficient of temperature, κφ interface curvature, n normal vector. In Eq. (6), the Continuum Surface Force (CSF) model is employed to handle the surface tension [16]. The tangential surface force is added into Eq. (6) to realize thermocapillary convection at non-isothermal conditions.

3 3.1

Numerical Investigation Numerical Setup

In this paper, an axisymmetric region is adopted, and a nucleation site with the radius R = 0.1mm is located at the bottom center of the region, and the diameter and height of the region are both 40R. The computational area is shown in Fig. 1. The mesh convergence study has been undertaken and ultimately a mesh number of 200,000 is adopted for subsequent calculation. The bottom of the region is the heating surface with constant wall temperature measured by the Jacob number, which is a dimensionless number reflecting the degree of wall superheat, Ja =

ρl cp,l ΔT ρg hv

(7)

where cp is specific heat capacity at constant pressure, ΔT = Tw − Tsat , and Tw is wall temperature. Farstream g Axis LOX

z

Heating wall

r

Superheat layer

Nucleation site

Fig. 1. Schematic diagram of simulation cases

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Bubble Growth Rate

The bubble growth rate reflects the velocity of the motion of gas-liquid interface. It can be concluded from the results in Fig. 2 that the greater the degree of wall superheat, the faster the bubble growth, and the less the bubble departure time, since the hotter wall contributes to the higher heat flux rate at gas-liquid interface. For the same gravity level, the bubble departure diameter declines with the decrease of the Jacob number, whereas this trend is not obvious in Fig. 2. Different gravity levels are diminutively influential on the growth rates of bubbles, and the effects are only evident at low degree of wall superheat due to the long growth time, but they are crucial factors for the bubble departure diameter and the growth time. And the higher the microgravity level, the larger the bubble and the more the bubble detachment time.

Fig. 2. Bubble growth rates under different gravity levels and the Jacob numbers

Different degrees of liquid subcooling are also considered as pictured in Fig. 3. In ground gravity, it has small effects on the growth rates of bubbles, but when the gravity decreases, the difference of bubble growth rates becomes gradually distinct, since in the low gravity level the bubble always gets larger, the longer bubble growth time leads to the more immersion time that the part of the bubble far from the heating wall is immersed in the subcooling liquid, and consequently, the degree of liquid subcooling carries much more weight. What is noteworthy is that the change is more enormous from the saturation to subcooling at 1K than from that at 1K to 2K, which also contributes to the long immersion time. 3.3

Bubble Departure Diameter and Frequency

The simulation results of bubble departure diameter are listed in Fig. 4. There is intense competition between the surface tension and gravity. As gravity level has no influence on the surface tension, the surface tension over gravity turns greater with the decrease of the gravity. In other words, microgravity environment leads to the huge bubble and consequently the large detachment diameter. Furthermore, the differences of the effects of the degrees of wall superheat and liquid

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(a) g/g0 = 1

(b) g/g0 = 0.5

(c) g/g0 = 0.1

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(d) g/g0 = 0.05

Fig. 3. Bubble growth rates under different gravity levels and degrees of liquid subcooling

subcooling on bubble departure diameter lie in their change with the gravity. The degree of the wall superheat has a uniform impact for desperate gravity levels, while that of liquid subcooling has significant effects especially under high microgravity level. The possible reason is that the superheat layer is close to the heating wall and more sensible to the degree of wall superheat than that of liquid subcooling. Basically, bubble detachment diameter of liquid oxygen under different degrees of liquid subcooling and wall superheat is proportional to the -1/3 power of gravitational acceleration, which corresponds with the results in Ref. [8,9].

(a) Influence of Ja

(b) Influence of the degree of liquid subcooling

Fig. 4. The change of bubble departure diameter with various gravity levels

Ultimately, the bubble departure frequency is discussed here, as illustrated in Fig. 5. The greater the gravity level, the higher the bubble departure frequency, and the larger the bubble grows. Since surface tension is in a leading stage under microgravity which prevents the bubble from detaching from the heating wall, the microgravity environment gives rise to the difficulty of the departure of the bubble, bringing about the longer bubble growth time and lower bubble departure frequency. Nevertheless, different from bubble departure diameter, the

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degrees of wall superheat and liquid subcooling both have uniform effects on the bubble departure frequency regardless of the gravity level, which further explains the even impact of the superheat layer. And the difference lies in the distinct change rates of bubble departure frequency between the degree of wall superheat and liquid subcooling in a specific gravity level. The results also show that there is a 1/2 power of proportional relationship between bubble departure frequency times diameter and the gravitational acceleration. The results of the power of 1/2 are corresponded with that in Ref. [3,4].

(a) Influence of Ja

(b) Influence of the degree of liquid subcooling

Fig. 5. The impact of gravity levels on bubble departure frequency

4

Conclusion

In general, CLSVoF method have been incorporated into OpenFOAM in this study, and investigations have been undertaken on the single bubble pool boiling of liquid oxygen, and analyses have been conducted through the characteristics of the bubble growth and departure. The results can be then concluded, 1. The larger degree of wall superheat and lower degree of liquid subcooling contribute to the greater bubble growth rate and a longer detachment time. The degree of liquid subcooling only has an enormous impact on the bubble growth rates under microgravity. 2. The greater the microgravity level, the larger the bubble departure diameter, and the lower the bubble departure frequency. 3. The bubble departure diameter has an even change with the Jacob number under different gravity levels. On the contrary, no obvious distinction is there in the bubble departure diameter under ground gravity for various degrees of liquid subcooling. 4. The bubble departure diameter of the oxygen is proportional to g −1/3 , and bubble detachment frequency is in direct proportion to g 1/2 . Acknowledgments. This project is funded by National Natural Science Foundation of China (No. 51936006).

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References 1. Sakashita, H., Ono, A.: Boiling behaviors and critical heat flux on a horizontal plate in saturated pool boiling of water at high pressures. Int. J. Heat Mass Transf. 52(3–4), 744–750 (2009) 2. Miglani, A., Joo, D., Basu, S., Kumar, R.: Nucleation dynamics and pool boiling characteristics of high pressure refrigerant using thermochromic liquid crystals. Int. J. Heat Mass Transf. 60, 188–200 (2013) 3. Cole, R.: Bubble frequency and departure volumes at subatmospheric pressures. AIChE J. 13(4), 779–783 (1967) 4. Ivey, H.J.: Relationships between bubble frequency, departure diameter and rise velocity in nucleate boiling. Int. J. Heat Mass Transf. 10(8), 1023–1040 (1967) 5. Peebles, F.N., Garber, H.: Study on motion of gas bubbles in liquids. Chem. Eng. Prog. 49, 88–97 (1953) 6. Zuber, N.: Nucleate boiling the region of isolated bubbles and the similarity with natural convection. Inter. J. Heat Mass Transfer 6(1), 53–78 (1963) 7. Mohanty, R.L., Das, M.K.: A critical review on bubble dynamics parameters influencing boiling heat transfer. Renew. Sustain. Energy Rev. 78, 466–494 (2017) 8. van Stralen, S.J., Zijl, W.: Fundamental developments in bubble dynamics. In: Proceedings of the 6th International Heat Transfer Conference, pp. 429–450 (1978) 9. Gorenflo, D., Knabe, V., Beiling, V.: Bubble density on surfaces with nucleate boiling-its influences on heat transfer. In: Proceedings of the 8th International Heat Transfer Conference, pp. 1995–2000 (1986) 10. Zhang, X., Chen, J., Xiong, W., Jin, T.: Visualization study of nucleate pool boiling of liquid nitrogen with quasi-steady heat input. Cryogenics 72, 14–21 (2015) 11. Zhang, X., Xiong, W., Chen, J., Wang, Y., Tang, K.: CFD simulations and experimental verification on nucleate pool boiling of liquid nitrogen. Phys. Procedia 67, 569–575 (2015) 12. Yamamoto, T., Okano, Y., Dost, S.: Validation of the S-CLSVOF method with the density-scaled balanced continuum surface force model in multiphase systems coupled with thermocapillary flows. Int. J. Numer. Meth. Fluids 83(3), 223–244 (2017) 13. Xiao, M., Yang, G., Huang, Y., Wu, J.: Evaluation of different interface-capturing methods for cryogenic two-phase flows under microgravity. Phys. Fluids 34(11), 112124 (2022) 14. Openfoam, the open source CFD toolbox (2020). http://www.openfoam.com/ documentation 15. Tanasawa, I.: Advances in condensation heat transfer. In: Hartnett, J.P. (ed.) Advances in Heat Transfer, vol. 21, pp. 55–139. Academic Press, San Diego (1991) 16. Brackbill, J.U., Kothe, D.B., Zemach, C.: A continuum method for modeling surface tension. J. Comput. Phys. 100(2), 335–354 (1992)

Thermal Design and Verification of Low-Temperature Storage Device for China Space Station Shuai Wang1,2(B) , Fankong Meng1,2 , Yufeng Fan1,2 , Meng Xiao1,2 , Bo Wang3,4 , Yin Zhang3,4 , Yawei Xu1,2 , Huning Yang1,2 , Changpeng Yang1,2 , Jianyin Miao1,2 , and Hongyang Zheng1,2 1 Beijing Institute of Space System Engineering, Beijing 100094, China

[email protected]

2 Beijing Key Laboratory of Space Thermal Technology, Beijing 100094, China 3 Anhui Province Key Laboratory of Low-Temperatures Technology, Hefei 230088, China 4 16th Institute of China Electronics Technology Group Corporation, Hefei 230088, China

Abstract. The low-temperature storage device of China Space Station is the first large-capacity and multi temperature zone low-temperature storage facility in Chinese Space Field. It consists of −80 °C, −20 °C and + 4 °C storage areas and uses both Stirling cryocooler and thermoelectric coolers, with vacuum insulation panel and polyurethane foam being used for heat insulation, and fluid circuit being used for heat dissipation. This paper will present the thermal design and verification of the low-temperature storage device. Keywords: CSS · Low-Temperature Device · Thermal-Design · Verification

1 Introduction Conducting experimental studies in manned spacecraft is a key approach to the development of space life science and biotechnology. In order to meet the on-orbit storage requirements of experimental samples, extensive efforts had been conducted on the research and development of low-temperature storage devices in space. For example, the International Space Station is equipped with storage devices of different specifications and temperature regions, with a maximum storage capacity of up to 300 L and a temperature range from room temperature to − 60 °C [1, 2]. China started late in this field. Small low-temperature storage devices are equipped on ShenZhou spacecraft and TianZhou cargo spacecraft, with a minimum temperature of −20 °C, which are mainly used for short-term storage of medical samples. According to China Space Station program, a low-temperature storage device will be equipped in the WenTian experimental module of the China Space Station [3], which can provide three independent sample storage areas of −80 °C, −20 °C and +4 °C, with the specific structures being shown in Fig. 1. The low-temperature storage device occupies 4 SPUs space in the laboratory cabinet, with a launch mass of about 120 kg and a power dissipation less than 500 W. Table 1 shows the main thermal performance indexes. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 914–920, 2023. https://doi.org/10.1007/978-981-99-6128-3_118

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Fig. 1. Front of low-temperature storage device

Table 1. Main thermal performance indexes of the low-temperature storage device Parameters

Box 1

Box 2

Box 3

Temperature

–80 °C

–20 °C

+4 °C

Temperature uniformity of cold box

At least three temperature measuring points shall be set at the untypical heat source in each low-temperature area

Net refrigeration capacity

≯ ± 5 °C

≯ ± 2 °C

≯ ± 1 °C

≮20 W

≮25 W

≮20 W

2 Thermal Design of the Low-Temperature Storage Device In order to reduce the complexity and improve the reliability of the system, the three storage areas of the Low-temperature storage device are designed independently. 2.1 Refrigeration Design A small high-efficiency Stirling cryocooler is used as the cold source for the temperature area of –80 °C. It consists of compressor, expander, cold ring, vibration absorber and so on, as Fig. 2 shown. It has an enveloping size of 310 mm × 140 mm × 180 mm, a mass less than 8 kg, and a maximum input power of 180 W, the refrigeration efficiency can reach as high as 30%. In order for high-efficiency cold conduction, it is installed by a heat spreader coupling plate in contact with the inner liner of the storage area at −80 °C, and the contact interface is filled with 0.1 mm-thick indium foil, see Fig. 2. Thermoelectric coolers are used as the cold source of + 4 °C and −20 °C areas, which have advantages in terms of volume and weight. Four pieces of thermoelectric

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Fig. 2. Stirling cooler and cold box connection

coolers are used for each temperature area (see Fig. 3), power supply is independent for each piece, with a maximum input power of 70 W. The hot end interface of the TEC is filled with silicon grease, with a heat transfer coefficient up to 4000 W/ (m2 ·K), and the cold end interface is filled with flexible thermal pad.

Fig. 3. Thermoelectric cooler and cold box connection

2.2 Insulation Design The composite insulation scheme of vacuum insulation panel and polyurethane foam is adopted. Integrated polyurethane foaming is conducted after the vacuum insulation panel (see Fig. 4) is pasted between the inner surface of the shell and the outer surface of the inner liner. The equivalent thermal conductivity of the thermal insulation component is about 0.014 W/(m·K). 2.3 Internal Heat Dissipation Design The space station fluid circuit system is used to dissipate heat to ensure that the refrigerator and cooler in operation is in the optimal temperature range. The cooling fluid

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Fig. 4. Vacuum Insulation Panel

temperature ranges from 20 °C to 26 °C and the flow rate is 60 L/h. Figure 5 shows the fluid circuit system.

Fig. 5. Fluid circuit system

3 Thermal Performance Verification In order to verify the thermal performance of the low-temperature storage device, a test was conducted in a high-low temperature chamber. The boundary conditions were set as follows: temperature: 23 °C, cooling liquid temperature: 23 °C, and flow rate: 60 L/h. Figure 6 show the composition of the test system and the test scene. Four thermistors were designed in each storage area for temperature measurement and closed-loop control. In addition to the thermistors, thermocouples were arranged for specific temperature measurement. The net refrigeration capacity was tested by setting simulated experimental sample (300 mL water) and pasting the electric heater in each temperature area. 3.1 Cooling Performance Test During the cooling process, real-time monitoring was conducted on the average temperature of each storage area and the cooling performance was evaluated.

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Fig. 6. State of test system

The cooling period of the −80 °C area was from 21:47 pm October 28, 2021 to 06:54 am October 29, 2021, with a total of 547 min and the cooling rate was 5.36 min/°C; The cooling period the −20 °C area was from 06:58 am October 28,2021 to 08:38 am October 29, 2021, with a total of 100 min and the cooling rate was 2.38 min/°C; The cooling period in the +4 °C area was from 08:36 am October 29, 2021 to 09:12 am October 29, 2021, with a total of 36 min and the cooling rate was 2 min/°C. The maintaining power was 101.7 W (−80 °C), 128.2 W (−20 °C) and 13.7 W (+4 °C) respectively, and the maintaining power consumption of the whole device was about 310 W, which met the requirement of index. 3.2 Temperature Uniformity Test After temperature equilibrium of each temperature area, the temperature uniformity of the four thermistors was calculated, which was respectively 3 °C (–80 °C), 1.5 °C (– 20 °C) and 1.25 °C (+4 °C). The following figures show the temperature curves of 4 temperature measuring points in each temperature area after equilibrium (Fig. 7). 3.3 Net Refrigeration Capacity Test Thin-film heater was pasted on the surface of inner linear in order for refrigeration capacity test, as shown in the following figure. The net refrigeration capacity tested was 27.5 W(–80 °C), 27 W(–20 °C) and 43.3 W(+4 °C) respectively (Fig. 8).

Thermal Design and Verification of Low-Temperature Storage Device

Fig. 7. Temperature at different positions of each box

Fig. 8. Net refrigeration capacity test

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4 Conclusion The test results show that the thermal performance of the Low-temperature storage device meets the technical requirements. However, under the influence of the ground gravity, the natural convection effect is conducive to the internal cooling of the storage area, which is different from that under the on-orbit microgravity environment. Therefore, it requires further simulation analysis and on-orbit test.

References 1. NASA. Space Station Research Experiments. http://www.nasa.gov/mission_pages/station/res earch/experiments_category/index.html. Accessed 31 Dec 2014 2. Zabel, P., Bamsey, M., Schubert, D.: Review and analysis of over 40 years of space plant growth systems. Life Sci. Space Res. 10, 1–16 (2016) 3. Gu, Y.D., Gao, M., Zhao, G.: Space research plan of China’s Space Station. Chin. J. Space Sci. 36(5), 595–599 (2016) 4. Robinson, J.A., Thumm, T.L., Thomas, D.A.: NASA utilization of the international space station and the vision for space exploration. Acta Astronaut. 61(1–6), 176–184 (2007) 5. Colangelo, G., De Parolis, M. N., Jimenez, J.: MELFI Cooling Performance Characterization and Verification, ICES ñ Toulouse, France, July 2000

Modeling of Pressure Reducing in a Cryogenic Tank by a Thermodynamic Vent System Considering Flashing Xujin Qin, Chuiju Meng, and Yonghua Huang(B) Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai 200240, China [email protected]

Abstract. Thermodynamic vent system (TVS) is a promising approach for controlling pressure propellants in cryogenic tanks under microgravity conditions. Prior simulations for the pressurization of cryogen in the tank with TVS basically treated the liquid as quasi-static by the venting operation. However, experiments showed that the liquid could be superheated during that period. The present work will present a model for the passive TVS for one-dimensional cryogenic tanks. Flash evaporation correlation is adopted for calculating the phase change between the superheated liquid and the vapor. The predicted pressure by the model that considers flash evaporation agrees with the experimental data better than that without considering the flashing phenomenon. The evaporation rates in the cases at different fill levels are compared based on the model. It is found that the flashing has a significant impact on the cycle period and mass mean loss rate prediction. Keywords: Thermodynamic Vent System · Flash Evaporation · Cryogenic Storage · Modeling

1 Introduction Long-term storage of cryogenic propellants in orbit will be a competitive solution in space exploration. One of the major challenges for storing cryogenic propellants in orbit is to guarantee venting vapor only for the purpose of reducing mass loss. Thermodynamic vent systems (TVS) have been considered a promising method for pressure control under such a condition. Simulations have been conducted in literatures on the pressure control performance of thermodynamic vent systems. Hastings et al. [1] developed a transient analytical model for the TVS installed in the multipurpose hydrogen test bed (MHTB). Another analytical model was developed by Wang et al. [2] for predicting the pressure variation of a TVS. The effects of pressure control band on cycle durations, running numbers and vapor mass loss was discussed. More recently, Majumdar et al. [3] considered thermal diffusion in the ullage and improved the model for the MHTB by treating the ullage as a series of nodes instead of a single lumped node. Zuo et al. [4] proposed a computational fluid model for the TVS that sits in a 3-D tank and connects an external © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 921–928, 2023. https://doi.org/10.1007/978-981-99-6128-3_119

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fluid circulation loop, a throttling device, and the mixing and jetting components. They found that during the complete thermodynamic venting process, the depressurization efficiencies decline significantly when the bulk liquid is nearly saturated. Unfortunately, the previous studies only considered mild evaporation on the liquidvapor interface either in the simulation or the experimental data analysis. Flash evaporation could happen when the liquid reaches superheat state during depressurization. This liquid-vapor transition behaviors differ significantly from mild evaporation, which calls for an improved model considering the flashing phenomenon into the system to better predict the tank pressure.

2 Physical and Numerical Model for a TVS Equipped Tank The numerical model is based on the mulitifunctional cryogenic test system [5], which consists of a storage tank and a passive TVS. The heat flux across the tank wall pressurises tank constineously. When the pressure gets higher than the setted upper limit, the TVS is activated. The TVS extract a small stream of the liquid to the J-T valve, where the liquid expands to a colder two-phase flow. The two-phase flow absorbs heat from the rest liquid and ullage in the heat exchanger. Thus, the pressure is reduced. The schematic of the TVS is shown in Fig. 1. Venting valve

Stratified wall 1 2

Heat exchanger

Ullage

Stratified ullage

… i … n

Liquid

1 2 … i … n

Bulk liquid

J-T valve

Fig. 1. Schematic of the TVS

Fig. 2. Schematic of the physical model of the tank containing a two-phase fluid

A schematic of the tank model is shown in Fig. 2. The tank is a cylinder with a height of 0.9 m and a diameter of 0.6 m. The thickness of the tank wall is 4 mm. The primary consumptions of the model are as follows: (1) Both the ullage and the wall are thermally stratified. (2) The free interface is a thin layer at a saturated temperature. (3) The temperature and pressure of the liquid is uniform. The Tank Wall. The wall is divided into two parts. The first part contacts the ullage, the second part contacts the liquid. The energy equation of the wall are given by: Twu(i+1,j)j) − Twu(i,j) Twu(i,j+1) + Twu(i,j−1) − 2Twu(i,j) φπ Dz − q˙ w,u(i,j) + = aw t (z)2 ρw cpw π Dδz

(ullage part)

(1)

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  1 Tw,l(i + 1) − Tw,l(i) Tw,u(i,n) − Tw,l(i) = π Dδλ − q ˙ (liquid part) w w,l(i) t ρw cpw (π DδHliq + π D2 δ/4) z

(2) where T w,u and T w,l are the temperatures of the wall contacting the ullage and the liquid, respectively; ρ w , aw , δ, cpw , and λw are density, thermal diffusion coefficient, thickness, isobaric specific heat capacity, and thermal conductivity of the wall, respectively; H liq is the liquid level; D is inner diameter of the tank; ϕ is heat flux from environment to the tank wall; z is axial distance; t is time; i and j stand for the time step and the axial direction node, respectively; q˙ w,u and q˙ w,l are the heat transfer rate between the wall and the ullage, and between the wall and the liquid, according to the correlations for natural convection on a vertical column [6]. The Ullage. The energy equations of the thermal diffusion of the ullage are given by: Tu(i + 1,j) − Tu(i,j) Tu(i,j + 1) + Tu(i,j - 1) − 2Tu(i,j) q˙ w,u(i,j) − q˙ TVS,u(i.j) = au + 2 t (z) ρu cpu π D2 z/4

(3)

where ρ u , T u , cpu and au are the density, temperature, isobaric specific heat capacity, and thermal diffusion coefficient of the ullage, respectively; q˙ TVS,u(i,j) is the cooling capacity provided by the TVS to the ullage nodes. The pressure is calculated using the average temperature and density of the ullage: Pu = P(T u , ρ u ), T u =

1 n

j=1  n

Tu(j) , ρ u =

mu Vu

(4)

The Liquid. The liquid is homogeneous as mentioned in the assumptions. The energy equations are as follows: q˙ w,l + q˙ int,l − q˙ TVS,l − (m dTl ˙ lu − m ˙ cond )γ − m ˙ vent hl = t ml cpl

(5)

where hl , cpl and T l are the enthalpy, isobaric specific heat capacity, and temperature of the liquid. γ is the heat of vaporization; q˙ TVS,l is heat rate that is extracted from the ˙ vent is the mass venting rate. liquid by the TVS; ml is the mass of the liquid; m The Liquid-Vapor Interface. Mild evaporation is considered on the free interface. The mass transfer rate is written as [3]: m ˙ lu =

q˙ u,int −˙qint,l , γ

Pliq ≥ Psat

(6)

where qu,int and qint,l is the heat received from the vapor and the heat rejected to the liquid, respectively. quint = Aint αu,int (Tu(n) − Tint ), qintl = Aint αint,l (Tint − Tl )

(7)

where Aint is the interface area; The heat transfer coefficients α u,int and α int,l are calculated using correlations for natural convection on a vertical column [6]; T u(n) is the temperature of the ullage adjacent to the interface. T int and T l are the temperature of the interface and the liquid, respectively.

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The Flash Evaporation. During the TVS venting process, the liquid becomes superheated once the pressure drops below the saturation pressure at the local temperature, which meets the condition for flash evaporation. For the purpose of simplification, the trigger condition of flashing is set to Pliq < Psat . The flash evaporation rate is calculated by an empirical correlation proposed by Watanabe [7]: sat m ˙ lu = Cm T Tsat ml S, Pliq < Psat

(8)

where C m is a coefficient which Watanabe takes a value of 0.0375; T sat is the difference of saturated and liquid temperature. ml is the liquid mass. S denotes non-dimensional area of the flashing bubbles. Watanabe suggests that the bubble radii is proportional to square root of time t (i.e., bubble growth time), S = 1 + Cs · t

(9)

where C s = 20 is a coefficient. When t = 0, S = 1. It means evaporation occurs only on the free interface. The Thermodynamic Vent System. The TVS in the model is considered to run in a passive way. The throttling orifice has a diameter of 0.5 mm. The heat exchanger is a copper tube with a length of 0.9 m, an inner diameter of 4 mm and a thickness of 1 mm. The mass flow rate through the throttle is given by,  2 ˙ m =  Cε π dori 2ρl (Pliq − Pori ) Q 1 − β4 4

(10)

And the flow loss in the heat exchanger tube is, Pori − Pamb =

2 ˙m 8λLhex Q 5 π 2 ρori dhex

(11)

where C denotes the outflow coefficient of the orifice; ε is the expansion coefficient of the flow stream; ρ l and ρ ori are the density of the flow stream in the inlet and outlet of the orifice, respectively; L hex is the length of the exchanger tube. Dori and d hex are diameter of the orifice and hydraulic diameter of the heat exchanger, respectively; β is the ratio of diameters of the orifice and its inlet tube; Pori and Pamb are the pressure at outlet of the orifice and the ambient pressure; λ is the friction factor of the heat exchanger tube. The heat exchanger contacts both the ullage and the liquid, as shown in Fig. 3. For the part that is submerged under the liquid, q˙ TVS,l = Ahex,l (Tl − Tvent,l )

1 1 αin

+

1 αout

(12)

where α out andα in are convection coefficient outside and inside the heat exchanger tube, respectively. The characteristic lengths for calculating α in and α out are d hex and L hex , respectively. Ahex,l is the heat transfer area between the exchange tube and the liquid. T vent,l is the temperatures of the venting fluid in the liquid section. The venting fluid from the

Modeling of Pressure Reducing in a Cryogenic Tank Ullage

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xi+1 xi x1 x0

Liquid

Fig. 3. Schematic of the heat exchanger

orifice is assumed a liquid-vapor mixture. Therefore, T vent,l is equal to the saturated temperature at the orifice outlet pressure, namely T sat (Pori ). For the section surrounded by the vapor: q˙ TVS,u(j) = Ahex,u(j) (Tu(j) − Tvent,u(j) )

1 1 αin

+

(13)

1 αout

The quality of the liquid-vapor flow stream in the tube increases after absorbing heat: xi + 1 = xi +

q˙ TVS,u(j) γ Qm

(14)

where Ahex,u(j) is the heat transfer area between the exchange tube and the ullage nodes; x is the quality of the flow stream. T vent,u(j) is the temperature of the venting mixture in jth node of the heat exchange,  Tsat (Pori ), x T 1 and P2 = P1 , there will be a flow of superfluid from container A to the warmer one B. However, Eq. (6) is only suitable for ideal laminar flow. Turbulent flow is more complicated, and the mathematical model would be given in the next chapter.

3 Mathematical Model of Turbulent Flow When the flow is transformed into turbulent flow, vortex will have an obvious influence on the flow. The steady flow equation for superfluid is as follows: 0=−

ρs ∇p + ρs S∇T − Fsn ρ

(7)

ρs and ρ are normal and total fluid density, F sn is the friction term, which was proposed by Gorter and Mellink [9]: Fsn = AGM ρn ρs (vs − vn )3

(8)

where AGM is the Gorter-Mellink (G-M) constant, ρs and ρn are the superfluid and normal fluid density, vs and vn are the velocity of superfluid and normal fluid. For the superfilter, the normal component fluid is trapped causing vn = 0. For the mass flow through a superleak the expression S

1 dP dT − − ρn AGM vs3 = 0 dx ρ dx

(9)

follows from Eq. (7). However, it is inaccurate to use straightforward application of G-M constant to the porous element flow. Modification of porous element has been made as follows [10] (Fig. 2): d was predicted by Blake-Kozeny equation [11]:   (10) d 2 = 25.2K 1 − ε2 /ε3 where K is the permeability of the element. m can be expressed by:  1/3 ρs3 S m = ω−4/3 εAl −1/3 ∫TT21 dT ρn AGM

(11)

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Fig. 2. Effective flow model [10]

The effect of Q on m is indirect, it mainly affects the change of m by controlling the temperature T2 . To better understand the mechanism and characteristics of turbulent flow, unsteady flow should be studied. Taking into account the effect of evaporation loss in the initial stage and the hydrostatic pressure effect, the equations of mass and energy conservation are derived as follows [10]: 

 ρST − P 1/3 d ρv dT2 V − m = ρs Aeff ρρn AGM leff dT dt  1/3 −P Q − ST2 ρs Aeff ρST ρρn AGM leff dT2 = dt MC + λV ddTρv M = ∫t0 mdt

(12)

(13) (14)

4 Results and Discussion Calling the Hepak physical database, these two non-linear equations are numerically solved. During the numerical solution process, most thermodynamic quantities are evaluated at T2 except the initial stage. To validate the numerical solution, the results were compared with the experimental data [10]. As can be seen from Fig. 3, the numerical results were in good agreement with the experimental data both for the temperature T 2 and mass flow rate m. It verified the accuracy of the numerical model. To better illustrate the influence of some parameters including Q, m and geometric parameters, more numerical results were shown below.

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Fig. 3. Numerical results compared with experimental data (a) T2−t, (b) m−t

Fig. 4. T2 and m as a function of time t with different Q (a) T 2 −t (b) m−t

As can be seen from Fig. 4, when the heat input Q is different, the increment rate of the temperature T 2 at the initial moment is different, with a larger Q corresponding to a larger rise. When Q < Qmax , T 2 and m can eventually stabilize at a value. When Q = Qmax , m can stabilize at mmax eventually. Further increase in Q would lead to a continuous increase in temperature T 2 . When Q > Qmax , m would firstly increase to mmax and then decrease because of the superfluid turbulence. Despite heat input Q, the temperature of superfluid helium bath is also a key factor affecting efficiency of FEP. With a certain value of heat input Q = 0.15 W, several different superfluid helium bath temperatures are calculated in Fig. 5. With lower superfluid helium bath temperature, the change of T 2 would be greater, which is determined by the initial temperature and physical parameters. As for the mass flow rate m, the results are somewhat similar to those of ideal laminar flow. According to two-fluid model of superfluid helium, for lower temperature, there is more superfluid component compared with normal component. Obviously, the mass flow rate m would increase as the temperature goes down.

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Fig. 5. T2 and m as a function of time t with different T1 (a) T 2 −t (b) m−t

In order to illustrate the influence of the geometric parameters, different Aeff and L eff were set for numerical calculation. The results were shown in Fig. 6:

Fig. 6. T2 and m as a function of time t with different Aeff and Leff (a) T2-t with different Aeff (b) m-t with different Aeff (c) T2-t with different Leff (d) m-t with different Leff

It could be seen from Fig. 6 that Larger Aeff would lead to smaller T 2 and larger m. As for L eff , it had opposite effects on T 2 and m, which conformed to the law of normal fluid flow. From above analysis, T 1 , Q and geometric parameters including Aeff and L eff would influence temperature T 2 and mass flow rate m.

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5 Conclusions For the purpose of performing numerical analyses on unsteady flow of Helium II through a porous element, a modified model has been developed from the original two-fluid model. The validity of the model has been verified by comparing the numerical results with the past experimental data. For some special parameters, the analysis has been performed to demonstrate the capability of the present model applied in the other conditions with different Q and T 1 . Several conclusions were summarized as follows: 1. T 2 and m can eventually be stabilized at a certain value When Q ≤ Qmax ; m would reach maximum and drop next because of superfluid turbulence when Q > Qmax ; 2. With lower T 1 , T 2 and m would grow rapidly and increased more compared with the initial stage; 3. Larger Aeff and smaller L eff would lead to smaller T 2 and larger m.

References 1. Frederking, T., Yuan, S., Carandang, R.M.: Fountain effect pump phenomena for liquid helium transfer: thermodynamic system studies. Cryogenics 26(2), 93–96 (1986) 2. London, H.: Thermodynamics of the Thermomechanical Effect of Liquid He II. Proc. Roy. Soc. Lond. 171(947), 484–496 (1939) 3. Takamatsu, K., Fujimoto, N., Rao, Y.F., et al.: Numerical study of flow and heat transfer of superfluid helium in capillary channels. Cryogenics 37(12), 829–835 (1997) 4. Dalban-Canassy, M., Van Sciver, S.W.: Steady counterflow He II heat transfer through porous media. Am. Inst. Phys. 1327–1334 (2010) 5. Baudouy, B., Juster, F.P., et al.: Heat transfer through porous media in static superfluid helium. In: AIP Conference Proceedings (2006) 6. London, F.: Superfluids. Dover Publications (1961) 7. Andronikashvili, E.: A direct observation of two kinds of motion in Helium II. Helium 4, 154–165 (1971) 8. Allen, J.F., Misener, A.D.: Flow of liquid Helium II. Nature 141(3558), 75 (1938) 9. Gorter, C.J., Mellink, J.H.: On the irreversible processes in liquid Helium II. Physica 15(3–4), 285–304 (1949) 10. Murakami, M., Hanyu, T.: Thermomechanical flow of He II through a porous element for a fountain effect pump. Cryogenics 32(4), 371–378 (1992) 11. Bird, R.B., Stewart, W.E., Lightfoot, E.N.: Transport Phenomena. Wiley (2002)

Low-Temperature Thermal Conductivity Test of Aerogels Under the Full-Pressure Range Xiafan Xu1,3 , Jianpeng Zheng2 , Hao Xu1,3 , Liubiao Chen1,3,4(B) , and Junjie Wang1,3 1 CAS Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Beijing 100190, China [email protected] 2 Beijing Institute of Aerospace Control Devices, Beijing 100854, China 3 University of Chinese Academy of Sciences, Beijing 100049, China 4 Institute of Optical Physics and Engineering Technology, Qilu Zhongke, Jinan 250100, Licheng District, China

Abstract. Aerogel is a nano-scale porous material with low density, low thermal conductivity, and flame retardant, which has been widely used in the field of high-temperature insulation. In recent years, aerogel is also used for cryogenic insulation, but the related parameters of its thermal insulation performance at low temperatures are not sufficient. In this paper, a low-temperature thermal conductivity measurement device for aerogel with the controllable temperature of both cold and hot ends and vacuum degree has been developed based on a mechanical refrigerator. The temperature test range of the device is 20 K to 300 K, and the vacuum test range is 10–5 Pa to atmospheric pressure. The measurement principle and structure of the device will be introduced, and the test thermal conductivity/apparent thermal conductivity of aerogels under different conditions will also be given. Keywords: Aerogel · Thermal conductivity · Cryogenic insulation · Full-pressure range

1 Introduction Aerogel is a highly porous amorphous solid material with a nano-scale three-dimensional network framework structure [1]. Its unique characteristics such as large specific surface area, low density, low thermal conductivity, low refractive index, and low dielectric constant [2] make it suitable for high-temperature or low-temperature thermal insulation [3, 4], impurity adsorption [5], catalyst supports [6], novel rechargeable batteries and other frontier fields. When aerogel is used in the thermal insulation field, pure aerogel is mostly powder or small block with high brutality, poor compressive and flexural resistance, so it is usually required to be compounded with other fibers to make insulation materials with excellent performance [7]. NASA [8] took the lead in successfully applying aerogel to space exploration as an insulation material. The whole piece of SiO2 aerogel was used as the insulation material on the Mars lander ‘pathfinder’, and the aerogel was used to fill the foam core to © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 979–985, 2023. https://doi.org/10.1007/978-981-99-6128-3_127

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form a composite thermal insulation material for the storage of low-temperature propellants. R. Begag et al. [9] developed a composite thermal insulation material composed of aerogel and multi-layer thermal insulation for cryogenic propellant on-orbit storage, and the results show that thermal insulation performance was improved in high vacuum, soft vacuum, and no vacuum. Some scholars have also conducted relevant tests on the thermal properties of aerogel at low temperature, thermal conductivity, effective thermal conductivity, and other physical parameters are usually used to measure the thermal insulation performance. Fesmire et al. [10, 11] tested the effective thermal conductivity of aerogel blanket in different gas environments when the hot and cold boundary temperatures were 78 K and 293 K, respectively, and gave the relationship between the effective thermal conductivity and the cold vacuum pressure. Previous studies on the insulation performance of aerogel were mostly conducted in a high vacuum environment or under atmospheric pressure, and the study temperature range was usually above the liquid nitrogen temperature [10–12]. In order to make up for the lack of attention to the intermediate pressure region and the temperature region below the liquid nitrogen temperature, this paper has researched the thermal insulation properties of aerogel under different vacuum degrees, and the measurement device with controllable temperature and vacuum degrees has been established. A mechanical refrigerator is used as the cold source to control the cold end, and the temperature test range of the device is 20 K to 300 K. The vacuum degree is controlled by filling gas into the vacuum chamber, and the vacuum test range is 10–5 Pa to atmospheric pressure. Besides, the device can be filled with different gases (helium, nitrogen, argon, etc.) to test and compare the thermal conductivity of aerogel in different atmospheres. It should be noted that the device can also test the insulation performance of other similar shape insulation materials such as multilayer insulation and foam insulation in the full-pressure range.

2 Experimental System 2.1 Experimental Principle The thermal conductivity test principle of aerogel is based on Fourier’s law. The steadystate method which requires the establishment of stable temperature distribution within the sample is adopted in the test process. During the experiment, after the overall temperature of the sample drops to the temperature control point and reaches a steady state, a uniform heat flow is applied to the sample to form a stable temperature difference in the vertical direction of the sample. Formula 1 is used to calculate the thermal conductivity at different temperatures. Ks =

˙ Q A ∗ (TH − TC ) ∗ L

Ks − thermal conductivity of the aerogel, W/(m·K); A − the surface area of the aerogel, m2 ; L− thickness of the aerogel, m; TC , TH − cold/hot boundary temperature of the aerogel, K; ˙ heating power, W. Q−

(1)

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2.2 Experimental Device Figure 1 shows the photo of the thermal conductivity measurement device. The system consists of a refrigerator, vacuum chamber, temperature sensors, a vacuum pump unit, a DC power supply, data acquisition units, and a gas tank. Figure 1 also shows the arrangement of the samples inside the vacuum chamber. The heat sink is connected to the cold head of the refrigerator, and the “sample-heating film-sample” in a sandwich structure is stacked on the heat sink. In order to make full contact between the sample, the heat sink, and the heating film, put a box-shaped copper shield on the sample and fasten it with screws, as shown in Fig. 2. Simultaneously, 20 screws are used to ensure that the copper shield is in close contact with the heat sink so that the cold boundary temperature of the two samples is as same as possible. In order to reduce heat leakage, the copper shield is wrapped with multi-layer insulation materials. Through the test, the temperature difference between the cold head and the top center of the copper shield is less than 1.5 K, which is negligible. A high-frequency pulse tube refrigerator was used as the cold source to carry out the test in the temperature range of 77 K–300 K while the filling gas in the vacuum chamber is air. Considering that when the vacuum degree is adjusted by filling the air that contains water vapor into the vacuum chamber when the temperature is lower than the dew point temperature, water will slowly separate out, and it is undeniable that it will affect the test results. Simultaneously, the main components in the air are nitrogen and oxygen. When the test is performed below the liquid nitrogen temperature range, the air will be liquefied in large quantities. Therefore, when testing in the 20 K–77 K range, helium is selected to be injected into the vacuum chamber to regulate the pressure, and the Gifford–McMahon (GM) cryocooler is selected as the cold source.

Fig. 1. Photo of the thermal conductivity measurement device, and arrangement of sample.

In order to effectively regulate the pressure inside the vacuum shield, as shown in Fig. 2, a tee is connected to the vacuum chamber, one end is connected to a Pirani gauge for testing the vacuum degree, and the other end is the gas pipeline, which is connected to the gas tank (air, helium, nitrogen, etc.). A needle valve is set on the gas pipeline, and the valve is closed when there is no need to transport gas into the vacuum chamber; When it is necessary to deliver gas to adjust the vacuum degree, slowly open the valve, observe the reading of the vacuum gauge, and close the valve when the set value is reached. During the experiment, some platinum resistance PT-100 and Rhodium iron thermometers with an accuracy of 0.1 K are used to measure the cold and hot boundary

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temperatures of the sample. All measured values of temperature are recorded by the data acquisition unit (artificial intelligence temperature controller AI-751ES6). The PID temperature controller that comes with the refrigerator controlled the temperature with an accuracy of 0.1 K. A compound vacuum gauge is used to measure the vacuum degree in the vacuum chamber. Before data recording, ensure that all experimental conditions have reached the final thermal equilibrium. The relative uncertainty in the measurement process is shown in Table 1.

Fig. 2. Photo of the copper shield, the needle and the Pirani gauge. Table 1. Main parameters and uncertainty of measurement apparatus. Apparatus

Range

Accuracy

Platinum resistance thermometer

55–673 K

±0.1 K

Rhodium iron thermometer

1.3–300 K

±0.1 K

Temperature inspection instrument

~

±0.05%F.S

Compound vacuum gauge

~

±10%F.S

Table 2. Related parameters of phosphor copper wire lead. AWG

Thermal conductivity(W/m*K)

Wires number

32

20 K

80 K

150 K

300 K

4

10

25

34

48

Since the sample size in the proposed experimental setup is small (150 mm*150 mm), further analysis is needed to determine whether the lead heat leakage of temperature sensors and the heating film can be ignored. In order to effectively reduce the heat leakage of the lead, the phosphor copper wire with low thermal conductivity is selected in the experiment, and the relevant parameters are shown in Table 2. Taking the test at 20 K as an example: the hot and cold boundary temperatures of the aerogel are 7.96 K and 23.65 K respectively, the heat flow is 0.031 W, temperature sensors and the heating film have three leads in total, and the lead length is 80 mm. Through calculation, it is obtained that the heat leakage of the lead wire is 7.53*10–4 W, and the heat leakage of the lead wire accounts for 2.43% of the heat flow, which is small and can be ignored.

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3 Results and Discussion Based on the above experimental device, the thermal conductivity of the aerogel blanket in the full pressure range was tested. The test results of 77 K–300 K are shown in Fig. 3. It can be seen from the Fig. 3 (a) that in the temperature range of 77 K–300 K, regardless of the vacuum degree, there is an almost linear relationship between the temperature and the thermal conductivity. Simultaneously, it can be seen that the thermal conductivity of aerogel at the same temperature increases with the decreasing vacuum degree. This is because the amount of gas in the aerogel increases, and the thermal conductivity of the gas in the aerogel increases continuously, which also causing the minimum temperature that the refrigerator can reach constantly elevates. Figure 3 (b) shows the change of thermal conductivity of aerogel with vacuum degrees at 300 K and 250 K respectively. In the flow field, the ratio of the average free path of gas molecules to the characteristic length is defined as Knudsen number. When its value is greater than 10, the flow field is in the free molecules state. Therefore, after calculation, when the vacuum degree in the vacuum chamber is lower than 2.33 Pa, the gas inside the aerogel is in a free molecules state, and the heat transfer of the gas is very small, resulting in almost constant thermal conductivity. When the vacuum degree is between 2.33 Pa and 5000 Pa, the gas in the aerogel gradually changes from the free molecular state to the continuum status, the gas heat conduction begins to increase and the convective heat transfer also begins to appear, resulting in a rapid increase in thermal conductivity; When the vacuum degree is higher than 5000 Pa, the gas in the aerogel has been completely transformed into a continuum status, and the influence of the vacuum on thermal conductivity begins to gradually decrease.

Fig. 3. Thermal conductivity of aerogels at different vacuum degrees (residual gas in the vacuum chamber is air).

Figure 4 shows the change of aerogel thermal conductivity in the temperature range of 20 K–300 K under a high vacuum (10–4 Pa), an almost linear change can be seen. The test shows that the aerogel thermal conductivity is 1.06*10–3 W/m*K at 15.8 K; while at 304.3 K, the aerogel thermal conductivity is 1.20*10–2 W/m*K. The relationship between aerogel thermal conductivity and temperature under high vacuum is obtained

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by linear fitting, as shown in Formula 2: Ks = aT + b

(2)

Ks − thermal conductivity of the aerogel, W/(m·K); T − the average temperature of the hot and cold ends of aerogel, K; a, b − the fitting coefficient, a=3.77 ∗ 10−5 W/(m·K2 ); b=8.92 ∗ 10−4 W/(m·K); The test results in the 20 K–77 K range are shown in Fig. 5. When the helium content in the vacuum chamber gradually increases, the thermal conductivity of the gas increases greatly, since the thermal conductivity of helium is much higher than that of other gases (the thermal conductivity of air and helium are 0.024 W/m*K and 0.147 W/m*K respectively at 273 K under atmospheric pressure). This also explains the large increase in the minimum temperature that can be achieved when the pressure is 20 Pa shown in the Fig. 5.

Fig. 4. Thermal conductivity of aerogel in the Fig. 5. Thermal conductivity of aerogels at 20 K–300 K range. different vacuum degrees (residual gas in the vacuum chamber is helium).

4 Conclusions A low-temperature thermal conductivity experimental system has been established to research the thermal insulation performance of aerogel under a full-pressure range. The temperature test range of the device is 20 K to 300 K, and the vacuum test range is 10–5 Pa to atmospheric pressure. In this paper, the device was used to test the thermal conductivity of the aerogel blanket in the temperature range of 77 K–300 K in the full pressure range (residual gas in the vacuum chamber was air). The apparent thermal conductivities of the aerogel blanket at the temperature in the range of 20 K–77 K and the pressure in the range of 0.1 Pa–20 Pa were further measured (residual gas in the vacuum chamber was helium). The measurement results show that the thermal conductivity of the aerogel blanket decreases with the decrease of temperature when the vacuum degree is the same. The thermal conductivity of the aerogel blanket has a small change range below 0.1 Pa and above 5000 Pa, while increases rapidly with the increase of vacuum in the range of 0.1 Pa to 5000 Pa when the temperature is the same. When the residual gas is helium, the apparent thermal conductivity is significantly greater than that of air.

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Acknowledgments. The project is supported by the National Natural Science Foundation of China (No. 12073058, No. U1831203), the China National Space Administration (No. D050104, No.D040305), the Key Research Program of Frontier Sciences, CAS (No. QYZDY-SSW-JSC028), and the Youth Innovation Promotion Association of Chinese Academy of Sciences (No. 2019030).

References 1. Kanamori, K., Aizawa, M., Nakanishi, K., Hanada, T.: New Transparent Methylsilsesquioxane Aerogels and Xerogels with Improved Mechanical Properties. Adv. Mater. 19(12), 1589–1593 (2007) 2. Hasan, M.A., Sangashetty, R., Esther, A.C.M., Patil, S.B., Sherikar, B.N., Dey, A.: Prospect of thermal insulation by silica aerogel: a brief review. J. Inst. Eng. (India): Ser. D 98(2), 297–304 (2017) 3. Hasan, M.A., Rashmi, S., Esther, A.C.M., et al.: Evaluations of silica aerogel-based flexible blanket as passive thermal control element for spacecraft applications. J. Mater. Eng. Perform. 27(3), 1265–1273 (2018) 4. Li, J., Guo, P., Hu, C., et al.: Fabrication of Large Aerogel-Like Carbon/Carbon Composites with Excellent Load-Bearing Capacity and Thermal-Insulating Performance at 1800 °C. ACS nano (2022) 5. Dolai, S., Bhunia, S.K., Beglaryan, S.S., et al.: Colorimetric polydiacetylene-aerogel detector for Volatile Organic Compounds (VOCs). ACS Appl. Mater. Interfaces. 9(3), 2891–2898 (2017) 6. He, X., Li, X., Zhao, J., Shi, F., Hu, J.: Synthesis of GO/PAAM/iminodiacetic acid functionalized alginate aerogels with enhanced catalytic activity in oxygen reaction. Materials Letters (2022) 7. Chakraborty, S., Pisal, A.A., Kothari, V.K., Rao, A.V.: Synthesis and characterization of fibre reinforced silica aerogel blankets for thermal protection. Adv. Mater. Sci. Eng. 2016, 2495623 (2016) 8. Sabri, F., Marchetta, J., Smith, K.M.: Thermal conductivity studies of a polyurea cross-linked silica aerogel-RTV 655 compound for cryogenic propellant tank applications in space. Acta Astronaut. 91, 173–179 (2013) 9. Begag, R., White, S., Fesmire, J.E., Johnson, W.L.: Hybrid aerogel-mli insulation system performance studies for cryogenic storage in space applications. MRS Proceedings 1306 (2011) 10. Fesmire, J.E., Ancipink, J.B., Swanger, A.M., et al.: Thermal conductivity of aerogel blanket insulation under cryogenic-vacuum conditions in different gas environments. IOP Conference Series: Materials Science and Engineering 278, 012198 (2017) 11. Fesmire, J.E.: Aerogel-based insulation materials for cryogenic applications. IOP Conference Series: Materials Science and Engineering 502, 012188 (2019) 12. Hao Qiang, P., Zeng-Yao, L.: Experimental investigations on the thermal insulation performance of SiC opacifier doped silica aerogel at large temperature difference. Int. J. Thermal Sci. 160 (2021)

Analysis of the Performance Difference Between4 He and 3 He in Space JT Cryocoolers Ziyao Liu1,2 , Yuexue Ma1(B) , Jia Quan1 , Yanjie Liu1(B) , Juan Wang1 , Jianguo Li1 , and Jingtao Liang1,2 1 Key Laboratory of Technology on Space Energy Conversion, Technical Institute of Physics

and Chemistry, Chinese Academy of Sciences, Beijing, China {mayuexue,yjliu}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing, China

Abstract. Joule-Thomson (JT) cryocoolers play an essential role in space exploration missions. The JT cryocoolers with different working fluids have different operating temperatures. 4 He and 3 He are used in JT coolers to obtain temperatures below 4 K. Due to the different physical properties of 4 He and 3 He, 4 He is widely used to get the temperature above 2 K, while 3 He is used to obtain the temperature below 2 K. In this paper, a finite-element model (FEM) of the counter-flow tubein-tube heat exchangers (CFHX) is established. According to the Laval nozzle model, the mass flow rate and the cooling capacity of the JT process are analyzed. Finally, this paper combines these models and uses MATLAB for programming. The influence mechanism of the CFHX length and the JT orifice diameter on the 4 He and 3 He JT effects are obtained. Keywords: 4 He · 3 He · Joule-Thomson effect · Heat exchanger · JT cryocooler

1 Introduction The Space 2 K cryogenic technology is a crucial precooling stage of the sub-Kelvin refrigerators. Besides, the Space 2 K cryogenic technology is also the premise that the detector with an operating temperature of 2 K can normally work in space. Due to the high efficiency of the hybrid Joule-Thomson (JT) cryocooler at a temperature of about 2 K, many research institutes have recently accelerated the research on the space 2 K hybrid JT cryocooler. The space 2 K hybrid JT cryocooler is a challenging research program. Table 1 summarizes the development of hybrid 2 K JT cryocoolers [1–8]. Because 3 He is too scarce, some institutions utilize 4 He to develop the space 2 K hybrid JT cryocooler. For example, the National Institute of Standards and Technology (NIST) developed a hybrid 4 He JT cryocooler that reaches 1.7 K in the open cycle. Since the saturated vapor pressure of 4 He below 2 K is too low, its JT compressor has not been developed [6]. However, at the same temperature, the saturated vapor pressure of 3 He is higher than that of 4 He. Lower temperatures are easier to be achieved with 3 He than with 4 He. Some space exploration missions are designed to carry the space 3 He hybrid JT cryocoolers. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 986–993, 2023. https://doi.org/10.1007/978-981-99-6128-3_128

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As Table 1 shows, there is a significant difference between the performances of the 3 He JT cryocooler and the 4 He JT cryocooler. Based on the space 2.17 K JT cryocooler developed in our laboratory, a finite-element model (FEM) of the counter-flow tube-in-tube heat exchangers (CFHXs) is established in this paper. Furthermore, according to the Laval nozzle model, the mass flow and the cooling capacity of the JT effect process are investigated. Finally, combining these models and using MATLAB programming, the heat exchange performance of the final stage CFHX and the performance of the JT effect when 3 He or 4 He flows are analyzed. Table 1. Overview of hybrid 2 K JT cryocoolers designed for space applications. Missions/Institutions

Cryogenic performance

Cryocooler types

JT compressor

Astro-H

10.1 mW @ 1.72 K

2SC + 3 He JT

4-stage

SPICA

16 mW @ 1.70 K

2SC + 3 He JT

4-stage

19 mW @ 1.77 K

2PTC + 3 He JT

4-stage

RAL

14 mW @ 2.00 K

GM + 3 He JT

4-stage

SITP

13.9 mW @ 1.80 K

3PTC + 3 He JT

4-stage

1.4 mW @ 1.70 K

3PTC + 4 He JT

Open-loop

SIPT

1.80 K

3PTC + 4 He JT

4-stage

TSEC

2.17 K

2PTC + 4 He JT

4-stage

ATHENA

NIST

Note: RAL—The Rutherford Appleton Laboratory, TSEC—The Key Laboratory of Technology on Space Energy Conversion of CAS, SITP—The Shanghai Institute of Technical Physics of CAS

2 Model Establishment 2.1 The Finite-Element Model of the CFHX The final stage CFHX is a critical part of the hybrid JT cryocooler, which directly affects the temperature of the fluid before the JT process. The FEM of the CFHX is established, as shown in Fig. 1. The high-pressure fluid flows in the inner tube made of copper. The low-pressure fluid flows in the outer tube made of 304 steel. According to energy balances, the following equation can be obtained. ˙ hI ,i + qCh,i dTh,i The high − pressure fluid : H˙ h,i − H˙ h,i+1 = Q

(1)

˙ hI ,i + Q ˙ I ,i − Q ˙ I ,i+1 − Q ˙ IL,i = qCI ,i dTI ,i The inner tube : Q

(2)

˙ IL,i + Q ˙ OL,i = qCL,i dTL,i The low − pressure fluid : H˙ L,i+1 − H˙ L,i + Q

(3)

˙ O,i + Q ˙ rad ,i + Q ˙ s,i − Q ˙ OL,i − Q ˙ O,i+1 = qCO,i dTO,i The outer tube : Q

(4)

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Fig. 1. The schematic of FEM

˙ hI ,i is the heat transfer rate between the inner tube and the high-pressure fluid, where Q ˙ IL,i is the heat transfer rate between the inner tube and the low-pressure fluid, W. W. Q ˙ OL,i is the heat transfer rate between the outer tube and the low-pressure fluid, W. H˙ Q is the enthalpy flow rate of the inner or outer tube, W. These heat transfer rates can be calculated by the following formulas. ˙ hI ,i = αhI ,i ShI ,i ( Th,i + Th,i+1 − TI ,i ) Q 2   TL,i + TL,i+1 ˙ QIL,i = αIL,i SIL,i TI ,i − 2 ˙ OL,i = αOL,i SOL,i (TO,i − TL,i + TL,i+1 ) Q 2

(6) (7)

λh,i + λh,i+1 Nuh,i 2 dI

(8)

λL,i + λL,i+1 NuL,i = αOL,i 2 dO

(9)

αhI ,i = αIL,i =

(5)

where αhI ,i is the heat transfer coefficient between the inner tube and the high-pressure fluid, W/(m2 K). αIL,i is the heat transfer coefficient between the inner tube and the lowpressure fluid, W/(m2 K). αOL,i is the heat transfer coefficient between the outer tube and the low-pressure fluid, W/(m2 K). Nuh,i is the Nusselt number of the high-pressure fluid flowing in the inner tube. NuL,i is the Nusselt number of the low-pressure fluid flowing in the outer tube. These two Nusselt numbers can be calculated by the following formulas [9]. For the laminar flow,   0.9  1 d (10) Reξ Pr 3 Nu = 3.66 + 0.08 1 + 0.8 D d 0.194 ξ = 0.5 + 0.2903( ) D

(11)

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For the turbulent flow, Nu =

ξ 8 RePr

 2 1 + 12.7 ξ8 (Pr 3 − 1)

(12)

d 0.5 ) + 0.03( (13) D Re0.25 where d is the diameter of the tube. D is the spiral diameter of the CFHX. ξ is a reference value. In addition, the heat transfer along the tube is calculated by the following formulas. ξ=

0.3164

˙ I ,i = λI ,i AI ,i (TI ,i−1 − TI ,i )/dL Q

(14)

˙ O,i = λO,i AO,i (TO,i−1 − TO,i )/dL Q

(15)

where λ is the thermal conductivity of the inner or outer tube, W/(m · K). This model ignores the radiative heat exchange and the heat convection between the CFHX and the vacuum environment and assumes that the fluid has no pressure loss in the tube. 2.2 The Cooling Capacity Model Our laboratory uses the small orifice as the JT expansion impedance. The orifice can be regarded as a Laval nozzle. According to the Laval nozzle model, the mass flow rate of the ideal fluid is calculated by the following formulas [10].  κ+1 2(κ−1) P02 M 2 m ˙ = κ 0.5 ( ) (16) G(P0 , T0 ) = A 1+κ RT0 1 h(P1 , T1 ) + V 2 = h(P0 , T0 ) (17) 2 s(P1 , T1 ) = s(P0 , T0 )

(18)

π 2 d · 10−12 4 orifice

(19)

A=

Qc = m(h ˙ 2 − h1 )

(20)

where G is the mass flux at the orifice, kg/(m2 · s). P1 and T1 are the pressure and temperature of the fluid before the JT process, which are the same as the pressure and temperature of the high-pressure fluid at the CFHX outlet calculated from the FEM. However, the mass flux at the orifice is related to the pressure and temperature of the stagnation state before the JT process, P0 and T0 , which can be calculated by formulas 17 and 18. dorifice is the diamater of the orifice, μm. A is the cross-sectional area of the orifice, m2 . Qc is the cooling capacity of the JT effect, W. h1 is the enthalpy of the fluid before the JT process, J/kg. After the JT process, fluid flows into the evaporator. h2 is the enthalpy of the fluid flowing out of the evaporator, which is equal to the enthalpy of the low-pressure fluid flowing in the CFHX, J/kg. The enthalpy difference (HD) mentioned later refers to the difference between h2 minus h1 .

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3 Calculation Results and Discussions According to the previous experimental results of the 4 He JT cryocooler, the suction pressure of 5.33 kPa and the supply pressure of 1 MPa can be easily provided by the JT compression system, which are used in the following calculations. At this operating parameter, the 4 He JT effect can obtain a temperature of 2.20 K, while the 3 He JT effect obtains 1.42 K. Table 2 displays the assumed working parameters of 3 He/4 He in CFHX based on previous experimental results. The diameters of the CFHX and the JT orifice used for calculations are shown in Table 3. Table 2. The assumed working parameters of 3 He/4 He in CFHX Supply Pressure (MPa)

Suction Pressure (kPa)

Inlet temperature of high-pressure 3 He/4 He (K)

Inlet temperature of low-pressure 3 He (K)

Inlet temperature of low-pressure 4 He (K)

1.0

5.33

18

1.42

2.20

Table 3. The diameters of the CFHX and the orifice used for calculations Length (m)

doo (mm)

doi (mm)

dio (mm)

dii (mm)

dorifice (μm)

1–2

4

3

2

1.6

15–50

As the orifice becomes larger, the mass flow rate of the 3 He JT effect increases, as shown in Fig. 2. Although a higher mass flow causes a larger heat transfer coefficient between the fluid and the tube according to the FEM, the mass flowing in the CFHX per unit time also increases. Overall, the outlet temperature of the high-pressure fluid in a certain CFHX is higher with increasing the mass flow rate, as shown in Fig. 3. In addition, as the length of the final stage CFHX increases, the outlet temperature of the high-pressure fluid decreases with the same JT orifice. Thus, when the JT orifice is constant, the mass flow rate in the long CFHX is higher than that in the short CFHX because the outlet temperature of the high-pressure fluid in the long CFHX is lower than that in the short CFHX, based on the formula 16. Under the condition that the properties of the fluid after the JT process remain unchanged, the change trend of the HD is opposite to that of the outlet temperature, as shown in Fig. 4. According to formula 20, the cooling capacity is the product of the mass flow rate and the HD. If the final stage CFHX is longer, the cooling capacity with a certain JT orifice will be higher because both the HD and the mass flow are higher. The cooling capacity of the 3 He JT effect with a certain CFHX has a peak value with increasing the diameter of the JT orifice, as shown in Fig. 5. As the orifice gets larger, the cooling capacity with a certain CFHX gradually increases to a peak and then decreases. This is because, when the JT orifice is too large, the mass flow rate of the JT process is too high. The heat exchange in the CFHX has already exceeded its heat exchange capacity in this situation. Therefore,

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the large orifice is not always better. In addition, as the CFHX gets longer, the peak of the cooling capacity will be higher, and the orifice corresponding to the peak cooling capacity will be larger. However, the pressure drop in tubes also increases as the CFHX becomes longer, which directly affects the temperature obtained by the JT effect. Thus, the parameters of the CFHX and the diameters of the JT orifice should be considered comprehensively in the design of the hybrid JT cryocooler.

1.7 m CFHX 1.5 m CFHX 1.3 m CFHX 1.1 m CFHX

The outlet temp of the high-pressure fluid (K)

The mass flow rate (mg/s)

15

1.7 m 1.5 m

10

1.3 m 1.1 m 5

0 10

15

20

25

30

35

40

45

50

55

3.0

1.1 m

1.7 m CFHX 1.5 m CFHX 1.3 m CFHX 1.1 m CFHX

1.3 m 1.5 m

1.7 m

2.5

2.0

1.5 10

15

The diameter of the JT orifice (µm)

20

25

30

35

40

45

50

55

The diameter of the JT orifice (µm)

Fig. 2. The mass flow rates of the 3 He JT effect with different JT orifices

Fig. 3. The outlet temperatures of the high-pressure 3 He with different JT orifices

1.7 m CFHX 1.5 m CFHX 1.3 m CFHX 1.1 m CFHX

14

Cooling capacity (mW)

The enthalpy difference (J/g)

3

2

1

1.1 m 0 10

15

20

25

30

35

1.3 m

1.5 m 1.7 m

12

1.7 m CFHX 1.5 m CFHX 1.3 m CFHX 1.1 m CFHX

10

1.7 m

8

1.5 m 6

1.5 m 4

1.1 m 40

45

50

55

The diameter of the JT orifice (µm)

Fig. 4. The enthalpy differences of the 3 He JT effect with different JT orifices

2 10

15

20

25

30

35

40

45

50

55

The diameter of the JT orifice (µm)

Fig. 5. Cooling capacities of the 3 He JT effect with different JT orifices

The calculation results of the 4 He JT effect have similar change rules with that of the 3 He JT effect, as shown in Figs. 6 and 7. Compared with the 3 He JT process, the outlet temperature of the high-pressure 4 He is higher due to the higher temperature of the low-pressure 4 He entering the CFHX. The HD of the 4 He JT effect also decreases as the orifice gets larger. The cooling capacity of the 4 He JT effect has a similar trend as that of the 3 He JT effect. However, the peak of the cooling capacity of the 4 He JT effect with the same CFHX is higher, and the JT orifice corresponding to the peak is smaller. Thus, the appropriate JT orifices are different with different lengths of the CFHX and different working fluids.

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Cooling capacity (mW)

50

40

30

20

1.7 m 10

1.5 m 1.1 m

0 10

20

The enthalpy difference (J/g)

60

1.7 m CFHX 1.5 m CFHX 1.3 m CFHX 1.1 m CFHX

15

10

5

1.7 m 0

1.3 m

1.3 m 30

1.5 m

1.1 m 40

The diameter of the JT orifice (µm)

Fig. 6. Cooling capacities of the 4 He JT effect with different JT orifices

10

20

30

40

The diameter of the JT orifice (µm)

Fig. 7. The enthalpy differences of the 4 He JT effect with different JT orifices

4 Conclusions In this paper, the finite-element model of the CFHX is established. The FEM is combined with the Laval nozzle model to form the cooling capacity model of the JT effect based on MATLAB. The performance of the 3 He JT effect and the 4 He JT effect based on our space JT cryocooler is analyzed. With the same JT orifice and same diameters of the final stage CFHX, if the CFHX gets longer, the cooling capacity will be higher, and the pressure drop will also get higher. Thus, the CFHX should be long enough but have an acceptable pressure drop. Furthermore, as the JT orifice gets larger, the cooling capacity of the JT process with a certain CFHX increases to a peak and then decreases. The peaks of cooling capacities of the 3 He and 4 He JT effects with the same CFHX correspond to the different diameters of the JT orifices. The CFHX’s length, the peak of cooling capacity, and the working fluid of the JT process should be considered when selecting a suitable JT orifice. This paper provides a method for designing the final stage heat exchanger and selecting the JT orifice of the 3 He and 4 He hybrid JT cryocooler. Acknowledgments. The project is supported by the Strategic Priority Research Program of Chinese Academy of Sciences (Grant No. XDA18000000, No. XDA18040000) and 2009ZYHG0003.

References 1. Sato, Y., et al.: Development of mechanical cryocoolers for Astro-H/SXS. Cryogenics 50(9), 500–506 (2010) 2. Sugita, H., et al.: Development of mechanical cryocoolers for the Japanese IR space telescope SPICA. Cryogenics 48(5–6), 258–266 (2008) 3. Shinozaki, K., et al.: Cooling performance of Joule Thomson coolers in 300 K -50 mK cryochain demonstration for ATHENA X-IFU. IOP Conference Series Materials Science and Engineering 502(1), 012069 (2019) 4. Crook, M., Hills, M., Gilley, G., Rawlings, T., Pulker, C., Green, B.: Performance testing of a 2K Joule-Thomson closed-cycle cryocooler. Cryocooler 21, 433–441 (2021) 5. Dang, H., et al.: Investigations on a 1 K hybrid cryocooler composed of a four-stage Stirlingtype pulse tube cryocooler and a Joule-Thomson cooler. Part B: Experimental verifications. Cryogenics 123, 103452 (2022)

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6. Kotsubo, V., Ullom, J.N., Nam, S.W.: Compact 1.7 K cryocooler for superconducting nanowire single-photon detectors. Cryocooler 20, 295–304 (2018) 7. Dang, H., Zhang, T., Zhao, B.: A hybrid cryocooler achieving 1.8 K with He-4 as the only working medium and its application verification. Chin. Sci. Bull. 67(9), 896–905 (2022). (In Chinese) 8. Liu, Z., et al.: Development of a compact 2.17 K hybrid 4He JT cryocooler for space applications. Cryogenics 118, 103347 (2021) 9. Geng, L.H., Cao, H.S., Li, J.M., Jiang, P.X.: Design and performance analysis of an ejectorequipped Joule–Thomson cryocooler. Appl. Thermal Eng. 190, 116779 (2021) 10. Maytal, B.Z., Pfotenhauer, J.M.: Miniature Joule-Thomson cryocooling: principles and practice. Springer Science & Business Media (2012)

Cryogenic Instrumentation and Control

Recent Progress in Instrumentation for Large to Medium Scale Facilities Operating at 4 K Temperature Level Sergiy Putselyk(B) Magna ESS, Temsa Alle 1, 8261 Sinabelkirchen, Austria [email protected]

Abstract. An operation of cryogenic facilities working at 2–4 K temperature levels is hardly possible without a reliable and properly functioning cryogenic instrumentation. The key measurement components are temperature sensors, pressure transmitters, strain gauges and flow meters. Present paper describes a progress in the instrumentation for the large to medium scale industrial or scientific facilities operating at 2–4 K temperature level over the last 10–15 years. The focus of the paper is on the sensors, which have been successfully applied for serial applications and commercially available or being in the prototyping stage and therefore have the potential for industrial applications. Keywords: Cryogenic sensors · Temperature measurements

1 Introduction It is of a paramount importance to have a reliable instrumentation for any refrigerator, cryostats or cryomodule operating at 2–4 K temperature levels. To make the most optimal choice of sensors, it is important to follow the most recent developments in this field. This paper presents an overview on the recent progress on cryogenic instrumentation for large to medium scale industrial facilities. In order to fulfil requirements on a quite limited available space for the paper, only sensors for large to medium facilities are considered, e.g. for refrigerators, medium to large scale detectors or accelerator cryomodules. Present paper reviews the progress in the last 15 years, and for earlier reviews one can refer to e.g. [1]. The first chapter gives an overview on temperature sensors, followed by the one on flow meters. Strain gauges, pressure transmitter and level meters are considered in chapters 4 and 5, respectively.

2 Temperature Sensors Before one starts discussion on the temperature sensors, it is important to shortly remind the main requirements, which are applied on sensors for industrial applications, e.g.: This work was mainly done during the author work at Ferchau Engineering, Germany. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 997–1005, 2023. https://doi.org/10.1007/978-981-99-6128-3_129

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accuracies, see Table 1, maintenance free operation for at least 10 years, data acquisition rate, one measurement per 2–5 s, and low costs for sensors and reading electronics.

Table 1. Typical accuracy required for the temperature measurements in the range 1–5 K Accuracy

Application area

± 0.005 K or better

Specialized laboratories, e.g. on material properties

± 0.005 till ± 0.05 K

Laboratory measurements or industrial prototypes, but also for industrial monitoring with high accuracies

± 0.05 till ± 0.20 K

Practically near all industrial monitoring facilities

± 0.20 till ± 2.0 K

Monitoring of cool-down / warm-up of cryogenic facilities, temperature interlocks and alarms

It is also worth to underline two important details for the further discussion of the temperature sensors, i.e.: • the focus is given to the sensor installation. This is related to two facts, that a) sensor development reached a maturity, and manufacturing firms developed sensors for “plug-in-play” usage, and only installation must be correctly performed, b) most of “mistakes” are done on the sensor installation, because a comprehensive understanding of the physics behind is sometimes missing. • Sensor performance (particularly a long-term one) is not discussed in this paper. Although this parameter is important and allows comparison of sensor performances, these data are rarely published in the opened literature sources. For that reason, the author has to use own unpublished data or ones provided by his colleagues, and, therefore, for the reader it is often not possible to find and analyse these data. 2.1 Sensors From large varieties of temperature sensors, for industrial application, only three types are applied, i.e. CERNOX, Si-Diode and partly TVO (Vishay CLTS-2B has also significant potential, though more data on industrial applications must be present in opened literature sources). For the near future perspective it is possible that fiber-optic sensors for helium temperatures will be commercialized. At the present, such sensors are already applied at nitrogen temperatures and in some cases at helium ones [14]. It would be very helpful to have data published on large sensor numbers for better statistics as well as to have data on long-term stability. It is worth to note that at the present time, the temperature resolution is somehow worse than the one achieved with Cernox, TVO or Si-diodes.

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CERNOX This is the most popular resistive temperature sensor as for the industrial as well as for low temperature physics applications. Special thanks must be given to LakeShore Cryotronics for providing large variety of forms and sample holders, which can be matched for any required temperature ranges as well as application. Recent activities were mainly directed to characterizing the different packages, e.g. for mechanical [2], and electrical [3] and aging [4] stabilities, thermal cycling [5].

Fig. 1. Cernox (AA-serie according to Lake Shore notation) with PCB, developed at CERN [6].

Cernox sensors were successfully applied at LHC, CERN [6]. In order to adapt these sensors for LHC needs, special sensor holder and mounting procedure were applied, see Fig. 1. The main goals were: i) to automate the mounting procedures as much as possible for reduction of overall installation time, ii) to facilitate thermal anchoring of wires and iii) not to overheat the sensors in order to avoid the shift of calibration curves (so, only proven measurement devices were allowed). After the sensor mounting on holder made from Printed Circuit Board (PCB), it was calibrated at special test facilities. So, temperature difference between sensor and PCB was included in the calibration, and temperature difference between PCB and copper blocked brazed on process pipe was negligible due to large contacting surface area. As a limitation for the applications with restricted free space, the relative large size of PCB (100·10·2 mm) could be mentioned. This type of sensor mounting with slight modifications will be also applied at ITER project [7]. It is worth to note that it is possible to mount the Cernox sensor inside a small pipe, which is perpendicularly welded into the larger process pipe, see Fig. 2 as an example. So, precise temperature measurement inside process lines is possible [8]. Si-Diodes It is probably the most popular thermometer for industrial application. Its main advantages are low costs and possibility to use standard calibration curves. LakeShore Cryotronics offers different packages and holding clamps for proper mounting. For the case if a mounting inside the pipe is needed, a design applied at JLab, see Fig. 2, could be used [10].

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The Si-diode sensors have been extensively investigated in the past and no significant development was needed in the last 15 years (only some supplementary measurements, like sensor behavior under high radiation fields or extremely high pressures, were done).

Fig. 2. Diode mounted inside the process pipe, developed at JLab [10].

TVO This temperature sensor gained a significant attention over the last two decade. Besides having similar characteristics to CERNOX sensor, its main advantages are robustness to electrical disturbances and following the standard calibration curse, i.e. it is possible to calibrate the sensor on three points at 4 K, 77 K and 300 K. Main development activities have been done by three groups: • CERN and Temati Company [9]: Surface mounting using the thermal pad typically leads to good results if it is done properly, see Fig. 3 and 4. Different possibilities exists, e.g. a gluing (or soldering) thermal pad, typically done from Kapton (polyimide) with copper strips, on flat or round surfaces with following soldering of sensor bines on copper strips and putting Apiezon N between the sensor and thermal pad or gluing the sensor on the pipe or cable surface with Stycast. It is worth to note that gluing with Stycast must be done on only one surface, otherwise filling of the sensor with the glue from all sides will lead to significant thermal-mechanical stresses on sensor and calibration curve is shifted. Such mounting method, see Fig. 3, was also successfully used at three test benches for testing of XFEL cryomodules at DESY [11]. • Mounting inside the hole with Apiezon N grease used for thermal and mechanical contact between sensor and inner surfaces of the hole [12]: this mounting solution was developed by Y. Filippov and could be shortly described as follows: i) copper block with the 3 mm hole is brazed to the object (typically with stainless steel surface) in the vertical direction and 1mm hole for the air release is done at bottom of this block, ii) the TVO sensor with connecting wires is inserted into the 3 mm hole, iii) Apiezon N grease is filled into the space between the sensor and copper block, so while keeping the copper block slightly heated with heat gun, the vacuum grease is flowing from the top inside the 3 mm hole of the copper block, while air is released from the bottom through 1 mm hole. Similar copper block is used for the thermal anchoring of cables, with the only difference – instead of TVO sensor, small coils of measuring wires are inserted. This mounting method has the following advantages: i) simple installation method, i.e. TVO sensors with connecting cables can be prepared at laboratory and only insertion into the hole with the following vacuum grease filling

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is required for a work around the cryogenic facility, ii) error due to mounting method is negligible because similar mounting method is used for sensor calibration inside the calibration cryostat. Application of this mounting method could be found e.g. in several large distribution boxes at AMTF facility, DESY, as well as in the transfer line and string connection boxes of XFEL accelerator, DESY. • Mounting of TVO sensor inside the copper can [13]: this mounting solution was developed by M. Suesser and has similar advantages as the mounting method developed by Y. Filippov. Though there are significant differences, i.e. i) copper can allows thermal equilibration of temperature around the TVO sensor, ii) due to very thin layer of Stycast between TVO sensor and the can, good mechanical and thermal contact is guaranteed over 10 or more years as well as the sensor thermal time constant is minimized, iii) due to high tolerances on copper block (H7 for the hole) as well as for can (g6), no gluing is needed, i.e. only Apiezon N is applied, iv) it is possible to grave the sensor number on the copper can and for the cross-checking it is possible to take out the sensor, check the sensor number and to insert it back. It is worth to stress that due to very low energy deposition in the sensor, very thin Apiezon layer and large contact area, the temperature difference over this grease layer is estimated to be below 1 mK. This small temperature difference will be also present, if e.g. only 30% of surface area is filled with grease. It is important to note, that for the calibration, this sensor is inserted in the copper block with the similar tolerances, and temperature of copper block is well monitored. In this case, the temperature offsets in the vacuum grease, copper can, Stycast glue and sensor squeezing due to Stycast glue, are included into the calibration curve and therefore do not need to be considered. It is worth to note that it is possible to insert the copper can inside the small stainless steel tube, which is welded perpendicularly inside the larger one for temperature measurement inside the process lines, see Fig. 2 as an example, or into different designed copper blocks brazed or screwed to process line for temperature measurement outside the line. And last but not the least, it is possible to optimize. This cupper can for installation of any other sensors, e.g. CERNOX, Pt, Si-diodes. So it could be advantageous, if manufacturing firms have such copper can as possible option for sensor installation in their portfolio. It is worth also to mention, that mounting the TVO sensor inside the copper block brazed to steel surface was done for JT-60SA cryogenic system [8].

Fig. 3. TVO sensor mounted on sc cable or rod with use of thermal pad, developed at CERN [9].

And last but not the least, it is worth to say that for industrial applications, it is very important to have estimation on total costs for single measurement channel. Unfortunately, total costs are rarely mentioned in the opening literature sources, and therefore

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it is difficult to compare the sensor costs, installation work, and reading electronics. According to the author’s experience, LakeShore Cryotronics had successfully commercialized large variety of sensor and sensor holder and developed reading electronics to be placed in electrical cabinets of the industrial facilities. However, the sensor costs, and uncertainties of temperature measurements due to installation methods, were for several cryogenic groups not acceptable and other solutions were searched. TVO sensor provided very good alternative to the CERNOX one with respect to physical properties and in the early times also had advantaged due to prices. However, limitation on the buying possibilities as well as later changing of sensor properties (sensor temperature sensitivity was significantly reduced due to new production process) did not make this sensor a real competitor to the Cernox one. Moreover, nowadays, the costs of a bare chip are also comparable with the ones of Cernox. Installation methods, which were developed at CERN, by Y. Filippov and M. Suesser, see discussion above, decreased uncertainties of temperature measurement due to installation, and reduced the time needed for an installation work on cryostat. However, in comparison to the simple “screw” methods, which have been developed by LakeShore Cryotronics for example on the ET, MT, CO, BO, CU, CD packages, other three methods mentioned above, brought also significant additional costs, for example, brazing of copper block to the stainless steel surfaces, manufacturing with high tolerances (H7/g6), hand work with the vacuum grease. So, bevor one starts a designing of the temperature measurement channel for industrial applications working at the helium temperatures, it is worth to consider the overall costs and required accuracies. These two parameters will lead to the choice of sensor, mounting method and reading electronics. Based on the long-time experience with applications of different temperature sensors, author also proposed some modifications of copper can, originally developed by M. Suesser. Figure 5 shows very similar copper can, developed by M. Suesser. It is possible to insert not only TVO, but to adapt other sensor type, e.g. Pt100 or bare chip Cernox Sensor. This solution is shown in Fig. 6, left, where Pt100 is mounted inside the copper can. Other solution is to put the sensor, e.g. TVO, Pt100 or Cernox, in the slot of copper rectangular block, see Fig. 8, right. Due to two screws, the contact surface area between the rectangular copper block and object is significantly increased. Even in the case, if one screw is lost, other screw will have sufficient force to provide a good thermal and mechanical contact. If temperature difference over contact between the rectangular copper block and object is two high and inacceptable, it is possible to use Apiezon N grease or indium foil. It is also possible to adapt the contact surface area of the copper block to other geometries, e.g. circular or cylindrical. It is also necessary to calibrate this copper block with sensor to estimate the temperature error in copper block, Stycast glue and sensor. 2.2 Reading Electronics Many laboratories, e.g. DESY, CERN, JLab, or industrial firms, Linde Kryotechnik or Air Liquide, have their own developed low costs electronic reading devices. Unfortunately, their expertise is not easily shared, or direct module exchange is not possible due to different operating systems, e.g. LabView, EPICS, Siemens PCS7.

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Fig. 4. Mounting of TVO sensor on conductor or magnet laminates [9].

Fig. 5. Drawing of copper can used for TVO mounting.

Fig. 6. Examples of a round copper cans for Pt100 (left) and a rectangular one for TVO sensors (right)

In the last 10 years, significant progress was done by Gantner Instruments GmbH, which tried to develop robust industrial reading devices for CERNOX and TVO sensors, see e.g. model Q.bloxxA105CR. Their devices were installed at Wendelstein 7X stellerator or other laboratories. These models have a robust construction for the installation inside the electrical cabinets and have typical industrial systems, e.g. Ethernet TCP/IP, Profibus DP, EtherCAT or others under request. For several years, LakeShore Cryotronics Inc developed similar temperature monitors, see “240 series input modules”. These monitors have the potential for industrial applications; though it would be very helpful to have data published on a long-term operational performance at the large and medium scale cryogenic facilities.

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3 Flow Meters In the last 15 years, number of Coriolis flow meters was significantly increased. They are now more reliable as for twenty years ago and long-term stability was also demonstrated [15]. Practically near all Coriolis flow meters for helium temperatures were manufactured by Micro Motion company and it would be also helpful to test flow meters of other firms. For the case, if operation in magnetic fields or radiation is foreseen, Venturi flow meters could be applied, see e.g. ITER project [16], though significant work must be still done for the manufacturing and testing of Venturies. Other flow meter types are under early development stages and testing of serial prototypes and at industrial facilities over long time periods must be still done [17, 18].

4 Strain Gauges and Pressure Transmitters It is still not possible to buy commercialized strain gauges and pressure transmitters for helium temperatures. They are practically very seldom used in serial facilities and only applied on testing of prototype cryostats or cryomodules. According to the author’s knowledge, these sensors were successfully used for a testing of large superconducting magnets for Fusion facilities, e.g. LCT project at KIT, Germany. For the practical applications, one has to take either “old” and already tested versions (assuming it is still offered by manufacturing companies) or to choose arbitrary ones and to test them at the cryogenic temperatures before installation [19].

5 Level Meters The state-of-the-art for level meters for industrial facilities is practically still the same as for 30 years ago. NbTi superconducting level meters and pressure difference ones are applied. Other superconducting materials [20] or level meters based on other measuring principles [21] are still rarely used and not commercialized. It is worth to mention that NbTi-based level meters and pressure difference ones perform reliable and long-term performance was demonstrated, and, therefore, it is not expected that new sensors based on other technologies will be released on market soon.

6 Conclusion In the present paper, the recent progress in instrumentation for the large to medium scale industrial facilities operating at 2–4 K temperature levels is shortly presented and discussed.

References 1. Timmerhaus, K., et al.: Cryogenic Engineering, Fifty Years of Progress Springer Book, ISBN 978–0–387–46896–9, p.374 and Chapter 10 Lessons Learned in 50 Years of Cryogenic Thermometry, pp. 179–221 (2007)

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2. Courts, S.: Mechanical robustness of cryogenic temperature sensors packaged in a flat, hermetically-sealed package. IOP Conf. Ser. Mater. Sci. Eng. 101, 012153 (2015) 3. Scott, S., et al.: Effects of electrostatic discharge on three cryogenic temperature sensor models. Adv. Cryo. Eng. 59, p118 (2013) 4. Courts, S., et al.: Effect of room temperature aging on two cryogenic temperature sensor models used in aerospace applications. Adv. Cryo. Eng. 57, p1329 (2011) 5. Courts, S., et al.: Reliability and stability of three cryogenic temperature sensors models subjected to accelerated thermal cycling. Adv. Cryo. Eng. 57, p515 (2011) 6. Balle, C., et al.: Industrial-type cryogenic thermometer with built-in heat interception. Adv. Cryo. Eng. 41, 1715–1722 (1995), see also CERN report AT/95–35 (CR), LHC Note 346 7. Poncet, J., et al.: Cryogenic instrumentation for ITER project. Conf. Ser. Mater. Sci. Eng. 171, 012130 (2016) 8. Natsume, K., et al.: Cryogenic thermometry for refrigerant distribution system of JT-60SA. IOP Conf. Ser. Mater. Sci. Eng. 101, 012114 (2015) 9. Datskov, V., et al.: Precise thermometry for next generation LHC Superconducting magnets. IEEE Trans. Appl. Supercond. 24(3), 90000905 (2014) 10. Crea, J., et at.: Instrumentation, Controls, Design Overview, Serie Lecture on Cryogenic Engineering, Jefferson Laboratory, March 22 (2011) 11. Boeckmann, T., et al.: Serial testing of XFEL cryomodules: results of the cryogenic heat load measurements. IOP Conf. Ser. Mater. Sci. Eng. 278, 012184 (2017) 12. Filippov, Y., et al.: Operation of TVO temperature sensors in the range from 4.2 to up to 425 K. Cryogenics 44, 735–739 (2004) 13. Suesser, M.: Encasing of the TVO-sensor for heavy duty installations. Cryogenics 44, p255258 (2004) 14. Chiuchiolo, A., et al.: Cryogenic test facility instrumentation with fiber optic and fiber optic sensors for testing superconducting accelerator magnets. IOP Conf. Ser. Mater. Sci. Eng. 278, 012082 (2017) 15. Boeckmann, T., et al.: Long term stability of Coriolis flow meters: DESY experience. IOP Conf. Ser. Mater. Sci. Eng. 171, 012140 (2016) 16. Andre, J., et al.: Series supply of cryogenic venturi flowmeters for ITER project. IOP Conf. Ser. Mater. Sci. Eng. 101, 012155 (2015) 17. Jansen, A., et al.: Experimenatal validation of a self-calibrating cryogenic mass flowmeter. IOP Conf. Ser. Mater. Sci. Eng. 278, 012077 (2017) 18. Lovell, T., et al.: Cryogenic radio frequency two-phase flowmeter. Adv. Cryo. Eng. 51, p273 (2005) 19. Benson, C.P., et al.: Improved capacive stress transducers for high-field superconducting magnets. Adv. Cryo. Eng. 57, 1337 (2011) 20. Takede, M., et al.: Superconducting characteristics of short MgB2 wired of long level sensor for liquid hydrogen. IOP Conf. Ser. Mater. Sci. Eng. 101, 012156 (2015) 21. Karunanithi, R., et al.: Calibration and linearity verification of capacitance type cryolevel indicatiors using cryogenicallz multiplexed diode array. Adv. Cryo. Eng. 57, p499 (2011)

Architecture Design of Control System for the DALS Test Facility Cryogenic System Lei Xu1 , Zheng Sun1,2(B) , Liangbing Hu3 , Xu Shi1 , Haining LI3 , Jichao Dong3 , and Xilong Wang1 1 Dalian Institute of Chemical Physics (DICP, CAS), 457 Zhongshan Road, Dalian, China

{xulei2021,zhengsun,xushi,xilongwang}@dicp.ac.cn

2 University of Chinese Academy of Sciences, Shijingshan District, No.19(A) Yuquan Road,

Beijing, China 3 Institute of Advanced Science Facilities (IASF), Guangming Street, Shenzhen, China

{hulb,lihaining,dongjichao}@mail.iasf.ac.cn

Abstract. The Dalian Advanced Light Source (DALS), proposed by Dalian Institute of Chemical Physics (DICP), is a linear accelerator based on continuous wave superconducting radio frequency technology aiming to produce high-quality electron beam with repetition rate up to 1 MHz. Before the project is officially implemented, a DALS test facility (DATF) is under construction to qualify the performance of the key components, including the superconducting cavities and the cryomodules. A cryogenic system is designed to provide the cooling capacities for the DATF. The DATF is mainly composed of Horizontal Test Bench (HTB), Vertical Test Cryostat (VTC), Cryogenic Test Bench (CTB) and Injector Test Bench (ITB). The overall objective of cryogenic control system is basically a distributed system to guarantee the control and monitoring of the test facility. The system is mainly composed of Programmable Logic Controllers (PLCs) with local human machine interfaces (HMIs) and the Experimental Physics and Industrial Controls System (EPICS), it contains a total of seven control cabinets that can complete the functions of data acquisition and transmission, regulation/discrete/sequential control, monitoring layer operation, alarm display, data storage and login management of all important equipment. This paper reports on the architecture design and current progress of the control system of the DATF cryogenic system. Keywords: DATF Cryogenic system · PLC · Control · EPICS

1 Introduction The cryogenic system will provide the necessary cryogenic environment to supply the 2 K superfluid helium required for operation of the superconducting cavities and cryomodules in the DALS Test Facility (DATF). The major components include Test Facility Cryoplant (TFCP), Process Vacuum Pump System (PVPS), Auxiliary System (e. g. helium recovery & purification system), and Cryogenic Distribution System (CDS). The DATF will be arranged mainly in the Cryomodule Test Bench Hall, Injector Test Bench Hall and Cryo-Hall. The Module Test Bench Hall will house the cold box, Test © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1006–1012, 2023. https://doi.org/10.1007/978-981-99-6128-3_130

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Facility Distribution Box (TFDB), HTB, VTC. The ITB will be located in the Injector Test Bench Hall. Meanwhile, the Cryo-Hall will contain the warm compressor system (WCS), PVPS, helium recovery & purification system [1]. The layout of the entire DATF is shown in Fig. 1.

Fig. 1. Layout of the DATF

In order to automatically monitor the operating mode and status of the DATF cryogenic system, it is necessary to design a cryogenic instrumentation and control system to control and monitor the devices. The number of sensors and actuators are shown in Table 1. Table 1. Number of sensors and actuators. Instrument

Quantity

Pressures

76

Temperatures

527

Liquid levels

12

Flowmeters

15

Valves

104

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2 Control System Overview With the increasing requirements for minimizing human intervention to address the complexity of the process, the control system of large scale of cryogenic equipment has gradually evolved towards fully automatic operation. The cryogenic control system is designed to integrate all major subsystems in the DATF cryogenic system, including Auxiliary System (e. g. helium recovery & purification system), PVPS, TFCP, TFDB, ITB, HTB, VTB, etc., and to implement the control of all devices. The control system of the DATF cryogenic system is mainly composed of the Siemens Process Control System SIMATIC PCS7 and EPICS. The TFCP had been purchased from LINDE, which mainly used Siemens PLCs (S7–1500) and WinCC software to control the refrigerator. The control of PVPS and helium purification system, supplied by domestic manufacturers, is also managed by Siemens PLCs. The control systems in the CDS adopt Siemens S7– 1500/ET200M model PLCs. All subsystems of the DATF cryogenic system are integrated through the PROFINET network. EPICS is mainly used for information interaction with the main control system of the accelerator. The control system will provide the appropriate control algorithms to support the complex operation of the cryogenic system, enabling fully automated or semi-automated modes of operation. The supervisory control can be achieved in this control system, such as visualization of the process hierarchy. All parameters essential for operation in the DATF cryogenic system can be monitored and controlled. The system will allow for equipment protection and personnel protection through a hierarchy of alarms, interlocks, troubleshooting and tripping. And it can also realize the information interaction with central systems in the accelerator.

3 The Architecture of Control System In order to implement the necessary functions and integrate the subsystems, the control system of the DATF cryogenic system adopt the standard three layers distributed control architecture [2, 3]. It includes the supervision layer, control layer and device layer. The architecture of the control system for the DATF cryogenic system is shown in Fig. 2.

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Fig. 2. The architecture of the control system for the DATF cryogenic system

3.1 Supervision Layer The supervision layer is an important tool for operator to monitor, develop and commission the system. In this project, the supervision layer is equipped with two data servers (including one redundant server), eight operator stations and one mobile workstation, and contains control software, human-machine interface, data management, and remote communication module. The entire DATF cryogenic system is supervised through Data Servers (DS) running the Siemens WinCC SCADA system. In this layer, the system can implement the data archiving and display, visualization of the process hierarchy, access to the interlocks between devices, automatic adjustment of the control loops, direct access to the device files, etc. The master computer communicates with the PLC controllers through PROFINET protocol, ensuring the real-time communication and enabling rapid diagnosis functions. It also provides the dedicated interfaces to other control systems in the accelerator. 3.2 Control Layer The control layer is the core of remote device control, logic operation and signal acquisition and conversion. And it also provides safety interlock protection and hierarchical alarm for the system. The control layer continuously monitors the status of the cryogenic system and updates the control commands to maintain stable operation. In this project, the control layer is composed of Siemens S7–1500 PLCs and EPICS. The control duties are mainly executed within PLCs. Almost all process variables and status information of actuators and sensors are monitored and controlled by PLCs. The signal conversion, control algorithms (e. g. PID algorithms) and processing of variables are also implemented by PLCs.

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WINCC and Siemens PLC are mainly used to monitor the cryogenic control system, while the EPICS is used in the entire accelerator control system. These are two relatively independent control systems. In order to reduce the interference between the two systems, EPICS in the cryogenic control system is used only to provide a set of operating interfaces in the center control of accelerator system, to alarm and archive, and to analyze real-time and historical data. Therefore, the PLCs in the cryogenic control are also used to send the signals from sensors and actuators to the EPICS Input/Output Controller (IOC). In this project, each subsystem in the cryogenic control has its own control cabinet which is equipped with a switch to establish the Device-level Ring (DLR). The PLCs with HMI are connected to the switch based on the PROFINET protocol, while the switches communicate with each other through fiber optics. In addition, the TFCP control system is connected to the main control cabinet switch through PROFINET protocol, the recovery compressor control system is linked to the horizontal test bench switch through MODBUS protocol, the recovery & purification and PVPS control system are connected to the tank zone switch through PROFINET protocol. And the EPICS communicates with the main control cabinet by the optical fibers. 3.3 Device Layer The device layer primarily consists of various types of instrumentation in the field, such as industrial sensors (pressure, temperature, flow rate, liquid level, etc.) and actuators, for data acquisition and information interaction with the PLC. The device I/O signal numbers are listed in Table 2. Table 2. I/O signal numbers. (AI: Analog Input; AO: Analog Output; DI: Digital Input; DO: Digital Output; RTD: Resistance Temperature Detector.) Control cabinet

AI 4-20mA

AI RTD

AI CERNOX

AO

DI

DO

TFDB

55

43

16

35

8

28

VTC

23

8

36

8

1

1

HTB

40

26

150

13

1

1

ITB

54

34

183

21

2

2

Tank zone

40

10

0

24

8

1

Each field device is controlled remotely from the control room to minimize the workload of field personnel and to achieve a higher level of automation. Local manual control can be used when equipment requires maintenance or incident handling. Most devices communicate with PLCs through 4–20 mA signal lines. The temperature sensor and some valve positioners use PROFIBUS-DP protocol. The communication protocols used by each device is shown in Fig. 3.

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Fig. 3. The communication protocols used by devices

Instrument Selection. At present, we have completed the selection and procurement of most of the instruments, as shown in Table 3. Table 3. Instrument selection. Type

Description

Manufacturer

Pressure

PAA

KELLER

Level

1700

AMI

Flow

Room:780i/EL-FLOW

SIERRA/BRANK HORST

Cryogenic: Coriolis 1700

EMERSON

Temperature

CERNOX

LAKESHORE

Temperature Monitoring A number of temperature sensors are installed in the DATF cryogenic system. In order to reduce the communication and calculation in the PLCs, we choose LAKESHORE 240-8P as temperature monitoring [4]. The CERNOX sensors are connected to the 240-8P modules in the corresponding control cabinet, and then the 240-8P modules communicate with the PLC controllers through the PROFIBUS -DP protocol. Valve Positioner The valve positioners used in radiation environment require special protection. We choose Siemens SIPART PS2 as the valve controller. Each module communicates through PROFIBUS-PA protocol, which needs to be converted to PROFIBUS-DP protocol through a DP/PA Coupler and then communicates with the PLC.

3.4 Interfaces PLC to EPICS IOC In this project, the s7nodave driver based on Asyn and libnodave is selected to implement the communication between Siemens PLCs and EPICS IOCs. The driver and the PLCs exchange data via the ISO-TCP protocol. In this process, the EPICS records specify the memory address in the PLCs and the driver uses the ISO-TCP protocol to read or write the channel data [5].

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PLC to Other Accelerator Subsystems The PLCs and the master computer are connected through the switch in the main control cabinet via an optical fiber. The PROFINET protocol is used to enable communication between the accelerator main control system and the cryogenic test facility control system. The local control of the cryogenic system realized by the PLCs is all done by the control system of the DATF cryogenic system. The cryogenic control system uploads the data to the EPICS network, and other subsystems of the accelerator can be extracted through the EPICS network if they need the data of the cryogenic system.

4 Conclusion Up till now, the control system for the DATF cryogenic system is being designed and built. The hardware design has been completed, and the software design is in progress. The programming of the PLC side will be realized in the TIA Portal software, and the interface design and alarm display in the supervision layer will be implemented in the WINCC software. According to the project plan, the commissioning of the DATF Cryogenic system will be taken place in early 2023. Acknowledgments. The project is supported by the Dalian Municipal Government. The authors would like to express our gratitude to Prof. Guy Gistau Baguer and John Weisend II for their constructive comments and advices throughout process. We are also grateful that staff from IHEP, IASF and SHINE for their very valuable knowledge and experiences by hosting our visits or video meetings.

References 1. Sun, Z., Huang, L., Shi, X., Wang, X.L.: Conceptual design of DALS test facility cryogenic system. CEC.ICMC (2021) 2. Pezzetti, M.: Control of large helium cryogenic systems: a case study on CERN LHC. EPJ Tech. Instrum. 8.1 (2021) 3. Geyang, J., et al.: The cryogenic control system of SHINE. EPJ Tech. Instrum. 8.1 (2021) 4. Mattison, K., Boyes, M., Cyterski, M., Fairley, D., Lam, B.: LCLS-II cryomodule and cryogenic distribution control. In: International Conference on Accelerator and Large Experimental Physics Control Systems, Barcelona (2017) 5. s7nodave Device Support for EPICS website, https://oss.aquenos.com/epics/s7nodave/

Development of High-Density Signal Feedthroughs for Liquid Argon Calorimeters Maria Barba1(B) , Michel Chalifour1 , Martin Aleksa2 , and Johan Bremer1 1

CERN, TE-CRG-CL, 1211 Geneva 23, Switzerland [email protected] 2 CERN, EP-ADO, 1211 Geneva 23, Switzerland

Abstract. The next generation calorimeters for collider experiments will have to feature high granularity, very good energy resolution and a strong control over the systematic uncertainties and the acceptance. Recently, a highly granular noble liquid sampling calorimeter fulfilling these requirements has been proposed as a possible FCC experiment. It will operate at cryogenic temperatures and will be housed in a cryostat. The higher read-out granularity leads to a strong increase in the number of read-out channels and, consequently, to the need for high-density signal feedthroughs that ensure the passage of the signals from the cold calorimeter to the warm electronics outside. To be able to transfer up to 30 000 signal wires per feedthrough, new high-density flanges have been designed combining fiberglass laminate structures, epoxy glue and R flexible strip cables. Flange samples have indium seals with Kapton been tested at high pressure and low temperature, simulating the extreme conditions that can be reached in a liquid argon calorimeter such as the ATLAS electromagnetic calorimeters [1], used as reference. Details of the high-density flange design as well as experimental test results are presented and discussed. Keywords: signal feedthroughs · noble liquid calorimetry argon calorimetry · FCC experiments

1

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Introduction

Noble liquid calorimeters with high granularity have been proposed for FCC-ee [2] and FCC-hh [3] experiments. One of the consequences of this high granularity is the increase of signal channels, making the signal extraction to the outside of the cryostat one of the main challenges of this type of calorimeters. The present project focuses on the design and development of new connector-less high-density signal feedthroughs for the signal extraction to the warm electronics outside. As a reference, the existing signal feedthroughs of the ATLAS electromagnetic calorimeters are able to transfer up to 1920 signal channels per feedthrough [1]. Distributed along the perimeter of the cryostat, these signal feedthroughs are c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1013–1019, 2023. https://doi.org/10.1007/978-981-99-6128-3_131

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composed of a cold flange, separating the liquid argon from a vacuum zone, and a warm flange, separating the vacuum volume from the outside of the cryostat, as illustrated in Fig. 1. Pigtail coaxial cables connect the mother boards inside the calorimeter to the cold flange. On the other side, vacuum flexible strip cables ensure the signal transfer from the cold to the warm flange. The signal transfer through both flanges is ensured by a connector at the end of each cable. The body of the feedthrough is stainless steel made and has a diameter of 300 mm. Based on the existing signal feedthroughs of ATLAS, the flanges have now been re-designed to allow to transfer a higher number of signal cables (more than x10). This paper reports on the conceptual design of a high-density feedthrough flange, as well as on the experimental results of a reduced size sample tested under high pressure and low temperature conditions, simulating the extreme conditions that can be reached in the detector.

Fig. 1. Signal feedthrough of the ATLAS LAr EM barrel calorimeter.

2

High-Density Feedthrough Flanges

To reach a higher density of signal wires/cm2 , a high-density connector-less flange has been designed as illustrated in Fig. 2. This latter is composed of a stainless steel body grid with a diameter of 600 mm and a thickness of 20 mm. The spaces are filled with square structures (called “structural material” and indicated in blue in Fig. 2) with a high number of slits (30 mm long and 1 mm wide) and a maximum distance of 3 mm between each slit. In this case, the structural material used is fiberglass laminate in epoxy resin, also known as

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R G10. Kapton flexible strips, representing the vacuum flexible strip cables used in the ATLAS feedthroughs but without the connectors, pass through the slits of the structural material. The slits are then filled with epo-tek T7-110 [4] glue, ensuring the leak-tightness between the strips and the G10 structure. Finally, the leak-tightness between the structural material and the main flange body grid is ensured by an indium seal, also used in the past in [5]. More precisely, an indium wire of 1.5 mm of diameter is compressed by the structural material with a stainless steel compression plate. Several flange designs, where the distance between the slits can vary from 2 to 3 mm and the dimensions of the square structures containing the slits can vary from 96 × 65 mm to 217 × 155 mm, have been proposed and studied. In the present paper, experimental results obtained after testing a sample illustrated in Fig. 3 representing a single square structure of 96 × 65 mm and a distance of 3 mm between each slit are presented.

Fig. 2. Conceptual design of a high-density connector-less feedthrough flange.

3

The Cryogenic Experimental Facility

To simulate the operating conditions as well as the extreme conditions that can be reached in a liquid argon calorimeter, an experimental setup has been conceived to perform pressure and leak tests on flange samples at different temperatures. Based on the ATLAS EM calorimeters [1], the operating temperature is 87 K and the maximum pressure is 3.5 bar. The experimental setup, presented in Fig. 4, is composed of a main vacuum chamber (400 mm diameter and 1500 mm mm length) containing a stainless steel vessel (200 mm diameter and 300 mm length), with a copper pipe welded on its surface. The vessel is filled with gas helium (GHe) coming from a bottle and can be emptied via a pump of the central pumping circuit of the laboratory.

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Fig. 3. Schematic illustration of high-density connector-less flange sample.

The vessel is cooled by liquid nitrogen flowing through the copper pipe from the bottom to the top part. Furthermore, a double-wall thermal shield filled with cold gas nitrogen coming from the end part of the mentioned liquid nitrogen (LN2 ) circuit, surrounds the GHe vessel. Ten sheets of superinsulation (MLI) cover the top plate holding the thermal shield structure to face radiation inputs from the top flange of the vacuum chamber. A pumping group composed of a primary pump and a turbopump ensures a vacuum between 1×10−4 mbar at room temperature and 1×10−7 mbar at low temperature in the vacuum chamber. Moreover, a leak detector is connected at the end of the pumping circuit to detect any possible leak coming from the GHe vessel. The experimental setup is monitored with one PFEIFFER vacuum sensor, one WIKA TRONIC pressure sensor for the GHe circuit and the vessel, and up to 10 temperature sensors (Pt100 type in a 4-wire configuration) distributed between the GHe vessel, the liquid nitrogen circulation and the thermal shield. All the sensors are monitored with a LabView program. The data acquisition system allows also to regulate the temperature of the of the LN2 circuit at 87 K or below. In addition, a security system allows to stop the liquid nitrogen circulation in case of a breaking vacuum accident. Before performing every test, the vessel is purged three times with gaseous helium at room temperature. Then, the vessel is filled with 3.5 bar of GHe and the leak detector is connected to the pumping system of the vacuum chamber. The evolution of the temperature, the pressure in the vessel, the leakage, as well as the vacuum are measured and registered during 30 mins. After performing this leak and pressure test at room temperature, the cooling down starts by activating the liquid nitrogen circulation. This second step can last up to 7 h. During this process, the temperature and vacuum values are permanently registered but the leak detector is connected every hour during a few minutes to avoid any damage to the detector due to a long term exposure to helium in case of a leak. When the temperature at the bottom part of the vessel, where the flange sample is located,

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reaches a temperature below 87 K, the vessel is re-filled with GHe to 3.5 bar and the leak and pressure test at low temperature starts. Also in this case, the evolution of the temperature, the pressure in the vessel, the leakage, as well as the vacuum are measured and registered during 30 mins. After this, the warming up process can start by just stopping the liquid nitrogen circulation. The entire setup will warm up reaching room temperature in less than 2 days. Also during this process, the temperature, pressure and vacuum values are registered.

Fig. 4. Schematic illustration of the experimental setup.

4

Experimental Results

Figure 5 illustrates the evolution of the temperature of the flange sample as well as the pressure in the vessel during the leak and pressure test at room temperature. The average leak rate during the test is 2.25 × 10−8 (± 0.05 × 10−8 ) mbar.l/s. The stability of the pressure measurement as well as the leakage values show that the sample was not leaking and the vessel was not losing GHe during the entire test. Similarly, Fig. 6 illustrates the evolution of the temperature of the flange sample as well as the pressure in the vessel during the leak and pressure

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test at low temperature. In this case, the average leak rate during the test was lower due to the lower temperature: 3.15 × 10−9 (± 0.15 × 10−9 ) mbar.l/s. Again, the stability of the pressure measurement as well as the leakage values show that the sample was not leaking and the vessel was not losing GHe during the entire test. The test has been performed at around 84 K to be below 87 K but before reaching 83.8 K, corresponding to the melting point of argon, which is not representative for the liquid argon calorimeter application. In addition, the vacuum and leakage values measured during the cooling down and warming up processes, go from 8.5×10−4 mbar and 2.25 × 10−8 mbar.l/s at room temperature to 7.84 × 10−7 mbar and 3.15 × 10−9 mbar.l/s at low temperature, respectively.

Fig. 5. Evolution of the pressure ( ) and the temperature ( ) during a pressure test at room temperature with an average leak rate of 2.25×10−8 (± 0.05×10−8 ) mbar.l/s.

Fig. 6. Evolution of the pressure ( ) and the temperature ( ) during a pressure test at low temperature with an average leak rate of 3.15×10−9 (± 0.15×10−9 ) mbar.l/s.

Signal Feedthroughs

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Conclusions

High granularity requirements in liquid argon calorimeters for future FCC experiments lead to a strong increase in the number of wires extracting the signal from the calorimeter to the outside electronics. To deal with this, new highdensity connector-less feedthrough flanges have been designed. A reduced size flange sample made of fiberglass laminate in epoxy resin (G10) and epo-tek T7110 glue, and using an indium seal, has been constructed and tested. Experimental results have shown the performance of the different components under high pressure and low temperature conditions, demonstrating the viability of the connector-less flange concept for liquid argon calorimeters.

References 1. ATLAS Collaboration.: ATLAS liquid-argon calorimeter: Technical Design Report. CERN-LHCC-96-041 (1996) 2. Aleksa, M., Bedeschi, F., Ferrari, R., Sefkow, F., Tully, C.G.: Calorimetry at FCCee. Eur. Phys. J. Plus 136(10) (2021) 3. Aleksa, M., et al.: Calorimeters for the FCC-hh. CERN-FCC-PHYS-2019-0003 (2019) 4. Perin, A., Macias Jare˜ no, R., Metral, L.: Study of materials and adhesives for superconducting cables feedthroughs. LHC Project Report 504. In: Cryogenic Engineering Conference and International Cryogenic Materials Conference CEC/ICMC, Madison, Wisconsin (2001) 5. COMPASS Collaboration: The COMPASS experiment at CERN. In: Nuclear Instrument and Methods in Physics Research, Section A, vol. 577, pp. 455-518 (2007)

Helium Cryogenic Control System Design for CSMC Testing in CRAFT Changheng Xie1,2 , Zhiwei Zhou2(B) , and Qiang Yu2 1 Anhui University, Hefei 230031, China

[email protected]

2 Institute of Plasma Physics, and Hefei Institutes of Physical Science,

Chinese Academy of Sciences, Hefei 230031, China {zzw,qiang.yu}@ipp.ac.cn

Abstract. This paper focuses on the 1kw/4.5K helium refrigerator applied to Central Solenoid Model Coil (CSMC) test magnets under the Comprehensive Research Facility for Fusion Technology (CRAFT). The structure of CSMC and the process of cryogenic distribution system are briefly introduced. It mainly introduces the design of the control system under the CODAC (COntrol, Data Access and Communication) architecture through the analysis of the specific operation conditions required for testing the magnets including of the redundant control network, interlock function, and communication configuration and OPI (Operator Interface), etc. Through structural analysis and research on typical cryogenic applications, the purpose is to standardize cryogenic control logic design and program development, and improve the readability, scalability and maintainability of large-scale programs. System commissioning is scheduled to begin in July 2022. Keywords: CRAFT · Cryogenic control · Control design · CODAC

1 Introduction Led by the Chinese Academy of Sciences, CRAFT (Comprehensive Research Facility for Fusion Technology) is a major scientific research project, mainly carry out superconducting magnet and divertor research, which provides technical support for the development of fusion reactor design, cryogenic system and superconducting technology research and development [1]. Fusion reactor superconducting magnets include Toroidal Field (TF) coils, Central Solenoid (CS) coils and Poloidal Field (PF) coils. In order to meet the performance research of Nb3Sn superconducting magnets in high current and high field strength environments, the Central Solenoid Model Coil (CSMC) was constructed. CSMC consists of five superconducting coils in total, including Nb3Sn for high field and NbTi for low field. The magnets are planned to be forced-flow cooled by a helium circulating pump under 4.5K supercritical helium [2]. The forced-flow cooling of supercritical helium design maximum flow rate of 200 g/s, the advantage of using forced flow cooling is that the flow can be adjusted by changing the speed of the helium © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1020–1027, 2023. https://doi.org/10.1007/978-981-99-6128-3_132

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circulation pump, the thermal load is slowly transmitted to the refrigerator, effectively avoiding the heat load of the refrigerator to produce large fluctuations, improve the stability margin of the refrigerator [3]. This paper introduces the cooling process of the CSMC conductor, the cooling mode of the refrigerator, and the progress of the refrigerator control system based on the CODAC architecture.

2 Cryogenic System 2.1 Cryogenic System for the CSMC The refrigeration capacity of the main body of the system is provided by a 1 kW/4.5K helium refrigerator, which mainly includes recovery compressor, main compressor, cold box and CSMC test valve box. The refrigerator provides refrigeration mode, liquefaction mode and blend mode. The liquefaction rate at cryogenic is about 300 L/hr. The Fig. 1 shows the simplified process flow of the refrigerator. The Table 1. Shows the I/O number of control system.

Fig. 1. Simplified Process Flow of the 1kW/4.5K Helium Refrigerator.

Table.1. Number of I/O Points of the Refrigerator Subsystem. Item

DI

DO

AI

AI (RTD)

AO

Total

Oil Removal

16

16

24

8

8

72

Cold Box

112

64

72

32

32

312

CSMC Distribution Valve Box

39

16

32

24

32

143

Auxiliary System

48

64

48

24

40

224

Total

215

160

176

88

112

751

The test valve box mainly includes liquid nitrogen cold shield, liquid helium tank, heat exchanger, helium circulating pump and superconducting magnet. The liquid helium tank contains three coil heat exchangers, which are cooled by liquid helium immersion to absorb the heat load of the magnet and the helium circulating pump.

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2.2 Brief Process of Cooling Mode The Fig. 2 shows the brief process of magnet cooling.

Fig. 2. Brief process of magnet cooling.

The cooling process of the CSMC test conductor mainly includes the following modes [4]: (1) 300K-80K pre-cooling mode After being pressurized by the helium compressor, the high-pressure helium gas in the main flow path of the refrigerator is divided into two paths. One path first completes heat exchange with the liquid nitrogen in the liquid nitrogen tank through CV2002, and the other path of normal temperature helium gas is mixed with cooled helium by regulating CV2008. Mixed gas can realize the adjustment of helium in the temperature range of 300K-80K, and pre-cool the CSMC conductor, and the cold gas at the CSMC outlet returns to the 80K cold flow through CV2131. (2) 80 K-4.5 K cooling mode After the pre-cooling of CSMC magnet is completed, the refrigerator turbine is started. Before entering the valve box, the mainstream helium is divided into two ways, one way through CV2251 into the third stage of turbine and the other way is connected to the helium trough through the bypass valve CV2250. When the helium tank level reaches 15% and the conductor temperature drops to 4.5K, the CV2001 can be turned on to perform the pre-cooling of the helium circulating pump, and the return gas from the CSMC conductor outlet returns to the refrigerator through the CV2132.

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(3) 4.5K steady-state operation mode After the pre-cooling of the helium circulating pump is completed, the refrigerator continues to produce liquid helium until the helium tank level reaches 60%. At this time, the helium circulating pump is opened to cool the CSMC test conductor, and wait until the internal flow and temperature of the magnet reach the established requirements, and the experiment begins.

3 Control System Design The control system is a CODAC (COntrol, Data Access and Communication) software architecture based on EPICS (Experimental Physics and Industrial Control System) [5]. The Fig. 3 shows the architecture of CODAC.

Fig. 3. The CODAC Plant system I&C Architecture.

The main structure includes mini-CODAC which is a computer providing a reduced set of CODAC services and uses as a CODAC terminal. PSH (Plant System Host) is a computer providing general purpose management functions and a unified configuration and monitoring interface to CODAC for a given plant system I&C (Instrumentation and Control). The slow controllers mainly used in the cryogenic system which are based on the Siemens SIMATIC S7 PLC controller and providing control functions for processes with slow feedback. 3.1 Hardware Architecture Considering the stability of the system, the main controller selects S7-400H with redundancy, the CPU has PROFIBUS DP or PROFINET interface, as a centralized controller of distributed I/O. The auxiliary system controller is S7–1500, which communicates with the main controller through PROFINET. The subsystem selects ET200M distributed I/O controller and field layer connection selects PROFINET or PROFIBUS. The cryogenic acquisition module is M240, the compressor and temperature acquisition module are connected to the 400H by using PROFIBUS-DP. The Fig. 4 shows the hardware architecture of the control system.

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Fig. 4. Hardware architecture of the control system.

3.2 Redundant Network Configuration and Interlock Function Redundancy functions include CPU redundancy and I/O redundancy. This system redundancy is CPU redundancy, the redundant package includes two 414-5H CPUs, one as a master CPU and one as a redundant backup. Connecting the two through the synchronous module interface, it is established in a subnet in the STEP7 configuration, then assign the corresponding IP address according to the MAC code, assign names to the ET200M and verify. The above steps can establish a redundant control network for a distributed I/O. The interlock function is mainly used to restrict the operation of other associated devices when certain alarms are generated. Through some comparison of thresholds, or Boolean signals for the operation of specific equipment, the corresponding interlock alarm signal is generated, which is planned to be implemented in CFC (Continuous Function Chart) and GRAPH. 3.3 Communication Configuration ITER (International Thermonuclear Experimental Reactor) provides relevant tools to simplify the communication configuration between CODAC and PLC, through the SPSS (The Standard PLC Software Structure) generated blocks as a tool, using the binaries or sources mode, to achieve the communication between the PLC and the CODAC [6]. SPSS includes standardized FB, FC, DB and other STEP7 source files, which are first imported into STEP7. SDD (Self-Description Data) is a development tool which is used to configure CODAC, by completing the configuration of PV signals in SDD, SCL source files will be generated, after compiled, FC103 can be called in OB1 to complete the brief functions of communication. The main data blocks responsible for communication in SPSS are DB100 and DB101. The DB100 contains all input signals, such as hardware interface acquisition signals. The DB101 is responsible for all output signals, such as the valve opening after processing. Variables can be referenced from DB100 into the control program and mapped to DB101. Both DB100 and DB101 are called in FC103. Meanwhile, ITER divided and supplemented the corresponding block numbers, the divided functions are: CODAC Interface (SPSS), PLC Core Application, Hardware

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Access Layer, Exchange Data Blocks, SIEMENS Internal and Basic Software Architecture [7]. By classifying block numbers and functions, it is convenient for later system maintenance and integration.

4 Control Logic Design 4.1 Control Logic Classification The control system is divided into continuous control logic, supervisory control logic and interlock control [8]. The Fig. 5 is a schematic diagram of logical classification.

Fig. 5. Schematic diagram of control logic classification.

Continuous control logic Continuous control logic is mainly aimed at the control of conventional field equipment. After reading AI signals, it enters different functional blocks to realize the corresponding control algorithm. Function blocks can be written in the SCL (Structured Control Language) to complete the continuous control logic of the system. Supervisory control logic The supervisory control logic includes the development of the OPI interface and the implementation of the sequential control. The OPI interface is drawn out in the CSS (Control System Studio). Sequential control is used for automatic switching under multirun models, and it is planned to be programmed in GRAPH. Interlock control logic After the PLC collects the corresponding signal, according to the set threshold, when the signal is lower than or higher than the threshold, the corresponding alarm signal is triggered. Through the logical combination of different numbers of alarm signals, the on-site controller is driven to complete the corresponding actions. The Fig. 6 shows the composition and implementation of the control logic.

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Fig. 6. Implementation of Control Logic in CODAC and PLC.

4.2 OPI Interface OPI layer is being developed in the Control System Studio (CSS) tool to complete functions such as monitoring interface, data storage and alarm processing. At present, the OPI of the main compressor, oil removal system, cold box, and CSMC conductor interface has been completed. The AI and PID control panel should be designed later, and the interface display parts will be adjusted according to the debugging situation in CSS. The Fig. 7 shows the OPI interface of CSMC conductor.

Fig. 7. OPI design of CSMC test magnet in CSS.

Acknowledgments. This work is supported by The Comprehensive Research Facility for Fusion Technology Program of China under Contract. (No.2018–000052-73–01- 001228).

References 1. Institute of Plasma Physics, Chinese Academy of Sciences. comprehensive Research Facility for Fusion Technology [EB/OL], http://craft.ipp.ac.cn/cn/jsyyy/index_26.aspx 2. Cheng, A., Zhang, Q., Fu, B., et al.: CFETRCS model coil forced flow cooling process design. Vacuum Cryogenic, 21(4), 246-246 (2015)

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3. Cheng, A., Zhang, Q., Fu, B., Lu, X.: Process design of cryogenic distribution system for CFETR CS model coil. Plasma Sci. Technol 18(2), 202–205 (2016) 4. Lu, X., Zhang, Q., Cheng, A., et al.: Conceptual design of cryogenic distribution control system for CFETR Central Solenoid model coil. In: 2017 36th Chinese Control Conference (CCC) (2017) 5. Li, F., Zhang, J., Li, L.: Progress in EPICS and its application. Electron. World, (8), 3 (2021) 6. Bhimvarapu, A., Evrard, B.: Standard PLC Software Structure (SPSS) User Manual. France: ITER CODAC Team (2019) 7. Simelio, A., De Frutos Bolzoni, L.. Kusurkar, A.: PLC Software Engineering Handbook. France: ITER CODAC Team (2019) 8. Kadlec, M.: 34.0I. PC - IOTN – 000103: PLC SW Application architecture guidelines for PBS 34. France: ITER CODAC Team (2017)

Cryogenics for Superconducting Materials, Devices and Systems

Thermal Hydraulic Analysis of Background Magnet of 15T Conductor Performance Test Facility Libiao Hu1,2 , Qiangwang Hao1(B) , Yi Shi1 , Yu Wu1 , Kaihong Wu1,2 , and Chao Dai1 1 Institute of Plasma Physics, Chinese Academy of Science, Hefei, People’s Republic of China

[email protected] 2 University of Science and Technology of China, Hefei, People’s Republic of China

Abstract. The 15T Conductor Performance Test Facility (Super-X for short) is one of the comprehensive research facilities for fusion technology (CRAFT). The background magnet is wound by Nb3 Sn cable-in-conduit conductors (CICC) with a maximum magnetic field of 15.6 T, and is cooled by 4.3 K supercritical helium forced flow. The background magnet needs to work in a liquid helium temperature environment, in this paper, according to the heat load and cooling loop layout of the background magnet, thermal hydraulic analysis was carried out on the background magnet under stable operation, excitation and quench conditions to ensure the cooling scheme is reasonable. Keywords: Superconducting magnets · CICC · Thermal hydraulic analysis

1 Introduction Super-X is an important part of CRAFT. The two main scientific objectives of Super-x are confirmed. On the one hand, it is to study and test the electromagnetic behavior of superconducting conductors under multi-field coupling condition. On the other hand, it is to evaluate the reliability and safety of superconducting conductors in fusion reactor operation environment. Therefore, it is very important for the development of future superconducting fusion reactors. The background magnet of Super-X is large in size and has a high operating current. The operating mode is steady-state operation with a low field coil operating current of 14 kA and a high field mid-field coil operating current of 8.5 kA, with a total stored energy of 497 MJ. There is a risk of overheating and overpressure once quench occurs, so a thermal hydraulic analysis of the excitation and quench process is required to ensure the stability of the magnet.

2 Magnet Layout and Thermal-Hydraulic Parameters The background magnet of Super-X is a left-right separated structure with three coils on one side, as shown in Fig. 1, coil C is the high field coil (HFC), coil B is the medium field coil (MFC), and coil A is the low field coil (LFC). LFCs are pancake coils, HFCs © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1031–1037, 2023. https://doi.org/10.1007/978-981-99-6128-3_133

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and MFCs are layer coils. In the design of the cooling loop of superconducting coils wound by CICC, taking into account the length of the conductor and the location of the joints, the helium inlet and outlet locations are generally designed at the beginning or end of a pancake or layer for easy manufacturing and installation. A-1

A-32

LFC

B-6 MFC

C-4 C-3 C-2 C-1

B-1 HFC

Fig. 1. Background magnet and cooling channels layout

In order to take full advantage of the performance of superconducting conductors and reduce the cost, three conductors are developed for HFC, MFC and LFC. The Fig. 2 shows the cross section of the CICCs used for the background magnet, and thermal-hydraulic parameters of the conductors are shown in Table 1. 15.2

14 14.8

High Jc Nb3Sn cable

4-R2

316LN jacket

316LN jacket

24.6

30.8

316LN jacket

1.8

4-R2

High Jc Nb3Sn cable

2.5

33.5

4-R2

High Jc Nb3Sn cable

2.6

Fig. 2. Cross section of CICCs used for background magnets (mm)

Table 1. Thermal-hydraulic parameters of the conductors Items

Unit

HFC

Strand diameter

mm

MFC

LFC

0.82

0.82

0.82

Cable layout

3sc × 4 × 5 × 6

(1sc + 2cu) × 4 × 4 ×5

(1sc + 2cu) × 4 × 5 ×6

Number of strands

360

240

360 (continued)

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Table 1. (continued) Items

Unit

HFC

MFC

LFC

Operating current

kA

8.5

8.5

14.0

30%

30%

30%

30.8 × 14

14.8 × 24.6

33.5 × 15.2

Void friction Conductor layout

mm

The design of the cooling loop is based on the layout of the background coil, and then the heat load and current sharing temperature (Tcs) of the background coil are calculated. The temperature margin is ensured to be greater than 1.5K [1]. The final determined parameters of the cooling loop are shown in Table 2. Table 2. Parameters of the cooling loop Cooling channel

A1–32

B1–6

C1–4

Current sharing temperature (K)

6.6

6.4

6.3

Temperature Margin (K)

1.5

1.5

1.5

Inlet pressure (bar)

5.0

5.0

5.0

Inlet temperature(K)

4.3

4.3

4.3

Mass flow rate(g/s)

1.0

0.8

1.5

1.3

1.2

1.1

Pressure drop (bar)

0.5

0.8

0.5

0.5

0.5

0.5

Number of cooling channels

64

12

2

2

2

2

Length (m)

224

225

124

147

174

198

3 Excitation Analysis Superconductors will have energy losses when it works with alternating current or magnetic field. Energy loss generates heat, leading to temperature rise. During the excitation process, the backfield magnet generates hysteresis losses and coupling losses. According to the Bean model, the current density distribution inside the superconductor is the critical current density Jc or zero, and Jc remains constant in the region penetrated by the magnetic field. The hysteresis loss is calculated by integrating the magnetization curve. Practical superconducting wires are usually multifilament composite structures. CICC consists of multiple superconducting strands and copper strands twisted together, and the superconducting strands contain many superconducting filaments. The hysteresis loss power of the CICC is the sum of the hysteresis loss power of each filament [2], so the hysteresis loss power per unit length of CICC can be expressed as: Ph =

2 Jc Asc B˙ 3π

(1)

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where deff is the effective diameter of the superconducting filament, 20 μm for LFC, 40 μm for MFC and HFC, Asc is the cross-sectional area of the superconducting portion of the CICC, and B˙ is the changing rate of the magnetic field. The coupling loss power of CICC in the transverse magnetic field is related to the square of the magnetic field change rate [3], as shown in Eq. 2: P=

2θ ˙ 2 Be μ0

(2)

where 2θ is the total time constant, which is 200ms in the calculation, B˙e is the changing rate of the External magnetic field, and μ0 is the permeability of vacuum. P is the coupling loss power per unit volume. The loss power shown in Table 3 is used as an external heat source for calculations in the GANDALF software to obtain the temperature variation in the magnet cooling channel during the excitation process. Table 3. Energy loss of the background magnet during excitation Excitation time (s) Hysteresis loss (W)

Coupling loss (W)

2800

1400

LF

8.52

17.04

MF

6.71

13.42

HF

13.09

26.18

LF

1.52

6.09

MF

0.29

1.17

HF

0.30

1.22

Analytical calculations were performed for the C1 cooling channel of HFC, which has the highest magnetic field and the smallest stability margin. The temperature variation of the C1 channel during the excitation process is shown in Fig. 3. The excitation time is 1440s for the left figure and 2800s for the right.

Fig. 3. Temperature of C1 channel during excitation

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It can be seen that the excitation speed and the temperature rise of the magnet are positively correlated. The temperature margins for the two excitation speeds at the end of excitation are shown in Fig. 4. During the excitation process, the temperature margins are all greater than 1.9 K.

Temperature margin (K)

3.0

Δt=1400 Δt=2800

2.8 2.6 2.4 2.2 2.0 1.8

0

20

40

60

80

100

120

Length (m)

Fig. 4. Temperature margin of C1 channel at the end of excitation

4 Quench Analysis The background magnet is powered by two sets of power equipment, HFC and MFC are connected in series with a current of 8.5 kA, and LFC with a current of 14 kA. The stored energy of the background magnet is 497 MJ, so it is necessary to thermodynamically analyze the quench process to avoid irreversible damage to the magnet. The quench analysis was calculated for the cooling channel (C1, B1, A20) with the highest magnetic field in the HFC MFC and LFC respectively, the quench detection time was 2 s. The current discharge curve after quench is shown in Fig. 5, The minimum quenching energy of mechanical disturbances in the C1, B1 and A20 cooling channels is shown in Table 4. The disturbance time is 1 ms and length is 1 cm [4]. Figure 6 shows the maximum pressure and temperature of the cooling channel with the largest magnetic field in HFC, MFC and LFC in the quench process under the effect of mechanical disturbance. The analysis shows that the hot spot temperature of the LFC is the highest, around 100 K, and the pressure of the HFC is the highest, around 12.1 Bar. The initial normal zone length (LINZ ) has a direct effect on the maximum pressure and temperature in the cooling channel during quenching. It is assumed that the LINZ is in the middle of the cooling channel. As shown in Fig. 7, according to the analysis and calculation results, the maximum pressure is positively correlated with LINZ , while the maximum temperature is negatively correlated with it, and the influence gradually decreases when LINZ exceeds 30 m, and the maximum pressure of supercritical helium in the long LINZ quench is about 15 MPa.

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Fig. 5. Current discharge curve after quench

Table 4. Minimum quench energy of cooling channels Cooling channels

MQE (mJ/cc)

t, L

C1

1420

1ms, 1cm

B1

1026

1ms, 1cm

A20

1841

1ms, 1cm

Fig. 6. Variation of pressure and temperature with time after quench

Thermal Hydraulic Analysis of Background Magnet of 15T Conductor T

95 P P

Tmax (K)

90 85

T

P

T

80 T 75

P T P T T P

70

65 PP P 60

0

T T

T

T T

20

40

12

Temperature Pressures 10 T LF P LF T MF P MF 8 T HF P HF

T

T

T

T

T

T

T

T

T

T

T

60

16 14

P

P

P

P P

P

P P

P

P P

80

100

6

Pmax (MPa)

100

1037

4 2 0

120

140

160

LINZ (m)

Fig. 7. Variation of the peak quench temperature and pressure as a function of the initial normal zone length

5 Conclusion In this paper, the excitation and quench processes of the background magnet are studied, the AC loss of the magnet during excitation is calculated, and it is used as a heat source to calculate the temperature of the cooling channel. The results show that the temperature margin of the cooling channel is greater than 1.9 K for both excitation times of 1440 s and 2800 s. The peak temperature and pressure of the background magnet were analyzed in the quench process under mechanical perturbation. The peak temperature is 100 K and the peak pressure is 12.1 bar. The effect of LINZ on the peak pressure was also investigated, and the results showed that the peak pressure reached about 15 MPa when the LINZ was greater than 30 m. Acknowledgments. The project is supported by Comprehensive Research Facility for Fusion Technology Program of China (Number 2018–000052-73–01-001228).

References 1. Hu, L., et al.: Cooling loop design of background magnet of 15T conductor performance test facility. IEEE Trans. Appl. Supercond. 31(8), 1–4 (2021) 2. Fang, J.: The Theoretical and Experimental Research on HT-7U CICC Stability [D]. Hefei Institutes of Physical Science, Chinese Academy of Sciences (2002) 3. Schild, T., Ciazynski, D.: A model for calculating a.c. losses in multistage superconducting cables. Cryogenics 36, 1039–1049 (1996) 4. ITER_D_2NBKXY v1.2

Conceptual Cooling Design for 14T MRI Superconducting Magnet System Qiangwang Hao1(B)

, Libiao Hu1,2 , Kaihong Wu1,2 , and Yu Wu1

1 Institute of Plasma Physics, Chinese Academy of Sciences, Hefei, China

[email protected] 2 University of Science and Technology of China, Hefei, China

Abstract. At present, Ultra-high field MRI system is considered the right way to explore the human brain because of the brain activity would be seen at a resolution of hundreds of micrometers. In 2017, the Chinese government launched an ambitious project to design and manufacture a 14T MRI system to support neuroscience research in future. The superconducting magnets made of Nb3 Sn superconductor and NbTi superconductor is designed to generate a homogeneous field of 14 T with a warm bore of 900 mm. In order to ensure the magnet could be operated safely and stability with a higher temperature margin, the superconducting magnet system include a main coil made of Nb3 Sn superconductor and shielding coils made of NbTi superconductor which will be immersed and cooled by the sub-cooled helium. In this paper, the concept design of the low-temperature structural and the cryogenic system will be introduced. Keywords: Superconducting magnet · Cryogenic system · MRI

1 Introduction Magnetic resonance imaging (MRI) has become the reference approach over the last three decades to investigate the human brain functional anatomy in vivo because it provides high resolution images without ionizing. As we know, the sensitivity, the spatial and temporal resolution of the images increase together with the magnetic field B0. Ultrahigh field MRI system is considered the right way to explore the human brain in the future [1]. Brain connections and brain activity would be seen at a resolution of hundreds of micrometers, ions, metabolites, neurotransmitters would be detected and measured, giving access in vivo and noninvasively to brain chemistry or genes at work in the developing brain [2]. Currently, many countries in the world are conducting ultra-high field MRI system research. Approximately 60 ultra-high field human scanners are in operation world-wide with a majority of them being 7.0T systems [3]. Several systems with 9.4T are being planned or have been installed [4]. For example, a 9.4 T MRI system with a warm aperture magnet of 650 mm located at the Center for Magnetic Resonance Research, Minneapolis, is already in operation. Meanwhile, CMRR also have become the first in the world to perform magnetic resonance imaging of the human body in February of © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1038–1043, 2023. https://doi.org/10.1007/978-981-99-6128-3_134

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2018 with the 10.5T MRI system. In France, a large aperture 11.75 T MRI system with a 900mm warm bore for human brain research was built and installed in the NEUROSPIN laboratory at CEA Saclay [5]. Based on the unique advantages of ultra-high magnetic field, the Chinese government launched an ambitious project to design and manufacture a 14T MRI system in 2017 in object to explore the human brain in future. The 14T magnet system consist of a main coil manufactured from Nb3 Sn superconductor and two shielding coils manufactured from NbTi superconductor. It will give a central field of 14 T in a horizontal warm bore (900 mm) with a 0.05-ppm-homogeneity in the central area. Figure 1 gives the magnetic field distribution of the magnet. Table 1 summarizes the main specifications for the superconducting magnet. The external dimensions of the magnet are approximately 4.4 m in diameter and 4 m in length. The stored energy of all the coil is 564 MJ and the inductance is 612 H.

Fig. 1. The magnetic field distribution of the 14T magnet system

Table 1. Specification for the superconducting magnet system Items

Value

Magnetic field

14T

Operation current

1358A

Warm bore

900mm

Homogeneity

DSV@22 mm 0.5 ppm DSV@10 mm 0.05 ppm

5G line

12m(R) x 14.9 m(Z)

Inductance

612H

Stored energy

564MJ

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2 Concept Design of the Low-Temperature Structural The magnet cryostat consists of a superconducting coil, a thermal shield, and a vacuum vessel. Figure 2 shows the concept structure design of the 14T magnet system. It shows that the length of the cryostat is about 4.8 m, and the outer diameter is about 6 m. The main coil and shielding coils are mounted on the same cold mass structure. The end plate is welded to the cold-mass structure and suspended into the cryostat by the way of low conduction rods made of glass epoxy G-10 to reduce the heat load. The coil will be cooled with pressurized sub-cooled helium at 1.2 bar and 3.8 K. Each coil has a distinct helium tank but the two volumes of the shielding coil are linked. All of the tanks are existing in one vacuum vessel. The thermal shield made of aluminum alloy cooled by LN2 is suspended on the cold mass by the way of insulating spacers. The inner thermal shield and the outer thermal shield each have its own cooling circuit. The circuits for both thermal shields have the same inner diameter and are welded all around the thermal shield. In order to reduce the radiation heat load from the vacuum vessel, 40 layers blankets of multi-layer insulation are installed on the thermal shield. Meanwhile, the helium tanks are also covered with only 20 layers to reduce the radiation heat load from the thermal shield.

Fig. 2. The concept structure design of the 14T magnet system

The heat loads of the magnet system include support conduction, thermal radiation, and heat through current lead. The heat flow through the support conduction with constant cross-sectional area, A, is given by  A T1 k(T )dT (1) Q= L T2 where L and k(T) denote the length and temperature-dependent thermal conductivity of support, respectively. The heat transfer by thermal radiation can be approximately estimated by Q=

σ (TH4 − TL4 ) 1−εH εH AH

+

1 1 AL ( εL

+

2N εN

− N)

(2)

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where A is surface area, σ is the Stefan–Boltzmann constant and E is the emissivity of the surface. N is the number of multi-layer insulation. Heat through current lead occurs by heat conduction as well as Ohmic heat generation. During initial cool-down process no current is supplied, therefore only heat conduction is considered and Eq. (1) is used for this application. Table 2 gives the expected heat loads on magnet system. Table 2. 14T MRI system heat loads(estimated). Temperature (K)

Item

Heat load(W)

Total (W)

80K

Cryostat

320

540

Transfer lines

180

Current leads

40

Magnet cryostat

45

Tranfer lines

35

Current leads

10

3.8K

90

3 Concept Design of the Cryogenic System The cryogenic system consists of a 500 W/4.5 K helium refrigerator, a helium distribution system and the magnet system. The 500 W/4.5 K helium refrigerator is connected to helium distribution system by transfer lines. The cryogenic system is showed in Fig. 3. 3.1 500 W/4.5 K Helium Refrigerator The 500 W/4.5 K helium refrigerator adopts Collins helium liquefaction cycle and possesses of four cooling stages as follows: liquid nitrogen pre-cooling, the first stage expansion (including three piston expanders), the second stage expansion (one piston expander) and J-T valve throttle cooling. The original nominal cooling capacity of this refrigerator was approximately equivalent to 500 W at 4.5 K, and it could provide 150 L/h helium liquefying capability [6]. The main parameters of the refrigerator are presented in Table 3.

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Value

Coolant

LHe&SHe

Mass flow rates

20-40 g/s

LHe evaporation rate

Liquefying capability 150 L/h

Operation temperature

4.2 K LHe/4.5 K SHe

3.2 Helium Distribution System The distribution system is an extremely important sub-assembly of the cryogenic system. The main goal of the distribution system is to keep helium delivered from the refrigerator to each component with a high degree of reliability, and to ensure the parameters such as temperature, pressure and mass flow rate of helium can meet the requirements of each device in the system. In addition, it can produce the 3.8 K helium to immerse and cool down the superconducting magnet. The distribution system is shown in Fig. 3. It consists of a LN2 tank, two helium tanks, several heat exchanges and valves. The main function of the liquid nitrogen tank is to cool the helium gas during the pre-cooling stage. The 4.5 K liquid helium from the refrigerator was reduced to 3.8 K via being sub-cooled in a 3.5 K tank. In order to keep the temperature of the sub-cooled helium not exceeding 3.5 K, a vacuum system is used to depressurize the sub-cooled tank to maintain a pressure of 47000 Pa.

Fig. 3. Concept design of the cryogenic system for the 14T superconducting system

3.3 Cooling Procedure of the Cryogenic System Before the liquid nitrogen is supplied into the liquid nitrogen tanks and used to cooldown the thermal shields of all cryostats, a vacuum pumping system is initially used to reduce the preesure in all cryostats below 10−3 Pa.

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The first stage of the superconducting magnet cool-down (from 300 K down to 80 K) will be performed by using the helium gas from the compressors which is cooled by the liquid nitrogen through the EX-1 and EX-3 mixed with the room temperature helium gas. The cold helium gas from the magnet system directly goes back to the compressors of the refrigerator. To avoid extra thermal stress damage to magnets, when the temperature of magnets is higher than 100 K, the maximal temperature difference between output and input gas should not exceed 30 K. At the same time, the thermal shield and high temperature superconducting current lead are cooled by liquid nitrogen to 80 K. The valve CV-11 would be closed at the end of the 300 K-80 K cool down, and all the systems including helium distribution system and magnet system will be cooled by the refrigerator. The vacuum system will be activated when the liquid helium in the 3.5 K LHe bath reaches a volume capacity of 50%. The magnet system will be cooled to 3.8K by the sub-cooled helium.

4 Conclusion The conceptual cooling design for the 14T MRI superconducting system is presented. The expected heat loads on magnet system are 540 W/80 K and 90 W/3.8 K. Further studies will be conducted in order to begin the detailed design of the low-temperature structure and cryogenic system.

References 1. Boublay N, Schott A M and Krolak-Salmon P.: 2016 Neuroimaging correlates of neuropsychiatric symptoms in Alzheimer’s disease: a review of 20 years of research. Eur. J.Neurol. 23 1500–9(2016) 2. Denis Le Bihan1, Thierry Schild.: Human brain MRI at 500MHz, scientific perspectives and technological challenges. Supercond. Sci. Technol. 30 (2017) 033003 3. Kraff, O., Fischer, A., Nagel, A.M., Monninghoff, C., Ladd, M.E.: MRI at 7 Tesla and above: demonstrated and potential capabilities. J Magn Reson Imaging 41, 13–33 (2015) 4. Alentin G. Kemper, Federico De Martino, Thomas C. Emmerling, Essa Yacoub, Rainer Goebel.: High resolution data analysis strategies for mesoscale human functional MRI at 7 and 9.4T. NeuroImage 164, 48–58(2018) 5. L. Quettier et al.: Commissioning Completion of the Iseult Whole Body 11.7 T MRI System. IEEE Transactions on Applied Superconductivity 30(4), 1–5(2020) 6. Jinqing Peng, Yu Wu, Huajun Liu, Yi Shi, Jinglin Chen, Zhibin Ren.: The cryogenic system for ITER CC superconducting conductor test facility. Cryogenics 51(1), 62–67(2011)

Thermodynamic Analysis of a Forced-Flow-Cooling System for a CICC Coil in a High Field Superconducting Magnet Jingxin Zheng1,2 , Junjie Li1,2(B) , and Zhengrong Ouyang1,2 1 High Magnetic Field Laboratory, Hefei Institutes of Physical Science, Chinese Academy of

Sciences, Hefei 230031, China [email protected], [email protected] 2 University of Science and Technology of China, Hefei 230026, China

Abstract. A high field superconducting magnet with a 100 mm room-temperature bore is being designed and installed in the High Magnetic Field Laboratory in China. To provide an optional cryogenic supply plan for a compact CICC coil in this superconducting magnet, a forced-flow-cooling system based on the GM/JT cycle is designed. In this paper, the design of the cryogenic system is briefly described. Meanwhile, the influence of stage temperature of GM cryocooler and helium gas inlet pressure on the performance of the system is thermodynamically analyzed. The results of the analysis make it clear that, the temperature of GM cryocooler precooling stage plays a very important role on the system performance. And different operational helium gas inlet pressure and its thermal characteristics in the cooling system are also discussed. Based on these studies, the forced-flowcooling system for the CICC coil can be optimized and we can get a good reference for choosing GM cryocoolers. Keywords: Thermodynamic analysis · Forced-flow-cooling system · CICC · GM/JT

1 Introduction In recent years, with the development of compact high-field superconducting magnet coils made of cable-in-conduit conductors (CICCs), higher requirements for the stability and compactness of supercritical helium forced-flow-cooling systems have been put forward. A high field superconducting magnet with a 100 mm room-temperature bore is being designed and installed in the High Magnetic Field Laboratory in China. A 10 T superconducting magnet with the same warm bore size was fabricated and cooled by a GM cryocooler in Institute of Electrical Engineering, Chinese Academy of Sciences [1]. However, GM cryocooler can’t satisfy the demand for forced-flow-cooling. So, a forced-flow-cooling system based on the GM/JT cycle which is a type of hybrid JT helium refrigeration cycle, is designed to provide supercritical helium at 4.5K. The hybrid JT helium refrigeration cycle is normally used to liquefy helium gas and to cool some electronic detectors in space because it can make operation economical [2]. There © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1044–1050, 2023. https://doi.org/10.1007/978-981-99-6128-3_135

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are already some studies on Helium liquefier based on GM/J-T concept, such as the Alternating Gradient Synchrotron (AGS) of American Brookhaven National Laboratory ([email protected]) [3], Japanese Refrigerator for Maglev Vehicle ([email protected]) [4] and Harbin Institute of Technology ([email protected]). In the area of space astronomy, liquid helium temperature cryocooler for space application has become a key field for cryogenic research at National Aeronautics and Space Administration (NASA), Japan Aerospace Exploration Agency (JAXA) and European Space Agency (ESA) [5]. For instance, mechanical cryocooler used in Planck is a 4He JT cooler precooled by adsorption cryocooler [6]. JWST uses a three-stage pulse tube cryocooler to precool JT loop [7]. But few researchers have applied hybrid JT helium refrigeration cycle to forced flow cooling for CICC superconducting coils. Therefore, it is significant and feasible to do the research on the GM/JT cycle for the CICC coil forced-flow-cooling system. The maximum heat load of the CICC coil is about 3.5 W under the most severe condition where the ac-loss of 2.53 W was generated continuously. So, a GM cryocooler with a cooling capacity of [email protected] is used to design the cooling system based on GM/JT refrigeration cycle to absorb heat load and ensure that the outlet temperature of the coil is below 6K. In this paper, the process design of the forced flow cooling system for CICC coil in a high field superconducting magnet is briefly described. Meanwhile, the cooling system is analyzed by thermodynamic method, and the parameters like gas inlet pressure and stage temperatures of GM cryocooler are investigated systematically. The calculation results and the effect of precooling temperature and inlet pressure on the cooling system performance is systematically presented. Based on these studies, the forced-flow-cooling system for the CICC coil can be further optimized.

2 The Process Design of the Forced-Flow-Cooling System

Table 1. Specifications of the CICC coil. Parameter

Coil A (CICC)

Conductor

Nb3Sn

Inner diameter (mm)

120

Outer diameter (mm)

312

Coil height (mm)

416

Length (m)

141

Magnetic field contribution (T)

3T

The specifications of the CICC coil in the high field superconducting magnet are shown in Table 1. The static heat load of the system is currently estimated to be 0.97 W. Considering the ac-loss, the heat load of the CICC coil Qcoil will reach 3.5 W. And the outlet temperature of the CICC coil must be below 6K. The GM cryocooler with a cooling capacity of [email protected] are used to design a cooling system based on GM/JT

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refrigeration cycle to absorb heat load. The schematic view of the cooling system is shown in Fig. 1. The cooling system is mainly comprised of a room temperature compressor, six heat exchangers, a two stage GM cryocooler and JT valve. The circulating helium gas is compressed by a room-temperature compressor and cooled successively by a 1st-stage heat exchanger, a 1st-stage precooling heat exchanger, a 2nd-stage heat exchanger, a 2nd-stage precooling heat exchanger, and a 3rd-stage heat exchanger. Subsequently, the helium gas is cooled to 4.5K by the 4th-heat exchanger in the helium tank and flows through the coil to ensure the uniformity of the temperature inside the coil. Finally, liquid helium is generated after the flow from the superconducting joint of the coil through the JT valve. The evaporated helium gas in the helium tank flows through the low-pressure side of the heat exchanger and returns to the compressor. The components except the first stage heat exchanger are housed in a radiation shield which is cooled by the first cold stage of the GM cryocooler. The Temperature-Entropy (T-S) diagram of the cooling system is shown as Fig. 2. In the cooling process, without considering the heat leakage of the system, the work and heat exchange with the outside world mainly include the compression work of the compressor, the precooling capacity provided by the cold stage of GM cryocooler, the heat generated by the CICC coil and the cooling capacity stored in the helium tank. Therefore, from the view of power and heat exchange, it is the most direct way to improve the system performance to reduce the power consumption of compressor and the cold stage pre-cooling capacity and to increase the liquefaction rate of helium. In this paper, the improvement of system performance is studied only from the way of reducing precooling capacity and increasing cooling capacity in the helium tank. The consumption of precooling capacity is mainly related to the precooling temperature when the system flow is constant. It can be seen from Fig. 2 that the liquefaction rate of helium after throttling is closely related to the inlet pressure. In conclusion, precooling temperature and gas pressure are the key factors affecting the performance of forced flow cooling system based on GM/JT cycle.

3 Thermodynamic Calculation and Analysis Before analyzing the thermophysical characteristics of the cycle, some assumptions must be made: (1) The present thermodynamic analysis is based on the steady state conditions and other pressure losses except the inherent pressure drop of the superconducting coil are not considered; (2) The outlet temperature of the high-pressure side of the compressor is 300K, the pressure is 10bar, and the pressure after throttling is 1.3bar; (3) The temperature after helium gas exits the precooling heat exchangers are consistent with that of the cold stages of the cryocooler; The outlet temperature of the helium cooled by the 4th heat exchanger in the helium tank is the same as the temperature of the saturated helium at low pressure; (4) The efficiencies of all heat exchangers are constant, and does not change with the change of working medium pressure, temperature and mass flow and they are set to initial value 0.97;

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Fig.1. Schematic view of the forced-flow-cooling Fig.2. Temperature-entropy diagram of system the forced-flow-cooling system

(5) The JT expansion is a perfect isenthalpic process and other heat in-leak in the system is negligible. The thermophysical properties of the helium gas is taken from the NIST. From the perspective of thermodynamics, the steady-state conditions of each component of helium cycle system are analyzed by the first law of thermodynamics. The specific enthalpy of unsaturated helium gas in heat exchanger can be determined by temperature and pressure and the theoretical maximum heat load of heat exchanger can be calculated by formula (1) (2) and (3) [2]: ˙ hp,max = m ˙ ∗ (h(Thp,in , php,in ) − h(Tlp,in , php,out )) Q

(1)

˙ lp,max = m∗(h(T ˙ Q hp,in , plp,out ) − h(Tlp,in , plp,in ))

(2)

˙ hp,max , Q ˙ lp,max ) ˙ max = min(Q Q

(3)

     ˙ actual = m∗ ˙ h Thp,in , php,in − h Thp,out , php,out or Q      ˙ actual˙ = m∗ Q h Tlp,out , plp,out − h Tlp,in , plp,in

(4)

In the formula, the lower corner marks hp and lp represent the high-pressure and low-pressure sides of the heat exchanger; Lower corner mark in and out represent the ˙ renders the heat load of the heat exchanger, W. import and export of heat exchanger; Q The efficiency of heat exchanger is defined as the ratio of the actual heat load to the theoretical maximum heat load, i.e. ηHX =

˙ actual Q ˙ max Q

(5)

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The energy conservation of helium gas in the precooling heat exchanger, helium tank and CICC coil can be calculated by Eq. (5) and (6):   ˙ stage (6) m ˙ ∗ hPreHX ,in − hPreHX ,out = Q ˙ tank = m ˙ ∗ (htank,out −htank,in ) Q

(7)

˙ Coil = m ˙ ∗ (hcoli,out −hcoil,in ) Q

(8)

Therefore, the helium states at the high-pressure and low-pressure sides of the heat exchangers can be correlated through (1) (2) (3) (4) and (5). Equation (6) relates the cooling capacity required in the pre-cooling process to the state of helium at the inlet and outlet of the pre-cooling heat exchanger. Equation (7) relates the cooling capacity left in the helium tank to the state of helium entering and leaving the helium tank and Eq. (8) relates the heat load of the CICC coil to the state of helium entering and leaving the coil. The states of each point in the cycle can be calculated by combining the hypothetical conditions and the analysis of the circulating components. 3.1 Calculation Results The numerical calculation results are shown in Fig. 3 and 4. When the heat load in the CICC coil is constant, the heat transfers at the 1st cooling stage, Q1st , and the 2nd ˙ When the mass flow rate cooling stage, Q1st , increased with the helium gas flow rate m. m ˙ is greater than 0.6 g/s, the outlet temperature of the coil, Tcoil,out , is 5.99K and the cooling capacity left in the helium tank, Qtank , is greater than [email protected]. Thus, the gas temperature Tcoil,out can be up to 6 K at the helium gas flow rate of 0.6 g/s and the heat transfers at the 1st cooling stage, Q1st , and the 2nd cooling stage, Q2nd respectively are 19.6 W and 14.03 W.

Fig. 3. Heat transfers at the two cooling stages Fig. 4. Tcoil,out and Qtank as a function of and coil as a function of the He gas flow rate the He gas flow rate

3.2 Effect of Precooling Temperature Under the condition that the second-stage cooling temperature is kept constant, the change of each stage precooling power with the outlet temperature of the first-stage

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precooling heat exchanger, T1st,out is shown in Fig. 5. It can be seen that the heat transfers at the 1st cooling stage, Q1st , decreases linearly with the increasing of the precooling temperature, while the 2nd cooling stage, Q2nd , increases linearly. Therefore, although the increase of the first stage precooling temperature can reduce the precooling capacity required by the first stage precooling, but need to pay more lower temperature, higher quality precooling capacity in the second stage, which is unfavorable to improve the efficiency of the system. Keep the first-stage precooling temperature unchanged, and the changes of the first-stage precooling capacity and cooling capacity with the increase of the second-stage precooling temperature are shown in Fig. 6. The precooling capacity of the first stage is slightly increased, but the precooling capacity of the second stage is greatly reduced. The Qtank decreases at the same time. When the second-stage precooling temperature increases to 14.5K, the cooling capacity in the helium tank is only enough to pre-cool the high-pressure helium to 4.5K, and there is no more surplus cooling capacity.

Fig.5. Effect of T1st,out on heat transfers at the Fig. 6. Effect of T2nd,out on heat transfers two cooling stages at the two cooling stages and helium tank

In general, appropriately increasing the second stage precooling temperature can reduce the requirements for the GM cryocooler, which is beneficial to improving the efficiency of the whole system. 3.3 Effect of the Inlet Pressure According to T-S figure of the cooling system (Fig. 2), the helium liquefaction rate after throttling is closely related to the gas inlet pressure and the helium liquefaction rate after throttling will directly determine the cooling capacity in the helium tank. The dotted line in Fig. 2 shows the change of cooling capacity by changing the inlet pressure. It can be seen that the gas inlet pressure has a decisive influence on the cooling capacity. When the inlet pressure is increased, the changes of the heat transfers at the 1st cooling stage, the 2nd cooling stage, and the cooling capacity left in the helium tank is shown as Fig. 7. With the increase of the gas inlet pressure, the Qtank increases. The Q1st has a small rise, but the Q2nd increases significantly. Therefore, although increasing the gas inlet pressure is beneficial to the improvement of the cooling capacity left in the helium tank, the higher the inlet pressure is not the better. The pre-cooling capacity of the secondary cold stage is the main limiting factor.

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Fig. 7. Effect of the inlet pressure on heat transfers at the two cooling stages and helium tank

4 Conclusion The process design of the forced flow cooling system for CICC coil based on the GM/JT cycle in a high field superconducting magnet is briefly described in this paper and the thermodynamic analysis is presented. When the mass flow rate m ˙ is 0.6g/s, the cooling system can take away 3.5W heat load of CICC coil and outlet temperature of the coil is 5.99K and the cooling capacity left in the helium tank, Qtank , is [email protected]. The analysis results show that the temperatures of stages of GM cryocooler are essential parameters for precooling capacity input. Particularly, Increasing the second stage precooling temperature can reduce the demand for precooling capacity, but the cooling capacity left in the helium tank will decrease. Based on these studies, the forced-flowcooling system for the CICC coil can be optimized and we can get a good reference for choosing GM cryocoolers.

References 1. Dai, Y., et al.: Fabrication of A 10 tesla cryogen-free superconducting magnet. IEEE Trans. Appl. Supercond. 21(3), 1608–1611 (2011). https://doi.org/10.1109/TASC.2010.2096550 2. Liu, Z.M., et al.: Thermodynamic analysis of G-M/J-T Helium refrigeration cycle for a cryoadsorption test facility. J. Fusion Energy 34(5), 1188–1192 (2015) 3. Jia, L.X., et al.: A five-watt G-M/J-T refrigerator for LHe target at BNL. In: Proc. AIP Conf., 1902, pp. 776–781. https://doi.org/10.1063/1.1472094 4. Fujimoto, S., et al.: Development of a 4K GM/JT refrigerator for maglev vehicle. In: Proc. 16th Int. Cryogenic Eng. Conf./Int. Cryogenic Mater. Conf., pp. 331–334 (1996) 5. Ma, Y., et al.: Development of a space 100mW@5K closed loop JT Cooler. Cryogenics 104(710), 102983 (2019) 6. Morgante, G., et al.: Cryogenic characterization of the Planck sorption cooler system flight model. J. Instrumentation 4.12 (2010) 7. Ross, R.G.: A study of the use of 6K ACTDP cryocoolers for the MIRI instrument on JWST. Springer, US (2005)

Results and Analysis of the First Year of Operation of the UKRI STFC Daresbury Vertical Test Facility Andrew J. May1,2(B) , Shrikant Pattalwar1,2 , David Mason3 , Keith Middleman1,2 , Mark D. Pendleton3 , Paul A. Smith1,2 , Stuart Wilde1,2 , Ayomikun Akintola1,2 , Alexander R. Bainbridge1,2 , Rachael Buckley1,2 , Gary Collier1,2 , Peter Corlett1,2 , Keith Dumbell1,2 , Michael Ellis1,2 , Mark Hancock3 , Jane Hathaway3 , Sean Hitchen1,2 , Carl Hodgkinson3 , Philip Hornickel1,2 , Gary Hughes1,2 , Conor Jenkins1,2 , Geraint Jones3 , Michael Lowe3 , Peter McIntosh1,2 , George Miller3 , Jennifer Mutch3 , Andrew Moss1,2 , Adrian Oates3 , Alan E. Wheelhouse1,2 , Alastair A. J. White1,2 , and James Wilson3 1

Accelerator Science and Technology Centre, STFC Daresbury Laboratory, Keckwick Lane, Warrington WA4 4AD, UK [email protected] 2 Cockcroft Institute, Keckwick Lane, Warrington WA4 4AD, UK 3 Technology Department, STFC Daresbury Laboratory, Keckwick Lane, Warrington WA4 4AD, UK

Abstract. The first year of operation has been completed of the UKRISTFC Daresbury Laboratory’s novel vertical test facility. In each run, 3 jacketed superconducting RF cavities are tested in the VTF in a horizontal configuration at 2 K. A 24-month program to test 84+4 704 MHz high-β cavities for the European Spallation Source as part of the UK’s IKC is currently ongoing. Here, we report the results and analysis from the first full year of facility operation.

Keywords: Vertical test facility

1

· superconducting cavities

Introduction

The UKRI-STFC Daresbury Laboratory’s Vertical Test Facility (VTF) [1] has now been operating regularly for over a year. The VTF supports 2 K characterisation of three jacketed SRF cavities in a single cool-down run, with the 2-year ESS high-beta cavity testing program well underway. Higher order modes and passband modes are measured at low power. Q vs E field measurements are made at higher accelerating gradients (up to 200 W input power). A sizable saving in liquid helium (LHe) use is given by the novel cryogenic architecture of the facility. c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1051–1057, 2023. https://doi.org/10.1007/978-981-99-6128-3_136

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VTF Cryostat Design

The standard approach for vertical testing of SRF cavities is full submersion in a large LHe bath, with cooling to ≤2 K using a vacuum pump/cold compressor to reduce the partial pressure above the bath. SRF characterisation is then undertaken. This method has been used successfully for many programs, including testing XFEL cavities at DESY [2,3]. Whilst well-established, this technique requires both a large LHe cryoplant and, for the activities at UKRI STFC, would require ∼8500 L of LHe per 3-cavity testing cycle. Keeping in view the ever-diminishing global supply of He, along with the related increases in cost, an alternative approach has been developed for SRF testing of jacketed cavities which requires much less LHe and a significantly smaller liquefier throughput [1,4–6]. The cryostat architecture is based on a cavity support insert (CSI) whereby 3 jacketed cavities are mounted horizontally below a single header tank, with each being fed by a common fill/pumping line; this is shown in Figs. 1 and 2 which present a photograph and a CAD model respectively of an assembled CSI. In using this architecture, much less LHe is required per testing run (∼1500 L, all of which is recovered by the closed-circuit, as described in Sect. 4) compared with the conventional approaches.

Fig. 1. Photograph of CSI on stand with three jacketed cavities installed

Fig. 2. CAD model of CSI with 3 jacketed cavities installed; also shown are the positions of the active coils

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The CSI is mounted into a vessel which comprises the thermal radiation shields, magnetic shielding, and OVC (outer vacuum chamber). The manufacturuer of the cryostat was Criotec Impianti S.p.A.

3

Safety

Safety is paramount in the operation of any cryogenic facility. As such, significant efforts have been devoted to understanding the potential failure modes for the facility, and introducing strategies for mitigation in consideration of the appropriate regulations. For a detailed report on the safety of this system, the interested reader may refer to Ref. [5].

4

Cryogenic Infrastructure

The liquefier - an Air Liquide H´elial ML - was commissioned in 2018. The cryoplant supplies 50 K GHe (produced by the first heat exchanger stage of the liquefier) and 4.2 K LHe. LN2 precooling of the cold box first stage HEX supports a cold GHe supply rate of ∼2 g/s and a liquefaction rate of ∼130 L/hour. A storage dewar with a capcity of 3000 L is utilised for storage of LHe. At present, the cryoplanr is operated with a total He inventory of 2700 L LHe-equivalent. Liquid and gaseous cryogen supply to the CSI and shield cooling circuits is managed by a 2 K valve box located immediately next to the VTF cryostat. Helium boil-off from the CSI is pumped on by a single pumping set which is comprised of a roots blower booster pump (Leybold RA7001 SO) backed by a rotary vane main pump (SV1200). The returning helium from both the LHe and GHe cooling circuits is recovered into a single common gas bag, before being compressed through into high pressure gas storage, then purified by a purifier within the ALAT cold box, and finally reliquefied; in this way, a completely closed-loop system is implemented.

5

Cryogenic Performance

Following the facility commissioning [1], the first full year of regular operations has been carried out. As of April 2022, 31 commissioning and cavity test runs have been carried out, providing 64 cavity testing slots for ESS high-beta cavities. The following section reports some of the observations, lessons learnt, and improvements from that first year. 5.1

Testing Cycles

Two identical CSIs are operated to support high testing throughput; one can be prepared whilst the other is under test and then the two swapped to begin Run n+1 immediately following the completion of Run n. Test cycles are now reliably carried out on a two-week schedule; runs are divided into the following modes, each with clearly defined procedures and quality control checks:

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– Mode-1 Assembly of cavities on CSI – Mode-2 Loading of CSI into bunker cryostat – Mode-3 Cooldown of shield and cavity to ∼50 K using cold GHe; 48 h typically required for cooldown – Mode-4 Cooldown of cavities to 4.2 K using LHe; 6 h typically required for cooldown and fill – Mode-5 4.2 K RF operations (see Sect. 6) – Mode-6 Cooldown of cavities to 2 K utilising sub-atmospheric pumps; 5 h is typically required for fill and cooldown – Mode-7 2 K RF operations (see Sects. 5.2 and 6) – Mode-8 Warmup to 300 K; 72 h typically required (heaters in combination with recirculating pumps are used to increase rate of warm-up) – Mode-9 Removal of CSI from bunker – Mode-10 Disassembly of cavities on CSI stand (may then return to Mode-1) It may be considered that is is possible to have overlap between Modes 9–1 and Modes 2–8 on the alternate CSI, improving the throughout of the facility. 5.2

Pressure and Temperature Stability at 2 K

The 2 K pumping system provides >1 g/s at 30 mbar (equivalently 2 K). A bypass valve in parallel with the pump set (as shown in Fig. 3) is operated upon by a PID control loop. This has been shown to give pressure stability (under static loading) at the level of ±0.1 mbar in the CSI (corresponding to a temperature stability of ±1 mK of the LHe). Planning is currently underway for installation of an additional pump stack to provide additional throughput capacity, as well as increase the level of redundancy in the VTF. The reliably demonstrated cryogenic performance is sufficient to allow ∼40 s of RF power dissipation up to 200 W during high-gradient testing.

Fig. 3. Bypass valve (CV2213) shown operating on PID control in parallel with 2 K main pump (P6320A) and booster pump (P6310A) (NB: direction of flow is left to right through pumps and right to left through bypass valve)

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Shield Cooling

The thermal shields are cooled by ∼50 K GHe from the coldbox, which is fed from the clean gas buffer. When the liquefier is running, there is an additional demand on the clean gas buffer which the purifier throughput cannot match and so it is not possible to run both the shield cooling with gas and liquefier simultaneously for periods in excess of a few hours. As such, GHe is is used to cool the shields from room temperature down to ∼100 K during the first 2 days of the run (Mode-3). Shield cooling is then switched via the cryogenic valve box to be provided by the boil-off of LHe from the dewar. The latter is run alternately with daily LHe top ups carried out in Modes 4–7 and cools the shield 24 h at 70 °C to achieve a consistent ‘dry’ condition. Tensile testing was carried out in general accordance with ASTM D638. The testing speed was 5 mm/min, and specimens were stabilized at either 77K or 292K prior to testing. 77K tests were conducted with specimens immersed in liquid nitrogen at atmospheric pressure.

ASTM

DMA

Fig. 1. Test coupons: (a) dimensions in mm and (b) build orientation designation

In this series of tests we also decided to investigate the influence of build orientation on the mechanical performance. In practice, most parts will have a range of orientations of load relative to build plane. Test coupons for the materials were therefore produced in a variety of orientations which we designated as shown in Fig. 1(b).

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3 Test Results 3.1 DMA Testing The PA12-40GF material was tested in four build orientations using the DMA to obtain values of Storage Modulus over a full temperature sweep. Samples were evaluated at room temperature (292K) initially to determine sample variability. Figure 2(a) shows that the Z, Y and 45 build directions provide similar stiffness, however the X build direction is approximately 15% stiffer. The modulus at 84K (Fig. 2(b)) shows similar trends, except for the 45° sample which has lower relative stiffness compared to 292K.

Fig. 2. (a) Storage modulus of PA12-GF40 at room temperature and (b) at 84K (1-σ error bars)

Figure 3 shows typical results for modulus across the full temperature sweep for the four build directions with the characteristic shape being the same for all orientations. Figure 3 also shows the DMA modulus data from all 4 materials in the ‘X’ orientation. All materials show a clear transition around 130K as observed in other polymers [12], except the PA6-MF which shows an unexpected, almost linear response to temperature. 3.2 Tensile Testing As with the DMA testing, a number of PA12-GF40 samples were initially tested at room temperature to quantify variability, and results are shown in Fig. 4. In this case there is no statistically significant difference between the three build orientations tested. Figure 5 shows typical tensile test results comparing the filled and unfilled PA12, with a direct comparison of the room temperature result against the 77K result.

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Fig. 3. (a) Effect of build orientation on storage modulus and (b) ‘X’ orientation storage modulus comparison for the four materials tested

Fig. 4. (a) Yield stress and (b) Tensile modulus for PA12-GF40 at 292K (1-σ error bars shown), 3 build orientations

In practice, it was difficult to obtain a good result from the PA6-MF under cryogenic conditions. The material tended to shatter or fail in the grips, possibly indicating a strong notch sensitivity. However, the tensile modulus values did reach a linear zone and were reliable, so the values have been included later in the discussion, for interest.

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Fig. 5. Tensile test comparisons at 77K and 292K for (a) filled and (b) unfilled PA12

4 Discussion Figure 6 compares the DMA generated storage modulus and tensile modulus for three materials (all printed in the X-orientation) at 292K and approximately 80K respectively (noting that tensile tests were carried out at 77K compared to values taken at 84K for the DMA). Apart from one unexpected result for PA12-GF40, the results trend relatively consistently. The DMA modulus is around 75% of the tensile test results, and modulus values measured by both methods at 80K are 80% higher than those at room temperature, on average. It is also interesting to compare the properties measured against typical values given for injection moulding (IM) plastics. An unfilled PA12 for IM might be expected to have a modulus in the range of 1.5 to 1.8 GPa, an elongation of more than 10% and a yield strength of around 50 MPa. The AM version we tested showed similar properties at room

Fig. 6. Modulus comparison from DMA and tensile tests, three materials at two temperatures

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temperatures and under cryogenic conditions the stiffness and strength are significantly increased, with the trade-off being substantially lower elongation. In comparison, the glass filled PA12 also shows strength and stiffness improvements with a much smaller depreciation in ductility at cryogenic conditions. However, while the proportional improvement in strength is much greater, the improvement in stiffness is much smaller than for the unfilled material. The same effect is seen for the filled PA6, where the increase in stiffness is proportionally much lower than for the unfilled polymers. We must also note that compared to a typical glass filled PA12 IM material, the strength and stiffness at room temperature of the PA12-GF40 AM material is considerably lower. The heightened notch sensitivity of the PA6-MF at low temperature and the unusual nature of its overall temperature response (Fig. 3b) indicates that this composite may not be suitable for cryogenic applications.

5 Conclusions and Further Work As mentioned, this data represents the first steps in development of a library of material properties for cryogenic AM manufacturing. The AM polymers have different characteristics to typical IM material datasheets at room temperature and the performance under cryogenic conditions varies significantly from material to material. We have seen good consistency between coupons taken from the same build in the same SLS machine, but we will specify that test coupons be printed with any major engineering components for QA testing. Our next round of testing is focused on thermal expansion coefficient properties of these materials, in addition to other metal powder filled nylons that have become available. We expect to be reporting on these in the near future. Acknowledgments. The project is supported by New Zealand’s Ministry of Business, Innovation and Employment (Grant RTVU2004).

References 1. Badcock, R., et al.: High power density electric motors for large-scale transport. CEC/ICMC J3Or1A-03 [Invited], July 2021 2. Haran, K.S., et al.: High power density superconducting rotating machines—development status and technology roadmap. Supercond. Sci. Technol. 30(12), 123002 (2017) 3. Jeong, S., et al.: Holistic approach for cryogenic cooling system design of 3 MW electrical aircraft motors. In: Electric Aircraft Technologies Symposium, EATS 2021 (2021) 4. Chen, L., He, Y., Yang, Y., Niu, S., Ren, H.: The research status and development trend of additive manufacturing technology. Int. J. Adv. Manuf. Technol. 89(9–12), 3651–3660 (2017) 5. Daminabo, S.C., et al.: Fused deposition modeling-based additive manufacturing (3D printing): techniques for polymer material systems. Mat. Today Chem. 16, 100248 (2020) 6. Turner, B.N., Strong, R., Gold, S.A.: A review of melt extrusion additive manufacturing processes. Int. Process Des. Model. Rapid Prototyp. J. 20(3) (2014) 7. Gibson, I., Shi, D.: Material properties and fabrication parameters in selective laser sintering process. Rapid Prototyp. J. (1997)

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8. Kruth, J.-P., et al.: Benchmarking of different SLS/SLM processes as rapid manufacturing techniques. In: Proceedings of International Conference Polymers and Moulds Innovations. PMI 2005 (2005) 9. Gibson, I., Rosen, D., Stucker, B.: Direct digital manufacturing. In: Gibson, I., Rosen, D., Stucker, B. (eds.) Additive manufacturing technologies, pp. 375–397. Springer, Boston (2015). https://doi.org/10.1007/978-1-4419-1120-9_14 10. Weiss, K., Bagrets, N., Lange, C., Goldacker, W., Wohlgemuth, J.: Thermal and mechanical properties of selected 3D printed thermoplastics in the cryogenic. IOP Conf. Ser. Mater. Sci. Eng. 102 (2015) 11. Cruz, P., et al.: Tensile strengths of polyamide based 3D printed polymers in liquid nitrogen. IOP Conf. Ser. Mater. Sci. Eng. 102(1), 012020 (2015) 12. Clark, A.F., Reed, R.P., Hartwig, G. (eds.) Nonmetallic Materials and Composites at Low Temperatures. Springer, Boston (1979)

Improved Thermal Conductivity at Low Temperatures in Epoxy Nanocomposites by Hexagonal Boron Nitride Aerogels Zhicong Miao1,2 , Yalin Zhao1,2 , Zhengrong Zhou1,2 , Haojian su1,2 , Mingyue Jiang1,2 , Wanyin Zhao1 , Rongjin Huang1,2(B) , and Laifeng Li1 1 State Key Laboratory of Technologies in Space Cryogenic Propellants, Technical Institute of

Physics and Chemistry, Chinese Academy of Sciences, Beijing 100190, China [email protected] 2 Center of Materials Science and Optoelectronics Engineering, University of Chinese Academy of Sciences, Beijing 100049, China

Abstract. The development of large-scale superconducting magnets has motivated the current importance of manufacturing heat-dissipation polymers with high thermal conductivity at low temperatures. However, the extremely low thermal conductivity of epoxy resins from low temperature (0.03 W·m−1 ·K−1 at 77 K) to room temperature (0.21 W·m− ·K−1 at 300 K) limits its application in superconducting fields. Here, we prepared hexagonal boron nitride nanoribbon (BNNRs) aerogels with high crystallinity by the precursor pyrolysis method, and then the epoxy was immersed into the BNNRs aerogels to obtain the EP/BNNRs aerogel composites. The microstructure of BNNRs aerogel were investigated by transmission electron microscope (TEM) and scanning electron microscope (SEM). The crystallinity and molecule structure were characterized by X-ray diffraction (XRD) and Fourier transform infrared (FTIR). Under a low BNNRs aerogels filling content of 3.2 wt%, the thermal conductivity of the composites increased from 0.03 to 0.1 W·m−1 ·k−1 at 77 K, 0.21 to 0.38 W·m−1 ·k−1 at 300 K. In this work, we think the formation of thermal conduction path in BNNRs aerogels contributes to the improvement of thermal conductivity. Therefore, the as-designed material points the way to epoxy-based, thermally conductive for the applications of superconducting devices. Keywords: Epoxy · Superconducting · Boron nitride · Nanoribbons · Aerogel · Thermal conductivity

1 Introduction Epoxy resins is widely used in energy, aerospace and other fields due to its excellent insulating properties, high mechanical strength and stable chemical properties [1]. However, the extremely low thermal conductivity of epoxy limits their wide application [2, 3]. Therefore, improving the thermal conductivity of epoxy resins by inorganic fillers is a necessary way [4]. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1089–1094, 2023. https://doi.org/10.1007/978-981-99-6128-3_141

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In superconducting equipment, epoxy resins are an essential component which provides insulation protection and mechanical support for superconducting magnets [5, 6]. However, when the superconducting magnet works, a large thermal stress will be generated inside the epoxy resin structure despite in the extremely low temperature working condition. If the thermal stress accumulates for a long time, it will lead to the destruction of support structure, result in the quench of the superconducting magnet [7]. Hence, as the support structure, the epoxy resins matrix composite needs to have high thermal conductivity to prevent irreversible damage to the equipment caused by excessive thermal stress. Blending with high heat-dissipation inorganic fillers is an effective way to improve the thermal conductivity of epoxy resins, such as alumina, silicon carbide and boron nitride [8, 9]. But these fillers often require high loading fractions, result in other changes in properties of epoxy resins. Thus, an appropriate filler is required to balance the properties of epoxy resins. Recently, most of modified fillers used in the field of superconductivity are ceramic materials [10]. Three-dimensionally filled thermally conductive fillers have been studied a lot in recent years [11, 12]. It is an efficient way to improve thermal conductivity by constructing a three-dimensional network by using a ceramic filler with a lower dosage. In this work, we prepared a boron nitride nanoribbons (BNNRs) ceramic aerogel to construct a three-dimensional thermally conductive network. By vacuum impregnating epoxy resins (EP) into the aerogel structure, the BNNRs/EP nanocomposites were obtained. First, the macroscopic and microstructures of BNNRs aerogels were investigated, which can help us better understand the mechanism of the improved thermal conductivity of the BNNRs/EP nanocomposites. Then, by testing the thermal conductivity of the composite material from room temperature (300 K) to low temperature (77 K) under different loading fractions, we found that the thermal conductivity of the nanocomposite was significantly improved when the content of BNNRs aerogels increased. The strategy to construct the three-dimensional thermal conduct path way by BNNRs aerogels shows great potential in designing high thermal conductivity materials which can be used in superconducting devices.

2 Experiment Section 2.1 Materials Melamine, Boric acid and tert-Butanol were obtained from Macklin, Bisphenol F type epoxy resin monomer (DGEBF) and Diethyltoluenediamine (DETD) was purchased from Huntsman company (USA). In this work, all chemicals were obtained from commercial sources. 2.2 Synthesis of BNNRs Aerogel In brief, a mixture of melamine and boric acid with a molar ratio of 1:6 was added to the deionized water. The mixtures were stirred magnetically for 0.5 h at 80 °C to obtain a clear solution. Then the mixture was transferred to the molds of different shapes and

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then slowly cooled to room temperature until the white gel melamine diborate (M·2B) precursor was obtained. The gel was frozen-drying 72 h by lyophilizer to get a precursor gel. The gel was placed in a horizontal quartz tube and pyrolysis at 1300 °C under a flow of Argon. Finally, the BNNRs aerogel of different shapes was obtained. 2.3 Synthesis of BNNRs/EP Composites First, DETD and DGEBF at a weight ratio of 24/100 were mixed, then the mixture was degassed with stirring at 50 °C. After degassing is complete, the aerogels were fully immersed in epoxy and stored for 2 h in a vacuum oven at 50 °C. Then, the composites were cured according to a process of 80 °C for 12 h and 150 °C for 3 h. The loading fraction in BNNRs/EP nanocomposites is mainly determined by the concentration of the precursors. 2.4 Characterizations Fourier transform infrared spectroscopy (FT-IR; Excalibur HE3100) and X-ray diffraction (XRD; D8 focus) were used to analyze the structure of BNNRs. The morphologies of BNNRs and BNNRs/EP nanocomposites were observed with a field emission scanning electron microscope (SEM, Hitachi S-4800). Transmission electron microscope (TEM, JEM-2100F) was used to observe the crystalline structure of BNNRs. The thermal conductivity of nanocomposites was measured by the steady-state method, the calculation formula is K =Q

d T × S

where Q is the power supplied, S is the cross sectional of the sample, d is the thickness of the sample, and T is the measured temperature difference.

3 Result and Discussion 3.1 Characterization of BNNRs Aerogels Figure 1 shows the macrostructure of the aerogel. We can see that the aerogels have an ultra-light structure and a complete appearance, and can be prepared in different molds to obtain different shapes, indicating the good molding performance of BNNRs. Figure 2 is the TEM images of BNNRs. It can be seen that the size of BNNRs is between 200 nm and 1 um, and they are entangled with each other. The surface of BNNRs is composed of many tiny particles. Through high-magnitude image of BNNRs, the regular crystal plane distribution can be observed, the crystal plane spacing is about 0.332 nm, which is highly consistent with hexagonal boron nitride (h-BN). Figure 3a present the XRD curve of BNNRs. It can be seen that the feature diffraction peaks at 26°,43°, 57° and 77° related to the (002), (100), (101), (004) and (110) planes which are consistent with h-BN. Moreover, no spurious peaks were observed, it means the BNNRs have high purity. Figure 3b shows the FTIR of BNNRs. In Fig. 3b, the

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feature peak of the stretch and flexural vibration of B-N bond in 809 and 1381 cm−1 can be seen. In addition, the peak at 3428 cm−1 indicated that some hydroxyl and amino groups are remained during the pyrolysis procedure. Figure 4 shows the SEM images of the aerogels. It can be seen from the figure that BNNRs of different sizes are connected together by linking sites, forming a large network structure, which is the basis construction for the three-dimensional network of aerogels.

Fig. 1. Digital photographs of BNNRs aerogels.

Fig. 2. TEM images of BNNRs

Fig. 3. a) XRD curve of BNNRs, b) FTIR curve of BNNRs

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Fig. 4. SEM images of BNNRs aerogels

3.2 Thermal Conductivity of BNNRs/EP Composites Figure 5 displays the low temperature to room temperature thermal conductivity test of epoxy and BNNRs/EP composites. It can be seen from the figure that the thermal conductivity increases with the increase of temperature, and the thermal conductivity is significantly improved after filling with the BNNRs aerogels. The thermal conductivity of nanocomposites increased from 0.03 to 0.1 at 77 K, and increased from 0.21 to 0.38 W·m−1 ·k−1 at 300 K. Based on the previously characterization, the improvement of thermal conductivity is mainly attributed to the following aspects. First, the obtained BNNRs has high crystallinity, and its crystal structure is consistent with that of h-BN. Compared with h-BN nanosheets, the long band-like structure and large aspect ratio of BNNRs can form a more continuous structure, resulting in the decrease of the formation of the thermal interfaces and phonon scattering. Second, the traditional two-dimensional dispersion way is difficult to form a long continuous structure in the case of low content, while three-dimensional structure can ensure the stable existence of a long continuous thermal conduction path structure, thus making a great improvement of the thermal conductivity of the nanocomposites.

Fig. 5. Thermal conductivity curves of EP and EP/BNNRs aerogel composites

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4 Conclusion In this work, we successfully prepared BNNRs aerogels by precursor pyrolysis and combined with epoxy resins to obtain nanocomposites. We found that the obtained BNNRs was consistent with the crystal form of h-BN by TEM and XRD. Its molecular structure was further confirmed by FTIR. We measured the thermal conductivity of the BNNRs/EP nanocomposite, which has a certain increase of 230% at 77 K and 80% at 300 K. Combined with the SEM images, it can be concluded that the continuous thermal conduction network of the aerogel helps to improve the thermal conductivity of the composite material. Therefore, we believe that the BNNRs aerogels as a filler can be a potential candidate in improving the thermal conductivity of epoxy in the low temperature fields, which provides the reliability operation of superconducting devices.

References 1. Li, S., et al.: Epoxy-functionalized polysiloxane reinforced epoxy resin for cryogenic application. J. Appl. Polym. Sci. 136(2), 46930 (2019) 2. Ren, L., et al.: Engineering the coefficient of thermal expansion and thermal conductivity of polymers filled with high aspect ratio silica nanofibers. Compos. B Eng. 58, 228–234 (2014) 3. Teng, C.-C., et al.: Synergetic effect of thermal conductive properties of epoxy composites containing functionalized multi-walled carbon nanotubes and aluminum nitride. Compos. B Eng. 43(2), 265–271 (2012) 4. Gu, J., et al.: Thermal conductivity epoxy resin composites filled with boron nitride. Polym. Adv. Technol. 23(6), 1025–1028 (2012) 5. Lee, Y.J., et al.: High voltage dielectric characteristics of epoxy nano-composites in liquid nitrogen for superconducting equipment. IEEE Trans Appl Supercond. 25(3), 8600104 (2015) 6. Shin, H.J., et al.: Effects of impregnating materials on thermal and electrical stabilities of the HTS racetrack pancake coils without turn-to-turn insulation. IEEE Trans. Appl. Superconduct. 23(3) (2013) 7. Takematsu, T., et al.: Degradation of the performance of a YBCO-coated conductor double pancake coil due to epoxy impregnation. Physica C: Superconduct. Appl. 470(17), 674–677 (2010) 8. Jiang, W., et al.: Thermo-mechanical behaviors of epoxy resins reinforced with nano-Al2O3 particles. J. Ind. Eng. Chem. 18(2), 594–596 (2012) 9. Peng, C., et al.: Synthesis of SiO2/epoxy–benzoxazine ternary copolymer via sol–gel method: thermal and mechanical behavior. Mater. Des. 111, 453–462 (2016) 10. Bagrets, N., et al.: Thermal and mechanical properties of advanced impregnation materials for HTS cables and coils. IOP Conf. Ser. Mater. Sci. Eng. (2015) 11. Ji, C., et al.: Ice-templated MXene/Ag–epoxy nanocomposites as high-performance thermal management materials. ACS Appl. Mater. Interfaces 12(21), 24298–24307 (2020) 12. Zhang, C., et al.: Self-assembled boron nitride nanotube reinforced graphene oxide aerogels for dielectric nanocomposites with high thermal management capability. ACS Appl. Mater. Interfaces 12(1), 1436–1443 (2020)

Low-Temperature Tensile and Impact Properties of Fe24Mn0.45C High-Manganese Steel Mingyue Jiang1,2 , Chuanjun Huang1(B) , Yuguo Chai3 , Meiyan Liu3 , Zhicong Miao1,2 , Rongjin Huang1,2(B) , and Laifeng Li1,2 1 CAS Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Beijing 100049, China {cjhuang,huangrongjin}@mail.ipc.ac.cn 2 University of Chinese Academy of Sciences, Beijing 100049, China 3 Technology Institute of Shougang Group Co., Ltd., Beijing 100043, China

Abstract. High manganese steels possess the same (similar) excellent lowtemperature properties as nickel-based steels and are of greater research value for low-temperature applications because of their fully austenitic structure and the absence of low-temperature ductile-brittle transition. In this work, we tested the low-temperature tensile and impact properties of Fe24Mn0.45C high manganese steel and evaluated its value for use in liquefied natural gas tanks. It is observed the yield strength and tensile strength increased, but the elongation decreased as the temperature decreased. At liquid nitrogen temperature, the Fe-24Mn-0.45C steel exhibited a yield strength of 871 MPa, a tensile strength of 1248 MPa, and an elongation of 29.5%. The impact absorbed energy of the Fe-24Mn-0.45C steel was 131 J at room temperature and 80 J at liquid nitrogen temperature. The test results of low-temperature mechanical properties indicate that the high manganese steel is promising to replace 9% Ni steel for LNG tank applications. Keywords: Low-temperature · Mechanical Properties · High Manganese Steel · Liquefied Natural Gas Tanks

1 Introduction With the rapid development of economic and industrial production, the demand for energy is increasing day by day. At the same time, in order to cope with the consequent global warming problem, people use a lot of clean energy like liquefied natural gas (LNG) to replace traditional fossil energy such as coal and oil. The temperature of liquefied natural gas is about −165 °C. LNG is usually transported by sea in specially equipped ships and stored in LNG terminals in spherical vessels. The materials used in the construction of LNG vessels and tanks are required to have good low-temperature properties, such as high strength and resistance to fracture as well as low cost. Most of the steels currently used in the construction of LNG tanks are nickel-based steels with nickel content generally ranging from 5–9% [1, 2], of which the most famous and widely used material is 9Ni steel. 9Ni steel is a low-temperature steel with a nickel © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1095–1100, 2023. https://doi.org/10.1007/978-981-99-6128-3_142

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mass content of 8.5% to 9.5% [3] developed in the USA in 1944 and has been used for more than 50 years. The first 9Ni steel LNG tank was built in Japan in1969 in Negishi. The use of 9Ni steel began in the 1980s in China, the first domestic enterprise to master the production technology is TISCO, and the production of steel has been used in LNG tanks. The 9Ni steel is a low-carbon modulated steel and its structure is ferrite, martensite and a small amount of austenite [4, 5] with a significant ductile-brittle transition. China is a nickel-poor country, whereas the price of manganese is much lower than that of nickel. In previous reports on high manganese steels in low-temperature applications [6], it can be seen that manganese-based steels [7] have low-temperature properties that are comparable to those of nickel-based steels. For example, high manganese steel (0.35–0.55 wt.%C, 22.5–25.5 wt.% Mn, 0.3–0.7 wt.% Cu) used in LNG tanks has a yield strength of 640 MPa and tensile strength of 982 MPa at low temperature (−165 °C) and an elongation of 42% [8]. In the present work, we carried out a study on the mechanical properties of high-manganese steels for low-temperature applications.

2 Experiment The experimental steel was a high manganese steel produced by the Technology Institute of Shougang Group that is expected to be used at −196 °C. Its chemical composition is shown in Table 1. For comparison, the chemical composition of ASTM A 553 (9% Ni steel) was also shown in the table. We purchased high manganese steel with 25% manganese content and 0.55% carbon content from the market as a control group, named SC. The produced ingot was heated treatment at 1200 °C for 1.5 h, followed by hot rolling, which was designated SG. A Bruker D8 X-ray diffractometer was used to investigate the phase composition at different temperatures (−196 °C and 20 °C). The tensile and impact properties at different temperatures (−196 °C and 20 °C) were tested by using a universal testing machine and a Charpy impact testing machine respectively. The fracture morphology of the samples after tensile testing was observed using a model S4800 scanning electron microscope (SEM). Table 1. The composition of SG and A553(ASTM) [9] steels.

A553 SG

C (%)

Mn (%)

Ni (%)

Cr (%)

P (%)

S (%)

≤0.13

≤ 0.9

8.5–9.5

/

≤0.035

≤0.035

24

/

3.5

≤0.050

≤0.050

0.45

3 Results and Discussions XRD results for the SG steel at −196 °C and 20 °C are shown in Fig. 1. The three most intense diffraction peaks correspond to (111), (200) and (220) of the face-centered cubic structure’s austenite phase. This indicates that the SG steel is a face-centered

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cubic austenitic structure. Therefore, a decrease in temperature to 77 K doesn’t cause phase transformation for the Fe24Mn0.45C high manganese steel. The low-temperature toughness of metals with the face-centered cubic structure is believed to be better than those with the body-centered structure because of its numerous slip systems such that low-temperature ductile-brittle transitions do not normally occur [10].

Fig. 1. XRD patterns of Fe24Mn0.45C (SG) samples at −196 °C and 20 °C.

Fig. 2. Quasi-static tensile stress-strain curves of the SG and SC high-Mn steels were tested at − 196 °C (LN) and 20 °C (RT).

The tensile tests were carried out on SG and SC steel samples at −196 °C and 20 °C respectively, and the results showed that SG steels had a yield strength of 322 MPa, a tensile strength of 814 MPa and an elongation of 54.0% at 20 °C, and a yield strength of 871 MPa, a tensile strength of 1248 MPa and an elongation of 29.42% at −196 °C. The tensile curves are shown in Fig. 2(a), and the specific experimental data are shown in Table 2. It can be seen that the yield strength and tensile strength of SG are higher than SC at 20 °C. Although the yield strength of both steel samples is comparable at −196 °C, the tensile strength of SG steel samples is significantly greater than that of SC steel samples. In addition, the elongation of the SG steel is much higher than that of the SC steel both at room temperature and at low temperature. Specifically, the SG sample steel showed an increase in yield strength by 5.7% and tensile strength by 11.3% at room temperature and an increase in tensile strength by 22.53% at −196 °C. This indicates that the lowtemperature performance of Fe24Mn0.45C high manganese steel produced by the STRI is significantly better. 9Ni steel has a yield strength of more than 585 MPa and maximum

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tensile strength of 690–825 MPa according to ASTM at room temperature [11]. It shows that the SG steel samples meet the basic requirements for cryogenic pressure vessels and are promising as low-temperature steel for LNG tanks. The impact tests were conducted on SG samples at −196 °C and 20 °C to further examine the low-temperature toughness of the Fe24Mn0.45C steel. The impact results are displayed in Table 2. The impact energy was 131 J at 20 °C and 80 J at −196 °C, and the impact energy decreased by about 39%. Compared with 63.9 J impact energy at − 196 °C for a high Mn steel with Fe-24 wt.%Mn-0.46 wt.%C [12], the low-temperature toughness of the sample steel in this experiment is better. As the temperature decreases, the toughness of SG steel decreases. We can also verify this conclusion from the fracture morphology of the tensile samples at different temperatures. 9Ni steel grade is specified to have an impact energy of more than 80 J at −196 °C [13]. In comparison, the impact energy of the Fe24Mn0.45C steel at −196 °C is slightly lower than that of the 9Ni steel. Since the steel in this experiment only underwent hot rolling treatment, it can be improved by a variety of heat treatments to improve its performance. Table 2. Tensile properties and Charpy impact energy at -196 °C and 20 °C. Temperature (°C)

Yield strength (MPa)

Tensile strength (MPa)

Elongation (%)

Impact energy (J)

SG-RT

20

322

814

54.0

131

SG-LN

−196

871

1248

29.5

80

SC-RT

20

305

732

15.5

SC-LN

−196

957

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2.0

It is observed that the SG sample steel is a transgranular fracture at 20 °C from the tensile fracture morphology as shown in Fig. 3(a) and the morphology of the lacerated fracture shows the dimple fracture surface by SEM as shown in Fig. 3(b). At low temperatures, the fracture mode gradually changes from transgranular fracture to intergranular fracture, and the fracture surface has a very obvious splitting arc. Figure 3(c) shows that the center of the fracture (Fig. 3(d)) still has a dimple fracture surface at −196 °C, which shows a transgranular fracture, whereas the fracture morphology of the edge part as shown in Fig. 3(e) shows a cleavage tissue indicating that the intergranular fracture occurred. This also indicates that the high manganese steel is slowly changing from ductile fracture to brittle fracture with the decrease in the temperature.

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Fig. 3. Tensile fracture morphologies of the (a) whole and (b) central position at 20 °C, and the (c) whole, (d) central and (e) edge of position at −196 °C

4 Conclusion High manganese steels have excellent mechanical properties at low temperatures and are suitable for many low-temperature applications, such as LNG tanks. In this work, the tensile and impact properties of Fe-24 wt.%Mn-0.45 wt.%C-3.5 wt.%Cr high manganese steel were tested at room temperature (20 °C) and low temperature (−196 °C). In the tensile test, the yield strength and tensile strength increased, but the elongation decreased as the temperature decreased. At −196 °C, Fe-24Mn-0.45C high manganese steel exhibited a yield strength of 871.33 MPa, a tensile strength of 1248.31 MPa, and an elongation of 29.42%. In the impact test, the impact absorbed energy of Fe-24Mn0.45C was 131 J at room temperature and 80 J at low temperature. The test results of

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low-temperature mechanical properties indicate that high manganese steel is promising to replace 9Ni steel for LNG tank applications in the future. Acknowledgments. The project is supported by the Strategic Priority Research Program of the Chinese Academy of Sciences (No. XDB25040300) and the Key Research Program of the Chinese Academy of Sciences (Grant No. ZDRW-CN-2021-4-1).

References 1. Furuya, H., Saitoh, N., et al.: Development of 6% nickel steel for LNG storage tanks. In: International Conference on Offshore Mechanics and Arctic Engineering, vol. 44359, pp. 327331 (2011) 2. Kamo, T., Arimochi, K., et al.: Development of 7% Ni-TMCP steel plate for LNG storage tanks. In: International Conference on Offshore Mechanics and Arctic Engineering, vol. 44359, pp. 113–122 (2011) 3. Nippes, E.F., et al.: Study of the weld heat-affected zone toughness of 9% nickel steel. Weld. J. 65(9), 237–243 (1986) 4. Wang, D.J., et al.: Phase transformation behavior and microstructure characterization of 9Ni cryogenic steel. Adv. Mater. Res. 472–475, 1183–1187 (2012) 5. Hany, S., et al.: Advanced characterization of cryogenic 9Ni steel using synchrotron radiation, neutron scattering and 57Fe Mössbauer spectroscopy. Mater. Des. 146, 219–227 (2018) 6. Han, I.W., et al.: Microstructure and mechanical properties of cryogenic high-manganese steel weld metal. Int. J. Offshore Polar Eng. 27(03), 260–265 (2017) 7. Fan, X., et al.: Mechanical properties of cryogenic high manganese steel joints filled with nickel-based materials by SMAW and SAW. Mater. Lett. 304, 130596 (2021) 8. An, G., et al.: Fracture toughness characteristics of high-manganese austenitic steel plate for application in a liquefied natural gas carrier. Metals 11(12), 2047 (2021) 9. ASTM International: Annual Book of ASTM-Standards, Section 1, Iron and Steel Products (2005) 10. Kim, H., et al.: Interpretation of cryogenic-temperature Charpy fracture initiation and propagation energies by microstructural evolution occurring during dynamic compressive test of austenitic Fe-(0.4, 1.0) C-18Mn steels. Mater. Sci. Eng. A 641, 340–347 (2015) 11. Hickmann, K., Kern, A., et al.: Production and properties of high strength nickel alloy steel plates for low temperature applications. In: Proceedings of the 1st International Conference on Super-High Strength Steels (2005) 12. Nam, Y.H., et al.: Low-temperature tensile and impact properties of hydrogen-charged highmanganese steel. Int. J. Hydrogen Energy 44(13), 7000–7013 (2019) 13. Kern, A., et al.: Development of 9% nickel steel for LNG applications. Steel Res. Int. 78(3), 189–194 (2007)

Thermoelectric Properties of PbSe0.9 Te0.1 at Cryogenic Temperature Haojian Su1,2 , Zhicong Miao1,2 , Shanshan Wu1,2 , Siyi Zhang1,2 , Haoying Qi1,2 , Min Zhou1(B) , Rongjin Huang1,2(B) , and Laifeng Li1,2 1 CAS Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry,

Beijing 100190, China {mzhou,huangrongjin}@mail.ipc.ac.cn 2 Centre of Materials Science and Optoelectronics Engineering, University of Chinese Academy of Sciences, Beijing 100049, China

Abstract. Thermoelectric materials are attractive for the reversible conversion between heat and electrical energy. Bulk PbSe0.9 Te0.1 were prepared by two processes. The electrical conductivity, Seebeck coefficients and thermal conductivity of two samples prepared by different techniques were investigated in the temperature range from 15 K to 300 K. The phase was analyzed by XRD and the microstructure was observed by SEM. The obtained results showed that the unannealed sample exhibited the optimal thermoelectric properties. Keywords: thermoelectric materials · chalcogenide · cryogenic temperature

1 Introduction Thermoelectric (TE) materials are the semiconductors which could enable direct energy conversion between heat and electricity. Compared to traditional mechanical refrigeration or a generator, TE materials take advantage of no moving parts, low environmental impact and high reliability. The TE performance is determined by the dimensionless figure of merit zT = σS2 T/κ, where σ is the electrical conductivity, S is the Seebeck coefficient, κ is the thermal conductivity and T is the absolute temperature [1]. Good TE materials are crystalline materials of “phonon glass-electron crystal”, which has both good electrical and thermal transport properties. The conversion efficiency of conventional TE materials, such as Bi2 Te3 alloys, lead chalcogenides and half-Heusler compounds, has been improved recently. However, the progress of TE materials which are applied below room temperature was relatively limited. And there is even less research on TE materials in the deep low temperature (below 100 K). In fact, low and deep low temperature TE materials have wide potential application in cooling of precision devices, such as the spacecraft infrared devices to replace the traditional mechanical refrigeration. At low and deep low temperature, the research mainly focused on Bi-Sb alloys. For example, the z value of Bi85 Sb15 single crystals could reach to 11 × 10−3 K−1 at 100 K [2]. It was reported that Bi- Sb alloys © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1101–1107, 2023. https://doi.org/10.1007/978-981-99-6128-3_143

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with zT value of 0.07 at 140 K was obtained through partial substitution of Ge for Sb [3]. In addition, the p-type single crystal CsBi4 Te6 reached to 0.82 at 225 K by doped with SbI3 [4]. Lead chalcogenides (PbS, PbSe, PbTe) are NaCl-type semiconductors with a narrow gap with a forbidden bands of about 0.3–0.4 eV at room temperature. PbTe as a kind of traditional TE materials has been studied and applied in spacecraft more than 50 years. PbSe has also been focused as an ideal TE material. PbTe-PbSe solutions have also drawn wide attention as a TE materials of middle temperature region. However, the study of PbTe-PbSe solid solution are only obtained at high temperature. There is little research on the low temperature or even deep low temperature properties of this material. Thermoelectric materials are important for energy transformation. However, their properties and application at low temperature are rarely studied. To extend the application of thermoelectric materials at low temperature, we perform a comparative study on the thermoelectric properties of the PbTe-PbSe solutions at low temperatures. PbSe0.9 Te0.1 were prepared by two kinds of processes. The electrical conductivity, the Seebeck coefficient and the thermal conductivity of these samples were examined in the range of 4.2–300 K. The phase composition and the microstructure were investigated using X-ray diffraction and scanning electron microscopy techniques, respectively.

2 Experimental Polycrystalline PbSe0.9 Te0.1 samples were prepared through two different processes, as shown in Fig. 1. High-purity element powers (Pb, 99.999%; Se, 99.9%; Te, 99.99%) were used as starting materials. They were weighted according to the stoichiometric ratio of PbSe0.9 Te0.1 and was loaded into vacuum sealed quartz ampoules in order to prevent possible oxidation. The tubes were slowly heated to 1393 K over 18 h, then kept for 10 h, and slowly cooling to room temperature. The obtained alloys were powdered with an agate mortar and then consolidated by spark plasma sintering (SPS) under an axial pressure of 50 MPa at 773 K for 5 min. As a contrast, a sample was annealed, which was sealed in evacuated quartz tubes and then heated up to 600 K for 3 days. For the measurement of the electrical and thermal transport properties, all samples were cut and polished. The phase composition was analyzed by X-ray diffraction (Bruker, Germany) with Cu Kα radiation (λ = 0.154056 nm). The microstructure was observed by scanning electron microscopy (SEM; Hitachi, S-4800, Japan). The densities of the samples were measured by the Archimedes method. The electrical resistivity, Seebeck coefficient and thermal diffusivity coefficient test were simultaneously performed in the range of 15–300 K using a Physical Property Measurement System (Quantum Design, USA). The electrical conductivity was characterized by using a four-probe method. The Seebeck coefficient was determined from the measured temperature and electric potential difference between the two ends of a bar-shaped specimen.

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Fig. 1. Schematic diagram for the synthesis of PbSe0.9 Te0.1 .

3 Results and Discussion The room temperature PXRD pattern of PbSe0.9 Te0.1 prepared by two processes are shown in Fig. 2. All the samples are single phase without obvious impurities within our detection limit, which indicate that PbTe-PbSe solid solution have been prepared. There are no visible diffraction peaks corresponding to pure element or oxidation after consolidation. There is no detectable change found in the crystal structure expect the intensity of diffraction peaks. The difference in the intensity may be due to the change of grain size.

Fig. 2. The PXRD patterns of PbSe0.9 Te0.1 .

Figure 3 shows the SEM images of the fracture surface of the sintered bulks PbSe0.9 Te0.1 . The samples are very condensed, though there are some cracks, which is consistent with the density test results. Compared with unannealed sample, it is clear to see that the sample of annealing for 3 days have more long and straight stripes in it. This indicates that heat treatment has a significant effect on the microstructure of the PbSe0.9 Te0.1 sample. The stripes may be caused by the isoelectronic substitution

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[5], which may be more obvious after heat treatment. Changes in microstructure can determine the properties of materials.

Fig. 3. Typical SEM images of the PbSe0.9 Te0.1 prepared by (a) melting and SPS and (b) annealing for 3 days.

Figure 4 shows the temperature dependence of the electrical and thermal properties of PbSe0.9 Te0.1 . The electrical conductivity shows a decreasing trend with increasing temperature, which depicts heavily doped semiconducting behavior. The electrical conductivity of unannealed sample is higher than that of the annealed for 3 days sample below 250 K, which may be due to the microscopic cracks and the change of electric structure. The slight evaporation of Se results in over stoichiometry of Pb ions which may lead to suppression of positive ions concentration because of net charge balance [6], which could be another reason. In addition, the element homogeneity and relaxation of strain are also the direct results of high temperature annealing, which may lead to the decreased of electrical conductivity. The Seebeck coefficients of all the samples are positive across the whole temperature range, indicating the samples are p-type semiconductors. Nevertheless, the Seebeck coefficient is increased with increasing temperature. The Seebeck coefficient shows an opposite behavior to the electrical conductivity, due to its inversely related to the carrier concentration. The maximum value of the positive Seebeck coefficient is found for the unannealed sample. The difference in Seebeck coefficient for two samples is small below 200 K, but at higher temperature it started differ. According to the Seebeck coefficient formula, the Seebeck coefficient of materials is related to the scattering factor and carrier concentration. The formation of solid solution will cause the lattice distortion of materials, resulting in the increase of the point defects and grain boundaries.

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The power factor of all samples calculated based upon the Seebeck coefficient and electrical conductivity is shown in Fig. 4c. The power factor calculated from the Seebeck coefficient and electrical conductivity. It can be seen that the power factor presents an upward trend with the increasing temperature in the same way as the Seebeck coefficient. It is also noted that the power factor reduces after the annealed process. In brief, PF of unannealed sample is higher than that of the sample processed through annealed for 3 days. As shown in the figure, the maximum value of unannealed sample is 5679 × 10−12 W m−1 K−2 . The power factor for the sample of annealed for 3 days is much lower than the unannealed sample due to the low Seebeck coefficient. The grain size may be another reason to suppress the Seebeck coefficient and the power factor. As shown in Fig. 3, there are some differences of grain size between the two samples. Figure 4d presents the temperature dependence of thermal conductivity data of all samples. The heat conduction process of materials is mainly determined by carrier movement and lattice vibration. At low temperature, the thermal conductivity is mainly determined by the lattice thermal conductivity. There is no obvious difference between the two sample. The thermal conductivity of PbSe0.9 Te0.1 all decrease with the increase of temperature, due to enhanced phonon scattering. The minimum total thermal conductivity is about 1.1 W m−1 K−1 at 300 K. These values are slightly lower than that of undoped PbSe and PbTe.

Fig. 4. Temperature dependence of electrical and thermal properties of PbSe0.9 Te0.1 . (a) electrical conductivity, (b) Seebeck coefficient, (c) power factor and (d) total thermal conductivity.

According to the calculated power factor and the measured thermal conductivity, the temperature dependence of the dimensionless figure of merits of PbSe0.9 Te0.1 of all samples are shown in Fig. 5. The zT value increase slightly with an increase of temperature. The two lines are essentially coincident within the temperature range of 15 K to 100 K. Above 100 K, the dimensionless of figure of merits of the unannealed

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Fig. 5. Temperature dependence of zT for PbSe0.9 Te0.1 .

sample went up sharply. It is found that the tendency of the zT values is similar with the power factors.

4 Conclusions P-type PbSe0.9 Te0.1 bulk thermoelectric materials were prepared by two processes. XRD analysis confirms that all the samples with a single phase. The thermoelectric transport characteristics of the PbSe-PbTe solid solution at low temperature were studied for the first time. The measurements of Seebeck coefficients indicate that the samples are p-type thermoelectric materials. And the properties of the samples prepared by two processes were compared. The unannealed sample shows a high thermoelectric power factor and maximum zT. Overall, a significant change in the structural and thermoelectric properties were observed which validate the worthiness of the opted route. Acknowledgments. The project is supported by National Natural Science Foundation of China (Number. 51872299, 52071223), S&T Innovation 2025 Major Special Program of Ningbo (Number. 2019B10085), the National Key Research and Development Program of China (Number. 2019YFA0704904), the Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry (TIPC), Chinese Academy of Sciences (CAS) (Number. CRYO202106).

References 1. Zhu, T., et al.: Compromise and synergy in high-efficiency thermoelectric materials. Adv. Mater. 29(14), 1605884 (2017) 2. Yim, W.M., Amith, A.: Bi-Sb alloys for magneto-thermoelectric and thermomagnetic cooling. Solid State Electron. 15(10), 1141–1165 (1972) 3. Chen, Z., et al.: Thermoelectric properties of Ge-doped Bi85Sb15 alloys at low temperatures. J. Phys. Chem. Solids 75(4), 523–527 (2014) 4. Chung, D.Y., et al.: A new thermoelectric material: CsBi4Te6. J. Am. Chem. Soc. 126(20), 6414–6428 (2004)

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5. Fan, H., et al.: Enhanced low temperature thermoelectric performance and weakly temperaturedependent figure-of-merit values of PbTe–PbSe solid solutions. J. Alloy. Compd. 658, 885–890 (2016) 6. Butt, S., et al.: One-step rapid synthesis of Cu2Se with enhanced thermoelectric properties. J. Alloy. Compd. 786, 557–564 (2019)

Measurement of Apparent Strain of Foil Strain Gauge at Low Temperature Sara Sato(B) , Minoru Takeda, and Kazuma Maekawa Graduate School of Maritime Sciences, Kobe University, Kobe, Japan {sato.sara,takeda,maekawa}@maritime.kobe-u.ac.jp

Abstract. Marine transportation of large quantity of liquid hydrogen (LH2 : 20 K) from Australia to Japan by a tanker is expected to be an international hydrogen energy supply chain system. It is important to develop the LH2 loading/unloading technology as well as the LH2 storage/transportation technology to build the supply chain system. A new LH2 flowmeter, which is utilized a helical flow, will be noted as a key technology for LH2 loading/unloading. The helical flow based LH2 flowmeter consists of an GFRP pipe, a helical wall and foil strain gauges, which detects the pressure change inside the pipe due to the LH2 helical flow. An apparent strain of a foil strain gauge at low temperature is measured using a LHe/LN2 experimental device. Experimental results are discussed on the basis of the thermal and electrical characteristics of the stain gauges. Keywords: Foil strain gauge · Temperature dependence of apparent strain · GFRP pipe · Liquid hydrogen flowmeter · Helical flow

1 Introduction In order to establish an international hydrogen supply chain for a hydrogen-based energy society [1], liquid hydrogen (LH2 ) loading/unloading technology must be constructed. Therefore, it is necessary to develop a flowmeter that can be used even at LH2 temperature. Thus, a new LH2 flowmeter, which is utilized a helical flow, has been planed. This flowmeter consists of a GFRP pipe, a helical wall, and foil strain gauges, which detects the pressure change inside the pipe due to the LH2 helical flow as shown Fig. 1. For example, when LH2 flows through this flowmeter made with a pipe of 26 mm diameter and 1 mm thickness, a pressure of 0.010 MPaG at 30 L/min is expected. For the development of this flowmeter, it is important to investigate the temperature characteristics of the foil strain gauge at low temperature. Therefore, to understand the apparent strain of the cryogenic strain gauge as a function of temperature, strain measurements were made using a copper plate and LHe/LN2 .

© Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1108–1115, 2023. https://doi.org/10.1007/978-981-99-6128-3_144

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Strain gauge

Fig. 1. Schematic of LH2 flowmeter using strain gauge with helical flow.

2 Experimental 2.1 Measurement Principle of Strain Gauge Measurement of strain by means of strain gauges is utilized the fact that the resistance value changes due to elasticity of th e internal metallic resistive element caused by external force. The resistance value R of strain gauge is changed to R + R when stain ε is applied to the metallic element, then the following relationship is obtained. R = Ks · ε, R

(1)

where Ks: gauge factor. The gauge factor Ks is a coefficient that expresses the sensitivity of the strain gauge and, in general, approximately 2 [2]. Although the gauge factor increases linearly with decreasing temperature, the rate of change is less than 5% even when the temperature decreases from room temperature (300 K) to 77 K. Therefore, the temperature dependence of the gauge factor is usually not considered and is assumed to be constant. Strain gauges glued to the measured object, e.g. copper plate, detect strain due to temperature change (apparent strain, which includes strain of the glue) in addition to strain due to external force. Therefore, the following equation related to the measurement value εm , which is detected by the strain, can be derived. εm = εf + εT ,

(2)

where εf : external strain, εT : apparent strain. In this study, the experiment is made without applying any external force except for temperature change to the measured object, so Eq. (2) becomes Eq. (3). εm = εT ,

(3)

In addition, εT is a function of temperature and is expressed by the following equation.   α (4) + βs − βg , εT = Ks

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where α: temperature coefficient for resistance of the metallic resistive element, β s : linear expansion coefficient of the measured object, and β g : linear expansion coefficient of the metallic resistive element. By transforming Eq. (4) and defining εg as the apparent strain of the strain gauge itself due to temperature change, the following equation is derived. εg =

α − βg = εT − βs , Ks

(5)

εg shows the intrinsic value of the strain gauge, which is not affected by the quality of the measured object. Then, Eq. (5) becomes Eq. (6). εT = εg + βs ,

(6)

From Eqs. (3) and (6), the following equation is derived. εg = εm − βs ,

(7)

Using Eq. (7), εg can be obtained experimentally. 2.2 Foil Strain Gauge

3.7 mm

Kyowa Electronic Instruments KFL-5-120-C1-11F3M3 strain gauges, which is shown in Fig. 2, were used in the experiment. Table 1 shows specification of the foil strain gauge. The sample was made by attaching it to a copper plate (17 mm × 24 mm × 0.2 mm). In the case at temperature under the self-temperature compensation range, temperature compensation must be performed. EP-270 was the epoxy-based glue used to bond the strain gauges. It cures by mixing two different liquids at room temperature. The applicable temperature range is between −269 °C and 30 °C. Effective adhesion is obtained under condition of a pressure (50 ± 20 kPa) at room temperature for 24 h [2].

10 mm

Fig. 2. Photo of the foil strain gauge.

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Table 1. Specifications of the foil strain gauge [2]. Linear Expansion Coefficient

×10–6 /°C

11

Gauge Resistance



120

Gauge Factor

—

2.1

Temperature Compensated Range

°C

−196–50

2.3 Experimental Apparatus Figure 3 shows the low temperature experiment system used in this study. A sample was mounted on the sample holder, and a low temperature experiment probe was inserted into a LHe Dewar. The sample holder has a front side and a back side. The sample was mounted on a front side, and a thermometer was attached on a back side. Two types of thermometers, a platinum thermometer and a Cernox thermometer, were attached to the sample holder. LabVIEW was used as the temperature measurement software. The strain measurement system consists of a bridge box, a strain measurement unit, a control unit, and a PC. The bridge applied voltage for strain measurement is 2 V. Kyowa Electronic Instrument’s dynamic data acquisition software DCS-100 A was used for strain measurement. Current source ( behind the PC ) Low temperature experiment probe PC

Voltmeter

Feedthrough

Compound gauge

Vacuum jacket Sample holder

Suspension unit

Thermometer side

Sample side

LHe Dewar

Fig. 3. Low temperature experiment system.

2.4 Experimental Method The following two experiments were made as shown in Figs. 4 and 5. Experiment 1: Experiments using the 3-wire method with one active gauge

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The value of linear expansion of copper with the strain gauge was measured by means of this method. Experiment 2: Experiments using the orthogonal arrangement method with two active gauges The longer side of the strain gauge was used as the longitudinal direction, and the difference in output between the strain gauge mounted longitudinal directions and that mounted transverse directions was examined to confirm the 2D elastic uniformity.

r

Output voltage e

r

Rg

R r R R Bridge constant voltage

E

r

R

R

r r

r R

R

Output voltage e

Fig. 4. Experiments using the 3-wire method with one active gauge.

Bridge constant voltage E

Fig. 5. Experiments using the orthogonal arrangement method with two active gauges.

Experiments 1 and 2 were made using the following procedure. (1) Attach the strain gauge to the copper plate with glue of EP-270. (2) Attach the sample to the sample holder of the low temperature experiment probe. (3) Vacuum inside the low temperature experiment probe, and then fill it with a small amount of helium gas. (4) Insert the low temperature experiment probe into LHe Dewar and change the temperature using LHe/LN2 .

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(5) Simultaneously measure temperature and strain, and analyze using the measurement values obtained and the value of linear expansion of copper. The temperature was measured both in the temperature increase condition (Up) and in the temperature decrease condition (Down). The rate of temperature change was less than 0.1 K/s from room temperature to 200 K, and less than 0.05 K/s at the lower temperatures.

3 Experimental Results and Discussion The results of Experiment 1 and Experiment 2 are shown in Figs. 6 and 7, respectively. In Experiment 2, the Poisson’s ratio of copper was set to ν = 0.343, and Fig. 7 shows the values which corresponds to measurement values divided by (1 + ν). 500

Strain

-500 0

100

200

300

-1500 -2500

LHe Exp.1 Down LHe Exp.1 Up

-3500

Temperature (

Fig. 6. Temperature Up and Down results of experiment 1.

Strain (με)

500 -500 0

100

200

300

-1500 -2500 -3500

LHe Exp. 2 Down LHe Exp. 2 Up

Temperature (K)



Fig. 7. Temperature Up and Down results of experiment 2.

Hysteresis was observed from 20 K to 150 K in Experiment 1. On the other hand, no hysteresis was observed in Experiment 2. The cause of hysteresis is discussed as follows. There are two possible causes of hysteresis in Experiment 1. (i) Problem with the rate of temperature change

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(ii) Problem of the thermometer itself We decided to check these problems by using LN2 . In particular, for (ii), we used a Cernox thermometer in LHe experiment, thus we decided to measure the temperature by using a platinum thermometer as an addition one. Figure 8 summarizes the results of Experiment 1 using LHe and LN2 . No hysteresis was observed in the Experiment 1 using LN2 . It is probably that hysteresis did not occur in the experiment 1 using LN2 because the rate of temperature change was slower using LN2 than using LHe. Therefore, the cause of hysteresis in Experiment 1 is believed to be (i). From the above, it was determined that temperature Up value were more reliable than temperature Down value for Experiment 1 using LHe. 1500 500

Strain

-500 0 -1500 -2500 -3500

100

200

300 LHe Exp.1 Up LHe Exp.1 Down LN2 Exp.1 UP LN2 Exp.1 Down

Temperature (

Fig. 8. Temperature Up and Down results of experiment 1 using LHe and LN2 .

The apparent strain εg can be calculated from three values: the linear expansion of copper β s , the temperature Up value of Experiment 1 using LHe, and the average value of the temperature Up and Down of Experiment 2, and using the following equation. εg = (LHeExp.1Up) − (LHeExp.2Ave.) − βs ,

(8)

Figure 9 shows comparison between the experimental value and the calculated value of apparent strain. The calculated value was based on Eq. (7), the value provided by the manufacturer [3] and the value of linear expansion of SS400 (11.8 × 10−6 /°C). The temperature dependence of linear expansion of SS400 was estimated by calculation using that of SUS304 which data is well known because of no data of SS400. The experimental value based on Eq. (8) was shifted parallel to the y direction so that they are 0 με at 319 K for easy comparison. From room temperature to 77 K, experimental values are very close to the calculated values. The results show that it is possible to measure the apparent strain of strain gauges at low temperature with this method. Finally, we discuss the sudden increase in apparent strain below 20 K as is seen in reference [4]. At low temperatures, the linear expansion of copper is saturated to low value. On the other hand, the metallic resistance of the strain gauge continues to decrease with decreasing temperature. Therefore, the sudden increase in apparent strain is considered to be caused by the condition in which the metallic resistance decreases while the linear expansion remains unchanged.

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Experimental value Calculated value

2000 1000 0 0

100

200

Temperature (K)

300

400

Fig. 9. Comparison between experimental value and calculated value of ε g .

4 Summary This study can be summarized as follows: (1) Strain measurements were made using copper plates and LHe/LN2 to obtain the apparent strain of foil strain gauges at low temperatures as a function of temperature. (2) The experimental values of apparent strain were compared with the calculated values with discussion. (3) The experimental values were very close to the calculated values. (4) Below 20 K, the apparent strain increased suddenly. This is thought to be caused by the thermal and electrical characteristics of the strain gauges.

References 1. Kamiya, S., Nishimura, M., Harada, E.: Study on introduction of CO2 free energy to Japan with liquid hydrogen. Phys. Procedia 67, 11–19 (2015) 2. Kyowa Electronic Instruments Co., Ltd: General Catalog of Measuring Components (2019) 3. Kyowa Electronic Instruments Co., Ltd: Engineering Data Sheet (B), KFL-5-120-C1-11 F3M3, Lot No. Y070S-050A 4. Kyowa Electronic Instruments Co., Ltd: Technical Report, No. 265 (1980) in Japanese

Superconducting Materials and Devices

Breakdown Characteristics of Liquid Nitrogen/Tetrafluoromethane/ Polypropylene Laminated Paper Insulation System Utilized for Superconducting Energy Pipeline Zhihao Zhou1,2,3 , Qingquan Qiu1,2(B) , Yuping Teng1,2 , Liwei Jing1,2 , Naihao Song1,2 , Jingye Zhang1,2 , Guomin Zhang1,2,3 , and Liye Xiao1,2,3

3

1 Key Laboratory of Applied Superconductivity, Chinese Academy of Sciences, Beijing 100190, China 2 Institute of Electrical Engineering, Chinese Academy of Sciences, Beijing 100190, China [email protected] University of Chinese Academy of Sciences, Beijing 100049, China

Abstract. Liquid nitrogen/tetrafluoromethane (LN2 /CF4 ) mixture with a wide liquefied temperature range of 50 to 100 K might be a promising coolant for high-Tc superconducting (HTS) apparatus. However, certain features have not yet been clarified before further application, particularly the insulation properties regarding the composite insulation system consisting of such a cryogenic mixture and polypropylene laminated paper (PPLP). In this paper, AC and DC breakdown experiments and simulations of LN2 /CF4 - PPLP composite insulation system are carried out with various molar fractions of each constituent part taken into account using sample cables insulated by the foregoing composite insulation. Results indicate that such a mixture/PPLP composite insulation system possess a superior breakdown strength compared to pure LN2 /PPLP system; LN2 /CF4 can therefore be an advantageous choice for HTS power apparatus. Keywords: LN2 /CF4 Cryogenic Mixture · Composite Insulation · Breakdown Characteristic · Insulation Property · Superconducting Energy Pipeline

1

Introduction

In the past decades, the development of high-Tc superconducting (HTS) materials and cryogenic technologies have boosted corresponding applications of HTS apparatus in power grid such as transformer, fault current limiter and cable [1]. Besides the conventional cable, a kind of comprehensive device named energy pipeline is gradually regarded as a promising application, which can transmit c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1119–1125, 2023. https://doi.org/10.1007/978-981-99-6128-3_145

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electricity and liquid natural gas (LNG) at the same time [2]. One of the most important indices for assessing the reliability of such kind of power apparatus is breakdown strength of its main insulation [3]. A large number of studies have been carried out to investigate the breakdown characteristics of the commonly used polypropylene laminated paper (PPLP) immersed in liquid nitrogen (LN2 ) [4,5]; results indicate that LN2 can infiltrate into PPLP to form a composite insulation system which is suitable for use as main insulator. However, the pressure of saturated nitrogen is almost 0.8 MPa in the temperature range of LNG which may threaten the safe operation of HTS apparatus [6]. Fortunately, a kind of liquid mixture, LN2 /CF4 , possesses a wide liquid-phase temperature range from 50 to 100 K @ 0.5 MPa, which may be a promising coolant for energy pipeline [7]. Investigations regarding the molecular properties [8] and dielectric strengths [9] of LN2 /CF4 have documented recently but research about breakdown strength of a LN2 /CF4 -PPLP composite insulation system can hardly be found. Therefore, it is significant to carry out simulations and experiments concerning the electrical properties of LN2 /CF4 -PPLP and figure out the feasibility of using such a composite system as main insulator for energy pipeline. In this paper, AC and DC breakdown experiments are carried out; the liquid components of the tested composite insulation systems contains 60, 80 100 mol% of N2 respectively. Moreover, the electric field distributions are simulated through finite element method.

2 2.1

Experimental and Computational Detail Experimental Setup

All of the experiments are carried out using a home-made pipe Dewar as depicted in Fig. 1(a). The whole pipeline is hermetically sealed and immersed in a LN2 bath during the experiment. The mass of each component is weighted through a high-precision balance and the molar fraction of nitrogen can be determined. To investigate the insulation properties of the composite insulation system, a PPLP insulated sample cable is submerged in such a dielectric liquid mixture. In addition, high voltage from the standard high-voltage generator is connected to the cable core through a voltage lead whereas another ring-shape copper electrode surrounding the PPLP will be grounded together with the chamber. During each experiment, the high voltage applied to the sample cable would last for 60 s from 10 kV increasing up to the breakdown voltage with an interval of 2 kV. The mentioned PPLP insulated sample cable consists of five main layers in total. As depicted in Fig. 1(b), from inner to outer layers, they are the inner carbon paper, the 1st layer PPLP, the 2nd layer PPLP, stress cone and the outer carbon paper. Amongst them, carbon papers are used for smoothing the electric fields at both high-voltage and grounding electrodes; stress cone is adopted to adjust the distribution of electric field at cable ends. Thus, the concentrated electric field can be effectively screened, and eventually the electric field exerted

Breakdown Characteristics

Stress Cone

Outer

Gas Exit Manifold Gauges

Carbon Paper Grounding Ring

P1

CF4 N2

2nd Layer PPLP

Gas Entrance Pressure Gauge

1st Layer PPLP

N2/CF4

Carbon Paper

g

Inner

Gas Tank Gas Storage Cylinders

LN2 Bath

Electronic Balance

Pump

Copper Core

Experimental Setup

(b)

Electrical Insulation

(a) Line 2

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Ground

Line 1

Unit: mm

Polypropylene

Kraft

0.045

Kraft Liquid Medium

0.125 0.04 High Voltage

(c)

Polypropylene Liquid Medium

Copper Core

(d)

Fig. 1. Experimental setup and simulation model. (a) Photograph of experimental Dewar and the PPLP insulated sample cable. (b) Schematic diagram of sample cable for each layer. (c) Schematic diagram of the whole setup including the mixture preparation system. (d) Cross sections of the sample cable.

on the two-layer PPLP sample would be a relatively uniform field between the two layers of carbon paper. The two layers of PPLP in the middle of the cable constitute the solid component of the composite insulation system. It should be noted that butt gaps are evenly arranged during twining the PPLP in order to avoid been torn by bending [10]. In order to have a further understanding of electric field distribution in the foregoing composite insulation system, the PPLP as well as the liquid medium are modelled in COMSOL Multiphysics for finite element calculation. Details about the simulation model are shown in Fig. 1 (c) and (d). The main governing equation of this simulation are exhibited in Eqs. (1),   ∂ (1) ∇ · σ(−∇V ) + ε0 εr (−∇V ) = Q, ∂t where σ represents the electrical conductivity; V denotes the voltage potential; ε0 and εr are the vacuum and relative permittivity respectively; Q refers to the corresponding flux source of charge. As depicted in Fig. 1 (c), both the two mentioned cases, with and without butt gap, are considered in the simulation model. Correspondingly, two paths going through such two cases are defined and illustrated by notations Line 1 and 2. Line 1 represents the region without butt gap; therefore the two PPLP layers are tightly attached to each other. Line 2 refers to the region having a butt gap; thus the liquid medium would fulfill the gap. The high-voltage and grounding boundaries are arranged at the top and bottom of the model respectively. Other boundaries are assigned as electrical insulation. The mesh adopts quadrilateral

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Table 1. Physical properties of the model components used in the simulation. Item 60 mol%

80 mol%

100 mol% PP

Kraft

εr

1.59

1.52

1.45

1.03

σ

1.484 × 10−14 1.742 × 10−14 2 × 10−14 6.8 × 10−20 2.44 × 10−14

2.3

9.8

DC AC

9.7

27

9.6 9.5

26

9.4 25 9.3 24

23

AC Breakdown Voltage (kV)

DC Breakdown Voltage (kV)

28

DC

9.2

0.6

0.8

1.0

Molar Fraction of N2

(a)

9.1

Photographs of Cable after Breakdown

(b)

Fig. 2. Results of electrical breakdown experiments under AC and DC. (a) Breakdown strengths of composite insulation system. (b) Photographs of sample cable after breakdown.

meshes; the maximum and minimum unit is assigned as 0.2 and 0.1 times the thickness of the PPLP, respectively. Three main types of applied external voltage are taken into consideration including AC, DC and transient DC. Amongst them, transient DC takes the voltage rising process into account, which is a capacitive field. The frequency, denoted by f , of AC voltage is 50 Hz; the rising rate of voltage k is 1 kV/s for transient DC. The thicknesses of liquid medium (dL ), Kraft paper (dK ) and polypropylene (dP ) are 0.125, 0.04 and 0.045 mm respectively; the applied voltage (in the AC case it denotes the voltage apex) U is 30 kV (Table 1).

3 3.1

Results and Discussions Breakdown Strength

Insulation performances of LN2 /CF4 - PPLP composite insulation systems with 60, 80 and 100 mol% of N2 for liquid component under AC and DC voltages are investigated through electrical breakdown experiments, corresponding strengths are shown in Fig. 2 (a). Two main results can be concluded from the curves. To begin with, breakdown strengths of the composite insulation system under DC voltage are almost three times higher than that under AC voltage. It can be found that the breakdown level is around 27 kV for DC tests whereas the insulation

Breakdown Characteristics

108 72

106 104 102 100

70 80 mol% of N2

160 158 156 154

116

84

114 82

112 110 108

80

106 104

152

(a)

80 mol% of N2

×10-3

Loss Power Dielectric Loss Power (mW/cm3)

Electric Field (kV/mm)

140

120

AC-Line1 DC-Line1 AC-Line2 DC-Line2

0.05

0.10

0.15

0.20

0.25

8.0

Loss Energy

0.3

6.0

0.2

4.0

0.1

2.0

0.0

0.00

135

(b)

160

60

140

100 mol% of N2

0.4

80

145

Molar Fraction of N2

Molar Fraction of N2

100

150

78 60 mol% of N2

100 mol% of N2

155

0.00

Dielectric Loss Energy (J/cm3)

60 mol% of N2

162

86

LN2 PP Kraft

Electric Field in Kraft Region (kV/mm)

110

164

118

Electric Field in PP Region (kV/mm)

74

112

120

Electric Field in Liquid Region (kV/mm)

114

166

Electric Field in Kraft Region (kV/mm)

76

LN2 PP Kraft

116

Electric Field in PP Region (kV/mm)

Electric Field in Liquid Region (kV/mm)

118

1123

0.0 0.01

0.02

Position (mm)

Time (ms)

(c)

(d)

0.03

0.04

Fig. 3. Results of electric field distribution simulations. (a) Histograms of electric field in liquid, Kraft and PPLP regions under AC voltage. (b) Histograms of electric field in liquid region, Kraft region and PPLP regions under DC voltage. (c) Electric field distribution along paths Line 1 and 2 under AC and DC. (d) Power and energy of dielectric loss under AC.

strength is approximate 9.4 kV for AC tests. Moreover, a slight enhancement of breakdown voltage can be found with the decrease of molar fraction of N2 . Experimental phenomena can be quite different in terms of AC and DC breakdown as depicted in Fig. 2 (b). It can be found that DC breakdown is more destructive compared to AC breakdown. Firstly, the blast caused by DC breakdown would tear a big hole on the two-layer PPLP; on the contrast, only a small point can be found on the AC condition. Secondly, an intense deformation of the grounded ring-shape copper electrode can be found after DC breakdown whereas merely a small black hole can be observed after AC breakdown. 3.2

Electric Field Distribution

The differences of breakdown strengths may be attributed to the maximum electric field at the electrically weakest point. In order to have a better understanding of the distribution of electric field, simulations regarding the electric

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field, denoted by E, of the composite insulation system are performed. Figures 3 (a) and (b) summarize the histograms of E of the composite insulation system in different regions under capacitive AC and DC conditions respectively. In terms of such a capacitive field, E will be mainly distributed according to the permittivities of materials. In general, the liquid medium is the weak region of the whole composite insulator. A slight descending of E in liquid region can be observed with the reduce of molar fraction of N2 , which explains the slight enhancement of breakdown voltage. Moreover, a contrast of E distributions under AC and DC voltage may reveal the difference between their breakdown voltages. Take the pure LN2 as an instance, E along the paths Line 1 and 2 (marked in Fig. 1 (c)) under both AC and DC are plotted in Fig. 3 (c). According to the E distribution, there are two main reasons why AC breakdown voltage is much lower than DC. To begin with, the E distribution under DC is more reasonable than that under AC no matter along line 1 or 2. According to the simulation results, in the case that both apexes of applied voltage of AC and DC are 30 kV, the calculated E in Kraft region alone Line 1 under AC condition is approximate 10 kV/mm higher than that under DC. Moreover, the existence of dielectric loss would gradually age the insulators. Figure 3 (d) exhibits loss power and energy of the whole composite insulation system. It can be found that the frequency of loss power is two times of that of applied AC voltage (which is 50 Hz).

4

Conclusion

LN2 /CF4 can be a promising coolant for HTS apparatus and it is important to figure out the electrical breakdown strength of the LN2 /CF4 - PPLP composite insulation system before practical applications. The main findings of this work are summarized as follows. 1. Breakdown strength of the composite insulation system can be slightly enhanced (12% for DC and 4% for AC) by increasing the molar fraction of CF4 . The added CF4 can raise the overall permittivity of liquid mixture and thus lower the distributed electric field in liquid region, which explains the elevation of breakdown strength. 2. Breakdown strength of the composite insulation system under dc voltage is almost 3 times higher than that under ac voltage. An inferior E distribution and the existence of dielectric loss are the two main reasons why breakdown is easier to happen under AC voltage. 3. LN2 /CF4 - PPLP composite insulation system may be an advantageous choice for HTS power apparatus. The limitation of this current work lies in the lack of breakdown data, so it can hardly carry out statistical analysis. In the future, more breakdown experiments under different kinds of voltage forms should be carried out for Weibull distribution analysis. Moreover, the mechanisms behind the enhancement of the breakdown strength after adding CF4 can also become an exciting research subject.

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Acknowledgment. This work was supported by the National Key R&D Project of China (No. 2018YFA0704203 and No. 2018YFB0904400), the National Natural Science Foundation of China (Science Fund for Creative Research Groups, Grant No. 51721005) and the Key Research Program of Frontier Sciences, Chinese Academy of Sciences (Grant No. QYZDJ-SSW-JSC025).

References 1. Xiao, L., Lin, L.: Recent progress of power application of superconductor in China. IEEE Trans. Appl. Supercond. 17(2), 2355–2360 (2007) 2. Qiu, Q., et al.: Design and testing of a 10 kV/1 kA superconducting energy pipeline prototype for electric power and liquid natural gas transportation. Supercond. Sci. Technol. 33(9), 095007 (2020) 3. Zhang, C., et al.: Breakdown and flashover properties of cryogenic liquid fuel for superconducting energy pipeline. IEEE Trans. Appl. Supercond. 32(3), 1–7 (2022) 4. Seong, J.-K., Choi, W., Shin, W.-J., Hwang, J.-S., Lee, B.-W.: Experimental and analytical study on DC breakdown characteristics of butt gap condition in LN2 /PPLP composite system. IEEE Trans. Appl. Supercond. 23(3), 5401604– 5401604 (2013) 5. Nguyen, D.V., Le, T.V., Kim, S.H.: AC breakdown channel of PPLP multi-layer insulation for HTS cable. Cryogenics 108, 103072 (2020) 6. Span, R., Lemmon, E.W., Jacobsen, R.T., Wagner, W., Yokozeki, A.: A reference equation of state for the thermodynamic properties of nitrogen for temperatures from 63.151 to 1000 K and pressures to 2200 MPa. J. Phys. Chem. Ref. Data 29(6), 1361–1433 (2000) 7. Qiu, Q., et al.: General design of ±100 kV/1 kA energy pipeline for electric power and LNG transportation. Cryogenics 109, 103120 (2020) 8. Zhou, Z., et al.: Calculations and analyses of molecular features and properties of nitrogen/carbon tetrafluoride mixture. Comput. Theor. Chem. 1204, 113411 (2021) 9. Chen, J., et al.: Electrical insulation characteristics of LNE 2/CF4 mixture at cryogenic temperatures. IEEE Trans. Appl. Supercond. 31(1), 1–6 (2020) 10. Rezaeifar, F., Suzuki, Y., Kumada, A., Hidaka, K., Nishimura, T., Masuda, T.: Characterization of partial discharge in composite insulation system with PPLPR for HTS cable. IEEE Trans. Dielectr. Electr. Insul. 17(6), 1747–1753 (2010)

Preliminary Excitation Test of REBCO External Magnetic Field Coil in Liquid Nitrogen Masayoshi Ohya1(B) , Shohri Ikuta1 , Shoichiro Murata1 , Yugo Yamakawa1 , Shinsaku Imagawa2 , Akifumi Iwamoto2 , and Yasuyuki Shirai3 1 School of Engineering, Kwansei Gakuin University, Hyogo, Japan

[email protected]

2 National Institute for Fusion Science, Gifu, Japan 3 Graduate School of Energy Science, Kyoto University, Kyoto, Japan

Abstract. High-temperature superconducting (HTS) systems that harness the cold energy of liquid hydrogen have the potential to reduce CO2 emissions. We have begun research and development of liquid-hydrogen-cooled HTS coils and are planning to verify the stability of REBCO coils immersed in liquid hydrogen in the region above the critical current to obtain a cooling design guideline. A split-type field coil with a magnetic field applied to the test coil was designed and manufactured. The field coil comprised eight stacked single-pancake coils made of 4 mm wide REBCO wires. The inner diameter of the single pancake coil was 130 mm, outer diameter was 206 mm, and number of turns was 220–230. When conducting stability limit tests, the upper and lower coils are energized with the current in opposite directions. When conducting fault current tests, a test coil with a few turns is short-circuited and the field coil current should flow in the same direction. This current transformation method enables overcurrent tests several to several dozen times the critical current. The manufactured field coil was energized in liquid nitrogen. The measured and simulated I-V waveforms are in good agreement, and it is concluded that there are no problems in the design and fabrication of the field coil. We will conduct a liquid hydrogen test after a preliminary energization test of the field coil in liquid helium. Keywords: High temperature superconductors · Liquid hydrogen · Stability

1 Introduction Japan has set a goal of achieving carbon neutrality by 2050, and the use of hydrogen energy is key. Liquid hydrogen is considered an energy carrier, but the cost of liquefaction is an economic issue. The effective utilization of the cold heat of liquid hydrogen is required. We are currently developing high-temperature superconducting (HTS) equipment using liquid hydrogen as a coolant. For example, a zero-emission power generation system can be realized by cooling the field coils of a superconducting generator with liquid hydrogen and sending evaporated hydrogen gas to a hydrogen gas turbine to generate electricity [1]. Because no refrigerator is required, the cooling cost, which is the greatest hurdle in the practical application of superconducting equipment, can be reduced to almost zero. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1126–1132, 2023. https://doi.org/10.1007/978-981-99-6128-3_146

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Owing to its flammability, liquid hydrogen has not been widely studied as a coolant for superconducting equipment [1, 2]. An experimental facility was established at the Noshiro Rocket Testing Center of the Japan Aerospace Exploration Agency (JAXA) to study the coolant performance of liquid hydrogen under a wide range of temperature and pressure conditions over 14 years [3, 4, 6]. Furthermore, the safety issue of energizing a sample in liquid hydrogen has been resolved, and energization tests of HTS wires in liquid hydrogen have been conducted [5, 6]. Currently, we have begun research and development of liquid hydrogen-cooled REBCO coils and are studying coil design methods that effectively utilize the heat transfer characteristics of liquid hydrogen. The technical keyis, considering the steady-state heat transfer characteristics of liquid hydrogen, to enable coil design with a high load factor. A test method for coils immersed in liquid hydrogen was investigated, andthe design and manufacturing of a REBCO field coil, which is used to apply an external magnetic field to the test coil, were conducted.

2 Test Plan and Field Coil Design 2.1 Test Method In REBCO coils, the occurrence of hot spots leads to burn-out. This is due to the high current density and low thermal conductivity in the longitudinal direction of the wire. Because the generated voltage is small, the occurrence of a hot spot cannot be detected from the end-to-end voltage of the coil, resulting in a delay in the interruption of a transport current. The coil must withstand a certain amount of voltage to detect hot spots without burning. Because the critical heat flux of liquid hydrogen is approximately 10 times higher than that of liquid helium, it may be possible to establish a coil design in which heat generation and cooling are balanced in the nuclear boiling region, and thermal runaway does not occur before a hotspot is detected. We plan to verify the stability limit of various coils using REBCO wires with different copper stabilizer thicknesses in the region above the critical current in liquid hydrogen to obtain a cooling design guideline for liquid-hydrogen-cooled REBCO coils. However, the critical current of the REBCO wire at 20 K is high. For example, as shown in Fig. 1, the critical current of Fujikura’s FESC-SCH04 wire (4 mmw , artificial pins, Ic = 120 A@77 K, s.f.) exceeds 500 A at 20 K and 5 T//cc-axis [7]. The capacity of the DC power supply at the Noshiro Rocket Testing Center is 500 A. A field coil that applies a c-axis parallel magnetic field to the test coil is necessary for testing in this range. Therefore, a split-type field coil, shown in green in Fig. 2, was designed. The test coil is shown in red. The field coil comprised eight stacked single pancake coils made of 4 mm wide REBCO wires, which were divided into upper and lower sections across the axial center test space. The inner diameter of the single pancake coil was 130 mm, outer diameter was 206 mm, coil height was 4.6 mm, and number of turns was 220. The gap between each single pancake coil was 0.5 mm, the gap between the double pancake coils was 3.2 mm, and the center space was 15.7 mm. When conducting stability limit tests, the upper and lower coils were energized with the current in opposite directions to generate a magnetic field perpendicular to the wire surface of the test coil. The load line of this field coil is shown in a red line in Fig. 1, and the coil generates a magnetic field of 4.2 T when 300 A is supplied. By winding the test coil using a 2 mm wide wire or 4 mm wide

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Fig. 1. Load line of field coil and Ic-B//c characteristics of REBCO wires.

Fig. 2. Configuration and cross-sectional view of field coil and test coil.

wire without artificial pins (FYSC-SCH04, Ic = 120 A@77 K, s.f.) [7], stability limit tests can be conducted within the DC power supply capacity. 2.2 Mechanical Design of Field Coil A large axial counter-force is generated in the field coil because current flows in opposite directions in the upper and lower coils, as shown in Fig. 2. Simply stacking resinimpregnated single pancake coils may cause deterioration due to electromagnetic forces. We decided to insert each double pancake into a stainless (SUS) steel case to ensure strength. Mechanical analysis was performed by changing the SUS plate thickness and other parameters, and the axial electromagnetic force generated in the coil was suppressed to less than 5 MPa. The final coil reinforcement structure and the axial stress

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distribution at the coil radial center (r = 70 mm) are shown in Fig. 3. The thermal stress simulation result during cooling are also shown in the same figure, and the thermal stress generated in the coil is small because the coil and case are fixed with paraffin.

Fig. 3. Reinforcement structure of field coil and mechanical simulation results (axial stress).

3 Manufacture and Preliminary Test of REBCO Field Coil The specifications of the REBCO wire used in the field coils are listed in Table 1. The thickness of the Hastelloy substrate was 50 µm, the superconducting layer was EuBCO with BHO artificial pins, and the copper-plating thickness was 20 µm. Eight singlepancake coils were fabricated using this wire. The inner and outer diameters of the coils were 130 and 206 mm, respectively, and the number of turns was 220–230 turns. After vacuum impregnating the coils with epoxy resin, the inner circumference of two single

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pancake coils was connected with copper braided strands and inserted into a SUS coil case to withstand the large axial repulsive force generated when energized in the reverse direction. Four double-pancake coils were stacked and fixed using bolts (Table 2). Table 1. Specifications of REBCO wires (Fujikura FESC-SCH04) Material

Thickness(µm)

Substrate

Hastelloy

50.0

Superconductor

EuBCO-BHO

2.6

Protection

Ag

5.2

Stabilization

Cu

20.0

Insulation

Polyimide

27.5

Table 2. Specifications of REBCO field coil Inner diameter (mm)

130

Outer diameter (mm)

206

Turn number

220 ~ 230

3.1 Numerical Simulation Before conducting the energization tests of the field coil, the I-V characteristics of the coil at 77 K were simulated. Assuming a circular circuit divided into 20 sections in the width direction of each wire, the magnetic flux density matrix at each section was obtained by calculating and summing the magnetic fields generated by each circular circuit using the Biot-Savart formula. This matrix was incorporated into the MATLAB calculation code and the I-V waveforms were simulated by calculating the current and magnetic field distribution inside the wire when the current was increased quasi-statically in 0.1A increments. The code incorporates a fitting equation that can describe the Ic-B-θ characteristics of the Fujikura’s FESC-SCH04 wire at 77 K [8]. The Ic of the wire was 125A and the n value was 20. Figure 4 shows the results of I-V waveform analysis. Only the voltage of SP-D rises and the generated voltage is 7.72 mV when energized at 36.6 A. At this time, a voltage of 0.053 mV is generated at the innermost circumference of SP-D, corresponding to approximately 1 µV/cm. Based on the above results, 36.6 A was set as the expected critical current value of the coil. 3.2 Preliminary Energization Tests Figure 5(a) shows the preliminary energization test (sweep test) results for the field coil in liquid nitrogen. The current sweep rate was 0.1 A/s. Th voltage generation owing to

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Fig. 4. Simulated I-V characteristics of field coil and generated voltage distribution in SP-D @36.6 A.

the normal conductive connection in the inner circumference and the screening current was observed in all coils, but it was confirmed that the coil could be energized stably up to 37 A. The voltage rise of DP-3 was observed from approximately 33 A, and it was in good agreement with the simulation result shown in Fig. 4.. To remove the effect of the screening current, the voltage was measured after a sufficient time had elapsed after holding the transport current every 5 A. Figure 5 shows a comparison of the hold test results and the simulation results. The voltage generated in the hold test was sufficiently small, and it is concluded that the voltage generated in the sweep test is due to the screening current.

Fig. 5. Measured I-V characteristics of field coil, (a) sweep test, (b) hold test.

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Based on the above results, it is concluded that there are no problems in the design and manufacturing of the field coil. We plan to conduct a liquid-hydrogen test at the Noshiro Rocket Testing Center.

4 Conclusions We plan to conduct energization tests on REBCO coils immersed in liquid hydrogen to obtain cooling design guidelines for liquid-hydrogen-cooled REBCO coils. REBCO field coils were designed and manufactured for stability limit tests in the current-carrying region above the critical current. The REBCO-stacked coils were manufactured with eight single pancakes, and the preliminary excitation tests were performed in liquid nitrogen. The transport characteristics were confirmed to be equivalent to those obtained from numerical simulations. We plan to conduct a liquid-hydrogen test at the Noshiro Rocket Testing Center. Acknowledgments. This work is based on results obtained from a project, JPNP20004, subsidized by the New Energy and Industrial Technology Development Organization (NEDO). We would like to thank Editage (www.editage.com) for English language editing.

References 1. Hara, S., et al.: Development of liquid hydrogen cooling system for a rotor of superconducting generator IEEE Trans. Appl. Supercond. 31(5), 5202505 (2014). https://doi.org/10.1109/ TASC.2021.3065814 2. Stautner, W., et al.: AIP Conf. Proc. 1573(82), 82–90 (2014) 3. Haran, K.S., et al.: Supercond. Sci. Technol. 30, 123002 (2017) 4. Shirai, Y., et al.: Cryogenics 50(6–7), 410–416 (2010) 5. Shirai, Y., et al.: Cryogenics 55(6), 295–299 (2011) 6. Shirai, Y., et al.: DNB heat flux on inner side of a vertical pipe in forced flow of liquid hydrogen and liquid nitrogen. Cryogenics 92, 105–117 (2018). https://doi.org/10.1016/j.cry ogenics.2018.02.002 7. Tatsumoto, H., et al.: J. Phyasics: Conf. Ser. 234, 032056 (2010) 8. Matsumoto, T., et al.: IEEE Trans. Appl. Supercond. 29(5), 4700806 (2019) 9. Daibo, M.: Recent progress of 2G HTS wires and coils at Fujikura. In: IAdvSCMws (2019) 10. Lijima, Y.: Development of homogenous BMO-REBCO coated conductors for industrial production by hot-wall PLD process. In: ASC (2020)

Stress Calculation of 50 kJ High Temperature Superconducting Magnet Energy Storage Using FEM Ankit Anand(B) , Abhay Singh Gour, Tripti Sekhar Datta, and Vutukuru Vasudeva Rao Cryogenic Engineering Centre, Indian Institute of Technology Kharagpur, Kharagpur 721302, India [email protected]

Abstract. Superconducting Magnet Energy Storage(SMES) system is being used in various applications such as instantaneous voltage drop compensation, and dampening low-frequency oscillations in electrical power systems. It stores energy in the form of a magnetic field. The Lorentz force generated due to magnetic field can be very high due to the high current carrying capacity of superconductor. The stresses generated can lead to the degradation of Ic . Hence, during the design stage, it is very important to calculate the maximum stress and strain to select the tape support or support structure. In this paper, stress calculation in a 50 kJ SMES is carried out using load transfer methods with the assumption that the stress generated is within the limit, and does not degrade the operating current. The simulation is done using axi-symmetric geometry imported in COMSOL software. The volumetric force is estimated by the product of vectorial magnetic flux density generated in the magnet with its current density. Thereafter, the force value obtained from the simulation is used in static structural simulation with appropriate boundary conditions to estimate maximum stresses with and without support. The variation in the radial stress for the central coil is also studied for two cases i.e. with and without support. Keywords: FEM

1

· HTS · Magnetic field · SMES · Stress

Introduction

Superconducting Magnetic Energy Storage(SMES) system stores energy in the form of a magnetic field by the persistent flow of DC current in superconducting coil which offers zero DC resistance under superconducting state. SMES is being used for various power applications such as instantaneous voltage drop compensation, dampening low-frequency oscillations, regulating pulsating power supply, etc. The Lorentz force is generated in SMES due to high magnetic field produced by large circulating current. This force results in hoop and radial stress in superconducting coil which leads to the degradation of Ic [1], which in-turn leads to c Zhejiang University Press 2023  L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1133–1139, 2023. https://doi.org/10.1007/978-981-99-6128-3_147

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reduced energy storage capacity and unstable operation causing quenching of superconducting coil. Therefore during the design of SMES coil, reinforcement supporting structures to withstand hoop and radial stresses should be considered for safe and reliable operation. In 1997, Kalsi et al. [2] designed the first 5 kJ HTS SMES without support due to very low hoop stress. A 600 kJ and 800 kJ SMES with a support layer was designed by Park et al. [3] and Tixador et.al. [4] respectively. However, it does not discuss the effect of supporting layers on magnetic field density, operating current, tape length, and other related parameters. These [5–13] solenoid HTS SMES projects reported their electromagnetic design, but did not discuss their magnet coil stress. In general, external wrapping by copper or stainless steel tape of sufficient thickness is provided to reduce the stress in the coil. However, the comparison between supported and unsupported coil is not yet properly reported. The impact of support structure on other parameters such as operating current, and HTS tape length is also not properly presented. In this paper, a Finite Element Method (FEM) of stress calculation is described. Along with this, the effect of supporting layer on length, inductance, and the operating current is also discussed. Section 2 lists out the dimensional details used for analysis in this paper. Section 3 details the methodology adopted to obtain the dimension of a typical 50 kJ SMES with and without support using parametric sweep. It also discusses the methodology for stress estimation using FEM. Results and discussion are presented in section 4 and Sect. 5 concludes this paper.

Fig. 1. Constructional details of unsupported coil

Fig. 2. Constructional details of supported coil

Stress Calc. of HTS SMES using FEM

2

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Dimensions of 50 kJ SMES

The width and thickness including kapton insulation of 2G HTS tape (model No: SuNAM SCNk04200) used here are 4 and 0.17 mm respectively. Figure 1 and Fig. 2 show a 3D picture of assembled SMES coil having 3 Double Pancake (DP) coils for unsupported and supported coil respectively. The distance between coils are kept as 4 mm to provide cooling plates. Axi-symmetric geometry is used for faster numerical calculation. The difference between supported coil and unsupported coil comes from the additional copper tape having a thickness double the 2G HTS thickness co-wound with HTS tape. Figure 1(b) and 2(c) show the winding scheme of unsupported and supported coils. Copper tape present in the supported coil can also assist in cooling apart from structural support.

3

Methodology for Dimension Determination and Stress Calculation

To determine the dimension of the 50 kJ stored energy in superconductor magnet coil, a parametric sweep was performed,in which Inner Diameter(ID) and the distance between coils is fixed for comparison. The operating temperature of SMES coil was taken as 4.2 K. An FEM-based numerical solver COMSOL along with MATLAB is used for this parametric sweep. The input parameters (ID, number of coils(Nz ), and number of turns(Nr ))was provided to COMSOL from MATLAB, and the result from the simulation is fed back to MATLAB. The complete procedure is as follows: (a) A set of ID,N r,Nz is passed into COMSOL from MATLAB. A 2D axissymmetric model geometry is created using these parameters for faster calculation. Materiel properties and boundary conditions are assigned to this model. (b) An excitation of 1 A DC is given to the model. Mesh is created for this model. Model is solved by COMSOL and B⊥ is calculated. (c) Load line slope (m) can be calculated by dividing this set current I by B⊥ . This load was plotted on B⊥ -Ic characteristic curve as shown in the Fig. 3. The B⊥ -Ic curve at 4.2 K for HTS tape is taken from [14]. The intersection point of load line and B⊥ -Ic curve will give the maximum operating current(Iopmax ) for this coil. The intersection point can be obtained by solving both the equations simultaneously. This step was performed in MATLAB. The obtained Iopmax is multiplied with 0.9 to get operating current (Iop ) for safety reasons. (d) The Iop is given to comsol for evaluation of stored energy, inductance and B⊥ (e) All the evaluated parameters are transferred back to MATLAB for saving. Figure 4 shows the flowchart of the parametric sweep. After the completion of parametric sweep, two dimensions were chosen, one with support and the other

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Fig. 3. B-Ic characteristics at 4.2 K used for design.eps

Fig. 4. Energy and stress calculation flowchart

without supports. The chosen dimension and related parameters are given in Table 1. Both the coils have approximately 50 kJ stored energy. The ID and number of coils selected are same for comparison. The Outer Diameter(OD) comes out to be 285 and 369 mm for unsupported and supported coil respectively The operating current for unsupported and supported coil comes out to be 393 A and 512.88 A respectively. A quick calculation of effective current density is done using Eq. 1. Nr × Iop (1) Je = 0.5 × (OD − ID) × w Here, w is the height of single coil. Equation 1 gives effective current density values 578 and 285 MA/m2 for unsupported and supported coils respectively. The effective current density of supported layer has been reduced to half of unsupported coil. The significant reduction of current density has resulted in reduced central magnetic flux density and B⊥ . This also confirms the increase in operating current, however both the operating current and B⊥ are related to each other by B⊥ −Ic characteristic. Figure 5 and 6 (colored arrow) show the magnetic flux density distribution for unsupported and supported coils respectively.

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Table 1. Input parameters for stress calculation obtained from parametric sweep Parameters

Without support With support

Inner Diameter[mm]

200

200

Number of turns per pancake coil

250

188

Outer diameter[mm]

285

369

Thickness of 2G HTS tape including insulation[mm] 0.17

0.17

Copper support layer thickness[mm]

0.28

0

Number of pancake coils

6

6

Distance between pancake coils[mm]

4

4

B perpendicular max[T]

4.65

2.96

Central magnetic field[T]

3.024

2.5916

Inductance[H]

0.65234

0.38363

Operating current[A]

393.2

512.88

Energy stored[J]

50430

50456

Total HTS tape length[m]

1142.2

1007

Once the distribution of magnetic flux density(B) has been calculated in the whole coil, it is multiplied by current density(J) to calculate the volumetric force in the coil given by Eq. 2. This force will be present in the whole coil volume and is dependent upon local magnetic flux density values. Figure 5 and 6 (black arrows) show volume forces for unsupported and supported coils respectively. The arrow direction shows the direction of volumetric force in each coil. Fv = J × B

Fig. 5. Magnetic flux density distribution for unsupported coil

(2)

Fig. 6. Magnetic flux density distribution for supported coil

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As for boundary conditions for structural simulation, the top and bottom surfaces of each coil have been given roller support. Inner and outer diameters are free to deform. A homogeneous materiel is used for simulation. ∇ · s + Fv = 0

(3)

The model is solved for radial displacement ur . Equation 3 is solved numerically by COMSOL.

4

Results and Discussions

The HTS tape length required for a coil with support(1007 m) is less in comparison with a coil without support (1142.2 m) although the volume of the coil with support is more. The inclusion of a support layer that is double of the layer thickness of HTS tape, reduced the effective current density. This leads to a reduced magnetic flux density. Reduced effective current density and magnetic flux density resulted in reduced volumetric force. This has changed the stress distribution in the coil. Figure 7 and 8 show the variation of hoop and radial stresses in the radially outward direction from ID for both unsupported and supported coils. 0

108

5 4 3

7.5×107

2 5×107 2.5×10

1

7

0

0.02 0.04 0.06 0.08 Radial Distance from inner diameter(m)

Fig. 7. Hoop stress for both coils along radius

5

6

0

Radial stress(Pa)

1.25×10

Hoop Stress-Unsupported Magnetic flux density-unsupported Hoop Stress-Supported Magntic flux density-supported

8

Magnrtic Flux Density(T)

Hoop Stress(Pa)

1.5×108

−5×106 −1×107

−2×107

−2.5×107

Radial stress-Suppported coil Radial Stress-Unupported coil 0

0.02 0.04 0.06 Radial distance from inner diamater(m)

0.08

Fig. 8. Radial stress for both coils along radius

Conclusions

Two different SMES coil winding schemes with unsupported layer and other with support layer are described. The dimensions of supported and unsupported coils for same amount of stored energy is obtained using parametric sweep. For better comparison, inner diameter and number of coils of both are taken same. Magnetic flux distribution is calculated for both coils and used to calculate the Lorentz force in them. Hoop and radial stresses are calculated and plotted. Stress in HTS SMES coil can be reduced by co-winding a copper tape with HTS Tape. A comparison has been also done between unsupported and supported coils in

Stress Calc. of HTS SMES using FEM

1139

terms of stress, operating current, inductance and HTS tape length. There is an increase in operating current for supported coil due to reduction in B⊥ . The HTS tape length requirement is also decreased for supported coil due to increase in operating current, to store same amount of stored energy, although the volume of coil increased.

References 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14.

Zhang, Y., et al.: IEEE Trans. Appl. Superconductivity 26(4) (2016) Kalsi, S.S., et al.: IEEE Trans. Appl. Supercond. 7(2 PART 1), 971–976 (1997) Park, M.J., et al.: volume 17, pp. 1994–1997 (2007) Tixador, P., et al.: IEEE Trans. Appl. Supercond. 17(2), 1967–1972 (2007) Kreutz, R., et al.: IEEE Trans. Appl. Supercond. 13(2 II), 1860–1862 (2003) Hawley, C.J., Gower, S.A.: IEEE Trans. Appl. Supercond. 15(2 PART II), 1899– 1902 (2005) Wojtasiewicz, G., et al.: J. Phys: Conf. Ser. 43(1), 821–824 (2006) Wang, Q., et al.: IEEE Trans. Appl. Supercond. 16(2), 570–573 (2006) Bellin, B., Tixador, P., Deleglise, M., Vallier, J.C., Pavard, S., Bruzek, C.E.: J. Phys: Conf. Ser. 43(1), 817–820 (2006) Kim, W.S., et al.: IEEE Trans. Appl. Supercond. 16(2), 620–623 (2006) Shi, J., et al.: IEEE Trans. Appl. Supercond. 17(3), 3846–3851 (2007) Xiao, L., et al.: IEEE Trans. Appl. Supercond. 18(2), 770–773 (2008) Kim, A.R., et al.: IEEE Trans. Appl. Supercond. 19(3), 2023–2027 (2009) Braccini, V., et al.: Supercond. Sci. Technol. 24(3), 035001 (2010)

Simulation and Analysis of Cryogenic System for FFAG Superconducting Magnets Zhou Hongji(B)

, Fu Wei, Zhang Tianjun, Wang Chuan, Wang Fei, Liu Jingyuan, Zhu Xiaofeng, and Zhang Suping

China Institute of Atomic Energy, Beijing 102413, China [email protected]

Abstract. GeV-class proton beam with an average power of several megawatts have many important applications in particle physics towards the intensity frontier, as well as in the advanced energy, and material science. In 2019, China Institute of Atomic Energy (CIAE) proposed an isochronous FFAG conceptual design with capability of producing 2GeV/6MW CW proton beam. The high temperature superconducting magnets are adopted in the 2GeV/6MW FFAG design and are briefly introduced in this paper. The heat loads of superconducting magnets are analyzed by theoretical analysis method. Then, the centralized refrigeration scheme and distributed refrigeration scheme are analyzed and compared by numerical simulation method. Based on that, suggestions for further engineering design are outlined. Keywords: FFAG · Accelerator · THS · Cryogenic system

1 Introduction GeV-class proton beam with an average power of several megawatts have many important applications in particle physics towards the intensity frontier, as well as in the advanced energy, and material science [1]. By utilizing the strong focusing and large acceptance features of FFAG (Fixed-Field Alternating Gradient) in the theoretical framework of the fixed field and fixed frequency of isochronous cyclotron, a continue wave (CW) FFAG capable of producing 2 GeV/3 mA protons (with a beam power of 6 MW) has been proposed in China Institute of Atomic Energy (CIAE) in 2019 [2]. Figure 1 shows the proposed 2 GeV FFAG circular accelerator complex [2]. The complex consists of a 100 meV pre-injector cyclotron, an 800 meV injector ring cyclotron and a 2 GeV CW FFAG. The magnetic field in the extraction radius of 2 GeV in defocusing magnet and the focusing magnet are ~2.4 T and ~2.7 T respectively, which have already exceeded the saturation field of pure iron (∼2.14 T), thus a superconducting magnet design was used [3]. Due to the beam loss of high-power proton beams, the resulting high radiation will deposit a large amount of radiation dose and heat load on the superconducting magnet. HTS magnet are chosen, as they have better thermal stabilities than LTS magnet. Cooling method of these magnet are discussed in detail in this paper. © Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1140–1147, 2023. https://doi.org/10.1007/978-981-99-6128-3_148

Simulation and Analysis of Cryogenic System

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Fig. 1. 2 GeV High Power Circular Accelerator Complex.

2 The Magnet Design of CYCIAE-FFAG The diameter of the magnet ring is approximately 45 m. The magnet system of the 2GeV/6MW FFAG accelerator consists of ten “F-D-F” periodic magnet cells, each contains two bending (focusing or F) magnets, one reverse bending (defocusing or D) magnet, two smaller drift spaces. Both the D and F magnet are C type magnet. HTS SC coil pair will be installed between iron magnet poles. The basic design parameters of F and D magnets are show in Table 1. The 2nd generation HTS ReBCO superconducting tapes were chosen. The operating current of the magnet is designed at bout [email protected], 30 K, about 50% of the Ic. The large temperature margin of ∼20 K indicates a thermal stable design which is crucial for stable operation of SC magnets in high power accelerators. Table 1. The design parameters of THS magnet Type of magnet

Total weight /t

Focusing magnet

~196

Defocus magnet

~127

Ampere-turns

Magnet field /T

Half air gap /mm

325000

1.56–2.62

31–140

−310000

1.77–2.51

25–110

3 Thermal Load Analysis of High Temperature Superconducting Magnet 3.1 Cryostat of the Magnet Focusing magnet and Defocusing magnet have similar cryostat structure. Figure 2 shows a scheme of the components of cryostat. The HTS coils are cooling to 30K. Thermal shield about 60K surround the HTS coils, and 300K vacuum vessel surround the thermal shield. The HTS coils are fixed by eight vertical supports and four horizontal supports.

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Material of the thermal shield is copper or aluminum. Material of the twelve cold mass support are G10 or CFRP (carbon fiber reinforced polymer). A 60K thermal intercept is fixed to the 300K-30K cold mass supports, to decrease heat conduction heat to 30K. MLI (multilayer insulation) are used between the 60K thermal shield and 300K vacuum vessel. The thermal load of the magnet is divided into two parts: 30K heat load and 60K heat load. These two types thermal load will be cooled by different cold source.

a) The overall view

b) The sectional view

Fig. 2. The scheme of the components of cryostat(1/4model)

3.2 The Heat Load Calculation of 60K The heat load of 60K includes heat load of multilayer insulation, heat load of conduction, heat load of the current lead and heat load of the beam. The heat load of multilayer insulation is calculated by the equation proposed by Ralph Shutt [4]:     ξ q = [1.22 × 10−3 P · T0 0.25 − Ti 0.25 + 3.22 × 10−14 T0 4.25 − Ti 4.25 ] (1) N Q1 = q ∗ A

(2)

where q-W/cm2 . ξ -Safety margin, general engineering experience is 6. N -The film layer of multilayer insulation, the value is 30. P-The vacuum, the unit is micron, the value is 0.007 micron. T0 \Ti - The temperature of hot and cold surfaces, A-cold mass surface. T0 = 300K, Ti = 60K, from formula 1, we can calculate q = 2.19 W/m2 . The cold surface of F magnet is 24.8 m2 . The cold surface of D magnet is 20.9 m2 . We can calculate multi-layer heat flux of the F magnet and the D magnet. The multi-layer heat flux of the F magnet is 54.3 W. The multi-layer heat flux of the D magnet is 45.9 W. The heat load of conduction is calculated by the thermal conductivity of support is calculated by Fourier law of conduction: Q1 = −λA

∂t ∂x

(3)

where lambda-conduction coefficient. The heat flux of conduction consists of two parts: the support and the signal line (Table 2).

Simulation and Analysis of Cryogenic System

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Table 2. Calculating parameter of conduction Parameters

Axial support

Radial support

signal line

The material of support

CFRP

CFRP

Copper

High temperature

300K

300K

300K

Low temperature

60K

60K

60K

Length of support

100 mm

100 mm

300 mm

Equivalent thermal conductivity

3.69 W/(m·K)

3.69 W/(m·K)

400 W/(m·K)

sectional area

80 mm2

80 mm2

0.2 mm2

Number

8

8

80

The heat load of conduction

5.2 W

5.2 W

5.1 W

The operating current of superconducting magnet is 300 A. The copper current leads are used in the temperature range of 300K–60K. And the number of current leads is two. The heat load of the current lead is estimated to be 27.6 W (refer to Lankai Li’paper [5]). The beam loss is estimated at 1 W/m. And the heat load of F magnet beam loss is 1.8 W. The heat load of D magnet beam loss is 1.2 W. We calculated the heat loads of the focusing magnet (F magnet) and the defocusing magnet (D magnet) by theoretical formulas, and the calculation results are shown in Table 4. 3.3 The Heat Load Calculation of 30K The heat load of 30K includes heat load of radiation, heat load of conduction, heat load of the current lead, the heat load of coil and heat load of the beam. The heat load of radiation is calculated by the following formula:  = εAσ (T0 4 − Ti 4 ) ε-system emissivity, ε =

1 1 1 ε0 + εi

−1

(4)

, ε0 , εi the emissivity of hot and cold surfaces.

σ -Stefan-Boltzmann constant. The value is 5.67 × 10−8 W/(m2·K4). The coil emissivity ε 0 is 0.2. The thermal shield emissivity ε i is 0.1. Equivalent emissivity ε = 0.07. We suppose the temperature T0 = 80K and Ti = 30K. From formula 4, we can calculate the heat load of radiation is 4.03 W for F magnet, and the heat load of radiation is 3.40 W for D magnet. The heat load of conduction is calculated as the heat load of 60K. The calculating parameter and result of conduction are show in Table 3. The heat load of the current lead is the heat conduction of HTS current leads and the joule heat of the joint. Let’s suppose the joint resistance wasc1.0E−7 . There are three joint in every THS current lead. We can calculate the joule heat of the current lead is 0.23 W. The heat conduction calculating result of HTS current leads is show in Table 4.

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Z. Hongji et al. Table 3. Calculating parameter of conduction

parameters

Axial support

Radial support

signal line

HTS current leads

The material of support

CFRP

CFRP

Cu

Cu + HTS

High temperature

60K

60K

60K

60K

Low temperature

30K

30K

30K

30K

Length of support

100 mm

90 mm

300 mm

300 mm

Equivalent thermal conductivity

0.47 W/m·K

0.47 W/m·K

22 W/m·K

400 W/m·K

sectional area

80 mm2

80 mm2

0.2 mm2

6

number

8

8

80

2

The heat load of conduction

0.15 W

0.17 W

0.06 W

0.48 W

The heat load of coil is the joule heat of high temperature superconducting joint. We assume that the heat load of the coil is 0.23 W. The beam loss is estimated at 1 W/m. And the heat load of F magnet beam loss is 1.8 W. The heat load of D magnet beam loss is 1.2 W. We calculated the heat loads of the focusing magnet (F magnet) and the defocusing magnet (D magnet) by theoretical formulas, and the calculation results are shown in Table 4. Table 4. The summary of heat load Item

Heat load of 60K

Heat load of 30K

F magnet

F magnet

D magnet

D magnet

Heat load of radiation/W

/

/

4.03

3.4

Heat load of multilayer insulation/W

54.3

45.9

/

/

Heat load of conduction /W

15.5

15.5

0.38

0.38

Heat load of the current lead /W

27.6

27.6

0.71

0.71

Heat load of the beam loss /W

1.8

1.2

1.8

1.2

Heat load of the single magnet/W

99.2

90.2

7.15

5.92

Heat load of the coil/W

/

/

0.23

0.23

Heat load of the magnet unit /W

288.6

20.22

Heat load of the magnet system/W

2886

202.2

Simulation and Analysis of Cryogenic System

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4 Numerical Simulation of Coil Temperature The centralized refrigeration scheme uses a large cryogenic system to centralize refrigeration, and then transfers the cooling capacity to the superconducting coil with highpressure helium gas as the medium. The distributed refrigeration scheme uses two GM cryocoolers to provide cooling for the superconducting coils. The high-pressure helium cooling adopts the pulse tube refrigerator, and the conduction cooling adopts the GM refrigerator. The safety factor is calculated by dividing the rated cooling capacity of the refrigerator by the amount of cooling required by the theoretical calculation. 4.1 Comparison of Centralized Refrigeration and Distributed Refrigeration The centralized refrigeration scheme uses a large cryogenic system to centralize refrigeration, and then transfers the cooling capacity to the superconducting coil with high-pressure helium gas as the medium. Take the Stirling refrigerator on the market as an example, the total power of the refrigeration system is 110 kW, the primary cooling capacity of the cryogenic system safety factor is 1.2, and the secondary is 1.5. The calculation model is 1/2 coil model. The helium flow rate is 4.24 g/s. The inlet temperature is 30K and the helium outlet pressure is 5 bar. The current lead is copper. The temperature at the high temperature end of the current lead is 60K. The heat flow of the coil support structure is 0.19 W. The total heat load at the surface of the cold mass is 2.92 W (in addition to the support structure). The fluid calculation model is k-ε model. The thermal contact resistance between the coil and the support structure is 0.002 m2 K/W. Figure 3(left) show temperature contour of 1/2 coil model. Superconducting coil maximum temperature rise 0.34K for helium circulation cooling.

Fig. 3. Temperature contour of helium circulation cooling(left); temperature contour of distributed refrigeration(right).

The distributed refrigeration scheme uses two GM cryocoolers to provide cooling for the superconducting coils. The total power of the cryogenic system is 450 kW, and the safety factor of the primary cooling capacity is 1.1 and the safety factor of the second is 2.7. The calculation model is also 1/2 coil model. Despite the distributed cool condition, other parameters are the same as the above model. Figure 3(right) show temperature contour of 1/2 coil for distributed refrigeration. Superconducting coil maximum temperature rise 1.65K for GM cryocoolers cooling.

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As can be seen from the simulation results in Fig. 3 the cooling effect of highpressure helium cooling is better than the conduction cooling of GM cryocoolers for the high temperature superconducting coils of this size (the superconducting coil is about 3 m * 1 m). From the view of cooling effect, high-pressure helium is preferred for cooling coils of this size. Due to the high temperature superconducting material has good thermal stability and the maximum temperature rise of conduction cooling by GM cryocoolers is no more than 2K, it is acceptable to use GM cryocoolers to directly cool the superconducting coil of this size. 4.2 The Maximum Temperature Rise of Coils with Different Cooling Methods Since the thermal contact resistance between the HTS tapes is greatly affected by the coil production process, and we have not carried out relevant experimental research at present, we simulated the maximum temperature rise of the coil under different thermal contact resistance conditions under the two cooling modes. The simulation results are shown in Fig. 4. The overall temperature rise of helium refrigeration scheme (concentrated refrigeration) is less than that of GM refrigeration scheme (distributed refrigeration). As can be seen from the Fig. 4, with the increase of thermal contact resistance, the deterioration of cooling effect of the high-pressure helium cooling is less than that of the GM cryocoolers. If the GM refrigerator is used to cool the coil, the contact thermal resistance needs to be better controlled.

Fig. 4. Temperature rise of coil

5 Conclusion This report analyzes the heat loads of the FFAG magnet system and simulates the coil temperature distribution under different cooling schemes. From the perspective of theoretical analysis, the centralized refrigeration scheme is an ideal cooling scheme with high refrigeration efficiency and uniform temperature distribution of the coil. However,

Simulation and Analysis of Cryogenic System

1147

centralized refrigeration has the disadvantages of long transmission pipeline, complex system and poor independence. The ideal solution is to use a small and efficient helium refrigeration system for cooling. Acknowledgments. This work was supported by the National Natural Science Foundation of China (Grant No. 1213000472) and the Scientific Research Program for Young Talent Elite Project of China Institute of Atomic Energy (Grant No. YC202301000814).

References 1. Bian, T., Zhang, T., Li, M., et al.: Possible solution for the integer resonance crossing problem of GeV-class isochronous Fixed-Field Alternating Gradient accelerators. Nuclear Inst. Methods Phys. Re. A 1031, 166594 (2022) 2. Zhang, T., Li, M., Bian, T., et al.: A new solution for cost effective, high average power (2GeV, 6MW) proton accelerator and its R&D activities. In: Proceedings of 22nd International Conference on Cyclotrons and Their Applications, Cape Town, South Africa (2019) 3. Wang, C., Bian, T., Li, M., et al.: Preliminary study of the high temperature superconducting solution for 2GeV CW FFAG magnet. IEEE Trans. Appl. Superconduct. 30(4) (2020) 4. Green, M.A.: Heat transfer through a multilayer insulation system as a function of pressure in the cryostat vacuum space. Adv. Cryogen. Eng. (43), 1313–1318 (1998) 5. Li, L.: Design of cryogenic system and current leads for MICE coupling superconducting magnets, pp. 8–11 (2007)

Author Index

A Ai, Xin 402 Akintola, Ayomikun 508, 1051 Aldiyarov, Abdurakhman 958 Aleksa, Martin 1013 Anand, Ankit 1133 Ankang, Kan 529 Arnold, P. 101, 133, 172, 203 Atrey, M. D. 818 B Badcock, Rodney 1081 Bainbridge, Alexander R. 1051 Bao, Shiran 413 Barba, Maria 1013 Beßler, Y. 133, 203 Bi, Yujing 117 Bo, Wang 561 Bonetti, N. 92 Boros, M. 133 Bott, Stuart 249 Brake, H. J. M. ter 672 Bremer, Johan 286, 1013 Brodzinski, K. 92 Brodzinski, Krzysztof 78 Buckley, Rachael 1051 C Cai, J. H. 763, 787 Cai, Jinghui 755, 771, 795 Cai, Yu-hong 515 Cao, Jianfeng 942 Cao, Qiang 642 Casagrande, Fabio 233, 249 Chai, Yuguo 1095 Chalifour, Michel 1013 Chang, Zhengze 48, 308 Che, Bangxiang 942 Chen, Chaojie 642 Chen, H. L. 780, 787 Chen, Han 889

Chen, Hou lei 755 Chen, Huaiyu 357 Chen, Liubiao 156, 164, 537, 701, 732, 979 Chen, Mengjia 109 Chen, Shixiong 365, 372 Chen, Xuheng 402 Chen, Yu 463 Chen, Yuji 642 Chen, Zhichao 679 Cheng, Xin 889 Cheng, Yue 321 Chi, Chunyun 694, 740, 832 Chi, Chun-Yun 848 Chi, Zhang 561 Chuan, Wang 1140 Claudet, Serge 63 Collier, Gary 1051 Corlett, Peter 1051 Cui, Guanglong 56, 265 Cui, Wenhui 149, 716 Cui, Zheng 686, 862 D Dai, Chao 1031 Dai, Haibin 569 Dai, Wei 747 Datta, Tripti Sekhar 1133 Ding, Lei 583, 664, 679 Ding, Peizhi 402 Dong, Bing 213 Dong, Caiqian 649 Dong, Hongyu 1058 Dong, Jichao 1006 Dong, Xinbo 125, 336 Dong, Xueqiang 194 Duan, Yanbo 840 Dumbell, Keith 508, 1051 Dupont, Thierry 41 E Ellis, Michael

1051

© Zhejiang University Press 2023 L. Qiu et al. (Eds.): ICEC28-ICMC 2022, ATSTC 70, pp. 1149–1155, 2023. https://doi.org/10.1007/978-981-99-6128-3

1150

F Fan, Chengfei 329, 365, 372 Fan, Xiaoyu 156, 164 Fan, Yufeng 914 Fang, Xinqi 301 Fei, Wang 1140 Feng, T. S. 780 Feng, Tianshi 755, 795 Fengshuo, Wen 635 Ferrand, F. 92 Ferrand, Frederic 41 Fila, Adam 233 Fischer, Matthias 294 Fluder, C. 92 Fu, Bao 365, 372, 544 Fu, Yang 942 Fu, Zhengdong 929 Fydrych, J. 172 G Gahier, Vanessa 63, 78 Gan, Zhihua 271, 279, 301, 642, 656 Ganni, Venkatarao 233, 249 Gao, M. 780 Gao, Zhaozhao 701, 732 Gaoqiao, Luo 561 Garceau, Nathaniel 413 Ge, Jian 420 Ge, Rui 48, 308 Geng, Zongtao 686 Ghosh, Parthasarathi 881 Gistau Baguer, Guy 3 Gong, Linghui 85, 149, 179, 187, 716 Gong, Maoqiong 194 Gour, Abhay Singh 1133 Guillotin, Nicolas 41 Guo, Jia 537 Guo, Luna 156, 164 Guo, Wei 413 Guo, Y. J. 870 Guo, Zhimin 621 H Haifeng, Zhu 635 Han, Haiying 942 Han, Ruixiong 48, 308 Han, Shaofei 321 Han, Xian-hu 515 Han, Yemao 1065

Author Index

Han, Yinan 613 Hancock, Mark 1051 Hao, Qiangwang 1031, 1038 Hasan, Nusair 233, 249 Hathaway, Jane 1051 He, Ming 149, 716 He, Shen 125 He, Sheng 336 He, Shuisheng 508 He, Yuanxin 271 Herblin, L. 92 Herrmann, Robert 63 Hitchen, Sean 508, 1051 Hodgkinson, Carl 1051 Holland, H. J. 672 Hong, Guotong 694, 740, 832, 1072 Hong, Guo-Tong 848 Hongji, Zhou 1140 Hornickel, Philip 1051 Horvath, A. 133 Hu, Liangbin 70 Hu, Liangbing 56, 125, 265, 336, 1006 Hu, Libiao 1031, 1038 Huang, Chuanjun 501, 855, 1058, 1095 Huang, Jianhong 420 Huang, Jinyin 929 Huang, L. 810, 870 Huang, Li 708, 724 Huang, Qi 664 Huang, Rongjin 394, 501, 972, 1089, 1095, 1101 Huang, T. H. 810, 870 Huang, Tai He 708 Huang, Taihe 724 Huang, Yawei 256 Huang, Yonghua 427, 889, 907, 921, 965 Huang, Z. J. 763, 787 Huang, Zehua 724 Huang, Zhijie 771, 795 Hughes, Gary 1051 Hui, Hejun 583, 597, 628

I Ikuta, Shohri 1126 Imagawa, Shinsaku 1126 Ismail, Anel 958 Iuga, V. 92 Ivens, B. 92 Iwamoto, Akifumi 1126

Author Index

J Jenkins, Conor 508, 1051 Ji, Wei 156, 164 Jia, Huan 889 Jia, Peng 394 Jia, Qiming 716 Jiali, Wang 529 Jiang, Lifeng 929 Jiang, Mingyue 1089, 1095 Jiang, Qingfeng 544 Jiang, Wenbing 427 Jiang, Yongcheng 308 Jiang, Zhenhua 449, 456, 576, 583, 597, 628, 664, 679 Jiang, ZhenHua 649 Jiang, Zhiyi 472 Jiantang, Song 628, 635 Jiao, Kexin 694, 740, 832 Jiao, Ke-Xin 848 Jin, Tao 420 Jin, Yu 942 Jing, Liwei 1119 Jingyuan, Liu 1140 Jones, Geraint 1051 Jones, Shelly 233 Joseph, Nathan 233 Ju, Yonglin 117, 463, 479 Junker, S. 92 K Ke, Changlei 213 Kickulies, M. 133 Knauss, Eric 357 Knudsen, Peter 249 Koettig, Torsten 286 Kolev, N. 101 Kossel, Logan 442 Kuang, Dazhi 402 Kuizhang, Zhu 561 Kulmer, Karl-Heinz 294 Kunniyoor, Keerthi Raj 881 L Lai, Bihui 56, 125, 265, 336 Laumer, Brandon 233, 249 Lei, Gang 899, 951 Lei, Zhuoqun 463 Li, Chaolong 951 Li, Chunyu 435, 907 LI, Haining 1006 Li, Huaibing 544

1151

Li, Jian 935 Li, Jianguo 986, 1072 Li, Jing 179 Li, Junjie 402, 1044 Li, Ke 747 Li, Kongrong 213 Li, Kunyin 187 Li, Laifeng 314, 394, 487, 494, 501, 840, 855, 972, 1058, 1065, 1089, 1095, 1101 Li, Mei 48, 308 Li, Nanxi 449, 456 Li, Peng 642 Li, Penghui 308 Li, Ruijie 694, 832 Li, Rui-Jie 848 Li, Shanshan 343, 365, 372 Li, Shaopeng 48, 308 Li, Shu 479 Li, X. Y. 810 Li, Xiao Yong 708 Li, Xiao-jin 515 Li, Xiao-xia 515 Li, Xiaoyong 724 Li, Yanzhong 522, 935, 951 Li, Yongyan 456 Li, Zheng-qing 515 Li, Zhengyu 716 Li, Zhenyu 942 Li, Zicheng 664 Li, Ziwei 349 Liang, J. T. 780, 810 Liang, Jingtao 986 Liang, M. L. 780 Lin, Mingqiang 694, 740, 832 Lin, Ming-Qiang 848 Lin, Yilin 899 Lin, Yujie 279 Lin, Yuzhe 109 Liu, Bo 508 Liu, Dongli 279, 656 Liu, Huiming 394, 487, 855 Liu, Jianglai 463 Liu, Jiyun 156, 164 Liu, Lei 656 Liu, Liqiang 179, 213 Liu, Meiyan 1095 Liu, Ming 551 Liu, Ping 747 Liu, Shaoshuai 576, 583, 597, 628, 649, 664, 679 Liu, Sihao 357

1152

Liu, Sixue 929 Liu, Sumei 420 Liu, W. J. 870 Liu, Xuming 732 Liu, Y. J. 763 Liu, Y. L. 763, 787 Liu, Yanjie 771, 986 Liu, Yiyong 256 Liu, Yuan li 755 Liu, Yuanli 771, 795 Liu, Yuefeng 256 Liu, Ziyao 986 Liubiao, Chen 825 Lowe, Michael 1051 Ludbrook, Bart 1081 Lumsden, Grant 1081 Lv, Bingkun 840 Lv, Cui 149, 349 Lyngh, D. 133, 203 Lyu, Baoxi 314 Lyu, Bingkun 494, 501 M Ma, Changcheng 308 Ma, Min 515 Ma, Yuan 935 Ma, Yuexue 755, 986, 1072 Maekawa, Kazuma 1108 Mason, David 1051 May, Andrew J. 240, 508, 1051 McIntosh, Peter 1051 Mei, Enming 321 Meng, Chuiju 427, 921 Meng, Fankong 914 Meng, Qiumin 402 Miao, Jianyin 914, 929 Miao, Zhicong 1058, 1089, 1095, 1101 Middleman, Keith 1051 Miller, Franklin 442, 472 Miller, George 1051 Moldenhauer, Stefan 294 Monneret, Emmanuel 63 Moss, Andrew 1051 Mou, Jian 694, 740, 832, 848 Murata, Shoichiro 1126 Mutch, Jennifer 1051 N Naydenov, Boyan 286 Nguyen, Chinh 233

Author Index

Ni, Dongsheng 33, 321 Niu, Y. F. 763 Niu, Yuefeng 771 O Oates, Adrian 1051 Ohya, Masayoshi 1126 Onufrena, Aleksandra 286 Ouyang, Zhengrong 109, 321, 402, 1044 P Pan, Chongyao 544 Pan, Wei 33, 85 Pan, Xiaoshan 679 Parmar, Darshit 818 Pattalwar, Shrikant 508, 1051 Pei, Haiyue 279 Pendleton, Mark D. 1051 Pendleton, Mark 508 Peng, Nan 213, 551 Perin, A. 92 Perin, Antonio 41 Pezzetti, M. 92 Pfotenhauer, John M. 621 Pfotenhauer, John 442, 472 Ping, Zhu 343 Pirotte, O. 92 Puchleitner, Rainer 294 Putselyk, Sergiy 222, 294, 997 Q Qi, Haoying 1065, 1101 Qiang, Li 379 Qiang, Zhou 825 Qin, Hailing 179 Qin, Xiangqi 321 Qin, Xujin 427, 921 Qiu, Changxu 656 Qiu, Qingquan 1119 Qiu, Shun 213 Quan, J. 810 Quan, Jia 986 R Radebaugh, Ray 16 Rao, Vutukuru Vasudeva Ren, Feng 899 Rosenthal, E. 133 Rueegge, A. 101

1133

Author Index

S Salleger, Daniel 294 Sang, Minjing 308 Sato, Sara 1108 Segerup, M. 133, 203 Sha, Xinquan 649, 664 Shan, Xinran 1058 Shang, Jin 349, 357 Shanshan, Li 802 Shaoshuai, Liu 635 Shen, Fuzhi 855, 1065 Shen, Guoqiang 187 Shen, Jun 194, 747 Shen, Xian 271 Shen, Yunwei 656 Sheng, Lina 321 Shi, Chenfangda 899 Shi, Lei 402 Shi, Xu 56, 70, 265, 1006 Shi, Yaran 394, 494, 501 Shi, Yi 1031 Shirai, Yasuyuki 1126 Shuai, Wang 561 Singamneni, Sarat 1081 Sixue, Liu 825 Smith, Paul A. 508, 1051 Song, Jiantang 583, 597 Song, Naihao 1119 Song, Qinglu 387, 605 Song, Yuntao 420 Su, Haojian 1065, 1089, 1101 Su, He 349, 357 Sujun, Yang 825 Sun, C. 870 Sun, Dandan 387, 605 Sun, J. 810 Sun, Jian 724 Sun, Jiuce 109 Sun, Liangrui 308 Sun, Xingzhong 265 Sun, Zheng 56, 70, 125, 1006 Suping, Zhang 1140 T Takeda, Minoru 1108 Tang, Qingjun 755, 795 Tatsumoto, H. 133, 203 Teng, Yuping 1119 ter Brake, Marcel 286 Tereszkowski, P. 133, 203

1153

Tianjun, Zhang 1140 Ting, Liu 561 Tirolien, Thierry 286 Tolboom, A. H. 672 Tong, Yelong 929 Tychengulova, Aliya 958 V Vermeer, C. H.

672

W Wagner, Udo 78 Wan, Shiqing 544 Wang, Bo 271, 279, 301, 642, 914 Wang, Dechang 387, 605 Wang, Guopeng 1072 Wang, Haocheng 194, 387, 605 Wang, Haoren 271 Wang, Haosu 349 Wang, Huizhi 929 Wang, Jiasen 314 Wang, Jinzhen 349, 357 Wang, Juan 986 Wang, Jue 840, 855 Wang, Junjie 156, 164, 537, 701, 732, 979 Wang, Lei 951 Wang, Liguo 972 Wang, Ling 708 Wang, Lishi 321 Wang, Miaomiao 642 Wang, Nailiang 755 Wang, Pengcheng 642 Wang, Shiyue 456 Wang, Shuai 914 Wang, Simin 522 Wang, Tao 494 Wang, Wei 501, 840 Wang, Xiao-jun 515 Wang, Xilong 56, 70, 125, 265, 336, 1006 Wang, Xiuli 463 Wang, Xudong 321 Wang, Yan 942 Wang, Yanan 747 Wang, Ye 899 Wang, Yubo 301 Wang, Yun 724 Wei, Fu 1140 Wei, Jianjian 420 Wei, Lingjiao 1072

1154

Wei, Yaxue 387 Weijers, Huub 1081 Weisend II, J. G. 172 Weiwei, Wu 561 Wen, Jian 522, 951 Wenting, Wu 635 Wheelhouse, Alan E. 1051 White, Alastair A. J. 1051 White, Alastair 508 Wilde, Stuart 1051 Wilson, James 1051 Wright, Mathew 233 Wu, Beimin 321 Wu, Chen 529 Wu, Humin 271 Wu, Jiefeng 372 Wu, Jihao 149, 179, 349, 357 Wu, Jingyi 435, 889, 899, 907 Wu, Jinxing 387, 605 Wu, Kaihong 1031, 1038 Wu, Keping 329 Wu, Shanshan 855, 1101 Wu, Shixian 402 Wu, Wei 33, 187, 321 Wu, Wenting 597 Wu, Xianlin 942 Wu, Yinong 449, 456, 576, 583, 597, 628 Wu, YiNong 649 Wu, Yinong 664, 679 Wu, Yu 1031, 1038 Wu, Zhixiong 494, 1058 X Xi, Chen 802 Xi, Xiaotong 701 Xia, Jiaxu 279 Xiang, Hanzhen 583 Xianlin, Wu 825 Xiantong, Chen 802 Xiao, Liye 1119 Xiao, Meng 914 Xiao, Mingkun 889, 907 Xiao, Tang 635 Xiao, Tiancheng 642 Xiao, Tiantian 487, 855 Xiaofeng, Zhu 1140 Xiaolei, Yao 561 Xie, Changheng 1020 Xie, Jing 613 Xie, Ningning 187

Author Index

Xie, Shiyong 314, 394 Xie, Xiujuan 33, 85, 187, 321, 716 Xin, Jijun 1058 Xing, E. C. 763, 780 Xing, Enchun 755, 771, 795 Xinghua, Tian 561 Xiong, Lianyou 179 Xiong, Zhenyan 271 Xu, Dong 314, 494, 501, 972 Xu, Hao 537, 979 Xu, Hongbo 551 Xu, Honghao 271 Xu, Lei 56, 70, 1006 Xu, Miaofu 48, 308 Xu, Pan 522 Xu, Ran 889 Xu, Sheng 109 Xu, T. 870 Xu, Xiafan 537, 979 Xu, Xiangdong 85 Xu, Xiaowei 321 Xu, Yawei 914 Xue, Rui 33, 85, 187 Xuehua, Zhang 141 Xun, Y. Q. 780, 787 Y Yamakawa, Yugo 1126 Yan, Han 213 Yan, Jixiang 314, 394, 494 Yang, Biao 701, 732 Yang, Changpeng 914 Yang, Guang 435, 889, 899, 907 Yang, Huning 914 Yang, Jingyao 194 Yang, Kun 179 Yang, Longyu 862 Yang, Mingzhuo 694 Yang, Ming-Zhuo 848 Yang, Pengcheng 329, 365 Yang, Shaoqi 33, 85, 187, 321 Yang, Sheng-sheng 515 Yang, Xiaochen 308 Yao, Qinggao 321 Yao, Zhitao 256 Ye, Rui 308 Yerezhep, Darkhan 958 Yifan, Lin 529 Yin, Wang 597, 649 Yin, Zhang 561

Author Index

Ying, Kongkuai 576 Yinong, Wu 635 Yongqing, Zhang 561 You, Wei 321 Yu, Jin 825 Yu, Qiang 1020 Yu, Wenhui 613 Yuan, Jiao 271 Yuan, Zhou 825 Yuchen, Luo 141, 379 Yufeng, Fan 561 Z Zeng, Yong 724 Zezhang, Wang 141 Zhai, Yanfei 256 Zhang, Ankuo 613 Zhang, C. 763, 787 Zhang, Chen 771, 795 Zhang, Chuanjia 343 Zhang, Guomin 1119 Zhang, Hengcheng 855, 1065 Zhang, Hongxing 929 Zhang, J. 101, 172 Zhang, Jiehao 308 Zhang, Jingye 1119 Zhang, Lei 109 Zhang, Meimei 149, 716 Zhang, Qiyong 329, 343, 365, 372 Zhang, Shuyue 357 Zhang, Siyi 1101 Zhang, Xiangzhen 308 Zhang, Xiaohua 213 Zhang, Y. T. 780 Zhang, Yin 914 Zhang, Zhiming 724 Zhang, Zhuo 308 Zhao, Chenyang 449 Zhao, Liang 929 Zhao, Miguang 795 Zhao, Qinyu 279, 301, 642, 656

1155

Zhao, Tongxian 308 Zhao, Wanyin 1058, 1089 Zhao, Wenyi 456 Zhao, Yalin 1058, 1089 Zhao, Yanxing 194 Zhao, Yuchen 487, 1065 Zheng, Chen 686, 862 Zheng, Hongyang 914 Zheng, Jianpeng 979 Zheng, Jingxin 1044 Zheng, Liangwei 213 Zhengrong, OuYang 141 Zhenhua, Jiang 635 Zhou, Aimin 522 Zhou, Gang 33, 179 Zhou, Han 149, 716 Zhou, Min 1101 Zhou, Qiang 929 Zhou, Weiming 271 Zhou, Xiaoqing 435 Zhou, Yuan 394, 487, 840, 855, 1058, 1065 Zhou, Zhengrong 1058, 1089 Zhou, Zhihao 1119 Zhou, Zhiwei 1020 Zhu, Li 321 Zhu, Minchen 301 Zhu, Q. L. 810 Zhu, Qiang Long 708 Zhu, Qiang 544 Zhu, Qianglong 724 Zhu, Shaowei 109, 569, 590, 621 Zhu, Weiping 716 Zhu, Wenxin 965 Zhu, Zhigang 329, 343 Zhun, Li 141, 379 Zhuqing, Ni 561 Zink, Stefan 294 Zong, Yiwen 329, 365 Zou, Zhongyu 329 Zuo, C. 870 Zuo, Zhongqi 965