Heat pump system and air conditioner

Abstract

The present application relates to the technical field of refrigeration equipment, and provides a heat pump system and an air conditioner. The heat pump system comprises an outdoor heat exchange unit; the outdoor heat exchange unit has a first heat exchange flow channel and a second heat exchange flow channel; the diameter of the first heat exchange flow channel is larger than the diameter of the second heat exchange flow channel; the reversing unit has a first working state and a second working state; in the first working state, the refrigerant discharged by the compressor returns to the compressor after sequentially passing through the indoor heat exchange unit, the first throttling unit, the first heat exchange flow channel and the second heat exchange flow channel; in the second working state, the refrigerant discharged by the compressor returns to the compressor after sequentially passing through the second heat exchange flow channel, the first heat exchange flow channel, the first throttling unit and the indoor heat exchange unit; and the refrigerant is a non-azeotropic mixed refrigerant. The present application solves the problem that the existing conventional heat pump system is prone to uneven gas-liquid two-phase distribution and poor refrigerant backflow when the refrigerant is a non-azeotropic mixed refrigerant.

Description

Technical Field

[0001] This invention relates to the field of refrigeration equipment technology, and more particularly to a heat pump system and an air conditioner. Background Technology

[0002] While the new generation of low-GWP environmentally friendly refrigerants, represented by R454C, have significant environmental advantages in mitigating global warming, their inherent temperature glide characteristics and high liquid viscosity pose major challenges to the design and operation of existing heat pump systems.

[0003] Traditional heat pump systems are mostly designed for azeotropic, near-azeotropic, or pure refrigerant configurations. Their heat exchanger structures and piping layouts are ill-suited to the continuous dynamic temperature changes during the phase change of non-azeotropic refrigerants like R454C. Due to insufficient property matching, the gas-liquid two-phase distribution during evaporation or condensation is severely uneven, easily inducing stratification and flow deviation within the piping. The higher liquid viscosity significantly increases flow resistance, leading to poor refrigerant return, increased risk of localized liquid accumulation, and further weakening the lubricating oil's return efficiency. These problems, combined, not only drastically reduce the effective utilization rate of the heat exchange area but also deteriorate the overall system heat exchange performance, potentially threatening the long-term reliability of the compressor.

[0004] Therefore, there is an urgent need to develop a new type of heat pump system that can be specifically adapted to such non-azeotropic refrigerants with significant temperature slip and high liquid viscosity. Summary of the Invention

[0005] This invention provides a heat pump system and an air conditioner to solve the problem that existing traditional heat pump systems are prone to uneven gas-liquid two-phase distribution and poor refrigerant reflux when the refrigerant is a non-azeotropic mixture with significant temperature glide and high liquid viscosity.

[0006] This invention provides a heat pump system, comprising: Compressor and indoor heat exchange unit; An outdoor heat exchange unit has a first heat exchange channel and a second heat exchange channel; the diameter of the first heat exchange channel is larger than the diameter of the second heat exchange channel. The first throttling unit is disposed between the first heat exchange channel and the indoor heat exchange unit; The commutation unit has a first working state and a second working state; In the first operating state, the refrigerant discharged by the compressor passes sequentially through the indoor heat exchange unit, the first throttling unit, the first heat exchange channel, and the second heat exchange channel before returning to the compressor; In the second operating state, the refrigerant discharged by the compressor passes sequentially through the second heat exchange channel, the first heat exchange channel, the first throttling unit, and the indoor heat exchange unit before returning to the compressor; The refrigerant is a non-azeotropic mixture.

[0007] According to the heat pump system provided by the present invention, the temperature glide of the non-azeotropic refrigerant mixture is greater than 4K, and the liquid phase viscosity of the non-azeotropic refrigerant mixture increases as the temperature decreases.

[0008] According to the heat pump system provided by the present invention, the non-azeotropic refrigerant mixture includes R454C refrigerant or R455A refrigerant.

[0009] According to the heat pump system provided by the present invention, the outdoor heat exchange unit includes: The first heat exchanger has the first heat exchange channel inside; one end of the first heat exchanger is connected to the first throttling unit. The second heat exchanger has a second heat exchange channel inside. One end of the second heat exchanger is connected to the other end of the first heat exchanger, and the other end of the second heat exchanger is connected to the compressor.

[0010] According to the heat pump system provided by the present invention, the outdoor heat exchange unit includes: The first heat exchange tube assembly has the first heat exchange flow channel inside, and one end of the first heat exchange tube assembly is connected to the first throttling unit. The second heat exchange tube assembly has a second heat exchange flow channel inside; one end of the second heat exchange tube assembly is connected to the other end of the first heat exchange tube assembly, and the other end of the second heat exchange tube assembly is connected to the compressor; the second heat exchange tube assembly is arranged at the air outlet of the outdoor fan, and the first heat exchange tube assembly is arranged on the other side of the second heat exchange tube assembly away from the outdoor fan.

[0011] According to the heat pump system provided by the present invention, the outdoor heat exchange unit further includes: Multiple first heat dissipation fins are spaced apart on the first heat exchange tube group; Multiple second heat dissipation fins are spaced apart on the second heat exchange tube assembly; The fin spacing of the second heat dissipation fin is smaller than that of the first heat dissipation fin.

[0012] According to the heat pump system provided by the present invention, the fin spacing of the first heat dissipation fin is 1.6~2.0mm, and the fin spacing of the second heat dissipation fin is 1.3~1.6mm.

[0013] According to the heat pump system provided by the present invention, the diameter of the first heat exchange channel is 8.0~10mm, and the diameter of the second heat exchange channel is 5.0~7.0mm.

[0014] The heat pump system provided by the present invention further includes: A liquid reservoir, wherein the liquid phase chamber of the liquid reservoir is directly connected to the first heat exchange channel; The second throttling unit has its two ends connected to the liquid phase cavity and the second heat exchange channel, respectively. In the first operating state, the refrigerant discharged from the compressor passes sequentially through the indoor heat exchange unit, the first throttling unit, and the first heat exchange channel before entering the liquid phase chamber. The refrigerant discharged from the liquid phase chamber passes sequentially through the second throttling unit and the second heat exchange channel before returning to the compressor. In the second operating state, the refrigerant discharged from the compressor passes through the second heat exchange channel and the second throttling unit in sequence before entering the liquid phase chamber. The refrigerant discharged from the liquid phase chamber passes through the first heat exchange channel, the first throttling unit, and the indoor heat exchange unit in sequence before returning to the compressor.

[0015] According to the heat pump system provided by the present invention, the liquid reservoir further includes a gas phase chamber located above the liquid phase chamber, and a control valve is provided at the outlet of the gas phase chamber, the control valve having an open state and a closed state.

[0016] According to the heat pump system provided by the present invention, the reversing unit is a four-way reversing valve; the D port of the four-way reversing valve is connected to the exhaust port of the compressor, the E port of the four-way reversing valve is connected to the indoor heat exchange unit, the C port of the four-way reversing valve is connected to the second heat exchange channel, and the S port of the four-way reversing valve is connected to the suction port of the compressor.

[0017] A second aspect of the present invention provides an air conditioner comprising the heat pump system described in any of the preceding claims.

[0018] The heat pump system provided by this invention, through the switching of its first and second operating states by a reversing unit, can adjust the order in which the refrigerant flows through two heat exchange channels according to heating or cooling needs. This allows the first and second heat exchange channels, with different diameters, to be matched to specific requirements under different operating conditions. By actively adjusting the flow path according to operating conditions, the system can match the differentiated requirements of non-azeotropic refrigerant mixtures for flow resistance and heat exchange intensity at different phase change stages. This solves the problems of pipe accumulation and uneven heat exchange caused by traditional fixed single flow path designs, and overcomes the technical challenge of its inability to simultaneously achieve efficient operation in both heating and cooling modes. Attached Figure Description

[0019] To more clearly illustrate the technical solutions in this invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are some embodiments of this invention. For those skilled in the art, other drawings can be obtained from these drawings without creative effort.

[0020] Figure 1 This is one of the structural schematic diagrams of the heat pump system provided by the present invention.

[0021] Figure 2 yes Figure 1 A schematic diagram of the refrigerant flow path of the provided heat pump system under heating conditions.

[0022] Figure 3 yes Figure 1 A schematic diagram of the refrigerant flow path of the provided heat pump system under refrigeration conditions.

[0023] Figure 4 This is the second schematic diagram of the heat pump system provided by the present invention.

[0024] Figure label: 100. Compressor; 200. Indoor heat exchange unit; 300. Outdoor heat exchange unit; 311. First heat exchanger; 312. Second heat exchanger; 321. First heat exchange tube assembly; 322. Second heat exchange tube assembly; 400. First throttling unit; 500. Reversing unit; 600. Liquid receiver; 601. Gas phase chamber; 602. Liquid phase chamber; 700. Second throttling unit; 800. Control valve. Detailed Implementation

[0025] Embodiments of the present invention are described in detail below. Examples of these embodiments are shown in the accompanying drawings, wherein the same or similar reference numerals denote the same or similar elements or elements having the same or similar functions throughout. The embodiments described below with reference to the accompanying drawings are exemplary and are only used to explain the present invention, and should not be construed as limiting the present invention.

[0026] The specific terms used in this specification are for illustrative purposes only and are not intended to limit the illustrated embodiments. For example, expressions such as "same" and "identical" not only indicate a strictly identical state, but also indicate a state with tolerances or differences in the degree of functionality. For example, expressions indicating relative or absolute arrangement such as "in a certain direction," "along a certain direction," "side by side," "perpendicular," "centered on," "concentric," or "coaxial" not only strictly indicate such an arrangement, but also indicate a state of relative displacement by tolerances or angles or distances with the same degree of functionality.

[0027] In the description of this invention, it should be understood that the terms "center," "longitudinal," "lateral," "length," "width," "thickness," "upper," "lower," "front," "rear," "left," "right," "vertical," "horizontal," "top," "bottom," "inner," "outer," "clockwise," "counterclockwise," "axial," "radial," and "circumferential" indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings. They are used only for the convenience of describing this invention and simplifying the description, and are not intended to indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation. Therefore, they should not be construed as limitations on this invention.

[0028] Furthermore, features specified as "first" or "second" may explicitly or implicitly include one or more of those features. In the description of the embodiments of this application, "multiple" means two or more, unless otherwise explicitly specified. In the description of the embodiments of this application, the term "and / or" is merely a description of the relationship between related objects, indicating that three relationships can exist. For example, B1 and / or B2 can represent: B1 existing alone, B1 and B2 existing simultaneously, and B2 existing alone. Additionally, the character " / " in this document generally indicates that the preceding and following related objects have an "or" relationship.

[0029] In the description of this invention, it should be noted that, unless otherwise explicitly specified and limited, the terms "installation," "connection," and "linking" should be interpreted broadly. For example, they can refer to a fixed connection, a detachable connection, or an integral connection; they can refer to a mechanical connection or an electrical connection; they can refer to a direct connection or an indirect connection through an intermediate medium; and they can refer to the internal connection of two components. Those skilled in the art can understand the specific meaning of the above terms in this invention based on the specific circumstances.

[0030] Existing air-source heat pump systems, when used with the environmentally friendly refrigerant R454C (which has a low Global Warming Potential (GWP) in Europe), often encounter problems due to the refrigerant's large temperature glide, high liquid viscosity, and weak flammability (A2L class). This leads to incompatibility between traditional heat pump systems and their flow path structures and heat exchanger parameter configurations. Consequently, problems such as uneven gas-liquid phase distribution, liquid accumulation in pipes, stratification and flow deviation, and decreased heat exchange efficiency may occur during operation. It should be noted that A2L is a classification code for refrigerant safety levels in international standards (ASHRAE 34) and European standard (EN 378).

[0031] Meanwhile, traditional outdoor heat exchangers use a single flow path and uniform specifications, resulting in a large evaporation pressure drop under heating conditions, making it difficult to completely evaporate residual liquid and leading to a low COP for low-temperature heating. Under cooling conditions, the condensation flow path length is fixed, resulting in insufficient subcooling and limited cooling capacity. Therefore, the energy efficiency levels in both heating and cooling modes cannot be effectively balanced at the same time.

[0032] Furthermore, the outdoor heat exchanger is divided into an inner row and a windward row along the airflow direction, with significant differences between the two sides in terms of air velocity, wind resistance, refrigerant phase, and heat exchange load. Existing heat exchangers use a uniform tube diameter and fin spacing structure without adapting to these differences, resulting in structural defects such as increased pressure drop in the high liquid content zone, insufficient heat exchange capacity on the windward side, severe frost and blockage on the inner row, and increased fan power consumption.

[0033] To address at least one of the aforementioned problems, the present invention provides a novel heat pump system and air conditioner.

[0034] The following is combined Figures 1 to 4 This invention describes the heat pump system provided by the present invention.

[0035] like Figure 1 As shown, a specific embodiment of the first aspect of the present invention provides a heat pump system, which includes a compressor 100, an indoor heat exchange unit 200, an outdoor heat exchange unit 300, a first throttling unit 400, and a reversing unit 500. The refrigerant in this embodiment is a non-azeotropic mixture with a temperature glide greater than 4K, and the liquid phase viscosity of this refrigerant increases as the temperature decreases (e.g., using R454C or R455A refrigerant). This results in significant changes in component concentration and pressure drop sensitivity during the phase change process, creating specific requirements for differentiated design of the heat exchange channels.

[0036] The outdoor heat exchange unit 300 has a first heat exchange channel and a second heat exchange channel; the diameter of the first heat exchange channel is larger than the diameter of the second heat exchange channel. Thus, when the refrigerant flows through different heat exchange channels, the larger diameter channel can reduce the flow resistance and pressure drop of refrigerants with high liquid content and high viscosity, while the smaller diameter channel can increase the flow rate of refrigerants with low liquid content or gaseous state, thereby enhancing the convective heat transfer inside the pipe. The combination of the two can adapt to the phase difference between the front and rear sections of non-azeotropic mixed refrigerants during the evaporation or condensation process.

[0037] The first throttling unit 400 is located between the first heat exchange channel and the indoor heat exchange unit 200. It is used to throttle and reduce the pressure of the refrigerant and can adjust the flow rate by switching the flow path under different operating conditions.

[0038] The commutation unit 500 has a first operating state and a second operating state. By switching the operating state, the commutation unit 500 can change the flow path of the refrigerant between the components.

[0039] In the first operating state, the refrigerant discharged from the compressor 100 passes sequentially through the indoor heat exchange unit 200, the first throttling unit 400, the first heat exchange channel, and the second heat exchange channel before returning to the compressor 100. That is, the refrigerant first flows through the larger-diameter first heat exchange channel and then through the smaller-diameter second heat exchange channel. This suppresses the pressure drop of the high-liquid-content refrigerant in the initial stage of evaporation and enhances the heat exchange of the low-liquid-content refrigerant in the later stage, promoting the complete evaporation of any remaining liquid refrigerant and ensuring the dryness of the refrigerant returning to the compressor 100.

[0040] In the second operating state, the refrigerant discharged from the compressor 100 sequentially passes through the second heat exchange channel, the first heat exchange channel, the first throttling unit 400, and the indoor heat exchange unit 200 before returning to the compressor 100. That is, the refrigerant first flows through the smaller-diameter second heat exchange channel, where the higher flow velocity enhances the convective heat transfer efficiency of the gaseous cooling section. Then, the partially or completely condensed refrigerant enters the larger-diameter first heat exchange channel, where the flow velocity decreases, providing sufficient residence time for the liquid refrigerant to subcool, thereby increasing the subcooling degree of the refrigerant before entering the first throttling unit 400.

[0041] In this embodiment, by switching the reversing unit 500 between its first operating state and second operating state, the heat pump system can adjust the order in which the refrigerant flows through the two heat exchange channels according to the heating or cooling demand, so that the first heat exchange channel and the second heat exchange channel with different diameters can respectively match the specific needs under different operating conditions.

[0042] In heating mode (first operating state), the refrigerant first flows through a large-diameter first heat exchange channel, and then through a small-diameter second heat exchange channel. The large-diameter first heat exchange channel provides lower flow resistance for the two-phase flow in the initial stage of evaporation, solving the problem of large evaporation pressure loss in traditional heat pump systems during heating. The small-diameter second heat exchange channel increases the refrigerant flow rate at the end of evaporation, enhances heat exchange, and ensures complete evaporation of the refrigerant, thereby improving the heating COP at low temperatures and solving the problem of incomplete residual liquid evaporation in traditional heat pump systems during heating.

[0043] In refrigeration mode (second operating state), the refrigerant flow is reversed, first passing through the small-diameter second heat exchange channel and then through the large-diameter first heat exchange channel. The high-temperature gaseous refrigerant first enters the small-diameter channel, where the higher flow velocity enhances convective heat transfer during condensation, solving the problem of low heat exchange efficiency in traditional systems during refrigeration. The condensed liquid refrigerant then enters the large-diameter channel, where the flow velocity decreases, allowing for a longer residence time and sufficient subcooling, thereby improving the system's refrigeration capacity and solving the problem of insufficient subcooling in traditional systems during refrigeration.

[0044] This embodiment actively adjusts the flow path according to the operating conditions, enabling the system to match the different requirements of non-azeotropic mixed refrigerants for flow resistance and heat transfer intensity at different phase change stages. It solves the problems of liquid accumulation and uneven heat transfer caused by traditional fixed single flow path design, and overcomes the technical difficulty of its inability to achieve efficient operation in both heating and cooling modes.

[0045] Optionally, the first throttling unit 400 includes an electronic expansion valve or a thermostatic expansion valve. Preferably, the first throttling unit 400 is an electronic expansion valve.

[0046] Optionally, the indoor heat exchange unit 200 may include a plate heat exchanger.

[0047] In some embodiments of the present invention, the reversing unit 500 is a four-way reversing valve; the D port of the four-way reversing valve is connected to the discharge port of the compressor 100, and the D port serves as a fixed inlet for the high-pressure refrigerant in the system, used to receive the high-temperature, high-pressure refrigerant gas discharged from the compressor 100. The S port of the four-way reversing valve is connected to the suction port of the compressor 100, and the S port serves as a fixed return port for the low-pressure refrigerant in the system, so that the evaporated gaseous refrigerant returns to the compressor 100, thereby forming a closed loop of refrigerant circulation. The E port of the four-way reversing valve is connected to the indoor heat exchange unit 200, so that the refrigerant flows into the indoor heat exchange unit 200 through the E port in heating mode for condensation and heat release. The C port of the four-way reversing valve is connected to the second heat exchange channel, so that the refrigerant flows into the second heat exchange channel through the C port in cooling mode for condensation.

[0048] In the first operating state, the slide valve inside the four-way reversing valve moves to the first position, connecting the D port to the E port and simultaneously connecting the C port to the S port. At this time, the refrigerant discharged from the compressor 100 flows sequentially through the indoor heat exchange unit 200, the first throttling unit 400, the first heat exchange channel, and the second heat exchange channel before returning to the compressor 100. In the first operating state, the refrigerant first flows through the larger diameter first heat exchange channel and then through the smaller diameter second heat exchange channel, so that the high liquid content region in the front section of evaporation obtains a lower flow resistance, and the low liquid content region in the rear section of evaporation obtains a higher flow velocity to promote the evaporation of residual liquid, thereby matching the different requirements of the front and rear sections of the non-azeotropic refrigerant mixture for the flow cross-section during the evaporation process.

[0049] In the second operating state, the slide valve inside the four-way reversing valve moves to the second position, connecting the D port to the C port and simultaneously connecting the E port to the S port. At this time, the refrigerant discharged from the compressor 100 flows sequentially through the second heat exchange channel, the first heat exchange channel, the first throttling unit 400, and the indoor heat exchange unit 200 before returning to the compressor 100. In the second operating state, the refrigerant first flows through the smaller-diameter second heat exchange channel and then through the larger-diameter first heat exchange channel, so that the high-temperature gaseous region in the pre-condensation stage obtains a higher flow velocity to enhance condensation heat exchange, and the liquid subcooled region in the post-condensation stage obtains a lower flow resistance to reduce pressure loss, thereby matching the different requirements of the gaseous section and the subcooled section for the flow cross-section during the condensation process of the non-azeotropic refrigerant mixture.

[0050] In this embodiment, by switching between the first and second working states using a four-way reversing valve, the flow sequence of the refrigerant in the outdoor heat exchange unit 300 is changed, so that the first and second heat exchange channels respectively undertake the front or rear functions adapted to the corresponding modes in the two working states, thereby obtaining the refrigerant flow sequence corresponding to the mode in both heating and cooling modes.

[0051] In some embodiments of the present invention, the heat pump system further includes a liquid receiver 600 and a second throttling unit 700. The liquid phase chamber 602 of the liquid receiver 600 is directly connected to the first heat exchange channel, so that after the gas-liquid mixed refrigerant flowing out of the first heat exchange channel enters the liquid phase chamber 602, the liquid portion accumulates at the bottom of the liquid phase chamber 602 under gravity, achieving gas-liquid separation. The two ends of the second throttling unit 700 are respectively connected to the liquid phase chamber 602 and the second heat exchange channel, and are used to throttle and reduce the pressure of the liquid refrigerant discharged from the liquid phase chamber 602, so that it meets the pressure conditions required for evaporation in the second heat exchange channel.

[0052] In the first working state, the refrigerant discharged from the compressor 100 passes through the indoor heat exchange unit 200, the first throttling unit 400, and the first heat exchange channel in sequence before entering the liquid phase chamber 602. At this time, the refrigerant flowing out from the first heat exchange channel is in a gas-liquid mixed state after partial evaporation. After the gas-liquid mixed refrigerant enters the liquid phase chamber 602, the gas phase part and the liquid phase part are separated. The gas phase part can return to the compressor 100 through the gas phase outlet of the liquid receiver 600, and the liquid phase part accumulates at the bottom of the liquid phase chamber 602. The refrigerant discharged from the liquid phase chamber 602 passes through the second throttling unit 700 and the second heat exchange channel in sequence before returning to the compressor 100, forming a two-stage evaporation system. After the liquid refrigerant is depressurized by the second throttling unit 700, it enters the second heat exchange channel for secondary evaporation, thereby controlling the state and flow rate of the refrigerant entering the second heat exchange channel. This aims to ensure that the remaining liquid refrigerant is fully evaporated in the second heat exchange channel, so that the residual liquid refrigerant is completely dried in the second heat exchange channel, thereby improving the dryness of the refrigerant returning to the compressor 100.

[0053] In the second operating state, the second throttling unit 700 remains fully open, without throttling the refrigerant. The refrigerant discharged from the compressor 100 passes sequentially through the second heat exchange channel and the second throttling unit 700 before entering the liquid phase chamber 602. To ensure subsequent subcooling, the second throttling unit 700 should be fully open at this time, without producing a significant throttling and pressure reduction effect, mainly serving as a connecting channel. After the refrigerant condenses in the second heat exchange channel, it enters the liquid phase chamber 602 in a high-pressure liquid state or a gas-liquid mixture. The refrigerant discharged from the liquid phase chamber 602 passes sequentially through the first heat exchange channel, the first throttling unit 400, and the indoor heat exchange unit 200 before returning to the compressor 100. The high-pressure liquid refrigerant in the liquid phase chamber 602 flows into the first heat exchange channel for sufficient subcooling and heat release, ensuring that the refrigerant obtains sufficient subcooling before entering the first throttling unit 400, thereby improving cooling performance.

[0054] In this embodiment, by using the liquid receiver 600 to separate the gas-liquid mixture of refrigerant flowing out of the first heat exchange channel in the first working state, and by allowing the liquid phase to enter the second heat exchange channel through the second throttling unit 700 for secondary evaporation, the refrigerant can undergo gas-liquid separation and secondary throttling before flowing through the second heat exchange channel, thereby controlling the flow rate and state of the refrigerant entering the second heat exchange channel. This allows the first and second heat exchange channels to respectively perform the functions of partial evaporation and secondary evaporation during the evaporation process.

[0055] By allowing the refrigerant to flow sequentially through the second heat exchange channel and the second throttling unit 700 into the liquid receiver 600 in the second operating state, and then out through the first heat exchange channel, the condensed refrigerant can be throttled and depressurized before entering the first heat exchange channel for subcooling heat exchange. This allows the refrigerant to be further cooled by the first heat exchange channel before entering the first throttling unit 400.

[0056] By coordinating the liquid receiver 600 and the second throttling unit 700 in two operating states, the refrigerant can undergo gas-liquid separation or pressure reduction buffering in the liquid receiver 600 during the process of flowing between the first heat exchange channel and the second heat exchange channel, depending on the current operating state. This achieves deep subcooling in the two-stage evaporation process and the condensation process in the two operating states, respectively.

[0057] Furthermore, the receiver 600 also includes a gas phase chamber 601, located above the liquid phase chamber 602. This allows the gas-liquid mixture of refrigerant entering the receiver 600 to naturally stratify under gravity, with the gas phase rising and accumulating in the gas phase chamber 601, and the liquid phase settling downwards in the liquid phase chamber 602. A control valve 800 is installed at the outlet of the gas phase chamber 601, and the control valve 800 has open and closed states. The control valve 800 switches between open and closed states by the movement of its internal valve core. When the valve core moves to a position that connects the outlet of the gas phase chamber 601 to the downstream pipeline, the control valve 800 is in the open state; when the valve core moves to a position that blocks the outlet of the gas phase chamber 601, the control valve 800 is in the closed state.

[0058] In the first operating state, control valve 800 is open, and the gaseous refrigerant in the gas phase chamber 601 is discharged through control valve 800, merging with the gaseous refrigerant discharged from the second heat exchange channel, and finally returning to compressor 100. The gas-liquid mixed refrigerant flowing out of the first heat exchange channel and into the liquid receiver 600 has its gaseous portion discharged directly through control valve 800 without entering the liquid phase chamber 602, while the liquid portion flows out of the liquid phase chamber 602 and undergoes secondary evaporation through the second throttling unit 700 and the second heat exchange channel. This causes the gaseous refrigerant generated during evaporation to merge with the gaseous refrigerant discharged through control valve 800 before returning to compressor 100, thereby reducing the flow rate of gaseous refrigerant through the second throttling unit 700 and the second heat exchange channel, making the second throttling unit 700 and the second heat exchange channel mainly used for handling liquid refrigerant. At the same time, when control valve 800 is open, it can also balance the internal pressure of the liquid receiver 600.

[0059] In the second operating state, control valve 800 is closed. Refrigerant flowing from the second heat exchange channel and passing through the second throttling unit 700 enters the liquid receiver 600. The gas and liquid phases of the refrigerant undergo gas-liquid separation within the liquid receiver 600. The gas phase remains in the gas phase chamber 601, while the liquid phase flows through the liquid phase chamber 602 into the first heat exchange channel for subcooling. At this time, the outlet of the gas phase chamber 601 is blocked by control valve 800, preventing the refrigerant flowing from the second heat exchange channel from bypassing the outlet of the gas phase chamber 601. All refrigerant must flow through the liquid phase chamber 602 into the first heat exchange channel and then through the first throttling unit 400 into the indoor heat exchange unit 200.

[0060] In this embodiment, by opening the control valve 800 in the first operating state, the gaseous refrigerant in the gas phase chamber 601 can be discharged from the outlet of the gas phase chamber 601 and merge with the refrigerant at the outlet of the second heat exchange channel, thereby reducing the liquid load of the second heat exchange channel. By closing the control valve 800 in the second operating state, the liquid receiver 600 can maintain its gas-liquid separation function in condensation mode, preventing the gaseous refrigerant from being directly discharged from the outlet of the gas phase chamber 601, thereby ensuring that the refrigerant entering the first heat exchange channel remains in a liquid or gas-liquid mixed state for subcooling heat exchange. Through the controllable bidirectional flow path switching of the control valve 800, two-stage evaporation with low pressure loss in heating and extended condensation with high subcooling in cooling are achieved, completely solving the problem of energy efficiency imbalance in heating and cooling conditions.

[0061] Optionally, control valve 800 is a solenoid valve.

[0062] Optionally, the second throttling unit 700 includes an electronic expansion valve or a thermostatic expansion valve. Preferably, the second throttling unit 700 is an electronic expansion valve.

[0063] like Figure 4 As shown, in some embodiments of the present invention, the outdoor heat exchange unit 300 includes a first heat exchanger 311 and a second heat exchanger 312. The first heat exchanger 311 has a first heat exchange flow channel inside; the second heat exchanger 312 has a second heat exchange flow channel inside. That is, the first heat exchanger 311 and the second heat exchanger 312 are two independent heat exchangers, which are connected in series by pipelines to form the outdoor heat exchange unit 300. Thus, the first heat exchange flow channel and the second heat exchange flow channel are physically located in two independent heat exchangers, so that the heat exchange area, number of pipes and flow path layout of the first heat exchanger 311 and the second heat exchanger 312 can be independently set during manufacturing and assembly. Furthermore, intermediate components such as valves or liquid reservoirs 600 can be installed on the connecting pipeline between the two independent heat exchangers as needed.

[0064] To increase heat exchange efficiency, each heat exchanger is equipped with an outdoor fan. Specifically, the first heat exchanger 311 is equipped with a first outdoor fan, and the second heat exchanger 312 is equipped with a second outdoor fan. When the first outdoor fan is working, its impeller rotates around the axis of the first fan to drive air to flow through the first heat exchanger 311. When the second outdoor fan is working, its impeller rotates around the axis of the second fan to drive air to flow through the second heat exchanger 312. The two outdoor fans operate independently, so that the first heat exchanger 311 and the second heat exchanger 312 each obtain their respective windward air volume.

[0065] In this embodiment, by placing the first and second heat exchange channels within two independent heat exchangers, and assigning each heat exchanger a corresponding outdoor fan, the first heat exchanger 311 and the second heat exchanger 312 can each independently receive an airflow supply matching their heat exchange area and channel diameter, avoiding the problem of uneven airflow distribution that occurs when the two heat exchange channels share a single outdoor fan. Simultaneously, the two independent heat exchangers can be arranged in different spatial locations during installation, ensuring that the windward side of each heat exchanger directly faces the outdoor ambient air. This allows both heat exchangers to receive fresh air for heat exchange, preventing one heat exchanger from being located on the leeward side of the other, which could lead to an increase in the inlet air temperature or a decrease in airflow. Furthermore, this ensures that the refrigerant flowing through the first and second heat exchange channels receives corresponding air-side heat exchange conditions.

[0066] In other embodiments of the present invention, the outdoor heat exchange unit 300 includes a first heat exchange tube group 321 and a second heat exchange tube group 322. The first heat exchange tube group 321 and the second heat exchange tube group 322 are integral and together form an integrated outdoor heat exchange unit 300, belonging to two sections of the outdoor heat exchange unit 300. Therefore, only one outdoor fan is required. When the outdoor fan is running, its impeller rotates around the fan axis to drive air to flow sequentially through the two tube groups. The first heat exchange tube group 321 has a first heat exchange flow channel inside. The second heat exchange tube group 322 has a second heat exchange flow channel inside. The second heat exchange tube group 322 is arranged at the air outlet of the outdoor fan, and the first heat exchange tube group 321 is located on the other side of the second heat exchange tube group 322 away from the outdoor fan. That is, along the airflow direction, the outdoor fan, the second heat exchange tube group 322, and the first heat exchange tube group 321 are arranged in sequence. When the outdoor fan is running, the driven air first flows through the second heat exchange tube group 322 and then through the first heat exchange tube group 321, so that the second heat exchange tube group 322 is located on the windward side and the first heat exchange tube group 321 is located on the leeward side. This utilizes the sufficient airflow at the location of the second heat exchange tube group 322 to enhance its heat exchange, and at the same time utilizes the reduced wind speed at the location of the first heat exchange tube group 321 to reduce its wind resistance sensitivity.

[0067] In other words, this embodiment can divide the heat exchange tubes inside the outdoor heat exchanger into two parts without changing the existing outdoor heat exchanger structure. The heat exchange tubes with a large diameter are designated as the first heat exchange tube group 321, and the heat exchange tubes with a small diameter are designated as the second heat exchange tube group 322. That is, within the same outdoor heat exchanger core, the first heat exchange tube group 321 and the second heat exchange tube group 322 share the same set of end plates and frames. By replacing the original heat exchange tubes with uniform diameters with two sets of heat exchange tubes with different diameters, the modification process does not require additional heat exchanger shell and installation space, thereby reducing production costs and overall machine size.

[0068] By placing the first heat exchange tube assembly 321 on the side of the second heat exchange tube assembly 322 away from the outdoor fan, and arranging them one after the other along the airflow direction, the outdoor fan drives the air to pass through the second heat exchange tube assembly 322 first and then through the first heat exchange tube assembly 321. The second heat exchange tube assembly 322 directly receives fresh air on the windward side to enhance heat exchange in the small-diameter flow channel, while the first heat exchange tube assembly 321, on the leeward side, utilizes the air after passing through the second heat exchange tube assembly 322 to cooperate with the large-diameter flow channel for heat exchange. Thus, with only one outdoor fan, both heat exchange tube assemblies obtain airside conditions adapted to their respective flow channel diameters. Furthermore, since both tube assemblies share the same outdoor fan, compared to a scheme where each tube assembly has its own independent fan, the number of fans and the corresponding installation space can be reduced, thus lowering the overall footprint and manufacturing cost.

[0069] Furthermore, the outdoor heat exchange unit 300 also includes a plurality of first heat dissipation fins and a plurality of second heat dissipation fins. The plurality of first heat dissipation fins are axially spaced along the first heat exchange tube assembly 321, and contact the outer wall surface of the first heat exchange tube assembly 321 to expand the heat exchange area between the first heat exchange tube assembly 321 and the air. The plurality of second heat dissipation fins are axially spaced along the second heat exchange tube assembly 322, and contact the outer wall surface of the second heat exchange tube assembly 322 to expand the heat exchange area between the second heat exchange tube assembly 322 and the air.

[0070] The fin spacing of the second heat exchange fin is smaller than that of the first heat exchange fin, meaning that the number of second heat exchange fins per unit length is greater than that of first heat exchange fins. This results in a higher fin density in the corresponding area of ​​the second heat exchange tube group 322 compared to the corresponding area of ​​the first heat exchange tube group 321. Consequently, the second heat exchange tube group 322 can obtain a larger total heat exchange area on the windward side by utilizing higher wind speeds and denser fins. At the same time, the first heat exchange tube group 321 can reduce airflow resistance on the leeward side by utilizing a larger fin spacing, thereby reducing the potential airflow attenuation and frost blockage caused by lower wind speeds on the leeward side.

[0071] When the outdoor fan is running, it drives air to flow sequentially through the second heat exchange tube assembly 322 and the first heat exchange tube assembly 321. The air first passes through the second heat dissipation fins with a smaller fin spacing, and then through the first heat dissipation fins with a larger fin spacing. Due to the smaller fin spacing of the second heat dissipation fins, the air achieves higher heat exchange efficiency when flowing through the second heat exchange tube assembly 322. Due to the larger fin spacing of the first heat dissipation fins, the air achieves lower flow resistance when flowing through the first heat exchange tube assembly 321. This allows the two heat exchange tube assemblies to achieve heat exchange performance and wind resistance characteristics that match their fin spacing under their respective airflow velocity conditions.

[0072] By setting the fin spacing of the second heat dissipation fins to be smaller than that of the first heat dissipation fins, the second heat exchange tube group 322 on the windward side uses denser fins to enhance heat exchange under sufficient airflow conditions, while the first heat exchange tube group 321 on the leeward side uses sparser fins to reduce wind resistance and delay frost formation under reduced wind speed conditions. Thus, the outdoor heat exchange unit 300, equipped with only one outdoor fan, obtains fin densities on both the windward and leeward sides that are adapted to the airflow conditions at that location.

[0073] Furthermore, the fin spacing of the first heat dissipation fin is 1.6~2.0mm, and the fin spacing of the second heat dissipation fin is 1.3~1.6mm. That is, the fin spacing of the first heat dissipation fin is generally larger than that of the second heat dissipation fin in the numerical range, thereby making the fin density at the location of the first heat exchange tube group 321 lower than the fin density at the location of the second heat exchange tube group 322.

[0074] In this embodiment, by setting the fin spacing of the first heat dissipation fin to 1.6~2.0mm and the fin spacing of the second heat dissipation fin to 1.3~1.6mm, the windward and leeward sides obtain fin densities corresponding to the airflow conditions at that location. The windward side uses higher wind speeds in conjunction with denser fins to enhance heat exchange, while the leeward side uses sparser fins to reduce wind resistance and delay frost formation. Thus, with only one outdoor fan, the outdoor heat exchange unit 300 allows each of the two heat exchange tube groups to obtain fin spacing parameters that match the location of the wind field.

[0075] In some embodiments of the present invention, the diameter of the first heat exchange channel is 8.0~10 mm, and the diameter of the second heat exchange channel is 5.0~7.0 mm. The diameter of the first heat exchange channel is generally larger than the diameter of the second heat exchange channel in the numerical range, and there is a clear difference range between the two diameter ranges.

[0076] The first heat exchange channel has a diameter range of 8.0~10mm. This diameter range ensures that, in the first operating state (heating mode), when the high-liquid-content two-phase refrigerant flowing from the first throttling unit 400 enters the first heat exchange channel, this diameter range provides a low flow resistance channel for the high-liquid-phase viscosity refrigerant, reducing pressure loss as the refrigerant flows through the first heat exchange channel. Simultaneously, in the second operating state (cooling mode), the liquid or gas-liquid mixed refrigerant, after condensation in the second heat exchange channel, flows through the first heat exchange channel at a lower flow rate within this diameter range. This allows the refrigerant sufficient residence time within the first heat exchange channel for subcooling and heat release, thereby achieving subcooling.

[0077] The second heat exchange channel has a diameter range of 5.0~7.0 mm. This allows the liquid refrigerant, after being depressurized by the second throttling unit 700 and entering the second heat exchange channel in the first operating state (heating mode), to achieve a higher flow velocity and turbulence intensity within this diameter range. This enhances heat exchange between the refrigerant and the pipe wall, promoting the complete evaporation of any remaining liquid refrigerant. Simultaneously, in the second operating state (cooling mode), when the high-temperature gaseous refrigerant discharged from the compressor 100 enters the second heat exchange channel, this diameter range increases the flow velocity of the gaseous refrigerant, thereby enhancing its convective heat transfer coefficient and causing rapid condensation of the high-temperature gaseous refrigerant within the second heat exchange channel.

[0078] In this embodiment, by setting the diameter of the first heat exchange channel to 8.0~10mm and the diameter of the second heat exchange channel to 5.0~7.0mm, the two heat exchange channels form differentiated flow cross sections within their respective diameter ranges. When the refrigerant flows through the two heat exchange channels under two operating conditions, the larger diameter channel provides lower flow resistance under low flow rate or liquid conditions, while the smaller diameter channel provides a higher heat transfer coefficient under high flow rate or gaseous conditions. Thus, the diameter ranges of the two heat exchange channels are adapted to their respective functions under the two operating conditions.

[0079] In some embodiments of the present invention, the diameter of the first heat exchange channel is 8.0~10mm, the diameter of the second heat exchange channel is 5.0~7.0mm, and the fin spacing of the first heat dissipation fin is 1.6~2.0mm, and the fin spacing of the second heat dissipation fin is 1.3~1.6mm. That is, in this embodiment, the two sections of the outdoor heat exchange unit 300 adopt an asymmetrical structure in both the pipe diameter and fin spacing dimensions. The first section (corresponding to the first heat exchange channel) adopts a combination of large pipe diameter and large fin spacing, and the second section (corresponding to the second heat exchange channel) adopts a combination of small pipe diameter and small fin spacing.

[0080] The second heat exchange tube group 322 is located at the air outlet (windward side) of the outdoor fan. This position directly receives the air discharged from the outdoor fan, with sufficient air volume and high air velocity. The combination of small tube diameter and small fin spacing can take advantage of the high air velocity conditions and increase the refrigerant flow rate through the small diameter flow channel to enhance the heat transfer coefficient inside the tube. At the same time, the dense fins increase the heat transfer area on the air side, so that the air volume advantage on the windward side can be fully utilized at both the inside and outside of the tube.

[0081] The first heat exchange tube group 321 is located on the other side (leeward side) of the second heat exchange tube group 322 away from the outdoor fan. The air velocity at this location has already decreased after passing through the second heat exchange tube group 322, and this location is prone to frost formation under low temperature heating conditions. The combination of large tube diameter and large fin spacing can reduce the flow resistance of high liquid refrigerant inside the tube through the large diameter flow channel, and reduce the air flow resistance outside the tube through the large fin spacing and delay the time for the frost layer to cross the gap between adjacent fins.

[0082] The first heat exchange channel uses a diameter of 8.0~10mm and a fin spacing of 1.6~2.0mm. When the refrigerant flows through the high liquid content front section or the liquid subcooled section of this channel, the large pipe diameter reduces the flow pressure drop inside the pipe, while the large fin spacing reduces the air resistance on the leeward side, keeping the flow resistance inside and outside the pipe at a low level. The second heat exchange channel uses a diameter of 5.0~7.0mm and a fin spacing of 1.3~1.6mm. When the refrigerant flows through the high temperature gaseous section or the low liquid content rear section of this channel, the small pipe diameter increases the flow velocity and turbulence intensity inside the pipe, while the small fin spacing provides more heat exchange area under high wind speed conditions on the windward side, thus enhancing heat exchange both inside and outside the pipe.

[0083] In this embodiment, by combining differences in pipe diameter and fin spacing, the outdoor heat exchange unit 300 achieves internal pipe flow cross-sections and external fin densities that match the air velocity conditions and refrigerant phase conditions as it flows through the zone on both the windward and leeward sides. On the windward side, high wind speeds combined with small pipe diameters and small fin spacing enhance heat exchange, while on the leeward side, low wind speeds combined with large pipe diameters and large fin spacing reduce flow resistance and delay frost formation. This ensures that the refrigerant phase change flow requirements inside the pipes are matched to the external airflow conditions in both zones, reducing the dual heat exchange bottlenecks caused by a single pipe diameter or a single fin spacing.

[0084] A specific embodiment of the second aspect of the present invention provides an air conditioner that includes the heat pump system of any of the above embodiments. Therefore, it possesses at least the advantages described above, which will not be repeated here.

[0085] Optionally, the air conditioner also includes an outdoor unit housing, in which the compressor 100, outdoor heat exchange unit 300, liquid receiver 600, reversing unit 500 and first throttling unit 400 are all housed.

[0086] Optionally, the air conditioner also includes an indoor unit housing, on which the indoor heat exchange unit 200 is installed.

[0087] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, and not to limit them; although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some of the technical features; and these modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the spirit and scope of the technical solutions of the embodiments of the present invention.

Claims

1. A heat pump system, characterized in that, include: Compressor (100) and indoor heat exchange unit (200); The outdoor heat exchange unit (300) has a first heat exchange channel and a second heat exchange channel; the diameter of the first heat exchange channel is larger than the diameter of the second heat exchange channel. The first throttling unit (400) is disposed between the first heat exchange channel and the indoor heat exchange unit (200); The commutation unit (500) has a first working state and a second working state; In the first working state, the refrigerant discharged by the compressor (100) passes through the indoor heat exchange unit (200), the first throttling unit (400), the first heat exchange channel, and the second heat exchange channel in sequence before returning to the compressor (100). In the second operating state, the refrigerant discharged by the compressor (100) passes through the second heat exchange channel, the first heat exchange channel, the first throttling unit (400), and the indoor heat exchange unit (200) in sequence before returning to the compressor (100). The refrigerant is a non-azeotropic mixture.

2. The heat pump system according to claim 1, characterized in that, The temperature slip of the non-azeotropic refrigerant mixture is greater than 4K, and the liquid phase viscosity of the non-azeotropic refrigerant mixture increases as the temperature decreases.

3. The heat pump system according to claim 1, characterized in that, The non-azeotropic refrigerant mixture includes R454C refrigerant or R455A refrigerant.

4. The heat pump system according to claim 1, characterized in that, The outdoor heat exchange unit (300) includes: The first heat exchanger (311) has the first heat exchange channel inside; one end of the first heat exchanger (311) is connected to the first throttling unit (400); The second heat exchanger (312) has a second heat exchange channel inside. One end of the second heat exchanger (312) is connected to the other end of the first heat exchanger (311), and the other end of the second heat exchanger (312) is connected to the compressor (100).

5. The heat pump system according to claim 1, characterized in that, The outdoor heat exchange unit (300) includes: The first heat exchange tube assembly (321) has the first heat exchange flow channel inside, and one end of the first heat exchange tube assembly (321) is connected to the first throttling unit (400); The second heat exchange tube assembly (322) has the second heat exchange flow channel inside; one end of the second heat exchange tube assembly (322) is connected to the other end of the first heat exchange tube assembly (321), and the other end of the second heat exchange tube assembly (322) is connected to the compressor (100); the second heat exchange tube assembly (322) is arranged at the air outlet of the outdoor fan, and the first heat exchange tube assembly (321) is arranged on the other side of the second heat exchange tube assembly (322) away from the outdoor fan.

6. The heat pump system according to claim 5, characterized in that, The outdoor heat exchange unit (300) also includes: Multiple first heat dissipation fins are spaced apart on the first heat exchange tube group (321). Multiple second heat dissipation fins are spaced apart on the second heat exchange tube group (322). The fin spacing of the second heat dissipation fin is smaller than that of the first heat dissipation fin.

7. The heat pump system according to claim 6, characterized in that, The fin spacing of the first heat dissipation fin is 1.6~2.0mm, and the fin spacing of the second heat dissipation fin is 1.3~1.6mm.

8. The heat pump system according to claim 1, characterized in that, The diameter of the first heat exchange channel is 8.0~10mm, and the diameter of the second heat exchange channel is 5.0~7.0mm.

9. The heat pump system according to any one of claims 1 to 8, characterized in that, Also includes: The liquid reservoir (600) has a liquid phase chamber (602) that is directly connected to the first heat exchange channel; The second throttling unit (700) has two ends connected to the liquid phase cavity (602) and the second heat exchange channel, respectively. In the first working state, the refrigerant discharged by the compressor (100) passes through the indoor heat exchange unit (200), the first throttling unit (400), and the first heat exchange channel in sequence before entering the liquid phase chamber (602). The refrigerant discharged from the liquid phase chamber (602) passes through the second throttling unit (700) and the second heat exchange channel in sequence before returning to the compressor (100). In the second operating state, the refrigerant discharged from the compressor (100) passes through the second heat exchange channel and the second throttling unit (700) in sequence and then enters the liquid phase chamber (602). The refrigerant discharged from the liquid phase chamber (602) passes through the first heat exchange channel, the first throttling unit (400), and the indoor heat exchange unit (200) in sequence and then returns to the compressor (100).

10. The heat pump system according to claim 9, characterized in that, The reservoir (600) also has a gas phase chamber (601) located above the liquid phase chamber (602). The outlet of the gas phase chamber (601) is provided with a control valve (800), which has an open state and a closed state.

11. The heat pump system according to any one of claims 1 to 8, characterized in that, The reversing unit (500) is a four-way reversing valve; the D port of the four-way reversing valve is connected to the exhaust port of the compressor (100), the E port of the four-way reversing valve is connected to the indoor heat exchange unit (200), the C port of the four-way reversing valve is connected to the second heat exchange channel, and the S port of the four-way reversing valve is connected to the suction port of the compressor (100).

12. An air conditioner, characterized in that, Includes the heat pump system as described in any one of claims 1 to 11.

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