A pump-driven two-phase flow thermal control system and its control method

Abstract

This invention belongs to the field of thermal management technology and discloses a pump-driven two-phase flow thermal control system and its control method. In the pump-driven two-phase flow thermal control system, the compressor is connected to the first interface of a three-way proportional control valve via a gas cooler; the second interface of the three-way proportional control valve is connected to the inlet of a throttling valve via the first heat exchange channel of a homogenizer; the third interface of the three-way proportional control valve is connected to the inlet of the throttling valve; the outlet of the throttling valve is connected to the first inlet of a liquid storage tank; the gas outlet of the liquid storage tank is connected to the compressor; the liquid outlet of the liquid storage tank is sequentially connected to the second inlet of the liquid storage tank via a pump, the second heat exchange channel of the homogenizer, and a cooling module; the cooling module includes one or more cooling branches. This invention can solve the technical problems of insufficient cooling capacity under high heat flux density and the inability to achieve the expected temperature uniformity in existing two-phase flow thermal control systems.

Description

Technical Field

[0001] This invention belongs to the field of thermal management technology, and specifically relates to a pump-driven two-phase flow thermal control system and its control method. Background Technology

[0002] Most electronic devices have limited operating temperatures, and it is necessary to control the temperature of each component to prevent equipment failure. Therefore, the concept of a thermal control system was proposed to achieve temperature control of components with high temperature requirements.

[0003] As the functionality and complexity of electronic devices increase, the integration of components continues to improve, and the demand for heat dissipation also increases. Existing passive thermal control systems can no longer meet the thermal management needs of electronic devices. Furthermore, passive thermal control systems suffer from problems such as the inability to automatically adjust the temperature and insufficient driving force of capillary pump circuits.

[0004] Mechanically pump-driven two-phase flow cooling circuit systems have become a hot topic in the field of cooling technology research due to their superior stability, longer transmission distance, and higher temperature control accuracy; for example, in the aerospace field; research shows that pump-driven two-phase flow thermal control systems are superior to water-cooled systems in terms of cooling performance, system size, and pump power consumption.

[0005] Furthermore, due to the special nature of the heat-generating components in electronic devices and their operating temperature requirements, it is necessary to dissipate heat from multiple heat sources in a timely manner and ensure temperature uniformity. Compared to single-phase water circuit systems, two-phase evaporators offer significantly improved temperature uniformity and heat exchange capacity because the refrigerant temperature remains almost constant.

[0006] However, existing technologies are not perfect in controlling temperature uniformity. Since the temperature and pressure of a two-phase refrigerant are in a one-to-one correspondence, a combination of heat exchanger pressure control and power control is generally used. Current control methods can achieve the desired temperature uniformity (<±1℃) under low loads; however, they are prone to temperature runaway under high loads. Furthermore, due to the integration of electronic components, the required heat flux density for cooling is becoming increasingly demanding, exceeding 10W / cm². 2 Improving the cooling effect of the system under high heat flux density based on the existing system is also an urgent problem to be solved. Summary of the Invention

[0007] The purpose of this invention is to provide a pump-driven two-phase flow thermal control system and its control method to solve one or more of the aforementioned technical problems. The technical solution provided by this invention can solve the technical problems of insufficient cooling capacity under high heat flux density and the inability to achieve the expected temperature uniformity in existing two-phase flow thermal control systems.

[0008] To achieve the above objectives, the present invention adopts the following technical solution:

[0009] The present invention provides a pump-driven two-phase flow thermal control system, comprising: a compressor, a gas cooler, a three-way proportional regulating valve, a temperature equalization regenerator, a throttling valve, a liquid storage tank, a pump, and a cooling module;

[0010] The outlet of the compressor is connected to the first port of the three-way proportional control valve via the gas cooler, the second port of the three-way proportional control valve is connected to the inlet of the throttle valve via the first heat exchange channel of the isothermal regenerator, and the third port of the three-way proportional control valve is connected to the inlet of the throttle valve.

[0011] The outlet of the throttle valve is connected to the first inlet of the liquid storage tank, the gas outlet of the liquid storage tank is connected to the inlet of the compressor, and the liquid outlet of the liquid storage tank is connected to the second inlet of the liquid storage tank via the pump, the second heat exchange channel of the isothermal regenerator, and the cooling module in sequence.

[0012] The cooling module includes one or more cooling branches, each of which includes a heat exchange plate and a regulating valve.

[0013] The present invention provides a control method for a pump-driven two-phase flow thermal control system, comprising the following steps:

[0014] Obtain the parameter state space of the pump-driven two-phase flow thermal control system during operation;

[0015] The parameter state space is input into the predictive control model to obtain the overall working fluid flow prediction value, the inlet dryness prediction value of the cooling module, and the outlet dryness prediction value of each cooling branch as the control target.

[0016] Based on the overall working fluid flow prediction value as the control target, the pump speed is controlled; based on the inlet dryness prediction value of the cooling module as the control target and the outlet dryness prediction value of each cooling branch, the opening degree of the three-way proportional regulating valve and the opening degree of the regulating valve of each cooling branch are controlled.

[0017] A further improvement to the control method of the present invention is that the parameter state space is represented as follows:

[0018]

[0019] In the formula, P3 is the system low-pressure; D Plate,i L represents the characteristic cross-sectional dimension of the heat exchange plate in the i-th cooling branch; Plate,i Let be the characteristic length of the heat exchange plate of the i-th cooling branch; U represents the refrigerant flow rate of the i-th cooling branch; i W represents the heat dissipation power of the i-th cooling branch. i x represents the heat dissipation power of the i-th cooling branch; inFor the refrigerant inlet dryness of the cooling module; x out，i α represents the refrigerant outlet dryness of the i-th cooling branch; target For temperature uniformity; T Component，i The temperature of the controlled component.

[0020] A further improvement to the control method of the present invention lies in that,

[0021] The refrigerant inlet dryness of the cooling module is calculated by back-calculating its enthalpy value. The enthalpy value calculation expression is as follows:

[0022]

[0023] The refrigerant outlet dryness of the i-th cooling branch is calculated by back-calculating its enthalpy value. The enthalpy value calculation expression is as follows:

[0024]

[0025] In the formula, H in The enthalpy of the refrigerant at the cooling module inlet; H out，i Let be the enthalpy value of the refrigerant outlet in the i-th cooling branch; T1 is the mass flow rate of the refrigerant flowing through the high-pressure refrigerant in the regenerator; P1 is the refrigerant temperature at the high-pressure inlet of the regenerator; T2 is the refrigerant pressure at the high-pressure outlet of the regenerator; P2 is the refrigerant pressure at the high-pressure outlet of the regenerator. P3 is the refrigerant flow rate in the pump-driven circuit; P3 is the system low-pressure; U i Let be the heat dissipation power of the i-th cooling branch; Let be the refrigerant flow rate of the i-th cooling branch.

[0026] A further improvement to the control method of the present invention lies in that,

[0027] The model predictive controller is a model predictive controller used in a thermal management system, and also includes the following control logic:

[0028] x heat_transfer =f(P3, D) Plate，i L Plate，i W i );

[0029] x optimal =[x optimai，in x optimai，out ];

[0030]

[0031] x optimal =α TemUni，i ·x heat_transfer ;

[0032] Npump =f(T) Component，i x in，xout，i x optimal );

[0033] In the formula, α TemUni，i x is the temperature uniformity coefficient of the cooled component in the i-th cooling branch; heat_transfer x is the heat transfer coefficient calculated based on real-time parameters. optimai For optimal import and export dryness, including optimal import dryness x optimal，in With optimal export dryness x optimal，out N pump This is the optimal pump speed.

[0034] A further improvement to the control method of the present invention is that the step of controlling the opening of the three-way proportional regulating valve and the opening of the regulating valve of each cooling branch based on the predicted inlet dryness value of the cooling module as the control target and the predicted outlet dryness value of each cooling branch includes:

[0035] Based on the predicted value of the inlet dryness of the cooling module as the control target, the opening of the three-way proportional control valve is controlled by the inlet dryness PI controller so that the calculated value of the refrigerant inlet dryness of the cooling module approaches the predicted value of the inlet dryness.

[0036] Based on the predicted outlet dryness of each cooling branch as the control target, the opening of the regulating valve of each cooling branch is controlled by the outlet dryness PI controller of each cooling branch, so that the calculated value of the refrigerant outlet dryness of each cooling branch approaches the predicted outlet dryness value of each cooling branch.

[0037] A further improvement to the control method of the present invention is that the design and selection steps of the heat exchange cold plate include:

[0038] Step 1, obtain the extreme operating condition parameters; wherein, the extreme operating condition parameters include the extreme heat generation of the controlled component, the target temperature, the temperature difference of the cold plate heat exchange, the temperature uniformity target, the ambient temperature, the temperature difference of the air-cooled heat exchange, and the air-cooled oncoming wind speed.

[0039] Step 2: After using the model predictive controller to predict the optimal inlet and outlet dryness and flow rate under different cold plate parameters, the cold plate parameters are evaluated. The design is completed when the preset target is achieved; otherwise, the cold plate parameter update cycle begins. The cold plate evaluation parameter selected during the evaluation is the cold plate pressure drop.

[0040] A further improvement to the control method of the present invention lies in that, in the pump design and selection step,

[0041] The minimum displacement of the pump is expressed as,

[0042] In the formula, V pumpρ is the pump's minimum displacement; N is the pump's maximum speed; liquid The density of the refrigerant as a saturated liquid at the evaporation pressure; m pump The optimal pump drive flow rate is predicted by the model to control the controller output.

[0043] A further improvement to the control method of the present invention lies in that, in the compressor design and selection step,

[0044] The minimum displacement of the compressor is expressed as,

[0045] In the formula, V comp N is the minimum displacement of the compressor. comp ρ is the maximum speed of the compressor; gas The density of the refrigerant as a saturated gas at the evaporation pressure; m compressor This refers to the refrigerant flow rate in the compressor circuit.

[0046] A further improvement to the control method of the present invention is that the design and selection steps of the gas cooler and the isothermal regenerator include:

[0047] Obtain the design operating condition parameters; wherein, the design operating condition parameters include: ambient temperature, air-cooled heat exchange temperature difference, low-pressure heat exchange temperature difference, and air-cooled oncoming wind speed;

[0048] The intake pressure and air-cooled outlet temperature are obtained through the design operating parameters, and the optimal exhaust pressure is obtained through the air-cooled outlet temperature.

[0049] The refrigerant flow rate of the compressor circuit, the refrigerant flow rate of the pump drive circuit, the air-cooled outlet temperature, the discharge pressure, the low-pressure inlet dryness, and the low-pressure pressure are used as the input heat exchange conditions, and the low-pressure outlet dryness is used as the design target to design the isothermal regenerator.

[0050] The refrigerant flow rate of the compressor circuit, the exhaust temperature, the exhaust pressure, and the air velocity at the air cooling front are used as the input conditions for the gas cooler, and the gas cooler is designed with the air cooling outlet temperature as the design target.

[0051] Compared with the prior art, the present invention has the following beneficial effects:

[0052] This invention discloses a pump-driven two-phase flow thermal control system, which innovatively incorporates a homogenizer, a three-way proportional control valve, and a throttling valve. This addresses the technical problems of insufficient cooling capacity and unsatisfactory temperature uniformity in existing two-phase flow thermal control systems under high heat flux densities. Specifically, the opening of the three-way proportional control valve is adjustable. By adjusting its opening, the heat exchange capacity of the homogenizer can be controlled, thereby controlling the refrigerant inlet dryness of the cooling module. Furthermore, the cooling module consists of multiple heat exchange plates and corresponding control valves. Each set of heat exchange plates and control valves forms a cooling branch. The opening of the control valve is adjustable, and by controlling the opening of the control valve, the refrigerant flow rate through the cooling branch can be controlled, ultimately achieving the effect of controlling the outlet dryness.

[0053] In the control method of this invention, for the cooling problem of the thermal management system of electronic equipment with multiple heat sources, high heat flux density, and high temperature uniformity requirements, dual dryness control of the cold plate inlet and outlet of the pump-driven two-phase flow system is achieved through a uniform temperature regenerator and model predictive control. At the level of refrigerant dryness, a large refrigerant heat transfer coefficient is ensured to guarantee performance under high heat flux density. Since the dual dryness control has good refrigerant temperature uniformity and large flow rate, it can be combined with control to achieve high energy efficiency and high temperature uniformity control of the two-phase flow thermal control system.

[0054] In this invention, a combination of model predictive control and PI control can be used to achieve dual dryness control at the inlet and outlet of the cooling module, ensuring that the refrigerant operates in the high heat transfer coefficient range, ultimately leading to greater temperature uniformity control and improved system performance. Attached Figure Description

[0055] To more clearly illustrate the technical solutions in the embodiments of the present invention or the prior art, the drawings used in the description of the embodiments or the prior art are briefly introduced below; obviously, the drawings described below are some embodiments of the present invention, and those skilled in the art can obtain other drawings based on these drawings without creative effort.

[0056] Figure 1 This is a schematic diagram of a pump-driven two-phase flow thermal control system provided in an embodiment of the present invention;

[0057] Figure 2 This is a schematic diagram of the control logic of the pump-driven two-phase flow thermal control system in an embodiment of the present invention;

[0058] Figure 3 This is a schematic diagram of the heat transfer coefficient of the refrigerant at different dryness levels in an embodiment of the present invention;

[0059] Figure 4 This is a schematic diagram illustrating the design and selection process of each component in an embodiment of the present invention;

[0060] The reference numerals in the diagram are as follows: 1. Compressor; 2. Gas cooler; 3. Isostatic regenerator; 4. Throttling valve; 5. Liquid receiver; 6. Pump; 7. Cooling module. Detailed Implementation

[0061] To enable those skilled in the art to better understand the present invention, the technical solutions of the present invention will be clearly and completely described below with reference to the accompanying drawings of the embodiments of the present invention. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort should fall within the scope of protection of the present invention.

[0062] It should be noted that the terms "first," "second," etc., in the specification, claims, and accompanying drawings of this invention are used to distinguish similar objects and are not necessarily used to describe a specific order or sequence. It should be understood that such data can be interchanged where appropriate so that the embodiments of the invention described herein can be implemented in orders other than those illustrated or described herein. Furthermore, the terms "comprising" and "having," and any variations thereof, are intended to cover a non-exclusive inclusion; for example, a process, method, system, product, or apparatus that comprises a series of steps or units is not necessarily limited to those steps or units explicitly listed, but may include other steps or units not explicitly listed or inherent to such processes, methods, products, or apparatus.

[0063] The present invention will now be described in further detail with reference to the accompanying drawings:

[0064] Please see Figure 1 This invention provides a pump-driven two-phase flow thermal control system, comprising: a compressor 1, a gas cooler 2, a three-way proportional regulating valve, a temperature equalization regenerator 3, a throttling valve 4, a liquid storage tank 5, a pump 6, and a cooling module 7; wherein,

[0065] The outlet of the compressor 1 is connected to the first port of the three-way proportional control valve via the gas cooler 2, the second port of the three-way proportional control valve is connected to the inlet of the throttle valve 4 via the first heat exchange channel of the equalizer 3, and the third port of the three-way proportional control valve is connected to the inlet of the throttle valve 4.

[0066] The outlet of the throttle valve 4 is connected to the first inlet of the liquid storage tank 5, the gas outlet of the liquid storage tank 5 is connected to the inlet of the compressor 1, and the liquid outlet of the liquid storage tank 5 is connected to the second inlet of the liquid storage tank 5 in sequence via the pump 6, the second heat exchange channel of the equalizer 3, and the cooling module 7.

[0067] The cooling module 7 includes one or more cooling branches for exchanging heat with the controlled component (specifically, an external electronic device, etc.); the cooling branch includes a heat exchange cold plate and a regulating valve.

[0068] The principle of the invention in this embodiment is explained as follows: the three-way proportional regulating valve is adjustable. By adjusting it, the heat exchange of the isothermal regenerator can be controlled, thereby controlling the refrigerant inlet dryness of the cooling module. Furthermore, the cooling module consists of multiple heat exchange plates and corresponding regulating valves. Each set of heat exchange plates and regulating valves constitutes a cooling branch. The opening degree of the regulating valve is adjustable. By controlling the opening degree of the regulating valve, the refrigerant flow rate through the cooling branch can be controlled, ultimately achieving the effect of controlling the outlet dryness. Further explanation: ensuring a large refrigerant heat transfer coefficient at the refrigerant dryness level guarantees performance under high heat flux density. Due to the good temperature uniformity and large flow rate of the refrigerant, dual dryness control can be combined to achieve high energy efficiency and high temperature uniformity control in a two-phase flow thermal control system.

[0069] Further specific examples are explanatory, such as Figure 1 As shown, the outlet of the compressor 1 is connected to the inlet of the gas cooler 2, and the outlet of the gas cooler 2 is connected to port A of the isothermal regenerator 3. The circuit then splits into two paths: one bypasses to the high-pressure outlet B of the isothermal regenerator 3, and the other passes through the isothermal regenerator 3 for heat exchange. Specifically, the isothermal regenerator module can consist of an isothermal regenerator 3, a three-way proportional control valve, and a bypass branch. The high-pressure outlet B of the isothermal regenerator 3 is connected to the inlet of the throttle valve 4, and the outlet of the throttle valve 4 is connected to the inlet A of the liquid storage tank 5. The gaseous outlet B of the liquid storage tank 5 is connected to the suction end of the compressor 1, forming a complete gaseous circulation loop. The liquid outlet C of the liquid storage tank 5 is connected to the inlet of the pump 6 in the pump-driven circulation loop, the outlet of the pump 6 is connected to port C of the isothermal regenerator 3, the port D of the isothermal regenerator 3 is connected to the inlet of the cooling module 7, and the outlet of the cooling module 7 is connected to port D of the gas-liquid separator, forming a complete liquid pump cooling loop. Explanatoryly, the cooling module 7 consists of several cooling components (exemplary, such as cold plate heat exchange components, etc.), the number of which may be consistent with the number of heat sources.

[0070] Based on the above embodiments, the working process of a pump-driven two-phase flow thermal control system provided by the present invention specifically includes:

[0071] The high-temperature, high-pressure refrigerant at the outlet of compressor 1 is cooled down after heat exchange in gas cooler 2, and then flows into port A of homogenizer 3. Explain that the homogenizer module consists of three parts: a three-way proportional control valve, a bypass branch, and homogenizer 3. After the high-pressure, medium-temperature refrigerant at the outlet of gas cooler 2 enters the homogenizer module, it is first divided into two parts by the three-way proportional control valve. One part passes through the bypass branch, and the other part is cooled after passing through homogenizer 3. The two parts are mixed and then enter the throttling valve 4.

[0072] The low-pressure refrigerant passing through the throttle valve 4 enters port A of the liquid receiver 5 for gas-liquid separation. The gas returns to the suction port of the compressor 1 from port B of the liquid receiver 5, forming a complete gaseous circuit. The liquid in the liquid receiver 5 enters the pump 6 through port C. After being pressurized by the pump 6, the liquid refrigerant enters port C of the temperature equalization regenerator 3, where it is heated by the high-pressure, medium-temperature refrigerant. It then flows out through port D and enters the cooling module 7 to cool multiple heat sources. Specifically, the cooling module 7 can be composed of several cooling components (cold plates and regulating valves), the number of which is the same as the number of heat sources.

[0073] After passing through cooling module 7, the refrigerant enters liquid storage tank 5 through port D, completing the circulation.

[0074] Please see Figure 2 The present invention provides a control method for a pump-driven two-phase flow thermal control system, specifically a high-precision temperature control method based on dual dryness control, comprising:

[0075] Step 1: Obtain the parameter state space of the pump-driven two-phase flow thermal control system during operation;

[0076] Step 2: Based on the parameter state space, use the pre-trained predictive control model to obtain the overall working fluid flow rate as the control target and the inlet dryness and outlet dryness of each cooling branch.

[0077] Step 3: Based on the overall working fluid flow rate as the control target, control the pump speed; based on the inlet dryness and outlet dryness of each cooling branch as the control target, control the opening of the three-way proportional regulating valve and the opening of the regulating valve of each cooling branch.

[0078] In a further preferred embodiment of the control method of the present invention, step 1 includes the following specific steps:

[0079] The state parameters of the pump-driven two-phase flow thermal control system are collected in real time, and the refrigerant dryness at the inlet of the cooling module is calculated. The cold plate outlet dryness of each cooling branch is calculated based on the heat output, refrigerant flow rate, and inlet dryness of each cooling branch.

[0080] Specifically, the parameter state space is represented as follows:

[0081]

[0082] In the formula, P3 is the system low-pressure; D Plate，i L represents the characteristic cross-sectional dimension of the heat exchange plate in the i-th cooling branch; Plate，i Let be the characteristic length of the heat exchange plate of the i-th cooling branch; U represents the refrigerant flow rate of the i-th cooling branch; i W represents the heat dissipation power of the i-th cooling branch. i x is the heat dissipation power of the i-th cooling branch; in For the refrigerant inlet dryness of the cooling module; x out，i αx represents the refrigerant outlet dryness of the i-th cooling branch; arget For temperature uniformity; T Component，i The temperature of the controlled component.

[0083] Further, in theoretical research, the model predictive controller uses system state input and system disturbance input (interpretive, disturbances are added to theoretical data to approximate actual data, while no additional disturbances are added to the actual collected parameter data) to determine the current system operating state and provide optimal inlet dryness and optimal outlet dryness reference values, providing reference values ​​for the controller (exemplarily, a PI controller); at the same time, based on the overall working fluid flow rate, the optimal pump speed can be given for control.

[0084] In summary, the method provided by the embodiments of the present invention can be divided into three parts: upper-level model predictive control, middle-level dryness calculation, and lower-level PI execution control.

[0085] Further illustratively, since dryness cannot be directly measured, the quantities in this invention are calculated through virtual dryness measurement. Specifically, the calculation of the inlet dryness of the cooling module mainly relies on the isothermal regenerator module, which collects the temperature and pressure of the refrigerant at the high-pressure side of the isothermal regenerator; the temperature, pressure, and flow rate of the refrigerant at the high-pressure outlet of the isothermal regenerator; and the total flow rate, branch flow rate, calorific value, and actual temperature of the pump-driven cooling module.

[0086] In this embodiment of the invention, the formulas for calculating the dryness of the cooling module inlet and outlet in the dryness calculation module are as follows:

[0087]

[0088]

[0089] In the formula, H in The enthalpy of the refrigerant at the cooling module inlet; H out,i Let be the enthalpy value of the refrigerant outlet in the i-th cooling branch; T1 is the mass flow rate of the refrigerant flowing through the high-pressure refrigerant in the regenerator; P1 is the refrigerant temperature at the high-pressure inlet of the regenerator; T2 is the refrigerant pressure at the high-pressure outlet of the regenerator; P2 is the refrigerant pressure at the high-pressure outlet of the regenerator. P3 is the refrigerant flow rate in the pump-driven circuit; P3 is the system low-pressure; U i The heat dissipation power of the i-th cooling branch; Let be the refrigerant flow rate of the i-th cooling branch.

[0090] In a further preferred embodiment of the present invention, since the heat transfer coefficient of the refrigerant is different at different dryness, there is an optimal refrigerant inlet and outlet dryness from the perspective of heat exchange; when the model prediction controller predicts the optimal inlet and outlet dryness, it first considers the heat transfer coefficient of the refrigerant at different dryness to ensure that the system cools the electronic components at the optimal heat transfer coefficient.

[0091] Therefore, the model predictive controller used in this embodiment of the invention is based on the existing MPC controller for thermal management systems. [1-3] , and add new control relationships to it; among them, reference [1].Wang W, Zhao Z, Zhou Q, etal. Model predictive control for the operation of a transcritical CO2airsource heat pump water heater [J]. Applied Energy, 2021,300:117339; [2]. Wang H, Wang W, Song Y, et al. Data-driven model predictive control of transcriticalCO2systems for cabin thermal management in coollng mode[J].Applied ThermalEngineering, 2023:121337;[3].Yue B, Su B, Xiao F, et al.Energy-oriented ControlRetrofit for Existing HVAC system adopting data-driven MPC-Methodology, Implementation and Field test[J].Energy and Buildings, 2023:113286;

[0092] The new control relationships include:

[0093] xheat_transfer =f(P3, D) Plate，i L plate，i W i );

[0094] Given the optimal dryness fraction, in addition to considering the heat transfer coefficient of refrigerants with different dryness fractions, the influence of the temperature uniformity of the cooled components is also taken into account.

[0095] x optimal =[x optimai，in x optimai，out ];

[0096]

[0097] x optimal =α TemUni，i ·x heat_transfer ;

[0098] In the formula, α Temuni，i is the temperature uniformity coefficient of the cooled component in the i-th cooling branch, and its value range can be [0.7, 1].

[0099] When predicting the optimal pump speed, the temperature of the controlled component and the current and optimal inlet and outlet dryness of the cooling module are taken into account to ensure that the system can be adjusted to the set optimal dryness.

[0100] N pump =f(T) Component，i x in x out，i x optimal ).

[0101] In a further preferred embodiment of the present invention, the step of controlling the opening of the three-way proportional control valve and the opening of the control valve of each cooling branch based on the inlet dryness and outlet dryness of each cooling branch as the control target specifically includes:

[0102] Two PI controllers can be used to control the dryness of the inlet and outlet, ensuring that the cold plate has a large heat transfer coefficient and also ensuring the temperature uniformity of the cooled components.

[0103] Specifically, the PI controller performs dryness control, comprising two types: an inlet dryness controller and an outlet dryness controller. The inlet dryness PI controller controls the opening of the three-way proportional regulating valve in the isothermal regenerator to control the inlet dryness of the cooling module calculated by the dryness calculator. The outlet dryness PI controllers for each cooling branch regulate the flow distribution in each branch by controlling the opening of the valves in that branch, thereby controlling the outlet dryness of each branch. The actual dryness signal is calculated using the inlet dryness, flow rate, and heat generation of the device. The intermediate dryness calculator primarily serves to represent the difficult-to-measure dryness value through other modules, thus providing input values ​​for the PI controller.

[0104] In summary, in this embodiment of the invention, the heat exchange of the regenerator is adjusted by regulating the refrigerant flow rate through the high-pressure side of the regenerator, thereby adjusting the refrigerant dryness at the outlet of the low-pressure side regenerator; by controlling the valve opening of each branch, the refrigerant flow rate through this branch is controlled, thereby achieving refrigerant dryness control at the outlet of each branch.

[0105] In the application of this invention, the thermal control system is not limited to using CO2; other refrigerants can also achieve the same effect.

[0106] Please see Figure 3 The principle of the invention in this embodiment is explained as follows: one of the bases for predicting the optimal inlet and outlet dryness of the system using a model predictive controller is the difference in thermal conductivity of the refrigerant at different dryness levels; for example... Figure 3 As shown, the heat transfer coefficient of the refrigerant in the two-phase region varies with the dryness fraction. In the technical solution provided by this invention, one of the significances of the optimal inlet and outlet dryness fractions lies in using a model predictive controller to determine the current state of the cooled component, taking into account the parameters D of the system's cold plate. Plate，i L Plate，i Low-pressure P3 and heat generation of electronic components W i In terms of heat transfer coefficient, it improves the system's heat exchange efficiency; furthermore, it also improves the temperature uniformity of the cooled components by changing the inlet and outlet dryness, using α Temuni，i The temperature uniformity of the cooled components in each branch is measured by the target temperature uniformity coefficient α. target Current traffic flow on each branch The heat dissipation power U of the cooled component i The heating power of cold components (W) i Refrigerant inlet dryness x in The calculation shows that during the calculation process, the inlet and outlet dryness is converted into heat (W) through the cooled component. i Traffic flow of each branch Cooling module inlet refrigerant dryness x in Coupling means determining the optimal outlet dryness of each branch simultaneously with determining the optimal inlet dryness. The expression for the optimal dryness is:

[0107] x heat_transfer =f(P3, D) Plate，i L Plate，i W i );

[0108] α TemUni，i =f(W i α target );

[0109] x optimal =α TemUni，i ·x heat_transfer ;

[0110] Further explaining, another function of the model predictive controller is to control the pump speed, thereby controlling the flow rate. In the control system provided in this embodiment of the invention, the refrigerant flow rate is automatically distributed in each branch through the opening degree of each branch valve. If the pump operates at a constant speed, it is difficult to meet the cooling requirements of the more variable electronic components. Here, the model predictive controller is used to predict the optimal pump speed, so that the system is in the optimal operating state. Its main principle is to evaluate the refrigerant flow rate of the current pump-driven cycle based on the current temperature of each cooled component and the difference between the current inlet and outlet dryness of the cooling module and the optimal inlet and outlet dryness, thereby determining the optimal pump speed. The expression for the optimal speed is: N pump =f(T) Compinent,i ,x in ,x out,i ,x optimal ).

[0111] In the technical solution disclosed in this invention, the model predictive controller collects real-time system state parameters and calculates reference values, which are also optimal values, for the control system, such as inlet and outlet dryness. Before using the model predictive controller, it is necessary to develop a data-driven model. Based on the study of system dynamic characteristics, simulations and experiments for system identification are designed, and simulation and experimental data are collected. Based on big data, data analysis, data cleaning, and data integration are performed, removing irrelevant, repetitive, and smoothing noise from the original data. Response surface methodology is used to extract key factors affecting the inlet and outlet dryness of the heat exchanger cold plate. By selecting a suitable machine learning algorithm, a data-driven model capable of accurately predicting the system's operating conditions is trained. Finally, based on the data-driven model, a model predictive intelligent control algorithm is developed. The model predictive control strategy can solve the objective equation within the controller based on the real-time operating conditions and disturbances of the two-phase system, directly providing the optimal control sequence for the inlet and outlet dryness of the cold plate in the next few steps.

[0112] In this embodiment of the invention, the controlled variable during control system operation is the dryness of the cooling system inlet and outlet, but the dryness value is difficult to measure during actual operation. This embodiment of the invention uses a dryness calculation module to solve this problem. The dryness calculation module needs to collect real-time system status parameters, including: the mass flow rate of the high-pressure refrigerant flowing through the regenerator. Regenerator high-pressure inlet refrigerant temperature T1, pressure P1; high-pressure outlet refrigerant temperature T2, pressure P2; refrigerant flow rate in pump drive circuit. The refrigerant pressure at the low-pressure inlet of the regenerator is P3; the heat dissipation power of the cooled component in the i-th branch is U. i ; Refrigerant flow rate of the i-th cooling branch

[0113] In the dryness calculation, the dryness of the cooling module inlet is calculated by back-calculating its enthalpy value, which is calculated as follows:

[0114]

[0115] The dryness of the cooling module outlet is also calculated using the refrigerant enthalpy values ​​of its respective branch outlets:

[0116]

[0117] After calculating and obtaining the actual inlet and outlet temperatures of the cooling module, the dryness input value can be provided to the dryness PI controller to achieve dual dryness control.

[0118] In this embodiment of the invention, the refrigerant dryness calculation serves as a control connection, which is a unique feature of the technical solution. Refrigerant dryness control offers significant advantages in improving system performance and ensuring temperature uniformity of cooled components. However, in actual systems, refrigerant dryness cannot be directly measured. Since the actual input of the PI controller after the upper-level predictive model controller provides the control signal cannot be directly measured by a sensor, it can be achieved through a dryness calculator. The main function of the dryness calculator is to acquire and calculate indirectly measurable quantities by collecting system state parameters. This connects the upper-level predictive model controller with the lower-level dryness PI control actuator.

[0119] The PI control module of this invention can be composed of two parts; wherein,

[0120] The cooling module refrigerant inlet dryness PI controller uses the optimal inlet dryness output from the model predictor as a reference value and the refrigerant dryness calculated by the dryness calculator as the input value for control. Its output is the opening degree of the three-way proportional control valve in the isothermal regenerator. The refrigerant, driven by the pump, enters the regenerator in a saturated liquid state, and its dryness control is actually achieved by controlling the amount of heat applied to the refrigerant. The low-pressure liquid refrigerant exchanges heat with the high-pressure refrigerant in the regenerator, and the three-way proportional control valve controls the ratio of the refrigerant flow rate through the regenerator to that in the bypass branch. A higher high-pressure refrigerant flow rate through the regenerator results in a greater heating capacity for the low-pressure liquid refrigerant, leading to a higher refrigerant inlet dryness in the cooling module. Conversely, adjusting the three-way proportional control valve to reduce the high-pressure refrigerant flow rate through the regenerator results in a lower refrigerant inlet dryness level.

[0121] The refrigerant dryness PI controller at the outlet of each branch of the cooling module uses the optimal outlet dryness predicted by the model as a reference value. The dryness of each branch outlet calculated by the dryness calculator is the input value of the controller, and its output is the valve opening degree in each branch. Since there are many heat sources to be cooled in the system, the cooling module has many branches. The cooling demand in each branch may be different, so a multi-branch parallel connection is used to cool the equipment. When the valve opening increases, the refrigerant flow rate through that branch increases, and the refrigerant dryness at the cold plate outlet decreases; similarly, when the valve opening decreases, the refrigerant flow rate through that branch decreases, and the refrigerant dryness at the cold plate outlet increases. Therefore, the refrigerant dryness at the outlet of the cooling module can be controlled by this PI controller.

[0122] The PI controller acts as the actuator, and its control effect is crucial to the entire control logic. In this two-part PI controller system, the cooling module inlet dryness controller has only one PI controller, and its principle is easy to understand: refrigerant dryness is adjusted by regulating the flow ratio through the regenerator and the bypass. The outlet dryness controller consists of multiple branch dryness PI controllers, essentially regulating the distribution of flow across different branches. If the optimal outlet dryness predicted by the model predictive controller cannot be fully achieved through flow distribution, the predictive controller will adjust the pump speed to ensure that the cooling module outlet dryness PI controller can adjust the outlet dryness to the reference state.

[0123] Please see Figure 4 The present invention provides a design and selection method for various components of the above-mentioned pump-driven two-phase flow thermal control system, including:

[0124] In a pump-driven two-phase flow thermal control system, the design of the cold plate has the greatest impact on the cooling effect. However, the design of the cold plate differs from that of a typical heat exchanger. Typically, heat exchanger design requires given conditions such as the flow rate, inlet temperature, pressure, or enthalpy of the fluid flowing through both sides of the heat exchanger under extreme operating conditions, and then the heat exchanger is designed accordingly. However, in this embodiment of the invention, the cold plate design dimensions are coupled with the calculation of optimal dryness fraction, making its design and selection more complex. The design and selection of each component in the entire system includes the following steps:

[0125] Step 1: Determine the relevant parameters for extreme operating conditions to provide guidance for subsequent design. Among these, the extreme operating parameters related to the cold plate design include the extreme heat generation (W) of the equipment's electronic components. max Target temperature T target Cold plate heat exchange temperature difference ΔT, temperature uniformity target α target Ambient temperature T ambient Air-cooled heat exchange temperature difference ΔT1, air-cooled oncoming wind speed v airExplanatory, step 1 is the work of determining the operating conditions before the design of the entire system. It is necessary to determine the extreme operating conditions of the equipment and the requirements of the cooled components under these conditions. The following design is based on this and selects and designs the parameters of each component.

[0126] Step 2: During the cold plate parameter design and selection process, a model predictive controller needs to be used to determine the optimal inlet and outlet dryness fraction (x) under different cold plate parameters. optimai and flow rate m i After prediction, the parameter is evaluated. The design can be completed only if the target is met; otherwise, a cold plate parameter update cycle begins. This cyclical approach addresses situations where design conditions need to be updated based on cold plate parameters. The cold plate evaluation parameter is chosen as the cold plate pressure drop ΔP, primarily because the calculated flow rate and inlet / outlet dryness meet the cooling requirements. Therefore, pressure drop is used as the evaluation standard, as detailed below:

[0127] Step 2.1, based on the target temperature T in the extreme operating condition. target The system's low-pressure P3 is calculated based on the temperature difference ΔT between the heat exchanger and the cold plate.

[0128] Step 2.2, given the initial cold plate parameters L plate，i and D plate，i ;

[0129] Step 2.3, firstly, through the cold plate parameter L plate，i and D plate，i Based on the model prediction controller mentioned above, the optimal dryness fraction and pump-driven circulation flow rate are calculated:

[0130] x heat_transfer =f(P3, D) Plate，i L Plate，i W i );

[0131] α TemUni，i =f(W i α target );

[0132] x optimal =α TemUni，i ·x heat_transfer ;

[0133]

[0134] Step 2.4: After calculating the cold plate flow rate and the optimal inlet and outlet dryness, these boundary conditions are substituted into the model with initial cold plate parameters for design calculation, and the evaluation coefficient of the cold plate is given. Since the given flow rate and inlet and outlet dryness can meet the cooling requirements, the cold plate pressure drop ΔP is selected as the evaluation coefficient here.

[0135] Step 2.5: Determine whether the evaluation coefficient of the cold plate meets the design requirements. If yes, output the optimal cold plate parameters and proceed to step 3; otherwise, update the cold plate feature length L. plate，i and characteristic cross-sectional dimensions D plate，i And return to step 2.3;

[0136] Explanatory, step 2 is the focus of system design, specifically the design of the cold plates for each cooling branch. This is because the optimal inlet and outlet dryness (x) varies with the characteristic dimensions of the cold plates. optimal Since the parameters of a cold plate are variable, their design cannot be based on given boundary conditions and dimensional parameters as in traditional heat exchanger design. Here, a model predictive controller, as mentioned in the control process, is used to calculate the optimal inlet and outlet refrigerant dryness and flow rate for different cold plate parameters. Then, the current cold plate is evaluated based on its pressure drop and heat exchange. If it meets the requirements, the current parameters are used as the design parameters; if not, the parameters are recalculated until the required cold plate parameters are found.

[0137] Step 3, compressor and pump selection, the specific steps are as follows:

[0138] Step 3.1, based on the cold plate design parameters L designed in Step 2. plate，i D plate，i and system low pressure P3, ultimate heating power W max and temperature uniformity target α target The isoparametric data model predicts the controller to calculate the optimal inlet and outlet dryness of the system. optimal and optimal pump drive flow rate m pump ;

[0139] Step 3.2: Calculate the minimum pump displacement based on the calculated optimal pump drive flow rate.

[0140]

[0141] In the formula, V pump ρ is the pump's minimum displacement; N is the pump's maximum speed; liquid The density of the refrigerant as a saturated liquid at the evaporation pressure;

[0142] Step 3.3, based on the optimal inlet and outlet dryness fraction x calculated in Step 2. optimal and optimal pump drive flow rate m pump Calculate the compressor flow rate:

[0143] m compressor (1-x in，comp ) = m pump x optimai，out ;

[0144] m pump (H(x optimal，in)-H(saturated liquid))=m compressor (H(x in，comp )-H(P1,T1));

[0145] In the formula, x in，comp The dryness of the refrigerant entering the receiver tank from the compressor circuit;

[0146] The refrigerant flow rate (m) in the compressor circuit can be obtained by solving the two equations simultaneously. compressor ;

[0147] Step 3.4: Calculate the minimum displacement of the compressor based on the refrigerant flow rate in the compressor circuit.

[0148]

[0149] In the formula, V comp N is the minimum displacement of the compressor. comp ρ is the maximum speed of the compressor; gas This is the density of the refrigerant as a saturated gas at the evaporation pressure;

[0150] Explanatoryly, step 3 mainly involves selecting the displacement of the pumps and compressors in the system, primarily based on the maximum heat output under extreme operating conditions. The pump and compressor flow rates are coupled; here, the coupled solution for the flow rates of the pump drive circuit and compressor circuit is achieved through gas-liquid balance calculations in the receiver tank and heat exchange calculations using the isothermal regenerator, ultimately realizing the calculation and selection of the pumps and compressors. It is important to note that step 3 uses the air-cooled outlet temperature T1 and the optimal discharge pressure P. optimai This will be explained in step 4.

[0151] Step 4, Gas cooler and isothermal regenerator parameters and selection:

[0152] Step 4.1, determine the ambient temperature T under the design operating conditions. ambient Air-cooled heat exchange temperature difference ΔT1, low-pressure heat exchange temperature difference ΔT, air-cooled oncoming wind speed v air ;

[0153] Step 4.2: Obtain the intake pressure P3 and the air-cooled outlet temperature T1 using the above parameters, and then calculate the optimal exhaust pressure P of the system by setting the air-cooled outlet temperature. optimai ;

[0154] Step 4.3, change the refrigerant flow rate m in the compressor circuit. compressor Pump drive circuit refrigerant flow rate (m) pump Air-cooled outlet temperature T1, exhaust pressure P optimai Low-pressure inlet dryness x optimall，in Low-pressure P3 is used as the input for heat exchange, and the low-pressure outlet dryness fraction x optimal，outDesign the regenerator to meet the design objectives;

[0155] Step 4.4, change the refrigerant flow rate m in the compressor circuit. compressor Exhaust temperature T0, exhaust pressure P optimal Cool air, frontal wind speed v air As the input condition for the gas cooler, the gas cooler is designed with the gas outlet temperature T1 as the design target.

[0156] Explanatoryly, step 4 mainly involves the design of the isothermal regenerator and the gas cooler. As an important part of the isothermal regenerator module, the isothermal regenerator controls the refrigerant inlet dryness of the cooling module in the system of this invention. Its selection and design are very important for the normal operation of the system. The determination of the design conditions of the isothermal regenerator and the gas cooler is relatively complex because the system has a strong coupling relationship. It is necessary to calculate the refrigeration cycle based on the design of other components before it can be finally determined.

[0157] In summary, this invention discloses a control and component design and selection method for a pump-driven two-phase flow thermal control system, which requires real-time acquisition of system state parameters. The proposed upper-level control module, a model predictive controller, determines the optimal inlet and outlet dryness of the cooling module based on the system's characteristic dimensions and current system parameters, considering the heat transfer mechanism and the temperature uniformity of the cooled components. This predictive controller also monitors the temperature, actual, and target dryness of the cooled components, providing a pump speed that meets the optimal requirements. Then, a dryness calculator connected to the calculation module calculates the inlet and outlet dryness of the cooling module, achieving real-time monitoring of the system's dryness. Finally, a control execution module, consisting of multiple dryness PI controllers, controls the inlet and outlet dryness of the cooling module by adjusting three-way proportional control valves / cooling branch valves. Furthermore, a design and selection process for each component is designed for the specific control portion of this system to facilitate system development and application. The technical solution provided by the embodiments of the present invention improves the cooling efficiency of the pump-driven two-phase flow thermal control system, significantly improves the temperature uniformity of the cooled components, and also improves the performance of the system. This promotes the application and popularization of the pump-driven two-phase flow thermal control system and makes a significant contribution to environmental protection and alleviating the fossil energy crisis.

[0158] 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 it. Although the present invention has been described in detail with reference to the above embodiments, those skilled in the art should understand that modifications or equivalent substitutions can still be made to the specific implementation of the present invention. Any modifications or equivalent substitutions that do not depart from the spirit and scope of the present invention should be covered within the scope of protection of the claims of the present invention.

Claims

1. A pump-driven two-phase flow thermal control system, characterized in that, include: Compressor (1), gas cooler (2), three-way proportional control valve, isothermal regenerator (3), throttle valve (4), liquid storage tank (5), pump (6) and cooling module (7); The outlet of the compressor (1) is connected to the first port of the three-way proportional control valve via the gas cooler (2), the second port of the three-way proportional control valve is connected to the inlet of the throttle valve (4) via the first heat exchange channel of the equalizer (3), and the third port of the three-way proportional control valve is connected to the inlet of the throttle valve (4). The outlet of the throttle valve (4) is connected to the first inlet of the liquid storage tank (5), the gas outlet of the liquid storage tank (5) is connected to the inlet of the compressor (1), and the liquid outlet of the liquid storage tank (5) is connected to the second inlet of the liquid storage tank (5) via the pump (6), the second heat exchange channel of the equalizer (3), and the cooling module (7) in sequence. The cooling module (7) includes one or more cooling branches, each of which includes a heat exchange plate and a regulating valve.

2. A control method for the pump-driven two-phase flow thermal control system as described in claim 1, characterized in that, Includes the following steps: Obtain the parameter state space of the pump-driven two-phase flow thermal control system during operation; The parameter state space is input into the predictive control model to obtain the overall working fluid flow prediction value, the inlet dryness prediction value of the cooling module, and the outlet dryness prediction value of each cooling branch as the control target. Based on the predicted value of the overall working fluid flow rate as the control target, the pump speed is controlled; Based on the predicted inlet dryness of the cooling module and the predicted outlet dryness of each cooling branch, the opening of the three-way proportional control valve and the opening of the control valve of each cooling branch are controlled.

3. The control method according to claim 2, characterized in that, The parameter state space is represented as follows: In the formula, P3 is the system low-pressure; D Plate,i L represents the characteristic cross-sectional dimension of the heat exchange plate in the i-th cooling branch; Plate,i Let be the characteristic length of the heat exchange plate of the i-th cooling branch; U represents the refrigerant flow rate of the i-th cooling branch; i W represents the heat dissipation power of the i-th cooling branch. i x represents the heat dissipation power of the i-th cooling branch; in For the refrigerant inlet dryness of the cooling module; x out,i α represents the refrigerant outlet dryness of the i-th cooling branch; target The goal is to achieve temperature uniformity; Component,i The temperature of the controlled component.

4. The control method according to claim 3, characterized in that, The refrigerant inlet dryness of the cooling module is calculated by back-calculating its enthalpy value. The enthalpy value calculation expression is as follows: The refrigerant outlet dryness of the i-th cooling branch is calculated by back-calculating its enthalpy value. The enthalpy value calculation expression is as follows: In the formula, H in The enthalpy of the refrigerant at the inlet of the cooling module; H out,i Let be the enthalpy value of the refrigerant outlet in the i-th cooling branch; The mass flow rate of the high-pressure refrigerant flowing through the regenerator; T1 is the refrigerant temperature at the high-pressure inlet of the regenerator; P1 is the refrigerant pressure at the high-pressure inlet of the regenerator; T2 is the refrigerant temperature at the high-pressure outlet of the regenerator; P2 is the refrigerant pressure at the high-pressure outlet of the regenerator. P3 is the refrigerant flow rate in the pump-driven circuit; P3 is the system low-pressure; U i Let be the heat dissipation power of the i-th cooling branch; Let be the refrigerant flow rate of the i-th cooling branch.

5. The control method according to claim 3, characterized in that, The model predictive controller is a model predictive controller used in a thermal management system, and also includes the following control logic: x heat_transfer =f(P3,D Plate,i ,L Plate,i ,W i ); x optimal =[x optimal,in ,x optimal,out ]; x optimal =a TemUni,i ·x heat_transfer ; N pump =f(T Component,i ,x in ,x out,i ,x optimal ); In the formula, α TemUni,i x is the temperature uniformity coefficient of the cooled component in the i-th cooling branch; heat_transfer x is the heat transfer coefficient calculated based on real-time parameters. optimal For optimal import and export dryness, including optimal import dryness x optimal,in With optimal export dryness x oprimal,out N pump This is the optimal pump speed.

6. The control method according to claim 5, characterized in that, The steps of controlling the opening of the three-way proportional control valve and the opening of the control valve of each cooling branch based on the predicted inlet dryness value of the cooling module and the predicted outlet dryness value of each cooling branch as the control target include: Based on the predicted value of the inlet dryness of the cooling module as the control target, the opening of the three-way proportional control valve is controlled by the inlet dryness PI controller so that the calculated value of the refrigerant inlet dryness of the cooling module approaches the predicted value of the inlet dryness. Based on the predicted outlet dryness of each cooling branch as the control target, the opening of the regulating valve of each cooling branch is controlled by the outlet dryness PI controller of each cooling branch, so that the calculated value of the refrigerant outlet dryness of each cooling branch approaches the predicted outlet dryness value of each cooling branch.

7. The control method according to claim 5, characterized in that, The design and selection steps for the heat exchange cold plate include: Step 1, obtain the extreme operating condition parameters; wherein, the extreme operating condition parameters include the extreme heat generation of the controlled component, the target temperature, the temperature difference of the cold plate heat exchange, the temperature uniformity target, the ambient temperature, the temperature difference of the air-cooled heat exchange, and the air-cooled oncoming wind speed. Step 2: After using the model predictive controller to predict the optimal inlet and outlet dryness and flow rate under different cold plate parameters, the cold plate parameters are evaluated. The design is completed when the preset target is achieved; otherwise, the cold plate parameter update cycle begins. The cold plate evaluation parameter selected during the evaluation is the cold plate pressure drop.

8. The control method according to claim 5, characterized in that, In the design and selection steps of the pump... The minimum displacement of the pump is expressed as, In the formula, V pump ρ is the pump's minimum displacement; N is the pump's maximum speed; liquid The density of the refrigerant as a saturated liquid at the evaporation pressure; m pump The optimal pump drive flow rate is predicted by the model to control the controller output.

9. The control method according to claim 5, characterized in that, In the design and selection steps of the compressor The minimum displacement of the compressor is expressed as, In the formula, V comp N represents the minimum displacement of the compressor. comp ρ is the maximum speed of the compressor; gas The density of the refrigerant as a saturated gas at the evaporation pressure; m compressor This refers to the refrigerant flow rate in the compressor circuit.

10. The control method according to claim 5, characterized in that, The design and selection steps for the gas cooler and the isothermal regenerator include: Obtain the design operating condition parameters; wherein, the design operating condition parameters include: ambient temperature, air-cooled heat exchange temperature difference, low-pressure heat exchange temperature difference, and air-cooled oncoming wind speed; The intake pressure and air-cooled outlet temperature are obtained through the design operating parameters, and the optimal exhaust pressure is obtained through the air-cooled outlet temperature. The refrigerant flow rate of the compressor circuit, the refrigerant flow rate of the pump drive circuit, the air-cooled outlet temperature, the discharge pressure, the low-pressure inlet dryness, and the low-pressure pressure are used as the input heat exchange conditions, and the low-pressure outlet dryness is used as the design target to design the isothermal regenerator. The refrigerant flow rate of the compressor circuit, the exhaust temperature, the exhaust pressure, and the air velocity at the air cooling front are used as the input conditions for the gas cooler, and the gas cooler is designed with the air cooling outlet temperature as the design target.

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