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

Abstract

The present invention relates to the field of heat pump technology, and discloses a non-azeotropic mixed refrigerant heat pump system and a control method thereof. The heat pump system includes: a compressor and a refrigerant circulation loop, wherein the refrigerant circulation loop is sequentially provided with a condenser, a first throttling device, a regenerator, a second throttling device, and an evaporator; a first heat exchange channel of the regenerator is connected between the first throttling device and the second throttling device, and a second heat exchange channel thereof is connected between the outlet of the evaporator and the return air port of the compressor; a first liquid spray pipeline is connected between the first heat exchange channel and the outlet of the evaporator and is provided with a first control valve; a second liquid spray pipeline is connected between the outlet of the condenser and the outlet of the evaporator and is provided with a second control valve. The present invention is provided with two liquid spray pipelines, and by controlling their respective control valves, the exhaust gas temperature can be accurately controlled, thereby preventing the exhaust gas temperature from being too high due to low temperature operating conditions of the heat pump system, helping to protect the compressor from high temperature damage, extending the service life, and improving system reliability.

Description

Technical Field

[0001] The present invention relates to the technical field of heat pumps, and in particular to a non-azeotropic mixed working fluid heat pump system and a control method thereof. Background Art

[0002] With the increasing global focus on renewable energy utilization and energy efficiency improvements, heat pump systems, as efficient and environmentally friendly energy conversion devices, have been widely used in both civil and industrial fields. However, when heat pump systems operate in ultra-low temperature environments, the extremely low evaporation temperature causes the low-pressure side of the system to be very low, which not only increases the workload of the compressor, but also requires the compressor to overcome a larger pressure ratio to complete the refrigerant circulation process, and causes the compressor's exhaust temperature to be too high. This high pressure differential operating state and excessively high exhaust temperature will greatly increase the compressor's energy consumption and may shorten the compressor's service life.

[0003] Furthermore, because ultra-low-temperature heat pumps need to provide sufficient heat under extreme conditions, the compressor must operate at higher efficiency, resulting in higher exhaust temperatures. Excessively high exhaust temperatures can cause problems such as lubricant carbonization and sealing material aging. In severe cases, this can lead to compressor failure or even damage, compromising the stability and reliability of the entire system. Summary of the Invention

[0004] The first technical problem solved by the present invention is to provide a non-azeotropic mixed working fluid heat pump system, which effectively solves the problem in the related art that the compressor exhaust temperature is too high when the heat pump system operates in a low temperature environment.

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

[0006] The first technical problem mentioned above is solved by the following technical solution:

[0007] A non-azeotropic mixed working fluid heat pump system, comprising:

[0008] A compressor and a refrigerant circulation circuit connected to the compressor to form a loop, wherein the refrigerant circulation circuit is provided with a condenser, a first throttling device, a regenerator, a second throttling device, and an evaporator in sequence along the refrigerant flow direction; the regenerator has a first heat exchange channel and a second heat exchange channel for mutual heat exchange, 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 return air port of the compressor;

[0009] a first liquid spray pipeline connected between the outlet of the first heat exchange channel and the outlet of the evaporator, and a first control valve is provided on the first liquid spray pipeline;

[0010] The second liquid spray 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 spray pipeline.

[0011] Compared with the background technology, the non-azeotropic mixed working fluid heat pump system of the present invention has the following beneficial effects:

[0012] In response to the problem of high-temperature exhaust caused by component segregation and increased pressure ratio when a non-azeotropic mixture is operated at ultra-low temperature, the present invention is provided with a liquid spray 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 return air temperature of the compressor. It can be understood that when the return air temperature of the compressor decreases, the exhaust temperature of the compressor will also decrease accordingly. Thus, the present invention achieves the technical effect of suppressing the rise in exhaust temperature, avoiding the heat pump system from having an excessively high exhaust temperature due to low temperature working conditions, helping to protect the compressor from high temperature damage, extending its service life and improving the reliability of the system. It is particularly suitable for the problem of high-temperature exhaust caused by component segregation and increased pressure ratio when a non-azeotropic mixture is operated at ultra-low temperature. At the same time, by providing a first throttling device, a regenerator and a second throttling device, the present application can perform graded throttling and heat recovery on the mixed refrigerant before it flows through the evaporator, thereby reducing throttling losses and thereby increasing the heating capacity.

[0013] Furthermore, the first liquid injection line draws liquid refrigerant from the regenerator outlet and injects it directly into the gaseous refrigerant at the evaporator outlet. This lowers the refrigerant temperature, thereby reducing the compressor's return air temperature and, consequently, the compressor's exhaust temperature. Because the injection point is located after the regenerator, the refrigerant has already been throttled and subcooled. The pressure differential between the refrigerant and the compressor's return air inlet is small, leading to a preliminary temperature drop. Consequently, the injection volume is small, minimizing the impact on the system's heating capacity and potentially reducing the exhaust temperature by 5°C-10°C.

[0014] The second liquid spray pipeline is used to extract medium-temperature and high-pressure liquid refrigerant from the condenser outlet and directly inject it into the gaseous refrigerant at the evaporator outlet. Since the injection point of the second liquid spray pipeline is located after the condenser, the pressure difference of the refrigerant relative to the compressor return air port is large, so the injection volume is large. By activating the second liquid spray pipeline, the compressor return air temperature can be significantly reduced, and the exhaust temperature can be further reduced by 10℃-20℃. However, under this operating condition, the compressor power increases and the system energy efficiency decreases.

[0015] Therefore, under the above premise, the present invention achieves precise exhaust temperature control and a dynamic balance between heating performance through a phased liquid injection design for the first and second liquid injection pipelines. For example, under low-temperature conditions, the first control valve is preferentially opened for light liquid injection. If the exhaust temperature remains above the set threshold, the second control valve is gradually opened for deep liquid injection. In this way, by adjusting the opening of the two control valves, a stepped control of the exhaust temperature is achieved, which not only ensures system stability in extreme low temperatures but also avoids a significant decrease in the heating efficiency of the heat pump system due to excessive liquid injection.

[0016] In addition, the two liquid injection lines can be selectively opened under different operating conditions, thereby achieving refined control of the exhaust temperature for optimal temperature management. For example, under light or medium loads, because the injection point of the first liquid injection line is located after the regenerator, the refrigerant has been throttled and subcooled, and the pressure difference between the refrigerant and the compressor return air inlet is small, so the injection volume is small. Simply opening the first liquid injection line can reduce the exhaust temperature by 5°C-10°C, with minimal impact on the system's heating capacity. Under high load or extreme conditions, because the injection point of the second liquid injection line is located after the condenser, the pressure difference between the refrigerant and the compressor return air inlet is large, so the injection volume is large. By activating the second liquid injection line, the exhaust temperature is further reduced by 10°C-20°C, thereby further reducing the exhaust temperature while maintaining heating capacity, ensuring stable operation of the system in various environments.

[0017] In one embodiment, the non-azeotropic mixed working fluid heat pump system further includes:

[0018] The liquid accumulator is provided with a refrigerant inlet and a first refrigerant outlet, wherein the refrigerant inlet is connected to the outlet of the condenser, and the first refrigerant outlet is connected to the first heat exchange channel.

[0019] In one embodiment, the non-azeotropic mixed working fluid heat pump system further includes:

[0020] The defrost branch circuit is provided with a second refrigerant outlet in the liquid storage device. Both ends of the defrost branch circuit are respectively connected with the second refrigerant outlet and the inlet of the evaporator. An electrically controlled valve with adjustable opening is provided on the defrost branch circuit.

[0021] In one embodiment, the second refrigerant outlet is disposed at the top of the liquid reservoir, and the first refrigerant outlet is disposed at the bottom of the liquid reservoir.

[0022] In one embodiment, the non-azeotropic mixed working fluid heat pump system further includes:

[0023] an evaporator fan, for driving air to flow through the evaporator;

[0024] The gas-liquid separator is arranged on the refrigerant circulation loop and is located between the second heat exchange channel and the return air port of the compressor.

[0025] The second technical problem mentioned above is solved by the following technical solution:

[0026] A control method for a non-azeotropic mixed refrigerant heat pump system is applied to the non-azeotropic mixed refrigerant heat pump system according to the first aspect of the present invention, the control method comprising:

[0027] In response to a heating mode, obtaining an actual exhaust temperature T of the compressor;

[0028] Comparing the actual exhaust temperature T with the maximum exhaust temperature t2 and the minimum exhaust temperature t1 respectively to obtain a determination result;

[0029] Based on the determination result, an operating state of at least one of the first control valve and the second control valve is controlled.

[0030] Compared with the background technology, the control method of the non-azeotropic mixed working fluid heat pump system of the present invention has the following beneficial effects:

[0031] To address the problem of high-temperature exhaust caused by component segregation and increased pressure ratio when a non-azeotropic mixture operates at ultra-low temperatures, the control method of the present invention controls the operating states of a first control valve and a second control valve to select whether a liquid spray line sprays liquid toward the outlet of the evaporator. In the case of liquid spraying, that is, when at least one of the first control valve and the second control valve is open, liquid refrigerant in the liquid spray line can be delivered to the outlet of the evaporator. At this time, the liquid refrigerant mixes with the gaseous refrigerant flowing out of the evaporator outlet, reducing the temperature of the refrigerant, thereby effectively reducing the return air temperature of the compressor. It can be understood that when the return air temperature of the compressor decreases, the exhaust temperature of the compressor will also decrease accordingly. Thus, the present invention achieves the technical effect of suppressing the rise in exhaust temperature, avoiding excessively high exhaust temperature due to low-temperature operating conditions in the heat pump system, helping to protect the compressor from high-temperature damage, extending its service life, and improving system reliability. It is particularly suitable for addressing the problem of high-temperature exhaust caused by component segregation and increased pressure ratio when a non-azeotropic mixture operates at ultra-low temperatures. At the same time, by setting up a first throttling device, a heat regenerator and a second throttling device, the present application can perform graded throttling and heat regenerator on the mixed refrigerant before the mixed refrigerant flows through the evaporator, thereby reducing throttling losses and increasing the heating capacity.

[0032] Furthermore, when the first control valve is opened, the first liquid injection line draws liquid refrigerant from the regenerator outlet and injects it directly into the gaseous refrigerant at the evaporator outlet. This lowers the refrigerant temperature, thereby reducing the compressor's return air temperature and, consequently, the compressor's exhaust temperature. Because the injection point is located after the regenerator, the refrigerant has already been throttled and subcooled. The pressure differential between the refrigerant and the compressor's return air inlet is small, leading to a preliminary temperature drop. Consequently, the injection volume is small, minimizing the impact on the system's heating capacity and potentially reducing the exhaust temperature by 5°C-10°C.

[0033] When the second control valve is opened, the second liquid spray 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 spray point of the second liquid spray pipeline is located after the condenser, the pressure difference of the refrigerant relative to the compressor return air port is large, so the spray volume is large. By activating the second liquid spray pipeline, the compressor return air temperature can be significantly reduced, and the exhaust temperature can be further reduced by 10℃-20℃. However, under this working condition, the compressor power increases and the system energy efficiency decreases.

[0034] Therefore, under the aforementioned premise, the present invention achieves precise exhaust temperature control and a dynamic balance between heating performance through a phased liquid injection design for the first and second liquid injection pipelines. For example, in low-temperature conditions, the first control valve is preferentially opened for light liquid injection. If the exhaust temperature remains above the set threshold, the second control valve is gradually opened for deeper liquid injection. In this way, by adjusting the openings of the two control valves, a stepped exhaust temperature control is achieved, ensuring system stability at extremely low temperatures while avoiding a significant decrease in heating efficiency due to excessive liquid injection.

[0035] In addition, under different operating conditions, the first and second control valves can be selectively opened to achieve refined control of the exhaust temperature for optimal temperature management. For example, under light or medium loads, since the injection point of the first liquid injection pipeline is located after the regenerator, the refrigerant has been throttled and subcooled, and the pressure differential between the refrigerant and the compressor return air inlet is small, resulting in a small injection volume. Simply opening the first control valve can reduce the exhaust temperature by 5°C-10°C, with minimal impact on the system's heating capacity. Under high load or extreme conditions, since the injection point of the second liquid injection pipeline is located after the condenser, the pressure differential between the refrigerant and the compressor return air inlet is large, resulting in a larger injection volume. By opening the second control valve, the exhaust temperature is further reduced by 10°C-20°C, thereby further reducing the exhaust temperature while maintaining heating capacity, ensuring stable operation of the system under various environments.

[0036] In one embodiment, controlling the working state of at least one of the first control valve and the second control valve based on the determination result includes:

[0037] When the determination result is t1<T<t2, controlling the first control valve and the second control valve to maintain the current openings unchanged;

[0038] When the determination result is T≦t1 or T≧t2, the opening degree of at least one of the first control valve and the second control valve is adjusted.

[0039] In one embodiment, when the determination result is T≤t1 or T≥t2, adjusting the opening of at least one of the first control valve and the second control valve includes:

[0040] When the determination result is T≤t1, reducing the opening of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve and the second control valve are both closed;

[0041] When the scheduling result is T≥t2, increasing the opening of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range and is zero, or the first control valve and the second control valve are both opened to the maximum opening;

[0042] The preset exhaust temperature Te satisfies: t1<Te<t2.

[0043] In one embodiment, when the determination result is T≤t1, reducing the opening of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve and the second control valve are both closed, includes:

[0044] When the determination result is T≤t1, obtaining the opening of the second control valve;

[0045] In response to the opening of the second control valve being greater than zero, reducing the opening of the second control valve until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the second control valve is closed;

[0046] In response to the opening of the second control valve being zero, the opening of the first control valve is reduced until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve is closed.

[0047] In one embodiment, when the scheduling result is T≥t2, increasing the opening of at least one of the first control valve and the second control valve until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve and the second control valve are both opened to the maximum opening, includes:

[0048] When the determination result is T≥t2, obtaining the opening of the first control valve;

[0049] In response to the opening of the first control valve being less than the maximum opening, increasing the opening of the first control valve until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve reaches the maximum opening;

[0050] In response to the opening of the first control valve reaching the maximum opening, the opening of the second control valve is increased until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the second control valve reaches the maximum opening.

[0051] In one embodiment, when the determination result is T≤t1 or T≥t2, adjusting the opening of at least one of the first control valve and the second control valve includes:

[0052] When the determination result is T≤t1 or T≥t2, obtaining the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te;

[0053] Determining a target adjustment opening based on the temperature difference Ts;

[0054] adjusting the opening of at least one of the first control valve and the second control valve based on the target adjustment opening;

[0055] The target adjustment opening is proportional to the temperature difference Ts.

[0056] In one embodiment, the heat pump system further includes a liquid reservoir and a defrost branch, the liquid reservoir 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 reservoir is further provided with a second refrigerant outlet, and both ends of the defrost branch are respectively connected to the second refrigerant outlet and the inlet of the evaporator, and the defrost branch is provided with an electrically controlled valve with an adjustable opening;

[0057] The control method further includes:

[0058] In response to the defrost mode, the electrically controlled valve is controlled to be opened, and the second throttling device, the first control valve and the second control valve are controlled to be closed. BRIEF DESCRIPTION OF THE DRAWINGS

[0059] In order to more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the following briefly introduces the drawings required for use in the specific embodiments or the description of the prior art. Obviously, the drawings described below are some embodiments of the present invention. For ordinary technicians in this field, other drawings can be obtained based on these drawings without paying any creative work.

[0060] Figure 1 This is a schematic structural diagram of a non-azeotropic mixed working fluid heat pump system in a heating mode according to an embodiment of the present invention;

[0061] Figure 2 This is a schematic structural diagram of a non-azeotropic mixed working fluid heat pump system in a defrost mode according to an embodiment of the present invention;

[0062] Figure 3 This is a flow chart of a control method for a non-azeotropic mixed working fluid heat pump system according to an embodiment of the present invention;

[0063] Figure 4 This is a second flow chart of a control method for a non-azeotropic mixed working fluid heat pump system according to an embodiment of the present invention;

[0064] Figure 5 This is a third flow chart of a control method for a non-azeotropic mixed working fluid heat pump system according to an embodiment of the present invention;

[0065] Figure 6 This is a fourth flow chart of a control method for a non-azeotropic mixed working fluid heat pump system according to an embodiment of the present invention;

[0066] Figure 7 This is a fifth flow chart of a control method for a non-azeotropic mixed working fluid heat pump system according to an embodiment of the present invention;

[0067] Figure 8 This is a schematic structural diagram of a control device for a non-azeotropic mixed working fluid heat pump system according to an embodiment of the present invention;

[0068] Figure 9 This is a structural schematic diagram of a non-azeotropic mixed working fluid heat pump system according to another embodiment of the present invention.

[0069] Description of reference numerals:

[0070] 1. Compressor; 2. Condenser; 3. First throttling device; 4. Liquid storage tank; 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 spray pipeline; 13. Second liquid spray pipeline; 14. Defrost branch; 15. Evaporator fan; 100. Acquisition module; 200. Judgment module; 300. Control module. DETAILED DESCRIPTION

[0071] The following will be combined with the drawings in the embodiments of this application to clearly and completely describe the technical solutions in the embodiments of this application. Obviously, the embodiments described are only part of the embodiments of this application, not all of the embodiments. Based on the embodiments in this application, all other embodiments obtained by ordinary technicians in this field without making creative efforts are within the scope of protection of this application.

[0072] In the description of this application, it should be understood that the terms "center", "up", "down", "front", "back", "left", "right", "vertical", "horizontal", "top", "bottom", "inside", "outside", etc., indicating the orientation or position relationship, are based on the orientation or position relationship shown in the accompanying drawings, and are only for the convenience of describing this application and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore cannot be understood as a limitation on this application.

[0073] 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 number of the technical features being referred to. Thus, a feature specified as "first" or "second" may explicitly or implicitly include one or more of such features. Throughout this application, unless otherwise specified, "plurality" means two or more.

[0074] In the description of this application, it should be noted that, unless otherwise expressly specified or limited, the terms "mounted," "connected," and "connected" should be understood in a broad sense. For example, they can refer to fixed connections, detachable connections, or integral connections; mechanical connections or electrical connections; direct connections or indirect connections through an intermediate medium; and internal connections between two components. Those skilled in the art will understand the specific meanings of the above terms in this application based on the specific circumstances.

[0075] A non-azeotropic mixed working fluid heat pump system and a control method thereof provided by the present invention will be described in detail below with reference to the accompanying drawings.

[0076] like Figure 1 and Figure 2 As shown, the non-azeotropic mixed working fluid heat pump system according to the first embodiment of the present invention includes a compressor 1, a refrigerant circulation loop, a first liquid spray pipeline 12 and a second liquid spray pipeline 13.

[0077] The refrigerant circulation loop is connected to the compressor 1 to form a loop. A condenser 2, a first throttling device 3, a regenerator 6, a second throttling device 8 and an evaporator 9 are arranged in sequence along the refrigerant flow direction on the refrigerant circulation loop; the regenerator 6 has a first heat exchange channel 61 and a second heat exchange channel 62 for mutual heat exchange. 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 return air port of the compressor 1.

[0078] The first liquid spray 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 spray pipeline 12; the second liquid spray 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 spray pipeline 13.

[0079] The specific structure of the non-azeotropic mixed working fluid heat pump system according to the embodiment of the present invention is described as follows:

[0080] The heat pump system primarily 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 spray line 12, a second liquid spray line 13, and corresponding control valves. Compressor 1 compresses the low-temperature, low-pressure gaseous mixed refrigerant into a high-temperature, high-pressure gas and then feeds it into condenser 2, where it releases heat and becomes a gas-liquid mixture. Subsequently, the mixed refrigerant passes through the first throttling device 3, where it is throttled and depressurized, and enters the first heat exchange channel 61 of the regenerator 6, where it undergoes heat exchange with the low-temperature, low-pressure gaseous refrigerant returning from the evaporator 9, further reducing its temperature. Afterwards, the refrigerant is further depressurized by the second throttling device 8, enters the evaporator 9, and absorbs ambient heat before returning to compressor 1, completing a complete cycle.

[0081] The first liquid spray line 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 spray line 12 is used to extract liquid refrigerant from the outlet of the regenerator 6 and directly inject it into the gaseous refrigerant at the outlet of the evaporator 9, thereby reducing the temperature of the mixed refrigerant and effectively reducing the return air temperature of the compressor 1. It can be understood that when the return air temperature of the compressor 1 drops, the exhaust temperature of the compressor 1 will also drop accordingly, thereby achieving the technical effect of suppressing the increase in the exhaust temperature of the compressor 1. It should be pointed out that since the liquid spray point is located after the regenerator 6, the refrigerant has been throttled and supercooled, and the pressure difference of the refrigerant relative to the return air port of the compressor 1 is small, and the temperature is initially reduced. Therefore, its liquid spray volume is small, and the impact on the heating capacity of the system is small, and the exhaust temperature can be reduced by 5-10°C.

[0082] The second liquid spray line 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 spray line 13 is used to extract 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, thereby reducing the temperature of the mixed refrigerant and effectively reducing the return air temperature of the compressor 1. It can be understood that when the return air temperature of the compressor 1 drops, the exhaust temperature of the compressor 1 will also drop accordingly, thereby achieving the technical effect of suppressing the increase in the exhaust temperature of the compressor 1. It should be pointed out that since the injection point of the second liquid spray line 13 is located after the condenser 2, the pressure difference between the refrigerant and the return air port of the compressor 1 is large, so the injection amount is large. By activating the second liquid spray line 13, the return air temperature of the compressor 1 can be significantly reduced, and the exhaust temperature can be further reduced by 10°C-20°C. However, under 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 mixed fluid heat pump system of the present invention solves the problem of excessively high exhaust temperature of the compressor 1 causing decreased system stability under ultra-low temperature heating conditions by setting a first liquid spray pipeline 12 and a second liquid spray pipeline 13. At the same time, a phased liquid spray strategy is used to balance the contradiction between heating performance and exhaust temperature control.

[0083] Specifically, in order to solve the problem of high-temperature exhaust caused by component segregation and increased pressure ratio when non-azeotropic mixed working fluids are running at ultra-low temperatures, the present invention introduces two key liquid spray pipelines, namely the first liquid spray pipeline 12 and the second liquid spray pipeline 13. Among them, the first liquid spray 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 allows the liquid refrigerant to mix with the gaseous refrigerant flowing out of the outlet of the evaporator 9, thereby reducing the temperature of the refrigerant, thereby effectively reducing the return air temperature of the compressor 1, and then achieving the technical effect of suppressing the rise in the exhaust temperature of the compressor 1, while having little impact on the heating performance. The second liquid spray pipeline 13 is led out from the outlet of the condenser 2, 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, high-pressure liquid refrigerant is directly sprayed into the outlet of the evaporator 9. Since the spray point of the second spray pipeline 13 is located behind the condenser 2, the pressure difference of the refrigerant relative to the return air port of the compressor 1 is large, so the spray volume is large. By activating the second spray pipeline 13, the return air 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.

[0084] Under low-temperature conditions, the first control valve 7 is preferentially opened for light liquid injection. If the exhaust temperature remains above the set threshold, the second control valve 10 is gradually opened for deeper liquid injection. In this way, by adjusting the openings of the two control valves, step-by-step control of the exhaust temperature is achieved, ensuring system stability in extreme low temperatures while preventing a significant drop in heating performance due to excessive liquid injection.

[0085] Furthermore, under ultra-low temperature heating conditions, the heat pump system of the present invention achieves precise control of the exhaust temperature through the phased liquid spraying operation of the first liquid spraying pipeline 12 and the second liquid spraying pipeline 13. The working process of the heat pump system of the present invention is as follows:

[0086] like Figure 1 As shown, the mixed refrigerant flows out of the condenser 2, is throttled and reduced in pressure 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 and the low-temperature refrigerant in the second heat exchange channel 62 complete heat exchange, further reducing the temperature of the refrigerant entering the evaporator 9. After further pressure reduction by the second throttling device 8, the refrigerant enters the evaporator 9 and absorbs heat and evaporates.

[0087] When the monitored exhaust temperature of compressor 1 is slightly higher than the set maximum exhaust temperature, the heat pump system preferentially opens first liquid injection line 12 for a light liquid injection. At this point, first control valve 7 partially opens, injecting a small amount of liquid refrigerant from first heat exchange channel 61 of regenerator 6 into the gaseous refrigerant at the outlet of evaporator 9. The liquid refrigerant mixes with the gaseous refrigerant flowing out of the evaporator 9 outlet, lowering the refrigerant temperature and, consequently, the return air temperature from compressor 1, thereby reducing the exhaust temperature by 5°C-10°C. During this process, heating performance degradation is minimal, making this operation suitable for moderately low-temperature operating conditions.

[0088] When the ambient temperature drops further, for example, when the ambient temperature is lower than -20°C, the exhaust temperature of 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 condenser 2 is injected into the gaseous refrigerant at the outlet of evaporator 9 through the second liquid spray line 13. At this time, a double liquid spray action is performed in the heat pump system. On the one hand, the first control valve 7 opens and maintains a light liquid spray to continue to lower the return air temperature of 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 evaporator 9 to further lower the return air temperature of compressor 1, thereby further lowering the exhaust temperature, and at the same time diluting the superheat of the gaseous refrigerant at the outlet of evaporator 9, reducing the load fluctuation of compressor 1. In the above process, the heating performance is greatly attenuated, but the stability of the system is significantly improved. Therefore, the above operation is suitable for extreme low temperature conditions.

[0089] Therefore, the present invention achieves a dynamic balance between precise exhaust temperature control and heating performance through the phased liquid injection design of the first and second liquid injection lines 12, 13. Under ultra-low temperature conditions, the system can flexibly switch liquid injection strategies based on actual needs, avoiding the risk of compressor 1 failure due to excessively high exhaust temperatures while minimizing the impact on energy efficiency, significantly improving the adaptability and reliability of the non-azeotropic mixture heat pump system.

[0090] In summary, according to the non-azeotropic mixed refrigerant heat pump system of the embodiment of the present invention, by setting the first liquid spray pipeline 12 and the second liquid spray 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 return air temperature of the compressor 1 is lowered by lowering the refrigerant temperature, thereby effectively lowering 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 maintain stable operation in an extremely low temperature environment, solving the problem of performance degradation and system instability of the traditional heat pump system caused by too low evaporation temperature and too large compression ratio under low temperature conditions. In addition, the present invention does not change the basic structure of the original heat pump system by optimizing the refrigerant circulation path and introducing a diversion branch at a key node, which is convenient for integration and engineering application and has good practicality.

[0091] like Figure 1 and Figure 2 As shown, according to some embodiments of the present invention, the heat pump system further includes a liquid reservoir 4 .

[0092] The liquid accumulator 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 .

[0093] In this embodiment, the liquid reservoir 4 has at least two interfaces: a refrigerant inlet 41 for receiving the mixed refrigerant discharged from the condenser 2; and a first refrigerant outlet 42 for delivering the stored or conditioned liquid refrigerant to the first heat exchange channel 61 of the regenerator 6. The liquid reservoir 4 is located between the condenser 2 and the regenerator 6 in the system. After flowing out of the condenser 2, the mixed refrigerant first enters the liquid reservoir 4 for storage or buffering. The conditioned liquid refrigerant in the liquid reservoir 4 then enters the first heat exchange channel 61 of the regenerator 6, exchanging heat with the gaseous refrigerant in the second heat exchange channel 62.

[0094] It can be understood that the liquid reservoir 4 plays the role of refrigerant storage, buffering and stabilizing flow, which can effectively prevent the impact of insufficient or fluctuating liquid refrigerant on the subsequent throttling device and evaporator 9. The liquid reservoir 4 is particularly suitable for variable load operation or low temperature startup, thereby improving the stability and reliability of the system.

[0095] In related art, the refrigerant flow rate at the condenser outlet may fluctuate during system operation, especially when using a variable-frequency compressor or operating in multiple modes. However, the present invention incorporates a liquid reservoir 4 within the heat pump system to temporarily store liquid refrigerant within its internal space, acting as a buffer pool and preventing throttling device malfunctions or unstable liquid supply to the evaporator 9 due to transient flow rate fluctuations.

[0096] 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 in the liquid storage tank 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.

[0097] On the other hand, under ultra-low temperature heating conditions, the refrigerant circulation volume may increase. Reservoir 4 can serve as a refrigerant reserve container to prevent problems such as refrigerant supply shortages and ensure safe system operation. Furthermore, if the system is also equipped with a first liquid spray line 12 and a second liquid spray line 13, the presence of reservoir 4 can further ensure the sufficiency and controllability of the sprayed refrigerant, making exhaust temperature regulation more accurate and reliable.

[0098] like Figure 1 and Figure 2As shown, the heat pump system further includes a defrost branch 14. The liquid reservoir 4 is further provided with a second refrigerant outlet 43. The two ends of the defrost branch 14 are respectively connected to the second refrigerant outlet 43 and the inlet of the evaporator 9. The defrost branch 14 is provided with an electrically controlled valve 5 with an adjustable opening.

[0099] In this embodiment, the second refrigerant outlet 43 of the reservoir 4 serves as a dedicated outlet for defrost refrigerant. It is activated when the system enters the defrost phase, supplying high-temperature gaseous refrigerant to the evaporator 9 to accelerate the melting of the frost layer. The defrost branch 14 provides a refrigerant passage that bypasses conventional throttling devices, allowing high-temperature refrigerant from the reservoir 4 to flow directly to the evaporator 9, rapidly raising the evaporator 9 temperature and achieving efficient defrosting. The electrically controlled valve 5 controls the opening and closing of the defrost branch 14, and its opening can be adjusted according to actual needs, thereby precisely controlling the amount of refrigerant entering the evaporator 9 and avoiding excessive impact or energy waste.

[0100] When the system detects that the evaporator 9 is severely frosted, for example, through a pressure difference sensor, a temperature sensor or timing logic judgment, the defrost program is started. The specific process is as follows: Figure 2 As shown, the system enters the defrost phase, closing the first throttling device 3 on the main refrigerant path or adjusting it to a high opening. The electrically controlled valve 5 is opened, allowing the high-temperature gaseous refrigerant from the reservoir 4 to enter the evaporator 9 through the defrost branch 14. The high-temperature refrigerant rapidly releases heat in the evaporator 9, significantly raising the surface temperature of the evaporator 9 and causing the frost layer to melt quickly. After the frost layer is cleared, the electrically controlled valve 5 is closed, and the main refrigerant circulation path is restored, and the system re-enters normal heating mode.

[0101] Thus, by introducing the defrost branch 14 and the electrically controlled valve 5, and in conjunction with the second refrigerant outlet 43 of the liquid reservoir 4, the present invention can utilize the high-temperature gaseous refrigerant from the liquid reservoir 4 to directly heat the evaporator 9, significantly accelerating the melting of the frost layer and shortening the defrost time. Compared to the traditional reverse defrost method, this solution does not require changing the operating direction of the compressor 1, thereby reducing system energy loss.

[0102] Furthermore, the second refrigerant outlet 43 is provided at the top of the liquid reservoir 4 , and the first refrigerant outlet 42 is provided at the bottom of the liquid reservoir 4 .

[0103] It is understood that after the condensed refrigerant enters the liquid reservoir 4, due to gravity and density differences, the liquid refrigerant naturally settles at the bottom of the reservoir 4, while the gaseous or liquid-gas mixture floats to the top. The first refrigerant outlet 42 is located at the bottom to ensure that the output is pure liquid refrigerant, suitable for the main circulation path requiring a stable liquid working medium, such as the regenerator 6, the throttling device, and the evaporator 9. The second refrigerant outlet 43 is located at the top to facilitate the extraction of high-temperature and high-pressure gaseous refrigerant for defrosting.

[0104] like Figure 1 and Figure 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.

[0105] 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.

[0106] 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 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.

[0107] 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:

[0108] Step S100, in response to the heating mode, obtain the actual exhaust temperature T of the compressor 1;

[0109] 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;

[0110] 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.

[0111] For the control method of the zeotropic mixture refrigerant heat pump system according to the embodiments of the present invention, the specific control logic is introduced as follows:

[0112] 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 rely on the current actual exhaust temperature value. Among them, the exhaust temperature T reflects the working state inside the compressor 1 and the load situation it faces. Too high or too low exhaust temperature will affect the performance and lifespan of the system.

[0113] 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.

[0114] Specifically, the system continuously compares the actual exhaust temperature T with these two thresholds. If T > t1, it means the exhaust temperature is too high and measures need to be taken to reduce the temperature; if T < t2, it indicates 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 no adjustment of the liquid injection valve is required.

[0115] In step S300, based on the determination result in the previous step, the system will dynamically adjust the states of the first control valve 7 and the second control valve 10 to regulate the flow of refrigerant entering the outlet of the evaporator 9. For example, when T>t1, the first control valve 7 is opened first: if the exhaust temperature is only slightly higher than t1, it can be moderately cooled by light liquid spraying. At this time, it is only necessary to open the first control valve 7 and use the liquid refrigerant in the first heat exchange channel 61 of the regenerator 6 for preliminary cooling. If T continues to rise to exceed a certain set value, it is necessary to further open the second control valve 10 to introduce more liquid refrigerant to be sprayed directly into the outlet of the evaporator 9 to achieve a greater cooling effect.

[0116] Furthermore, the specific control process of the heat pump system of the present invention is as follows: After the system enters heating mode, it begins collecting real-time data on the exhaust temperature T of compressor 1 and transmits it to the control system. The system continuously compares the actual exhaust temperature T with the preset upper minimum exhaust temperature. During this process, the system can quickly determine whether the current exhaust temperature deviates from the normal range and determine the next step accordingly. Based on this determination, the system automatically adjusts the status of the first control valve 7 and the second control valve 10.

[0117] For example, if the exhaust temperature slightly exceeds the standard, the system only needs to fine-tune the opening of the first control valve 7 to inject an appropriate amount of liquid refrigerant into the outlet of the evaporator 9 to restore the exhaust temperature to normal. For another example, when faced with a significant overtemperature caused by extreme low temperatures or high load conditions, the system will simultaneously open both control valves to ensure that sufficient refrigerant is injected into the outlet of the evaporator 9, quickly suppressing the rising trend of the exhaust temperature.

[0118] Once the exhaust temperature is detected to have fallen back into the safe range, the system will gradually close or reduce the opening of the liquid injection valve, resuming the normal refrigerant circulation process and ensuring efficient system operation. It is important to note that the system continuously monitors the exhaust temperature throughout its operation cycle and makes timely adjustments to the control strategy based on the latest exhaust temperature feedback, forming a closed-loop control mechanism to ensure that the system always operates in optimal condition regardless of changing external conditions.

[0119] In summary, through the above-mentioned detailed control method, the non-azeotropic mixed working fluid heat pump system of the present invention can flexibly respond to various complex working conditions, ensuring that the exhaust temperature of the compressor 1 always remains within a reasonable range, thereby improving the reliability of the system and improving the overall energy efficiency performance.

[0120] Specifically, the control method of the present invention monitors the actual exhaust temperature of the compressor 1 in real time and compares it with the preset upper minimum exhaust temperature, thereby enabling precise regulation of the exhaust temperature, thereby helping to avoid problems such as carbonization of lubricating oil and aging of sealing materials caused by excessively high temperatures, while also preventing excessively low temperatures from affecting system efficiency.

[0121] 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.

[0122] 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:

[0123] 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;

[0124] 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.

[0125] 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. 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.

[0126] 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.

[0127] Among them, for the case of T ≤ t1, the control system may choose to reduce or shut off 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.

[0128] 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 gas of the compressor 1, thereby effectively suppressing the rise of the exhaust temperature.

[0129] It should be noted that the operation of adjusting the opening of the first control valve 7 and the opening of the second control valve 10 can be gradual, that is, gradually increasing or decreasing the opening, or it can be a one-time large-scale adjustment, and the present invention does not impose any special restrictions on this.

[0130] Thus, through 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 operating conditions, avoiding unnecessary adjustments. It also enables timely response under extreme conditions, ensuring safe and efficient system operation. Furthermore, by preventing the compressor 1 from operating at excessively high or low exhaust temperatures for extended periods, the present invention reduces the risk of damage to the equipment and helps extend the service life of the entire system.

[0131] like Figure 4 As shown, according to some embodiments of the present invention, when the determination result is T≤t1 or T≥t2, step S320 of adjusting the opening of at least one of the first control valve 7 and the second control valve 10 includes:

[0132] Step S321: When the judgment result is T≤t1, the opening of at least one of the first control valve 7 and the second control valve 10 is reduced until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve 7 and the second control valve 10 are both closed;

[0133] Step S322: When the scheduling result is T≥t2, increase the opening 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 first control valve 7 and the second control valve 10 are both opened to the maximum opening.

[0134] The preset exhaust temperature Te satisfies: t1<Te<t2.

[0135] In this embodiment, when the determination result is that the actual exhaust temperature T is not within the preset ideal range, the system adjusts the exhaust temperature by adjusting the opening 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.

[0136] Specifically, when the judgment result is T ≤ t1, the actual exhaust temperature of the system is lower than the set minimum exhaust temperature t1, which means that the exhaust temperature is too low. Since excessive refrigerant injection will cause further cooling, the system needs to reduce the injection amount to increase the exhaust temperature.

[0137] The control system begins to gradually reduce the opening of the first control valve 7 and / or the second control valve 10, and obtains the actual exhaust temperature of the compressor 1 and the current opening of the control valve again after each opening adjustment 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, which indicates that the exhaust temperature has returned to the ideal state; or, the first control valve 7 and the second control valve 10 are both completely closed, and no refrigerant is sprayed into the outlet of the evaporator 9 to prevent further cooling.

[0138] When the judgment 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. In this case, it is necessary to increase the injection amount to reduce the exhaust temperature.

[0139] The control system begins to gradually increase the opening of the first control valve 7 and / or the second control valve 10 and obtains the actual exhaust temperature of the compressor 1 and the current opening of the control valve again after each opening adjustment 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, indicating that the exhaust temperature has dropped to the ideal level; or, the first control valve 7 and the second control valve 10 are both opened to the maximum opening, ensuring that the spray cooling effect is maximized to quickly reduce the exhaust temperature.

[0140] It should be explained that the preset exhaust temperature Te is an ideal value between the upper and lower minimum exhaust temperatures. It represents the optimal operating point that will neither cause equipment overheating nor affect efficiency due to too low a temperature. For example, the preset exhaust temperature Te = (t1 + t2) / 2. In this way, by adjusting the actual exhaust temperature T to close to Te, the system can avoid potential risks while ensuring performance, such as carbonization of lubricating oil or aging of sealing materials. The above-mentioned preset error range is an artificially set allowable error range, the purpose of which is to improve the fault tolerance of the adjustment and prevent errors from affecting the normal control logic. For example, the preset error range can be -1°C to 1°C.

[0141] In summary, this approach not only improves the system's response speed and accuracy, but also enhances its adaptability and reliability under different operating conditions, ensuring long-term stable operation. Furthermore, thanks to the use of a graded regulation strategy, the spray volume can be flexibly adjusted according to actual needs, achieving optimal energy efficiency.

[0142] like Figure 5As shown, in some specific embodiments, when the determination result is T≤t1, the opening of at least one of the first control valve 7 and the second control valve 10 is reduced until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve 7 and the second control valve 10 are both closed, step S321, includes:

[0143] Step S3211: When the determination result is T≤t1, obtain the open / close state of the second control valve 10;

[0144] Step S3212: In response to the second control valve 10 being in the open state, reducing the opening of the second control valve 10 until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the second control valve 10 is closed;

[0145] Step S3213: In response to the second control valve 10 being in the closed state, reduce the opening of the first control valve 7 until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve 7 is closed.

[0146] In this embodiment, when the actual exhaust temperature T ≤ t1 is detected, it indicates that the current exhaust temperature is too low, which may cause problems such as liquid carryover in the suction air of compressor 1 and reduced heating efficiency. Therefore, the control system needs to gradually reduce the liquid injection amount to increase the return air temperature of compressor 1, thereby raising the exhaust temperature. Because the two control valves control the liquid injection amount in different paths, the system adopts a strategy of first closing the second control valve 10 and then adjusting the first control valve 7 to achieve more precise temperature regulation.

[0147] Specifically, the system first determines whether the second control valve 10 is currently open. If it is open, it indicates that the high-pressure liquid refrigerant from the outlet of the condenser 2 is currently being injected, which has a strong cooling capacity. If it is closed, it means that only the first control valve 7 is currently performing liquid injection regulation.

[0148] In response to the second control valve 10 being in the open state, the control system begins to gradually reduce the opening of the second control valve 10 by a certain percentage each time, such as 5% to 10%, while continuously monitoring the changes in the actual exhaust temperature. During this process, the system continuously calculates the difference between the actual exhaust temperature T and the target temperature Te.

[0149] The above adjustment will continue until any of the following conditions is met: the actual exhaust temperature T rises to be equal to 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, at which time the adjustment is stopped; the second control valve 10 is completely closed, at which time continued adjustment will be invalid and the next step needs to be performed.

[0150] In response to the second control valve 10 being in the closed state, if the actual exhaust temperature is still below the target value Te after the second control valve 10 has been closed, the system switches to regulating the first control valve 7. The control system begins to gradually reduce the opening of the first control valve 7, also using a step-by-step adjustment method, and monitors the actual exhaust temperature changes in real time.

[0151] The above adjustment process will continue until any of the following conditions are 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 completely closed, at which time all the liquid injection paths are closed and the system enters a non-liquid injection operation state.

[0152] It should be noted that, because the refrigerant flowing through the second control valve 10 is a high-pressure, high-temperature liquid refrigerant injected directly from the condenser 2, its presence significantly reduces the operating temperature of the evaporator 9, thereby affecting the heating efficiency of the entire system. It is understood that, in the event of excessively low exhaust gas temperatures, the above steps, by first reducing or closing the second control valve 10, can effectively reduce the amount of additional refrigerant entering the outlet of the evaporator 9, thereby minimizing interference with normal system operation and maintaining the stability of the main circulation refrigerant flow.

[0153] Furthermore, prioritizing adjustment of the second control valve 10 is also due to its greater impact on exhaust temperature. If the exhaust temperature can be raised to the target value Te simply by reducing the opening of the second control valve 10, further adjustment of the first control valve 7 is unnecessary, thus avoiding system fluctuations caused by over-adjustment. If the exhaust temperature still does not reach the desired level after fully closing the second control valve 10, further fine-tuning of the first control valve 7 is necessary to achieve more precise temperature control.

[0154] In summary, the present invention prioritizes closing the second control valve 10 to reduce unnecessary injection of high-pressure, high-temperature liquid refrigerant, thus avoiding excessive impact on the main circulation refrigerant volume and thus ensuring the system's heating efficiency. Furthermore, the aforementioned steps are adjusted in stages based on actual exhaust temperature changes, enabling rapid response to abnormal conditions while preventing new problems caused by over-adjustment, ensuring stable system operation.

[0155] like Figure 6 As shown, in other specific embodiments, when the scheduling result is T≥t2, the opening of at least one of the first control valve 7 and the second control valve 10 is increased until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve 7 and the second control valve 10 are both opened to the maximum opening, step S322 includes:

[0156] Step S3221: When the determination result is T≥t2, obtain the opening degree of the first control valve 7;

[0157] Step S3222: In response to the opening of the first control valve 7 being less than the maximum opening, the opening of the first control valve 7 is increased until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve 7 reaches the maximum opening;

[0158] Step S3223: In response to the opening of the first control valve 7 reaching the maximum opening, increase the opening of 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 second control valve 10 reaches the maximum opening.

[0159] In this embodiment, when the actual exhaust temperature T ≥ t2 is detected, it indicates that the current exhaust temperature is too high, which may lead to risks such as lubricant carbonization and sealing material aging, affecting the life of compressor 1 and system stability. Therefore, the control system needs to gradually increase the injection volume to reduce the return air temperature of compressor 1, thereby suppressing the rise in exhaust temperature. Because the two control valves control the injection flow rate and cooling capacity of different paths respectively, the system adopts a strategy of first adjusting the first control valve 7 and then adjusting the second control valve 10, achieving more precise and energy-saving temperature regulation.

[0160] Specifically, the system first determines whether the first control valve 7 has reached its maximum allowable opening. If it has not, there is room to further increase the injection volume. If it has, it is necessary to consider activating or continuing to open the second control valve 10 to enhance the cooling effect.

[0161] In response to the opening of the first control valve 7 being less than the maximum opening, the control system begins to gradually increase the opening of the first control valve 7 by a certain percentage each time, while continuously monitoring changes in the actual exhaust temperature. During this process, the system continuously calculates the difference between the actual exhaust temperature T and the target temperature Te.

[0162] 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, at which time the adjustment is stopped; the first control valve 7 has been fully opened to the maximum allowable opening, at which time continued adjustment will be invalid and the next step must be entered.

[0163] When the first control valve 7 is at its maximum opening, if the actual exhaust temperature is still above the target value Te, the system switches to regulating the second control valve 10. The control system begins to gradually increase the opening of the second control valve 10, also using a step-by-step adjustment method, while monitoring the actual exhaust temperature changes in real time.

[0164] 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, at which time the injection volume has reached the limit and the system enters the maximum cooling state.

[0165] It should be noted that when increasing the opening of first control valve 7 or second control valve 10, medium-temperature, medium-pressure refrigerant is injected into first control valve 7. Liquid refrigerant from first heat exchange channel 61 of regenerator 6 is relatively low in temperature but moderate in pressure. This has a moderate cooling effect on the exhaust gas temperature after being injected into the outlet of evaporator 9. In contrast, this has a minimal impact on the main refrigerant cycle, does not significantly reduce the amount of heat absorbed by evaporator 9, and thus maintains a high energy efficiency ratio.

[0166] The second control valve 10 injects high-pressure, medium-temperature refrigerant. The liquid refrigerant from the outlet of condenser 2 has a higher pressure and offers a stronger cooling capability. However, the larger the amount of liquid injected, the greater the impact on the heating efficiency of the heat pump system. Therefore, the second control valve 10 is activated as a supplementary measure only when the first control valve 7 is fully open but still unable to effectively control the exhaust gas temperature.

[0167] In addition, the above steps can also prevent the sudden drop in the amount of main circulation refrigerant due to the sudden opening of the second control valve 10, thereby avoiding system instability or a sharp drop in efficiency.

[0168] In summary, the above steps, by prioritizing the opening of the first control valve 7, can reduce unnecessary injection of high-pressure, high-temperature liquid refrigerant, avoiding excessive impact on the main circulation refrigerant volume, thereby ensuring the system's heating efficiency. Furthermore, this method, by performing adjustments in steps based on actual exhaust temperature changes, can both quickly respond to abnormal situations and prevent new problems caused by over-adjustment, thereby ensuring stable system operation.

[0169] like Figure 7 As shown, according to some embodiments of the present invention, when the determination result is T≤t1 or T≥t2, step S320 of adjusting the opening of at least one of the first control valve 7 and the second control valve 10 includes:

[0170] 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;

[0171] Step S324: determining the target adjustment opening based on the temperature difference Ts;

[0172] Step S325: Adjust the opening of at least one of the first control valve 7 and the second control valve 10 based on the target adjustment opening.

[0173] Among them, the target adjustment opening is proportional to the temperature difference Ts.

[0174] In this embodiment, when the system detects that T ≤ t1 or T ≥ t2, it enters the adjustment stage, and the controller calculates the absolute difference Ts = |T - Te| between the current actual exhaust temperature T and the preset exhaust temperature Te; where 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.

[0175] 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.

[0176] 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 return air temperature, thereby increasing the exhaust temperature.

[0177] 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.

[0178] 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.

[0179] 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, thus effectively preventing over-adjustment. In addition, the above steps can avoid unnecessary excessive liquid injection or flow interruption, maintain the stability of the main circulating refrigerant volume, and ensure the heating performance.

[0180] 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. The two ends of the defrosting branch 14 are respectively communicated with the second refrigerant outlet 43 and the inlet of the evaporator 9, and an electronically controlled valve 5 with an adjustable opening degree is provided on the defrosting branch 14.

[0181] The control method further includes:

[0182] Step S400: In response to the defrost mode, the electronically controlled valve 5 is controlled to be open, and the second throttling device 8, the first control valve 7 and the second control valve 10 are controlled to be closed.

[0183] In this embodiment, the above steps switch the refrigerant flow path during the defrost stage, close unnecessary liquid spray and throttling channels, and enable the dedicated defrost branch 14, thereby achieving a fast and efficient defrost process.

[0184] The specific control process of defrost mode is as follows:

[0185] First, the system determines whether it is necessary to start defrosting. The system can determine to start defrosting based on the surface temperature of the evaporator 9 being continuously lower than the set value, or it can also determine to start defrosting based on an abnormal drop in suction pressure, the system running time reaching a preset defrost cycle, or the sensor detecting severe frost on the surface of the evaporator 9. The present invention does not impose any special restrictions on this.

[0186] Once it is confirmed that the defrost mode has been entered, the control system immediately executes the next operation. The control system opens the electric control valve 5 on the defrost branch 14; the high-temperature refrigerant flows out from the second refrigerant outlet 43 of the liquid reservoir 4 and directly enters the inlet of the evaporator 9 through the defrost branch 14; the refrigerant releases a large amount of heat in the evaporator 9, causing the frost layer on the surface of the evaporator 9 to melt quickly.

[0187] In order to ensure that the refrigerant flows to the evaporator 9 for defrosting as much as possible and to avoid interference or energy waste, the system simultaneously performs the following operations: close the second throttling device 8, thereby blocking the refrigerant in the main circulation path from flowing to the evaporator 9, preventing the refrigerant from continuing to evaporate and absorb heat and affecting the defrosting effect; close the first control valve 7, thereby stopping the spraying of liquid to the outlet of the evaporator 9, avoiding the refrigerant diversion causing the defrosting efficiency to decrease; close the second control valve 10, which can also prevent the refrigerant from being sprayed from the outlet of the condenser 2 into the outlet of the evaporator 9, ensuring that the defrost branch 14 is the dominant path.

[0188] In summary, the above steps utilize high-temperature refrigerant to directly heat the evaporator 9, significantly accelerating the melting speed of the frost layer and shortening the defrosting time. In addition, the above operation can realize the defrosting operation without reversing the four-way valve, which significantly reduces the energy loss of the system.

[0189] The control device for the non-azeotropic mixed working fluid heat pump system provided by the present invention will be described in detail below with reference to the accompanying drawings. It should be noted that the control device is applied to the non-azeotropic mixed working fluid heat pump system of the first aspect of the present invention.

[0190] like Figure 8 As shown, the control device of the non-azeotropic mixed working fluid heat pump system according to the third aspect of the present invention includes:

[0191] An acquisition module 100 is configured to acquire an actual exhaust temperature T of the compressor 1 in response to a heating mode;

[0192] The judgment module 200 is used to compare the actual exhaust temperature T with the maximum exhaust temperature t2 and the minimum exhaust temperature t1 to obtain a judgment result;

[0193] The control module 300 is configured to control the working state of at least one of the first control valve 7 and the second control valve 10 based on the determination result.

[0194] Figure 9 The schematic structural diagram of an embodiment of a non-azeotropic mixed working fluid heat pump system provided by an embodiment of the present invention is shown. The specific embodiment of the present invention does not limit the specific implementation of the non-azeotropic mixed working fluid heat pump system.

[0195] like Figure 9 As shown, the heat pump may include: a processor 502 , a communications interface 504 , a memory 506 , and a communication bus 508 .

[0196] Processor 502, communication interface 504, and memory 506 communicate with each other via communication bus 508. Communication interface 504 is used to communicate with other devices, such as client devices or other server network elements. Processor 502 is used to execute program 510, which may specifically perform the steps of the aforementioned embodiment of the control method for a non-azeotropic mixed working fluid heat pump system.

[0197] Specifically, the program 510 may include program code including computer-executable instructions.

[0198] Processor 502 may be a central processing unit (CPU), an application-specific integrated circuit (ASIC), or one or more integrated circuits configured to implement embodiments of the present invention. The one or more processors included in the non-azeotropic mixture heat pump system may be processors of the same type, such as one or more CPUs, or may be processors of different types, such as one or more CPUs and one or more ASICs.

[0199] 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 (non-volatile memory), such as at least one disk storage.

[0200] The program 510 can be specifically called by the processor 502 to enable the non-azeotropic mixed working fluid heat pump system to execute the relevant steps in the above-mentioned control method embodiment for the non-azeotropic mixed working fluid heat pump system.

[0201] It can be understood by those skilled in the art that Figure 9 The structure shown is only for illustration and does not limit the structure of the above-mentioned equipment. For example, the non-azeotropic mixed working fluid heat pump system may also include Figure 9 More or fewer components than shown, or with Figure 9 Different configurations shown.

[0202] The embodiment of the present invention also provides a computer-readable storage medium. The above-mentioned method according to the embodiment of the present invention can be implemented in hardware, firmware, or implemented as a computer code that can be recorded in a storage medium, or implemented as a computer code that is originally stored in a remote storage medium or a non-temporary machine-readable storage medium and downloaded through a network and will be stored in a local storage medium, so that the method described herein can be stored in such software processing 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 storage 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 memory. 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. When the software or computer code is accessed and executed by a computer, a processor or hardware, the method shown in the above embodiment is implemented.

[0203] In the specific contents of the above-mentioned specific implementation methods, the various technical features can be combined in any non-contradictory manner. In order to make the description concise, not all possible combinations of the above-mentioned technical features are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0204] The specific contents of the above-mentioned specific embodiments merely represent several embodiments of the present invention. While the description is relatively specific and detailed, it should not be construed as limiting the scope of the present invention. It should be noted that a person skilled in the art may make various modifications and improvements without departing from the scope of the present invention, and these modifications and improvements fall within the scope of protection of the present invention. Therefore, the scope of protection of the present invention shall be based on the appended claims.

Claims

1. A non-azeotropic mixed working fluid heat pump system, characterized in that: include: A compressor (1) and a refrigerant circulation circuit connected to the compressor (1) to form a loop, wherein a condenser (2), a first throttling device (3), a regenerator (6), a second throttling device (8) and an evaporator (9) are sequentially provided on the refrigerant circulation circuit along the refrigerant flow direction; the regenerator (6) has a first heat exchange channel (61) and a second heat exchange channel (62) for mutual heat exchange, 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 return air port of the compressor (1); a first liquid spraying pipeline (12) 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 spraying pipeline (12); a second liquid spraying pipeline (13) 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 spraying pipeline (13); The liquid storage device (4) is provided with a refrigerant inlet (41) and a first refrigerant outlet (42), wherein 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 defrost branch (14) is provided with a second refrigerant outlet (43) in the liquid storage device (4), and the two ends of the defrost branch (14) are respectively connected to the second refrigerant outlet (43) and the inlet of the evaporator (9), and the defrost branch (14) is provided with an electrically controlled valve (5) with an adjustable opening.

2. The non-azeotropic mixed working fluid heat pump system according to claim 1, characterized in that: The second refrigerant outlet (43) is arranged at the top of the liquid reservoir (4), and the first refrigerant outlet (42) is arranged at the bottom of the liquid reservoir (4).

3. The non-azeotropic mixed working fluid heat pump system according to any one of claims 1 to 2, characterized in that: Also includes: an evaporator fan (15) for driving 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 return air port of the compressor (1).

4. A control method for a non-azeotropic mixed working fluid heat pump system, applied to the non-azeotropic mixed working fluid heat pump system according to any one of claims 1 to 3, 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, the operating state of at least one of the first control valve (7) and the second control valve (10) is controlled.

5. The control method of the non-azeotropic mixed working fluid heat pump system according to claim 4, characterized in that: The controlling of the working state of at least one of the first control valve (7) and the second control valve (10) based on the determination result comprises: When the determination result is t1<T<t2, controlling the first control valve (7) and the second control valve (10) to maintain the current opening unchanged; When the determination result is T≤t1 or T≥t2, the opening degree of at least one of the first control valve (7) and the second control valve (10) is adjusted.

6. The control method of the non-azeotropic mixed working fluid heat pump system according to claim 5, characterized in that: When the determination result is T≤t1 or T≥t2, adjusting the opening of at least one of the first control valve (7) and the second control valve (10) comprises: When the determination result is T≤t1, reducing the opening 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 a preset error range, or the first control valve (7) and the second control valve (10) are both closed; When the judgment result is T≥t2, the opening of at least one of the first control valve (7) and the second control valve (10) is increased until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve (7) and the second control valve (10) are both opened to the maximum opening; The preset exhaust temperature Te satisfies: t1<Te<t2.

7. The control method of the non-azeotropic mixed working fluid heat pump system according to claim 6, characterized in that: When the determination result is T≤t1, the opening of at least one of the first control valve (7) and the second control valve (10) is reduced until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve (7) and the second control valve (10) are both closed, comprising: When the determination result is T≤t1, obtaining the opening degree of the second control valve (10); In response to the opening of the second control valve (10) being greater than zero, reducing the opening of the second control valve (10) until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the second control valve (10) is closed; In response to the opening of the second control valve (10) being zero, the opening of the first control valve (7) is reduced until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve (7) is closed.

8. The control method of the non-azeotropic mixed working fluid heat pump system according to claim 6, characterized in that: When the determination result is T≥t2, the opening of at least one of the first control valve (7) and the second control valve (10) is increased until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve (7) and the second control valve (10) are both opened to the maximum opening, comprising: When the determination result is T≥t2, obtaining the opening of the first control valve (7); In response to the opening of the first control valve (7) being less than the maximum opening, increasing the opening of the first control valve (7) until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the first control valve (7) reaches the maximum opening; In response to the opening of the first control valve (7) reaching the maximum opening, the opening of the second control valve (10) is increased until the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te is within a preset error range, or the second control valve (10) reaches the maximum opening.

9. The control method of a non-azeotropic mixed working fluid heat pump system according to any one of claims 6 to 8, characterized in that: When the determination result is T≤t1 or T≥t2, adjusting the opening of at least one of the first control valve (7) and the second control valve (10) comprises: When the determination result is T≤t1 or T≥t2, obtaining the temperature difference Ts between the actual exhaust temperature T and the preset exhaust temperature Te; Determining a target adjustment opening based on the temperature difference Ts; adjusting the opening of at least one of the first control valve (7) and the second control valve (10) based on the target adjustment opening; The target adjustment opening is proportional to the temperature difference Ts.

10. The control method of a non-azeotropic mixed working fluid heat pump system according to any one of claims 4 to 8, characterized in that: The heat pump system further comprises a liquid reservoir (4) and a defrost branch (14), wherein the liquid reservoir (4) is provided with a refrigerant inlet (41) and a first refrigerant outlet (42), wherein 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 reservoir (4) is further provided with a second refrigerant outlet (43), and both ends of the defrost branch (14) are respectively connected to the second refrigerant outlet (43) and the inlet of the evaporator (9), and the defrost branch (14) is provided with an electrically controlled valve (5) with an adjustable opening; The control method further includes: In response to the defrost mode, the electrically controlled valve (5) is controlled to open, and the second throttling device (8), the first control valve (7) and the second control valve (10) are controlled to close.

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