CN211233433U

CN211233433U - Double-air-supply heat pump system

Info

Publication number: CN211233433U
Application number: CN201922053299.8U
Authority: CN
Inventors: 胡余生; 魏会军; 余冰; 徐嘉; 杨欧翔; 赵海红
Original assignee: Gree Electric Appliances Inc of Zhuhai; Gree Green Refrigeration Technology Center Co Ltd of Zhuhai
Current assignee: Gree Electric Appliances Inc of Zhuhai; Gree Green Refrigeration Technology Center Co Ltd of Zhuhai; Zhuhai Gree Energy Saving Environmental Protection Refrigeration Technology Research Center Co Ltd
Priority date: 2019-11-25
Filing date: 2019-11-25
Publication date: 2020-08-11
Anticipated expiration: 2029-11-25

Abstract

The utility model provides a two tonifying qi heat pump systems, including connecting the compression source that forms closed circulation in order, first heat exchanger, first throttling element, first economizer, second throttling element, second economizer, third throttling element, the second heat exchanger, the compression source is including parallelly connected first compression portion, second compression portion, wherein first compression portion has first exhaust port, first tonifying qi mouth and first air inlet, second compression portion has second air inlet and second gas vent, first tonifying qi branch pipe that first economizer has and second air inlet through connection, second tonifying qi branch pipe that second economizer has and first tonifying qi mouth through connection, refrigerant pressure varies in first tonifying qi branch pipe and the second tonifying qi branch pipe. The utility model provides a pair of two tonifying qi heat pump systems can realize the twice subcooling to the refrigerant, makes the system satisfy refrigeration demand and the heating demand under the extremely cold operating mode under the extremely hot operating mode.

Description

Double-air-supply heat pump system

Technical Field

The invention belongs to the technical field of air conditioning, and particularly relates to a double-air-supplementing heat pump system.

Background

The compression cycle with air supplement is a very effective way for improving the performance of a refrigeration/heat pump system, can overcome the defects of large heat pump heating capacity attenuation, poor reliability, low efficiency and the like of the traditional single-stage compressor in a low-temperature environment, and the air source heat pump technology based on the air source heat pump technology is widely applied to cold regions. In the prior art, a parallel air supply compression cycle heat pump system is disclosed, in which medium-pressure air supply and low-pressure air suction are compressed by different compressors (or different compression parts of a single compressor) respectively, so that air supply flow passage loss is reduced, the efficiency of the compressors is improved, liquid supercooling is realized by air supply, and the refrigerating/heating capacity of system circulation is increased. However, as the temperature of the low-temperature operation environment of the system is further reduced, for example, below-30 ℃, the working pressure ratio of the first compression part 107 (or the first compression part) is greatly increased, the exhaust temperature is sharply increased, and the compression efficiency and reliability of the system are both significantly reduced; in addition, as the temperature of the high-temperature operating environment of the system further increases, for example, above 60 ℃, the working pressure difference and load of the first compression part 107 increase greatly, which leads to increased wear of the bearing system of the compressor and also leads to reduced reliability and performance of the compressor.

Meanwhile, as global warming is getting more severe and extreme weather is gradually increasing, higher demands are made on the performance of the refrigeration system in extremely cold/hot weather. Researches show that the existing air supply technology still has the restriction factors of small environmental temperature range, insufficient heating capacity and low-temperature heating energy efficiency, and the use comfort, energy conservation and reliability of the air source heat pump air conditioning technology in severe cold/hot areas or extreme weather are seriously influenced. The invention is provided based on the defects in the prior art and the use requirements under extreme operation conditions.

Disclosure of Invention

Therefore, the technical problem to be solved by the present invention is to provide a dual air-supplement heat pump system, which can realize twice supercooling of a refrigerant by supplementing air twice to a first compression part and a second compression part which are connected in parallel, further reduce the enthalpy value of the refrigerant at an inlet of an evaporator, increase the enthalpy difference in the evaporation process of the refrigerant, improve the refrigeration capacity of the system, and make the system especially suitable for the refrigeration requirement under the extreme heat working condition; through twice air supplement, the exhaust flow of the compressor is improved, the flow of the refrigerant in the condenser is increased, the heating capacity of the system is improved, and the system is particularly suitable for the heating requirement under the extreme cold working condition.

In order to solve the above problems, the present invention provides a dual-air-supplement heat pump system, comprising a compression source, a first heat exchanger, a first throttling element, a first economizer, a second throttling element, a second economizer, a third throttling element, and a second heat exchanger, which are connected in sequence to form a closed cycle, wherein the compression source comprises a first compression part and a second compression part which are connected in parallel, wherein the first compression part is provided with a first exhaust port, a first air supplement port and a first air inlet connected with the second heat exchanger pipeline, the second compression part is provided with a second air inlet and a second air outlet which is connected with the first air outlet collecting pipeline, the first economizer is provided with a first air supply branch pipe which is communicated with the second air inlet, the second economizer is provided with a second air supply branch pipe which is communicated with the first air supply port, and the pressure of refrigerant in the first air supply branch pipe is different from that in the second air supply branch pipe.

Preferably, the compression source has a buffer chamber, and the first and second exhaust ports are merged in the buffer chamber.

Preferably, the buffer cavity is provided with a buffer cavity outlet, and the four-way valve further comprises a four-way valve, wherein the four-way valve is provided with a first port, a second port, a third port and a fourth port, the first port is connected with the buffer cavity outlet through a pipeline, the second port is connected with the first heat exchanger through a pipeline, the third port is connected with the first air inlet through a pipeline, and the fourth port is connected with the second heat exchanger through a pipeline.

Preferably, a first pipeline is arranged between a first air inlet of the first compression part and a third port of the four-way valve, a second pipeline is arranged between the first pipeline and a second air inlet of the second compression part, and a first on-off valve is arranged on the second pipeline.

Preferably, a third pipeline is arranged between the first air supply branch pipe and the second air inlet, and a second on-off valve is arranged on the third pipeline.

Preferably, a fourth pipeline is arranged between the second air supply branch pipe and the first air supply port, and a third shutoff valve is arranged on the fourth pipeline.

Preferably, the first economizer is one of a flash evaporator 1 or an intermediate heat exchanger 2; and/or the second economizer is one of a flash evaporator 1 or an intermediate heat exchanger 2.

Preferably, when the first economizer is an intermediate heat exchanger 2, a fifth pipeline is arranged between the first heat exchanger and the intermediate heat exchanger 2, and the fifth pipeline is connected with the first throttling element in parallel.

Preferably, the working volume of the first compression part is V1, the working volume of the second compression part is V2, and 0.02 ≦ V2/V1 ≦ 0.25.

Preferably, 0.07 ≦ V2/V1 ≦ 0.15; or V2/V1 is more than or equal to 0.02 and less than or equal to 0.15; or 0.07-V2/V1-0.25.

According to the double-air-supplement heat pump system, the refrigerant can be supercooled twice by supplementing air twice to the first compression part and the second compression part which are connected in parallel, so that the enthalpy value of the refrigerant at the inlet of the evaporator is further reduced, the enthalpy difference in the evaporation process of the refrigerant is increased, the refrigerating capacity of the system is improved, and the system is particularly suitable for the refrigerating requirement under the extreme heat working condition; through twice air supplement, the exhaust flow of the compressor is improved, the flow of the refrigerant in the condenser is increased, the heating capacity of the system is improved, and the system is particularly suitable for the heating requirement under the extreme cold working condition.

Drawings

FIG. 1 is a system schematic of a dual make-up heat pump system according to one embodiment of the present invention (in a refrigeration mode);

FIG. 2 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 1;

FIG. 3 is a system schematic diagram of a dual make-up heat pump system according to an embodiment of the present invention (in a heating mode);

FIG. 4 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 3;

FIG. 5 is a system schematic of a dual make-up heat pump system according to another embodiment of the present invention (in a refrigeration mode);

FIG. 6 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 5;

FIG. 7 is a system schematic of a dual make-up heat pump system according to another embodiment of the present invention (in a heating mode);

FIG. 8 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 7;

FIG. 9 is a system schematic of a dual make-up heat pump system according to yet another embodiment of the present invention (in a refrigeration mode);

FIG. 10 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 9;

FIG. 11 is a system schematic of a dual make-up heat pump system according to yet another embodiment of the present invention (in a heating mode);

FIG. 12 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 11;

FIG. 13 is a system schematic of a dual make-up heat pump system according to yet another embodiment of the present invention (in a refrigeration mode);

FIG. 14 is a pressure-enthalpy diagram of the heat pump system corresponding to FIG. 13;

FIG. 15 is a system schematic of a dual make-up heat pump system according to yet another embodiment of the present invention (in a refrigeration mode);

fig. 16 is a pressure-enthalpy diagram of the heat pump system corresponding to fig. 15;

fig. 17 is a graph of the effect of the volume ratio V2/V1 on the COP cycle for the heat pump system of fig. 1-4.

The reference numerals are represented as:

100. a first heat exchanger; 101. a first throttling element; 102. a first economic device; 1021. a flash evaporator; 1022. an intermediate heat exchanger; 103. a second throttling element; 104. a second economizer; 105. a third throttling element; 106. a second heat exchanger; 107. a first compression section; 108. a second compression section; 109. a buffer chamber; 110. a four-way valve; 111. a first on-off valve; 112. a second on-off valve; 113. a third shutoff valve; 114. the source is compressed.

Detailed Description

The first embodiment:

referring to fig. 1 to 2 in combination, according to an embodiment of the present invention, a dual air supplement heat pump system is provided, which includes a compression source 114, a first heat exchanger 100, a first throttling element 101, a first economizer 102, a second throttling element 103, a second economizer 104, a third throttling element 105, and a second heat exchanger 106, which are sequentially connected to form a closed cycle, wherein the compression source 114 includes a first compression part 107 and a second compression part 108 connected in parallel, the first compression part 107 has a first exhaust port, a first air supplement port and a first air inlet connected to the second heat exchanger 106 in a pipeline, the second compression part 108 has a second air inlet and a second exhaust port connected to the first exhaust port collection pipeline, the first economizer 102 has a first air supplement branch pipe connected to the second air inlet, the second economizer 104 has a second air supplement branch pipe connected to the first air supplement port, the refrigerant pressure in the first air supply branch pipe is different from that in the second air supply branch pipe. Further, the compression source 114 has a buffer chamber 109, and the first exhaust port and the second exhaust port are combined in the buffer chamber 109, so that the refrigerants of the first exhaust port and the second exhaust port are sufficiently mixed, and the phenomenon that the exhaust pressure of the compression source 114 varies due to the discontinuous exhaust of the first compression part 107 and the second compression part 108 is prevented. It can be understood that the specific cooling or heating condition in the dual air-supplement heat pump system is determined by the specific positioning of the first heat exchanger 100 and the second heat exchanger 106, for example, when the first heat exchanger 100 is used as an indoor side, the first heat exchanger 100 is used as a condenser, and when the second heat exchanger 106 is used as an indoor side, the second heat exchanger 106 is used as an evaporator, that is, the heat pump system is used in the specific cooling or heating condition, and the present invention is not limited in particular.

Preferably, the buffer cavity 109 has a buffer cavity outlet, and further includes a four-way valve 110, where the four-way valve 110 has a first port, a second port, a third port, and a fourth port, the first port is connected to the buffer cavity outlet through a pipeline, the second port is connected to the first heat exchanger 100 through a pipeline, the third port is connected to the first air inlet through a pipeline, and the fourth port is connected to the second heat exchanger 106 through a pipeline, and the four-way valve 110 is arranged to enable the heat pump system to have a dual operation mode of cooling and heating, at this time, a system schematic diagram of the heat pump system is shown in fig. 1 (the system is in a cooling condition), and a schematic flow diagram of the refrigerant during system operation is shown in fig. 1, where double-dot chain line arrows indicate a low pressure Ps (suction pressure and evaporation pressure) of the refrigerant, and single-dot chain line arrows indicate a second medium pressure P2 (i.e., a second air supplement pressure in the second air supplement branch pipe of the second economizer 104), The dashed arrow represents a first intermediate pressure P1 (a first vapor supplement pressure, i.e., the pressure of the refrigerant in the first vapor supplement branch pipe of the first economizer 102, at this time, P1 > P2), and the solid arrow represents a high refrigerant pressure Pd (discharge pressure, condensation pressure), which are the same as the following figures and will not be described again; the refrigerant circulating process: the low-pressure gaseous refrigerant 1 from the second heat exchanger 106 enters a first inlet of the first compression part 107, and in the compression process, a second medium-pressure refrigerant 10 from the system (discharged from the second economizer 104) is injected through a first supplementary gas port on the first compression part 107 to form a mixed gas 13, and is further compressed to a discharge pressure 2 with high temperature and high pressure and then discharged from the first compression part 107; the second compression unit 108 sucks the first medium-pressure refrigerant 7 from the system, compresses the refrigerant to a high-pressure discharge pressure 3, and discharges the refrigerant from the outlet of the second compression unit 108; the first discharge port of the first compression part outlet 2 and the second discharge port 3 of the second compression part 108 are mixed in the buffer chamber 109, and after flowing out of the compression source 114 from the buffer chamber discharge port 4, the mixture enters the first heat exchanger 100 to release heat (heat), and then is converted into a high-pressure liquid refrigerant 5; the two-phase refrigerant 6 throttled by the first throttling element 101 is formed into a first intermediate pressure P1; gas-liquid separation of the two-phase refrigerant 6 is achieved in the first economizer 102, and the gaseous refrigerant 7 flows into the second inlet port of the second compression part 108 through the first gas make-up branch of the first economizer 102; the liquid refrigerant 8 enters the second throttling element 103 through a liquid outlet of the first economizer 102, and is throttled and formed into the two-phase refrigerant 9 with the second intermediate pressure P2 again; the second economizer 104 realizes gas-liquid separation of the two-phase refrigerant 9, and the gaseous refrigerant 10 flows into the first gas supplementing port of the first compression part 107 through the second gas supplementing branch pipe of the second economizer 104; the liquid refrigerant 11 enters the third throttling element 105 through the liquid outlet of the second economizer 104, forms a low-pressure low-temperature two-phase refrigerant 12 after being throttled, then enters the second heat exchanger 106 to evaporate and absorb heat (refrigerate), forms a low-temperature low-pressure refrigerant 1, and is sucked by the first compression part 107.

Fig. 2 shows a pressure-enthalpy diagram (solid line) of the refrigeration cycle of the foregoing technical solution, and simultaneously shows a pressure-enthalpy diagram (dashed line) corresponding to the refrigeration cycle of the single-stage compressor under the same working condition, as can be clearly seen from the diagram, the working process of the single-stage compressor in the prior art is 1-4 ', since the second intermediate-pressure refrigerant 10 is supplemented into the first compression part 107, the working process of the first compression part 107 is 1-13-2, the single-stage compression is divided into two-stage compression, and the discharge temperature is reduced from T4' to T2, the refrigerant flow of the first compression part 107 is increased from m0(m0 is the discharge flow corresponding to the first compression part 107 without being supplemented with air) to m0+ m2(m2 is the flow corresponding to the second intermediate-pressure refrigerant 10); the working process of the second compression part 108 is 7-3, the temperature of the discharge gas 3 is lower than that of the discharge gas 2 of the first compression part, the temperature of the mixed discharge gas is further reduced from T2 to T4, and the total flow rate m of the compression source is further increased to m0+ m2+ m1(m1 is the flow rate of the refrigerant corresponding to the second compression part 108); 3-5 in the pressure-enthalpy diagram is a condensation process, so that heating is realized; 5-6 are the first throttling element 101 throttling process; 6-7, 6-8 is the working process of the first economizer 102, the refrigerant supercooling (first supercooling of the refrigerant) of the first compression part 107 is realized by separating the first medium pressure gaseous refrigerant 7, and the enthalpy difference h6-h7 of the refrigerant is increased; 8-9 is the throttling process of the second throttling element 103, 9-10 and 9-11 are the working processes of the second economizer 104, and the refrigerant in the first compression part 107 is further sub-cooled (the second sub-cooling of the refrigerant) by separating the second medium-pressure gaseous refrigerant 10, so that the enthalpy difference h9-h10 of the refrigerant is increased; 11-12 is the working process of the third throttling element 105, and 12-1 is the evaporation process, so as to realize refrigeration.

As can be seen from the foregoing, through comparative analysis of the pressure-enthalpy diagram (fig. 2) of the present application technical solution and the conventional single-stage compression cycle, the present application scheme realizes two times of supercooling enthalpy increases of the refrigerant in the first compression part 107 of the heat pump system, the enthalpy difference of the refrigerant in the evaporation process is increased to h1-h12 from h1-h5 compared with the single-stage compression cycle, the enthalpy difference is further increased compared with the single intermediate pressure air-filling system, and the refrigeration capacity of the system is increased; compared with single-stage compression circulation, the exhaust flow of the compressor is increased from m0 to m0+ m1+ m2, and the exhaust flow is further increased compared with a traditional single intermediate pressure air supply system, so that the heating capacity is improved; the compressor is through twice tonifying qi, and first compression portion is compressed by the single-stage compression and is decomposed into the second grade and compress, has reduced exhaust temperature, has improved the compression process, has improved compressor efficiency and reliability, has enlarged pressure simultaneously and has compared application scope for refrigerating system's operating temperature scope is bigger. The double-air-supplement heat pump system adopting the technical scheme can realize the double supercooling of the refrigerant by supplementing air for two times to the first compression part and the second compression part which are connected in parallel, further reduce the enthalpy value of the refrigerant at the inlet of the evaporator, increase the enthalpy difference in the evaporation process of the refrigerant, improve the refrigerating capacity of the system and ensure that the system is particularly suitable for the refrigerating requirement under the extreme heat working condition; through twice air supplement, the exhaust flow of the compressor is improved, the flow of the refrigerant in the condenser is increased, the heating capacity of the system is improved, and the system is particularly suitable for the heating requirement under the extreme cold working condition.

Fig. 3 and 4 respectively show a system schematic diagram of the heat pump system of the first embodiment of the present application in a heating operation condition and a corresponding pressure-enthalpy diagram. The first inlet port of the second compression part 108 is supplemented with refrigerant 10 of the second intermediate pressure P2, and the supplementary inlet port of the first compression part 107 is supplemented with refrigerant 7 of the first intermediate pressure P1 (in this case, P1 > P2). Fig. 4 shows the pressure-enthalpy diagram of the operation of the compressor of this embodiment, the operation of the first compression part 107 is still 1-13-2, but the supplementary gas is set as the first intermediate-pressure refrigerant 7, and the discharge flow rate of the first compression part 107 is m0+ m 1; the operation of the second compression part 108 is 10-3 and the exhaust gas flow rate is m 2. Because the pressure adjustment is supplemented to first compression part 107 for the higher first middling pressure P1 of pressure for the pressure ratio of first compression part 107 distributes more rationally, thereby has further improved the compression efficiency of first compression part 107, makes the compressor more be applicable to the operating mode of big pressure ratio, heats under the ultra-low temperature environment for example.

FIG. 17 shows the variation of COP lift amplitude of the dual air-replenishing circulation system with volume ratio V2/V1. For the dual make-up refrigeration cycle system of the present application, the ratio V2/V1 of the swept volume V2 of the second compression section 108 to the swept volume V1 of the first compression section 107 has a significant impact on system performance. During the cooling and heating operation, because the working pressure difference and the working pressure ratio of the compressor are different, how to set the volume ratio V2/V1 to improve the working process of the compressor and the air supplementing effect is very important for improving the performance of the cooling system. The volume ratio V2/V1 is set between 0.02 and 0.25, and the COP (coefficient of performance) of refrigeration and heating is relatively excellent; when the volume ratio V2/V1 is 0.02-0.15, the refrigeration energy efficiency is better; when the volume ratio is 0.07-0.25, the heating energy efficiency is better; particularly, when the volume ratio V2/V1 is 0.07-0.15, the refrigeration COP and the heating COP are both better.

The first compression part 107 and the second compression part 108 may be two independent compressors, or may be integrated into the same compressor; wherein the first compression part 107 is understood to be a secondary compressor (part) or a quasi-secondary compressor (part) with the function of air-supplementing and enthalpy-increasing; likewise, the buffer chamber 109 may exist independently of the compression source or may be configured to be integrated within the compression source. The first throttling element 101, the second throttling element 103 and the third throttling element 105 may be implemented by using conventional capillary tubes, expansion valves and other elements, and the present invention is not particularly limited.

Second embodiment:

in addition to the heat pump system of the first embodiment, a first pipeline is provided between the first air inlet of the first compression portion 107 and the third port of the four-way valve 110, a second pipeline is provided between the first pipeline and the second air inlet of the second compression portion 108, and the second pipeline is provided with a first on-off valve 111. Preferably, a third pipeline is arranged between the first air supply branch pipe and the second air inlet, and a second cut-off valve 112 is arranged on the third pipeline.

Specifically, fig. 5 shows a schematic view of a second embodiment of the refrigeration system of the present application (in a refrigeration condition). The difference from the first embodiment is that the second inlet port of the second compression portion 108 communicates with the second heat exchanger 106, and a first on-off valve 111 is provided on the communication path (i.e., the first branch gas supply pipe); a second intake port of the second compression portion 108 communicates with the first economizer 102, and a second cut-off valve 112 is provided on the communication path; the first on-off valve 111 is opened, the second on-off valve 112 is closed, that is, the second compression part 108 sucks the low-pressure refrigerant, and the first gas supply branch pipe is disconnected. The first on-off valve 111 and the second on-off valve 112 switch the first inlet refrigerant pressure of the second compression part 108, and may be implemented by a three-way valve or a four-way valve.

Fig. 6 shows the pressure-enthalpy diagram of the refrigerant cycle (in the cooling condition) of the second embodiment. Since the first compression part 107 sucks the refrigerant at m0 and the second compression part 108 sucks the refrigerant at m1, the refrigerant flow of the second heat exchanger 106 is m0+ m1, and the refrigerant flow of the evaporator is increased; 1-3 is the working process of the second compression part 108; 1-13-2 is the operation of the first compression section 107, and the exhaust flow rate is m0+ m2 due to the injection of the second intermediate pressure 10. In the operation mode of this embodiment, the enthalpy difference during the evaporation process of the refrigerant is also h1-h12, but the refrigerant flow rate of the evaporator is m0+ m2, and as can be seen from the cooling capacity calculation formula Q ═ (h1-h12) × (m0+ m2), the cooling capacity is further increased as compared with the first embodiment. The embodiment is particularly suitable for operating under the working conditions of large refrigerating capacity demand and moderate working pressure ratio (such as the refrigeration at the ambient temperature of 35-50 ℃), and has the characteristics of large refrigerating capacity and high operating efficiency.

Fig. 7 shows a schematic view of a second embodiment of the refrigeration system of the present application (in a heating condition). This embodiment has the following features compared to the first embodiment described above: a second air inlet of the second compression part 108 is communicated with the evaporator (the first heat exchanger 100), and a first on-off valve 111 is arranged on a communication path (a first air make-up branch pipe); a second air inlet of the second compression part 108 is communicated with the first air replenishing branch pipe, and a second cut-off valve 112 is arranged on the communication path; the first on-off valve 111 is opened, the second on-off valve 112 is closed, that is, the second compression part 108 sucks the low-pressure refrigerant, and the first branch gas replenishing pipe is disconnected from the second gas inlet of the second compression part 108. The first and second on-off

valves

111 and 112 switch the pressure of the refrigerant at the suction port of the second compression unit 108.

Fig. 8 shows the pressure-enthalpy diagram of the refrigerant cycle (in the heating condition) of the second embodiment. First compression unit 107 sucks the refrigerant at a flow rate of m0, and second compression unit 108 sucks the refrigerant at a flow rate of m 2; 1-3 is the working process of the second compression part 108; 1-13-2 is the operation of the first compression section 107, and the exhaust flow rate is m0+ m1 due to the injection of the first intermediate pressure 7. In the operation mode of this embodiment, the enthalpy difference during the evaporation of the refrigerant is also h1 to h12, but the refrigerant flow rate of the evaporator is m0+ m2, and the cooling capacity calculation formula Q ═ m1 to h12 (m0+ m2), as can be seen, the cooling capacity is further increased, and the heating capacity Qc is equal to Q + W and the heating capacity is further increased by the conservation of the cycle energy, as compared with the first embodiment. The embodiment is particularly suitable for operating under the working conditions of large heating quantity requirement and high working pressure ratio (such as heating at the ambient temperature of minus 7 ℃ to minus 20 ℃), and has the characteristics of large heating quantity and high operating efficiency.

The third embodiment:

on the basis of the first embodiment, a fourth pipeline is arranged between the second air supply branch pipe and the first air supply port, and a third shutoff valve 113 is arranged on the fourth pipeline.

Specifically, fig. 9 shows a schematic view of a third embodiment of the refrigeration system of the present application (in a refrigeration condition). This embodiment has the following features compared to the first embodiment: a third cut-off valve 113 is arranged on a communication path between the first air supplement port of the first compression part 107 and the second air supplement branch pipe of the second economizer 104; when the third shutoff valve 113 is opened, the heat pump system is the first embodiment (refrigeration condition), which is not described herein; when the third shutoff valve 113 is closed, the gas supplementing passage of the first compression part 107 is closed, and the refrigerant 7 of the first intermediate pressure P1 is sucked into the second compression part 108.

Fig. 10 shows the pressure-enthalpy diagram of the refrigerant cycle (in the cooling condition) of the third embodiment. The first compression section 107 has a refrigerant suction flow rate of m0, the second compression section 108 has a refrigerant suction flow rate of m1, and the total discharge flow rate is m0+ m 1; 7-3 is the working process of the second compression part 108, 1-2 is the working process of the first compression part 107, and the system cycle is a single air make-up parallel compression refrigeration cycle (at this time, the third shutoff valve 113 is disconnected). The embodiment has the advantages of simple cycle process and small air supply flow loss, is particularly suitable for running under the working conditions of higher energy efficiency requirement, lower refrigeration/heating quantity requirement and lower working pressure ratio (such as refrigeration at the ambient temperature of 30-35 ℃), and has the characteristic of high system refrigeration/heating efficiency.

Fig. 11 shows a schematic view of a third embodiment of the refrigeration system of the present application (in a heating condition). At this time, a third shut-off valve 113 is provided on a communication path between the first gas replenishing port of the first compression part 107 and the second gas replenishing branch pipe of the second economizer 104; when the third shutoff valve 113 is opened, the heat pump system is the first embodiment (heating condition), which is not described herein; when the third shutoff valve 113 is closed, the first compression portion 107 gas replenishing passage is closed, and the second compression portion 108 sucks the refrigerant 10 of the second intermediate pressure P2.

Fig. 12 shows the pressure-enthalpy diagram of the refrigerant cycle (in the heating condition) of the third embodiment. The first compression section 107 has a refrigerant suction flow rate of m0, the second compression section 108 has a refrigerant suction flow rate of m2, and the total discharge flow rate is m0+ m 2; 10-3 is the working process of the second compression part 108, 1-2 is the working process of the first compression part 107, and the system cycle is a single air make-up parallel compression refrigeration cycle (at this time, the third shutoff valve 113 is disconnected). The embodiment has the advantages of simple cycle process and small air supply flow loss, is particularly suitable for operating under the working conditions of higher energy efficiency requirement, lower refrigeration/heating quantity requirement and lower working pressure ratio (such as when the environment temperature is minus 7 ℃ to 7 ℃) and has the characteristic of high system refrigeration/heating efficiency.

The fourth embodiment:

on the basis of the second embodiment, a fourth pipeline is further arranged between the second air supplement branch pipe and the first air supplement port, and a third shutoff valve 113 is arranged on the fourth pipeline.

Specifically, fig. 13 shows a schematic view of a fourth embodiment of the heat pump system of the present application (in a refrigeration mode). This embodiment has the following features compared to the second embodiment described above: a third shutoff valve 113 is arranged on a communication path (a second gas supply branch pipe) between the first gas supply port of the first compression part 107 and the second economizer 104; when the third on-off valve 113 is opened, the second on-off valve 112 is closed, and the first on-off valve 111 is opened, the system is the second embodiment; when the third on-off valve 113 is opened, the second on-off valve 112 is opened, and the first on-off valve 111 is closed, the system is the first embodiment; when the third on-off valve 113 is closed, the second on-off valve 112 is opened, and the first on-off valve 111 is closed, the system is the third embodiment; when the third cut-off valve 113 is closed, the second cut-off valve 112 is closed, and the first cut-off valve 111 is opened, both the two gas make-up passages of the compression source 114 are closed, and the second compression portion 108 sucks the low-pressure refrigerant 1.

Fig. 14 shows the pressure-enthalpy diagram of the refrigerant cycle (in the cooling condition) of the fourth embodiment. The system cycle is a single-machine compression refrigeration cycle, the working process of the first compression part 107 is 1-2, and the working process of the second compression part 108 is 1-3. The system circulation of the embodiment does not have the air supplementing function, is particularly suitable for the operation working conditions with small pressure ratio and unsuitable for air supplementing, such as the refrigeration working conditions with the environmental temperature lower than 30 ℃ or the heating working conditions with the environmental temperature higher than 7 ℃ and the like, and can avoid the system flow loss caused by air supplementing.

Fifth embodiment:

preferably, the first economizer 102 is one of a flash evaporator 1021 or an intermediate heat exchanger 1022; and/or the second economizer 104 is one of a flash evaporator 1021 or an intermediate heat exchanger 1022.

Specifically, the flash evaporator 1021 is adopted as the first economizer 102 or the second economizer 104 in the first to fourth embodiments, in this embodiment, the intermediate heat exchanger 1022 is adopted as the first economizer 102, and in this case, a fifth pipeline is further provided between the first heat exchanger 100 and the intermediate heat exchanger 1022, and the fifth pipeline is connected in parallel with the first throttling element 101.

Fig. 15 shows a fifth embodiment of the present application, featuring a refrigeration cycle system wherein the first intermediate pressure P1 refrigerant is provided using an intermediate heat exchanger 1022 and the second intermediate pressure P2 refrigerant is provided using a flash evaporator 1021. A part of the high-pressure high-temperature liquid refrigerant 5 is branched off and throttled by a first throttling element 101 to be medium-pressure medium-temperature refrigerant 6; the refrigerants of the first compression parts 107 of 6 and 5 exchange heat in the intermediate heat exchanger 1022; 6 evaporates and absorbs heat to become gaseous refrigerant 7 of the first intermediate pressure P1 while refrigerant 5 of the first compression part 107 releases heat to be further cooled to 8, thereby achieving the first supercooling of refrigerant; the process of the flash evaporator 1021 to generate the second intermediate pressure P2 gaseous refrigerant is the same as described above and will not be described further.

Fig. 16 shows a cycle pressure-enthalpy diagram of the fifth embodiment, 6 to 7 are processes of absorbing heat by evaporation of the refrigerant of the

second compression part

108, and 5 to 8 are processes of cooling by heat release of the refrigerant of the first compression part 107, and the intermediate heat exchanger 1022 achieves heat exchange of the refrigerant m1 of the second compression part 108 and the refrigerant m0+ m2 of the first compression part 107. The advantage of this approach is that the auxiliary air make-up (for the second compression section 108) and the refrigerant in the first compression section 107 are independently controlled, and the air make-up flow is more easily regulated.

Sixth embodiment:

according to an embodiment of the present invention, there is also provided a control method for a dual gas-supply heat pump system, for controlling the dual gas-supply heat pump system, including the following steps:

acquiring operation modes of a heat pump system, wherein the operation modes can comprise an ultrahigh-temperature refrigeration mode/an ultralow-temperature heating mode, a high-temperature refrigeration mode/a low-temperature heating mode and a conventional refrigeration mode/a conventional heating mode;

the flow path switching of the four-way valve 110 is controlled and the on/off switching of the first on/off valve 111, the second on/off valve 112, and the third on/off valve 113 is controlled according to the operation mode.

Specifically, the mode definitions of the ultra-high temperature cooling mode/ultra-low temperature heating mode, the high temperature cooling mode/low temperature heating mode, the higher temperature cooling mode/lower temperature heating mode, and the conventional cooling mode/conventional heating mode are based on the specific temperature range of the temperature-regulated space, for example, the ambient temperature in the ultra-high temperature cooling mode is higher than 50 ℃, and the ambient temperature in the ultra-low temperature heating mode is lower than-20 ℃; the environment temperature corresponding to the high-temperature refrigeration mode is between 35 ℃ and 50 ℃, and the environment temperature corresponding to the low-temperature heating mode is between-20 ℃ and-7 ℃; the environment temperature corresponding to the higher temperature refrigeration mode is between 30 ℃ and 35 ℃, and the environment temperature corresponding to the lower temperature heating mode is between 7 ℃ below zero and 7 ℃; the normal cooling mode corresponds to an ambient temperature of less than 30 deg.c and the normal heating mode corresponds to an ambient temperature of greater than 7 deg.c. Preferably, when the operation mode is an ultra-high temperature refrigeration mode, the first port and the second port of the four-way valve 110 are controlled to be communicated, the third port and the fourth port are controlled to be communicated, and the first on-off valve 111, the second on-off valve 112 and the third on-off valve 113 are controlled to be communicated; or, when the operation mode is the ultra-low temperature heating mode, the first port and the fourth port of the four-way valve 110 are controlled to be communicated, the second port and the third port are controlled to be communicated, and the first on-off valve 111, the second on-off valve 112 and the third on-off valve 113 are controlled to be communicated.

Preferably, when the operation mode is a high-temperature cooling mode, the first port and the second port of the four-way valve 110 are controlled to be communicated, the third port and the fourth port are controlled to be communicated, and the second on-off valve 112, the first on-off valve 111 and the third on-off valve 113 are controlled to be communicated; or, when the operation mode is the low-temperature heating mode, the first port and the fourth port of the four-way valve 110 are controlled to be communicated, the second port and the third port are controlled to be communicated, and the second on-off valve 112, the first on-off valve 111 and the third on-off valve 113 are controlled to be communicated.

Preferably, when the operation mode is a higher-temperature cooling mode, the first port and the second port of the four-way valve 110 are controlled to be communicated, the third port and the fourth port are controlled to be communicated, and the first on-off valve 111, the third on-off valve 113 and the second on-off valve 112 are controlled to be cut off and communicated; or, when the operation mode is the lower temperature heating mode, the first port and the fourth port of the four-way valve 110 are controlled to be communicated, the second port and the third port are controlled to be communicated, and the first on-off valve 111, the third on-off valve 113 are controlled to be closed, and the second on-off valve 112 is controlled to be communicated.

Preferably, when the operation mode is a normal cooling mode, the first port and the second port of the four-way valve 110 are controlled to be communicated, the third port and the fourth port are controlled to be communicated, and the second on-off valve 112, the third on-off valve 113 are controlled to be cut off, and the first on-off valve 111 is controlled to be communicated; or, when the operation mode is the normal heating mode, the first port and the fourth port of the four-way valve 110 are controlled to be communicated, the second port and the third port are controlled to be communicated, and the second on-off valve 112, the third on-off valve 113 are controlled to be closed, and the first on-off valve 111 is controlled to be communicated.

In summary, the dual air-supplement heat pump system provided by the present invention has the parallel compressor for compressing dual air supplement, except for the second compression part 108 for the first air supplement, the second air supplement is performed into the first compression part 107, so as to reduce the exhaust temperature of the first compression part 107 under the extreme working condition, improve the compression efficiency of the first compression part 107, and improve the working efficiency of the first compression part 107; meanwhile, the refrigerant is supercooled twice, compared with the prior art, the enthalpy value of the refrigerant at the inlet of the evaporator is further reduced, and the enthalpy difference in the evaporation process of the refrigerant is increased, so that the refrigerating capacity of the system is improved; through twice air supply, the exhaust flow of the compressor is improved, and the flow of the refrigerant in the condenser is increased, so that the heating capacity of the system is improved. Secondly, aiming at different operating conditions, a plurality of operating modes and switching methods are provided, and the operating modes can be partially or completely used according to the using range or the requirement of the system, so that the comprehensive operating efficiency of the refrigeration system is improved. In addition, the optimal range of the volume ratio of the compressor is provided, the flow distribution characteristics of the first medium-pressure refrigerant and the second medium-pressure refrigerant are improved, the air supplementing effect of the auxiliary compression part is optimal, and the refrigerating/heating performance improving effects of the compressor and the refrigerating system are optimal. Compared with a traditional three-stage compression double-air-supply system, the technical scheme has the technical advantages of simple structure, short flow path inside the compressor, small flow loss, high efficiency and low cost.

It is readily understood by a person skilled in the art that the advantageous ways described above can be freely combined, superimposed without conflict.

The present invention is not limited to the above preferred embodiments, and any modifications, equivalent substitutions and improvements made within the spirit and principle of the present invention should be included in the protection scope of the present invention. The above is only a preferred embodiment of the present invention, and it should be noted that, for those skilled in the art, several improvements and modifications can be made without departing from the technical principle of the present invention, and these improvements and modifications should also be regarded as the protection scope of the present invention.

Claims

1. A double-air supplement heat pump system is characterized by comprising a compression source (114), a first heat exchanger (100), a first throttling element (101), a first economizer (102), a second throttling element (103), a second economizer (104), a third throttling element (105) and a second heat exchanger (106) which are sequentially connected to form a closed cycle, wherein the compression source (114) comprises a first compression part (107) and a second compression part (108) which are connected in parallel, the first compression part (107) is provided with a first exhaust port, a first air supplement port and a first air inlet connected with a pipeline of the second heat exchanger (106), the second compression part (108) is provided with a second air inlet and a second exhaust port connected with a collecting pipeline of the first exhaust port, the first economizer (102) is provided with a first air supplement branch pipe communicated with the second air inlet, and the second economizer (104) is provided with a second air supplement branch pipe communicated with the first air supplement port, the refrigerant pressure in the first air supply branch pipe is different from that in the second air supply branch pipe.

2. The heat pump system of claim 1, wherein the compression source (114) has a buffer chamber (109), and the first and second exhaust ports are combined in the buffer chamber (109).

3. A heat pump system according to claim 1 or 2, wherein said buffer chamber (109) has a buffer chamber outlet, further comprising a four-way valve (110), said four-way valve (110) having a first port in line with said buffer chamber outlet, a second port in line with said first heat exchanger (100), a third port in line with said first air inlet, and a fourth port in line with said second heat exchanger (106).

4. The heat pump system according to claim 3, wherein a first pipeline is provided between a first inlet of the first compression part (107) and a third port of the four-way valve (110), a second pipeline is provided between the first pipeline and a second inlet of the second compression part (108), and a first on-off valve (111) is provided on the second pipeline.

5. The heat pump system according to claim 4, wherein a third pipeline is arranged between the first branch gas supply pipe and the second gas inlet, and a second cut-off valve (112) is arranged on the third pipeline.

6. The heat pump system according to any one of claims 1, 2, 4, and 5, wherein a fourth pipeline is arranged between the second branch air supplement port and the first air supplement port, and a third shut-off valve (113) is arranged on the fourth pipeline.

7. The heat pump system of claim 6, wherein the first economizer (102) is one of a flash evaporator (1021) or an intermediate heat exchanger (1022); and/or the second economizer (104) is one of a flash evaporator (1021) or an intermediate heat exchanger (1022).

8. The heat pump system according to claim 7, wherein when the first economizer (102) is an intermediate heat exchanger (1022), there is also a fifth circuit between the first heat exchanger (100) and the intermediate heat exchanger (1022), the fifth circuit being in parallel with the first throttling element (101).

9. The heat pump system according to claim 6, wherein the displacement volume of the first compression part (107) is V1, and the displacement volume of the second compression part (108) is V2, 0.02 ≦ V2/V1 ≦ 0.25.

10. The heat pump system of claim 9, wherein 0.07 ≦ V2/V1 ≦ 0.15; or V2/V1 is more than or equal to 0.02 and less than or equal to 0.15; or 0.07-V2/V1-0.25.