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Article(id=1319293591441224379, tenantId=1146029695717560320, journalId=1316781260334530699, issueId=1319293542573392487, articleNumber=null, orderNo=null, doi=10.3969/j.issn.1005-0329.2025.07.001, pmid=null, cstr=null, oa=null, hot=null, price=null, onlineType=0, articleFormat=0, articleType=null, articleTypeStr=null, receivedDate=1725984000000, receivedDateStr=2024-09-11, revisedDate=1739548800000, revisedDateStr=2025-02-15, acceptedDate=null, acceptedDateStr=null, onlineDate=1792473703489, onlineDateStr=2026-10-20, pubDate=1753804800000, pubDateStr=2025-07-30, doiRegisterDate=null, doiRegisterDateStr=null, onlineIssueDate=1792473703489, onlineIssueDateStr=2026-10-20, onlineJustAcceptDate=null, onlineJustAcceptDateStr=null, onlineFirstDate=null, onlineFirstDateStr=null, sourceXml=null, magXml=null, createTime=1792473703489, creator=13701087609, updateTime=1792473703489, 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=1, endPage=8, ext={EN=ArticleExt(id=1319293591890014908, articleId=1319293591441224379, tenantId=1146029695717560320, journalId=1316781260334530699, language=EN, title=Experimental study on heating performance of low‑temperature heat pump water heater with vapor injection, columnId=null, journalTitle=Fluid Machinery, columnName=null, runingTitle=null, highlight=null, articleAbstract=

To address the heating capacity attenuation of air‑source heat pumps under low‑temperature conditions，an experimental setup with a flash tank for vapor injection was established to investigate the influence of parameters such as intermediate refrigerant pressure and superheat degree on system performance under superheated vapor injection conditions.The results show that the flow rate and state of the injected refrigerant are key factors.When the vapor injection pressure increases，the superheat degree decreases and the refrigerant dryness approaches 1.As the electronic expansion valve opening increases from 10% to 40%，the vapor injection pressure rises from 1.48 MPa to 1.79 MPa under minimum operating conditions and from 1.01 MPa to 1.71 MPa under low‑temperature operating conditions.Both the vapor injection enthalpy increment and heating coefficient of performance （COP） exhibit a trend of first increasing and then decreasing，with an overall improvement in heating capacity.When the electronic expansion valve opening is approximately 25%，the COP reaches 3.64 under minimum operating conditions and a maximum value of 2.62 under low‑temperature operating conditions.Maintaining the injected refrigerant close to saturated vapor state is an important control factor，while the theoretically optimal enthalpy‑increasing pressure is not universally applicable.This study can provide references for the design，operation，and control of low‑temperature heat pump water heaters.

, authors=Jiangnan LIN, Haoning CAO, Zhenya ZHANG, Tingxiang JIN, Gailian LI, authorsList=Jiangnan LIN, Haoning CAO, Zhenya ZHANG, Tingxiang JIN, Gailian LI, authorCompany=null, correspAuthors=Tingxiang JIN, 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=1319293595501310662, articleId=1319293591441224379, tenantId=1146029695717560320, journalId=1316781260334530699, language=CN, title=补气增焓低温热泵热水器制热性能试验研究, columnId=1317484286475883136, journalTitle=流体机械, columnName=试验研究, runingTitle=null, highlight=null, articleAbstract=

针对空气源热泵低温制热衰减问题，搭建带闪发器的补气增焓实验台，研究补气过热条件下中间支路制冷剂压力与过热度等参数对系统性能的影响。结果表明，中间补气制冷剂流量和状态是关键因素；补气压力上升时，过热度降低，制冷剂干度趋近于1；电子膨胀阀开度从10%增至40%时，最小运行工况下补气压力从1.48 MPa增大至1.79 MPa，低温运行工况下补气压力从1.01 MPa增大至1.71 MPa；补气增焓量和制热性能系数（COP）均呈现先增后减的趋势，整体制热能力提升；当电子膨胀阀开度约为25%时，最小运行工况下的COP为3.64，低温运行工况下的COP最大值为2.62。保持补气制冷剂接近饱和蒸汽状态是重要调控要素，而理论最佳增焓压力不具普遍适用性。研究可为低温热泵热水器设计、运行和控制提供参考。

, authors=林江男, 曹浩宁, 张振亚, 金听祥, 李改莲, authorsList=林江男, 曹浩宁, 张振亚, 金听祥, 李改莲, authorCompany=null, correspAuthors=金听祥, authorNote=

林江男（1988），女，硕士研究生，主要从事制冷空调节能新技术研究，E-mail：。

, correspAuthorsNote=
金听祥（1976），男，教授，主要从事制冷空调设备新技术及关键部件等相关研究工作，通信地址：450002河南省郑州市高新区科学大道136号郑州轻工业大学能源与动力工程学院，E-mail：。
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林江男（1988），女，硕士研究生，主要从事制冷空调节能新技术研究，E-mail：。

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林江男（1988），女，硕士研究生，主要从事制冷空调节能新技术研究，E-mail：。

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ArticleFig(id=1319293701415874962, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=EN, label=Fig.4, caption=Variation trends of heating COP and vapor injection superheat under two operating conditions, figureFileSmall=PsZOXRQygsNEzyBYAT/yHQ==, figureFileBig=INRv5p42ZBmfT9Kyj1QiRQ==, tableContent=null), ArticleFig(id=1319293701474595219, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=CN, label=图4, caption=2种运行工况下制热性能系数与补气过热度的变化趋势, figureFileSmall=PsZOXRQygsNEzyBYAT/yHQ==, figureFileBig=INRv5p42ZBmfT9Kyj1QiRQ==, tableContent=null), ArticleFig(id=1319293701545898388, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=EN, label=Tab.1, caption=

Specification parameters of main test equipments

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设备名称参数
压缩机变频补气增焓转子压缩机，排气容积10.2 cm3/min，工质R410A，转速900~7 200 r/min，额定制热量4.25 kW，额定功率1.02 kW
管翅式
换热器		管径8 mm，排间距21 mm，列间距20 mm；
管排数为29，管列数为3，流程数为9，外形尺寸591 mm×550 mm×45 mm
电子膨胀阀1适配工质R410A，最高工作压力4.5 MPa
电子膨胀阀2适配工质R410A，最高工作压力4.5 MPa
蓄热水箱盘管式微通道换热器，容量200 L
单向阀适用工质R410A，适用温度-30~50 ℃
闪发器外径35 mm；容积200 cm3
), ArticleFig(id=1319293701608812949, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=CN, label=表1, caption=

主要试验设备规格参数

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设备名称参数
压缩机变频补气增焓转子压缩机，排气容积10.2 cm3/min，工质R410A，转速900~7 200 r/min，额定制热量4.25 kW，额定功率1.02 kW
管翅式
换热器		管径8 mm，排间距21 mm，列间距20 mm；
管排数为29，管列数为3，流程数为9，外形尺寸591 mm×550 mm×45 mm
电子膨胀阀1适配工质R410A，最高工作压力4.5 MPa
电子膨胀阀2适配工质R410A，最高工作压力4.5 MPa
蓄热水箱盘管式微通道换热器，容量200 L
单向阀适用工质R410A，适用温度-30~50 ℃
闪发器外径35 mm；容积200 cm3
), ArticleFig(id=1319293701701087638, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=EN, label=Tab.2, caption=

Test conditions of the vapor‑injection heat pump water heater system

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项目电子膨胀阀
开度/%		水侧空气侧
进水温度/℃出水温度/℃干球温度/℃湿球温度/℃
最小运行工况10，15，20，25，30，35，4095576
低温运行工况10，15，20，25，30，35，40955-7-8
), ArticleFig(id=1319293701780779416, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=CN, label=表2, caption=

补气增焓热泵热水器系统的试验工况

, figureFileSmall=null, figureFileBig=null, tableContent=
项目电子膨胀阀
开度/%		水侧空气侧
进水温度/℃出水温度/℃干球温度/℃湿球温度/℃
最小运行工况10，15，20，25，30，35，4095576
低温运行工况10，15，20，25，30，35，40955-7-8
), ArticleFig(id=1319293701843693977, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=EN, label=Tab.3, caption=

Test data of pressure and heating COP under minimum operating conditions

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开度/%冷凝压力/MPa蒸发压力/MPa实测中间压力/MPa理论中间压力/MPaCOP
104.390.851.481.933.30
154.320.851.531.923.43
204.220.851.581.893.52
254.190.851.631.893.64
304.110.851.671.873.58
354.050.851.721.863.55
403.970.851.791.853.53
), ArticleFig(id=1319293701919191450, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=CN, label=表3, caption=

最小运行工况下压力与制热COP测试数据

, figureFileSmall=null, figureFileBig=null, tableContent=
开度/%冷凝压力/MPa蒸发压力/MPa实测中间压力/MPa理论中间压力/MPaCOP
104.390.851.481.933.30
154.320.851.531.923.43
204.220.851.581.893.52
254.190.851.631.893.64
304.110.851.671.873.58
354.050.851.721.863.55
403.970.851.791.853.53
), ArticleFig(id=1319293701977911707, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=EN, label=Tab.4, caption=

Test data of pressure and heating COP under low‑temperature operating conditions

, figureFileSmall=null, figureFileBig=null, tableContent=
开度/%冷凝压力/MPa蒸发压力/MPa实测中间压力/MPa理论中间压力/MPaCOP
104.330.531.011.512.38
154.270.531.181.502.42
204.230.531.291.492.55
254.190.531.421.492.62
304.160.531.531.482.56
354.130.531.641.482.52
404.090.531.711.472.49
), ArticleFig(id=1319293702053409180, tenantId=1146029695717560320, journalId=1316781260334530699, articleId=1319293591441224379, language=CN, label=表4, caption=

低温运行工况下压力与制热COP测试数据

, figureFileSmall=null, figureFileBig=null, tableContent=
开度/%冷凝压力/MPa蒸发压力/MPa实测中间压力/MPa理论中间压力/MPaCOP
104.330.531.011.512.38
154.270.531.181.502.42
204.230.531.291.492.55
254.190.531.421.492.62
304.160.531.531.482.56
354.130.531.641.482.52
404.090.531.711.472.49
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补气增焓低温热泵热水器制热性能试验研究
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林江男 , 曹浩宁 , 张振亚 , 金听祥 , 李改莲
流体机械 | 试验研究 2025,53(7): 1-8
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流体机械 |试验研究 2025 , 53 (7) : 1 -8
补气增焓低温热泵热水器制热性能试验研究
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林江男 , 曹浩宁, 张振亚, 金听祥 , 李改莲
作者信息
  • 郑州轻工业大学 能源与动力工程学院，郑州 450002
通讯作者:
金听祥（1976），男，教授，主要从事制冷空调设备新技术及关键部件等相关研究工作，通信地址：450002河南省郑州市高新区科学大道136号郑州轻工业大学能源与动力工程学院，E-mail：。
作者简介:

林江男（1988），女，硕士研究生，主要从事制冷空调节能新技术研究，E-mail：。

Experimental study on heating performance of low‑temperature heat pump water heater with vapor injection
Jiangnan LIN , Haoning CAO, Zhenya ZHANG, Tingxiang JIN , Gailian LI
Affiliations
  • College of Energy and Power Engineering，Zhengzhou University of Light Industry，Zhengzhou 450002，China
出版时间: 2025-07-30 doi: 10.3969/j.issn.1005-0329.2025.07.001
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摘要
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针对空气源热泵低温制热衰减问题，搭建带闪发器的补气增焓实验台，研究补气过热条件下中间支路制冷剂压力与过热度等参数对系统性能的影响。结果表明，中间补气制冷剂流量和状态是关键因素；补气压力上升时，过热度降低，制冷剂干度趋近于1；电子膨胀阀开度从10%增至40%时，最小运行工况下补气压力从1.48 MPa增大至1.79 MPa，低温运行工况下补气压力从1.01 MPa增大至1.71 MPa；补气增焓量和制热性能系数（COP）均呈现先增后减的趋势，整体制热能力提升；当电子膨胀阀开度约为25%时，最小运行工况下的COP为3.64，低温运行工况下的COP最大值为2.62。保持补气制冷剂接近饱和蒸汽状态是重要调控要素，而理论最佳增焓压力不具普遍适用性。研究可为低温热泵热水器设计、运行和控制提供参考。

低温热泵热水器  /  补气增焓  /  中间补气压力  /  补气过热度  /  制热性能系数

To address the heating capacity attenuation of air‑source heat pumps under low‑temperature conditions，an experimental setup with a flash tank for vapor injection was established to investigate the influence of parameters such as intermediate refrigerant pressure and superheat degree on system performance under superheated vapor injection conditions.The results show that the flow rate and state of the injected refrigerant are key factors.When the vapor injection pressure increases，the superheat degree decreases and the refrigerant dryness approaches 1.As the electronic expansion valve opening increases from 10% to 40%，the vapor injection pressure rises from 1.48 MPa to 1.79 MPa under minimum operating conditions and from 1.01 MPa to 1.71 MPa under low‑temperature operating conditions.Both the vapor injection enthalpy increment and heating coefficient of performance （COP） exhibit a trend of first increasing and then decreasing，with an overall improvement in heating capacity.When the electronic expansion valve opening is approximately 25%，the COP reaches 3.64 under minimum operating conditions and a maximum value of 2.62 under low‑temperature operating conditions.Maintaining the injected refrigerant close to saturated vapor state is an important control factor，while the theoretically optimal enthalpy‑increasing pressure is not universally applicable.This study can provide references for the design，operation，and control of low‑temperature heat pump water heaters.

low‑temperature heat pump water heater  /  vapor injection  /  intermediate vapor injection pressure  /  vapor injection superheat degree  /  heating coefficient of performance
林江男, 曹浩宁, 张振亚, 金听祥, 李改莲. 补气增焓低温热泵热水器制热性能试验研究. 流体机械, 2025 , 53 (7) : 1 -8 . DOI: 10.3969/j.issn.1005-0329.2025.07.001
Jiangnan LIN, Haoning CAO, Zhenya ZHANG, Tingxiang JIN, Gailian LI. Experimental study on heating performance of low‑temperature heat pump water heater with vapor injection[J]. Fluid Machinery, 2025 , 53 (7) : 1 -8 . DOI: 10.3969/j.issn.1005-0329.2025.07.001
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0　引言
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热泵作为一种高效、节能环保的制热技术，可在耗少量电能的同时将低品位热能转化为高品位热能，是降低建筑空调能耗的有效手段之一［1］。但在冬季低温条件下，室外温度的降低会使蒸发温度（压力）随之下降，导致制冷剂循环流量减小，制热量下降；蒸发压力的下降还会进一步造成系统压缩比增大，排气温度升高，降低压缩机运行的可靠性和安全性，严重阻碍热泵系统在全国范围内的推广应用［2-3］。基于上述问题，现有研究中提出了双级压缩技术［4-6］、复叠循环［7-8］及补气增焓技术［9］等多种解决方案。相较于多级压缩技术和复叠循环，补气增焓技术因应用形式简单且具有较好的低温适应性，受到学者们广泛关注。
补气增焓技术是指在传统制冷循环的基础上，通过增加中间补气支路，将冷凝器出口的部分制冷剂经补气口流入压缩机混合压缩腔，以增加制冷剂循环总量，同时降低排气温度。此外，还通过构建准二级压缩系统，减少压缩机的耗功，使COP得到显著提升［10］。典型补气增焓热泵系统主要由补气增焓压缩机、蒸发器、冷凝器、闪发器（或过冷器）及多个节流部件组成。与过冷器补气热泵系统相比，闪发器补气热泵系统的结构紧凑、管路连接简单，且闪发器作为关键部件，其传热效率优于过冷器［11］，有助于降低蒸发器入口的焓值。
近年来，相关研究人员对带闪发器的补气增焓热泵技术进行了广泛研究，包括系统结构优化、压缩机改造和控制策略设计等。KANG等［12］在单级补气系统的基础上提出多级补气结构，不同中间压力的制冷剂分别被喷射入压缩机的中间压缩腔，依次完成补气、混合与压缩过程，相关性能研究结果表明，双级补气系统在制热量和COP方面优于单级补气系统。WANG等［13］提出一种新型端面喷射结构，即将补气孔的位置设在压缩机的上法兰或下法兰上，由单向阀控制补气孔的启闭，理论上，采用该结构可以完全避免补气回流，同时增强压缩机的变工况适应性，研究结果表明，通过增大补气孔的面积及选择合理的补气孔位置，系统制热量和COP相较于传统喷射结构分别提升12.5%~18.2%和0.8%~3.5%，但余隙容积的存在会降低压缩等熵效率，使系统的功耗增加。WANG等［14］采用极值寻求控制方法，分别将系统功耗和补气压力设为目标优化量和控制量，并基于控制回路反馈的功耗值调节补气压力，实现了系统实时功耗的最小化。TELLO‑OQUENDO等［15］开发了一种基于半经验模型的预测方法，用于预测补气涡旋式压缩机的性能参数，该方法特别注重压缩机的参数，而不依赖于系统的具体形式，研究表明，该方法能够将蒸发器质量流量、补气质量流量及压缩机耗电量的预测值与实际测量值之间的最大偏差均控制在5%以内。此外，还有研究者对补气涡旋式压缩机的工作过程进行了建模和仿真分析，进一步深化了对这一领域的认知［16］。这种基于半经验模型预测方法的提出，为补气涡旋式压缩机的性能预测提供了一种更为精确和实用的工具。
柴沁虎等［17］构建了涡旋式压缩机的压缩和补气过程的分析模型，深入探讨了不同运行阶段补气涡旋式压缩机的动态特性，通过模拟分析，发现补气孔的位置对压缩机的制热性能有着显著的影响。特别是，最合理的开孔位置在吸气腔刚刚闭合处，这一位置能够最大程度地保证系统在低温环境下的制热能力，并且有助于提高系统的性能系数（COP）。这项研究强调了在设计补气涡旋式压缩机时，合理布局补气孔位置的重要性，以实现在保持系统性能系数的同时，优化其在低温条件下的制热性能。张科等［18］基于补气涡旋式压缩机的实际运行情况，构建了一套分析框架，旨在模拟该压缩机的压缩和补气行为，通过将模拟得到的数据与实际测量数据进行比较，发现系统性能参数随着蒸发温度的变化趋势与试验数据相吻合，并且模拟值与实际值之间的误差均未超过5%，表明所建立的模型能够较为精确地反映压缩机的动态特性。这些研究揭示了通过优化系统结构、改进压缩机设计、制定有效的控制策略及精确预测性能参数，可以显著提高补气增焓热泵的效能和能源效率。尽管这些研究取得了一定的进展，但同样发现了一些关键问题，例如余隙容积对压缩机等熵效率的影响，以及补气孔位置对制热量和COP的影响。
基于上述研究的总结和思考，本文旨在深入探讨补气增焓热泵技术中的关键问题，特别是补气孔位置对系统性能的影响，以及如何通过进一步结构优化和控制策略改进来克服现有技术的局限性，因此，提出一种新的补气结构和控制方法，以实现更高的制热量和COP，同时降低系统的功耗，为补气增焓热泵技术的进一步发展提供理论和实践指导。
1　理论分析
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1.1　补气增焓低温热泵循环工作原理
补气增焓低温热泵系统原理及循环压焓图如图1所示。压缩机排出的高温高压气态制冷剂，流入蓄热水箱与冷却水进行热量交换，在冷凝盘管内放热后成为中温高压的液态制冷剂，然后通过电子膨胀阀1进行第一次节流后进入闪发器，在闪发器的出口分流为两路：主路制冷剂经电子膨胀阀2的二次节流后，进入室外蒸发器吸热制冷，随后流经气液分离器，返回压缩机吸气口进行初压缩；补气支路制冷剂经单向阀与初压缩后的制冷剂气体混合，进行再压缩，并完成整个制热循环。
1.2　补气增焓低温热泵循环的热力学分析
补气增焓低温热泵系统通过中间补气，提高循环制冷剂流量和系统能效。假设流经蒸发制冷主路的制冷剂流量为m，中间补气支路的制冷剂流量为α，则蒸发器制冷量可表示为：
Q1=mh1-h4'=mh1-h4+mh4-h4'
式中，Q1为蒸发器制冷量，kW；m为蒸发制冷主路的制冷剂质量流量，kg/s；h1，h4，h4'分别为点1，4，4'处的焓值，kJ/kg。
由图1（b）可以看出，在热泵系统中引入补气增焓技术，可将一次节流后的部分液体制冷剂进行冷却，使得蒸发制冷主路的比焓差由h1-h4增至h1-h4'。保持蒸发制冷主路中制冷剂的质量流量m不变，可有效增强蒸发器的换热效果。此时，该系统制冷量按照4'-4-1的换热计算模型进行评估，即式（1）。
根据闪发器内能量守恒关系，可得：
mh4-h4'=mh5-h3'=αh6-h5
式中，h3'，h5，h6分别为点3'，5，6处的焓值，kJ/kg；α为中间补气支路的制冷剂质量流量，kg/s。
联立式（1）（2）可得：
Q1=mh1-h4+αh6-h5
由式（3）可知，在蒸发主路中，制冷剂的质量流量与无补气系统保持一致的条件下，该系统的补气增焓量可表示为：
QΔh=αh6-h5
式中，QΔh为系统补气增焓量，kW。
压缩机耗功：
P=mh7-h1+m+αh2'-h8
式中，P为压缩机的总功率，kW；h2'，h7，h8分别为点2'，7，8处的焓值，kJ/kg。
由图1（a）可知，补气增焓型压缩机的总耗功主要由吸气压缩的低压级和经补气混合后的高压级两部分组成。其中，低压级部分仅对流经室外蒸发器的制冷剂进行初压缩，高压级部分则对经初压缩、中间补气混合后的制冷剂进行再压缩。
补气增焓热泵热水器系统的制热量可表示为：
Qh=m+αh2'-h3=cpMW(T2-T1)/τ
式中，Qh为热泵制热量，kW；h3为点3处的焓值，kJ/kg；cp为平均进出水温度下水的比热容，kJ/（kg·℃）；MW为蓄热水箱中被加热水的质量，kg；T1，T2分别为被加热水的初始温度和终止温度，℃；τ为热泵运行时间，s。
系统的制热COP可表示为：
COP=Qh/P
由式（7）可知，系统制热COP为制热量与耗功之比。根据能量守恒定律，该系统的制热量来源为2个部分：压缩机耗功和室外蒸发器换热量。由式（5）可知，随着中间补气量的增加，压缩机耗功不断增大，在提升制热量过程中逐渐占据主导地位，成为系统性能提高的不利因素，此时补气增焓量成为影响系统性能系数的关键因素。由式（4）可知，补气增焓量取决于中间补气量与制冷剂单位质量流量的焓差h6-h5。由于补气过热度和冷凝器出口过冷度均会随着补气量的增加而减小，导致比焓差也会有所减少。因此，当中间补气量变化的权重大于比焓差的变化量时，补气增焓效果增强，系统COP提高，反之COP降低。由此可见，中间补气量的合理调控是实现系统制热性能最佳的关键策略。
现有研究中［19-21］，补气支路制冷剂流量的控制主要参考中间压力，其最优值的计算公式如下：
pm=pepc
式中，pm为中间压力的理论最优值，MPa；pe为蒸发压力，MPa；pc为冷凝压力，MPa。
鉴于压缩机补气结构各异等原因，当压缩机类型及运行工况不同时，系统均有最佳中间补气量。在此条件下，不仅能够确保充足的制冷剂经补气口进入压缩机，完成再压缩，避免因制冷剂供给过少而导致补气增焓效果不理想，同时，还解决制冷剂过多引起的耗功增大、补气带液等问题。
2　试验装置及方法
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2.1　补气增焓低温热泵热水器系统
依据补气增焓热泵原理，搭建补气增焓热泵热水机组实验台，进行低温环境下的制热性能测试研究。主要试验设备规格参数如表1所示。
制热测试试验在标准焓差试验室的室外侧进行，其工作温度设定范围为-10~20 ℃。机组实验台及其测试系统如图2所示。
2.2　测试方案
我国北方地区冬季气温普遍低于冰点，维持在0 ℃以下。根据GB/T 23137—2020《家用和类似用途热泵热水器》，选取低温运行工况（空气侧干球温度-7 ℃，湿球温度-8 ℃，水侧初始温度9 ℃）作为性能评估测试点。另选取最小运行工况（空气侧干球温度7 ℃，湿球温度6 ℃，水侧初始温度9 ℃）作为对比分析测试点。为了深入分析补气增焓热泵系统中，中间补气支路参数对系统整体制热性能的影响，通过调节上游电子膨胀阀的开度来改变中间补气支路参数，其开度依次设定为10%，15%，20%，25%，30%，35%，40%，以此探究不同室外环境温度、不同中间支路参数下补气增焓热泵热水器系统的运行规律，相关试验工况见表2。为了确保试验结果的准确性和可靠性，排除其他潜在的干扰因素，将室外侧风机维持在满负荷运行状态，同时压缩机在固定频率下运转。
由图2可看出，补气增焓低温热泵热水器系统中的各级节流装置均采用电子膨胀阀，且配备相应的控制器分别独立控制：电子膨胀阀1主要调控中间补气压力等中间支路参数，电子膨胀阀2主要控制压缩机的吸气过热度保持恒定。此外，通过压力传感器和温度传感器分别对中间补气温度、中间补气压力、蒸发器入口温度、蒸发器入口压力、压缩机吸气温度、压缩机吸气压力、压缩机排气温度、压缩机排气压力等参数进行测量。根据补气增焓低温热泵热水器运行特性，在制热状态下，系统测试重点关注中间补气压力和中间补气过热度2个因素。
补气增焓效果的主要评价依据为制热COP。当标准焓差室工况稳定后，在保证压缩机吸气口制冷剂气体过热度（5 K）恒定的前提下，主要通过改变电子膨胀阀1的开度，调控中间补气参数，对比运行参数，评价制热COP和补气增焓效果。
3　试验数据及分析
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3.1　中间补气压力对制热性能系数的影响
该测试方案主要是验证系统冷凝压力、蒸发压力与中间补气压力之间的关系，寻找最优中间压力值。
3.1.1　最小运行工况试验
在最小运行工况下，风机满负荷运转，压缩机在固定频率下运行。通过改变中间补气压力，测取系统制热能效变化，其具体试验数据如表3所示。
最佳补气增焓压力的理论值Pm均在1.9 MPa左右；而实测中间压力受到冷凝压力的影响，其数值在一定范围内变化，与理论值有所差距。同时，随着电子膨胀阀1开度的逐步增大，中间压力呈现上升趋势，但机组的制热COP呈现出非线性变化趋势：初期随中间压力的增加而升高，达到某一峰值后，随着压力的继续增加而降低，形成类似抛物线的最优值区间。这表明在某一中间压力下，热泵系统能够实现最佳的性能系数表现。通过试验数据，可以确定在最小运行工况下，制热COP达到峰值的具体中间压力范围为1.6~1.8 MPa，由此可知，在热泵系统的设计和操作中，通过精确调整中间压力至一个特定的高效区间，可以显著提升系统的热效率。
3.1.2　低温运行工况试验
在低温运行工况下，风机满负荷运转，压缩机在固定频率下运行。试验通过改变中间补气压力，测试系统制热COP变化，其具体试验数据如表4所示。可以看出，在低温运行工况下，系统的蒸发压力保持稳定，均为0.53 MPa。受此影响，补气增焓压力的理论值稳定在1.5 MPa的水平，与实测值相比存在一定的偏差。此外，随着中间补气压力的调整，机组制热COP呈现先增后减的变化趋势。
在2种不同运行工况下，制热COP与中间补气压力之间的变化关系较为稳定，但实测数据与理论预测值存在较大差异。通过对比压缩机设计参数，得到与试验数据推导结果大致相符的结论，且中间压力主要受补气口在压缩机气缸壁上的位置影响。由于加工精度、补气状态和补气压力控制等因素，补气口的实际位置难以满足中间压力理论值的预设要求，其实测值会在一定范围内变化［21-22］。
经由中间补气压力对制热COP影响的试验验证，确定最佳补气压力理论值在低温热泵系统控制方面具有指导意义。然而，二者之间量化控制关系的建立存在一定挑战，不宜将理论值作为补气增焓热泵系统控制策略的关键变量。
3.2　补气过热度对制热能力的影响
在带闪发器的补气增焓热泵系统中，当气液分离效率较高时，补入压缩机的气态制冷剂接近于饱和蒸气状态。然而，已有试验研究表明［23］，当闪发器内液位较低时，会有少量气态制冷剂经补气口进入压缩机。此时，靠近压缩机补气口处的气态制冷剂会出现过热现象，其过热度可通过制冷剂的压力和温度参数测得与控制。在上述测试中间补气压力对制热COP的影响时，电子膨胀阀1的动作不仅能调控中间补气压力，也会使得补气过热度发生改变，直接影响比焓差h6-h5和补气增焓量。
图3示出最小运行工况和低温运行工况下，制热量与补气过热度的变化趋势曲线。可以看出，在电子膨胀阀开度增加到40%时，制热量达到最大值，对应的补气过热度达到最小值。
随着电子膨胀阀开度逐渐增加，补气过热度逐渐减小，系统制热量的变化曲线基本呈现增长速率不断减小的趋势，即过热度较大时，制热量增长率为正，随着过热度不断减小，制热量增长速率减缓。出现该变化规律的主要原因：（1）开阀初期，补气量增加、补气过热度逐渐减小，使得高数量级的比焓差h6-h5减小，但补气量的增长速率较大，总体呈现出补气增焓量不断增加的趋势，同时，压缩机耗功也不断增加，导致系统制热量的增长率为正数，且数值较大；（2）随着开阀比例的持续增大，冷凝器出口过冷度开始减小，同时，压缩腔内冷热流体混合冷却作用增强，补气过热度进一步减小，比焓差h6-h5的变化权重增大，导致补气增焓量在达到最大值后开始减小，使得系统制热量的增速明显减缓。基于上述分析可知，在补气过热度不断减小的过程中，补气增焓量大致呈现先增后减的变化趋势，存在使系统制热量增速达到最快的极值点。
3.3　补气过热度对制热性能系数的影响
由上述分析可知，随着电子膨胀阀1的动作，补气增焓量的变化规律和制热COP大致相同。结合理论分析，影响制热COP的主要因素有压缩机耗功和补气增焓量，后者对COP提升起着增益作用。补气过热度作为衡量补气增焓效果的一个重要参数，也可以在一定程度上判断制热COP的变化趋势。图4示出最小运行工况和低温运行工况下，制热COP与补气过热度的变化趋势曲线。
可以看出，在电子膨胀阀开度增加到25%时，制热量达到最大值，对应的补气过热度持续降低。机组的补气过热度接近线性变化，而制热COP呈现先增后减的非线性变化趋势。随着电子膨胀阀的开度逐渐增大，经补气口吸入的制冷剂质量流量随之增加，中间冷却作用增强，补气过热度减小。当补气过热度接近0（即干度x约为1）时，补气增焓量达到最大值，对应的制热COP最佳。理论研究表明［24］，当进入补气口的制冷剂处于饱和蒸气状态时，压缩机的吸气效率较高，系统补气增焓效果最佳，图4中的试验数据也支持了该理论。
在开度从30%增至40%过程中，电子膨胀阀的开度持续增加，而补气过热度保持不变，表明此时中间补气支路中制冷剂的状态已进入两相区，补气增焓量开始减少，制热量的相对增长率也逐渐减小，而整个过程中压缩机耗功随补气量的增加而不断增大，因此，制热COP出现拐点，开始减小。
总的来说，在电子膨胀阀的开度不断增加的过程中，中间补气量也随之逐渐增大，这导致压缩机补气口的制冷剂状态从过热状态逐渐转变为气液两相状态。当制冷剂达到饱和蒸汽状态时，补气增焓的效果最为显著，此时制热COP达到最高。
4　结论
收起
（1）低温热泵系统中，中间补气压力对热泵系统制热性能的优化控制有指导作用，但因位置、制冷剂的气液两相状态等因素影响，理论值与实测值差异较大，不宜作为控制核心。
（2）随着电子膨胀阀开度的增加，中间补气支路的制冷剂过热度对补气增焓量的影响较大，使其呈现出先增后降的变化趋势，即存在最佳开度。
（3）在不同运行工况下，当补气过热度逐渐减小并趋近于0（干度x约为1）时，机组制热COP均达到最优值区间，表明补气口处制冷剂的过热度在一定程度上是提升制热效率的关键。
（4）制热系统的COP与中间补气支路中制冷剂的过热度紧密相关。当中间补气支路的制冷剂过热度降低时，其增焓效果更佳，直接提升了系统的制热性能。进一步可看出，中间补气支路中制冷剂的具体状态，对于整个系统的制热效率有着决定性的影响。

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doi: 10.3969/j.issn.1005-0329.2025.07.001
  • 接收时间：2024-09-11
  • 首发时间：2026-10-20
  • 出版时间：2025-07-30
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  • 收稿日期：2024-09-11
  • 修回日期：2025-02-15
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    郑州轻工业大学 能源与动力工程学院，郑州 450002

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金听祥（1976），男，教授，主要从事制冷空调设备新技术及关键部件等相关研究工作，通信地址：450002河南省郑州市高新区科学大道136号郑州轻工业大学能源与动力工程学院，E-mail：。
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