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Article(id=1319293359122920364, tenantId=1146029695717560320, journalId=1316781260334530699, issueId=1319293303430955712, articleNumber=null, orderNo=null, doi=10.3969/j.issn.1005-0329.2025.08.004, pmid=null, cstr=null, oa=null, hot=null, price=null, onlineType=0, articleFormat=0, articleType=null, articleTypeStr=null, receivedDate=1739548800000, receivedDateStr=2025-02-15, revisedDate=1749916800000, revisedDateStr=2025-06-15, acceptedDate=null, acceptedDateStr=null, onlineDate=1792473648099, onlineDateStr=2026-10-20, pubDate=1756483200000, pubDateStr=2025-08-30, doiRegisterDate=null, doiRegisterDateStr=null, onlineIssueDate=1792473648099, onlineIssueDateStr=2026-10-20, onlineJustAcceptDate=null, onlineJustAcceptDateStr=null, onlineFirstDate=null, onlineFirstDateStr=null, sourceXml=null, magXml=null, createTime=1792473648099, creator=13701087609, updateTime=1792473648099, updator=13701087609, issue=Issue{id=1319293303430955712, tenantId=1146029695717560320, journalId=1316781260334530699, year='2025', volume='53', issue='8', pageStart='1', pageEnd='118', issueExtLink='null', onlineDate='null', pubDate='1756483200000', pubDateStr='2025-08-30', beforeIssueId=null, nextIssueId=null, price=null, status=1, issueComplete=1, articleOrder=1, issueType=-1, specialIssue=null, createTime=1792473634822, creator='13701087609', updateTime=1792473634822, updator='13701087609', preIssue=null, nextIssue=null, articleTotal=null, ext=null, issueFiles=null, downloadFileDto=null}, startPage=27, endPage=35, ext={EN=ArticleExt(id=1319293359299081133, articleId=1319293359122920364, tenantId=1146029695717560320, journalId=1316781260334530699, language=EN, title=Investigation on matching characteristics of hydraulic and booster systems of hydraulic‑driven hydrogen compressor, columnId=null, journalTitle=Fluid Machinery, columnName=null, runingTitle=null, highlight=null, articleAbstract=

To investigate the effects of key parameters such as gas & oil piston diameter and control signal of proportional valve on the actual cycle indicated energy and stability of hydrogen compressor，a numerical model of hydraulic and booster coupled systems of a two‑stage 90 MPa hydraulic‑driven hydrogen compressor is established.The effects of key parameters such as gas & oil piston diameter and control signal of proportional valve on the actual cycle indicated energy and stability of hydrogen compressor are studied.The influence behavior of above parameters acting on the displacement，velocity，pressure and temperature inside gas cylinder，and pressure inside oil cylinder are analyzed. The reliability of the prediction method is verified by the experimental test. The results show that as the gas‑piston diameter increases，the maximum velocity gradually increases. When the gas‑piston diameter of stage 1 is 140 mm，its maximum reversing velocity is 0.199 m/s，which increases by nearly 28% compared to the velocity at 130 mm. For the process of suction and discharge of hydraulic‑driven hydrogen compressor，the changes of gas & oil piston diameter have the opposite influence for the pressure and temperature inside gas and pressure inside oil cylinder. When the gas‑piston diameter and oil‑piston diameter of stage 1 are 130 mm and 160 mm，the maximum pressure inside oil cylinder is15.23 MPa，and the maximum discharge pressure and temperature are 30.3 MPa and 390.5 K，respectively. Based on the above study，optimal matching of the gas & oil piston diameter and control signal of proportional valve are carried out，which bring about the energy conservation of 90 MPa hydraulic‑driven hydrogen compressor. The research can provide reference for the optimization design of hydrogen compressors.

, authors=Xudong TIAN, Shengli WANG, Chengyan ZHANG, Bo TAO, Minghao DAI, authorsList=Xudong TIAN, Shengli WANG, Chengyan ZHANG, Bo TAO, Minghao DAI, authorCompany=null, correspAuthors=Shengli WANG, 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=1319293363615019978, articleId=1319293359122920364, tenantId=1146029695717560320, journalId=1316781260334530699, language=CN, title=液驱氢压机液压与增压系统耦合匹配特性研究, columnId=1317484287230853622, journalTitle=流体机械, columnName=设计计算, runingTitle=null, highlight=null, articleAbstract=

为了研究气缸活塞直径、油缸活塞直径、换向阀控制信号等关键参数对氢压机实际循环指示功和运行稳定性的影响，构建了90 MPa两级液驱氢压机液压与增压耦合系统的数值模型，分析上述参数对活塞行程、速度、气缸内压力和温度、油缸压力的影响规律，并试验验证了数值方法的可靠性。结果表明，随着气缸活塞直径的增大，最大速度逐渐增大，当一级气缸活塞直径为140 mm时，其最大换向速度为0.199 m/s，对比直径130 mm时的速度增加约28%；在氢压机增压缸进气和排气两阶段，气缸/油缸直径的变化引起两阶段的气体压力、温度及油压的变化规律相反，当一级气缸活塞直径和油缸活塞直径分别为130 mm和160 mm时，一级油压最大为15.23 MPa，气缸内最大排气压力和温度分别为30.3 MPa和390.5 K。基于上述研究，对90 MPa氢压机气缸直径、油缸直径及换向阀控制信号进行了优化匹配，实现了氢压机增压系统节能化设计。研究可为氢压机的优化设计提供参考。

, authors=田旭东, 王胜利, 张成彦, 陶波, 代明昊, authorsList=田旭东, 王胜利, 张成彦, 陶波, 代明昊, authorCompany=null, correspAuthors=王胜利, authorNote=

田旭东（1966），男，教授级高级工程师，主要从事压缩机制冷设备节能技术、人工环境模拟技术和产品性能可靠性试验技术研究工作，E-mail：。

, correspAuthorsNote=
王胜利（1990），男，博士，助理研究员，主要从事高压无油往复式压缩机设计技术研究工作，通信地址：230031安徽 省合肥市长江西路888号合肥通用机械研究院有限公司，E-mail：。
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田旭东（1966），男，教授级高级工程师，主要从事压缩机制冷设备节能技术、人工环境模拟技术和产品性能可靠性试验技术研究工作，E-mail：。

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田旭东（1966），男，教授级高级工程师，主要从事压缩机制冷设备节能技术、人工环境模拟技术和产品性能可靠性试验技术研究工作，E-mail：。

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and their effects on pressure in the oil‑cylinder, figureFileSmall=Xy8Kx65jFr7tJ/aOwkIOZA==, figureFileBig=fYwN3m7ZHJY3mK1ECXUNjA==, tableContent=null), ArticleFig(id=1319293369004699655, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293359122920364, language=CN, label=图13, caption=不同控制信号及其对油缸内压力的影响, figureFileSmall=Xy8Kx65jFr7tJ/aOwkIOZA==, figureFileBig=fYwN3m7ZHJY3mK1ECXUNjA==, tableContent=null), ArticleFig(id=1319293369088585736, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293359122920364, language=EN, label=Fig.14, caption=Optimized hydraulic‑driven hydrogen compressor system, figureFileSmall=cAZz4GVB7M0a61aEB+YVxw==, figureFileBig=ivlYztBIZav65YBVGMvGxg==, tableContent=null), ArticleFig(id=1319293369147305993, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293359122920364, language=CN, label=图14, caption=优化后的液驱氢压机系统, figureFileSmall=cAZz4GVB7M0a61aEB+YVxw==, figureFileBig=ivlYztBIZav65YBVGMvGxg==, tableContent=null), ArticleFig(id=1319293369231192074, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293359122920364, language=EN, label=Tab.1, caption=

Predicted and measured oil pressure of each stage

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进气压力/MPa级数实测压力/MPa预测压力/MPa相对误差/%
5.76一级13.4614.034.23
二级20.6321.514.26
10.43一级13.3313.460.98
二级20.0920.763.33
13.00一级14.7415.233.22
二级19.5820.394.14
15.09一级15.0715.060.07
二级19.4619.691.82
19.38一级15.715.941.53
二级18.6219.122.69
), ArticleFig(id=1319293369294106635, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293359122920364, language=CN, label=表1, caption=

数值计算和实测的各级油缸内压力对比

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进气压力/MPa级数实测压力/MPa预测压力/MPa相对误差/%
5.76一级13.4614.034.23
二级20.6321.514.26
10.43一级13.3313.460.98
二级20.0920.763.33
13.00一级14.7415.233.22
二级19.5820.394.14
15.09一级15.0715.060.07
二级19.4619.691.82
19.38一级15.715.941.53
二级18.6219.122.69
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液驱氢压机液压与增压系统耦合匹配特性研究
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田旭东 , 王胜利 , 张成彦 , 陶波 , 代明昊
流体机械 | 设计计算 2025,53(8): 27-35
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流体机械 |设计计算 2025 , 53 (8) : 27 -35
液驱氢压机液压与增压系统耦合匹配特性研究
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田旭东 , 王胜利 , 张成彦, 陶波, 代明昊
作者信息
  • 合肥通用机械研究院有限公司 高端压缩机及系统技术全国重点实验室，合肥 230031
通讯作者:
王胜利（1990），男，博士，助理研究员，主要从事高压无油往复式压缩机设计技术研究工作，通信地址：230031安徽 省合肥市长江西路888号合肥通用机械研究院有限公司，E-mail：。
作者简介:

田旭东（1966），男，教授级高级工程师，主要从事压缩机制冷设备节能技术、人工环境模拟技术和产品性能可靠性试验技术研究工作，E-mail：。

Investigation on matching characteristics of hydraulic and booster systems of hydraulic‑driven hydrogen compressor
Xudong TIAN , Shengli WANG , Chengyan ZHANG, Bo TAO, Minghao DAI
Affiliations
  • Hefei General Machinery Research Institute Co.，Ltd.，State Key Laboratory of High‑end Compressorand System Technology，Hefei 230031，China
出版时间: 2025-08-30 doi: 10.3969/j.issn.1005-0329.2025.08.004
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摘要
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为了研究气缸活塞直径、油缸活塞直径、换向阀控制信号等关键参数对氢压机实际循环指示功和运行稳定性的影响，构建了90 MPa两级液驱氢压机液压与增压耦合系统的数值模型，分析上述参数对活塞行程、速度、气缸内压力和温度、油缸压力的影响规律，并试验验证了数值方法的可靠性。结果表明，随着气缸活塞直径的增大，最大速度逐渐增大，当一级气缸活塞直径为140 mm时，其最大换向速度为0.199 m/s，对比直径130 mm时的速度增加约28%；在氢压机增压缸进气和排气两阶段，气缸/油缸直径的变化引起两阶段的气体压力、温度及油压的变化规律相反，当一级气缸活塞直径和油缸活塞直径分别为130 mm和160 mm时，一级油压最大为15.23 MPa，气缸内最大排气压力和温度分别为30.3 MPa和390.5 K。基于上述研究，对90 MPa氢压机气缸直径、油缸直径及换向阀控制信号进行了优化匹配，实现了氢压机增压系统节能化设计。研究可为氢压机的优化设计提供参考。

液驱氢压机  /  液压与增压系统  /  耦合匹配特性  /  数值模拟

To investigate the effects of key parameters such as gas & oil piston diameter and control signal of proportional valve on the actual cycle indicated energy and stability of hydrogen compressor，a numerical model of hydraulic and booster coupled systems of a two‑stage 90 MPa hydraulic‑driven hydrogen compressor is established.The effects of key parameters such as gas & oil piston diameter and control signal of proportional valve on the actual cycle indicated energy and stability of hydrogen compressor are studied.The influence behavior of above parameters acting on the displacement，velocity，pressure and temperature inside gas cylinder，and pressure inside oil cylinder are analyzed. The reliability of the prediction method is verified by the experimental test. The results show that as the gas‑piston diameter increases，the maximum velocity gradually increases. When the gas‑piston diameter of stage 1 is 140 mm，its maximum reversing velocity is 0.199 m/s，which increases by nearly 28% compared to the velocity at 130 mm. For the process of suction and discharge of hydraulic‑driven hydrogen compressor，the changes of gas & oil piston diameter have the opposite influence for the pressure and temperature inside gas and pressure inside oil cylinder. When the gas‑piston diameter and oil‑piston diameter of stage 1 are 130 mm and 160 mm，the maximum pressure inside oil cylinder is15.23 MPa，and the maximum discharge pressure and temperature are 30.3 MPa and 390.5 K，respectively. Based on the above study，optimal matching of the gas & oil piston diameter and control signal of proportional valve are carried out，which bring about the energy conservation of 90 MPa hydraulic‑driven hydrogen compressor. The research can provide reference for the optimization design of hydrogen compressors.

hydraulic‑driven hydrogen compressor  /  hydraulic and booster systems  /  matching characteristic  /  numerical simulation
田旭东, 王胜利, 张成彦, 陶波, 代明昊. 液驱氢压机液压与增压系统耦合匹配特性研究. 流体机械, 2025 , 53 (8) : 27 -35 . DOI: 10.3969/j.issn.1005-0329.2025.08.004
Xudong TIAN, Shengli WANG, Chengyan ZHANG, Bo TAO, Minghao DAI. Investigation on matching characteristics of hydraulic and booster systems of hydraulic‑driven hydrogen compressor[J]. Fluid Machinery, 2025 , 53 (8) : 27 -35 . DOI: 10.3969/j.issn.1005-0329.2025.08.004
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0　引言
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近年来，氢燃料电池汽车等氢能产业的发展引起全世界广泛关注［1］。预计到2030年，在德国、日本、韩国及美国加利福尼亚等国家和地区销售的每12辆汽车中就有1辆由氢气提供动力［2-3］。根据《氢能产业发展中长期规划（2021-2035年）》，2025年会形成较为完整的氢能产业体系和供应链，氢燃料电池车保有量将达到5万辆［4］。加氢站作为氢能在交通领域进行大规模应用的重要基础设施，主要包括储氢装置、压缩设备、加注设备、站控系统等模块。其中，液驱氢压机和隔膜氢压机是加氢站氢气增压最常用的两种设备。与隔膜氢压机相比，液驱氢压机具有设备成本低、结构简单、维护时间短且费用低、可带载启停等显著优势［5-7］，但液驱氢压机仍存在系统噪声大和能耗大等问题。《工信部等七部门关于加快推动制造业绿色化发展的指导意见》中明确要求：加快推动新兴产业绿色高起点发展，建立健全支撑制造业绿色发展的技术。90 MPa液驱氢压机作为新兴装备，实现其低能耗是加氢站等氢能产业绿色发展的必经之路。
液驱氢压机的能耗是由液压与增压系统综合决定，影响两者耦合匹配的因素众多，例如气缸/油缸活塞直径、行程、活塞线速度、气缸余隙容积、级间缓冲罐容积、管路长度及换向阀控制信号等关键参数，学者们对这些因素进行了研究并取得了一定进展。ZHOU等［8］通过引入可压缩油模型、漏油模型，建立了自换向液压驱动活塞压缩机运行过程的动力学模型，并采用相关仿真软件模拟压缩机的运行过程，验证了模型的可行性。YE等［9-11］基于CFD方法，采用流固耦合传热法、动网格技术，构建了液压驱动活塞压缩机的瞬态计算模型，研究往复循环频率对增压缸内热力学参数和壁面传热特性的影响规律，并将绝热压缩理论计算的排气温度与瞬态模型预测的排气温度进行了对比分析。LIGEN等［12］基于试验数据构建了空气驱动氢气压缩机仿真模型，研究了压缩机排气压力、流量、排气温度和容积效率的动态演变过程。张璟等［13］数值研究了300 MPa超高压液驱活塞压缩机液封结构，提高了活塞密封寿命和密封效果，并通过CFD仿真分析了密封液流场及油液各处的压力分布和循环流量。然而，上述研究主要针对液驱压缩机组增压系统开展，忽略了液驱压缩机组液压系统的影响，缺乏两者耦合匹配节能技术的相关研究。
为了解决上述问题，部分学者建立了液驱压缩机组液压与增压耦合系统的简化数值模型和试验测试模型，并进行了初步探索。WANG等［14］试验研究了两级液压驱动活塞式压缩机在不同运行压力下的p-V曲线，计算和分析了压缩机在不同工作压力下的性能。刘纯华等［15-16］基于液驱活塞式压缩机液压系统仿真模型，研究分析了油缸活塞直径、管路长度及控制信号对液压系统动态性能的影响，并提出液驱活塞式压缩机液压系统优化改进方案。陈志鹏等［17］建立了液驱往复式压缩机活塞冲击信号的数学模型，模拟活塞撞击气缸极限位置的振动信号，可为氢气压缩机状态检测、故障诊断与寿命预测提供运行工况参数。但是，上述数值研究对液压系统进行了较多简化，未实现对液压系统的真实建模。
本文针对90 MPa液驱氢压机机组，建立液压与增压系统耦合匹配数值模型，创新性地实现对液压系统的真实建模；研究气缸活塞直径、油缸活塞直径、换向阀控制信号等关键参数对氢压机机组运行稳定性的影响，分析上述参数对活塞行程、速度、气缸内压力和温度、油缸压力的影响规律，进一步探究液压与增压系统耦合匹配节能技术；同时，对氢压机气缸直径、油缸直径及控制信号进行优化匹配，并试验验证数值方法的可靠性。
1　液压与增压耦合系统数值建模
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1.1　A10VO负载敏感泵控制原理
90 MPa液驱氢压机的液压系统采用力士乐A10VO负载敏感泵，其控制原理见图1。
负载敏感阀实际为O型中位机能的三位三通阀，为表示方便，以二位三通阀来示意。此外，在负载敏感阀和压力补偿阀的右侧分别设有阻尼，其主要作用是减小负载敏感阀和压力补偿阀的阀芯在运动过程的液压冲击，提高系统稳定性。
A10VO负载敏感油泵的工作原理：当驱动电机未启动时，油泵排量为0，负载敏感阀在预载弹簧力的作用下处于右位工作状态，变量调节缸内油压为0，油泵斜盘在液压行程限位器的预载弹簧力作用下倾角最大，这表明油泵将在最大排量下启动；驱动电机启动初期，油泵输出压力不足以克服负载敏感阀右端的弹簧力，油泵仍以最大排量工作。同时，随着系统压力的升高，负载敏感阀的阀芯受力有3种情况：（1）当F左油=F右油+F弹时，负载敏感阀O型中位机工作，变量调节缸油口关闭，油泵的排量处于恒定值；（2）当F左油＜F右油+F弹时，负载敏感阀阀芯左移，右位工作；（3）当F左油＞F右油+F弹时，负载敏感阀阀芯右移，左位工作，高压油进入变量调节缸，油泵斜盘倾角减小，排量降低。
1.2　液压与增压系统仿真建模
液驱氢压机机组由液压驱动系统、增压压缩机、缓冲罐、冷却器、电控系统、手动/气动阀门和管路等模块组成，如图2所示。其中，压缩机采用两级增压，一、二级各两缸，一、二级压缩机由2个独立A10VO负载敏感油泵驱动，油泵排量分别为125 ml/r和140 ml/r。
根据液驱氢压机及负载敏感泵工作原理，分别搭建了增压系统模型和液压系统模型，如图3所示。在一级压缩机入口设有恒压恒温入口流量源〔见图3（a）〕，其依次通过单向阀和余隙腔体与一级气缸相连接，左右气缸通过单向阀与级间缓冲罐相连接，级间缓冲罐与二级余隙腔体相连接，二级气缸通过单向阀与排气缓冲罐相连，最后排气缓冲罐与气体泄压阀连通。其中，一、二级余隙容积分别为0.05，0.03 L，级间和排气缓冲罐容积均为10 L。恒温恒压入口流量源设置各工况实测进气压力和温度，泄压阀设定压力为实测压缩机排气压力。
为了实现液驱系统真实建模，本文对两负载敏感泵的液压系统分开进行建模。根据A10VO负载敏感泵控制原理图和油泵厂家提供的油泵三维图，建立各级液压系统的数值模型〔见图3（b）〕，主要包括电机、油泵、安全阀、换向阀、压力补偿阀、负载敏感阀和油泵排量调节机构。
1.3　耦合系统数学分析模型
压缩机的增压原理：中间的油缸活塞在液压油推动下带动两端的气缸活塞同时进行压缩或膨胀。增压缸运动机构受力分析如图4所示。两端活塞及活塞杆直径相同，当运动机构以速度v向右运动时，左端气缸处于膨胀或进气阶段，则右端气缸则处于压缩或排气阶段，此时，左端油缸处于进油阶段，右端油缸处于排油阶段。
假设气缸、油缸活塞环与缸体内壁间的摩擦力作用于气缸、油缸活塞的重心，即忽略摩擦力的力矩。在不计运动机构液阻及气阻情况下，根据牛顿第二定律，其力平衡方程：
pg1-pg2A1+po1-po2A-2Ffp-Ffo=mdvdt
式中，pg1，pg2分别为左端、右端气缸压力，Pa；A1，A分别为气缸、油缸活塞端面面积，m2；po1，po2分别为左端、右端油缸压力，Pa；Ffp，Ffo分别为气缸、油缸活塞环与缸体内壁间的摩擦力，Pa。
对于气缸活塞组件，在其往复运动过程中，活塞组件和缸体内壁之间会产生摩擦力，且其作为阻力作用在活塞上，摩擦力大小与活塞运动速度、摩擦系数和接触面润滑状况等因数有关。文中假设摩擦力在压缩机整个工作周期中是恒定的。因此，摩擦力大小可根据下述经验公式估算［18］：
Ffp=Ac16(psi+pdi)(1ηm-1)
式中，Ac为气缸截面面积，m2；psi为名义进气压力，Pa；pdi为名义排气压力，Pa；ηm为机械效率。
对于油缸活塞组件，其摩擦力的大小与缸体内径、密封接触宽度、油膜高度和相对运动速度有关。在计算油缸活塞密封面的摩擦力时，假设密封接触表面光滑且接触面上的压力是均匀分布的，借鉴文献［19］对液压缸摩擦特性的研究，油缸活塞摩擦力可表示为：
Ffo=πDpμvh0∫0B4h0h-3h0h2dx
式中，Dp为油缸内径，m；h0为油膜高度，m；μ为动力黏性系数，Pa·s；B为密封接触宽度，m。
2　液驱氢压机试验系统设计
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2.1　试验系统
为了验证数值方法的准确性，对90 MPa液驱氢压机进行了试验研究。搭建的试验系统如图5所示，主要包含气路、油路和冷却水路三部分。气体输入端并联有置换用气管路和氢气入口管路，管路中设有用于检测气体流量、压力和温度的检测模组；在一级压缩机前端、二级压缩机前后两端分别设有冷却器GC01，GC02和GC03；二级压缩机压缩后的高压气体进入高压气瓶，高压气体经减压阀减压后进入低压气瓶，低压气瓶与氢气入气管路相连，形成闭式循环管路。
试验系统中，90 MPa氢压机的进气、级间和排气压力的采集，分别使用量程为0~25 MPa（精度±0.02%），0~60 MPa（精度±0.02%）和0~120 MPa（精度±0.05%）的压力变送器；气体温度的采集，均采用量程-10~200 ℃（精度±0.5%）的温度变送器；进气流量的采集，使用量程0~1 000 kg（精度±0.2%）科里奥利质量流量计；液压系统压力的采集，采用量程为0~40 MPa（精度±0.5%）DG系列工业型压力变送器。
2.2　试验工况
试验介质为氮气，进气温度16.6 ℃，冷却水进水和出水温度分别为19.5 ℃和25.6 ℃。大气压力102.17 kPa，环境温度17.9 ℃。试验时进气压力选取5，10，12.5，15，20 MPa共5个工况点，对90 MPa液驱氢压机组进行变工况性能试验研究，各工况进气压力偏差控制在±1 MPa。
3　结果与分析
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3.1　模型验证
实测进气压力的平均值分别为5.76，10.43，13.00，15.09，19.38 MPa，相对应的排气压力的平均值分别为92.34，90.38，89.55，91.3，89.45 MPa。选取上述5个进气工况进行数值模拟
对氢压机各级最大油压和一级排气压力进行了预测，并与实测值进行对比，结果如表1和图6所示。可以看出，各级最大油压和一级排气压力的数值预测值与试验值变化一致，一级和二级最大油压的计算值略高于实测值，最大误差仅为4.26%；一级排气压力的预测值略小于实测值，最小误差为1.28%，最大误差仅为4.92%，总体计算结果较为精准。
3.2　气缸活塞直径的影响
在保证一、二级压缩机油缸直径d1=160 mm，d2=185 mm不变情况下，一级压缩机气缸活塞直径D1选取130，135，140 mm三种工况进行研究，进气压力为实测的13.00 MPa，液压系统设置参数均保持不变。以一级压缩机为研究对象，研究其活塞行程、活塞速度、一级油缸压力、一级气缸压力和温度随气缸活塞直径的变化规律。
图7示出不同气缸活塞直径下，活塞的行程和速度曲线。当D1=130 mm时，活塞行程为260 mm；当D1=135 mm时，活塞行程为226 mm；当D1=140 mm时，活塞行程为172 mm。在液压系统参数设置不变工况下，随着气缸活塞直径增大，活塞行程逐渐减小、最大速度逐渐增大，且在活塞换向时其速度畸变越明显。当D1=130 mm时，最大换向速度为0.156 m/s；当D1=135 mm时，最大换向速度为0.177 m/s；当D1=140 mm时，最大换向速度为0.199 m/s，相对于130 mm时的换向速度增加约28%，表明液压系统与增压系统间的匹配效果变差。
氢压机工作时包括进气、压缩、排气和膨胀共4个阶段，如图8所示。在进气、压缩和膨胀阶段，气缸内压力和温度随着气缸活塞直径的增加而增大；在氢压机排气阶段，气缸内压力和温度随着气缸活塞直径的增加而减小，压力最大和最小值分别为30.3 MPa和29.7 MPa，温度最大和最小值分别为390.5 K和382.4 K。在进气阶段，油缸内压力随着气缸活塞直径的增加而增大；排气阶段与之相反，随着气缸活塞直径的增加而减小，油缸内压力最大和最小值分别为15.9 MPa和15.23 MPa；压缩阶段无显著变化；在膨胀阶段前期，油缸内压力随着气缸活塞直径的增加而减小，压缩阶段后期，油缸内压力随着气缸活塞直径的增加而增大。
氢压机的工作过程由压缩过程a→b、排气过程b→c、膨胀过程c→d和进气过程d→a共4个阶段组成。实际循环指示功为图9中a→b→c→d围成的封闭曲面面积。
值得注意的是，在液压系统参数不变的前提下，随着气缸活塞直径的增大，氢压机的进、排气量均有所减小。其具体原因：当气缸活塞直径增大时，活塞所受的气体力会伴随着活塞直径的增大而增大，而在负载敏感泵输出压力不变的工况下，活塞往复行程会出现下降，导致氢压机的余隙容积变大、压缩效率降低，进而引起氢压机的进、排气量均有所减小。这意味着当实际应用不需要大排量时，可以采用增加缸径、减小行程的方法来节省能耗。
3.3　油缸活塞直径的影响
在保持氢压机一级气缸活塞直径D1=130 mm不变情况下，选取一级油缸活塞直径d1分别为160，165，170 mm三种工况进行研究，其结果见图10。可以看出，与气缸活塞直径变化不同，油缸活塞直径变化对活塞行程和速度的影响较小，不同油缸活塞直径下的行程曲线最大值基本重合；活塞速度曲线略有不同，3种工况下最大行程和速度分别为274 mm和0.15 m/s。
不同油缸活塞直径下，气缸内压力、温度和油缸内压力如图11所示。可以看出，在氢压机压缩阶段，气缸内压力和温度随着油缸活塞直径的增加而减小；在排气阶段，气缸内压力和温度随着油缸活塞直径的增加而增大，气缸内压力最大和最小值分别为31.5 MPa和30.3 MPa，温度最大和最小值分别为394.4 K和390.4 K；在进气和膨胀阶段，油缸活塞直径的变化对气缸内压力和温度的影响不大。在排气阶段，油缸内压力随着油缸活塞直径的增加而增大；在进气和膨胀阶段则与之相反，随着油缸活塞直径的增加而减小，油缸内压力的最大和最小值分别为15.3 MPa和14.8 MPa。在增压系统参数不变的前提下，随着油缸活塞直径由165 mm增加到170 mm，实际循环指示功逐渐增大，如图12所示。
3.4　比例换向阀控制信号的影响
选取3种换向阀控制信号开展研究，控制信号1为矩形信号，控制信号2和3为梯形信号，如图13（a）所示。信号2和3相对于信号1增加了缓冲阶段，信号2和3的缓冲时间分别为0.1 s和0.2 s，信号3的缓冲斜率度小于信号2。图13（b）示出不同控制信号下油缸内压力的变化曲线。可以看出，控制信号1在比例阀换向时压力变动最大，控制信号2其次，控制信号3最小。这表明：在比例换向阀变向时，当控制信号无缓冲阶段，液压油对换向阀的冲击最大；当存在缓冲阶段，冲击有所降低，且缓冲信号斜度越小，液压油对换向阀的冲击越小，液压系统越稳定。
3.5　氢压机优化设计
基于上述研究，发现气缸活塞直径、油缸活塞直径、换向阀控制信号这3个关键参数的变化，对90 MPa液驱氢压机活塞行程、速度、气缸内压力和温度、油缸内压力和实际循环指示功的影响程度不同。其中，气缸活塞直径变化影响最大，油缸活塞直径变化影响次之，而换向阀控制信号的变化则主要影响换向时液压系统的稳定性。其主要原因：文中选用的一、二级油泵的排量分别为125 ml/r和140 ml/r，而在氢压机工作时，一、二级油缸所需油量均小于该油泵排量，油泵排出的多余的液压油会被各级溢流阀溢流掉。因此，在油缸活塞直径变化范围不大的情况下，其对氢压机活塞行程、速度、气缸内压力和温度、油缸内压力和实际循环指示功的影响基本可以忽略。然而，与油缸活塞直径不同，气缸活塞直径一旦变化，上述参数均会受到不同程度的影响。因此，气缸活塞直径变化相对于油缸活塞直径变化影响较大。
基于上述研究，文中对90 MPa液驱氢压机进行了优化设计，优化匹配了气缸活塞直径、油缸活塞直径及换向阀控制信号。优化后，一级气缸和一级油缸活塞直径分别选取130 mm和160 mm，控制信号选取控制信号3，加工成撬后的液驱氢压机系统如图14所示。成撬完成后，对90 MPa液驱氢压机进行了500 h运行试验，并在试验后进行了拆机检查，未发现异常。
4　结论
收起
（1）气缸活塞直径、油缸活塞直径、换向阀控制信号这3种关键参数的变化，对90 MPa液驱氢压机活塞行程、速度、气缸内压力和温度、油缸内压力和实际循环指示功的影响程度不同。其中，气缸活塞直径变化影响最大，油缸活塞直径变化影响次之，而换向阀控制信号的变化则主要影响换向时液压系统的稳定性。当控制信号无缓冲阶段时，液压油对换向阀的冲击最大；当存在缓冲阶段时，冲击有所降低，换向缓冲时间由0 s提升至0.2 s时，液压油对换向阀的冲击明显减小。
（2）在液压系统参数不变情况下，随着气缸活塞直径增大，活塞行程逐渐减小，最大速度逐渐增大，且在活塞换向时其速度畸变越明显。当一级气缸活塞直径为130，135，140 mm时，活塞行程分别为260，226，172 mm，换向时的最大速度分别为0.156，0.177，0.199 m/s，换向速度增加约28%，表明液压系统与增压系统间的匹配效果变差。
（3）在氢压机排气阶段，气缸内压力和温度随着气缸活塞直径的增加而减小，气缸内压力最大和最小值分别为30.3 MPa和29.7 MPa，温度最大和最小值分别为390.5 K和382.4 K；气缸内压力随着油缸活塞直径的增加而增大，其最大和最小值分别为31.5 MPa和30.3 MPa，温度最大和最小值分别为394.4 K和390.4 K。
（4）在氢压机进气阶段，气缸内压力随着气缸活塞直径的增加而增大，而油缸活塞直径的变化对气缸内压力的影响相对较小；油缸内压力随着气缸活塞直径的增加而增大，随着油缸活塞直径的增加而减小。排气阶段则与之相反，油缸内压力随着气缸活塞直径的增加而减小，其最大和最小值分别为15.9 MPa和15.23 MPa；油缸内压力随着油缸活塞直径的增加而增大，其最大和最小值分别为15.3 MPa和14.8MPa。

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2025年第53卷第8期
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doi: 10.3969/j.issn.1005-0329.2025.08.004
  • 接收时间：2025-02-15
  • 首发时间：2026-10-20
  • 出版时间：2025-08-30
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  • 收稿日期：2025-02-15
  • 修回日期：2025-06-15
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    合肥通用机械研究院有限公司 高端压缩机及系统技术全国重点实验室，合肥 230031

通讯作者:

王胜利（1990），男，博士，助理研究员，主要从事高压无油往复式压缩机设计技术研究工作，通信地址：230031安徽 省合肥市长江西路888号合肥通用机械研究院有限公司，E-mail：。
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