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Article(id=1217836024105976730, tenantId=1146029695717560320, journalId=1210938733613449225, issueId=1217836019408360416, articleNumber=null, orderNo=null, doi=10.19666/j.rlfd.202501049, pmid=null, cstr=null, oa=null, hot=null, price=null, onlineType=0, articleFormat=0, articleType=null, articleTypeStr=null, receivedDate=1737561600000, receivedDateStr=2025-01-23, revisedDate=null, revisedDateStr=null, acceptedDate=null, acceptedDateStr=null, onlineDate=1768284334446, onlineDateStr=2026-01-13, pubDate=1764000000000, pubDateStr=2025-11-25, doiRegisterDate=null, doiRegisterDateStr=null, onlineIssueDate=1768284334446, onlineIssueDateStr=2026-01-13, onlineJustAcceptDate=null, onlineJustAcceptDateStr=null, onlineFirstDate=null, onlineFirstDateStr=null, sourceXml=null, magXml=null, createTime=1768284334446, creator=13701087609, updateTime=1768284334446, updator=13701087609, issue=Issue{id=1217836019408360416, tenantId=1146029695717560320, journalId=1210938733613449225, year='2025', volume='54', issue='11', pageStart='1', pageEnd='168', issueExtLink='null', onlineDate='null', pubDate='null', beforeIssueId=null, nextIssueId=null, price=null, status=1, issueComplete=1, articleOrder=1, issueType=-1, specialIssue=null, createTime=1768284333326, creator=13701087609, updateTime=1768284453982, updator=13701087609, preIssue=null, nextIssue=null, ext={EN=IssueExt(id=1217836525543408117, tenantId=1146029695717560320, journalId=1210938733613449225, issueId=1217836019408360416, language=EN, specialIssueTitle=, coverIllustrator=null, specialIssueEditor=, specialIssueAbout=), CN=IssueExt(id=1217836525543408118, tenantId=1146029695717560320, journalId=1210938733613449225, issueId=1217836019408360416, language=CN, specialIssueTitle=, coverIllustrator=null, specialIssueEditor=, specialIssueAbout=)}, issueFiles=null}, startPage=1, endPage=11, ext={EN=ArticleExt(id=1217836024361829287, articleId=1217836024105976730, tenantId=1146029695717560320, journalId=1210938733613449225, language=EN, title=Integration and performance analysis of high back pressure gas turbine combined cycle power generation system, columnId=1217836020075254754, journalTitle=Thermal Power Generation, columnName=Advanced power cycle technology, runingTitle=null, highlight=null, articleAbstract=

Based on the concept of high efficiency of gas turbine variable back pressure operation regulation, a high back pressure gas turbine combined cycle power generation system scheme is proposed, in which a pre-compressor and an expander are set in front and behind the main top cycle respectively to maintain and regulate the gas turbine exhaust pressure. Key parameters of the combined cycle are designed based on the initial parameters of the F-class gas turbine, and the case and characteristic analysis are carried out for the temperature of recirculated gas (divided into two conditions: cooling to normal temperature and not cooling), the main top cycle pressure ratio and the gas turbine back pressure. The results show that, the combined cycle efficiency of the recirculated flue gas cooling is not as high as that of the non-cooled flue gas cooling, which is 58.07% and 58.94% when the turbine back pressure is 0.30 MPa. The exergy loss rate of the main compressor is lower because the exit temperature of the main compressor is higher when the recirculated flue gas temperature is higher. When the gas turbine back pressure is 0.30 MPa, the maximum pressure ratio of the combined cycle system efficiency is 17.0, the corresponding combined cycle efficiency is 58.97%, and the specific work is 563.87 kJ/kg. Considering the specific work comprehensively, the recommended main top pressure ratio is 15.4, and when the turbine back pressure is from 0.03 MPa to 0.35 MPa, the variation range of the combined circulation efficiency under the two conditions is about 56.00%~58.57% and 55.81%~59.12%, respectively, which increases at first and then decreases, and the variation range of the combined efficiency is not large at high back pressure. At the same time, based on the practical engineering application, the design of a single waste heat boiler is considered, and its thermal characteristics and possible flexible, low-cost and efficient utilization of renewable energy are analyzed, which provides a new system scheme reference for the flexible and efficient modern combined cycle with multi-energy complementarities.

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针对燃气轮机（燃机）变背压运行调节具有高变工况效率的特性，在构建联合循环时主动提高设计工况时燃机的背压，提出了一种高背压燃机联合循环发电系统方案，即在主顶循环之前和之后分别设置前置压气机和膨胀机用于保持和调节燃机的排气压力。基于F级燃机透平初参数设计了此联合循环的关键参数，针对关键参数再循环烟气的温度（分为冷却到常温和不冷却2种情况）、主顶循环压比和燃机背压开展了案例和特性规律分析。分析结果表明，再循环烟气冷却时的联合循环效率没有不冷却时高，在透平背压为0.30 MPa时，其分别为58.07%和58.94%，主要原因为再循环烟气温度较高时主压气机的出口温度较高，燃烧室㶲损失较低。燃机背压为0.30 MPa时联合循环系统效率最佳压比为17.0，其对应的联合循环效率为58.97%，比功为563.87 kJ/kg。综合考虑比功推荐主顶压比为15.4，其透平背压在0.03~0.35 MPa时，冷却到常温和不冷却2种情况下联合循环效率变化区间分别在56.00%~58.57%和55.81%~59.12%，均为先上升后降低，高背压时联合效率变化范围不大。同时，基于实际工程应用，考虑了单一余热锅炉的设计情况，分析了其热力特性和可能的可再生能源灵活低成本高效利用方式，可为多能互补灵活高效的现代化联合循环提供新的系统方案参考。

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张国强（1980），男，副教授，主要研究方向为能源动力集成及其优化，。
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薛鹏（1996），男，硕士研究生，主要研究方向为燃气蒸汽联合循环系统集成及优化，。

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薛鹏（1996），男，硕士研究生，主要研究方向为燃气蒸汽联合循环系统集成及优化，。

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薛鹏（1996），男，硕士研究生，主要研究方向为燃气蒸汽联合循环系统集成及优化，。

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combined cycle efficiency and solar power efficiency with the flue gas temperature at inlet of the expander, figureFileSmall=5Fv3OEWhUucp6tsOzuO76Q==, figureFileBig=Onb6jx+FQqt/j3ZCNUJdKA==, tableContent=null), ArticleFig(id=1217836036521116150, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=CN, label=图11, caption=联合循环效率和太阳能发电效率随膨胀机入口烟气温度变化曲线, figureFileSmall=5Fv3OEWhUucp6tsOzuO76Q==, figureFileBig=Onb6jx+FQqt/j3ZCNUJdKA==, tableContent=null), ArticleFig(id=1217836036621779452, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=EN, label=Tab.1, caption=

Main parameters of the gas turbine combined cycle unit

, figureFileSmall=null, figureFileBig=null, tableContent=
项目常规燃机联合循环机组（PG9351FA机组）高背压燃机联合循环机组
燃料低位发热量/(kJ·kg–1)48 68548 685
压损/%燃烧室3.5，主余热锅炉1.16，副余热锅炉0.98，冷却器0.25
其他损失/%机械损失1，发电机损失1，燃烧室热损失0.5
环境条件及参数15 ℃ 101.3 kPa，60%RH
压气机设计参数压气机效率前置压气机0.900，第一段0.888，第二、三、四段0.920
冷却空气比例/%18.4518.45
膨胀机设计参数膨胀机出口压力/kPa103.4103.4
膨胀机效率0.90.9
透平设计参数透平效率第一、二段0.900、第三段0.908
蒸汽轮机效率高压缸0.87，中压缸0.91，低压缸0.89
余热锅炉节点温差/℃高/中/低压分别为30/25/25
), ArticleFig(id=1217836036722442753, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=CN, label=表1, caption=

燃机联合循环机组主要参数

, figureFileSmall=null, figureFileBig=null, tableContent=
项目常规燃机联合循环机组（PG9351FA机组）高背压燃机联合循环机组
燃料低位发热量/(kJ·kg–1)48 68548 685
压损/%燃烧室3.5，主余热锅炉1.16，副余热锅炉0.98，冷却器0.25
其他损失/%机械损失1，发电机损失1，燃烧室热损失0.5
环境条件及参数15 ℃ 101.3 kPa，60%RH
压气机设计参数压气机效率前置压气机0.900，第一段0.888，第二、三、四段0.920
冷却空气比例/%18.4518.45
膨胀机设计参数膨胀机出口压力/kPa103.4103.4
膨胀机效率0.90.9
透平设计参数透平效率第一、二段0.900、第三段0.908
蒸汽轮机效率高压缸0.87，中压缸0.91，低压缸0.89
余热锅炉节点温差/℃高/中/低压分别为30/25/25
), ArticleFig(id=1217836036810523143, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=EN, label=Tab.2, caption=

Design parameters and thermodynamic calculation results of the gas turbine combined cycle units

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项目常规联合循环机组方案1机组（0.10 MPa背压）方案2机组（0.10 MPa背压）
燃料流量/(kg·s–1)14.331 314.331 314.331 3
透平进口初温/℃1 3271 3271 327
主压气机压比15.415.415.4
进口空气流量/(kg·s–1)639.94288.81288.81
再循环烟气流量/(kg·s–1)0338.29352.02
压气机出口温度/℃390.96396.84458.83
透平排烟温度/℃611.42627.95637.53
燃机出功/MW260.975254.122243.112
燃机效率/%37.3936.4134.83
蒸汽轮机出功/MW147.105153.84166.68
联合循环出功/MW406.495406.27407.95
联合循环效率/%58.2458.2158.45
), ArticleFig(id=1217836036890214923, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=CN, label=表2, caption=

燃机联合循环机组设计参数及热力计算结果对比

, figureFileSmall=null, figureFileBig=null, tableContent=
项目常规联合循环机组方案1机组（0.10 MPa背压）方案2机组（0.10 MPa背压）
燃料流量/(kg·s–1)14.331 314.331 314.331 3
透平进口初温/℃1 3271 3271 327
主压气机压比15.415.415.4
进口空气流量/(kg·s–1)639.94288.81288.81
再循环烟气流量/(kg·s–1)0338.29352.02
压气机出口温度/℃390.96396.84458.83
透平排烟温度/℃611.42627.95637.53
燃机出功/MW260.975254.122243.112
燃机效率/%37.3936.4134.83
蒸汽轮机出功/MW147.105153.84166.68
联合循环出功/MW406.495406.27407.95
联合循环效率/%58.2458.2158.45
), ArticleFig(id=1217836036969906703, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=EN, label=Tab.3, caption=

Thermodynamic calculation results at typical operating points of the high back pressure gas turbine combined cycle units

, figureFileSmall=null, figureFileBig=null, tableContent=
项目工况点1工况点2工况点3
方案1方案2方案1方案2方案1方案2
透平背压/MPa0.030.030.100.100.300.30
进口空气流量/(kg·s–1)288.81288.81288.81288.81288.81288.81
再循环烟气流量/(kg·s–1)345.11349.46338.29352.02415.79421.50
透平排烟温度/℃625.74637.48627.95637.53632.00643.43
压气机出口温度/℃397.96458.26396.84458.83496.94551.77
燃机出功/MW238.40223.78254.12243.11265.31260.14
燃机效率/%34.1632.0636.4134.8338.0137.27
主余热锅炉节点温差/℃（高/中/低）22.2/17.2/17.221.5/16.5/16.530.0/25.0/25.030.0/25.0/25.021.7/16.7/16.721.4/16.4/16.4
副余热锅炉节点温差/℃高/中/低压分别为30.0/25.0/25.0
蒸汽轮机出功/MW154.17167.6153.84166.68141.46152.82
联合循环出功/MW390.88389.52406.27407.95405.33411.39
比功/(kJ·kg–1)616.61610.28647.86636.6575.26579.17
联合循环效率/%56.0055.8158.2158.4558.0758.94
), ArticleFig(id=1217836037087347220, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=CN, label=表3, caption=

高背压燃机联合循环机组典型工况点热力计算结果对比

, figureFileSmall=null, figureFileBig=null, tableContent=
项目工况点1工况点2工况点3
方案1方案2方案1方案2方案1方案2
透平背压/MPa0.030.030.100.100.300.30
进口空气流量/(kg·s–1)288.81288.81288.81288.81288.81288.81
再循环烟气流量/(kg·s–1)345.11349.46338.29352.02415.79421.50
透平排烟温度/℃625.74637.48627.95637.53632.00643.43
压气机出口温度/℃397.96458.26396.84458.83496.94551.77
燃机出功/MW238.40223.78254.12243.11265.31260.14
燃机效率/%34.1632.0636.4134.8338.0137.27
主余热锅炉节点温差/℃（高/中/低）22.2/17.2/17.221.5/16.5/16.530.0/25.0/25.030.0/25.0/25.021.7/16.7/16.721.4/16.4/16.4
副余热锅炉节点温差/℃高/中/低压分别为30.0/25.0/25.0
蒸汽轮机出功/MW154.17167.6153.84166.68141.46152.82
联合循环出功/MW390.88389.52406.27407.95405.33411.39
比功/(kJ·kg–1)616.61610.28647.86636.6575.26579.17
联合循环效率/%56.0055.8158.2158.4558.0758.94
), ArticleFig(id=1217836037175427607, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=EN, label=Tab.4, caption=

Exergic calculation results

, figureFileSmall=null, figureFileBig=null, tableContent=
项目常规燃机联合循环机组方案1机组（0.10 MPa背压）方案2机组（0.10 MPa背压）
㶲/MW比例/%㶲/MW比例/%㶲/MW比例/%
输入㶲天然气726.0799.98726.0799.99726.0799.99
空气0.110.020.050.010.050.01
输入㶲总和726.18100.00726.12100.00726.12100.00
㶲损失压气机15.132.0814.822.0415.712.16
燃烧室190.3226.21188.1025.90179.2424.69
透平17.722.4417.222.3718.052.49
余热锅炉33.344.5935.144.8438.295.27
蒸汽轮机13.871.9114.451.9915.622.15
凝汽器8.361.1512.921.789.321.28
余热锅炉排烟19.472.6815.942.2018.242.51
其他㶲损21.482.9621.262.9323.703.26
㶲损失总和319.6844.02319.8544.05318.1743.82
输出㶲净输出功406.5055.98406.2755.95407.9556.18
共计726.18100.00726.12100.00726.12100.00
), ArticleFig(id=1217836037284479516, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=CN, label=表4, caption=

㶲计算结果

, figureFileSmall=null, figureFileBig=null, tableContent=
项目常规燃机联合循环机组方案1机组（0.10 MPa背压）方案2机组（0.10 MPa背压）
㶲/MW比例/%㶲/MW比例/%㶲/MW比例/%
输入㶲天然气726.0799.98726.0799.99726.0799.99
空气0.110.020.050.010.050.01
输入㶲总和726.18100.00726.12100.00726.12100.00
㶲损失压气机15.132.0814.822.0415.712.16
燃烧室190.3226.21188.1025.90179.2424.69
透平17.722.4417.222.3718.052.49
余热锅炉33.344.5935.144.8438.295.27
蒸汽轮机13.871.9114.451.9915.622.15
凝汽器8.361.1512.921.789.321.28
余热锅炉排烟19.472.6815.942.2018.242.51
其他㶲损21.482.9621.262.9323.703.26
㶲损失总和319.6844.02319.8544.05318.1743.82
输出㶲净输出功406.5055.98406.2755.95407.9556.18
共计726.18100.00726.12100.00726.12100.00
), ArticleFig(id=1217836037385142818, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=EN, label=Tab.5, caption=

Thermal calculation results of the single waste heat boiler units and double-waste-heat-boiler units

, figureFileSmall=null, figureFileBig=null, tableContent=
项目/单位单余热锅炉A单余热锅炉B双余热锅炉
燃料流量/(kg·s–1)14.331 314.331 314.331 3
透平进口初温/℃1 3271 3271 327
主压气机压比15.415.415.4
进口空气流量/(kg·s–1)288.81288.81288.81
再循环烟气流量/(kg·s–1)422.54422.54421.50
透平排烟温度/℃643.51643.51643.43
前置压气机耗功/MW34.0634.0634.06
膨胀机做功/MW28.4836.0069.86
燃机出功/MW224.08224.08224.34
蒸汽轮机出功/MW187.65187.65152.82
联合循环出功/MW404.08411.59411.39
联合循环效率/%57.8958.9758.94
), ArticleFig(id=1217836037464834600, tenantId=1146029695717560320, journalId=1210938733613449225, articleId=1217836024105976730, language=CN, label=表5, caption=

单余热锅炉机组和双余热锅炉机组热力计算结果

, figureFileSmall=null, figureFileBig=null, tableContent=
项目/单位单余热锅炉A单余热锅炉B双余热锅炉
燃料流量/(kg·s–1)14.331 314.331 314.331 3
透平进口初温/℃1 3271 3271 327
主压气机压比15.415.415.4
进口空气流量/(kg·s–1)288.81288.81288.81
再循环烟气流量/(kg·s–1)422.54422.54421.50
透平排烟温度/℃643.51643.51643.43
前置压气机耗功/MW34.0634.0634.06
膨胀机做功/MW28.4836.0069.86
燃机出功/MW224.08224.08224.34
蒸汽轮机出功/MW187.65187.65152.82
联合循环出功/MW404.08411.59411.39
联合循环效率/%57.8958.9758.94
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高背压燃气轮机联合循环发电系统的集成与性能分析
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薛鹏 1 , 高四艺 1 , 杨雨 1 , 冯志恒 1 , 李永毅 2 , 张国强 1
热力发电 | 先进动力循环技术 2025,54(11): 1-11
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热力发电 | 先进动力循环技术 2025, 54(11): 1-11
高背压燃气轮机联合循环发电系统的集成与性能分析
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薛鹏1 , 高四艺1, 杨雨1, 冯志恒1, 李永毅2, 张国强1
作者信息
  • 1.华北电力大学能源动力与机械工程学院，北京　102206
  • 2.华北电力大学动力工程系，河北　保定　071003

    • 薛鹏（1996），男，硕士研究生，主要研究方向为燃气蒸汽联合循环系统集成及优化，。

通讯作者:

张国强（1980），男，副教授，主要研究方向为能源动力集成及其优化，。
Integration and performance analysis of high back pressure gas turbine combined cycle power generation system
Peng XUE1 , Siyi GAO1, Yu YANG1, Zhiheng FENG1, Yongyi LI2, Guoqiang ZHANG1
Affiliations
  • 1.Department of Energy, Power and Mechanical Engineering, North China Electric Power University, Beijing 102206, China
  • 2.School of Energy Power and Mechanical Engineering, North China Electric Power University, Baoding 071003, China
出版时间: 2025-11-25 doi: 10.19666/j.rlfd.202501049
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针对燃气轮机（燃机）变背压运行调节具有高变工况效率的特性，在构建联合循环时主动提高设计工况时燃机的背压，提出了一种高背压燃机联合循环发电系统方案，即在主顶循环之前和之后分别设置前置压气机和膨胀机用于保持和调节燃机的排气压力。基于F级燃机透平初参数设计了此联合循环的关键参数，针对关键参数再循环烟气的温度（分为冷却到常温和不冷却2种情况）、主顶循环压比和燃机背压开展了案例和特性规律分析。分析结果表明，再循环烟气冷却时的联合循环效率没有不冷却时高，在透平背压为0.30 MPa时，其分别为58.07%和58.94%，主要原因为再循环烟气温度较高时主压气机的出口温度较高，燃烧室㶲损失较低。燃机背压为0.30 MPa时联合循环系统效率最佳压比为17.0，其对应的联合循环效率为58.97%，比功为563.87 kJ/kg。综合考虑比功推荐主顶压比为15.4，其透平背压在0.03~0.35 MPa时，冷却到常温和不冷却2种情况下联合循环效率变化区间分别在56.00%~58.57%和55.81%~59.12%，均为先上升后降低，高背压时联合效率变化范围不大。同时，基于实际工程应用，考虑了单一余热锅炉的设计情况，分析了其热力特性和可能的可再生能源灵活低成本高效利用方式，可为多能互补灵活高效的现代化联合循环提供新的系统方案参考。

高背压  /  燃气轮机  /  联合循环  /  循环效率  /  压气机  /  系统集成与优化

Based on the concept of high efficiency of gas turbine variable back pressure operation regulation, a high back pressure gas turbine combined cycle power generation system scheme is proposed, in which a pre-compressor and an expander are set in front and behind the main top cycle respectively to maintain and regulate the gas turbine exhaust pressure. Key parameters of the combined cycle are designed based on the initial parameters of the F-class gas turbine, and the case and characteristic analysis are carried out for the temperature of recirculated gas (divided into two conditions: cooling to normal temperature and not cooling), the main top cycle pressure ratio and the gas turbine back pressure. The results show that, the combined cycle efficiency of the recirculated flue gas cooling is not as high as that of the non-cooled flue gas cooling, which is 58.07% and 58.94% when the turbine back pressure is 0.30 MPa. The exergy loss rate of the main compressor is lower because the exit temperature of the main compressor is higher when the recirculated flue gas temperature is higher. When the gas turbine back pressure is 0.30 MPa, the maximum pressure ratio of the combined cycle system efficiency is 17.0, the corresponding combined cycle efficiency is 58.97%, and the specific work is 563.87 kJ/kg. Considering the specific work comprehensively, the recommended main top pressure ratio is 15.4, and when the turbine back pressure is from 0.03 MPa to 0.35 MPa, the variation range of the combined circulation efficiency under the two conditions is about 56.00%~58.57% and 55.81%~59.12%, respectively, which increases at first and then decreases, and the variation range of the combined efficiency is not large at high back pressure. At the same time, based on the practical engineering application, the design of a single waste heat boiler is considered, and its thermal characteristics and possible flexible, low-cost and efficient utilization of renewable energy are analyzed, which provides a new system scheme reference for the flexible and efficient modern combined cycle with multi-energy complementarities.

high back pressure  /  gas turbine  /  combined cycle  /  cycle efficiency  /  compressor  /  system integration and optimization
薛鹏, 高四艺, 杨雨, 冯志恒, 李永毅, 张国强. 高背压燃气轮机联合循环发电系统的集成与性能分析. 热力发电, 2025 , 54 (11) : 1 -11 . DOI: 10.19666/j.rlfd.202501049
Peng XUE, Siyi GAO, Yu YANG, Zhiheng FENG, Yongyi LI, Guoqiang ZHANG. Integration and performance analysis of high back pressure gas turbine combined cycle power generation system[J]. Thermal Power Generation, 2025 , 54 (11) : 1 -11 . DOI: 10.19666/j.rlfd.202501049
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随着全球经济的稳步增长，能源和环境问题日益严峻，为实现“双碳”战略目标，化石能源优化利用是关键，可再生能源规模化利用是根本[1]。截至2024年，我国火电机组占总装机容量比例降至35.7%，占总发电量63.2%，并网风电和光伏发电占总装机容量比例为42.1%，占总发电量18.2%[2]。可再生能源发电因其波动性和不可预测性，需要火电为电力系统提供调峰保障[3]。燃气轮机（燃机）联合循环系统相较于燃煤机组，具有效率高、爬坡速率高和启动速度快的特点，但在负荷下降时燃机的运行性能下降，尤其在70%负荷以下下降时，运行性能下降更明显[4-5]。因此，对燃机联合循环系统的设计和运行优化及燃机联合循环系统全工况热力性能的研究非常重要。
对于燃机联合循环系统，热力性能直接影响电厂的天然气消耗量，从而影响发电成本。因此许多学者考虑从流程和参数方面优化来提高热力性能。为了最大限度地利用燃机的废热，对燃机联合循环系统的多种配置进行参数分析，得到三压再热余热锅炉具有最高的整体功率和热效率，并得到效率随压比的变化关系[6]；通过建立底循环参数优化的热工模型，对主蒸汽压力/温度进行优化计算，并对实际蒸汽系统的偏差因素进行定量分析，优化了汽、水参数，获得较好的联合循环效率[7-8]；利用所推导的余热锅炉排汽温度计算关系，运用数值分析的方法对各级蒸汽以及再热余热锅炉的压力、温度进行优化[9]；通过MATLAB建立双压再热余热锅炉模型，分析计算不同热力参数对其的影响[10]；文献[11]提出了一种新的优化算法，研究了节点温差、接近点温差和余热锅炉压力与联合循环效率的关系，结果表明节点温差和接近点温差的增加将导致联合循环效率降低。
不同于上述研究，Allam等人提出了一种以超临界CO2为工质、高压、低压比回热布雷顿循环，即Allam循环。Allam循环在工作压力30 MPa、透平进口温度1 100 ℃时的供电效率，比现有的F级燃机天然气联合循环高2.3百分点[12-14]；结合再热装置提出一种再热Allam循环，对再热Allam循环进行热力分析和优化，结果表明再热Allam循环的效率降低5%，但是净功率输出和比功为传统Allam循环的2.1和2.2倍[15]。
为了提升联合循环变工况下的效率，越来越多的学者开始进行联合循环系统结构上的创新和运行策略的优化。针对压气机进口温度参数调控，通过提升进气温度从而实现流量调节方式，可以提升变工况效率。如基于余热加热压气机进气来降低压气机进气流量，提出了一种带有进气加热的燃机联合循环系统[16]，分析显示可显著提高部分负荷效率。烟气回注运行策略将一部分烟气重新注入环境空气中来提高压缩机入口温度，以降低涡轮机入口温度操作策略作为辅助手段[17]，研究表明这种方式可显著提高冷热电联产系统的效率和节能潜力。通过利用烟气再循环燃机系统中余热锅炉排放烟气的低温热量，可提高整体系统全工况的效率并节省燃料[18]。
同样，改变进气压力以降低流量，也可以提升变工况效率。如在燃机循环的上游设置一台不与燃机共轴的辅助压气机为主压气机提供合适的进气压力和空气流量[19]，分析显示该方法可以提高联合循环的灵活性和效率。燃机进气节流（inet air throttle，IAT）控制方法相比于纯燃料控制策略具有更好的运行经济性。然而，考虑燃机通流匹配，变工况时压气机流量的减少必然带来压比降低，因此将会导致排烟温度过高或高强度的燃料调节。通过背压调节，可以避免这一问题[20]。基于余热锅炉的出口设置引风机方式，通过调节引风机入口压头从而降低燃气透平的排气背压，提高透平的膨胀比，提出了一种燃气-蒸汽联合循环系统变背压改造方案[21-22]，分析显示该方法可提升部分负荷下联合循环系统的发电效率。在余热锅炉出口设置引风机的基础上，在压气机之前设置前置膨胀机用以调节进气流量，可实现进口参数调控和变背压相结合，使得联合循环在变工况下高效运行[23]。
综合而言，已有的研究主要基于大背压下联合循环的集成与优化，为了主动适应机组运行时流量的调节，本文提出了一种具有大流量调节范围潜力的高背压燃机联合循环系统，并对其设计工况下透平背压对系统性能的影响开展了能量分析和㶲分析，分析了设计压比和燃机背压2个关键参数的热力特性规律，进一步考虑了其实际工程应用情景和可再生能源灵活利用的方式。本研究为多能互补的灵活高效的现代化联合循环提供了新的系统方案参考。
1　高背压燃气轮机联合循环发电系统
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1.1　系统集成思路
高背压燃机联合循环基于现有F级燃机及联合循环的关键参数，将空气压缩过程分为新空气预压缩过程和新空气与燃机高压排气混合后的主压缩过程两部分，燃烧室出口高温烟气膨胀过程分为膨胀比恒定的主膨胀过程和排烟余压回收膨胀过程两部分。该系统设计的目的是方便在变负荷时主压气机能有较高的流量调节范围，即：在设计工况时，主压气机进气压力和主透平的排气压力较高，变工况时2个压力逐步降低，但膨胀比基本不变。2个压力的调节可以分别通过前置压机和后置膨胀透平的转速和进气角度进行控制。为了全面分析不同设计背压对联合循环性能的影响，本研究也对背压小于环境压力情况进行了分析。
1.2　联合循环发电系统流程
1.2.1　常规循环流程
高背压燃机联合循环发电机组中，新空气经前置压气机预压缩与新空气和燃机高压排气混合后经主压气机进行主压缩过程。主压气机的高压排气进入燃烧室使燃料充分燃烧，燃烧室出口高温烟气经透平进行膨胀比不变的主膨胀过程。主透平的高压排气经余热锅炉后分为再循环烟气和排放烟气两部分，排放烟气进入膨胀机，膨胀至常压后经余热回收排入大气。再循环烟气经冷却器冷却后进入主压气机（方案1），或不经冷却器冷却进入主压气机（方案2）。为方便区分2个余热锅炉，将排放再循环烟气的余热锅炉记作主余热锅炉，将排放烟气的余热锅炉记作副余热锅炉，2个余热锅炉共用1个汽轮机。高背压燃机联合循环流程如图1所示。
1.2.2　透平背压小于环境压力的联合循环流程
在透平设计背压低于101.3 kPa且保持膨胀比不变时，不参与再循环的余热锅炉排放烟气经压气机压缩至102.0 kPa后排放到大气。此时，排放烟气的温度较高，可达到150~250 ℃，为了利用这部分热量，用排放烟气去加热前置膨胀机的进口空气，使得空气膨胀后达到15 ℃。机组在图1的基础上，不参与再循环的余热锅炉进口前设置压气机，主压气机前设置膨胀机，增加换热器，其余流程不变，改造后流程如图2所示。
2　系统模拟方法与案例分析
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2.1　模拟方法与主要参数
本文对高背压燃机联合循环机组使用商业软件Aspen Plus进行设计工况流程模拟及热力计算。采用PG9351FA联合循环发电机组作为参比系统，开展高背压燃机联合循环机组的热力性能分析[24]，其主要参数见表1。
2.2　案例分析
对常规燃机联合循环机组与高背压燃机联合循环机组中2个方案进行能量分析以及㶲分析。其中，前置压气机及余压回收膨胀机视为燃机部分。
2.2.1　能量分析
常规燃机联合循环机组与高背压燃机联合循环机组在透平背压为环境压力时的设计参数及热力计算结果对比见表2。因方案1再循环烟气冷却过程中存在压损，由表2可知，其联合循环效率低于常规联合循环效率0.03百分点。方案2相较于方案1，再循环烟气直接进入主压气机使得进汽温度较环境温度高68 ℃，压气机出口温度高约62 ℃，从而使得燃烧室燃烧㶲损失降低，但同时也会导致压气机耗功增加。总体来看，得益于烟气余热的再利用和燃烧㶲损失降低，方案2机组的联合循环出功较方案1提高1.68 MW。
取3个典型的工况点，透平背压分别0.03、0.10、0.30 MPa进行热力计算并进行方案对比，其中透平背压为环境压力时，机组只有1个余热锅炉，节点温差取30/25/25 ℃；透平背压不为环境压力时，机组有主余热锅炉和副余热锅炉，副余热锅炉的排放烟气排入大气，节点温差取30/25/25 ℃，主余热锅炉的节点温差根据联合循环选取，结果见表3。由表3可知，当透平背压为0.03 MPa时，因余热锅炉出口烟气的压力需压缩提高至环境压力后排入大气，烟气压缩过程耗功且排放烟气温度可达200 ℃以上，后利用排放烟气加热空气，回收部分排烟热量，排放烟气温度最终约为150 ℃，因此透平背压为0.03 MPa时联合循环效率比透平背压0.10 MPa时的降低2百分点以上。当透平背压为0.30 MPa时，与透平背压为0.10 MPa相比，再循环烟气流量增大，烟气余热回收量增大，使得方案2的联合循环效率提高0.49百分点，然而方案1中再循环烟气冷却过程中存在压损，使得方案1在透平背压为0.30 MPa时比透平背压为0.10 MPa的联合循环效率降低0.14百分点。
2.2.2　㶲分析
常规联合循环机组与高背压联合循环机组在透平背压为环境压力时的㶲计算结果见表4。
由2.2.1能量分析可知，在透平背压为环境压力时，方案2的效率最高。方案2与常规燃机联合循环相比，因压气机出口温度高，使得燃烧室㶲损减少11.08 MW，压气机和透平㶲损略有增大，同时余热锅炉的进气温度较高，余热锅炉的㶲损增大4.95 MW，蒸汽轮机与凝汽器㶲损略有增大。综合表4数据来看，方案2机组的㶲效率相较于常规联合循环机组高0.2百分点。
3　高背压燃气轮机联合循环系统特性
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3.1　主顶循环设计压比对系统性能的影响
针对高背压燃机联合循环系统，保持主透平进气温度为1 327 ℃，选取0.03、0.10、0.30 MPa 3个典型透平设计背压，分析主压气机设计压比对系统性能的影响。
图3为3个透平设计背压下主压气机设计压比对压气机排气温度的影响规律。由图3可见，当设计压比在13.0~21.0时，压气机出口温度随压比递增。透平背压0.03、0.10 MPa的2条温度曲线几乎重合，这是由于这2个背压下主压气机的进气温度相同；透平背压为0.30 MPa时，由于前置压气机的影响，主压气机进气温度较高，导致压气机出口温度较高。方案2与方案1相比，再循环烟气不经冷却器进入主压气机，导致主压气机进口温度较高，使得压气机出口温度较高。在透平背压为0.30 MPa、设计压比为21.0时，方案2压气机出口温度可达623.1 ℃。
图4为3个透平设计背压下主压气机设计压比对燃机出功的影响规律。本节中，前置压气机耗功和余压回收膨胀机出功归为燃机出功。在燃料相同的情况下，当设计压比在13.0~21.0时，燃机出功随压比递增。相较于透平背压分别为0.10、0.03 MPa时，由于需引风机做功使得透平背压升到0.01 MPa（高于进气膨胀所做的功），导致燃机出功降低；透平背压为0.30 MPa时，由于没有参与再循环的循环工质总压比增加（包含了预压缩和尾部膨胀），因此燃机出功升高。方案2与方案1相比，再循环烟气不经冷却器进入主压气机，导致主压气机进口温度较高，使得压气机耗功较大。在方案1背压0.30 MPa、压比由13.0增加至21.0时，燃机出功由256.98 MW增加至277.91 MW。
图5为3个透平设计背压下主压气机压比对底循环出功的影响规律。当压比由13.0增加至21.0时，底循环出功递减。透平背压0.03、0.10 MPa的2根温度曲线几乎重合，这是由于这2个背压下透平的排气温度几乎相同；透平背压为0.30 MPa时，部分透平排气经膨胀机膨胀至0.103 MPa后进入副余热锅炉，使得副余热锅炉的进气温度大幅度降低，导致底循环出功较低。方案2与方案1相比，在透平初温和膨胀比不变的情况下，再循环烟气不经冷却器进入主压气机，导致主压气机的进气成分改变（CO2含量增加），使得透平排烟温度较高，因此底循环出功较大。
图6为2个方案在3个透平设计背压下主压气机设计压比对联合循环效率和比功的影响规律。由图6a)可见，方案1当压比由13.0增至21.0时，比功均递减，与常规联合循环变化规律相同；透平背压为0.03、0.10 MPa的联合循环效率递增，还未出现最佳压比，透平背压为0.30 MPa的联合循环效率先增后减，存在最佳压比。与背压为0.10 MPa的情况相比，当透平背压为0.03 MPa时，由于需引风机做功使得透平背压升到0.01 MPa（高于进气膨胀所做的功），导致联合循环效率较低，联合循环功率也低，压比相同，工质流量变化不大，同时也使得比功较低。当透平背压为0.30 MPa时，没有参与再循环工质的总膨胀比，远高于参与再循环的膨胀比，两者综合（折合等效总压比较高）使得在主压气机设计压比存在最佳值；同样，由于折合等效压比较高，使得比功最低。具体来看，当压比由13.0增至21.0时，透平背压为0.03 MPa和0.10 MPa均未出现最佳压比，这是由于主压气机进气的CO2含量递增所导致；透平背压为0.30 MPa时，主燃气轮机的循环压比若为15.4，未参与再循环的烟气工质，其总膨胀比应为46.2，因此，主压气机出现最佳设计压比，约为16.2，此时联合循环效率为58.08%，比功为574.02 kJ/kg。
由图6b)可见，方案2当压比由13.0增至21.0时，比功递减，透平背压为0.03 MPa和0.10 MPa时，联合循环效率递增（背压为0.03 MPa时，在压比为21.0附近已开始出现最佳压比的情况），透平背压为0.30 MPa的联合循环效率先增加后减小。效率和比功变化规律的原因同方案1。方案2在透平背压为0.30 MPa的最佳压比为17.0，此时联合循环效率为58.97%，比功为563.87 kJ/kg。
方案2相较于方案1，再循环烟气不经冷却器进入主压气机，不存在再循环烟气冷却过程中的压损，因此效率较高。同时，因主压气机的进气成分不同，导致方案1与方案2在透平背压为0.30 MPa时的最佳压力不同。
3.2　主顶循环设计背压对系统性能的影响
针对高背压燃机联合循环系统，综合考虑比功，主压气机设计压比推荐为15.4，保持主透平膨胀比和燃料量不变，分析透平背压对系统设计工况性能的影响。
图7为透平设计背压对联合循环、燃机和底循环出功的影响规律。本节中前置压气机耗功和余压回收膨胀机出功归为燃机部分。当燃料量不变，透平背压由0.03 MPa增至0.35 MPa时，燃机出功、底循环出功和联合循环出功均先增加后减小。
透平设计背压由0.03 MPa增至0.10 MPa时，由于引风机耗功减小，导致燃机出功递增；透平背压由0.10 MPa增至0.35 MPa时，没有参与再循环的烟气的总膨胀比逐渐增加，但流量却逐渐减小，且主压气机进气的CO2体积分数增加（图8），使得透平排烟温度增大，总体来看燃机出功先增加后减小。其中图7b)的燃机出功递增，这是由于再循环烟气温度较高使得燃机出功最高点延后所致。
透平设计背压由0.03 MPa增加至0.10 MPa时，由于透平排气温度几乎相同，使得底循环出功略微变化；透平背压由0.10 MPa增加至0.35 MPa时，由于没有参与再循环的烟气的总膨胀比逐渐增加，导致副余热锅炉的进气温度递减，使得底循环出功递减。
总体来看，因燃机功率高于底循环功率，对联合循环影响更大。因此，联合循环出功与燃机功率类似，先增后减，但变化趋势更为明显。
方案2与方案1相比，再循环烟气不经冷却器进入主压气机，使得主压气机进气温度较高，使得压气机耗功较大，导致燃机出功较低；透平排烟温度较高，使得底循环出功较高；不存在再循环烟气冷却过程中的压损，因此联合循环出功较高。
图9为透平设计背压对联合循环效率和比功的影响规律。由图9可见，透平背压由0.03 MPa增至0.35 MPa时，联合循环效率和比功先增加后减小。因燃料量不变，联合循环效率变化原因同图7中联合循环出功的变化。透平背压由0.03 MPa增加至0.10 MPa时，由于联合循环效率增大（压比相同，工质流量变化不大），导致比功增大；透平背压由0.10 MPa增至0.35 MPa时，由于没有参与再循环的烟气的总膨胀比逐渐增加，折合等效总压比增加，比功逐渐减小。
由前述可知，方案2的联合循环效率比方案1的略高，都存在最佳值：方案1在透平背压0.16 MPa时，联合循环效率最高为58.57%，比功621.88 kJ/kg；方案2在透平背压为0.20 MPa时，联合循环效率最高为59.12%，比功为604.35 kJ/kg。方案2的比功略低于方案1，主要是受工质成分影响导致流量变化所致，但在较高背压之后比功略高于方案1，主要是联合循环效率的相对提高带来的影响。
4　高背压联合循环系统应用展望
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本文提出的高背压联合循环系统中有2个余热锅炉，即主余热锅炉和副余热锅炉，增加了制造成本。为了适应工业要求，该系统也可以只使用主余热锅炉，但是随着透平背压增大，余压回收膨胀机的排烟温度低于环境温度，热力特性较差。因此在本节对只使用主余热锅炉的系统进行热力特性分析及优化。
4.1　系统流程及介绍
图10为单余热锅炉的高背压联合循环系统流程。由图10可见，系统中新空气经前置压气机预压缩与新空气和燃机高压排气混合后经主压气机进行主压缩过程，主压气机的高压排气进入燃烧室使燃料充分燃烧，燃烧室出口高温烟气经透平进行膨胀比不变的主膨胀过程，主透平的高压排气经余热锅炉后分为再循环烟气和排放烟气两部分，再循环烟气分2种情况，即经冷却器冷却后进入主压气机或不经冷却器冷却后进入主压气机，系统若与可再生能源相结合，排放烟气由太阳能加热后进入膨胀机，若不与可再生能源相结合，则排放烟气直接进入膨胀机，排放烟气膨胀至常压后经余热回收排入大气。
在透平背压为0.30 MPa、压气机压比为15.4时，再循环烟气不经冷却器冷却情况下，单余热锅炉机组和双余热锅炉机组热力计算结果见表5。其中，膨胀机入口温度不经太阳能加热的单余热锅炉机组记作单余热锅炉A，膨胀机入口温度提高至210 ℃的单余热锅炉机组记作单余热锅炉B。
由表5可见，在不考虑前置压气机耗功以及余压回收膨胀机做功的情况下，燃机数据几乎相同。因单余热锅炉机组的排放烟气是经余热锅炉后排放，余压回收膨胀机的入口烟温约为90 ℃，双余热锅炉机组的排放烟气是经余压回收膨胀机膨胀后进入余热锅炉，余压回收膨胀机的入口烟温约为640 ℃，所以单余热锅炉机组的蒸汽轮机功率相对高34.83 MW，而膨胀功相对低41.38 MW。当提高单余热锅炉机组膨胀机入口烟温时，余压回收膨胀机做功提高7.52 MW，蒸汽轮机出功不变，联合循环效率为58.97%，较双余热锅炉机组的联合循环效率58.94%高0.03百分点，且经济性更优。
4.2　高背压联合循环系统耦合太阳能热发电系统
高背压联合循环系统可与太阳能热发电系统相耦合[25]，利用太阳能热发电系统加热余热锅炉的排放烟气，从而增大余压回收膨胀机的出功，使得联合循环效率增大，同时降低了太阳能储存设备的投资。耦合方案中的太阳能热不参与联合循环效率计算，太阳能发电效率是高背压联合循环系统出功增加部分除以输入系统的总太阳能的结果。
图11为膨胀机入口烟气温度对联合循环效率和太阳能发电效率的影响规律。由图11可见，随着膨胀机入口烟气温度升高，联合循环效率和太阳能发电效率递增。当烟气温度达到240 ℃时，联合循环效率为59.29%，太阳能发电效率为18.77%，此时的单余热锅炉机组联合循环效率比双余热锅炉机组高0.35百分点。
5　结论
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基于常规联合循环对于压气机进气和燃机排气的压力进行提升改造，本文提出了一种高背压燃机联合循环系统，对其进行系统集成优化研究的主要结论如下。本研究为未来高效，特别是具有变工况高效潜力的联合循环发电系统提供了设计参考。
1）当高背压联合循环系统的主压气机设计压比由13.0提高至21.0，联合循环效率先升高后降低，压比存在最佳值（方案1为16.2，方案2为17.0），比功逐渐降低。当透平设计背压为0.30 MPa、主压气机设计压比为17.0时，方案2的联合循环效率为58.97%，透平设计背压为0.30 MPa、主压气机设计压比为16.2时，方案1的联合循环效率为58.08%。当主压气机设计压比取最佳值时，方案2效率比方案1高0.89百分点。方案2的联合循环效率始终高于方案1。
2）当主透平设计背压小于环境压力时，联合循环效率低于常压背压时循环效率；当透平设计背压由0.10 MPa增加到0.35 MPa，联合循环效率先升高后降低，背压存在最佳值，其中方案1为0.16 MPa，方案2为0.20 MPa。方案2的联合循环效率高于方案1，方案1和方案2的联合循效率变化区间分别在56.00%~58.57%和55.81%~59.12%，效率变化幅度不大，未来设计时推荐背压0.30 MPa及更高的背压。
3）考虑到经济性，高背压联合循环系统可使用1个余热锅炉，但系统联合循环效率会降低约1.05百分点；但当利用低温太阳热来提高膨胀机入口的排烟温度时，如入口温度为210 ℃时，效率为58.97%，已高于双预热锅炉的效率；同时，该系统还可灵活适应太阳能热源的不稳定性，省去储能系统，大幅度降低太阳能热发电成本。
基金
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  • 国家科技重大专项(J2019-I-0009-0009)
文献
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2025年第54卷第11期
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doi: 10.19666/j.rlfd.202501049
  • 接收时间：2025-01-23
  • 首发时间：2026-01-13
  • 出版时间：2025-11-25
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  • 收稿日期：2025-01-23
基金
National Science and Technology Major Project(J2019-I-0009-0009)
国家科技重大专项(J2019-I-0009-0009)
作者信息
    1.华北电力大学能源动力与机械工程学院，北京　102206
    2.华北电力大学动力工程系，河北　保定　071003

通讯作者:

张国强（1980），男，副教授，主要研究方向为能源动力集成及其优化，。
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https://castjournals.cast.org.cn/joweb/rlfd/CN/10.19666/j.rlfd.202501049
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2种不同金属材料的力学参数
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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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