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Article(id=1319293494888354291, tenantId=1146029695717560320, journalId=1316781260334530699, issueId=1319293383915454866, articleNumber=null, orderNo=null, doi=10.3969/j.issn.1005-0329.2025.09.004, pmid=null, cstr=null, oa=null, hot=null, price=null, onlineType=0, articleFormat=0, articleType=null, articleTypeStr=null, receivedDate=1732723200000, receivedDateStr=2024-11-28, revisedDate=1749484800000, revisedDateStr=2025-06-10, acceptedDate=null, acceptedDateStr=null, onlineDate=1792473680468, onlineDateStr=2026-10-20, pubDate=1759161600000, pubDateStr=2025-09-30, doiRegisterDate=null, doiRegisterDateStr=null, onlineIssueDate=1792473680468, onlineIssueDateStr=2026-10-20, onlineJustAcceptDate=null, onlineJustAcceptDateStr=null, onlineFirstDate=null, onlineFirstDateStr=null, sourceXml=null, magXml=null, createTime=1792473680468, creator=13701087609, updateTime=1792473680468, updator=13701087609, issue=Issue{id=1319293383915454866, tenantId=1146029695717560320, journalId=1316781260334530699, year='2025', volume='53', issue='9', pageStart='1', pageEnd='118', issueExtLink='null', onlineDate='null', pubDate='1759161600000', pubDateStr='2025-09-30', beforeIssueId=null, nextIssueId=null, price=null, status=1, issueComplete=1, articleOrder=1, issueType=-1, specialIssue=null, createTime=1792473654011, creator='13701087609', updateTime=1792473654011, updator='13701087609', preIssue=null, nextIssue=null, articleTotal=null, ext=null, issueFiles=null, downloadFileDto=null}, startPage=21, endPage=28, ext={EN=ArticleExt(id=1319293496612213237, articleId=1319293494888354291, tenantId=1146029695717560320, journalId=1316781260334530699, language=EN, title=Study on thermodynamic performance of two‑stage dual‑cylinder hydraulically driven piston compressor under varying inlet pressures, columnId=null, journalTitle=Fluid Machinery, columnName=null, runingTitle=null, highlight=null, articleAbstract=

To investigate the thermodynamic performance variation of a two‑stage hydraulically driven piston compressor for hydrogen refueling stations under varying suction pressures，a thermal simulation model of the two‑stage hydraulically driven piston compressor coupling the hydraulic system and gas compression system was constructed，and the model was experimentally validated.The results show that the maximum error between the simulated and experimental thermodynamic performance parameters is 2.7%，indicating good agreement.When the second‑stage discharge pressure remains unchanged and the first‑stage suction pressure decreases from 20.0 MPa to 3.0 MPa，the first‑stage pressure ratio remains essentially constant，while the second‑stage pressure ratio increases；the discharge temperature of the first‑stage cylinder remains around 120.0 ℃；the discharge temperature of the second‑stage cylinder increases from 54.1 ℃ to 224.1 ℃，the first‑stage shaft power decreases from 60.2 kW to 14.9 kW，the second‑stage shaft power increases from 12.7 kW to 35.5 kW，and the mass flow rate decreases from 55.5 kg/h to 7.1 kg/h.The two‑stage dual‑cylinder hydraulically driven unit is suitable for constant suction pressure conditions.When operating under varying suction pressure conditions，issues such as excessively high discharge temperature，unbalanced shaft power，and low motor efficiency occur.To address this problem，an operational strategy of using high speed under low‑pressure conditions and low speed under high‑pressure conditions can be adopted.The study provides a theoretical basis for the optimization of hydraulically driven piston compressor systems.

, authors=Qiang QI, Chen ZHANG, Jingyao WANG, Xiaohan JIA, Shengdong REN, authorsList=Qiang QI, Chen ZHANG, Jingyao WANG, Xiaohan JIA, Shengdong REN, authorCompany=null, correspAuthors=Shengdong REN, authorNote=null, correspAuthorsNote=null, copyrightStatement=null, copyrightOwner=null, extLink=null, articleAbsUrl=null, sourceXml=null, magXml=null, pdfUrl=null, pdf=null, pdfFileSize=null, pdfExtLink=null, richHtmlUrl=null, mobilePdfUrl=null, reviewReport=null, pdfFirstPage=null, abstractGraph=null, abstractGraphContent=null, abstractVideo=null, citation=null, cebUrl=null, magXmlContent=null, mapNumber=null, fund=null), CN=ArticleExt(id=1319293498793251338, articleId=1319293494888354291, tenantId=1146029695717560320, journalId=1316781260334530699, language=CN, title=两级双缸液驱活塞压缩机变进气压力下的热力学性能研究, columnId=1317484287230853622, journalTitle=流体机械, columnName=设计计算, runingTitle=null, highlight=null, articleAbstract=

为了研究加氢站用两级液驱活塞压缩机在变吸气压力下的热力学性能变化规律，构建了液压系统和气体压缩系统耦合的两级液驱活塞压缩机热力仿真模型，并对模型进行了试验验证。结果表明，仿真与试验的热力性能参数最大误差为2.7%，吻合良好。当二级排气压力不变，一级进气压力从20.0 MPa减小到3.0 MPa时，一级压比基本保持不变，二级压比增大；一级气缸的排气温度保持在120.0 ℃左右；二级气缸的排气温度则从54.1 ℃增大至224.1 ℃，一级轴功率从60.2 kW降至14.9 kW，二级轴功率从12.7 kW增至35.5 kW，质量流量从55.5 kg/h减小到7.1 kg/h。两级双缸液驱机组适用于恒定吸气压力工况，在变吸气压力工况下运行时，会出现排气温度过高，轴功率不平衡，电机效率低的问题。针对该问题，可采取低压工况下采用高转速，在高压工况下采用低转速的运行策略。研究可为液驱活塞压缩机系统优化提供理论基础。

, authors=齐强, 张辰, 王静瑶, 贾晓晗, 任省栋, authorsList=齐强, 张辰, 王静瑶, 贾晓晗, 任省栋, authorCompany=null, correspAuthors=任省栋, authorNote=

齐强（1998），男，博士，主要从事液驱活塞压缩机研究，E-mail：。

, correspAuthorsNote=
任省栋（1997），男，博士，主要从事氢气压缩机与加氢站研究，通信地址：710049陕西省西安市咸宁西路28号西安交通大学压缩机研究所，E-mail：。
, copyrightStatement=作者向本刊所投稿件一经本刊录用，则该论文的复制权、发行权、信息网络传播权、翻译权、汇编权等权利视为自动转让给本刊。本刊为国内各主要数据库收录期刊，本刊所发表的稿件将同时通过互联网进行网络出版或提供信息服务。, copyrightOwner=null, extLink=null, articleAbsUrl=null, sourceXml=x+iGgwW3Xy1Kutc1CTPm4g==, magXml=lTjwhbja0+RhemG2CiA7wg==, pdfUrl=null, pdf=null, pdfFileSize=null, pdfExtLink=null, richHtmlUrl=null, mobilePdfUrl=null, reviewReport=null, pdfFirstPage=null, abstractGraph=kqyVzlH+3alCTM6fJIx11Q==, abstractGraphContent=null, abstractVideo=null, citation=null, cebUrl=null, magXmlContent=s6zau/MjmNUrJGdzzd+RWw==, mapNumber=null, fund=null)}, authors=[Author(id=1319293499132989967, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, orderNo=0, firstName=null, middleName=null, lastName=null, nameCn=null, orcid=null, stid=null, country=null, authorPic=null, dead=0, email=Qiqiang@stu.xjtu.edu.cn, emailSecond=null, emailThird=null, correspondingAuthor=0, authorType=1, ext={EN=AuthorExt(id=1319293499225264657, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, authorId=1319293499132989967, language=EN, stringName=Qiang QI, firstName=Qiang, middleName=null, lastName=QI, prefix=null, suffix=null, authorComment=null, nameInitials=null, affiliation=null, department=null, xref=null, address=School of Energy and Power Engineering，Xi'an Jiaotong University，Xi'an 710049，China, bio=null, bioImg=null, bioContent=null, aboutCorrespAuthor=null), CN=AuthorExt(id=1319293499309150738, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, authorId=1319293499132989967, language=CN, stringName=齐强, firstName=null, middleName=null, lastName=null, prefix=null, suffix=null, authorComment=null, nameInitials=null, affiliation=null, department=null, xref=null, address=西安交通大学 能源与动力工程学院，西安 710049, bio={"content":"

齐强（1998），男，博士，主要从事液驱活塞压缩机研究，E-mail：。

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齐强（1998），男，博士，主要从事液驱活塞压缩机研究，E-mail：。

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language=EN, label=Fig.1, caption=Schematic diagram of the main structure of hydraulically driven piston compressor, figureFileSmall=YAcuqWesoDCnG/TWs2TYRg==, figureFileBig=kqyVzlH+3alCTM6fJIx11Q==, tableContent=null), ArticleFig(id=1319293502626845232, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=CN, label=图1, caption=液驱活塞压缩机主机结构示意, figureFileSmall=YAcuqWesoDCnG/TWs2TYRg==, figureFileBig=kqyVzlH+3alCTM6fJIx11Q==, tableContent=null), ArticleFig(id=1319293502874309169, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=EN, label=Fig.2, caption=Thermodynamic simulation model of the two‑stage hydraulically driven piston compressor system, figureFileSmall=8obnvKBleOW6qPvdQTgbeQ==, figureFileBig=5Cxnn5CMv5mSmIPUI3TCyg==, tableContent=null), ArticleFig(id=1319293502933029426, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=CN, label=图2, caption=两级液驱活塞压缩机系统热力仿真模型

1-气源；2-单向阀；3-压缩腔；4-换热器；5-信号源；6-排气口；7-级间气罐；8-背压阀；9-质量流量计；10-冷水源；11-水泵；12-水箱；13-电机；14-气动室；15-液压油缸；16-三位四通阀；17-功率仪；18-双向油泵；19-油箱；20-溢油阀；21-单向油泵。

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Thermodynamic simulation model parameter settings

, figureFileSmall=null, figureFileBig=null, tableContent=
参数数值参数数值
主油泵排量/（L⋅h-1）15 660气管直径/mm11
补油泵排量/（L⋅h-1）3 480水管直径/mm25
进气压力/MPa10冷却水温度/℃10
排气压力/MPa65进气温度/℃25
), ArticleFig(id=1319293505973899840, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=CN, label=表1, caption=

热力仿真模型参数设置

, figureFileSmall=null, figureFileBig=null, tableContent=
参数数值参数数值
主油泵排量/（L⋅h-1）15 660气管直径/mm11
补油泵排量/（L⋅h-1）3 480水管直径/mm25
进气压力/MPa10冷却水温度/℃10
排气压力/MPa65进气温度/℃25
), ArticleFig(id=1319293506091340353, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=EN, label=Tab.2, caption=

Test instrument parameters

, figureFileSmall=null, figureFileBig=null, tableContent=
设备测量范围精度
压力表/MPa0~350.1%FS
0~700.1%FS
0~1000.1%FS
接近开关/mm0~52.5%FS
油流量计/（L⋅h-1）900~18 0002.5%FS
质量流量计/（kg⋅h-1）0~7 2000.35%FS
温度表/℃-40~2500.5%FS
), ArticleFig(id=1319293506187809346, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=CN, label=表2, caption=

试验仪器参数

, figureFileSmall=null, figureFileBig=null, tableContent=
设备测量范围精度
压力表/MPa0~350.1%FS
0~700.1%FS
0~1000.1%FS
接近开关/mm0~52.5%FS
油流量计/（L⋅h-1）900~18 0002.5%FS
质量流量计/（kg⋅h-1）0~7 2000.35%FS
温度表/℃-40~2500.5%FS
), ArticleFig(id=1319293506271695427, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=EN, label=Tab.3, caption=

Simulation and experimental data of thermal performance

, figureFileSmall=null, figureFileBig=null, tableContent=
参数仿真模型试验误差
一级排气温度/℃156.2158.11.2%
二级排气温度/℃132.1130.01.6%
质量流量/（kg⋅h-1）320.1311.62.7%
轴功率/kW17.3916.962.5%
容积效率87.03%84.91%2.5%
), ArticleFig(id=1319293506334609988, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293494888354291, language=CN, label=表3, caption=

热力性能仿真及试验数据

, figureFileSmall=null, figureFileBig=null, tableContent=
参数仿真模型试验误差
一级排气温度/℃156.2158.11.2%
二级排气温度/℃132.1130.01.6%
质量流量/（kg⋅h-1）320.1311.62.7%
轴功率/kW17.3916.962.5%
容积效率87.03%84.91%2.5%
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两级双缸液驱活塞压缩机变进气压力下的热力学性能研究
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齐强 , 张辰 , 王静瑶 , 贾晓晗 , 任省栋
流体机械 | 设计计算 2025,53(9): 21-28
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流体机械 |设计计算 2025 , 53 (9) : 21 -28
两级双缸液驱活塞压缩机变进气压力下的热力学性能研究
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齐强 , 张辰, 王静瑶, 贾晓晗, 任省栋
作者信息
  • 西安交通大学 能源与动力工程学院，西安 710049
通讯作者:
任省栋（1997），男，博士，主要从事氢气压缩机与加氢站研究，通信地址：710049陕西省西安市咸宁西路28号西安交通大学压缩机研究所，E-mail：。
作者简介:

齐强（1998），男，博士，主要从事液驱活塞压缩机研究，E-mail：。

Study on thermodynamic performance of two‑stage dual‑cylinder hydraulically driven piston compressor under varying inlet pressures
Qiang QI , Chen ZHANG, Jingyao WANG, Xiaohan JIA, Shengdong REN
Affiliations
  • School of Energy and Power Engineering，Xi'an Jiaotong University，Xi'an 710049，China
出版时间: 2025-09-30 doi: 10.3969/j.issn.1005-0329.2025.09.004
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摘要
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为了研究加氢站用两级液驱活塞压缩机在变吸气压力下的热力学性能变化规律，构建了液压系统和气体压缩系统耦合的两级液驱活塞压缩机热力仿真模型，并对模型进行了试验验证。结果表明，仿真与试验的热力性能参数最大误差为2.7%，吻合良好。当二级排气压力不变，一级进气压力从20.0 MPa减小到3.0 MPa时，一级压比基本保持不变，二级压比增大；一级气缸的排气温度保持在120.0 ℃左右；二级气缸的排气温度则从54.1 ℃增大至224.1 ℃，一级轴功率从60.2 kW降至14.9 kW，二级轴功率从12.7 kW增至35.5 kW，质量流量从55.5 kg/h减小到7.1 kg/h。两级双缸液驱机组适用于恒定吸气压力工况，在变吸气压力工况下运行时，会出现排气温度过高，轴功率不平衡，电机效率低的问题。针对该问题，可采取低压工况下采用高转速，在高压工况下采用低转速的运行策略。研究可为液驱活塞压缩机系统优化提供理论基础。

液驱活塞压缩机  /  热力仿真模型  /  流量  /  轴功率

To investigate the thermodynamic performance variation of a two‑stage hydraulically driven piston compressor for hydrogen refueling stations under varying suction pressures，a thermal simulation model of the two‑stage hydraulically driven piston compressor coupling the hydraulic system and gas compression system was constructed，and the model was experimentally validated.The results show that the maximum error between the simulated and experimental thermodynamic performance parameters is 2.7%，indicating good agreement.When the second‑stage discharge pressure remains unchanged and the first‑stage suction pressure decreases from 20.0 MPa to 3.0 MPa，the first‑stage pressure ratio remains essentially constant，while the second‑stage pressure ratio increases；the discharge temperature of the first‑stage cylinder remains around 120.0 ℃；the discharge temperature of the second‑stage cylinder increases from 54.1 ℃ to 224.1 ℃，the first‑stage shaft power decreases from 60.2 kW to 14.9 kW，the second‑stage shaft power increases from 12.7 kW to 35.5 kW，and the mass flow rate decreases from 55.5 kg/h to 7.1 kg/h.The two‑stage dual‑cylinder hydraulically driven unit is suitable for constant suction pressure conditions.When operating under varying suction pressure conditions，issues such as excessively high discharge temperature，unbalanced shaft power，and low motor efficiency occur.To address this problem，an operational strategy of using high speed under low‑pressure conditions and low speed under high‑pressure conditions can be adopted.The study provides a theoretical basis for the optimization of hydraulically driven piston compressor systems.

hydraulically driven piston compressor  /  thermal simulation model  /  flow rate  /  shaft power
齐强, 张辰, 王静瑶, 贾晓晗, 任省栋. 两级双缸液驱活塞压缩机变进气压力下的热力学性能研究. 流体机械, 2025 , 53 (9) : 21 -28 . DOI: 10.3969/j.issn.1005-0329.2025.09.004
Qiang QI, Chen ZHANG, Jingyao WANG, Xiaohan JIA, Shengdong REN. Study on thermodynamic performance of two‑stage dual‑cylinder hydraulically driven piston compressor under varying inlet pressures[J]. Fluid Machinery, 2025 , 53 (9) : 21 -28 . DOI: 10.3969/j.issn.1005-0329.2025.09.004
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0　引言
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氢能作为一种资源丰富、绿色低碳、应用广泛的二次能源，正逐步成为全球能源转型的重要载体［1］。《氢能产业发展中长期规划（2021-2035年）》提出要统筹建设氢能基础设施，稳步构建储运体系和加氢网络。氢能网络包括制氢、储氢、加氢和用氢4个环节［2］。加氢站作为加氢环节的主要组成部分，是连接上游制氢、储氢与下游氢燃料电池汽车等应用的重要枢纽［3］。作为加氢站的核心设备，压缩机的性能直接决定加氢站的运营状况［4］。液驱活塞压缩机以结构简单、成本较低、故障率低的优点，成为加氢站的主流机型［5］。
为了解液驱活塞压缩机内部热力学性能，相关人员对此展开研究。YE等［6］研究了活塞往复循环频率对氢气参数变化和换热特性的影响。WANG等［7］对压缩机在不同工作压力下的热力性能进行了计算和分析。杜宗钢［8］通过建立液驱氢气压缩机数值仿真模型，分析了气缸内氢气瞬态流动特性，获得了运行参数和吸气条件对压缩机性能的影响规律。黄振辉等［9］结合加氢站具体情况，确定了液驱活塞压缩机的选型方法。陈志鹏等［10］提出了基于变分时域分解、Hilbert变换和自相关函数相结合的氢气液驱压缩机的压缩周期参数非直接测量方法。但是，目前对高压工况下的研究仍然较少。
对于气体系统，曲柄连杆驱动的活塞压缩机与液驱活塞压缩机均为往复压缩机，共性较多，针对曲柄连杆驱动的活塞压缩机的性能研究对液驱活塞压缩机的性能研究具有较强的借鉴意义。DINU等［11］通过数值仿真，研究了进气压力、排气压力对活塞式天然气压缩机进排气量的影响规律。鲍怀谦等［12］对往复压缩机工作过程进行瞬态模拟，获得压缩机气缸内部流体的流动状态变化、气缸内流体压力脉动值随曲柄转角的变化。孙旭等［13］通过建立包含进排气阀、气缸的CFD模型，开展往复压缩机流量调控工况耦合气阀故障的压缩机热力学性能研究。王金铭等［14］等构建了多级双作用往复压缩机曲轴力学模型，分析了在正常气量调节工况，以及气缸级内和级间发生调节故障时，整周期交变载荷的变化规律。韩宝坤等［15］等建立压缩机吸气过程的瞬态动力学仿真，得到不同厚度的吸气阀片在不同状态下阀舌表面与气缸应力分布、阀舌位移与入口流量随时间的变化特性。赵杰等［16］对往复压缩机管路进行气固耦合分析、模态分析和谐响应分析，预测了气固耦合和非耦合作用下管路振动响应。对于无级气量调节系统，江志农［17-18］等通过仿真获得了执行机构动态响应和控制系统调控参数对气量调节的影响规律。张进杰等［19］在模型中添加吸气阀泄漏等故障，得到了不同故障对无级气量调节效果的影响规律。对于往复压缩机实际运行中，进气压力逐渐降低的变工况研究仍存在不足。
针对液压系统，刘辞英等［20］提出了一种阻尼槽-缓冲容腔-导油槽非对称配流盘结构，并进行内部流场仿真，通过压力脉动特性试验验证模型的准确性。赵叔吉等［21］建立了基于推移液压缸实验平台的液压系统仿真模型，为减小液压缸位置控制误差，对超调量进行了理论分析，发现载荷质量、初速度、黏性阻力和摩擦阻力均会对超调量产生影响。韩博等［22］建立了多液压缸系统故障仿真模型，分析了不同故障模式下多液压缸系统的影响因素，识别在典型故障模式下各分液压缸的运行参数变化，从而调整系统设计方案，实现多缸、多故障模式下不同任务和功能的仿真分析。但是，气体系统与液压系统的工作过程均是单独开展，未将两者耦合研究。
针对上述问题，文中提出一种液驱活塞压缩机热力仿真模型，并搭建两级液驱活塞压缩机试验台。对比仿真模型与试验数据中的热力学参数和压力变化曲线，验证该模型的有效性。利用热力仿真模型分析不同进气压力下的压缩机热力学性能变化影响。研究可为液压驱动活塞式氢气压缩机的可靠运行和设计优化提供参考。
1　液驱活塞压缩机热力仿真模型
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1.1　液驱压缩机系统构成
液驱活塞压缩机系统指利用液压油作为驱动介质，通过驱动侧活塞带动气体压缩侧的活塞运动，实现气体的吸入和压出的机械系统。系统主要包括液压油路、气路和冷却水路三部分。液压油路由驱动电机、油泵、电磁换向阀、单向阀和溢流阀等部件构成，具备驱动、换向、溢油和补油等功能。气路由压缩机、气罐和阀门等部件构成，实现气体增压功能。冷却水路则由电机、水泵、换热器和水箱等部件组成，负责降低气体温度。液驱压缩机作为系统的核心部件，由气缸、油缸、气侧活塞、油侧活塞和活塞杆等组成，如图1所示。
1.2　热力仿真模型建立
液驱压缩机系统的热力仿真模型如图2所示。在此模型中，气体和液体管路均假设为直管，并考虑气体在压缩过程中对气缸传热的影响，忽略油路的传热变化；考虑到环境换热量较小，忽略环境传热的影响；机组运行前进行10 MPa密封性能测试，确保气缸内气路没有明显泄漏，因此忽略泄漏的影响；与高速往复压缩机50 r/s的运行频率不同，液驱活塞压缩机的运行频率为0.25 r/s，可忽略压力波动的影响。冷却介质为水，固体介质为碳钢，液体介质为液压油。
液压油路中，使用2个尾部相连的双作用单连接液压油缸15，替代传统的双作用双连接液压缸。通过控制信号5直接改变三位四通阀16的供液方向，实现换向功能。当主油路油压超过30 MPa时，高压油通过3个溢油阀20进入油箱，实现溢油功能。当主油路管道压力低于3 MPa时，低流量单向油泵21通过单向阀2向主路补油，实现补油功能。在气路中，气动室14与压缩腔3的组合替代传统的气缸设计。气动室14用于计算活塞的位移，压缩腔3则用于计算压力变化。气阀由带弹簧的单向阀2替代。冷却水路采用恒定温度的冷水源10以维持系统的冷却效果。部分参数设置见表1，未涉及的参数均采用系统默认值。
2　热力性能参数计算
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活塞压缩机的热力性能参数主要包括排气温度、质量流量、轴功率和容积效率，是压缩机研究的关键组成部分。
2.1　排气温度
排气温度是指压缩机每一级压缩后排出气体的温度，计算公式如下：
Td=Ts+273.15εn-1n-273.15
式中，Td为排气温度，℃；Ts为吸气温度，℃；ε为压力比；n为压缩过程指数。
2.2　质量流量
质量流量指单位时间内通过特定截面的流体质量，计算公式如下：
Q=ρvπd2
式中，Q为流量，m3/h；ρ为液体密度，kg/m3；；v为液体流速，m/h；d为管道半径，m。
2.3　轴功率
轴功率是指驱动级传动给压缩机主轴的功率，由于模型忽略摩擦力，其计算公式如下：
Pi=160 000∑ psQnn-1ε(n-1)/n-1Zs+Zd2Zs
式中，Pi为轴功率，kW；ps为进气压力，MPa；Q为流量，m3/h；Zs为吸气压缩系数；Zd为排气压缩系数。
2.4　容积效率
容积效率是指压缩机实际流量与理论流量之比，通常用百分比表示，计算公式如下：
ηv=QcQt
式中，ηv为容积效率；Qc为实际流量，m3/h；Qt为理论流量，m3/h。
3　液驱活塞压缩机试验研究
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3.1　试验台及流程图
加氢站一般为开式循环系统，由于场地有限且没有稳定气源，本试验选择闭式循环模式，如图3所示。
为了减少液压油对油缸冲击所引发的振动，压缩机及其管路均被牢固地固定在撬座上。试验装置中，气体依次流经低压气罐、一级压缩机、一级换热器、级间气罐、二级压缩机、二级换热器、高压气罐，形成完整回路。为了控制气体流动，使用背压阀进行背压与节流。当背压阀节流后的气体流量与低压气罐中排出的气体流量达到平衡时，系统进入稳态运行。
3.2　测点布置及仪器仪表
隔离腔中安装有接近开关，用于控制三位四通阀的换向。气缸上布置压力表，用于获取缸内动态气压。油管中安装油流量计，通过积分计算得到活塞位移。压缩机各级出口处布置温度表，测量各级气缸的排气温度。此外，在配电柜出口架设功率仪，测量输入机组的功率。整个数据采集系统由USB2872多功能采集卡和计算机构成，采集卡采集压力表、温度表等输出的电压信号，计算机进行数据处理。仪器参数见表2。
4　结果分析与讨论
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4.1　仿真模型验证
采用排气温度、质量流量等热力性能参数与压力变化曲线等热力过程对仿真模型进行验证。验证过程中压缩介质为空气。
4.1.1　热力性能仿真及试验数据对比
热力性能仿真及试验数据见表3。
由表3可知，热力仿真模型的计算结果与试验结果的最大误差为2.7%，吻合良好，表明该模型能够有效反映压缩机的动态特性。热力仿真模型中的质量流量比试验的数据大，原因是忽略了管道阻力等因素，导致压缩机的进气量高于试验台实际情况。压比相似的情况下，质量流量增大会使指示功率增加。
4.1.2　热力过程仿真及试验数据对比
图4示出气缸内压力的变化曲线，仿真模型与试验数据吻合良好，变化规律一致。图中V表示运行过程中的行程，V0表示总行程。一级压力的平均误差为5.7%，二级压力的平均误差为6.2%。
气缸的进气压力呈现出先下降后上升的趋势，二级气缸表现明显。主要原因有2点：一是闭式系统通过节流阀将气体从排气端引回吸气端，无法模拟稳定的吸气气源，导致气压波动；二是级间气罐容积较小，缓冲效果差，难以平衡一级气缸排气压力在级间管路中引发的压力波动。此外，气缸的排气压力呈上升趋势，一级气缸表现明显。这是由于气阀阻力较大，导致气体无法及时排出。
与等熵压缩相比，试验与仿真的进气压力偏低，而排气压力偏高，导致较大的吸气损失和排气损失，从而增加系统的功耗。主要原因是转速过高，以及管道和气阀等设备的阻力较大，使气体在管路中的摩擦损失较多。
4.2　变进气压力下的热力过程与热力性能分析
在实际工况中，加氢站通常使用长管拖车作为低压气源，压缩机在运行过程中，一级进气压力会从20.0 MPa逐渐降低到5.0 MPa。本文以变化的进气压力作为研究对象，对高压工况下的液驱活塞压缩机热力性能变化展开研究。考虑到空气的主要成分氮气和氧气，与氢气气体组成相似都是双原子，具有近似的绝热压缩指数，使得其压缩过程的热力性能相似，所以计算中使用氢气作为压缩介质。
4.2.1　过程分析
进气压力对气缸压力的影响如图5所示。当一级进气压力从20.0 MPa减小到3.0 MPa时，一级绘制的封闭区域逐渐减小，二级绘制的封闭区域逐渐增大，说明随着一级进气压力的逐步降低，一级耗功逐渐减小，而二级耗功逐渐增大；一级的压缩基本行程不变，而二级的压缩行程加长，说明一级压比基本保持不变，而二级压比增大。
4.2.2　排气温度分析
由表1可知，冷却后压缩机的进气温度为25 ℃。进气压力对排气温度的影响如图6所示。当一级进气压力从20.0 MPa减小到3.0 MPa时，一级气缸的排气温度维持在120.0 ℃上下；而二级气缸的排气温度则从54.1 ℃上升至224.1 ℃，呈抛物线型，差值逐渐增大。二级压比的增加导致二级排气温度升高。此外，当进气压力降至3 MPa时，排气温度达到224.1 ℃，远超管道内氢气温度小于180 ℃的要求。因此，压比较高时，建议采用吸气预冷或加大级间冷却的方法降低排气温度，从而避免排气温度过高带来的安全隐患。
4.2.3　能耗分析
进气压力对轴功率的影响如图7所示。
从图可见，当一级进气压力从20.0 MPa减小到3.0 MPa时，一级轴功率从60.2 kW下降至14.9 kW；二级轴功率则从12.7 kW上升到38.5 kW。一级轴功率随着进气压力的降低而减少，而二级轴功率随着二级压比的降低而增加。当气体流量恒定时，两级的压比越接近，压缩气体能耗越小。进气压力为10 MPa时，压缩机一、二级压比差异较小，轴功率相近，此时电机工作稳定且效率较高。在其他进气压力工况下压比差异增大，导致一、二级的轴功率不平衡，电机处于变工况运行，效率低且能耗高。说明两级双缸液驱压缩机适用于恒定工况。
4.2.4　质量流量分析
进气压力对质量流量的影响如图8所示。当一级进气压力从20.0 MPa减小到3.0 MPa时，质量流量从55.5 kg/h减小到7.1 kg/h，差值保持不变容积效率也保持不变。这表明，提升一级进气压力可以显著提高系统流量。根据电机的轴功率与效率曲线图可知，电机的轴功率越接近额定功率，电机效率越高。在低压工况下采用高转速，而在高压工况下采用低转速，可以使压缩气体所需的轴功率保持在额定功率附近，从而减少电机损耗，提高系统效率。
5　结论
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当两级压缩过程中二级排气压力保持不变时，一级进气压力从20.0 MPa减小到3.0 MPa，主要结论如下：
（1）二级气缸的排气温度从54.1 ℃上升至224.1 ℃。当吸气压力较低时，可以通过吸气预冷或加大级间冷却的方法降低排气温度，从而避免排气温度过高带来的安全隐患。
（2）一级轴功率从60.2 kW下降至14.9 kW，二级轴功率从12.7 kW上升到38.5 kW。在设计工况下，压缩机一、二级压比差异较小，轴功率相近，电机运行效率高，说明两级双缸液驱压缩机适用于恒定工况。
（3）质量流量从55.5 kg/h减小到7.1 kg/h，容积效率与一级进气压力无明显相关性。建议采用低压工况下高转速、高压工况下低转速的策略，使压缩气体所需的轴功率保持在额定功率附近，从而提高系统效率。

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doi: 10.3969/j.issn.1005-0329.2025.09.004
  • 接收时间：2024-11-28
  • 首发时间：2026-10-20
  • 出版时间：2025-09-30
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  • 收稿日期：2024-11-28
  • 修回日期：2025-06-10
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    西安交通大学 能源与动力工程学院，西安 710049

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任省栋（1997），男，博士，主要从事氢气压缩机与加氢站研究，通信地址：710049陕西省西安市咸宁西路28号西安交通大学压缩机研究所，E-mail：。
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2种不同金属材料的力学参数
科
Family		 属数
Number of
genus		 种数
Number of
species		 占总种数比例
Percentage of
total species (%)		 属
Genus		 种数
Number of
species		 占总种数比例
Percentage of total
species (%)
鹅膏菌科Amanitaceae 2 11 5.26 鹅膏菌属 Amanita 10 4.78
小菇科 Mycenaceae 2 12 5.74 丝盖伞属 Inocybe 5 2.39
多孔菌科 Polyporaceae 8 14 6.70 蜡蘑属 Laccaria 5 2.39
红菇科 Russulaceae 3 23 11.00 小皮伞属 Marasmius 6 2.87
小菇属 Mycena 11 5.26
光柄菇属 Pluteus 5 2.39
红菇属 Russula 17 8.13
栓菌属 Trametes 5 2.39
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