Non-azeotropic mixed working medium heat pump system and control method thereof

Abstract

The invention relates to the technical field of heat pumps, and discloses a non-azeotropic mixed working medium heat pump system and a control method thereof.The heat pump system comprises a compressor and a refrigerant circulation loop, and a condenser, a first throttling device, a heat regenerator, a second throttling device and an evaporator are sequentially arranged on the refrigerant circulation loop; a first heat exchange channel of the heat regenerator is connected between the first throttling device and the second throttling device, and a second heat exchange channel of the heat regenerator is connected between an outlet of the evaporator and an air return port of the compressor; the first liquid spraying pipeline is connected between the first heat exchange channel and an outlet of the evaporator and is provided with a first control valve; and the second liquid spraying pipeline is connected between the outlet of the condenser and the outlet of the evaporator and is provided with a second control valve. The two liquid spraying pipelines are arranged, the exhaust temperature can be accurately controlled by controlling the control valves of the two liquid spraying pipelines, the situation that the exhaust temperature of the heat pump system is too high due to the low-temperature working condition is avoided, the compressor can be protected against high-temperature damage, the service life is prolonged, and the system reliability is improved.

Description

Technical Field

[0001] The present invention relates to the technical field of heat pumps, and particularly to azeotropic mixture refrigerant heat pump system and its control method. Background Art

[0002] With the increasing global attention to the utilization of renewable energy and energy efficiency improvement, heat pump systems, as efficient and environmentally friendly energy conversion devices, have been widely used in civil and industrial fields. However, when a heat pump system operates in an ultra-low temperature environment, due to the extremely low evaporation temperature, the pressure on the low-pressure side of the system is very low, which not only increases the working burden of the compressor, but also makes the compressor need to overcome a greater pressure ratio to complete the refrigerant circulation process, and will cause the exhaust temperature of the compressor to be too high. The above-mentioned high-pressure difference working state and too high exhaust temperature will greatly increase the energy consumption of the compressor and may shorten the service life of the compressor.

[0003] In addition, since an ultra-low temperature heating heat pump needs to provide sufficient heat under extreme conditions, the compressor must operate at a higher efficiency, resulting in an increase in the exhaust temperature. Too high exhaust temperature may cause problems such as carbonization of lubricating oil and aging of sealing materials, and in severe cases, it will cause compressor failure or even damage, affecting the stability and reliability of the entire system. Summary of the Invention

[0004] The first technical problem to be solved by the present invention is to provide an azeotropic mixture refrigerant heat pump system, which effectively solves the problem that the exhaust temperature of the compressor is too high when the heat pump system operates in a low temperature environment in the related art.

[0005] The second technical problem to be solved by the present invention is to provide a control method for an azeotropic mixture refrigerant heat pump system, which effectively solves the problem that the exhaust temperature of the compressor is too high when the heat pump system operates in a low temperature environment in the related art.

[0006] The above first technical problem is solved by the following technical solutions: An azeotropic mixture refrigerant heat pump system, comprising: A compressor and a refrigerant circulation loop connected to the compressor and forming a loop, and a condenser, a first throttling device, a regenerator, a second throttling device and an evaporator are sequentially arranged on the refrigerant circulation loop along the refrigerant flow direction; the regenerator has a first heat exchange channel and a second heat exchange channel for heat exchange with each other, the first heat exchange channel is connected between the first throttling device and the second throttling device, and the second heat exchange channel is connected between the outlet of the evaporator and the suction port of the compressor; A first liquid injection pipeline is connected between the outlet of the first heat exchange channel and the outlet of the evaporator, and a first control valve is arranged on the first liquid injection pipeline; A second liquid injection pipeline is connected between the outlet of the condenser and the outlet of the evaporator, and a second control valve is provided on the second liquid injection pipeline.

[0007] Compared with the background art, the non-azeotropic mixture refrigerant heat pump system of the present invention has the following beneficial effects: Aiming at the high-temperature exhaust problem caused by component segregation and increased pressure ratio when the non-azeotropic mixture refrigerant operates at ultra-low temperature, the present invention is provided with a liquid injection pipeline, which can transport the liquid refrigerant in the system to the outlet of the evaporator. At this time, the liquid refrigerant is mixed with the gaseous refrigerant flowing out of the evaporator outlet, reducing the temperature of the refrigerant, thereby effectively reducing the suction temperature of the compressor. It can be understood that when the suction temperature of the compressor drops, correspondingly, the discharge temperature of the compressor will also drop. Furthermore, the present invention realizes the technical effect of suppressing the rise of the discharge temperature, avoids the excessive discharge temperature of the heat pump system due to low-temperature working conditions, helps to protect the compressor from high-temperature damage, extends its service life and improves the reliability of the system, and is particularly suitable for the high-temperature exhaust problem caused by component segregation and increased pressure ratio when the non-azeotropic mixture refrigerant operates at ultra-low temperature. At the same time, by setting the first throttling device, the regenerator and the second throttling device, the present application can perform staged throttling and regeneration on the mixed refrigerant before it flows through the evaporator, thereby reducing the throttling loss and further increasing the heating capacity.

[0008] Furthermore, the first liquid injection pipeline is used to extract the liquid refrigerant from the outlet of the regenerator and directly inject it into the gaseous refrigerant at the outlet of the evaporator. By reducing the temperature of the refrigerant, the suction temperature of the compressor is reduced, and then the discharge temperature of the compressor is reduced. Among them, since the liquid injection point is behind the regenerator, the refrigerant has been throttled and subcooled, and the pressure difference of the refrigerant relative to the compressor suction port is small and the temperature is initially reduced. Therefore, the liquid injection volume is small, and the influence on the heating capacity of the system is small, and the discharge temperature can be reduced by 5°C - 10°C.

[0009] The second liquid injection pipeline is used to extract the medium-temperature and high-pressure liquid refrigerant from the outlet of the condenser and directly inject it into the gaseous refrigerant at the outlet of the evaporator. Since the liquid injection point of the second liquid injection pipeline is behind the condenser, the pressure difference of the refrigerant relative to the compressor suction port is large. Therefore, the liquid injection volume is large. By enabling the second liquid injection pipeline, the suction temperature of the compressor can be significantly reduced, and the discharge temperature can be further reduced by 10°C - 20°C. However, under this condition, the compressor power increases and the system energy efficiency decreases.

[0010] Therefore, on the premise described above, through the phased liquid injection design of the first liquid injection pipeline and the second liquid injection pipeline, the present invention can achieve precise control of the exhaust temperature and dynamic balance of the heating performance. For example, under low-temperature conditions, the first control valve is preferentially opened for mild liquid injection. If the exhaust temperature is still higher than the set threshold, the second control valve is gradually opened for deep liquid injection. In this way, by adjusting the opening degrees of the two control valves, stepped control of the exhaust temperature is achieved, ensuring the stability of the system under extremely low temperatures and avoiding a significant decrease in the heating energy efficiency of the heat pump system due to excessive liquid injection.

[0011] In addition, the two liquid injection pipelines can also be selectively opened under different working conditions, thereby achieving refined control of the exhaust temperature to achieve the best temperature management effect. For example, under light load or medium load, since the liquid injection point of the first liquid injection pipeline is behind the regenerator and the refrigerant has passed through throttling and subcooling, the pressure difference between the refrigerant and the compressor suction port is small, so the liquid injection volume is small. Only opening the first liquid injection pipeline can reduce the exhaust temperature by 5°C - 10°C, and the impact on the heating capacity of the system is small; while under high load or extreme conditions, since the liquid injection point of the second liquid injection pipeline is behind the condenser and the pressure difference between the refrigerant and the compressor suction port is large, the liquid injection volume is large. By enabling the second liquid injection pipeline, the exhaust temperature is further reduced by 10°C - 20°C, thereby further reducing the exhaust temperature while ensuring the heating capacity and ensuring the stable operation of the system in various environments.

[0012] In one embodiment, the zeotropic mixture refrigerant heat pump system further includes: A liquid receiver, provided with a refrigerant inlet and a first refrigerant outlet. The refrigerant inlet is connected to the outlet of the condenser, and the first refrigerant outlet is connected to the first heat exchange channel.

[0013] In one embodiment, the zeotropic mixture refrigerant heat pump system further includes: A defrosting branch. The liquid receiver is further provided with a second refrigerant outlet. The two ends of the defrosting branch are respectively communicated with the second refrigerant outlet and the inlet of the evaporator, and an electronically controlled valve with adjustable opening degree is provided on the defrosting branch.

[0014] In one embodiment, the second refrigerant outlet is arranged at the top of the liquid receiver, and the first refrigerant outlet is arranged at the bottom of the liquid receiver.

[0015] In one embodiment, the zeotropic mixture refrigerant heat pump system further includes: An evaporator fan for driving air to flow through the evaporator; A gas-liquid separator, arranged on the refrigerant circulation loop and between the second heat exchange channel and the suction port of the compressor.

[0016] The above second technical problem is solved by the following technical solution: A control method for azeotropic refrigerant mixture heat pump system, which is applied to the azeotropic refrigerant mixture heat pump system described in the first aspect of the present invention. The control method includes: In response to the heating mode, obtaining the actual exhaust temperature T of the compressor; Comparing the actual exhaust temperature T with the maximum exhaust temperature t2 and the minimum exhaust temperature t1 respectively to obtain a determination result; Based on the determination result, controlling the working state of at least one of the first control valve and the second control valve.

[0017] Compared with the background art, the beneficial effects of the control method for the azeotropic refrigerant mixture heat pump system of the present invention are as follows: Aiming at the high-temperature exhaust problem caused by component segregation and increased pressure ratio when the azeotropic refrigerant mixture operates at ultra-low temperature, the control method of the present invention controls the working states of the first control valve and the second control valve, so as to select whether the liquid injection pipeline sprays liquid at the outlet of the evaporator. Among them, in the case of liquid injection, that is, when at least one of the first control valve and the second control valve is opened, the liquid refrigerant in the liquid injection pipeline can be transported to the outlet of the evaporator. At this time, the liquid refrigerant is mixed with the gaseous refrigerant flowing out of the evaporator outlet, reducing the temperature of the refrigerant, thereby effectively reducing the suction gas temperature of the compressor. It can be understood that when the suction gas temperature of the compressor decreases, correspondingly, the exhaust temperature of the compressor will also decrease. Furthermore, the present invention realizes the technical effect of suppressing the rise of the exhaust temperature, avoids the too high exhaust temperature of the heat pump system due to low-temperature working conditions, helps to protect the compressor from high-temperature damage, extends its service life and improves the reliability of the system. It is particularly applicable to the high-temperature exhaust problem caused by component segregation and increased pressure ratio when the azeotropic refrigerant mixture operates at ultra-low temperature. At the same time, by setting the first throttling device, the regenerator and the second throttling device, the present application can perform staged throttling and heat regeneration on the mixed refrigerant before the mixed refrigerant flows through the evaporator, thereby reducing the throttling loss and further increasing the heating capacity.

[0018] Furthermore, when the first control valve is opened, the first liquid injection pipeline is used to extract the liquid refrigerant from the outlet of the regenerator and directly inject it into the gaseous refrigerant at the outlet of the evaporator. By reducing the temperature of the refrigerant, the suction gas temperature of the compressor is reduced, and then the exhaust temperature of the compressor is reduced. Among them, since the liquid injection point is behind the regenerator, the refrigerant has been throttled and subcooled, and the pressure difference between the refrigerant and the compressor suction port is small, and the temperature is initially reduced. Therefore, the liquid injection amount is small, the influence on the heating capacity of the system is small, and the exhaust temperature can be reduced by 5°C - 10°C.

[0019] When the second control valve is opened, the second liquid injection pipeline is used to extract high-temperature and high-pressure liquid refrigerant from the condenser outlet and directly inject it into the gaseous refrigerant at the evaporator outlet. Since the liquid injection point of the second liquid injection pipeline is behind the condenser, the pressure difference of the refrigerant relative to the compressor suction port is large, so the liquid injection volume is large. By enabling the second liquid injection pipeline, the compressor suction temperature can be significantly reduced, and the discharge temperature can be further reduced by 10°C - 20°C. However, the compressor power increases and the system energy efficiency decreases under this working condition.

[0020] Therefore, on the above premise, the present invention can achieve precise control of the discharge temperature and dynamic balance of the heating performance through the staged liquid injection design of the first liquid injection pipeline and the second liquid injection pipeline. For example, in low-temperature working conditions, the first control valve is preferentially opened for mild liquid injection. If the discharge temperature is still higher than the set threshold, the second control valve is gradually opened for deep liquid injection. In this way, by adjusting the opening degrees of the two control valves, a stepped control of the discharge temperature is achieved, which not only ensures the stability of the system under extremely low temperatures but also avoids a significant decrease in the heating energy efficiency due to excessive liquid injection.

[0021] In addition, under different working conditions, the first control valve and the second control valve can be selectively opened to achieve refined control of the discharge temperature and obtain the best temperature management effect. For example, under light load or medium load, since the liquid injection point of the first liquid injection pipeline is behind the regenerator, the refrigerant has undergone throttling and subcooling, and the pressure difference of the refrigerant relative to the compressor suction port is small, so the liquid injection volume is small. Only by opening the first control valve can the discharge temperature be reduced by 5°C - 10°C, and the impact on the system heating capacity is small; while under high load or extreme conditions, since the liquid injection point of the second liquid injection pipeline is behind the condenser, the pressure difference of the refrigerant relative to the compressor suction port is large, so the liquid injection volume is large. By opening the second control valve, the discharge temperature can be further reduced by 10°C - 20°C, thereby further reducing the discharge temperature while ensuring the heating capacity and ensuring the stable operation of the system in various environments.

[0022] In one embodiment, controlling the working states of at least one of the first control valve and the second control valve based on the determination result includes: When the determination result is t1 < T < t2, controlling the first control valve and the second control valve to maintain their current opening degrees unchanged; When the determination result is T ≤ t1 or T ≥ t2, adjusting the opening degree of at least one of the first control valve and the second control valve.

[0023] In one embodiment, when the determination result is T ≤ t1 or T ≥ t2, adjusting the opening degree of at least one of the first control valve and the second control valve includes: When the determination result is T ≤ t1, reduce the opening degree of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or both the first control valve and the second control valve are closed; When the scheduling result is T ≥ t2, increase the opening degree of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is zero within the preset error range, or both the first control valve and the second control valve are opened to the maximum opening degree; Wherein, the preset exhaust gas temperature Te satisfies: t1 < Te < t2.

[0024] In one embodiment, when the determination result is T ≤ t1, reducing the opening degree of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or both the first control valve and the second control valve are closed, includes: When the determination result is T ≤ t1, obtain the opening degree of the second control valve; In response to the opening degree of the second control valve being greater than zero, reduce the opening degree of the second control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the second control valve is closed; In response to the opening degree of the second control valve being zero, reduce the opening degree of the first control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the first control valve is closed.

[0025] In one embodiment, when the scheduling result is T ≥ t2, increasing the opening degree of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or both the first control valve and the second control valve are opened to the maximum opening degree, includes: When the determination result is T ≥ t2, obtain the opening degree of the first control valve; In response to the opening degree of the first control valve being less than the maximum opening degree, increase the opening degree of the first control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the first control valve reaches the maximum opening degree; In response to the opening degree of the first control valve reaching the maximum opening degree, increase the opening degree of the second control valve until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or the second control valve reaches the maximum opening degree.

[0026] In one embodiment, when the determination result is T≤t1 or T≥t2, adjusting the opening degree of at least one of the first control valve and the second control valve includes: When the determination result is T≤t1 or T≥t2, obtain the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te; Based on the temperature difference Ts, determine the target adjustment opening degree; Based on the target adjustment opening degree, adjust the opening degree of at least one of the first control valve and the second control valve; Wherein, the target adjustment opening degree is proportional to the temperature difference Ts.

[0027] In one embodiment, the heat pump system further includes a liquid accumulator and a defrosting branch. The liquid accumulator is provided with a refrigerant inlet and a first refrigerant outlet. The refrigerant inlet is connected to the outlet of the condenser, and the first refrigerant outlet is connected to the first heat exchange channel; the liquid accumulator is further provided with a second refrigerant outlet. Both ends of the defrosting branch are communicated with the second refrigerant outlet and the inlet of the evaporator respectively, and an electronically controlled valve with an adjustable opening degree is provided on the defrosting branch; The control method further includes: In response to the defrosting mode, control the electronically controlled valve to open, and control the second throttling device, the first control valve and the second control valve to close. Description of the Drawings

[0028] In order to more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the following will briefly introduce the drawings required for use in the description of the specific embodiments or the prior art. Obviously, the following drawings are some embodiments of the present invention. For those of ordinary skill in the art, other drawings can be obtained based on these drawings without creative efforts.

[0029] Figure 1 It is a schematic structural diagram of the zeotropic mixture refrigerant heat pump system in the heating mode according to the embodiment of the present invention; Figure 2 It is a schematic structural diagram of the zeotropic mixture refrigerant heat pump system in the defrosting mode according to the embodiment of the present invention; Figure 3 It is one of the schematic flowcharts of the control method of the zeotropic mixture refrigerant heat pump system according to the embodiment of the present invention; Figure 4It is the second schematic flow chart of the control method for the non-azeotropic mixture refrigerant heat pump system according to the embodiment of the present invention; Figure 5 It is the third schematic flow chart of the control method for the non-azeotropic mixture refrigerant heat pump system according to the embodiment of the present invention; Figure 6 It is the fourth schematic flow chart of the control method for the non-azeotropic mixture refrigerant heat pump system according to the embodiment of the present invention; Figure 7 It is the fifth schematic flow chart of the control method for the non-azeotropic mixture refrigerant heat pump system according to the embodiment of the present invention; Figure 8 It is the schematic structural diagram of the control device for the non-azeotropic mixture refrigerant heat pump system according to the embodiment of the present invention; Figure 9 It is the schematic structural diagram of the non-azeotropic mixture refrigerant heat pump system according to another embodiment of the present invention.

[0030] Explanation of reference numerals: 1. Compressor; 2. Condenser; 3. First throttling device; 4. Liquid receiver; 41. Refrigerant inlet; 42. First refrigerant outlet; 43. Second refrigerant outlet; 5. Electric control valve; 6. Regenerator; 61. First heat exchange channel; 62. Second heat exchange channel; 7. First control valve; 8. Second throttling device; 9. Evaporator; 10. Second control valve; 11. Gas-liquid separator; 12. First liquid injection pipeline; 13. Second liquid injection pipeline; 14. Defrosting branch; 15. Evaporator fan; 100. Acquisition module; 200. Judgment module; 300. Control module. Detailed implementation manners

[0031] Next, the technical solutions in the embodiments of the present application will be clearly and completely described in conjunction with the accompanying drawings in the embodiments of the present application. Obviously, the described embodiments are only a part of the embodiments of the present application, rather than all the embodiments. Based on the embodiments in the present application, all other embodiments obtained by those of ordinary skill in the art without creative efforts shall fall within the protection scope of the present application.

[0032] In the description of the present application, it should be understood that the orientation or positional relationship indicated by the terms "center", "upper", "lower", "front", "rear", "left", "right", "vertical", "horizontal", "top", "bottom", "inner", "outer", etc. is based on the orientation or positional relationship shown in the accompanying drawings, and is only for the convenience of describing the present application and simplifying the description, rather than indicating or implying that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore should not be construed as a limitation to the present application.

[0033] The terms "first" and "second" are used for descriptive purposes only and should not be construed as indicating or implying relative importance or implicitly specifying the quantity of the indicated technical features. Thus, features defined with "first" and "second" may explicitly or implicitly include one or more of such features. In the description of this application, unless otherwise specified, the meaning of "a plurality" is two or more.

[0034] In the description of this application, it should be noted that, unless otherwise clearly specified and defined, the terms "mounted", "connected", and "coupled" should be construed in a broad sense. For example, it may be a fixed connection, a detachable connection, or an integral connection; it may be a mechanical connection or an electrical connection; it may be directly connected or indirectly connected through an intermediate medium, and it may be the communication inside two components. For those of ordinary skill in the art, the specific meanings of the above terms in this application can be understood according to specific circumstances.

[0035] The following will introduce in detail azeotropic-mixture refrigerant heat pump system and its control method according to the present invention with reference to the accompanying drawings.

[0036] As Figure 1 and Figure 2 shown, the azeotropic-mixture refrigerant heat pump system according to the first aspect embodiment of the present invention includes a compressor 1, a refrigerant circulation loop, a first liquid injection pipeline 12, and a second liquid injection pipeline 13.

[0037] The refrigerant circulation loop is connected to the compressor 1 to form a loop. Along the refrigerant flow direction on the refrigerant circulation loop, a condenser 2, a first throttling device 3, a regenerator 6, a second throttling device 8, and an evaporator 9 are sequentially provided; the regenerator 6 has a first heat exchange channel 61 and a second heat exchange channel 62 for heat exchange with each other. The first heat exchange channel 61 is connected between the first throttling device 3 and the second throttling device 8, and the second heat exchange channel 62 is connected between the outlet of the evaporator 9 and the suction port of the compressor 1.

[0038] The first liquid injection pipeline 12 is connected between the outlet of the first heat exchange channel 61 and the outlet of the evaporator 9, and a first control valve 7 is provided on the first liquid injection pipeline 12; the second liquid injection pipeline 13 is connected between the outlet of the condenser 2 and the outlet of the evaporator 9, and a second control valve 10 is provided on the second liquid injection pipeline 13.

[0039] The specific structure of the azeotropic-mixture refrigerant heat pump system according to the embodiment of the present invention is introduced as follows: The heat pump system mainly includes a compressor 1, a condenser 2, a first throttling device 3, a regenerator 6, a second throttling device 8, an evaporator 9, a first liquid injection pipeline 12 and a second liquid injection pipeline 13, as well as corresponding control valves. The compressor 1 compresses the low-temperature and low-pressure gaseous mixed refrigerant into a high-temperature and high-pressure gas and then sends it into the condenser 2, where heat is released and it becomes a gas-liquid mixture. Subsequently, the mixed refrigerant passes through the first throttling device 3 for throttling and pressure reduction, and enters the first heat exchange channel 61 of the regenerator 6, where it exchanges heat with the low-temperature and low-pressure gaseous refrigerant returning from the evaporator 9 to further reduce the temperature. After that, the refrigerant is further depressurized via the second throttling device 8 and enters the evaporator 9 to absorb ambient heat, and finally returns to the compressor 1 to complete a full cycle.

[0040] The first liquid injection pipeline 12 is connected between the outlet of the first heat exchange channel 61 of the regenerator 6 and the outlet of the evaporator 9, and the flow rate is adjusted by the first control valve 7. The first liquid injection pipeline 12 is used to extract the liquid refrigerant from the outlet of the regenerator 6 and directly inject it into the gaseous refrigerant at the outlet of the evaporator 9, reducing the temperature of the mixed refrigerant, thereby effectively reducing the suction temperature of the compressor 1. It can be understood that when the suction temperature of the compressor 1 decreases, correspondingly, the discharge temperature of the compressor 1 will also decrease, thereby achieving the technical effect of suppressing the rise of the discharge temperature of the compressor 1. It should be noted that since the liquid injection point is behind the regenerator 6, the refrigerant has undergone throttling and subcooling, and the pressure difference of the refrigerant relative to the suction port of the compressor 1 is small, and it is preliminarily cooled, so the liquid injection amount is small, and the impact on the heating capacity of the system is small, and the discharge temperature can be reduced by 5-10 °C.

[0041] The second liquid injection pipeline 13 is connected between the outlet of the condenser 2 and the outlet of the evaporator 9, and the flow rate is adjusted by the second control valve 10. The second liquid injection pipeline 13 is used to extract the medium-temperature and high-pressure liquid refrigerant from the outlet of the condenser 2 and directly inject it into the gaseous refrigerant at the outlet of the evaporator 9, reducing the temperature of the mixed refrigerant, thereby effectively reducing the suction temperature of the compressor 1. It can be understood that when the suction temperature of the compressor 1 decreases, correspondingly, the discharge temperature of the compressor 1 will also decrease, thereby achieving the technical effect of suppressing the rise of the discharge temperature of the compressor 1. It should be noted that since the liquid injection point of the second liquid injection pipeline 13 is behind the condenser 2, the pressure difference of the refrigerant relative to the suction port of the compressor 1 is large, so the liquid injection amount is large. By enabling the second liquid injection pipeline 13, the suction temperature of the compressor 1 can be significantly reduced, and the discharge temperature can be further reduced by 10 °C - 20 °C, but in this working condition, the power of the compressor 1 increases and the energy efficiency of the system decreases. Based on the above specific structure, the specific working principle of the heat pump system of the present invention is as follows: The non-azeotropic mixture refrigerant heat pump system of the present invention solves the problem that the discharge temperature of the compressor 1 is too high under ultra-low temperature heating conditions, resulting in a decrease in system stability, by setting the first liquid injection pipeline 12 and the second liquid injection pipeline 13, and at the same time balances the contradiction between heating performance and discharge temperature control through a staged liquid injection strategy.

[0042] Specifically, in view of the high-temperature exhaust gas problem caused by component segregation and increased pressure ratio when the zeotropic mixture refrigerant operates at ultra-low temperatures, the present invention introduces two key liquid injection pipelines, namely the first liquid injection pipeline 12 and the second liquid injection pipeline 13. Among them, the first liquid injection pipeline 12 is connected between the first heat exchange channel 61 of the regenerator 6 and the outlet of the evaporator 9, and is provided with a first control valve 7. The above structure enables the liquid refrigerant to mix with the gaseous refrigerant flowing out of the outlet of the evaporator 9, reducing the temperature of the refrigerant, thereby effectively reducing the suction temperature of the compressor 1, and further achieving the technical effect of suppressing the rise of the exhaust temperature of the compressor 1, while having a relatively small impact on the heating performance. The second liquid injection pipeline 13 is led out from the outlet of the condenser 2 and connected to the outlet of the evaporator 9, and is equipped with a second control valve 10. When the ambient temperature is extremely low, by opening the second control valve 10, the high-pressure liquid refrigerant is directly sprayed into the outlet of the evaporator 9. Since the liquid injection point of the second liquid injection pipeline 13 is behind the condenser 2, the pressure difference of the refrigerant relative to the suction port of the compressor 1 is large, so the liquid injection volume is large. By enabling the second liquid injection pipeline 13, the suction temperature of the compressor 1 can be significantly reduced, and the exhaust temperature can be further reduced by 10°C - 20°C. It can be understood that although the above operation will have a greater impact on the heating efficiency of the system, it can ensure the stability of the system under extremely low temperatures.

[0043] Under low-temperature working conditions, the first control valve 7 is preferentially opened for mild liquid injection. If the exhaust temperature is still higher than the set threshold, the second control valve 10 is gradually opened for deep liquid injection. In this way, by adjusting the opening degrees of the two control valves, a stepped control of the exhaust temperature is achieved, which not only ensures the stability of the system under extremely low temperatures but also avoids a significant decrease in heating performance due to excessive liquid injection.

[0044] Furthermore, under ultra-low temperature heating conditions, the heat pump system of the present invention realizes precise control of the exhaust temperature through the staged liquid injection operation of the first liquid injection pipeline 12 and the second liquid injection pipeline 13. Among them, the working process of the heat pump system of the present invention is specifically as follows: As Figure 1 shown, the mixed refrigerant flows out of the condenser 2, is throttled and depressurized by the first throttling device 3, and then enters the first heat exchange channel 61 of the regenerator 6. In the regenerator 6, the high-temperature refrigerant in the first heat exchange channel 61 exchanges heat with the low-temperature refrigerant in the second heat exchange channel 62, further reducing the temperature of the refrigerant entering the evaporator 9. After being further depressurized by the second throttling device 8, the refrigerant enters the evaporator 9 to absorb heat and evaporate.

[0045] When the monitored exhaust temperature of the compressor 1 is slightly higher than the set maximum exhaust temperature, the heat pump system preferentially opens the first liquid injection pipeline 12 for mild liquid injection. At this time, the first control valve 7 is partially opened, and a small amount of liquid refrigerant is injected into the gaseous refrigerant at the outlet of the evaporator 9 from the first heat exchange channel 61 of the regenerator 6. The liquid refrigerant is mixed with the gaseous refrigerant flowing out of the outlet of the evaporator 9, and the temperature of the refrigerant decreases, thereby reducing the suction temperature of the compressor 1, and further reducing the exhaust temperature by 5°C - 10°C. During the above process, the attenuation of the heating performance is small, so the above operation is suitable for medium and low temperature conditions.

[0046] When the ambient temperature further decreases, for example, when the ambient temperature is lower than -20°C, the exhaust temperature of the compressor 1 continues to rise, and the system automatically opens the second control valve 10. The high-temperature and high-pressure liquid refrigerant extracted from the outlet of the condenser 2 is injected into the gaseous refrigerant at the outlet of the evaporator 9 through the second liquid injection pipeline 13. At this time, the heat pump system performs a dual liquid injection action. On the one hand, the first control valve 7 is opened and maintains mild liquid injection to continue reducing the suction temperature of the compressor 1; on the other hand, the second control valve 10 is also opened and the liquid refrigerant is mixed with the gaseous refrigerant flowing out of the outlet of the evaporator 9 to further reduce the suction temperature of the compressor 1, and further reduce the exhaust temperature. At the same time, the superheat degree of the gaseous refrigerant at the outlet of the evaporator 9 is diluted, and the load fluctuation of the compressor 1 is reduced. During the above process, the attenuation of the heating performance is large, but the system stability is significantly improved, so the above operation is suitable for extreme low temperature conditions.

[0047] Therefore, through the staged liquid injection design of the first liquid injection pipeline 12 and the second liquid injection pipeline 13, the present invention realizes the precise control of the exhaust temperature and the dynamic balance of the heating performance. Under ultra-low temperature conditions, the system can flexibly switch the liquid injection strategy according to actual needs, avoiding the risk of failure of the compressor 1 due to excessive exhaust temperature, minimizing the impact on energy efficiency, and significantly improving the adaptability and reliability of the non-azeotropic mixture heat pump system.

[0048] In summary, for the non-azeotropic mixture heat pump system according to the embodiment of the present invention, by setting the first liquid injection pipeline 12 and the second liquid injection pipeline 13, liquid refrigerants in different states are respectively introduced and sprayed into the outlet of the evaporator 9, so that the liquid refrigerant is mixed with the gaseous refrigerant flowing out of the outlet of the evaporator 9, and the suction temperature of the compressor 1 is reduced by reducing the refrigerant temperature, thereby effectively reducing the exhaust temperature and avoiding problems such as carbonization of lubricating oil and aging of sealing materials caused by high temperature. At the same time, the above structure enables the system to still operate stably in an extremely low temperature environment, solving the problems of performance degradation and system instability of traditional heat pump systems due to too low evaporation temperature and too large compression ratio under low temperature conditions. In addition, by optimizing the refrigerant circulation path and introducing a shunt branch at key nodes, the present invention does not change the basic structure of the original heat pump system, is convenient for integration and engineering application, and has good practicability.

[0049] As shown in Figure 1 and Figure 2 FIG. 4, according to some embodiments of the present invention, the heat pump system further includes a liquid receiver 4.

[0050] The liquid receiver 4 is provided with a refrigerant inlet 41 and a first refrigerant outlet 42. The refrigerant inlet 41 is connected to the outlet of the condenser 2, and the first refrigerant outlet 42 is connected to the first heat exchange channel 61.

[0051] In this embodiment, the liquid receiver 4 has at least two interfaces. Among them, the refrigerant inlet 41 is used to receive the mixed refrigerant discharged from the condenser 2; the first refrigerant outlet 42 is used to transport the stored or regulated liquid refrigerant to the first heat exchange channel 61 of the regenerator 6. The liquid receiver 4 is located between the condenser 2 and the regenerator 6 in the system. After the mixed refrigerant flows out of the condenser 2, it first enters the liquid receiver 4 for storage or buffering. The liquid refrigerant regulated by the liquid receiver 4 enters the first heat exchange channel 61 of the regenerator 6 and exchanges heat with the gaseous refrigerant in the second heat exchange channel 62.

[0052] It can be understood that the liquid receiver 4 functions as a refrigerant storage, buffer, and flow stabilizer, which can effectively prevent the impact on the subsequent throttling device and evaporator 9 caused by insufficient or fluctuating liquid refrigerant. The liquid receiver 4 is particularly suitable for variable load operation or low-temperature startup, thereby improving the stability and reliability of the system.

[0053] In the related art, during the operation of the system, especially in the variable-frequency compressor or multi-mode operation state, the refrigerant flow rate at the outlet of the condenser may fluctuate. In the present invention, by providing a liquid receiver 4 in the heat pump system, the internal space is used to temporarily store the liquid refrigerant, acting as a buffer pool to avoid misoperation of the throttling device or unstable liquid supply to the evaporator 9 caused by instantaneous flow rate changes.

[0054] In summary, in this embodiment, on the one hand, the refrigerant enters the first heat exchange channel 61 of the regenerator 6 after being buffered by the liquid receiver 4. Since the flow rate is more stable and the temperature distribution is more uniform, it helps to enhance the heat exchange effect with the gaseous refrigerant in the second heat exchange channel 62, thereby improving the energy efficiency ratio of the entire system.

[0055] On the other hand, in the ultra-low temperature heating condition, the refrigerant circulation amount may increase. The liquid receiver 4 can be used as a refrigerant storage container to prevent problems such as insufficient refrigerant supply and ensure the safe operation of the system. In addition, if the system is also equipped with a first liquid injection pipeline 12 and a second liquid injection pipeline 13, the presence of the liquid receiver 4 can further ensure the sufficiency and controllability of the liquid injection refrigerant, making the exhaust temperature regulation more accurate and reliable.

[0056] As shown in Figure 1 and Figure 2As shown, further, the heat pump system further includes a defrosting branch 14. The liquid receiver 4 is further provided with a second refrigerant outlet 43. Both ends of the defrosting branch 14 are respectively communicated with the second refrigerant outlet 43 and the inlet of the evaporator 9. An electronically controlled valve 5 with adjustable opening degree is provided on the defrosting branch 14.

[0057] In this embodiment, the second refrigerant outlet 43 of the liquid receiver 4 serves as a dedicated outlet for defrosting refrigerant and is enabled when the system enters the defrosting stage, providing high-temperature gaseous refrigerant to the evaporator 9 to accelerate the melting of the frost layer. The defrosting branch 14 provides a refrigerant passage bypassing the conventional throttling device, allowing the high-temperature refrigerant from the liquid receiver 4 to directly flow to the evaporator 9 for quickly raising the temperature of the evaporator 9 and achieving efficient defrosting. The electronically controlled valve 5 is used to control the opening and closing of the defrosting branch 14 and can adjust its opening degree according to actual needs, thereby precisely controlling the amount of refrigerant entering the evaporator 9 and avoiding excessive impact or energy waste.

[0058] When the system detects that the surface of the evaporator 9 is severely frosted, for example, judged by a differential pressure sensor, a temperature sensor or a timing logic, the defrosting program is started, and the specific process is as follows: As Figure 2 shown, when the system enters the defrosting stage, the first throttling device 3 on the main refrigerant path is closed or adjusted to a high opening state. The electronically controlled valve 5 is opened, allowing the high-temperature gaseous refrigerant from the liquid receiver 4 to enter the evaporator 9 through the defrosting branch 14. The high-temperature refrigerant quickly releases heat in the evaporator 9, significantly raising the surface temperature of the evaporator 9 and prompting the frost layer to quickly melt. After the frost layer is removed, the electronically controlled valve 5 is closed, and then the main refrigerant circulation path is restored, and the system re-enters the normal heating mode.

[0059] In this way, by introducing the defrosting branch 14 and the electronically controlled valve 5 and cooperating with the second refrigerant outlet 43 of the liquid receiver 4, the present invention can directly heat the evaporator 9 with the high-temperature gaseous refrigerant from the liquid receiver 4, significantly accelerating the frost layer melting speed and shortening the defrosting time. Compared with the traditional reverse defrosting method, this solution does not need to change the running direction of the compressor 1, reducing the system energy loss.

[0060] Further, the second refrigerant outlet 43 is arranged at the top of the liquid receiver 4, and the first refrigerant outlet 42 is arranged at the bottom of the liquid receiver 4.

[0061] It can be understood that after the condensed refrigerant enters the liquid receiver 4, due to the gravity effect and density difference, the liquid refrigerant naturally deposits at the lower part of the liquid receiver 4, while the gaseous or gas-liquid mixture floats on the upper part. The first refrigerant outlet 42 is arranged at the bottom to ensure that the output is pure liquid refrigerant, which is suitable for the main circulation path requiring stable liquid working medium, such as the regenerator 6, the throttling device, the evaporator 9, etc. The second refrigerant outlet 43 is arranged at the top to facilitate the extraction of the gaseous refrigerant in the high-temperature and high-pressure state for defrosting.

[0062] As Figure 1 andFigure 2 As shown, according to some embodiments of the present invention, the heat pump system further includes an evaporator fan 15 and a gas-liquid separator 11.

[0063] The evaporator fan 15 is used to drive air to flow through the evaporator 9; the gas-liquid separator 11 is arranged on the refrigerant circulation loop and is located between the second heat exchange channel 62 and the suction port of the compressor 1.

[0064] The control method of the zeotropic mixture refrigerant heat pump system according to the present invention will be introduced in detail below with reference to the accompanying drawings. It should be noted that this control method is applied to the zeotropic mixture refrigerant heat pump system of the first aspect of the present invention.

[0065] As Figure 3 shown, the control method of the zeotropic mixture refrigerant heat pump system according to the second aspect of the present invention includes: Step S100, in response to the heating mode, obtain the actual exhaust temperature T of the compressor 1; Step S200, compare the actual exhaust temperature T with the maximum exhaust temperature t2 and the minimum exhaust temperature t1 respectively to obtain a determination result; Step S300, based on the determination result, control the working state of at least one of the first control valve 7 and the second control valve 10.

[0066] The control method of the zeotropic mixture refrigerant heat pump system according to the embodiments of the present invention is specifically introduced as follows: In step S100, after the heating mode is started, the control system first monitors the exhaust temperature T of the compressor 1 in real time. This step is the basis of the entire control strategy, and all subsequent decisions depend on the current actual exhaust temperature value. Among them, the exhaust temperature T reflects the working state inside the compressor 1 and the load it faces. Excessive or too low exhaust temperature will affect the performance and life of the system.

[0067] In step S200, the control system sets two key thresholds, namely the maximum exhaust temperature t2 and the minimum exhaust temperature t1. These two parameters define the allowable exhaust temperature range, and exceeding this range may cause system instability or component damage.

[0068] Specifically, the system continuously compares the actual exhaust temperature T with these two thresholds. If T > t1, it means that the exhaust temperature is too high and measures need to be taken to reduce the temperature; if T < t2, it means that the exhaust temperature is too low and the liquid injection amount needs to be considered to be reduced to maintain the system efficiency; if t2 ≤ T ≤ t1, it means that the current exhaust temperature is within the ideal range and the liquid injection valve does not need to be adjusted.

[0069] In step S300, according to the determination result in the previous step, the system dynamically adjusts the states of the first control valve 7 and the second control valve 10 to regulate the refrigerant flow rate entering the outlet of the evaporator 9. For example, when T > t1, the first control valve 7 is preferentially opened: if the exhaust temperature is only slightly higher than t1, moderate cooling can be achieved by slightly spraying liquid. At this time, only the first control valve 7 needs to be opened, and the liquid refrigerant in the first heat exchange channel 61 of the regenerator 6 is used for preliminary cooling. If T continues to rise above a certain set value, the second control valve 10 needs to be further opened to introduce more liquid refrigerant directly into the outlet of the evaporator 9 to achieve a greater cooling effect.

[0070] Furthermore, the specific control process of the heat pump system of the present invention is as follows: after the system enters the heating mode, it starts to collect the exhaust temperature T data of the compressor 1 in real time and transmits it to the control system. The system continuously compares the actual exhaust temperature T with the preset minimum exhaust temperature. During this process, the system can quickly identify whether the current exhaust temperature deviates from the normal range and decide the next operation accordingly. According to the above determination result, the system automatically adjusts the states of the first control valve 7 and the second control valve 10.

[0071] For example, when the exhaust temperature slightly exceeds the standard, the system only needs to finely adjust the opening degree of the first control valve 7 and inject an appropriate amount of liquid refrigerant into the outlet of the evaporator 9 to restore the normal exhaust temperature. Another example is that when facing a significant over-temperature phenomenon caused by extremely low temperature or high-load working conditions, the system will simultaneously open both control valves to ensure that enough refrigerant is sprayed into the outlet of the evaporator 9 to quickly suppress the rising trend of the exhaust temperature.

[0072] Once it is detected that the exhaust temperature drops back to the safe range, the system will gradually close or reduce the opening degree of the liquid spraying valve, restore the conventional refrigerant circulation process, and ensure the efficient operation of the system. It should be noted that the system will continuously monitor throughout the operation cycle and make timely adjustments to the control strategy according to the latest exhaust temperature feedback information, forming a closed-loop control mechanism, so as to ensure that the system can always be in the best working state regardless of how the external conditions change.

[0073] In summary, through the above detailed control method, the zeotropic mixture refrigerant heat pump system of the present invention can flexibly respond to various complex working conditions, ensure that the exhaust temperature of the compressor 1 always remains within a reasonable range, improve the reliability of the system, and enhance the overall energy efficiency performance.

[0074] Specifically, the control method of the present invention can accurately adjust the exhaust temperature by real-time monitoring of the actual exhaust temperature of the compressor 1 and comparing it with the preset minimum exhaust temperature, which helps to avoid problems such as lubricating oil carbonization and sealing material aging caused by too high temperature, and also prevents too low temperature from affecting the system efficiency.

[0075] Meanwhile, since the control method dynamically adjusts the operating states of the first control valve 7 and the second control valve 10 according to the actual exhaust temperature, it can ensure that the heat pump system maintains the optimal operating state under different environmental conditions. The above hierarchical regulation strategy can flexibly switch the liquid injection amount according to actual needs, which can not only provide sufficient cooling effect when necessary, but also reduce energy consumption when not needed.

[0076] As Figure 4 shown, according to some embodiments of the present invention, the step S300 of controlling the operating state of at least one of the first control valve 7 and the second control valve 10 based on the determination result includes: Step S310, when the determination result is t1 < T < t2, control the first control valve 7 and the second control valve 10 to keep the current opening unchanged; Step S320, when the determination result is T ≤ t1 or T ≥ t2, adjust the opening of at least one of the first control valve 7 and the second control valve 10.

[0077] In this embodiment, when the determination result is t1 < T < t2, in this case, the actual exhaust temperature T is within the preset ideal range, that is, between the maximum exhaust temperature t2 and the minimum exhaust temperature t1. At this time, the control system will keep the current openings of the first control valve 7 and the second control valve 10 unchanged, and the purpose is to maintain the existing state, which helps to ensure the stable operation of the system and avoid energy waste or system fluctuations caused by unnecessary adjustment actions.

[0078] When the determination result is T ≤ t1 or T ≥ t2, that is, the actual exhaust temperature T is lower than the set minimum exhaust temperature t1 or higher than the maximum exhaust temperature t2, it indicates that the current operating state of the system exceeds the ideal range and needs to be adjusted.

[0079] Among them, for the case of T ≤ t1, the control system may choose to reduce or close the liquid injection amount to increase the exhaust temperature. For example, reduce the opening of the first control valve 7 or even completely close the first control valve 7, and handle the second control valve 10 in the same way to prevent too much refrigerant from entering the outlet of the evaporator 9 and causing the temperature to drop further.

[0080] For the case of T ≥ t2, the system will take measures to reduce the exhaust temperature. For example, the system first moderately opens the first control valve 7 to introduce an appropriate amount of liquid refrigerant for preliminary cooling. If the exhaust temperature continues to rise above a certain threshold, the second control valve 10 needs to be further opened to increase the liquid injection amount to quickly cool the return air of the compressor 1, thereby effectively suppressing the rise of the exhaust temperature.

[0081] It should be noted that the operations of adjusting the opening degrees of the above-mentioned first control valve 7 and second control valve 10 can be progressive, that is, gradually increasing or decreasing the opening degrees, or can be a one-time large adjustment. The present invention does not make special restrictions here.

[0082] In this way, through the precise control of the first control valve 7 and the second control valve 10, the present invention enables the heat pump system to maintain stable operation under normal working conditions, avoiding unnecessary adjustments; while under extreme conditions, it can respond in a timely manner to ensure the safe and efficient operation of the system. In addition, by avoiding the compressor 1 from operating at too high or too low exhaust temperatures for a long time, the present invention reduces the risk of damage to the equipment and helps to extend the service life of the entire system.

[0083] As Figure 4 shown, according to some embodiments of the present invention, when the determination result is T ≤ t1 or T ≥ t2, the step S320 of adjusting the opening degree of at least one of the first control valve 7 and the second control valve 10 includes: Step S321, when the determination result is T ≤ t1, reducing the opening degree of at least one of the first control valve 7 and the second control valve 10 until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range, or both the first control valve 7 and the second control valve 10 are closed; Step S322, when the scheduling result is T ≥ t2, increasing the opening degree of at least one of the first control valve 7 and the second control valve 10 until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range, or both the first control valve 7 and the second control valve 10 are opened to the maximum opening degree.

[0084] Among them, the preset exhaust temperature Te satisfies: t1 < Te < t2.

[0085] In this embodiment, when the determination result is that the actual exhaust temperature T is not within the preset ideal range, the system will adjust the exhaust temperature by adjusting the opening degrees of the first control valve 7 and the second control valve 10 until it returns to the preset exhaust temperature Te within the ideal range.

[0086] Specifically, when the determination result is T ≤ t1, the actual exhaust temperature of the system is lower than the set minimum exhaust temperature t1, that is, the exhaust temperature is too low. Since too much refrigerant injection will cause further cooling, in order to increase the exhaust temperature, the system needs to reduce the liquid injection volume.

[0087] The control system begins to gradually reduce the opening degrees of the first control valve 7 and / or the second control valve 10, and after each opening degree adjustment, it obtains the actual exhaust temperature of the compressor 1 and the opening degree of the current control valve again until either of the following conditions is met: the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range, indicating that the exhaust temperature has returned to the ideal state; or, both the first control valve 7 and the second control valve 10 are fully closed, and no more refrigerant is sprayed into the outlet of the evaporator 9 to prevent further cooling.

[0088] When the determination result is T ≥ t2, the actual exhaust temperature is higher than the set maximum exhaust temperature t2, that is, the exhaust temperature is too high, which may cause damage to the compressor 1. At this time, it is necessary to increase the liquid injection volume to reduce the exhaust temperature.

[0089] The control system begins to gradually increase the opening degrees of the first control valve 7 and / or the second control valve 10 and after each opening degree adjustment, it obtains the actual exhaust temperature of the compressor 1 and the opening degree of the current control valve again until either of the following conditions is met: the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range, indicating that the exhaust temperature has dropped to the ideal level; or, both the first control valve 7 and the second control valve 10 are opened to the maximum opening degree to ensure maximum utilization of the liquid injection cooling effect to quickly reduce the exhaust temperature.

[0090] It should be noted that the preset exhaust temperature Te is an ideal value between the upper and lower minimum exhaust temperatures. It represents the best operating point that neither causes the equipment to overheat nor affects the efficiency due to too low temperature. For example, the preset exhaust temperature Te = (t1 + t2) / 2. In this way, by adjusting the actual exhaust temperature T to be close to Te, the system can avoid potential risks such as lubricating oil carbonization or sealing material aging while ensuring performance. And the above preset error range is an artificially set allowable error range, the purpose of which is to improve the fault tolerance rate of the adjustment and avoid errors affecting the normal control logic. For example, the preset error range can be from -1°C to 1°C.

[0091] In summary, this method not only improves the response speed and accuracy of the system, but also enhances the adaptability and reliability of the system under different working conditions, ensuring long-term stable operation. At the same time, due to the adoption of the hierarchical adjustment strategy, the liquid injection volume can be flexibly adjusted according to actual needs, so as to achieve the best energy utilization efficiency.

[0092] Such as Figure 5As shown, in some specific embodiments, when the determination result is T ≤ t1, the step S321 of reducing the opening degree of at least one of the first control valve 7 and the second control valve 10 until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or both the first control valve 7 and the second control valve 10 are closed, includes: Step S3211: When the determination result is T ≤ t1, obtain the opening and closing state of the second control valve 10; Step S3212: In response to the second control valve 10 being in the open state, reduce the opening degree of the second control valve 10 until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the second control valve 10 is closed; Step S3213: In response to the second control valve 10 being in the closed state, reduce the opening degree of the first control valve 7 until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the first control valve 7 is closed.

[0093] In this embodiment, when it is detected that the actual exhaust gas temperature T ≤ t1, it indicates that the current exhaust gas temperature is relatively low, which may cause problems such as liquid carry - over in the suction of the compressor 1 and a decrease in heating efficiency. Therefore, the control system needs to gradually reduce the liquid injection amount to increase the return gas temperature of the compressor 1, thereby increasing the exhaust gas temperature. Since the two control valves control the liquid injection amounts of different paths respectively, the system adopts the strategy of first closing the second control valve 10 and then adjusting the first control valve 7 to achieve more precise temperature regulation.

[0094] Specifically, the system first determines whether the current second control valve 10 is in the open state. If it is in the open state, it indicates that the currently injected high - pressure liquid refrigerant is from the outlet of the condenser 2 and its cooling capacity is relatively strong; if it is in the closed state, it means that the liquid injection is currently only adjusted by the first control valve 7.

[0095] In response to the second control valve 10 being in the open state, the control system starts to gradually reduce the opening degree of the second control valve 10. Each time it is reduced by a certain proportion, such as 5% - 10% each time, and continuously monitors the change of the actual exhaust gas temperature; during this process, the system continuously calculates the difference between the actual exhaust gas temperature T and the target temperature Te.

[0096] The above adjustment will continue until any of the following conditions is met: the actual exhaust gas temperature T rises to equal the preset exhaust gas temperature Te, or the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, at which time the adjustment stops; the second control valve 10 is completely closed, at which time continuing the adjustment will be ineffective and the next step of operation needs to be entered.

[0097] In response to the second control valve 10 being in the closed state, after the second control valve 10 has been closed, if the actual exhaust temperature is still lower than the target value Te, the system then adjusts the first control valve 7. The control system starts to gradually reduce the opening degree of the first control valve 7, also using a step-by-step adjustment method, and monitors the change of the actual exhaust temperature in real time.

[0098] The above adjustment process will continue until any of the following conditions is met: the actual exhaust temperature T rises back to equal the preset exhaust temperature Te, or the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range; the first control valve 7 is also fully closed, at this time the liquid injection path is completely closed, and the system enters a no-liquid-injection operation state.

[0099] It should be explained that since the refrigerant flowing through the second control valve 10 is directly sprayed from the condenser 2 as high-pressure and high-temperature liquid refrigerant, its presence will significantly reduce the working temperature of the evaporator 9, thereby affecting the heating efficiency of the entire system. It can be understood that in the case of too low exhaust temperature, the above steps can effectively reduce the amount of additional refrigerant entering the outlet of the evaporator 9 by first reducing or closing the second control valve 10, so as to minimize the interference with the normal operation of the system and maintain the stability of the main cycle refrigerant amount.

[0100] In addition, the reason for preferentially adjusting the second control valve 10 is also that it has a greater impact on the exhaust temperature. If the exhaust temperature can be raised back to the target value Te only by reducing the opening degree of the second control valve 10, there is no need to further adjust the first control valve 7, thus avoiding system fluctuations caused by over-adjustment. If the exhaust temperature still does not reach the ideal state after the second control valve 10 is fully closed, it is necessary to further finely adjust the first control valve 7 to achieve more precise temperature control.

[0101] In summary, the present invention preferentially closes the second control valve 10, reduces the injection of unnecessary high-pressure and high-temperature liquid refrigerant, and avoids causing too much impact on the main cycle refrigerant amount, thereby ensuring the heating efficiency of the system. At the same time, the above steps are adjusted step by step according to the change of the actual exhaust temperature, which can not only quickly respond to abnormal situations, but also prevent new problems caused by over-adjustment, ensuring the stable operation of the system.

[0102] As Figure 6 shown, in some other specific embodiments, when the scheduling result is T≥t2, increasing the opening degree of at least one of the first control valve 7 and the second control valve 10 until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range, or the step S322 in which both the first control valve 7 and the second control valve 10 are opened to the maximum opening degree, includes: Step S3221, when the determination result is T≥t2, obtain the opening degree of the first control valve 7; Step S3222: In response to the opening degree of the first control valve 7 being less than the maximum opening degree, increase the opening degree of the first control valve 7 until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the first control valve 7 reaches the maximum opening degree; Step S3223: In response to the opening degree of the first control valve 7 reaching the maximum opening degree, increase the opening degree of the second control valve 10 until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, or the second control valve 10 reaches the maximum opening degree.

[0103] In this embodiment, when it is detected that the actual exhaust gas temperature T ≥ t2, it indicates that the current exhaust gas temperature is too high, which may lead to risks such as lubricating oil carbonization and sealing material aging, affecting the life of the compressor 1 and the system stability. Therefore, the control system needs to gradually increase the liquid injection amount to reduce the suction gas temperature of the compressor 1, thereby suppressing the rise of the exhaust gas temperature. Since the two control valves control the liquid injection flow rate and cooling capacity of different paths respectively, the system adopts the strategy of first adjusting the first control valve 7 and then adjusting the second control valve 10 to achieve more precise and energy-saving temperature regulation.

[0104] Specifically, the system first determines whether the current first control valve 7 has reached its maximum allowable opening degree. If it has not reached the maximum opening degree, it means there is still room to further increase the liquid injection amount; if it has reached the maximum opening degree, it is necessary to consider enabling or continuing to open the second control valve 10 to enhance the cooling effect.

[0105] In response to the opening degree of the first control valve 7 being less than the maximum opening degree, the control system starts to gradually increase the opening degree of the first control valve 7. Each time, a certain proportion is increased, and the change of the actual exhaust gas temperature is continuously monitored; during this process, the system continuously calculates the difference between the actual exhaust gas temperature T and the target temperature Te.

[0106] The above adjustment process will continue until either of the following conditions is met: the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within the preset error range, and the adjustment is stopped at this time; the first control valve 7 has been fully opened to the maximum allowable opening degree, and further adjustment will be ineffective at this time, and the next step of operation needs to be entered.

[0107] When the first control valve 7 is already at the maximum opening degree, if the actual exhaust gas temperature is still higher than the target value Te, the system then adjusts the second control valve 10. The control system starts to gradually increase the opening degree of the second control valve 10, also using a step-by-step adjustment method, and monitors the change of the actual exhaust gas temperature in real time.

[0108] The above adjustment process will continue until any of the following conditions is met: the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within the preset error range; the second control valve 10 is also fully opened to the maximum opening degree, at which time the liquid injection amount has reached the limit and the system enters the maximum cooling state.

[0109] It should be noted that during the process of increasing the opening degree of the first control valve 7 or the second control valve 10, the medium-temperature and medium-pressure refrigerant is injected by the first control valve 7. The temperature of the liquid refrigerant from the first heat exchange channel 61 of the regenerator 6 is relatively low, but the pressure is moderate. After being injected into the outlet of the evaporator 9, it has a moderate cooling effect on the exhaust temperature. In contrast, the impact on the main cycle refrigerant is relatively small, and it will not significantly reduce the heat absorption of the evaporator 9, thus maintaining a relatively high energy efficiency ratio.

[0110] The second control valve 10 injects high-pressure and medium-temperature refrigerant. The liquid refrigerant from the outlet of the condenser 2 has a greater pressure and stronger cooling ability after being injected, but its liquid injection amount is larger and it has a greater impact on the heating energy efficiency of the heat pump system. Therefore, the second control valve 10 is only enabled as a supplementary means when the first control valve 7 is fully opened but still cannot effectively control the exhaust temperature.

[0111] In addition, the above steps can also prevent a sudden decrease in the main cycle refrigerant amount caused by suddenly opening the second control valve 10, avoiding instability or a sharp drop in efficiency of the system.

[0112] In summary, by preferentially opening the first control valve 7, the above steps can reduce unnecessary injection of high-pressure and high-temperature liquid refrigerant, avoid excessive impact on the main cycle refrigerant amount, and thus ensure the heating efficiency of the system. At the same time, this method adjusts step by step according to the change of the actual exhaust temperature, which can not only quickly respond to abnormal situations but also prevent new problems caused by over-adjustment, thereby ensuring the stable operation of the system.

[0113] As Figure 7 shown, according to some embodiments of the present invention, when the determination result is T≤t1 or T≥t2, the step S320 of adjusting the opening degree of at least one of the first control valve 7 and the second control valve 10 includes: Step S323, when the determination result is T≤t1 or T≥t2, obtain the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te; Step S324, based on the temperature difference Ts, determine the target adjustment opening degree; Step S325, adjust the opening degree of at least one of the first control valve 7 and the second control valve 10 based on the target adjustment opening degree.

[0114] Among them, the target adjustment opening degree is proportional to the temperature difference Ts.

[0115] In this embodiment, when the system detects that T ≤ t1 or T ≥ t2, it enters the adjustment stage. The controller calculates the absolute difference Ts = |T - Te| between the current actual exhaust temperature T and the preset exhaust temperature Te. Among them, Ts reflects the degree to which the current exhaust temperature deviates from the ideal state, and it can be used to determine the amplitude of subsequent valve adjustment.

[0116] The system determines the target opening degrees of the first control valve 7 and / or the second control valve 10 that need to be adjusted according to the magnitude of Ts. Specifically, the change in the opening degree is directly proportional to Ts, that is, if Ts is small, only a small adjustment of the valve opening is required; if Ts is large, a large increase or decrease in the valve opening is required.

[0117] According to the target adjustment opening degree obtained in the previous step, the control system makes corresponding opening adjustments to the first control valve 7 and / or the second control valve 10, and the adjustment direction depends on the position of T relative to Te. Among them, if T < Te, the valve opening is reduced. The second control valve 10 is preferentially closed, and then the first control valve 7 is reduced. The purpose of the above operations is to increase the suction temperature, thereby increasing the exhaust temperature.

[0118] If T > Te, the valve opening is increased. The first control valve 7 is preferentially opened, and then the second control valve 10 is opened. The purpose of the above operations is to enhance the liquid injection cooling effect, thereby reducing the exhaust temperature.

[0119] The above adjustment process continues until the actual exhaust temperature T returns to Te, that is, Ts = 0, or Ts is within the preset error range, or the control valve has reached the maximum / minimum allowable opening degree.

[0120] In this way, by introducing Ts, the system can make corresponding intensity adjustments according to the deviation degree of the exhaust temperature, ensuring the pertinence and accuracy of the control method. Among them, the greater the temperature deviation, the stronger the adjustment force; the smaller the deviation, the softer the adjustment, thereby effectively preventing over-adjustment. In addition, the above steps can avoid unnecessary over-liquid injection or flow interruption, maintain the stability of the main circulation refrigerant volume, and ensure the heating performance.

[0121] According to some embodiments of the present invention, the heat pump system further includes a liquid receiver 4 and a defrosting branch 14. The liquid receiver 4 is provided with a refrigerant inlet 41 and a first refrigerant outlet 42. The refrigerant inlet 41 is connected to the outlet of the condenser 2, and the first refrigerant outlet 42 is connected to the first heat exchange channel 61. The liquid receiver 4 is further provided with a second refrigerant outlet 43. Both ends of the defrosting branch 14 are respectively communicated with the second refrigerant outlet 43 and the inlet of the evaporator 9. An electronically controlled valve 5 with an adjustable opening degree is provided on the defrosting branch 14.

[0122] The control method further includes: Step S400: In response to the defrost mode, control the electronic control valve 5 to open, and control the second throttling device 8, the first control valve 7, and the second control valve 10 to close.

[0123] In this embodiment, the above steps switch the refrigerant flow path during the defrosting stage, close unnecessary liquid injection and throttling channels, and enable the dedicated defrosting branch 14, thereby realizing a fast and efficient defrosting process.

[0124] The specific control process of the defrost mode is as follows: First, the system determines whether defrosting needs to be started. Among them, the system can determine to start defrosting based on the surface temperature of the evaporator 9 continuously being lower than the set value. Or, it can also determine to start defrosting based on the abnormal drop of the suction pressure, the system operation time reaching the preset defrost cycle, or detecting that the surface of the evaporator 9 is severely frosted through a sensor, etc. The present invention does not make special limitations here.

[0125] Once it is confirmed to enter the defrost mode, the control system immediately executes the next operation. The control system opens the electronic control valve 5 on the defrosting branch 14; the high-temperature refrigerant flows out from the second refrigerant outlet 43 of the liquid receiver 4 and directly enters the inlet of the evaporator 9 through the defrosting branch 14; the refrigerant releases a large amount of heat in the evaporator 9, prompting the frost layer on the surface of the evaporator 9 to melt quickly.

[0126] To ensure that as much refrigerant as possible flows to the evaporator 9 for defrosting, and at the same time avoid interference or energy waste, the system synchronously executes the following operations: close the second throttling device 8, thereby blocking the refrigerant flow in the main circulation path to the evaporator 9 and preventing the refrigerant from continuing to evaporate and absorb heat, which affects the defrosting effect; close the first control valve 7, thereby stopping the liquid injection at the outlet of the evaporator 9 and avoiding the reduction of the defrosting efficiency caused by the refrigerant diversion; close the second control valve 10, which can also prevent the refrigerant from being sprayed into the outlet of the evaporator 9 from the outlet of the condenser 2, ensuring that the defrosting branch 14 is the main path.

[0127] In summary, the above steps use high-temperature refrigerant to directly heat the evaporator 9, significantly accelerating the frost layer melting speed and shortening the defrosting time. And the above operations can achieve the defrosting operation without the four-way valve reversing, significantly reducing the energy consumption of the system.

[0128] Next, the control device of the zeotropic mixture refrigerant heat pump system given by the present invention will be introduced in detail with reference to the accompanying drawings. It should be noted that this control device is applied to the zeotropic mixture refrigerant heat pump system of the first aspect of the present invention.

[0129] As Figure 8 shown, the control device of the zeotropic mixture refrigerant heat pump system according to the third aspect of the present invention includes: An acquisition module 100, configured to obtain the actual exhaust temperature T of the compressor 1 in response to the heating mode; A judgment module 200, configured to compare an actual exhaust temperature T with a maximum exhaust temperature t2 and a minimum exhaust temperature t1 respectively to obtain a judgment result; A control module 300, configured to control an operating state of at least one of a first control valve 7 and a second control valve 10 based on the judgment result.

[0130] Figure 9 The figure shows a schematic structural diagram of an embodiment of a non-azeotropic mixture refrigerant heat pump system provided by an embodiment of the present invention. The specific implementation of the non-azeotropic mixture refrigerant heat pump system in the specific embodiments of the present invention is not limited.

[0131] As Figure 9 shown, the heat pump may include: a processor 502, a communication interface 504, a memory 506, and a communication bus 508.

[0132] Wherein: the processor 502, the communication interface 504, and the memory 506 communicate with each other through the communication bus 508. The communication interface 504 is configured to communicate with network elements of other devices such as clients or other servers. The processor 502 is configured to execute a program 510, and specifically may execute relevant steps in the above-mentioned embodiment of the control method for the non-azeotropic mixture refrigerant heat pump system.

[0133] Specifically, the program 510 may include program codes, and the program codes include computer-executable instructions.

[0134] The processor 502 may be a central processing unit CPU, or a specific integrated circuit ASIC (Application Specific Integrated Circuit), or one or more integrated circuits configured to implement the embodiments of the present invention. One or more processors included in the non-azeotropic mixture refrigerant heat pump system may be of the same type of processor, such as one or more CPUs; or may be of different types of processors, such as one or more CPUs and one or more ASICs.

[0135] The memory 506 is used to store the program 510. The memory 506 may include a high-speed RAM memory, and may also include a non-volatile memory, such as at least one disk memory.

[0136] The program 510 may specifically be called by the processor 502 to enable the non-azeotropic mixture refrigerant heat pump system to execute relevant steps in the above-mentioned embodiment of the control method for the non-azeotropic mixture refrigerant heat pump system.

[0137] Those of ordinary skill in the art can understand,Figure 9 The structure shown is only schematic and does not limit the structure of the above-mentioned device. For example, the zeotropic mixture refrigerant heat pump system may further include more or fewer components than those shown in Figure 9 or have a configuration different from that shown in Figure 9 .

[0138] Embodiments of the present invention also provide a computer-readable storage medium. The methods according to the embodiments of the present invention can be implemented in hardware, firmware, or be implemented as computer code that can be recorded on a storage medium, or be implemented as computer code originally stored in a remote storage medium or a non-transitory machine-readable storage medium and downloaded through a network and to be stored in a local storage medium, so that the methods described herein can be stored as such software processes on a storage medium using a general-purpose computer, a dedicated processor, or programmable or dedicated hardware. Among them, the storage medium can be a magnetic disk, an optical disk, a read-only memory, a random access memory, a flash memory, a hard disk, or a solid-state drive, etc.; further, the storage medium can also include a combination of the above-mentioned types of memories. It can be understood that a computer, a processor, a microprocessor controller, or programmable hardware includes a storage component that can store or receive software or computer code, and when the software or computer code is accessed and executed by the computer, the processor, or the hardware, the methods shown in the above embodiments are implemented.

[0139] In the specific content of the above specific embodiments, the technical features can be combined arbitrarily without contradiction. For the sake of brevity of description, not all possible combinations of the above technical features are described. However, as long as the combinations of these technical features do not exist in contradiction, they should all be considered as the scope described in this specification.

[0140] The specific content of the above specific embodiments only expresses several embodiments of the present invention, and its description is relatively specific and detailed, but it should not be understood as a limitation to the scope of the patent of the present invention. It should be noted that for those of ordinary skill in the art, without departing from the concept of the present invention, several modifications and improvements can still be made, and these all belong to the protection scope of the present invention. Therefore, the protection scope of the patent of the present invention should be subject to the appended claims.

Claims

1. Azeotropic mixture refrigerant heat pump system, characterized in that, Comprising: A compressor (1) and a refrigerant circulation loop connected to the compressor (1) and forming a circuit. Along the refrigerant flow direction on the refrigerant circulation loop, a condenser (2), a first throttling device (3), a regenerator (6), a second throttling device (8), and an evaporator (9) are sequentially provided; the regenerator (6) has a first heat exchange channel (61) and a second heat exchange channel (62) that exchange heat with each other. The first heat exchange channel (61) is connected between the first throttling device (3) and the second throttling device (8), and the second heat exchange channel (62) is connected between the outlet of the evaporator (9) and the suction port of the compressor (1); A first liquid injection pipeline (12) is connected between the outlet of the first heat exchange channel (61) and the outlet of the evaporator (9), and a first control valve (7) is provided on the first liquid injection pipeline (12); A second liquid injection pipeline (13) is connected between the outlet of the condenser (2) and the outlet of the evaporator (9), and a second control valve (10) is provided on the second liquid injection pipeline (13).

2. The non-azeotropic mixture refrigerant heat pump system according to claim 1, wherein Further comprising: A liquid receiver (4) is provided with a refrigerant inlet (41) and a first refrigerant outlet (42). The refrigerant inlet (41) is connected to the outlet of the condenser (2), and the first refrigerant outlet (42) is connected to the first heat exchange channel (61).

3. The non-azeotropic mixture refrigerant heat pump system according to claim 2, wherein Further comprising: A defrosting branch (14). The liquid receiver (4) is further provided with a second refrigerant outlet (43). Both ends of the defrosting branch (14) are communicated with the second refrigerant outlet (43) and the inlet of the evaporator (9) respectively, and an electronically controlled valve (5) with adjustable opening degree is provided on the defrosting branch (14).

4. The non-azeotropic mixture refrigerant heat pump system according to claim 3, characterized in that, The second refrigerant outlet (43) is arranged at the top of the liquid receiver (4), and the first refrigerant outlet (42) is arranged at the bottom of the liquid receiver (4).

5. The zeotropic mixture refrigerant heat pump system according to any one of claims 1 to 4, characterized in that Further comprising: An evaporator fan (15) for driving air to flow through the evaporator (9); A gas-liquid separator (11) is arranged on the refrigerant circulation loop and is located between the second heat exchange channel (62) and the suction port of the compressor (1).

6. A control method for a non-azeotropic mixture refrigerant heat pump system, which is applied to the non-azeotropic mixture refrigerant heat pump system described in any one of claims 1 to 5, characterized in that, The control method includes: In response to the heating mode, obtaining the actual exhaust temperature T of the compressor (1); Comparing the actual exhaust temperature T with the maximum exhaust temperature t2 and the minimum exhaust temperature t1 respectively to obtain a determination result; Based on the determination result, controlling the working state of at least one of the first control valve (7) and the second control valve (10).

7. The control method of the zeotropic mixture refrigerant heat pump system according to claim 6, characterized in that, The controlling the working state of at least one of the first control valve (7) and the second control valve (10) based on the determination result includes: When the determination result is t1 < T < t2, controlling the first control valve (7) and the second control valve (10) to maintain their current opening degrees unchanged; When the determination result is T ≤ t1 or T ≥ t2, adjusting the opening degree of at least one of the first control valve (7) and the second control valve (10).

8. The control method of the zeotropic mixture refrigerant heat pump system according to claim 7, characterized in that When the determination result is T≤t1 or T≥t2, adjusting the opening degree of at least one of the first control valve (7) and the second control valve (10) includes: When the determination result is T≤t1, reducing the opening degree of at least one of the first control valve (7) and the second control valve (10) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or both the first control valve (7) and the second control valve (10) are closed; When the determination result is T≥t2, increasing the opening degree of at least one of the first control valve (7) and the second control valve (10) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or both the first control valve (7) and the second control valve (10) are opened to the maximum opening degree; Wherein, the preset exhaust gas temperature Te satisfies: t1<Te<t2.

9. The control method of the zeotropic mixture refrigerant heat pump system according to claim 8, characterized in that, When the determination result is T≤t1, reducing the opening degree of at least one of the first control valve (7) and the second control valve (10) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or both the first control valve (7) and the second control valve (10) are closed, includes: When the determination result is T≤t1, obtaining the opening degree of the second control valve (10); In response to the opening degree of the second control valve (10) being greater than zero, reducing the opening degree of the second control valve (10) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or the second control valve (10) is closed; In response to the opening degree of the second control valve (10) being zero, reducing the opening degree of the first control valve (7) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or the first control valve (7) is closed.

10. The control method of the zeotropic mixture refrigerant heat pump system according to claim 8, characterized in that, When the determination result is T≥t2, increasing the opening degree of at least one of the first control valve (7) and the second control valve (10) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or both the first control valve (7) and the second control valve (10) are opened to the maximum opening degree, includes: When the determination result is T≥t2, obtaining the opening degree of the first control valve (7); In response to the opening degree of the first control valve (7) being less than the maximum opening degree, increasing the opening degree of the first control valve (7) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or the first control valve (7) reaches the maximum opening degree; In response to the opening degree of the first control valve (7) reaching the maximum opening degree, increasing the opening degree of the second control valve (10) until the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te is within a preset error range, or the second control valve (10) reaches the maximum opening degree.

11. The control method of the zeotropic mixture refrigerant heat pump system according to any one of claims 8 to 10, characterized in that, When the determination result is T≤t1 or T≥t2, adjusting the opening degree of at least one of the first control valve (7) and the second control valve (10) includes: When the determination result is T≤t1 or T≥t2, obtaining the temperature difference Ts between the actual exhaust gas temperature T and the preset exhaust gas temperature Te; Based on the temperature difference Ts, determining the target adjustment opening degree; Adjusting the opening degree of at least one of the first control valve (7) and the second control valve (10) based on the target adjustment opening degree; Wherein, the target adjustment opening degree is proportional to the temperature difference Ts.

12. The control method of the zeotropic mixture refrigerant heat pump system according to any one of claims 6 to 10, characterized in that, The heat pump system further includes a liquid receiver (4) and a defrosting branch (14). The liquid receiver (4) is provided with a refrigerant inlet (41) and a first refrigerant outlet (42). The refrigerant inlet (41) is connected to the outlet of the condenser (2), and the first refrigerant outlet (42) is connected to the first heat exchange channel (61); The liquid receiver (4) is further provided with a second refrigerant outlet (43). Both ends of the defrosting branch (14) are communicated with the second refrigerant outlet (43) and the inlet of the evaporator (9) respectively. An electronically controlled valve (5) with adjustable opening degree is provided on the defrosting branch (14); The control method further includes: In response to the defrosting mode, controlling the electronically controlled valve (5) to open, and controlling the second throttling device (8), the first control valve (7) and the second control valve (10) to close.

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