A modeling method for thermal management system of electric vehicle based on moving boundary method

Abstract

The present invention discloses a method for modeling a thermal management system of an electric vehicle based on a moving boundary method, comprising the following steps: simplifying the one-dimensional models of an evaporator, a condenser and a plate heat exchanger; deriving the heat exchange equations of each phase region of the evaporator and the plate heat exchanger; deriving the heat exchange equations of each phase region of the condenser; establishing the overall equations of the evaporator, the condenser and the plate heat exchanger; establishing a dynamic model of the thermal management system; and calculating the entire operating parameter space of the system under given working conditions. The thermal management model established by the present invention based on the moving boundary method takes into account the different heat exchange processes in different phase regions of the heat exchanger, and the accuracy is improved; the established thermal management model is used to design and optimize the thermal management system of an electric vehicle, effectively reducing the number of experiments and saving costs; the thermal management system modeling method can be applied to the modeling of most types of heat exchangers, and can also be used for the modeling of electric vehicle thermal management systems coupled with air conditioning systems and battery cooling systems.

Description

Technical Field

[0001] The present invention relates to the field of electric vehicle thermal management design and optimization, and in particular to a modeling method of an electric vehicle thermal management system based on a moving boundary method. Background Art

[0002] Electric vehicles are developing rapidly, especially pure electric vehicles, which will become the mainstream trend in the future due to their advantages such as low noise, zero emissions and high energy utilization. Since electric vehicles and traditional internal combustion vehicles have great differences in driving mode and structure, the thermal management system must also be improved accordingly. The most important parts of the thermal management system of electric vehicles are the cabin thermal management system and the battery cooling system. The cabin thermal management system has a great impact on the thermal comfort of the passengers, and the battery cooling system ensures that the battery pack works in a suitable temperature range, which has a great impact on the performance of the battery pack.

[0003] Among them, the cooling of the battery pack is the focus of the thermal management system of electric vehicles. The battery pack will generate a lot of waste heat when the vehicle is driving and charging. If the battery pack works in a high temperature environment for a long time, it will seriously damage the life of the battery pack, and even cause spontaneous combustion, endangering driving safety. Traditional battery pack cooling mainly includes two forms: air cooling and liquid cooling. The air cooling structure is simple and low in cost, and is used in early electric vehicles; the liquid cooling form has excellent thermal conductivity, but the early liquid cooling structure is connected to the radiator alone for heat dissipation, the heat dissipation rate is slow, and the waste heat cannot be effectively utilized. In recent years, the liquid cooling form has been coupled with the air conditioning system of electric vehicles, which can not only cool quickly, but also utilize the waste heat of the battery pack when needed. It is an efficient and energy-saving cooling system.

[0004] The air conditioning system is also an important part of the thermal management system of electric vehicles. With the anxiety about the winter endurance of electric vehicles, the traditional cabin PTC heating can no longer meet the demand for efficient and energy-saving heating due to its low thermal efficiency. At this time, the heat pump system has attracted attention due to its good second-law efficiency. The heat pump system can make heat flow from the low-temperature outside of the car to the cabin by consuming only a small amount of reverse cycle net work, and the structure only needs to make few modifications to the traditional automobile air conditioning system, so it has strong adaptability. At present, the electric vehicle thermal management system structure that couples the battery cooling system and the heat pump system through a plate heat exchanger to achieve the purpose of comprehensive utilization of the heat of the whole vehicle is gradually being adopted. This structure can use the waste heat of the battery pack to heat the cabin and still have a high heating capacity at a lower outdoor temperature (Dynamic simulation of an air-source heat pump water heater using novel modified evaporator model).

[0005] To design and optimize a thermal management system, it is necessary to accurately understand the entire operating parameter space of the thermal management system. There are currently two methods for the development and design of thermal management systems: prototype experiment method and system simulation method. If the prototype experiment method is used for development, the experimental cost of covering the entire parameter space is high and the cycle is long, so it is rarely used now. The method of establishing a system simulation model can greatly shorten the development cycle, help understand the system performance and guide the development of the thermal management system, and is widely used. At present, the simulation model of the thermal management system has the following main problems: (1) The model is usually composed of lumped parameter models of each component, but the lumped parameter model of the heat exchanger has low accuracy and is greatly affected by the differences in the heat exchanger;

[0006] (2) Lack of efficient and accurate simulation modeling methods for thermal management systems. Summary of the invention

[0007] The purpose of the present invention is to provide a dynamic modeling method for an electric vehicle thermal management system based on a moving boundary method, which couples a battery cooling system with a heat pump system, and a dynamic modeling method for a thermal management system that takes into account both battery pack temperature control and cabin temperature control, and can accurately simulate system operating state parameters. According to the needs of the actual operation of the system, the evaporator, condenser and plate heat exchanger are equivalent to a one-dimensional heat exchange pipeline and divided into different calculation phase areas, wherein the evaporator and plate heat exchanger are divided into two phase areas, namely, a two-phase area and a superheated area, and the condenser is divided into three phase areas, namely, a superheated area, a two-phase area and a supercooled area. According to the principles of mass conservation, energy conservation and tube wall energy conservation, the calculation equations for the outlet pressure and specific enthalpy of the refrigerant side and the outlet temperature of the fluid side outside the tube in each phase area are given. According to the calculation equations of each phase area, the structural parameters of the heat exchanger and the inlet side state parameters, the heat exchange models of the evaporator, condenser and plate heat exchanger are established. According to the heat exchange model of the heat exchanger and the steady-state calculation model of the compressor and expansion valve, a dynamic simulation model of the thermal management system is coupled to provide guidance for the development and optimization design of the thermal management system.

[0008] The present invention is achieved by at least one of the following technical solutions.

[0009] A method for modeling a thermal management system of an electric vehicle based on a moving boundary method comprises the following steps:

[0010] (1) Simplified one-dimensional model of evaporator, condenser and plate heat exchanger: According to the actual operation state of the thermal management system: the inlet refrigerant of the evaporator and plate heat exchanger is two-phase and the outlet refrigerant is superheated, the inlet refrigerant of the condenser is superheated and the outlet refrigerant is subcooled, the evaporator and plate heat exchanger are divided into two phase zones: two-phase zone and superheated zone, and the condenser is divided into three phase zones: superheated zone, two-phase zone and subcooled zone;

[0011] (2) Derive the heat transfer state equations for each phase of the evaporator, plate heat exchanger, and condenser;

[0012] (3) Establish the overall equation of the evaporator, condenser and plate heat exchanger:

[0013] (4) Establish a dynamic model of the thermal management system:

[0014] (5) Calculate the entire operating parameter space of the system under given working conditions: Input parameters into the established thermal management system dynamic model to calculate the operating parameter space of the thermal management system.

[0015] Preferably, the heat exchanger simplification process in step (1) is to simplify the multi-channel heat exchanger into a single-channel one-dimensional heat exchange pipeline according to the flow direction; calculate the total length L of the refrigerant in the tube along the flow direction, the cross-sectional area A, and the heat exchange perimeter s along the flow direction with the wall according to the heat exchanger structural parameters. r and the equivalent heat exchange perimeter s between the fluid outside the tube and the wall along the flow direction f The heat exchanger is equivalent to a one-dimensional heat exchange tube with a length of L. The refrigerant flow direction is defined as the +Z direction, and the fluid outside the tube exchanges heat with the refrigerant in countercurrent.

[0016] Preferably, the state equation of step (2) is obtained by integrating the conservation equation along the pipe according to the Leibniz law.

[0017] Preferably, for the two-phase region of the evaporator and the plate heat exchanger, the conservation equations adopt the conservation of mass, conservation of energy and conservation of tube wall energy, and the derived state equations are as follows:

[0018]

[0019] Among them, l e is the length of the two-phase zone, A is the cross-sectional area of ​​the pipe, s r is the heat exchange circumference along the flow direction and the wall, s f is the equivalent heat exchange circumference of the fluid outside the tube and the wall along the flow direction, is the expansion valve outlet flow, is the outlet flow rate of the two-phase zone, is the average density of the refrigerant in the two-phase region, ρ g and ρ l are the refrigerant saturated vapor density and saturated liquid density, α r is the heat transfer coefficient of the refrigerant in the tube, α f is the heat transfer coefficient of the fluid outside the tube, h l and h g are the specific enthalpies of the saturated liquid and vapor of the refrigerant, respectively. For the superheated zone, the conservation equation uses the conservation of tube wall energy, and the derived state equation is the same as equation (3), where ρ w , Cp w and Aw are the density, specific heat capacity and cross-sectional area of ​​the tube wall, respectively, e is the evaporation pressure, is the average void fraction of the refrigerant in the two-phase region, h oc is the specific enthalpy at the evaporator outlet, T w , T r and They are the tube wall temperature, the refrigerant temperature in the two-phase region and the average temperature of the fluid outside the tube.

[0020] Preferably, for the two-phase region of the condenser, the conservation equations use conservation of mass, conservation of energy and conservation of tube wall energy, and the derived state equations are respectively as follows:

[0021]

[0022]

[0023] in and are the refrigerant mass flow rates at the inlet and outlet of the two-phase region, A is the cross-sectional area of ​​the pipe, and h l and h g are the refrigerant saturated liquid and vapor specific enthalpies, α r is the heat transfer coefficient of the refrigerant in the tube, s r is the heat exchange perimeter with the wall along the flow direction; ρ g and ρ l are the saturated vapor density and saturated liquid density of the refrigerant, l c1 and l c2 are the lengths of the superheated zone and the two-phase zone, is the average void fraction of the refrigerant in the two-phase region; for the superheated and subcooled regions, the conservation equation uses the tube wall energy conservation, and the derived state equation is the same as equation (6), ρ w , Cp w and A w are the density, specific heat capacity and cross-sectional area of ​​the tube wall, α f is the heat transfer coefficient of the fluid outside the tube, s f is the equivalent heat exchange circumference of the fluid outside the tube and the wall along the flow direction, is the average density of the refrigerant in the two-phase region of the condenser, P c is the condensation pressure, T w , T r and They are the tube wall temperature, the refrigerant temperature in the two-phase region and the average temperature of the fluid outside the tube.

[0024] Preferably, the average density of the refrigerant in the two-phase region is obtained based on the assumption that the average void fraction obtained by integrating along the tube path is constant; the heat transfer coefficient is obtained through heat transfer experimental testing or through empirical correlation fitting; the density, specific enthalpy, kinematic viscosity, specific heat capacity and other physical properties of the refrigerant in the tube and the fluid outside the tube are queried in real time through the Refprop tool.

[0025] Preferably, step (3) is specifically: combining the state equations of each phase region to obtain the overall state space equation of the heat exchanger, wherein the state equations of the evaporator and the plate heat exchanger are:

[0026]

[0027] Among them, x e =[P e l e T we1 T we2 ] T is the defined state variable, and the input variable is M e and f e (x e ,u e ) The specific form of the elements in is obtained from the derived heat transfer equations of each phase region; is the defined state variable, l e is the length of the two-phase region, P e is the evaporation pressure, T we1 and T we2 are the wall temperatures of the two-phase zone and superheat zone of the evaporator, Represents the mass flow rate of refrigerant through the expansion valve, is the refrigerant mass flow rate at the evaporator outlet, is the mass flow rate of the external fluid, h rei is the specific enthalpy of the refrigerant at the evaporator inlet, T fei represents the external fluid inlet temperature;

[0028] The state equation of the condenser is:

[0029]

[0030] The state variable defined is x c =[P c l c1 l c2 T wc1 T wc2 T wc3 ] T , the input variables are M c and f c (xc ,u c ) is obtained from the derived heat transfer equations of each phase region, P c is the condensing pressure, l c1 and l c2 are the lengths of the overheating zone and the two-phase zone, T wc1 , T wc2 and T wc3 are the wall temperatures of the condenser superheated zone, two-phase zone and subcooled zone, respectively. is the mass flow rate of refrigerant at the condenser outlet, is the mass flow rate of refrigerant through the compressor, represents the mass flow rate of the external fluid, h rci is the specific enthalpy of the refrigerant at the condenser inlet, T fci is the external fluid temperature at the condenser inlet.

[0031] Preferably, step (4) is specifically: coupling the overall equations of the evaporator, condenser and plate heat exchanger, and combining the compressor and expansion to obtain a dynamic model of the thermal management system:

[0032]

[0033] In the steady-state equations for the compressor and expansion valve, and are the refrigerant mass flow rates through the compressor and expansion valve, respectively, and f com is the compressor frequency, V com is the compressor displacement, η com is the compressor volumetric efficiency, η is is the isentropic efficiency of the compressor, C v is the flow coefficient of the expansion valve, A v is the expansion valve flow cross-sectional area, ρ oe and ρ oc are the refrigerant densities at the outlets of the evaporator and condenser, respectively, h ic is the specific enthalpy of the refrigerant at the condenser inlet, h oe is the specific enthalpy of the refrigerant at the evaporator outlet, Δh is is the enthalpy increment calculated by the compressor according to isentropic compression, P c is the condensing pressure, P e is the evaporation pressure.

[0034] Preferably, the inlet and outlet refrigerant specific enthalpies of the expansion valve in step (4) are assumed to be equal.

[0035] Preferably, the input parameters in step (6) include compressor speed, condenser inlet air temperature, mass flow rate of air passing through the condenser, flow area of ​​the expansion valve port, evaporator inlet air temperature and mass flow rate, inlet temperature and mass flow rate of the water side of the plate heat exchanger, and battery heating power and temperature; the system initial conditions need to be specified: refrigerant mass flow rate through the compressor, the expansion valve before the evaporator and the expansion valve before the plate heat exchanger, evaporation pressure and condensation pressure, two-phase zone length and tube wall temperature of each phase zone of the evaporator and the plate heat exchanger, superheat zone length, two-phase zone length and tube wall temperature of each phase zone of the condenser;

[0036] The compressor speed and the heat exchanger inlet air temperature are measured by corresponding sensors, and the mass flow rate on the air / water side is obtained by converting the fan / water pump speed.

[0037] The heat exchanger simplification process in step (1) is to simplify the multi-channel heat exchanger into a single-channel one-dimensional heat exchange pipeline according to the flow direction.

[0038] The state equations in step (2) and step (3) are obtained by integrating the conservation equations along the pipe according to the Leibniz law.

[0039] The average density of the refrigerant in the two-phase region in the step (2) and the step (3) is obtained based on the assumption that the average void fraction obtained by integrating along the tube path is constant.

[0040] The heat transfer coefficients in step (2) and step (3) are obtained through heat transfer experimental tests or through fitting of empirical correlation equations.

[0041] In the step (4), the physical parameters such as density, specific enthalpy, kinematic viscosity, specific heat capacity, etc. of the refrigerant in the tube and the fluid outside the tube can be obtained by query.

[0042] In step (5), the inlet and outlet refrigerant specific enthalpies of the expansion valve are assumed to be equal.

[0043] The compressor speed and the heat exchanger inlet air temperature in step (6) are measured by corresponding sensors, and the mass flow rate on the air / water side is obtained by converting the fan / water pump speed.

[0044] Compared with the prior art, the present invention has the following beneficial effects:

[0045] 1) The thermal management model established based on the moving boundary method takes into account the different heat transfer processes in different phase zones of the heat exchanger, and its accuracy is improved; the established thermal management model is used to design and optimize the thermal management system of electric vehicles, effectively reducing the number of tests and saving costs.

[0046] 2) The model has a low order and specifically removes the dynamic equations of the compressor and expansion valve with short response time, and uses steady-state equations instead of calculations, which saves computing costs while retaining accuracy, provides conditions for the intervention of system control, and is conducive to the implementation of complex control algorithms.

[0047] 3) Since the heat exchange model is derived on the basis of the basic conservation equations and is universal, the thermal management system modeling method can be applied to the modeling of most types of heat exchangers, and can also be applied to the modeling of most electric vehicle thermal management systems coupled with air conditioning systems and battery cooling systems. BRIEF DESCRIPTION OF THE DRAWINGS

[0048] Figure 1a It is a schematic diagram of the summer operation of a thermal management system for an electric vehicle;

[0049] Figure 1b It is a schematic diagram of winter operation of a thermal management system for an electric vehicle;

[0050] Figure 2 It is a schematic diagram of the phase division of the evaporator and the plate heat exchanger;

[0051] Figure 3 It is a schematic diagram of the phase division of the condenser;

[0052] Figure 4 It is a schematic diagram of the coupling mode of various components in the thermal management system;

[0053] Figure 5 It is a thermal management system comparison model used for verification, built using standardized modeling software;

[0054] Figure 6a It is the evaporation pressure verification result of the thermal management model;

[0055] Figure 6b It is the condensation pressure verification result of the thermal management model;

[0056] Figure 6c It is the verification result of the evaporator heat transfer of the thermal management model;

[0057] Figure 6d It is the verification result of the condenser heat transfer of the thermal management model;

[0058] Figure 7a It is the dynamic verification result of the evaporation pressure of the thermal management model under the compressor step input;

[0059] Figure 7b It is the dynamic verification result of the thermal management model under the compressor step input before the evaporator expansion valve flow. DETAILED DESCRIPTION

[0060] The modeling method of the electric vehicle thermal management system of the present invention is realized by simulation through dynamic system modeling software. At the same time, in order to verify the accuracy of the modeling method, a verification model is built using standardized modeling software for refrigeration systems.

[0061] In order to make the purpose, technical solutions and advantages of the present invention more clear and explicit, the present invention is further described in detail below with reference to the accompanying drawings and embodiments.

[0062] The schematic diagram of the electric vehicle thermal management system described in this embodiment is as follows Figure 1a and Figure 1b As shown, the system includes: compressor 1, in-vehicle heat exchanger 2, out-vehicle heat exchanger 3, plate heat exchanger 4, four-way reversing valve 5, liquid collector 6, main circuit expansion valve 7, summer plate exchange expansion valve 8, winter plate exchange expansion valve 9 and battery cooling system 10. System input signals include: compressor 1 speed, inlet air temperature of in-vehicle heat exchanger 2, mass flow rate of air passing through in-vehicle heat exchanger 2, valve port flow area of ​​main circuit expansion valve 7, summer plate exchange expansion valve 8 and winter plate exchange expansion valve 9, inlet air temperature and mass flow rate of out-vehicle heat exchanger 3, inlet temperature and mass flow rate of water side of plate heat exchanger 4, heating power and temperature of battery cooling system 10.

[0063] The main working principle of this thermal management system is similar to that of a vapor compression refrigeration cycle: The working principle of the summer system is as follows Figure 1a As shown, the system absorbs the heat from the cabin and the battery through the in-car heat exchanger 2 and the plate heat exchanger 4 and discharges it to the outside to cool the cabin and the battery pack. At this time, the in-car heat exchanger 1 is the evaporator and the outdoor heat exchanger 3 is the condenser. The working principle of the winter system is as follows Figure 1b As shown, the system absorbs heat from the environment and the battery pack through the outdoor heat exchanger 3 and the plate heat exchanger 4 to heat the cabin and utilize the waste heat of the battery. At this time, the indoor heat exchanger 2 is a condenser and the outdoor heat exchanger 3 is an evaporator.

[0064] The electric vehicle thermal management system modeling method based on the moving boundary method in this embodiment includes the following steps:

[0065] (1) Simplified one-dimensional model of evaporator, condenser and plate heat exchanger:

[0066] The heat exchanger is equivalent to a one-dimensional heat exchange tube with a tube length of L. The refrigerant flow direction is defined as the +Z direction, and the fluid outside the tube exchanges heat with the refrigerant in countercurrent. The total length of the heat exchange tube in the heat exchanger is simplified to the total length L of the refrigerant in the one-dimensional tube along the flow direction, the total cross-sectional area of ​​the refrigerant side of a single tube is simplified to the cross-sectional area A of the one-dimensional heat exchange tube, and the total heat exchange perimeter of the refrigerant side of a single tube is simplified to the heat exchange perimeter s of the one-dimensional heat exchange tube. r , the equivalent heat exchange perimeter s of the fluid outside the heat exchange tube along the flow direction and the wall is calculated according to the ratio of the heat exchange area to the length f .

[0067] According to the working state of the thermal management system in winter, the heat exchanger is analyzed: the inlet refrigerant of the evaporator and plate heat exchanger is two-phase, and the outlet refrigerant is superheated, and the inlet refrigerant of the condenser is superheated, and the outlet refrigerant is subcooled. Therefore, the evaporator and plate heat exchanger are divided into two phase zones: two-phase zone and superheated zone, and the condenser is divided into three phase zones: superheated zone, two-phase zone and subcooled zone. Figure 2 As shown, the condenser phase area is divided as follows Figure 3 shown.

[0068] (2) Derive the heat transfer equations for each phase of the evaporator and plate heat exchanger:

[0069] There are three types of conservation equations for heat exchangers: mass conservation, energy conservation, and tube wall energy conservation, which can be expressed as:

[0070]

[0071] In the conservation equation, A is the flow cross-sectional area in the tube; α r is the heat transfer coefficient of the refrigerant in the tube; s r is the circumference of the refrigerant heat exchange cross section; ρ, h, P and T r are the density, mass flow rate, specific enthalpy, pressure and temperature of the refrigerant, T w is the average temperature of the pipe wall, m w is the wall mass, Cp w is the specific heat capacity of the tube wall, l is the tube length, α f is the heat transfer coefficient of the fluid outside the tube, t represents time, z represents the length coordinate along the length of the tube, s f is the circumference of the heat transfer section of the fluid outside the tube, T f is the average temperature of the fluid outside the tube.

[0072] The above three conservation equations contain unsolvable differential terms, which can be transformed into soluble differential terms by integrating along the pipeline using the Leibniz rule:

[0073]

[0074] In formula (4), f(z,t) represents a function with length coordinate z and time variable t, z 1 (t) represents the upper limit of the length coordinate integral related to time, z 2 (t) represents the lower limit of the length coordinate integral related to time, f(z 2 (t), t) represents the value of the function at the lower limit of the length coordinate integral, f(z 1 (t), t) represents the value of the function at the upper limit of the length coordinate integral.

[0075] For the two-phase region, the conservation equations use mass conservation, energy conservation and wall energy conservation, and the derived state equations are as follows:

[0076]

[0077] In the above formula, l e is the length of the two-phase region, is the expansion valve outlet flow, is the outlet flow rate of the two-phase zone, is the average density of the refrigerant in the two-phase region, ρ g and ρ l are the refrigerant saturated vapor density and saturated liquid density, α r is the heat transfer coefficient of the refrigerant in the tube, α f is the heat transfer coefficient of the fluid outside the tube, h oc is the specific enthalpy of the refrigerant at the outlet of the expansion valve, h l and h g are the refrigerant saturated liquid and vapor specific enthalpies, ρ w , Cp w and A w are the density, specific heat capacity and cross-sectional area of ​​the tube wall, respectively, e is the evaporation pressure, is the average void fraction of the refrigerant in the two-phase region, h oc is the specific enthalpy at the evaporator outlet, T w 、T r and are the tube wall temperature, the refrigerant temperature in the two-phase region and the average temperature of the fluid outside the tube respectively;

[0078] For the overheating zone, the conservation equation uses the tube wall energy conservation, and the derived state equation is the same as equation (7);

[0079] Heat transfer coefficient α of the heat exchanger r and α f It can be measured experimentally or obtained through empirical correlation. For empirical correlation, the heat transfer coefficient of the two-phase zone is calculated using the Chen heat transfer correlation, and the heat transfer coefficient of the superheated zone and the fluid outside the tube is calculated using the Dittus-Bolter heat transfer correlation.

[0080] (3) Derive the heat transfer equations for each phase of the condenser:

[0081] For the two-phase region, the conservation equations use mass conservation, energy conservation and wall energy conservation, and the derived state equations are shown in equations (8)-(10), respectively, where and are the refrigerant mass flow rates at the inlet and outlet of the two-phase region, l c1 and l c2are the lengths of the superheated zone and the two-phase zone, is the average void fraction of the refrigerant in the two-phase region. Assuming that the average void fraction obtained by integrating along the tube path is constant, is the average density of the refrigerant in the two-phase region of the condenser, P c is the condensation pressure, T w , T r and are the tube wall temperature, the refrigerant temperature in the two-phase zone, and the average temperature of the fluid outside the tube, respectively. For the superheated zone and the subcooled zone, the conservation equation adopts the tube wall energy conservation, and the derived state equation is the same as equation (9).

[0082] Heat transfer coefficient α of the heat exchanger r and α f Obtained through empirical correlation. For the empirical correlation, the heat transfer coefficient of the two-phase zone of the condenser is calculated using the Shah heat transfer correlation, and the heat transfer coefficient of the superheated zone, subcooled zone and the fluid outside the tube is calculated using the Dittus-Bolter heat transfer correlation.

[0083]

[0084] in and are the refrigerant mass flow rates at the inlet and outlet of the two-phase region, A is the cross-sectional area of ​​the pipe, and h l and h g are the refrigerant saturated liquid and vapor specific enthalpies, α r is the heat transfer coefficient of the refrigerant in the tube, s r is the heat exchange perimeter with the wall along the flow direction; ρ g and ρ l are the saturated vapor density and saturated liquid density of the refrigerant, l c1 and l c2 are the lengths of the superheated zone and the two-phase zone, is the average void fraction of the refrigerant in the two-phase region; for the superheated and subcooled regions, the conservation equation uses the tube wall energy conservation, and the derived state equation is the same as equation (6), ρ w , Cp w and A w are the density, specific heat capacity and cross-sectional area of ​​the tube wall, α f is the heat transfer coefficient of the fluid outside the tube, s f is the equivalent heat exchange circumference of the fluid outside the tube and the wall along the flow direction, is the average density of the refrigerant in the two-phase region of the condenser, P c is the condensation pressure, T w , T r and They are the tube wall temperature, the refrigerant temperature in the two-phase region and the average temperature of the fluid outside the tube.

[0085] (4) Establish the overall equation of the evaporator, condenser and plate heat exchanger:

[0086] The state equations of each phase region are combined to obtain the overall state space equation of the heat exchanger. The state space equations of the evaporator and the plate heat exchanger are similar and can be described by equation (11). The defined state variable is x e =[P e l e T we1 T we2 ] T , the input variables are M e and f e (x e ,u e The specific form of the elements in ) can be obtained from the derived heat transfer equations of each phase region. is the defined state variable, l e is the length of the two-phase region, P e is the evaporation pressure, T we1 and T we2 are the wall temperatures of the two-phase zone and superheat zone of the evaporator, Represents the mass flow rate of refrigerant through the expansion valve, is the refrigerant mass flow rate at the evaporator outlet, is the mass flow rate of the external fluid, h rei is the specific enthalpy of the refrigerant at the evaporator inlet, T fei represents the external fluid inlet temperature;

[0087]

[0088] The state equation of the condenser can be described by equation (12), where the state variable is defined as x c =[P c l c1 l c2 T wc1 T wc2 T wc3 ] T , the input variables are M c and f c (x c ,u c ) can be obtained from the derived heat transfer equations of each phase region, where P c is the condensing pressure, l c1 and l c2 are the lengths of the overheating zone and the two-phase zone, T wc1 , T wc2 and T wc3are the wall temperatures of the condenser superheated zone, two-phase zone and subcooled zone, respectively. is the mass flow rate of refrigerant at the condenser outlet, is the mass flow rate of refrigerant through the compressor, represents the mass flow rate of the external fluid, h rci is the specific enthalpy of the refrigerant at the condenser inlet, T fci is the external fluid temperature at the condenser inlet.

[0089]

[0090] The status parameters of the refrigerant side and air side / water side of the heat exchanger can be obtained by querying the Refprop software.

[0091] (5) Establish a dynamic model of the thermal management system:

[0092] By coupling the overall equations of the evaporator, condenser and plate heat exchanger, and combining the steady-state equations of the compressor and expansion valve as shown in equations (13)-(15), the dynamic model of the thermal management system can be obtained, in which the coupling relationship between the components is as follows: Figure 4 As shown. The coupling model includes: compressor model 21, condenser model 22, evaporator model 23, plate heat exchanger model 24, evaporator front expansion valve model 25, plate heat exchanger front expansion valve model 26. Among them, the flow of compressor model 21, evaporator front expansion valve model 25 and plate heat exchanger front expansion valve model 26 are jointly determined by the evaporation pressure and the condensation pressure; the evaporator model 23 and the plate heat exchanger model 24 have a common node at the outlet, which represents a common evaporation pressure. Therefore, the mass conservation equation of the plate heat exchanger model 24 is used as the equation closure condition to participate in the thermal management system model calculation, rather than the state equation. The steady-state equations of the compressor and expansion valve are introduced because the response time of the compressor and the expansion valve is very small compared to the heat exchanger. Only the steady-state equations are needed to provide the flow calculation and the specific enthalpy increment of the refrigerant after passing through the compressor. The evaporation pressure is jointly determined by the state equations of the evaporator and the plate heat exchanger; the condensation pressure is determined by the state equation of the condenser. In the steady-state equations of the compressor and the expansion valve, and are the refrigerant mass flow rate through the compressor and expansion valve, f com is the compressor frequency, V com is the compressor displacement, η com is the compressor volumetric efficiency, Δh is is the enthalpy increment calculated by the compressor according to isentropic compression, η is is the isentropic efficiency of the compressor, C v is the flow coefficient of the expansion valve, A v is the expansion valve flow cross-sectional area, ρ oe and ρoc are the refrigerant densities at the outlets of the evaporator and condenser, respectively, h ic is the specific enthalpy of the refrigerant at the condenser inlet, h oe is the specific enthalpy of refrigerant at the evaporator outlet, P c is the condensing pressure, P e is the evaporation pressure. The system model is built through dynamic system modeling software.

[0093]

[0094] In the steady-state equations for the compressor and expansion valve, and are the refrigerant mass flow rates through the compressor and expansion valve, respectively, and f com is the compressor frequency, V com is the compressor displacement, η com is the compressor volumetric efficiency, η is is the isentropic efficiency of the compressor, C v is the flow coefficient of the expansion valve, A v is the expansion valve flow cross-sectional area, ρ oe and ρ oc are the refrigerant densities at the outlets of the evaporator and condenser, respectively, h ic is the specific enthalpy of the refrigerant at the condenser inlet, h oe is the specific enthalpy of the refrigerant at the evaporator outlet, Δh is is the enthalpy increment calculated by the compressor according to isentropic compression, P c is the condensing pressure, P e is the evaporation pressure.

[0095] (6) Calculate the entire operating parameter space of the system under given working conditions:

[0096] The established dynamic model of the thermal management system is used for calculation. The input parameters are the speed of compressor 1, the inlet air temperature of the in-vehicle heat exchanger 2, the mass flow rate of the air through the in-vehicle heat exchanger 2, the valve port flow area of ​​the main circuit expansion valve 7, the summer plate exchange expansion valve 8 and the winter plate exchange expansion valve 9, the inlet air temperature and mass flow rate of the outdoor heat exchanger 3, the inlet temperature and mass flow rate of the water side of the plate heat exchanger 4, and the heating power and temperature of the battery cooling system 10. The initial conditions of the system need to be specified: the mass flow rate of the refrigerant through the compressor 1, the main circuit expansion valve 7, the summer plate exchange expansion valve 8 and the winter plate exchange expansion valve 9; the evaporation pressure and the condensation pressure; the two-phase zone length and the pipe wall temperature of each phase zone of the outdoor heat exchanger 3 and the plate heat exchanger 4; the superheat zone length, the two-phase zone length and the pipe wall temperature of each phase zone of the in-vehicle heat exchanger 2. By setting the input parameters and initial conditions, the operating parameter space of the thermal management system can be calculated.

[0097] A comparison model of the electric vehicle thermal management system is established. The comparison model is built using standardized modeling software. When building the comparison model, it is not necessary to consider the calculation principle inside the components. It is only necessary to give the structural parameters of the components and the connection relationship between the components. The comparison model established is required to have the same component structural parameters and connection relationship as the model established by the modeling method. Figure 5 As shown, it is used to verify the accuracy of the modeling method.

[0098] Figure 5 The comparison model shown is composed of a compressor 31, an in-vehicle heat exchanger 32, an out-vehicle heat exchanger 33, a plate heat exchanger 34, a four-way reversing valve 35, a liquid collector 36, a main circuit expansion valve 37, a summer plate heat exchanger expansion valve 38, a winter plate heat exchanger expansion valve 39 and a battery cooling system 40. The connection relationship between the components is as follows Figure 1a The only difference is the spatial layout.

[0099] Given stable working conditions: compressor speed is 1000-1200rpm, condenser inlet air temperature is 10-30℃, condenser inlet air mass flow rate is 0.2kg / s, evaporator inlet air temperature is 10-20℃, evaporator inlet air mass flow rate is 0.3kg / s, battery heating power is 700W, and battery temperature is 25℃. The comparison results are as follows Figure 6a , Figure 6b He Ru Figure 6c , Figure 6d As shown, the steady-state accuracy of the model is verified.

[0100] Given a step input: the system runs stably for 100s at a compressor speed of 1200rpm, then the compressor speed is stepped to 1000rpm and runs for 150s, and finally the compressor speed is stepped to 1200rpm until the end. The comparison results are as follows Figure 7a and Figure 7b As shown, the dynamic accuracy of the model is verified.

[0101] The preferred embodiments of the present invention disclosed above are only used to help illustrate the present invention. The preferred embodiments do not describe all the details in detail, nor do they limit the invention to the specific implementation methods described. Obviously, many modifications and changes can be made according to the content of this specification. This specification selects and specifically describes these embodiments in order to better explain the principles and practical applications of the present invention, so that those skilled in the art can understand and use the present invention well. The present invention is limited only by the claims and their full scope and equivalents.

Claims

1. A modeling method for electric vehicle thermal management system based on moving boundary method, characterized in that: The following steps are involved: (1) Simplified one-dimensional model of evaporator, condenser and plate heat exchanger: According to the actual operation state of the thermal management system: the inlet refrigerant of the evaporator and plate heat exchanger is two-phase and the outlet refrigerant is superheated, the inlet refrigerant of the condenser is superheated and the outlet refrigerant is subcooled, the evaporator and plate heat exchanger are divided into two phase zones: two-phase zone and superheated zone, and the condenser is divided into three phase zones: superheated zone, two-phase zone and subcooled zone; (2) Derive the state equations for heat transfer in each phase of the evaporator, plate heat exchanger, and condenser. The state equations are obtained by integrating the conservation equations along the tube path according to the Leibniz law; (3) Establish the overall equations of the evaporator, condenser and plate heat exchanger, combine the state equations of each phase region, and obtain the overall state space equation of the heat exchanger, where the state equations of the evaporator and plate heat exchanger are: Among them, x e =[P e l e T we1 T we2 ] T is the defined state variable, and the input variable is M e and f e (x e ,u e ) The specific form of the elements in is obtained from the derived heat transfer equations of each phase region; is the defined state variable, l e is the length of the two-phase region, P e is the evaporation pressure, T we1 and T we2 are the wall temperatures of the two-phase zone and superheat zone of the evaporator, Represents the mass flow rate of refrigerant through the expansion valve, is the refrigerant mass flow rate at the evaporator outlet, is the mass flow rate of the external fluid, h rei is the specific enthalpy of the refrigerant at the evaporator inlet, T fei represents the external fluid inlet temperature; The state equation of the condenser is: The state variable defined is x c =[P c l c1 l c2 T wc1 T wc2 T wc3 ] T , the input variables are M c and f c (x c ,u c ) is obtained from the derived heat transfer equations of each phase region, P c is the condensing pressure, l c1 and l c2 are the lengths of the overheating zone and the two-phase zone, T wc1 、T wc2 and T wc3 are the wall temperatures of the condenser superheated zone, two-phase zone and subcooled zone, respectively. is the mass flow rate of refrigerant at the condenser outlet, is the mass flow rate of refrigerant through the compressor, represents the mass flow rate of the external fluid, h rci is the specific enthalpy of the refrigerant at the condenser inlet, T fci is the external fluid temperature at the condenser inlet; (4) Establishing a dynamic model of the thermal management system. Step (4) is specifically as follows: coupling the overall equations of the evaporator, condenser and plate heat exchanger, and combining the compressor and expansion to obtain the dynamic model of the thermal management system: In the steady-state equations for the compressor and expansion valve, and are the refrigerant mass flow rates through the compressor and expansion valve, respectively, and f com is the compressor frequency, V com is the compressor displacement, η com is the compressor volumetric efficiency, η is is the isentropic efficiency of the compressor, C v is the flow coefficient of the expansion valve, A v is the expansion valve flow cross-sectional area, ρ oe and ρ oc are the refrigerant densities at the evaporator and condenser outlets, h ic is the specific enthalpy of the refrigerant at the condenser inlet, h oe is the specific enthalpy of the refrigerant at the evaporator outlet, Δh is is the enthalpy increment calculated by the compressor according to isentropic compression, P c is the condensing pressure, P e is the evaporation pressure; (5) Calculate the entire operating parameter space of the system under given working conditions: Input parameters into the established thermal management system dynamic model to calculate the operating parameter space of the thermal management system.

2. The electric vehicle thermal management system modeling method based on the moving boundary method according to claim 1 is characterized in that: The heat exchanger simplification process in step (1) is to simplify the multi-channel heat exchanger into a single-channel one-dimensional heat exchange pipeline according to the flow direction; according to the heat exchanger structural parameters, the total length L of the refrigerant in the tube along the flow direction, the cross-sectional area A, and the heat exchange perimeter s along the flow direction with the wall are calculated. r and the equivalent heat exchange perimeter s between the fluid outside the tube and the wall along the flow direction f ; The heat exchanger is equivalent to a one-dimensional heat exchange tube with a tube length of L. The refrigerant flow direction is defined as the +Z direction, and the fluid outside the tube exchanges heat with the refrigerant in countercurrent.

3. The electric vehicle thermal management system modeling method based on the moving boundary method according to claim 1 is characterized in that: For the two-phase region of the evaporator and plate heat exchanger, the conservation equations use mass conservation, energy conservation and tube wall energy conservation, and the derived state equations are as follows: Among them, l e is the length of the two-phase zone, A is the cross-sectional area of ​​the pipe, s r is the heat exchange circumference along the flow direction and the wall, s f is the equivalent heat exchange circumference of the fluid outside the tube and the wall along the flow direction, is the expansion valve outlet flow, is the outlet flow rate of the two-phase zone, is the average density of the refrigerant in the two-phase region, ρ g and ρ l are the refrigerant saturated vapor density and saturated liquid density, α r is the heat transfer coefficient of the refrigerant in the tube, α f is the heat transfer coefficient of the fluid outside the tube, h l and h g are the specific enthalpies of the saturated liquid and vapor of the refrigerant, respectively. For the superheated zone, the conservation equation uses the tube wall energy conservation, and the derived state equation is the same as equation (3), where ρ w , Cp w and A w are the density, specific heat capacity and cross-sectional area of ​​the tube wall, respectively, e is the evaporation pressure, is the average void fraction of the refrigerant in the two-phase region, h oc is the specific enthalpy at the evaporator outlet, T w 、T r and They are the tube wall temperature, the refrigerant temperature in the two-phase region and the average temperature of the fluid outside the tube.

4. The electric vehicle thermal management system modeling method based on the moving boundary method according to claim 1 is characterized in that: For the two-phase region of the condenser, the conservation equations use mass conservation, energy conservation and tube wall energy conservation, and the derived state equations are as follows: in and are the refrigerant mass flow rates at the inlet and outlet of the two-phase region, A is the cross-sectional area of ​​the pipe, and h l and h g are the refrigerant saturated liquid and vapor specific enthalpies, α r is the heat transfer coefficient of the refrigerant in the tube, s r is the heat exchange perimeter with the wall along the flow direction; ρ g and ρ l are the saturated vapor density and saturated liquid density of the refrigerant, l c1 and l c2 are the lengths of the superheated zone and the two-phase zone, is the average void fraction of the refrigerant in the two-phase region; for the superheated and subcooled regions, the conservation equation uses the tube wall energy conservation, and the derived state equation is the same as equation (6), ρ w , Cp w and A w are the density, specific heat capacity and cross-sectional area of ​​the tube wall, α f is the heat transfer coefficient of the fluid outside the tube, s f is the equivalent heat exchange circumference of the fluid outside the tube and the wall along the flow direction, is the average density of the refrigerant in the two-phase region of the condenser, P c is the condensation pressure, T w 、T r and They are the tube wall temperature, the refrigerant temperature in the two-phase region and the average temperature of the fluid outside the tube.

5. The electric vehicle thermal management system modeling method based on the moving boundary method according to claim 3 or 4 is characterized in that: The average density of the refrigerant in the two-phase region is obtained based on the assumption that the average void fraction obtained by integrating along the tube path is constant; the heat transfer coefficient is obtained through heat transfer experimental tests or through empirical correlation fitting; the density, specific enthalpy, kinematic viscosity, and specific heat capacity physical property parameters of the refrigerant in the tube and the fluid outside the tube are all queried in real time through the Refprop tool.

6. The electric vehicle thermal management system modeling method based on the moving boundary method according to claim 5 is characterized in that: In step (4), the inlet and outlet refrigerant specific enthalpies of the expansion valve are assumed to be equal.

7. The electric vehicle thermal management system modeling method based on the moving boundary method according to claim 1 is characterized in that: The input parameters in step (5) include the compressor speed, the condenser inlet air temperature, the mass flow rate of the air passing through the condenser, the flow area of ​​the expansion valve port, the evaporator inlet air temperature and mass flow rate, the inlet temperature and mass flow rate of the water side of the plate heat exchanger, and the battery heating power and temperature; the system initial conditions need to be specified: the refrigerant mass flow rate through the compressor, the expansion valve before the evaporator and the expansion valve before the plate heat exchanger, the evaporation pressure and the condensation pressure, the two-phase zone length and the tube wall temperature of each phase zone of the evaporator and the plate heat exchanger, the superheat zone length of the condenser, the two-phase zone length and the tube wall temperature of each phase zone; The compressor speed and the heat exchanger inlet air temperature are measured by corresponding sensors, and the mass flow rate on the air / water side is obtained by converting the fan / water pump speed.

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