A self-temperature balancing battery pack structure and its adaptation verification method

Abstract

A self-equalizing battery pack structure and its adaptation verification method are disclosed. Currently, due to insufficient heat exchange between the battery and the external environment caused by ambient temperature, it is difficult to accurately control the temperature of the battery pack, and the temperature control performance is difficult to quantify, estimate, and configure. In this invention, an upper groove is machined along the length of one end of the heat sink, and an upper liquid metal layer is disposed within the upper groove. The top surface of the upper liquid metal layer is flush with the upper side of the heat sink. A lower groove is machined along the length of the lower side of the heat sink, located directly below the upper groove. A lower liquid metal layer is disposed within the lower groove, and the bottom surface of the lower liquid metal layer is flush with the lower side of the heat sink. Multiple upper hollow fins and multiple lower hollow fins are integrally connected to the other end of the heat sink. The upper hollow fins are evenly arranged along the length of the heat sink on the upper side, and the lower hollow fins are evenly arranged along the length of the heat sink on the lower side.

Description

Technical Field

[0001] This invention specifically relates to a self-equalizing battery pack structure and its adaptation verification method. Background Technology

[0002] As a high-energy-density energy storage device, lithium-ion batteries have been widely used in mobile electronic devices and electric vehicles. The rapid development of new energy vehicles is accelerating their replacement of traditional gasoline-powered vehicles. However, regardless of whether it's lithium iron phosphate or ternary lithium, lithium-ion batteries will perform poorly at excessively high or low temperatures, causing issues such as insufficient energy supply, rapid degradation, and even various safety problems. It is generally believed in the industry that the operating temperature range of lithium batteries should be controlled between 25℃ and 45℃. Temperatures above or below this range will lead to a decrease in battery charge / discharge performance and cycle life, and may even cause thermal runaway. When an electric vehicle is undergoing prolonged charging and discharging, the temperature inside the battery pack will continuously rise. Especially when the battery is discharging at a high rate, the temperature rises sharply. Since the internal space of the battery pack is limited, if effective heat dissipation measures are not taken, battery performance and lifespan will be reduced, and in severe cases, smoke, spontaneous combustion, and explosion may occur. Therefore, it is necessary to incorporate a battery cooling system to dissipate heat and ensure that the lithium-ion power battery remains within a safe temperature range during operation. This is of significant practical importance for improving vehicle range and driving safety.

[0003] Currently, thermal management is broadly divided into liquid cooling and air cooling. Although liquid cooling has a stronger temperature control capability than air cooling, it is more expensive and difficult to maintain. Air cooling is currently the most commonly used thermal management method. However, due to the ambient temperature, the heat exchange capacity between the battery and the outside environment is insufficient to support accurate control of the battery pack temperature, resulting in poor temperature control and reduced vehicle range.

[0004] Vapor chambers offer significant advantages in heat transfer efficiency. However, traditional vapor chambers have long condensation sections, requiring substantial installation space and making them unsuitable for the compact design of automotive battery packs. This patent proposes an irregularly shaped vapor chamber with a hollow fin structure in the condensation section. This design shortens the condensation section while maintaining the battery pack's maximum temperature and temperature uniformity within a reasonable range, reducing the extension length to 30% of the total length and optimizing the spatial layout. Furthermore, current vapor chamber simulations largely involve phase transition processes. Phase transition simulations require high mesh fineness and small time steps to avoid computational divergence, often resulting in lengthy computation times. Therefore, this invention proposes an optimized method for vapor chamber simulation, shortening the simulation cycle while maintaining accuracy. Summary of the Invention

[0005] To overcome the shortcomings of existing technologies, a self-equalizing battery pack structure and its adaptation verification method are provided to solve the above problems.

[0006] A wind-cooled heat spreader includes a heat dissipation plate body, an upper liquid metal layer, a lower liquid metal layer, an outer protective layer, multiple upper hollow fins, and multiple lower hollow fins. The heat dissipation plate body is a strip-shaped plate body. An upper groove is machined along the length of one end of the heat dissipation plate body, and an upper liquid metal layer is disposed in the upper groove. The top surface of the upper liquid metal layer is flush with the upper side of the heat dissipation plate body. A lower groove is machined along the length of the lower side of the heat dissipation plate body, and the lower groove is located directly below the upper groove. A lower liquid metal layer is disposed in the lower groove. The bottom surface of the lower liquid metal layer is flush with the lower side of the heat sink body. The other end of the heat sink body is integrally connected with multiple upper hollow fins and multiple lower hollow fins. The multiple upper hollow fins are evenly arranged on the upper side of the heat sink body along the length direction of the heat sink body, and the multiple lower hollow fins are evenly arranged on the lower side of the heat sink body along the length direction of the heat sink body. The upper hollow fins and lower hollow fins are arranged in a one-to-one correspondence. An outer protective layer is provided on the outer wall of the heat sink body, the multiple upper hollow fins and the multiple lower hollow fins.

[0007] As a preferred embodiment: the heat dissipation plate includes a condensing section, a steam section, and an evaporating section. A condensing cavity is machined along the length of the condensing section, a steam cavity is machined along the length of the steam section, and an evaporating cavity is machined along the length of the evaporating section. The condensing section is connected to one end of the steam cavity in the steam section through the condensing cavity, and the other end of the steam cavity is connected to one end of the evaporating cavity in the evaporating section. The plurality of upper hollow fins are integrally connected to the upper side of the condensing section, and each upper hollow fin is connected to the condensing cavity. The plurality of lower hollow fins are integrally connected to the lower side of the condensing section, and each lower hollow fin is connected to the condensing cavity.

[0008] As a preferred option, the length of the condensing section is less than or equal to one-third of the length of the evaporating section, and the length of the steam section is less than or equal to one-fifth of the length of the condensing section.

[0009] As a preferred embodiment: each upper hollow fin is provided with a first cavity, and each lower hollow fin is provided with a second cavity.

[0010] A self-heating battery pack structure, comprising a wind-cooled heat spreader as described in embodiments one, two, three, or four, includes a ventilation cover, a socket, and a self-heating battery pack structure. The socket is horizontally positioned, and the self-heating battery pack structure is disposed inside the socket. The bottom of the self-heating battery pack structure is detachably connected to the socket. The ventilation cover is disposed on the socket, and the top of the self-heating battery pack structure is detachably connected to the ventilation cover. An air duct is provided on the inner wall of the ventilation cover, and an air outlet and an air inlet are respectively machined on the ventilation cover. The air outlet and air inlet are respectively connected to the air duct. The self-heating battery pack structure is composed of multiple heat spreader vertical plates and multiple battery packs. The multiple heat spreader vertical plates are arranged vertically side by side, and multiple battery packs are disposed between every two adjacent heat spreader vertical plates. Each heat spreader vertical plate is formed by multiple wind-cooled heat spreaders arranged vertically.

[0011] Each air-cooled vapor chamber includes a heat sink body, an upper liquid metal layer, a lower liquid metal layer, an outer protective layer, multiple upper hollow fins, and multiple lower hollow fins. The heat sink body is a strip-shaped plate. An upper groove is machined along the length of one end of the heat sink body, and the upper liquid metal layer is placed within the upper groove. The top surface of the upper liquid metal layer is flush with the upper side of the heat sink body. A lower groove is machined along the length of the lower side of the heat sink body, directly below the upper groove. A lower liquid metal layer is placed within the lower groove, and the bottom surface of the lower liquid metal layer is flush with the lower side of the heat sink body. The heat sink is flush with the heat sink, and multiple upper hollow fins and multiple lower hollow fins are integrally connected to the other end of the heat sink. The upper hollow fins are evenly arranged on the upper side of the heat sink along the length of the heat sink, and the lower hollow fins are evenly arranged on the lower side of the heat sink along the length of the heat sink. The upper hollow fins and lower hollow fins are arranged in a one-to-one correspondence. An outer protective layer is wrapped around the outer wall of the heat sink, the upper hollow fins and the lower hollow fins. The upper hollow fins and lower hollow fins in each air-cooled heat sink face the ventilation cover.

[0012] As a preferred option: the air outlet includes multiple circular openings, which are sequentially machined along the width of the ventilation cover, and each circular opening contains a fan.

[0013] A method for adapting and verifying a self-temperature-equalizing battery pack structure, used to form the self-temperature-equalizing battery pack structure described in specific embodiments five or six, characterized by comprising the following steps:

[0014] S1. Modeling: Use modeling software to create a 3D model of the air-cooled heat exchanger;

[0015] S2. Determination of Optimization Parameters: The total number of upper and lower hollow fins is taken as optimization parameter A, the width of the upper hollow fins is taken as optimization parameter B, the length of the upper hollow fins is taken as optimization parameter C, and the horizontal angle of the upper hollow fins is taken as optimization parameter D. The average temperature of the outer shell in the evaporation section of the heat sink is taken as the target parameter. Three different levels are selected for each of optimization parameters A, B, C, and D. An orthogonal experiment is designed, and L9(3) is selected according to the orthogonal table in the orthogonal experiment table. 4 The value of );

[0016] S3. Simulation Experiment: The simulation calculation was completed in COMSOL software. A 2D hollow finned heat exchanger simulation model of the porous core and steam chamber was established. The lower end of the air-cooled heat exchanger was placed in contact with the heat source that needs to be dissipated. The air-cooled heat exchanger dissipates heat and cools through multiple upper hollow fins and multiple lower hollow fins.

[0017] Physics field setup: The contact surfaces of the steam chamber (1-2-1) with the upper liquid metal layer (2) and the lower liquid metal layer (3) are the core interfaces. Laminar flow nodes are used to solve for the laminar flow in the steam section (1-2). The specified pressure of the laminar flow is... Equal to the saturated vapor pressure at the core interface :

[0018]

[0019] The calculation results above show that when water and gas are in equilibrium in the vaporization section (1-2), the vapor pressure increases with increasing temperature, thus driving the steam to move from the high-temperature region to the low-temperature region. For the liquid flow on the surfaces of the upper and lower liquid metal layers, the Bringkman equation interface is used for solution. The velocity at the cavity-core interface is calculated based on the vapor flow rate on the cavity side. The heat transfer of the plate wall, upper liquid metal layer, lower liquid metal layer, and vapor section in the air-cooled homogenizer is solved using a porous medium heat transfer interface. Thermodynamic nodes are used to create material properties. A vapor system using the ideal gas law is established for the gas phase, and an IAPWA model liquid system is established for the liquid phase inside the air-cooled homogenizer. A vapor pressure function is created for the liquid system to describe the saturation pressure intensity.

[0020] The average temperature of the evaporation section in the air-cooled vapor chamber was obtained through simulation. After data acquisition, the data was analyzed and a graph of the changes in the level of influencing factors was drawn. The results showed that the total number of multiple upper hollow fins and multiple lower hollow fins was the dominant factor affecting the average temperature of the outer wall of the evaporation section in the air-cooled vapor chamber. The second most influential factor was the length of the upper hollow fins, followed by the width of the upper hollow fins. The horizontal angle of the upper hollow fins had the least influence.

[0021] Based on the simulation results, the maximum working fluid velocity on the surface of the upper liquid metal layer in the air-cooled vapor chamber, the maximum steam velocity, and the maximum working fluid velocity on the surface of the upper liquid metal layer in the air-cooled vapor chamber were obtained. When multiple upper hollow fins and multiple lower hollow fins were added to the vapor chamber compared to the finless vapor chamber, the maximum steam velocity and the maximum working fluid velocity on the surface of the upper liquid metal layer in the air-cooled vapor chamber were both improved.

[0022] After configuring multiple upper hollow fins and multiple lower hollow fins, the overall internal flow velocity of the air-cooled heat exchanger is improved, the heat transfer rate from the evaporation section to the condensation section is increased, and the heat is discharged from the condensation section to the outside. At the same time, the return velocity of the condensation section is also strengthened, which accelerates the circulation speed from the evaporation section to the condensation section and ensures that the heat exchange performance of the air-cooled heat exchanger is in the optimal quantitative state.

[0023] S4. Equivalent thermal conductivity: Extracting the optimal factors under the optimal quantization state, the average temperature T of the evaporation section of the air-cooled vapor chamber is obtained in the COMSOL simulation results. e The average temperature T of the condensation section (1-1) c The average temperature T was calculated. e and average temperature T c The difference ; Calculate the thermal conductivity of an air-cooled vapor chamber; Formula for calculating thermal conductivity:

[0024]

[0025] In the above formula: The thermal conductivity is expressed in W·(m·℃). -1 ; The input power of the heat source is expressed in watts (W). The straight-line distance between the midpoint of the evaporation section and the condensation end of the vapor chamber is expressed in meters (m). This represents the average interface area at the midpoint between the evaporation and condensation sections of the vapor chamber, expressed in m². 2 ; The temperature difference between the evaporation and condensation sections of the heat sink, expressed in °C; optimal factors for an air-cooled vapor chamber. =0.08m; =0.000325m 2 ; =1.401℃; Calculated, the equivalent thermal conductivity of the air-cooled vapor chamber under optimal conditions is... 3162.58 W·(m 2 ·℃) -1 ;

[0026] S5. Verify the simulation model: Import the battery pack structure into FLUENT, use transient mode, enable the ENERGY model, and determine the battery density, specific heat capacity, thermal conductivity in all directions, and volumetric heat generation rate. After determining the ambient temperature, simulate the battery pack's temperature rise under uncooled discharge at 0.33C, 0.5C, and 1C rates, recording the highest volumetric temperature of the cells in real time. Compare the recorded simulation data with experimental data:

[0027] Under a discharge state of 0.33C, the highest experimental temperature and the highest simulated temperature of the battery pack were obtained;

[0028] Under a 0.5C discharge condition, the highest experimental temperature and the highest simulated temperature of the battery pack were obtained;

[0029] Under 1C discharge conditions, the highest experimental temperature and the highest simulated temperature of the battery pack were obtained;

[0030] When the experimental and simulated maximum volume temperature error of the battery pack does not exceed 5% at discharge rates of 0.33C, 0.5C, and 1C, it can be determined that the thermal simulation model of the battery pack has a high degree of consistency with the actual situation.

[0031] S6. Adjusting the convective heat transfer coefficient of the condensing section in the heat sink: Using the outlet flow rate and inlet temperature as variables, and obtaining the optimal values ​​of parameters A, B, C, and D, the average temperature of the condensing section of the air-cooled heat exchanger during the entire 1C discharge process of the battery is taken as the surface temperature of the condensing section. At the same inlet temperature, the convective heat transfer coefficient increases with the increase of the outlet flow rate; while at the same outlet flow rate, the convective heat transfer coefficient increases with the decrease of temperature. Simultaneously, within this range, the convective heat transfer coefficient of the condensing section of the heat exchanger can be controlled between 12.8 and 75.8 W·(m³). 2 ·℃) -1 Within the range;

[0032] S7. Self-regulating process of the self-equalizing battery pack structure: The convective heat transfer coefficient of the condensation section in each heat sink is reduced from 5 W·(m²) 2 ·℃) -1 per 10W·(m 2 ·℃) -1 The interval was increased to 55 W·(m 2 ·℃) -1 Maximum temperature of the battery pack Temperature difference The pattern of change is as follows:

[0033] With the convective heat transfer coefficient of the condensation section in each heat sink... Increased maximum battery pack temperature Gradually decrease, temperature difference Gradually increase;

[0034] After multiple temperature-equalizing vertical plates and multiple battery packs are arranged alternately, the convective heat transfer coefficient of the battery packs shows an increasing trend; the highest battery temperature The temperature difference is gradually reduced to 2.0~5.0℃. Through this optimization process, the temperature difference requirement of the battery pack is finally met.

[0035] The beneficial effects of this invention are as follows:

[0036] I. The air-cooled heat spreader in this invention has a reasonable structure. Through the cooperation between the heat dissipation plate, the upper liquid metal layer, the lower liquid metal layer, the outer protective layer, multiple upper hollow fins and multiple lower hollow fins, a multi-layer zoned heat dissipation effect can be achieved, which is conducive to adapting to the existing battery pack for clamping heat dissipation in the wrapped state.

[0037] Second, the adaptation verification method of the self-equalizing battery pack structure in this invention can accurately determine the adaptable structural form of the self-equalizing battery pack structure through modeling, determining indicators, and comparative calculation. This method can facilitate the actual configuration after pre-calculation, save materials, and ensure the reasonable heat dissipation structure of the self-equalizing battery pack structure, making it suitable for widespread use.

[0038] Third, the arrangement of multiple upper hollow fins and multiple lower hollow fins in this invention improves the heat exchange capacity between the heat sink and the outside environment. At the same time, because the upper and lower hollow fins themselves have high heat exchange performance, the evaporation and condensation cycle speed inside the air-cooled heat exchanger is improved. This allows the self-heating battery pack structure formed by the air-cooled heat exchanger to maintain the battery pack temperature at a suitable operating temperature under varying external ambient temperatures. This not only reduces the calculation time for battery pack phase change heat dissipation simulation, but also effectively improves the research efficiency of related studies, providing stable and reliable subsequent research data. Attached Figure Description

[0039] Figure 1 This is a schematic diagram of the main structure of an air-cooled heat exchanger.

[0040] Figure 2 for Figure 1 Enlarged structural diagram at point A;

[0041] Figure 3 This is a schematic diagram of the first three-dimensional structure of an air-cooled heat exchanger.

[0042] Figure 4 This is a schematic diagram of the second three-dimensional structure of an air-cooled heat exchanger.

[0043] Figure 5 This is a schematic diagram of the first three-dimensional structure of a self-equalizing battery pack;

[0044] Figure 6 This is a schematic diagram of the second three-dimensional structure of the self-equalizing battery pack structure.

[0045] Figure 7 This is an exploded view of the self-equalizing battery pack structure.

[0046] Figure 8 The average temperature T in evaporation sections 1-3 of the air-cooled vapor chamber. e Schematic diagram of the curves showing changes in factor levels;

[0047] Figure 9 This is a diagram comparing the experimental and simulated maximum temperature values ​​of the battery pack.

[0048] Figure 10 This is a schematic diagram illustrating the change in the convective heat transfer coefficient.

[0049] In the diagram: 1-Heat dissipation plate; 1-1-Condensation section; 1-1-1-Condensation chamber; 1-2-Steam section; 1-2-1-Steam chamber; 1-3-Evaporation section; 1-3-1-Evaporation chamber; 2-Upper liquid metal layer; 3-Lower liquid metal layer; 4-Outer protective layer; 5-Upper tank; 6-Lower tank; 7-Upper hollow fins; 7-1-First cavity; 8-Lower hollow fins; 8-1-Second cavity; 10-Ventilation cover; 20-Socket; 30-1-Vertical heat equalization plate; 30-2-Battery pack; 12-Air outlet; 12-1-Circular opening; 12-2-Fan. Detailed Implementation

[0050] The following specific examples illustrate the implementation of the present invention. Those skilled in the art can easily understand other advantages and effects of the present invention from the content disclosed in this specification. The present invention can also be implemented or applied through other different specific embodiments, and various details in this specification can also be modified or changed based on different viewpoints and applications without departing from the spirit of the present invention.

[0051] Specific implementation method one: Combining Figure 1 , Figure 2 , Figure 3 , Figure 4 , Figure 5 , Figure 6 and Figure 7This embodiment describes an air-cooled heat exchange plate comprising a heat dissipation plate 1, an upper liquid metal layer 2, a lower liquid metal layer 3, an outer protective layer 4, multiple upper hollow fins 7, and multiple lower hollow fins 8. The heat dissipation plate 1 is a strip-shaped plate. An upper groove 5 is machined along the length of one end of the heat dissipation plate 1, and an upper liquid metal layer 2 is disposed within the upper groove 5. The top surface of the upper liquid metal layer 2 is flush with the upper side of the heat dissipation plate 1. A lower groove 6 is machined along the length of the lower side of the heat dissipation plate 1, located directly below the upper groove 5. A lower... The bottom surface of the liquid metal layer 3 is flush with the bottom side of the heat sink 1. The other end of the heat sink 1 is integrally connected with multiple upper hollow fins 7 and multiple lower hollow fins 8. The multiple upper hollow fins 7 are evenly arranged on the upper side of the heat sink 1 along the length direction of the heat sink 1, and the multiple lower hollow fins 8 are evenly arranged on the lower side of the heat sink 1 along the length direction of the heat sink 1. The upper hollow fins 7 and lower hollow fins 8 are arranged in a one-to-one correspondence. An outer protective layer 4 is provided on the outer wall of the heat sink 1, the multiple upper hollow fins 7 and the multiple lower hollow fins 8.

[0052] Furthermore, the heat dissipation plate 1 includes a condensation section 1-1, a steam section 1-2, and an evaporation section 1-3. One end of the condensation section 1-1 is connected to one end of the steam section 1-2, and the other end of the steam section 1-2 is connected to one end of the evaporation section 1-3. The plurality of upper hollow fins 7 are integrally connected to the upper side of the condensation section 1-1, and the plurality of lower hollow fins 8 are integrally connected to the lower side of the condensation section 1-1.

[0053] The heat dissipation plate 1 includes a condensing section 1-1, a steam section 1-2, and an evaporating section 1-3. A condensing cavity 1-1-1 is machined along its length in the condensing section 1-1, a steam cavity 1-2-1 is machined along its length in the steam section 1-2, and an evaporating cavity 1-3-1 is machined along its length in the evaporating section 1-3. The condensing section 1-1 is connected to one end of the steam cavity 1-2-1 in the steam section 1-2 through the condensing cavity 1-1-1, and the other end of the steam cavity 1-2-1 is connected to one end of the evaporating cavity 1-3-1 in the evaporating section 1-3. A plurality of upper hollow fins 7 are integrally connected to the upper side of the condensing section 1-1, and each upper hollow fin 7 is connected to the condensing cavity 1-1-1. A plurality of lower hollow fins 8 are integrally connected to the lower side of the condensing section 1-1, and each lower hollow fin 8 is connected to the condensing cavity 1-1-1.

[0054] Furthermore, the length of the condensing section 1-1 is less than or equal to one-third of the length of the evaporating section 1-3, and the length of the steam section 1-2 is less than or equal to one-fifth of the length of the condensing section 1-1.

[0055] Furthermore, each upper hollow fin 7 is provided with a first cavity 7-1, and each lower hollow fin 8 is provided with a second cavity 8-1. The first cavity 7-1 is connected to the condensation cavity 1-1-1, and the second cavity 8-1 is connected to the condensation cavity 1-1-1.

[0056] Among them, the steam cavity 1-2-1 is an irregularly shaped cavity, with the width at one end being the same as that of the condensation cavity 1-1-1, and the width at the other end being the same as that of the evaporation cavity 1-3-1. The width ratio of the two ends of the steam cavity 1-2-1 is 1:2. The angle between the surface of the upper hollow fin 7 and the surface of the heat dissipation plate 1 is 75 degrees. Similarly, the angle between the surface of the lower hollow fin 8 and the surface of the heat dissipation plate 1 is 75 degrees. The ratio of the straight-line distance between two adjacent upper hollow fins 7 to the length of the upper hollow fin 7 itself is 1:2, and the ratio of the straight-line distance between two adjacent lower hollow fins 8 to the length of the lower hollow fin 8 itself is 1:2. :2, Meanwhile, the width of the first cavity 7-1 is the same as the width of the second cavity 8-1. The optimal ratio of the width of the first cavity 7-1 to the width of the condensation cavity 1-1-1 is 1:2. The length range of the steam cavity 1-2-1 is between 4 and 6 mm. The sum of the thickness of the upper liquid metal layer 2, the thickness of the lower liquid metal layer 3, and the width of the steam cavity 1-2-1 is four times the width of the first cavity 7-1. Under this size relationship, the air-cooled heat spreader can continuously maintain a high level of heat dissipation. It is compatible with multiple battery packs 30-2 and can be used in the isolated heat transfer and heat dissipation conditions of multiple battery packs 30-2.

[0057] Specific Implementation Method Two: Combining Figures 1 to 7 This embodiment describes a self-regulating battery pack structure comprising a ventilation cover 10, a socket 20, and a self-regulating battery pack structure 30. The socket 20 is horizontally positioned, and the self-regulating battery pack structure 30 is disposed within the socket 20. The bottom of the self-regulating battery pack structure 30 is detachably connected to the socket 20. The ventilation cover 10 is mounted on the socket 20, and the top of the self-regulating battery pack structure 30 is detachably connected to the ventilation cover 10. An air duct is provided on the inner wall of the ventilation cover 10, and an air outlet 12 and an air inlet are respectively machined on the ventilation cover 10. The air outlet 12 and the air inlet are respectively connected to... The air ducts are interconnected. The self-heating battery pack structure 30 is composed of multiple heat-equalizing vertical plates 30-1 and multiple battery packs 30-2. The multiple heat-equalizing vertical plates 30-1 are arranged vertically side by side. Multiple battery packs 30-2 are arranged between every two adjacent heat-equalizing vertical plates 30-1. Each heat-equalizing vertical plate 30-1 is formed by multiple air-cooled heat-equalizing plates arranged vertically. One side of one air-cooled heat-equalizing plate in each heat-equalizing vertical plate 30-1 is in close contact with one side of another air-cooled heat-equalizing plate, and the other side of the air-cooled heat-equalizing plate is in close contact with one side of a third air-cooled heat-equalizing plate.

[0058] Each air-cooled vapor chamber includes a heat dissipation plate 1, an upper liquid metal layer 2, a lower liquid metal layer 3, an outer protective layer 4, multiple upper hollow fins 7, and multiple lower hollow fins 8. The heat dissipation plate 1 is a strip-shaped plate. An upper groove 5 is machined along the length of one end of the heat dissipation plate 1. The upper liquid metal layer 2 is disposed in the upper groove 5, and the top surface of the upper liquid metal layer 2 is flush with the upper side of the heat dissipation plate 1. A lower groove 6 is machined along the length of the lower side of the heat dissipation plate 1. The lower groove 6 is located directly below the upper groove 5, and the lower liquid metal layer 3 is disposed in the lower groove 6. The bottom surface of the lower liquid metal layer 3 is flush with the lower surface of the heat dissipation plate 1. The heat sink 1 is flush with the side panels and has multiple upper hollow fins 7 and multiple lower hollow fins 8 integrally connected to the other end. The multiple upper hollow fins 7 are evenly arranged on the upper side of the heat sink 1 along the length direction of the heat sink 1, and the multiple lower hollow fins 8 are evenly arranged on the lower side of the heat sink 1 along the length direction of the heat sink 1. The upper hollow fins 7 and lower hollow fins 8 are arranged in a one-to-one correspondence. The outer protective layer 4 is wrapped around the outer wall of the heat sink 1, the multiple upper hollow fins 7 and the multiple lower hollow fins 8. The multiple upper hollow fins 7 and the multiple lower hollow fins 8 in each air-cooled heat spreader are arranged facing the ventilation cover 10.

[0059] Furthermore, the air outlet 12 includes multiple circular openings 12-1, which are sequentially formed along the width direction of the ventilation cover 10, and each circular opening 12-1 is provided with a fan 12-2.

[0060] Specific implementation method three: Combining Figures 1 to 10 This embodiment describes the adaptation verification method for the self-equalizing battery pack structure, which includes the following steps:

[0061] S1. Modeling: Use modeling software to create a 3D model of the air-cooled heat exchanger;

[0062] S2. Determination of Optimization Parameters: The total number of upper hollow fins 7 and lower hollow fins 8 is taken as optimization parameter A, the width of upper hollow fins 7 is taken as optimization parameter B, the length of upper hollow fins 7 is taken as optimization parameter C, and the horizontal angle of upper hollow fins 7 is taken as optimization parameter D. The average shell temperature of evaporation sections 1-3 in the heat dissipation plate 1 is taken as the target parameter. Three different levels are selected for each of optimization parameters A, B, C, and D. An orthogonal experiment is designed, and L9(3) is selected according to the orthogonal table in the orthogonal experiment table. 4 The value of );

[0063] S3. Simulation Experiment: The simulation calculation was completed in COMSOL software. A 2D hollow finned heat exchanger simulation model of the porous core and steam chamber was established. The lower end of the air-cooled heat exchanger was placed in contact with the heat source that needs to be dissipated. The air-cooled heat exchanger is cooled by multiple upper hollow fins 7 and multiple lower hollow fins 8.

[0064] Physics setup: The contact surfaces of steam chamber 1-2-1 with the upper and lower liquid metal layers 2 and 3 are respectively the core-cavity interfaces. Laminar flow nodes are used to solve for the laminar flow in steam section 1-2, with the specified pressure for laminar flow. Equal to the saturated vapor pressure at the core interface :

[0065]

[0066] The calculation results above show that when water and gas phases are in equilibrium in steam section 1-2, the vapor pressure increases with increasing temperature, thus driving the steam to move from the high-temperature region to the low-temperature region. For the liquid flow on the surfaces of the upper and lower liquid metal layers 2 and 3, the Bringkman equation interface is used for solution. The velocity at the cavity-core interface is calculated based on the vapor flow rate on the cavity side. The heat transfer of the heat transfer between the plate wall, upper liquid metal layer 2, lower liquid metal layer 3, and steam section 1-2 in the air-cooled homogenizer is solved using a porous medium heat transfer interface. Thermodynamic nodes are used to create material properties. An ideal gas law steam system is established for the gas phase, and an IAPWA model liquid system is established for the liquid phase inside the air-cooled homogenizer. A vapor pressure function is created for the liquid system to describe the saturation pressure intensity.

[0067] The average temperature of evaporation section 1-3 in the air-cooled vapor chamber was obtained through simulation. After data acquisition, it was analyzed and a graph of the changes in the level of influencing factors was drawn. It was concluded that the total number of multiple upper hollow fins 7 and multiple lower hollow fins 8 is the dominant factor affecting the average temperature of the outer wall of evaporation section 1-3 in the air-cooled vapor chamber. The second most influential factor is the length of the upper hollow fin 7, followed by the width of the upper hollow fin 7. The horizontal angle of the upper hollow fin 7 has the least influence.

[0068] Based on the simulation results, the maximum working fluid velocity on the surface of the upper liquid metal layer 2 in the air-cooled heat exchanger, the maximum flow velocity in the steam section 1-2, and the maximum working fluid velocity on the surface of the upper liquid metal layer 2 in the air-cooled heat exchanger were obtained. When multiple upper hollow fins 7 and multiple lower hollow fins 8 were added compared to the finless heat exchanger, the maximum flow velocity in the steam section 1-2 and the maximum working fluid velocity on the surface of the upper liquid metal layer 2 in the air-cooled heat exchanger were both improved.

[0069] After configuring multiple upper hollow fins 7 and multiple lower hollow fins 8, the overall internal flow velocity of the air-cooled heat exchanger is improved, the heat transfer rate from the evaporation section 1-3 to the condensation section 1-1 is increased, and the heat is discharged outward from the condensation section 1-1. At the same time, the reflux velocity of the condensation section 1-1 is also strengthened, which accelerates the circulation speed from the evaporation section 1-3 to the condensation section 1-1, ensuring that the heat exchange performance of the air-cooled heat exchanger is in the optimal quantitative state.

[0070] S4. Equivalent thermal conductivity: Extracting the optimal factors under the optimal quantization state, the average temperature T of the air-cooled vapor chamber in evaporation section 1-3 in the COMSOL simulation results is obtained. e The average temperature T of condensation section 1-1 c The average temperature T was calculated. e and average temperature T c The difference ; Calculate the thermal conductivity of an air-cooled vapor chamber; Formula for calculating thermal conductivity:

[0071]

[0072] In the above formula: The thermal conductivity is expressed in W·(m·℃). -1 ; The input power of the heat source is expressed in watts (W). The straight-line distance between the midpoint of the evaporation section and the condensation end of the vapor chamber is expressed in meters (m). This represents the average interface area at the midpoint between the evaporation and condensation sections of the vapor chamber, expressed in m². 2 ; The temperature difference between the midpoint of the evaporation section 1-3 and the condensation section 1-1 of the heat sink body 1, in °C; the optimal factor for the air-cooled vapor chamber. =0.08m; =0.000325m 2 ; =1.401℃; Calculated, the equivalent thermal conductivity of the air-cooled vapor chamber under optimal conditions is... 3162.58 W·(m 2 ·℃) -1 ;

[0073] S5. Verify the simulation model: Import the battery pack structure into FLUENT, use transient mode, enable the ENERGY model, and determine the battery density, specific heat capacity, thermal conductivity in all directions, and volumetric heat generation rate. After determining the ambient temperature, simulate the temperature rise of the 30-2 battery pack under uncooled discharge at 0.33C, 0.5C, and 1C rates, recording the highest volumetric temperature of the cells in real time. Compare the recorded simulation data with experimental data:

[0074] Under a discharge state of 0.33C, the highest experimental temperature and the highest simulated temperature of battery pack 30-2 were obtained;

[0075] Under a 0.5C discharge condition, the experimental maximum temperature value and the simulated maximum temperature value of battery pack 30-2 were obtained;

[0076] Under 1C discharge conditions, the highest experimental temperature and the highest simulated temperature of battery pack 30-2 were obtained;

[0077] When the experimental and simulated maximum volume temperature error of battery pack 30-2 does not exceed 5% at discharge rates of 0.33C, 0.5C, and 1C, it can be determined that the thermal simulation model of battery pack 30-2 has a high degree of consistency with the actual situation.

[0078] S6. Adjusting the convective heat transfer coefficient of the condensing section 1-1 in the heat sink 1: Using the outlet flow rate and inlet temperature as variables, and obtaining the optimal values ​​of parameters A, B, C, and D, the average temperature of the condensing section 1-1 of the air-cooled heat exchanger during the entire 1C discharge process of the battery is taken as the surface temperature value of the condensing section 1-1. At the same inlet temperature, the convective heat transfer coefficient increases with the increase of the outlet flow rate; while at the same outlet flow rate, the convective heat transfer coefficient increases with the decrease of temperature. Simultaneously, within this range, the convective heat transfer coefficient of the condensing section of the heat exchanger can be controlled between 12.8 and 75.8 W·(m³). 2 ·℃) -1 Within the range;

[0079] S7. Self-regulating process of self-equalizing battery pack structure 30: The convective heat transfer coefficient of the condensation section 1-1 in each heat sink 1 is reduced from 5 W·(m²) 2 ·℃) -1 per 10W·(m 2 ·℃) -1 The interval was increased to 55 W·(m 2 ·℃) -1 The maximum temperature of the battery pack is 30-2. Temperature difference The pattern of change is as follows:

[0080] With the convective heat transfer coefficient of the condensation section 1-1 in each heat sink 1, Increased battery pack maximum temperature 30-2 Gradually decrease, temperature difference Gradually increase; when From 5 W·(m 2 ·℃) -1 Increased to 55 W·(m 2 ·℃) -1 At that time, the battery The temperature decreased from 53.3℃ to 44.8℃, a drop of 8.5℃. However, the temperature rose from 1.4℃ to 5.1℃, an increase of 3.7℃;

[0081] After multiple temperature-equalizing vertical plates (30-1) and multiple battery packs (30-2) are arranged alternately, the convective heat transfer coefficient of battery pack 30-2 shows an increasing trend; the highest battery temperature The convective heat transfer coefficient gradually decreases, that is, after adding the condensation section of the two rows of heat spreaders in the middle battery, it decreases from 55 W·(m²). 2 ·℃) -1 Increased to 65 W·(m 2 ·℃) -1 Maximum battery temperature The temperature was reduced to 44.1℃, and the temperature difference was reduced to 4.0℃. Through this optimization step, the temperature difference requirement of battery pack 30-2 was finally met.

[0082] The self-equalizing battery pack structure mentioned in this embodiment has the same composition and connection relationship as that in specific embodiment two.

[0083] Specific Implementation Method Four: This implementation method is a further limitation of Specific Implementation Methods One, Two, Three, or Four.

[0084] S1. Modeling: Use modeling software to create a 3D model of the air-cooled heat exchanger;

[0085] S2. Determination of Optimization Parameters: The total number of upper hollow fins 7 and lower hollow fins 8 is taken as optimization parameter A, the width of upper hollow fins 7 is taken as optimization parameter B, the length of upper hollow fins 7 is taken as optimization parameter C, and the horizontal angle of upper hollow fins 7 is taken as optimization parameter D. The average shell temperature of evaporation sections 1-3 in the heat sink 1 is taken as the target parameter. Three different levels are selected for each of optimization parameters A, B, C, and D. An orthogonal experiment is designed. Based on the relevant data in Table 1 of the orthogonal experiment, L9(3) is selected. 4 The value of );

[0086] Table 1 Orthogonal Experiment Table

[0087]

[0088] S3. Simulation Experiment: The simulation calculation was completed in COMSOL software. A 2D hollow finned heat exchanger simulation model of the porous core and steam chamber was established. The lower end of the air-cooled heat exchanger was placed in contact with the heat source that needs to be dissipated. The air-cooled heat exchanger is cooled by multiple upper hollow fins 7 and multiple lower hollow fins 8.

[0089] Physics setup: The contact surfaces of steam chamber 1-2-1 with the upper and lower liquid metal layers 2 and 3 are respectively the core-cavity interfaces. Laminar flow nodes are used to solve for the laminar flow in steam section 1-2, with the specified pressure for laminar flow. Equal to the saturated vapor pressure at the core interface :

[0090]

[0091] The calculation results above show that when water and gas phases are in equilibrium in steam section 1-2, the vapor pressure increases with increasing temperature, thus driving the steam to move from the high-temperature region to the low-temperature region. For the liquid flow on the surfaces of the upper and lower liquid metal layers 2 and 3, the Bringkman equation interface is used for solution. The velocity at the cavity-core interface is calculated based on the vapor flow rate on the cavity side. The heat transfer of the heat transfer between the plate wall, upper liquid metal layer 2, lower liquid metal layer 3, and steam section 1-2 in the air-cooled homogenizer is solved using a porous medium heat transfer interface. Thermodynamic nodes are used to create material properties. An ideal gas law steam system is established for the gas phase, and an IAPWA model liquid system is established for the liquid phase inside the air-cooled homogenizer. A vapor pressure function is created for the liquid system to describe the saturation pressure intensity.

[0092] The average temperature of evaporation section 1-3 in the air-cooled vapor chamber was obtained through simulation. After data acquisition, it was analyzed and a graph of the changes in the level of influencing factors was drawn. It was concluded that the total number of multiple upper hollow fins 7 and multiple lower hollow fins 8 is the dominant factor affecting the average temperature of the outer wall of evaporation section 1-3 in the air-cooled vapor chamber. The second most influential factor is the length of the upper hollow fin 7, followed by the width of the upper hollow fin 7. The horizontal angle of the upper hollow fin 7 has the least influence.

[0093] Based on the simulation results, the maximum working fluid velocity on the surface of the upper liquid metal layer 2 in the air-cooled heat exchanger, the maximum flow velocity in the steam section 1-2, and the maximum working fluid velocity on the surface of the upper liquid metal layer 2 in the air-cooled heat exchanger were obtained. When multiple upper hollow fins 7 and multiple lower hollow fins 8 were added compared to the finless heat exchanger, the maximum flow velocity in the steam section 1-2 and the maximum working fluid velocity on the surface of the upper liquid metal layer 2 in the air-cooled heat exchanger were both improved.

[0094] After configuring multiple upper hollow fins 7 and multiple lower hollow fins 8, the overall internal flow velocity of the air-cooled heat exchanger is improved, the heat transfer rate from the evaporation section 1-3 to the condensation section 1-1 is increased, and the heat is discharged outward from the condensation section 1-1. At the same time, the reflux velocity of the condensation section 1-1 is also strengthened, which accelerates the circulation speed from the evaporation section 1-3 to the condensation section 1-1, ensuring that the heat exchange performance of the air-cooled heat exchanger is in the optimal quantitative state.

[0095] In S3, it is assumed that the water and gas phases are in equilibrium at this location. The vapor pressure increases with temperature, which drives the steam from the high-temperature region to the low-temperature region. For the liquid flow within the porous core, the Bringkman equation interface is used for solving, and the velocity at the cavity-core interface is calculated based on the vapor flow rate on the cavity side. The heat transfer of each part of the geometry (plate wall, porous core, steam cavity) is solved using the porous media heat transfer interface. Material properties are created using thermodynamic nodes. A steam system using the ideal gas law is established for the gas phase. An IAPWA model liquid system is established for the liquid phase inside the isothermal plate; a vapor pressure function is created to describe the saturation pressure for the liquid system. To facilitate the application of properties in the model, two materials, liquid water and water vapor, are created using the thermodynamic system's generated material option, and copper from the material library is used to describe the plate wall properties.

[0096] The average temperature of evaporation sections 1-3 was obtained through simulation. The experimental results are shown in Table 2. Analysis was performed, and the changes in the levels of influencing factors were plotted. Figure 8 The simulation results show that the number of fins has the most significant impact on the average temperature of the outer shell of the evaporation section of the vapor chamber, followed by fin length, then fin width, while the horizontal angle of the fins has the least impact. The simulation results also show that the highest working fluid velocity in the liquid wick of the finless vapor chamber is 4.35 × 10⁻⁶. -3 m·s -1 In the steam domain, the highest flow velocity can reach 4.82 m·s. -1 The optimal factor is the working fluid flow rate within the vapor chamber of the temperature-sensing plate, which is 5.16 × 10⁻⁶. -3 m·s -1 In the steam domain, the highest flow velocity can reach 6.53 m / s. -1 Compared to a finless vapor chamber, the addition of hollow fins increased the maximum working fluid velocity in the vapor domain of the optimal vapor chamber by 1.71 m / s. -1 The working fluid flow rate in the suction core increased by 0.81 × 10⁻⁶. -3 m·s -1 After adding fins, the overall flow velocity inside the vapor chamber is increased, allowing heat from the evaporation section to reach the condensation section and dissipate more quickly. The reflux velocity in the condensation section is also enhanced, accelerating the evaporation-condensation cycle and optimizing the heat exchange performance of the vapor chamber. Table 2 details the specifics.

[0097] Table 2:

[0098]

[0099] S4. Equivalent thermal conductivity: Extracting the optimal factors under the optimal quantization state, the average temperature T of the air-cooled vapor chamber in evaporation section 1-3 in the COMSOL simulation results is obtained. e The average temperature T of condensation section 1-1c The average temperature T was calculated. e and average temperature T c The difference ; Calculate the thermal conductivity of an air-cooled vapor chamber; Formula for calculating thermal conductivity:

[0100]

[0101] In the above formula: The thermal conductivity is expressed in W·(m·℃). -1 ; The input power of the heat source is expressed in watts (W). The straight-line distance between the midpoint of the evaporation section and the condensation end of the vapor chamber is expressed in meters (m). This represents the average interface area at the midpoint between the evaporation and condensation sections of the vapor chamber, expressed in m². 2 ; The temperature difference between the midpoint of the evaporation section 1-3 and the condensation section 1-1 of the heat sink body 1, in °C; the optimal factor for the air-cooled vapor chamber. =0.08m; =0.000325m 2 ; =1.401℃; Calculated, the equivalent thermal conductivity of the air-cooled vapor chamber under optimal conditions is... 3162.58 W·(m 2 ·℃) -1 ;

[0102] S5. Verify the simulation model: Import the battery pack structure into FLUENT, use transient mode, enable the ENERGY model, and determine the battery density, specific heat capacity, thermal conductivity in all directions, and volumetric heat generation rate. After determining the ambient temperature, simulate the temperature rise of the 30-2 battery pack under uncooled discharge at 0.33C, 0.5C, and 1C rates, recording the highest volumetric temperature of the cells in real time. Compare the recorded simulation data with experimental data:

[0103] Under a discharge state of 0.33C, the highest experimental temperature and the highest simulated temperature of battery pack 30-2 were obtained;

[0104] Under a 0.5C discharge condition, the experimental maximum temperature value and the simulated maximum temperature value of battery pack 30-2 were obtained;

[0105] Under 1C discharge conditions, the highest experimental temperature and the highest simulated temperature of battery pack 30-2 were obtained;

[0106] Combination Figure 9The comparison results are as follows: at 0.33C discharge, the highest temperature in the experiment is 41.14℃, and the highest temperature in the simulation is 40.84℃; at 0.5C discharge, the highest temperature in the experiment is 46.57℃, and the highest temperature of the simulated battery pack is 48.03℃; at 1C discharge, the highest temperature in the experiment is 56.7℃, and the highest temperature in the simulation is 56.28℃.

[0107] When the experimental and simulated maximum volume temperature error of battery pack 30-2 does not exceed 5% at discharge rates of 0.33C, 0.5C, and 1C, it can be determined that the thermal simulation model of battery pack 30-2 has a high degree of consistency with the actual situation.

[0108] S6. Adjusting the convective heat transfer coefficient of the condensing section 1-1 in the heat sink 1: Using the outlet flow rate and inlet temperature as variables, and obtaining the optimal values ​​of parameters A, B, C, and D, the average temperature of the condensing section 1-1 of the air-cooled heat exchanger during the entire 1C discharge process of the battery is taken as the surface temperature value of the condensing section 1-1. At the same inlet temperature, the convective heat transfer coefficient increases with the increase of the outlet flow rate; while at the same outlet flow rate, the convective heat transfer coefficient increases with the decrease of temperature. Simultaneously, within this range, the convective heat transfer coefficient of the condensing section of the heat exchanger can be controlled between 12.8 and 75.8 W·(m³). 2 ·℃) -1 Within the range;

[0109] S7. Self-regulating process of self-equalizing battery pack structure 30: The convective heat transfer coefficient of the condensation section 1-1 in each heat sink 1 is reduced from 5 W·(m²) 2 ·℃) -1 per 10W·(m 2 ·℃) -1 The interval was increased to 55 W·(m 2 ·℃) -1 Based on the data in Table 3, the highest temperature of the battery pack 30-2 is... Temperature difference The pattern of change is as follows:

[0110] Table 3. Influence of Convection Heat Transfer Coefficient in the Condensation Section

[0111]

[0112] With the convective heat transfer coefficient of the condensation section 1-1 in each heat sink 1, Increased battery pack maximum temperature 30-2 Gradually decrease, temperature difference Gradually increase; when From 5 W·(m 2 ·℃) -1 Increased to 55 W·(m 2 ·℃)-1 At that time, the battery The temperature decreased from 53.3℃ to 44.8℃, a drop of 8.5℃. However, the temperature rose from 1.4℃ to 5.1℃, an increase of 3.7℃;

[0113] The convective heat transfer coefficient after adding the condensation section of the two rows of heat spreaders in the middle battery increased from 55 W·(m²) 2 ·℃) -1 Increased to 65W·(m 2 ·℃) -1 Maximum battery temperature The temperature was reduced to 44.1℃, and the temperature difference was reduced to 4.0℃. Through this optimization step, the temperature difference requirement of battery pack 30-2 was finally met.

[0114] Specific Implementation Method Five: This implementation method is a further limitation of Specific Implementation Methods One, Two, Three, or Four. The air-cooled heat dissipation plate heat dissipation automotive power battery pack adaptation verification method further includes a quantitative determination process of the capillary force and permeability of the upper liquid metal layer 2 and the lower liquid metal layer 3. The specific process is as follows:

[0115] First, determine the capillary force and permeability of the upper liquid metal layer 2 and the lower liquid metal layer 3;

[0116] The capillary force of the upper liquid metal layer 2 is the same as that of the lower liquid metal layer 3, and the permeability of the upper liquid metal layer 2 is the same as that of the lower liquid metal layer 3. The calculation process for the capillary force of the upper liquid metal layer 2 is as follows:

[0117] Formula 1 is the formula for capillary force measurement:

[0118] (1)

[0119] In the above formula, This refers to the liquid density of the upper liquid metal layer 2 or the lower liquid metal layer 3, which is the density of the liquid contained in the upper liquid metal layer 2 or the lower liquid metal layer 3. This refers to the difference in liquid level in a capillary force measuring device;

[0120] The permeability of the upper liquid metal layer 2 is obtained using Formula 2, which is the formula for measuring permeability. The permeability is calculated using Darcy's formula. Formula 2 is as follows:

[0121] (2)

[0122] In the above formula, To allow for the pressure difference between the two ends of the multi-hole core for liquid flow, For the height of the porous material, Viscosity of the flowing liquid The permeability of the sample. The average flow velocity of the liquid per unit interface. , ;

[0123] The smaller the size of the metal powder, the smaller the interconnecting pores formed between the sintered powders, and therefore the smaller the effective capillary radius of the upper liquid metal layer 2. The smaller; capillary force It is due to the surface tension of the liquid Contact angle and the effective capillary radius of the upper liquid metal layer 2 The decision is made theoretically using Formula 3, which is:

[0124] (3).

Claims

1. A self-regulating battery pack structure, characterized in that: The device includes a ventilation cover (10), a socket (20), and a self-regulating battery pack structure (30). The socket (20) is horizontally positioned, and the self-regulating battery pack structure (30) is located inside the socket (20). The bottom of the self-regulating battery pack structure (30) is detachably connected to the socket (20). The ventilation cover (10) is located on the socket (20), and the top of the self-regulating battery pack structure (30) is detachably connected to the ventilation cover (10). An air duct is provided on the inner wall of the ventilation cover (10). The air outlet (12) and air inlet are respectively processed. The air outlet (12) and air inlet are respectively connected to the air duct. The self-equalizing battery pack structure (30) is composed of multiple equalizing vertical long plates (30-1) and multiple battery packs (30-2). The multiple equalizing vertical long plates (30-1) are arranged vertically side by side. Multiple battery packs (30-2) are arranged between every two adjacent equalizing vertical long plates (30-1). Each equalizing vertical long plate (30-1) is formed by multiple air-cooled equalizing plates arranged vertically. Each air-cooled heat exchanger includes a heat dissipation plate (1), an upper liquid metal layer (2), a lower liquid metal layer (3), an outer protective layer (4), multiple upper hollow fins (7), and multiple lower hollow fins (8). The heat dissipation plate (1) is a strip-shaped plate. An upper groove (5) is machined along the length of one end of the heat dissipation plate (1). An upper liquid metal layer (2) is installed in the upper groove (5). The top surface of the upper liquid metal layer (2) is flush with the upper side of the heat dissipation plate (1). A lower groove (6) is machined along the length of the lower side of the heat dissipation plate (1). The lower groove (6) is located directly below the upper groove (5). A lower liquid metal layer (3) is installed in the lower groove (6). The bottom surface of the lower liquid metal layer (3) is flush with the bottom surface of the heat dissipation plate (1). The lower side of the heat dissipation plate (1) is flush with the heat dissipation plate (1). The other end of the heat dissipation plate (1) is integrally connected with multiple upper hollow fins (7) and multiple lower hollow fins (8). The multiple upper hollow fins (7) are evenly arranged on the upper side of the heat dissipation plate (1) along the length direction of the heat dissipation plate (1), and the multiple lower hollow fins (8) are evenly arranged on the lower side of the heat dissipation plate (1) along the length direction of the heat dissipation plate (1). The upper hollow fins (7) and the lower hollow fins (8) are arranged one-to-one. The outer protective layer (4) is wrapped around the outer wall of the heat dissipation plate (1), the multiple upper hollow fins (7) and the multiple lower hollow fins (8). The multiple upper hollow fins (7) and the multiple lower hollow fins (8) in each air-cooled heat dissipation plate are arranged facing the ventilation cover (10). The heat dissipation plate (1) includes a condensation section (1-1), a steam section (1-2), and an evaporation section (1-3). A condensation cavity (1-1-1) is machined along its length in the condensation section (1-1), a steam cavity (1-2-1) is machined along its length in the steam section (1-2), and an evaporation cavity (1-3-1) is machined along its length in the evaporation section (1-3). The condensation section (1-1) is connected to the steam cavity (1-2-1) in the steam section (1-2) via the condensation cavity (1-1-1). -1) is connected to one end of the steam chamber (1-2-1), and the other end of the steam chamber (1-2-1) is connected to one end of the evaporation chamber (1-3-1) in the evaporation section (1-3). The plurality of upper hollow fins (7) are integrally connected to the upper side of the condensation section (1-1), and each upper hollow fin (7) is connected to the condensation chamber (1-1-1). The plurality of lower hollow fins (8) are integrally connected to the lower side of the condensation section (1-1), and each lower hollow fin (8) is connected to the condensation chamber (1-1-1). Each upper hollow fin (7) has a first cavity (7-1) processed along its length direction, and each lower hollow fin (8) has a second cavity (8-1) processed along its length direction; The steam chamber (1-2-1) is an irregularly shaped cavity. The width of one end of the steam chamber (1-1-1) is the same as that of the condensation chamber (1-1-1), and the width of the other end of the steam chamber (1-2-1) is the same as that of the evaporation chamber (1-3-1). The width ratio of the two ends of the steam chamber (1-2-1) is 1:

2. The angle between the surface of the upper hollow fin (7) and the surface of the heat dissipation plate (1) is 75 degrees, and the angle between the surface of the lower hollow fin (8) and the surface of the heat dissipation plate (1) is 75 degrees. The ratio of the straight-line distance between two adjacent upper hollow fins (7) to the length of the upper hollow fin (7) itself is 1. :2, the ratio of the straight-line distance between two adjacent lower hollow fins (8) to the length of the lower hollow fin (8) is 1:2, the width of the first cavity (7-1) is the same as the width of the second cavity (8-1), the ratio of the width of the first cavity (7-1) to the width of the condensation cavity (1-1-1) is 1:2, the length range of the steam cavity (1-2-1) is between 4 and 6 mm, and the sum of the thickness of the upper liquid metal layer (2), the thickness of the lower liquid metal layer (3) and the width of the steam cavity (1-2-1) is four times the width of the first cavity (7-1).

2. The self-equalizing battery pack structure according to claim 1, characterized in that: The air outlet (12) includes multiple circular openings (12-1), which are sequentially formed along the width direction of the ventilation cover (10), and each circular opening (12-1) is equipped with a fan (12-2).

3. A method for adapting and verifying a self-temperature-equalizing battery pack structure, used to form the self-temperature-equalizing battery pack structure as described in claim 1 or 2, characterized in that: Includes the following steps: S1. Modeling: Use modeling software to create a 3D model of the air-cooled heat exchanger; S2. Determination of Optimization Parameters: The total number of upper hollow fins (7) and lower hollow fins (8) is taken as optimization parameter A, the width of upper hollow fins (7) is taken as optimization parameter B, the length of upper hollow fins (7) is taken as optimization parameter C, and the horizontal angle of upper hollow fins (7) is taken as optimization parameter D. Taking the average shell temperature of the evaporation section (1-3) in the heat dissipation plate (1) as the target parameter, three different levels are selected for each of optimization parameters A, B, C and D. Orthogonal experiments are designed, and L9(3) is selected according to the orthogonal table in the orthogonal experiment table. 4 The value of ); S3. Simulation Experiment: The simulation calculation was completed in COMSOL software. A 2D hollow finned heat exchanger simulation model of the porous core and steam chamber was established. The lower end of the air-cooled heat exchanger was placed in contact with the heat source that needs to be dissipated. The air-cooled heat exchanger was cooled by multiple upper hollow fins (7) and multiple lower hollow fins (8). Physics field setup: The contact surfaces of the steam chamber (1-2-1) with the upper liquid metal layer (2) and the lower liquid metal layer (3) are the core interfaces. Laminar flow nodes are used to solve for the laminar flow in the steam section (1-2). The specified pressure p of the laminar flow is equal to the saturated vapor pressure p at the core interface. H2O,sat (T): p=p H20,sat (T) The calculation results from the above formula show that when water and gas are in equilibrium in the steam section (1-2), the vapor pressure increases with the increase of temperature, thereby driving the steam to move from the high temperature region to the low temperature region. For the liquid flow on the surface of the upper liquid metal layer (2) and the lower liquid metal layer (3), the Bringkman equation interface is used for solution. The velocity on the cavity-core interface is calculated based on the vapor flow rate on the cavity side. The heat transfer of the plate wall, upper liquid metal layer (2), lower liquid metal layer (3) and steam section (1-2) in the air-cooled heat exchange plate is solved using a porous medium heat transfer interface. The average temperature of the evaporation section (1-3) in the air-cooled heat exchanger was obtained through simulation. After data acquisition, it was analyzed and a graph of the changes in the level of influencing factors was drawn. It was concluded that the total number of multiple upper hollow fins (7) and multiple lower hollow fins (8) is the dominant factor affecting the average temperature of the outer wall of the evaporation section (1-3) in the air-cooled heat exchanger. The second most influential factor is the length of the upper hollow fins (7), followed by the width of the upper hollow fins (7). The least influential factor is the horizontal angle of the upper hollow fins (7). Based on the simulation results, the maximum working fluid velocity on the surface of the upper liquid metal layer (2) of the air-cooled heat exchanger without fins, the maximum flow velocity in the steam section (1-2), and the maximum working fluid velocity on the surface of the upper liquid metal layer (2) of the air-cooled heat exchanger were obtained. After configuring multiple upper hollow fins (7) and multiple lower hollow fins (8), the heat exchange performance of the air-cooled heat exchange plate is ensured to be in the optimal quantitative state. S4. Equivalent thermal conductivity: Extracting the optimal factors under the optimal quantization state, the average temperature T of the evaporation section (1-3) of the air-cooled vapor chamber is obtained in the COMSOL simulation results. e The average temperature T of the condensation section (1-1) c The average temperature T was calculated. e and average temperature T c The difference ΔT; Calculate the thermal conductivity of the air-cooled vapor chamber; Formula for calculating thermal conductivity: In the above formula: K is the equivalent thermal conductivity, W·(m·℃). -1 Q is the input power of the heat source, in W; L is the straight-line distance between the midpoint of the evaporation section and the condensation end of the heat spreader, in m. A represents the average interface area at the midpoint between the evaporation and condensation sections of the vapor chamber, in m². 2 ; ΔT is the temperature difference between the middle position of the evaporation section (1-3) and the condensation section (1-1) of the heat sink (1), in °C; the optimal factor for the air-cooled heat spreader is L = 0.08m; A = 0.000325m 2 ΔT = 1.401℃; Calculations show that the equivalent thermal conductivity K of the air-cooled vapor chamber under optimal conditions is 3162.58 W·(m²). 2 ·℃) -1 ; S5. Verify the simulation model: Import the battery pack structure into FLUENT, use transient mode, enable the ENERGY model, and determine the battery density, specific heat capacity, thermal conductivity in all directions, and volumetric heat generation rate. After determining the ambient temperature, simulate the temperature rise of the battery pack (30-2) under uncooled discharge at 0.33C, 0.5C, and 1C rates, and record the highest volumetric temperature of the cells in real time. Compare the recorded simulation data with the experimental data: Under a discharge state of 0.33C, the highest experimental temperature and the highest simulated temperature of the battery pack (30-2) were obtained; Under a 0.5C discharge condition, the highest experimental temperature and the highest simulated temperature of the battery pack (30-2) were obtained; Under 1C discharge conditions, the highest experimental temperature and the highest simulated temperature of the battery pack (30-2) were obtained; When the experimental and simulated maximum temperature error of the battery pack (30-2) does not exceed 5% at discharge rates of 0.33C, 0.5C, and 1C, it can be determined that the thermal simulation model of the battery pack (30-2) has a high degree of consistency with the actual situation. S6. Adjusting the convective heat transfer coefficient of the condensing section (1-1) in the heat sink (1): Taking the outlet flow rate and inlet temperature as variables, and obtaining the optimal parameters A, B, C, and D at their optimal values, the average temperature of the condensing section (1-1) of the air-cooled heat exchanger during the entire 1C discharge process of the battery is taken as the surface temperature value of the condensing section (1-1). At the same inlet temperature, the convective heat transfer coefficient increases with the increase of the outlet flow rate; while at the same outlet flow rate, the convective heat transfer coefficient increases with the decrease of the temperature; at the same time, within this range, the convective heat transfer coefficient of the condensing section of the heat exchanger is controlled between 12.8 and 75.8 W·(m³). 2 ·℃) -1 Within the range; S7. Self-regulating process of the self-equalizing battery pack structure (30): The convective heat transfer coefficient of the condensation section (1-1) in each heat sink (1) is reduced from 5 W·(m²) to 5 W·(m²). 2 ·℃) -1 per 10W·(m 2 ·℃) -1 The interval was increased to 55 W·(m 2 ·℃) -1 The highest temperature T of the battery pack (30-2) max The variation law of temperature difference ΔT is as follows: As the convective heat transfer coefficient h of the condensation section (1-1) in each heat sink (1) increases, the highest temperature T of the battery pack (30-2) increases. max As the temperature gradually decreases, the temperature difference ΔT gradually increases. After multiple uniformly heated vertical plates (30-1) and multiple battery packs (30-2) are arranged alternately, the convective heat transfer coefficient of the battery pack (30-2) shows an increasing trend; the highest battery temperature T max The temperature difference is gradually reduced to 2.0–5.0℃. Through this optimization process, the temperature difference requirement of the battery pack (30-2) is finally met.

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