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Article(id=1319293693757083720, tenantId=1146029695717560320, journalId=1316781260334530699, issueId=1319293542573392487, articleNumber=null, orderNo=null, doi=10.3969/j.issn.1005-0329.2025.07.009, pmid=null, cstr=null, oa=null, hot=null, price=null, onlineType=0, articleFormat=0, articleType=null, articleTypeStr=null, receivedDate=1735056000000, receivedDateStr=2024-12-25, revisedDate=1744819200000, revisedDateStr=2025-04-17, acceptedDate=null, acceptedDateStr=null, onlineDate=1792473727883, onlineDateStr=2026-10-20, pubDate=1753804800000, pubDateStr=2025-07-30, doiRegisterDate=null, doiRegisterDateStr=null, onlineIssueDate=1792473727883, onlineIssueDateStr=2026-10-20, onlineJustAcceptDate=null, onlineJustAcceptDateStr=null, onlineFirstDate=null, onlineFirstDateStr=null, sourceXml=null, magXml=null, createTime=1792473727883, creator=13701087609, updateTime=1792473727883, updator=13701087609, issue=Issue{id=1319293542573392487, tenantId=1146029695717560320, journalId=1316781260334530699, year='2025', volume='53', issue='7', pageStart='1', pageEnd='118', issueExtLink='null', onlineDate='null', pubDate='1753804800000', pubDateStr='2025-07-30', beforeIssueId=null, nextIssueId=null, price=null, status=1, issueComplete=1, articleOrder=1, issueType=-1, specialIssue=null, createTime=1792473691838, creator='13701087609', updateTime=1792473691838, updator='13701087609', preIssue=null, nextIssue=null, articleTotal=null, ext=null, issueFiles=null, downloadFileDto=null}, startPage=61, endPage=68, ext={EN=ArticleExt(id=1319293693950021705, articleId=1319293693757083720, tenantId=1146029695717560320, journalId=1316781260334530699, language=EN, title=The design and analysis of high hub‑tip ratio underground gas storage compressor stages, columnId=null, journalTitle=Fluid Machinery, columnName=null, runingTitle=null, highlight=null, articleAbstract=

A large hub to hub ratio model stage is proposed to address the issue of the existing conventional hub to hub ratio model stage for domestic compressors being unable to meet the rotor stability requirements of the seven stage two‑stage scheme and can only adopt a dual cylinder design.Based on the medium hub‑tip ratio model stage of a double‑cylinder gas storage compressor，the design and development of a high hub‑tip ratio model stage has been completed by combining the forward design method，automatic optimization method and test method.The geometry，macroscopic performance，and internal flow field characteristics of different hub‑tip ratio model stages were also compared and analyzed，and it revealed the flow characteristics of the high hub‑tip ratio model stage.The study shows that the polytropic efficiency of the 0.46 high hub‑tip ratio model stage is 4.3% lower than that of the 0.39 hub‑tip ratio model stage，and the efficiency drop of the impeller part is the largest，accounting for 67.4% of the efficiency drop of the whole model stage.Adjusting the blade profile，the number of blades，and reducing the Mach number inside the impeller and the bladeless diffuser have a greater potential to improve the efficiency of the model stage.The deviation of the efficiency between the CFD and the experimental results is all within 3%，which has strong reference value.The developed model stage hub ratio has been increased by 18%，which solves the rotor stability problem of the seven‑stage two‑section scheme and realizes the design and development of a single‑cylinder gas storage compressor with a small floor area.This study addresses the stability issues of single‑cylinder seven‑stage two‑stage rotors，and fills the research gap in the application of large wheel hub ratio compressors in China，and provides valuable reference significance for the aerodynamic design，development and application of the high hub‑tip ratio model of centrifugal compressors.

, authors=Lixin YAO, Wei PEI, Jian CHENG, Shangming JI, Lin PEI, Xiao TAI, Yongrui SUN, Yang GUO, authorsList=Lixin YAO, Wei PEI, Jian CHENG, Shangming JI, Lin PEI, Xiao TAI, Yongrui SUN, Yang GUO, authorCompany=null, correspAuthors=null, 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=1319293698391789672, articleId=1319293693757083720, tenantId=1146029695717560320, journalId=1316781260334530699, language=CN, title=大轮毂比储气库压缩机模型级开发与分析, columnId=1317484287230853622, journalTitle=流体机械, columnName=设计计算, runingTitle=null, highlight=null, articleAbstract=

针对国产压缩机现有的常规轮毂比模型级无法满足七级两段方案的转子稳定性的要求，只能采用双缸设计的问题，提出了一种大轮毂比模型级。以双缸储气库压缩机中等轮毂比模型级为基础，结合正向设计方法，自动优化方法和试验方法，完成了大轮毂比模型级的设计开发。并对不同轮毂比模型级几何形状、宏观性能以及各部件的内部流场特性进行了对比分析，揭示了大轮毂比模型级的流动特性。研究表明，0.46大轮毂比模型级相比0.39轮毂比模型级多变效率低4.3%，叶轮部分的效率降最大，占整级效率降的67.4%，调节叶轮的叶型、叶片数以及降低叶轮与无叶扩压器内部马赫数对提升模型级效率具有更大的潜力。所开发的模型级轮毂比提高了18%，解决了七级两段方案的转子稳定性问题，实现了占地面积小的单缸储气库压缩机设计与开发。CFD与试验结果的效率偏差均在3%以内，具备较强的参考价值，为离心压缩机的大轮毂比模型级的气动设计、开发与应用提供了宝贵的借鉴意义。

, authors=么立新, 裴威, 程健, 吉尚明, 裴林, 邰骁, 孙永瑞, 郭杨, authorsList=么立新, 裴威, 程健, 吉尚明, 裴林, 邰骁, 孙永瑞, 郭杨, authorCompany=null, correspAuthors=null, authorNote=

么立新（1976），男，硕士，正高级工程师，主要从事离心压缩机模型级气动开发、泵水力模型开发相关工作，通信地址：110869辽宁省沈阳市铁西区开发大道16号甲沈鼓集团股份有限公司，E-mail：。

, correspAuthorsNote=null, 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=UkaSYB//ySd/EJ1srAlKWg==, magXml=VZwoajnT3WM+XH0Ql9xLXw==, pdfUrl=null, pdf=null, pdfFileSize=null, pdfExtLink=null, richHtmlUrl=null, mobilePdfUrl=null, reviewReport=null, pdfFirstPage=null, abstractGraph=l1LZcqjpmTETG7L1bUZugg==, abstractGraphContent=null, abstractVideo=null, citation=null, cebUrl=null, magXmlContent=Re+vk97+KjjmimS4JPD5Ug==, mapNumber=null, fund=null)}, authors=[Author(id=1319293778138092266, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, orderNo=0, firstName=null, middleName=null, lastName=null, nameCn=null, orcid=null, stid=null, country=null, authorPic=null, dead=0, email=yaolixin@shengu.com.cn, emailSecond=null, emailThird=null, correspondingAuthor=0, authorType=1, ext={EN=AuthorExt(id=1319293778217784044, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, authorId=1319293778138092266, language=EN, stringName=Lixin YAO, firstName=Lixin, middleName=null, lastName=YAO, prefix=null, suffix=null, authorComment=null, nameInitials=null, affiliation=null, department=null, xref=null, address=ShenGu Group Co.，Ltd.，Shenyang 110869，China, bio=null, bioImg=null, bioContent=null, aboutCorrespAuthor=null), CN=AuthorExt(id=1319293778297475821, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, authorId=1319293778138092266, language=CN, stringName=么立新, firstName=null, middleName=null, lastName=null, prefix=null, suffix=null, authorComment=null, nameInitials=null, affiliation=null, department=null, xref=null, address=沈鼓集团股份有限公司，沈阳 110869, bio={"content":"

么立新（1976），男，硕士，正高级工程师，主要从事离心压缩机模型级气动开发、泵水力模型开发相关工作，通信地址：110869辽宁省沈阳市铁西区开发大道16号甲沈鼓集团股份有限公司，E-mail：。

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么立新（1976），男，硕士，正高级工程师，主要从事离心压缩机模型级气动开发、泵水力模型开发相关工作，通信地址：110869辽宁省沈阳市铁西区开发大道16号甲沈鼓集团股份有限公司，E-mail：。

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High hub/tip diameter ratio centrifugal compressor data

, figureFileSmall=null, figureFileBig=null, tableContent=
项目机型
PWC PT6RR GemSolar T41GE
ν10.680.500.680.36
D2/mm259.08253.49205.74629.92
转速/(r⋅min-1)23.035.440.815.0
压比2.003.402.852.50
效率/%83797692
), ArticleFig(id=1319293786535088954, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=CN, label=表1, caption=

高轮毂比离心压缩机数据

, figureFileSmall=null, figureFileBig=null, tableContent=
项目机型
PWC PT6RR GemSolar T41GE
ν10.680.500.680.36
D2/mm259.08253.49205.74629.92
转速/(r⋅min-1)23.035.440.815.0
压比2.003.402.852.50
效率/%83797692
), ArticleFig(id=1319293786598003515, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=EN, label=Tab.2, caption=

Parameters of stage

, figureFileSmall=null, figureFileBig=null, tableContent=
参数Hub1Hub2
叶轮进口直径D0/mm186.4208.5
叶轮进口轮毂直径D1h/mm156.0184.0
叶轮叶片进口平均直径D1rms/mm179.6201.9
叶轮出口直径D2/mm400.0400.0
叶轮出口宽度b2/mm10.810.3
叶轮叶片数Z2/片1313
叶片出口平均安装角β2A/（°）-63.5-55.0
叶轮进口面积A0/m20.008 20.007 6
平均叶轮轮毂比ν(D1rms/D2）0.450.50
叶轮轮毂比ν10.390.46
宽径比b2/D20.0270.026
比转速Ns0.370.37
), ArticleFig(id=1319293786660918076, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=CN, label=表2, caption=

模型级参数

, figureFileSmall=null, figureFileBig=null, tableContent=
参数Hub1Hub2
叶轮进口直径D0/mm186.4208.5
叶轮进口轮毂直径D1h/mm156.0184.0
叶轮叶片进口平均直径D1rms/mm179.6201.9
叶轮出口直径D2/mm400.0400.0
叶轮出口宽度b2/mm10.810.3
叶轮叶片数Z2/片1313
叶片出口平均安装角β2A/（°）-63.5-55.0
叶轮进口面积A0/m20.008 20.007 6
平均叶轮轮毂比ν(D1rms/D2）0.450.50
叶轮轮毂比ν10.390.46
宽径比b2/D20.0270.026
比转速Ns0.370.37
), ArticleFig(id=1319293786740609853, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=EN, label=Tab.3, caption=

Boundary conditions

, figureFileSmall=null, figureFileBig=null, tableContent=
参数数值
进口总压力Pin/Pa100 000
进口总温度Tin/K293
出口质量流Gout/（kg⋅s-1）0.492
叶轮转速n/（r⋅min-1）9 831
), ArticleFig(id=1319293786795135806, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=CN, label=表3, caption=

边界条件

, figureFileSmall=null, figureFileBig=null, tableContent=
参数数值
进口总压力Pin/Pa100 000
进口总温度Tin/K293
出口质量流Gout/（kg⋅s-1）0.492
叶轮转速n/（r⋅min-1）9 831
), ArticleFig(id=1319293786858050367, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=EN, label=Tab.4, caption=

Verification results of grid independence

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网格数量/万归一化效率压比
501.021 61.279
751.025 21.281
1001.030 01.283
1301.033 11.284
1601.032 71.283
1901.031 31.283
2301.032 61.283
2601.034 81.284
), ArticleFig(id=1319293786920964928, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=CN, label=表4, caption=

网格无关性验证结果

, figureFileSmall=null, figureFileBig=null, tableContent=
网格数量/万归一化效率压比
501.021 61.279
751.025 21.281
1001.030 01.283
1301.033 11.284
1601.032 71.283
1901.031 31.283
2301.032 61.283
2601.034 81.284
), ArticleFig(id=1319293786975490881, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=EN, label=Tab.5, caption=

Comparison of aerodynamic parameters between Hub1 and Hub2

, figureFileSmall=null, figureFileBig=null, tableContent=
参数Hub1Hub2
级压比ε1.281.28
叶轮出口多变效率ηp20.9260.897
叶轮出口多变效率差△ηp2/%—-2.9
扩压器出口多变效率差△ηp4/%—-4.0
弯道出口多变效率差△ηp5/%—-4.0
回流器出口多变效率差△ηp6/%—-4.2
级出口多变效率差△ηp7/%—-4.3
扩压因子DF0.8200.690
级进口绝对马赫数Ma00.1500.162
叶轮进口叶尖马赫数Ma10.3350.371
叶轮出口绝对马赫数Ma20.3450.380
级出口绝对马赫数Ma70.1420.135
), ArticleFig(id=1319293787050988354, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293693757083720, language=CN, label=表5, caption=

Hub1与Hub2气动参数对比

, figureFileSmall=null, figureFileBig=null, tableContent=
参数Hub1Hub2
级压比ε1.281.28
叶轮出口多变效率ηp20.9260.897
叶轮出口多变效率差△ηp2/%—-2.9
扩压器出口多变效率差△ηp4/%—-4.0
弯道出口多变效率差△ηp5/%—-4.0
回流器出口多变效率差△ηp6/%—-4.2
级出口多变效率差△ηp7/%—-4.3
扩压因子DF0.8200.690
级进口绝对马赫数Ma00.1500.162
叶轮进口叶尖马赫数Ma10.3350.371
叶轮出口绝对马赫数Ma20.3450.380
级出口绝对马赫数Ma70.1420.135
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大轮毂比储气库压缩机模型级开发与分析
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么立新 , 裴威 , 程健 , 吉尚明 , 裴林 , 邰骁 , 孙永瑞 , 郭杨
流体机械 | 设计计算 2025,53(7): 61-68
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流体机械 |设计计算 2025 , 53 (7) : 61 -68
大轮毂比储气库压缩机模型级开发与分析
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么立新（1976），男，硕士，正高级工程师，主要从事离心压缩机模型级气动开发、泵水力模型开发相关工作，通信地址：110869辽宁省沈阳市铁西区开发大道16号甲沈鼓集团股份有限公司，E-mail：。

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么立新（1976），男，硕士，正高级工程师，主要从事离心压缩机模型级气动开发、泵水力模型开发相关工作，通信地址：110869辽宁省沈阳市铁西区开发大道16号甲沈鼓集团股份有限公司，E-mail：。

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么立新 , 裴威, 程健, 吉尚明, 裴林, 邰骁, 孙永瑞, 郭杨
作者信息
  • 沈鼓集团股份有限公司，沈阳 110869
作者简介:

么立新（1976），男，硕士，正高级工程师，主要从事离心压缩机模型级气动开发、泵水力模型开发相关工作，通信地址：110869辽宁省沈阳市铁西区开发大道16号甲沈鼓集团股份有限公司，E-mail：。

The design and analysis of high hub‑tip ratio underground gas storage compressor stages
Lixin YAO , Wei PEI, Jian CHENG, Shangming JI, Lin PEI, Xiao TAI, Yongrui SUN, Yang GUO
Affiliations
  • ShenGu Group Co.，Ltd.，Shenyang 110869，China
出版时间: 2025-07-30 doi: 10.3969/j.issn.1005-0329.2025.07.009
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摘要
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针对国产压缩机现有的常规轮毂比模型级无法满足七级两段方案的转子稳定性的要求，只能采用双缸设计的问题，提出了一种大轮毂比模型级。以双缸储气库压缩机中等轮毂比模型级为基础，结合正向设计方法，自动优化方法和试验方法，完成了大轮毂比模型级的设计开发。并对不同轮毂比模型级几何形状、宏观性能以及各部件的内部流场特性进行了对比分析，揭示了大轮毂比模型级的流动特性。研究表明，0.46大轮毂比模型级相比0.39轮毂比模型级多变效率低4.3%，叶轮部分的效率降最大，占整级效率降的67.4%，调节叶轮的叶型、叶片数以及降低叶轮与无叶扩压器内部马赫数对提升模型级效率具有更大的潜力。所开发的模型级轮毂比提高了18%，解决了七级两段方案的转子稳定性问题，实现了占地面积小的单缸储气库压缩机设计与开发。CFD与试验结果的效率偏差均在3%以内，具备较强的参考价值，为离心压缩机的大轮毂比模型级的气动设计、开发与应用提供了宝贵的借鉴意义。

储气库  /  压缩机  /  模型级  /  轮毂比  /  低比转速

A large hub to hub ratio model stage is proposed to address the issue of the existing conventional hub to hub ratio model stage for domestic compressors being unable to meet the rotor stability requirements of the seven stage two‑stage scheme and can only adopt a dual cylinder design.Based on the medium hub‑tip ratio model stage of a double‑cylinder gas storage compressor，the design and development of a high hub‑tip ratio model stage has been completed by combining the forward design method，automatic optimization method and test method.The geometry，macroscopic performance，and internal flow field characteristics of different hub‑tip ratio model stages were also compared and analyzed，and it revealed the flow characteristics of the high hub‑tip ratio model stage.The study shows that the polytropic efficiency of the 0.46 high hub‑tip ratio model stage is 4.3% lower than that of the 0.39 hub‑tip ratio model stage，and the efficiency drop of the impeller part is the largest，accounting for 67.4% of the efficiency drop of the whole model stage.Adjusting the blade profile，the number of blades，and reducing the Mach number inside the impeller and the bladeless diffuser have a greater potential to improve the efficiency of the model stage.The deviation of the efficiency between the CFD and the experimental results is all within 3%，which has strong reference value.The developed model stage hub ratio has been increased by 18%，which solves the rotor stability problem of the seven‑stage two‑section scheme and realizes the design and development of a single‑cylinder gas storage compressor with a small floor area.This study addresses the stability issues of single‑cylinder seven‑stage two‑stage rotors，and fills the research gap in the application of large wheel hub ratio compressors in China，and provides valuable reference significance for the aerodynamic design，development and application of the high hub‑tip ratio model of centrifugal compressors.

gas storage  /  compressor  /  model stage  /  hub‑tip ratio  /  low specific speed
么立新, 裴威, 程健, 吉尚明, 裴林, 邰骁, 孙永瑞, 郭杨. 大轮毂比储气库压缩机模型级开发与分析. 流体机械, 2025 , 53 (7) : 61 -68 . DOI: 10.3969/j.issn.1005-0329.2025.07.009
Lixin YAO, Wei PEI, Jian CHENG, Shangming JI, Lin PEI, Xiao TAI, Yongrui SUN, Yang GUO. The design and analysis of high hub‑tip ratio underground gas storage compressor stages[J]. Fluid Machinery, 2025 , 53 (7) : 61 -68 . DOI: 10.3969/j.issn.1005-0329.2025.07.009
正文
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0　引言
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随着国家能源需求增长和国际形势变化，地下储气库在能源安全保障中的作用愈发重要，战略储备型储气库建设成为天然气供应安全的重大战略举措。目前，国内储气库地面工程技术已基本成熟，但关键地面装备（尤其是压缩机）仍依赖进口，存在易损件更换频繁、供货周期长、成本高等问题［1］。为推进设备国产化，以双台子储气库为例，其进口机组采用单缸两段式压缩机（配套设备少，占地面积小），国产机组采用分体式双缸压缩机（配套设备多，占地面积大）［2］。国产压缩机采用双缸设计的主要原因：现有模型级轮毂比较小，单缸七级两段方案转子稳定性无法通过。
基于以上现状，目前急需研制高效、宽范围、高能头大轮毂比模型级。轮毂比定义如下：
ν1=D1h/D2
式中，D1h为叶轮进口轮毂直径；D2为叶轮的出口直径。
基于多级离心式清水泵的轮毂比为0.25左右的情况，为了提高多级离心泵运行的可靠性，设计中采用加粗泵轴及研制大轮毂比（0.33~0.40则称为大轮毂比）叶轮的方式，以此提高转子的刚性来实现。根据常规多级离心压缩机的轮毂比为0.33左右与泵进行比例评估，则离心压缩机的大轮毂比范围区间为0.44~0.53。
20世纪40年代，航空燃气轮机的发展主要集中在多级轴流压气机技术的完善。MCDONALD等［3］提出了9级轴流压气机和末级单级离心压气机的回热式涡轴发动机。PWC［4］，RR［5］，Solar［6］，GE［7］等公司对于带有AxC（axial centrifugal）压气机的航空燃气涡轮发动机的研究，带动了大轮毂比离心压缩机叶轮的应用、研究与发展，各机型高轮毂比离心压缩机数据见表1。根据表1中各公司离心压气机中应用的离心叶轮的各项数据与相关折算，可以初步得出轮毂比越小、直径越大、叶轮出口周向速度越低，模型级效率越高的结论。
大轮毂比离心压气机因进、出口半径比大，导致子午流道曲率半径（尤其是轮盖侧）显著减小，易引发叶轮出口沿叶高方向的流动分离。同时，小流道曲率导致罗斯比数降低，加剧叶轮出口低能流体横向迁移，掺混损失增加，性能恶化，高效设计十分困难［8-9］。闫慧慧等［10］研究发现，通过适当降低进口轮毂比及减小叶轮叶根进口角，合理选择叶轮出口相对宽度等，对提升压气机的效率和压比有利；龙双喜等［11］在给定的轴向尺寸约束条件下，提出了离心叶轮和径向扩压器子午流道倾斜的设计方法，提高某高轮毂比离心压气机性能；OANA等［12］通过叶片前缘激波控制、载荷优化等手段，使单级跨音速离心压缩机效率提高1%；高轮毂比导致叶轮进、出口马赫数升高是效率低的主因，可通过增加叶片数量缓解叶片滑移，但会提高制造成本［13］。
现有研究揭示了大轮毂比导致效率下降的共性规律，但缺乏对储气库压缩机叶轮-扩压器耦合流动机制的针对性分析；此外，部分研究存在优化不充分、缺少试验结果支撑等问题。针对单缸储气库压缩机用大轮毂比模型级的开发需求和上述问题，以双缸储气库压缩机的中等轮毂比模型级为基础，采用CFD分析与试验验证相结合的研究方法，开发用于单缸储气库压缩机的大轮毂比模型级，分析两者的结构几何形状、宏观性能，并分析2种模型级进口、叶轮等区域的内部流场的特性，实现理论-试验-工程应用的闭环验证。
1　研究对象
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本文研究对象包括应用于双缸储气库压缩机的中等轮毂比模型级（ν1=0.39，以下简称Hub1）和应用于单缸储气库压缩机中大轮毂比的模型级（ν1=0.46，以下简称Hub2）。2个模型级的部分几何与无量纲参数见表2。
2　研究方法
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为了完成单缸七级储气库离心压缩机中具有大轮毂比的低比转速模型级的开发，采用数值方法开展了大轮毂比模型级的改进与优化设计，并通过试验方法完成了模型级的宏观性能验证。
2.1　数值方法
储气库模型级的数值计算与分析采用NUMECA软件的Fine/Turbo完成，采用稳态雷诺平均Navier-Stokes方程，湍流模型选取S-A方程，多重网格迭代求解。
计算域包括进口延长段、叶轮、无叶扩压器、弯道、回流器及出口延长段，如图1所示。
数值计算采用AutoGrid5生成的单通道六面体网格，壁面第一层网格尺寸为0.01 mm，网格节点数189万，模型级的全通道网格如图2所示。
边界条件给定进口总温、总压，出口给定质量流量，叶轮部分采用面积定义法，具体见表3。
计算得到的y+分布如图3所示，整级的y+值均小于4，满足S-A湍流模型对y+的要求。
2.2　网格无关性验证
研究初期完成了网格无关性验证，网格节点数为50万~260万，表4列出网格无关性验证结果。可以看出，网格数50万~100万时，效率与压比逐渐上升；网格数100~260万时，效率和能头趋于稳定。综合考虑网格数量及计算结果，本文选取效率结果相对保守的190万方案，开展后续的CFD计算与优化。
2.3　试验方法
离心式压缩机试验装置如图4所示。
测试压缩机通过齿轮箱由一台315 kW的直流电机驱动。质量流量由位于进气管中的进口集流器测量，并由出口管处的2个阀门控制。入口温度和压力测量分别采用2根玻璃管温度计与4个单点总压探针，出口温度和压力测量分别采用2根玻璃管温度计与3个单点加1个梳齿总压探针，其中叶轮实物图见图5。总压探针位于级出口平直段，在管路中测量进、出口总温度，进口总压探针位于进口管路中。
3　研究内容
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3.1　设计方法
大轮毂比模型级的设计开发过程主要包括采用正问题设计方法的初步设计和自动优化两部分。
初步设计以Hub1模型级为基础，根据此模型级轮毂比0.46的几何约束要求，使用Concepts-NREC软件的Axcent模块完成模型级进、出口轮毂直径的修改，并根据进、出口面积相同的原则，完成模型级进、出口叶顶侧直径的计算与修改。此外，轮毂比增大后，模型级进、出口处子午型线径向尺寸明显增大，与原有的叶轮叶片前缘和回流器叶片出口位置并不匹配，需要进行相应的修改，其中，叶轮叶片前缘部分依据Z值不变的原则完成修改，回流器尾缘线根据调整后的子午型线进行适当地抬高，具体见图6。
基于Fine/Turbo，Concepts-NREC与MINAMO组建的全三维可视化优化设计平台，通过对模型级叶轮、回流器等结构的多轮优化，获得了大轮毂比模型级方案Hub2。
3.2　数值结果对比分析
为了更清晰地了解轮毂比从模型级Hub1的0.39增大到Hub2的0.46对宏观性能的影响，对比分析2种轮毂比模型级的整级和从入口到不同位置的多变效率。多变效率计算公式如下：
ηp=1k/(k-1)lg(P/Pin)lg(T/Tin)
式中，ηp为多变效率；k为绝热指数；P，T分别为总压和总温。
表5列出Hub1与Hub2气动参数对比。可以看出，Hub1，Hub2两个模型级CFD结果的叶轮出口多变效率分别为0.926，0.897；Hub2与Hub1效率差为-2.9%（占整级效率差67.4%）；扩压器出口多变效率差为-4.0%（占整级效率差93%，其中扩压器部分占整级效率差25.6%），弯道出口多变效率差为-4.0%，回流器出口、级出口的多变效率差略微增加，整级效率差达到-4.3%。
图7示出Hub1，Hub2两者归一化处理后的多变效率η（η=ηp/ηpd，其中，ηp为实际多变效率，ηpd为多变效率设计值）与压比ε的变工况曲线。Hub2与Hub1的曲线相比有以下特点：（1）Hub2多变效率整体低于Hub1，最高效率点与Hub1均位于110%工况点；（2）两者压比曲线在120%工况左侧非常接近，Hub2略高，120%工况右侧Hub2低于Hub1；（3）Hub2的曲线范围相比Hub1约窄了15%，其中，左侧约窄5%，右侧约窄10%。
3.3　内部流动特性分析
（1）在级压比、叶轮外径、D4/D2（D4为扩压器出口截面半径）与叶片数保持不变，以及尽量保证工况范围的约束条件下，随叶轮平均轮毂比的增大，Hub2方案的叶轮叶片径向尺寸显著缩短、轮盖侧流道的曲率半径显著减小［10］，结构尺寸相比Hub1存在进口面积A0和b2略小、β2A略大的问题，具体见表2。
（2）尽管Hub2叶轮进口直径D0位置马赫数均匀性优于Hub1，主流区翼展方向马赫数曲线斜率（约-0.036）也低于Hub1（-0.041），但受轮毂比增大及A0减小的影响，Hub2的(D0-D1h)/2=12.25 mm，仅约为Hub1的4/5，所以其叶轮进口处的相对马赫数仍显著高于Hub1，其叶轮的进口入射损失更大［12］，具体见图8。
（3）受叶片进口尺寸的影响，Hub2叶片进口直径D1位置盖侧叶尖马赫数显著高于Hub1，轮毂侧则显著偏低，均匀性显著劣于Hub1，具体见图8、图9和表5。
（4）Hub2，Hub1的扩压因子DF［14］分别为0.690，0.820，这个比值也被称为德哈勒比率［15］，其取值在0.7~0.95范围区间是被接受的。基于此种考虑，Hub2没有处于最佳取值范围内，所以其性能低于Hub1。
（5）扩压器出口效率差为-4.0%，整个扩压器产生约1.1%效率降，这主要是Hub2叶轮出口绝对马赫数值显著高于Hub1，此类低比转速模型级摩擦损失占比较大所致，具体见图10、图11。
（6）Hub2与Hub1弯道出口多变效率差与无叶扩压器出口相同，这是因为Hub2的无叶扩压器充分扩张后，其出口马赫数已经与Hub1很接近，弯道出口更是基本一致，具体见图10。
（7）回流器出口效率差仅增大0.2个百分点，这主要是由于轮毂比从Hub1的0.39增大到Hub2的0.46后，回流器叶片径向尺寸与叶片弦长显著缩短，在此类模型级具有回流器大转折角的情况下，其导流效果相比Hub1明显不足，导致回流器出口沿周向绝对马赫数分布均匀性劣于Hub1，参混损失增大，具体见图10与图12。
（8）Hub2回流器出口至级出口效率差仅增大0.1个百分点，这主要是回流器出口均匀性差，并一直影响到级出口，参混损失高于Hub1所致，具体见图12、图13。
3.4　试验验证
对原有Hub1与新开发的Hub2的试验数据进行分析，其结果对比如图14所示。
从图14（a）可以看出，两模型级的试验与CFD曲线趋势比较相近，100%工况时，两者试验与CFD的效率偏差均在3%以内。同时具有以下特点：
（1）Hub2多变效率整体上低于Hub1，最高效率点位于小流量区，而Hub1最高效率点位于曲线中部区域；
（2）两者压比曲线比较接近，在流量低于150%时，Hub2略高于Hub1；
（3）相比Hub1，Hub2的曲线范围约窄了18%，其中，左侧约窄14%，右侧约窄4%。
由于试验台位的进、出口均为大气，进口压力与温度随天气变化，在本研究中，2个模型级的试验进口压力范围在100~101.7 kPa之间，比CFD高0~1.7%；进口温度范围在294.9~299.0 K之间，比CFD高约0.6%~2.0%。Hub2与Hub1的试验与CFD的出口压力（Pout）与出口温度（Tout）对比图见图14（b）。可以看出，出口压力与CFD趋势符合较好，结果比较接近；由于试验进口温度较高，出口温度试验也均高于CFD。此外，Hub2出口温度明显高于Hub1，与其效率低相匹配。
3.5　产品应用
基于本文开发的大轮毂比模型级设计的单缸七级两段方案已成功应用于某储气库压缩机中。相比于常规双缸设计方案，在满足多种压比工况条件下的采气工况、注气工况的同时，该方案占地面积更小，成本更低，实现了单缸双段储气库离心压缩机的国产化。应用现场如图15所示。
4　结论
收起
（1）大轮毂比0.46模型级相比轮毂比0.39模型级多变效率低4.3%，叶轮部分的效率降最大，占整级效率降的67.4%，无叶扩压器部分效率降占整级效率降的25.6%，两者均是受马赫数增大所致，两者效率降占比总计约93.0%，所以后续此种低比转速、大轮毂比模型级的优化，可以重点放在叶轮的叶型、叶片数，以及降低叶轮与无叶扩压器内部马赫数等方面。
（2）大轮毂比0.46模型级的扩压因子DF值为0.690，与最佳取值范围0.7~0.95很接近，后期可以考虑从调整DF值方面入手，对此模型级开展进一步优化设计，提升其效率。
（3）2个模型级的CFD与试验的多变效率与压比曲线在总体趋势上吻合较好，在100%工况时，CFD与试验结果的效率偏差均在3%以内，说明CFD结果比较准确，具备较强的参考价值，可为后续大轮毂比模型级开发研究提供借鉴。

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doi: 10.3969/j.issn.1005-0329.2025.07.009
  • 接收时间：2024-12-25
  • 首发时间：2026-10-20
  • 出版时间：2025-07-30
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  • 收稿日期：2024-12-25
  • 修回日期：2025-04-17
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鹅膏菌科Amanitaceae 2 11 5.26 鹅膏菌属 Amanita 10 4.78
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