As an emerging power source, the energy storage battery has been widely applied in electric vehicles, grid peak shaving, and renewable energy integration. One of the critical challenges in the application of an energy storage battery is the temperature control within the battery module. An excessive temperature rise and uneven temperature distribution may lead to accelerated aging, capacity degradation, and even thermal runaway of the energy storage battery. In this study, I propose a novel thermal management unit for an energy storage battery that combines flat heat pipes, finned heat sinks, and phase change materials (PCM). To evaluate the thermal performance of the proposed unit, I conduct theoretical analysis and numerical simulation using ANSYS-Icepak. I systematically investigate the effects of fin number, fin thickness, fin spacing, heat pipe arrangement, battery discharge rate, wind speed, and PCM on the cooling performance and temperature uniformity of the energy storage battery module. The results indicate that the optimized configuration can significantly reduce the maximum temperature and the temperature difference, thereby improving the reliability and lifespan of the energy storage battery. This research provides a practical solution for the thermal design of energy storage battery systems.
1. Introduction
The rapid development of new energy technologies has promoted the widespread use of energy storage equipment. As an important energy storage device, the energy storage battery must maintain a proper operating temperature range to ensure high performance, long cycle life, and safety. The temperature of the energy storage battery is influenced by internal heat generation during charging and discharging, as well as by ambient conditions. If the heat cannot be dissipated efficiently, the temperature of the energy storage battery may exceed the recommended limit, causing irreversible damage. Therefore, a well-designed thermal management system is essential for the reliable operation of an energy storage battery.
Various cooling methods have been proposed for thermal management of the energy storage battery, including air cooling, liquid cooling, phase change material (PCM) cooling, and heat pipe cooling. Air cooling is simple and low-cost, but its cooling capacity is limited. Liquid cooling provides better heat transfer performance but adds complexity and weight. PCM cooling can absorb a large amount of latent heat during phase transition, yet its long-term performance depends on heat dissipation to the environment. Heat pipe cooling offers extremely high equivalent thermal conductivity and excellent temperature uniformity, making it a promising solution for the energy storage battery. In this paper, I combine the advantages of heat pipes and PCM to develop an efficient thermal management unit for a large-format energy storage battery.
I first analyze the heat transfer principles of the heat pipe and the heat generation characteristics of the energy storage battery. Then I design the thermal management unit and build a numerical model in ANSYS-Icepak. By conducting a series of simulations, I evaluate the influence of key parameters and determine the optimal design. The findings of this study are expected to provide guidance for engineers who are developing thermal management systems for high-capacity energy storage battery modules.
2. Heat Transfer Analysis of Heat Pipe for Energy Storage Battery
2.1 Structure and Working Principle of the Heat Pipe
A heat pipe is a sealed vessel containing a working fluid and a wick structure. It transfers heat through the evaporation and condensation of the working fluid. As illustrated in the classic theoretical model, the heat pipe consists of three sections: the evaporator section, the adiabatic section, and the condenser section. The evaporator section absorbs heat from the heat source (in this case, the energy storage battery), causing the working fluid to evaporate. The vapor then travels to the condenser section, where it releases latent heat and condenses back into liquid. The liquid returns to the evaporator through capillary forces in the wick. This circulation loop enables the heat pipe to transport large amounts of heat with very low temperature differences.
For the thermal management of an energy storage battery, I select a flat heat pipe with a width of 11 mm, a thickness of 3 mm, and a length of 240 mm. The working fluid is acetone, which has a boiling point of 56.25 °C at atmospheric pressure, making it suitable for battery operating temperatures between 0 °C and 60 °C. The physical properties of acetone are given in Table 1.
| Property | Value |
|---|---|
| Molecular weight | 58.1 |
| Normal boiling point (°C) | 56.25 |
| Melting point (°C) | -93.15 |
| Critical temperature (K) | 508.15 |
| Critical pressure (kPa) | 4761 |
| Critical density (kg/m³) | 273 |
| Saturation pressure (Pa) | 0.6×10⁵ |
| Latent heat of vaporization (kJ/kg) | 536 |
| Liquid density (kg/m³) | 768 |
| Vapor density (kg/m³) | 1.05 |
| Liquid viscosity (N·s/m²) | 0.269×10⁻³ |
| Vapor viscosity (N·s/m²) | 86×10⁻⁷ |
| Liquid thermal conductivity (W/(m·K)) | 0.175 |
| Vapor thermal conductivity (W/(m·K)) | 0.0126 |
| Surface tension (N/m) | 21.2×10⁻³ |
| Liquid specific heat (J/(kg·K)) | 2220 |
| Vapor specific heat (J/(kg·K)) | 1328 |
2.2 Advantages of Heat Pipe for Energy Storage Battery Thermal Management
The heat pipe offers several advantages that are particularly beneficial for the thermal management of the energy storage battery:
- Extremely high thermal conductivity: The effective thermal conductivity of a heat pipe can be several orders of magnitude higher than that of pure copper or aluminum. For example, a sodium heat pipe operating at 103 K can achieve a thermal conductivity of 10⁶ W/(m·K). Traditional finned heat sinks have a heat transfer coefficient of about 37.4 W/(m²·K), while a heat pipe can reach 268.8 W/(m²·K) in gas-to-gas heat exchange, as shown in Table 2.
- Large heat transport capacity: The heat pipe utilizes latent heat of vaporization, which allows it to transport a large amount of heat with a small cross-sectional area.
- Excellent temperature uniformity: Because the heat pipe operates under nearly isothermal conditions, the temperature difference between the evaporator and condenser is usually less than 5 °C, sometimes only 2–3 °C.
- Wide operating temperature range: Heat pipes can operate from -200 °C to 2000 °C depending on the working fluid.
| Heat transfer element | Heat transfer coefficient K (W/(m²·K)) |
|---|---|
| Plain tube | 20 |
| Finned tube | 37.4 |
| Heat pipe | 268.8 |
3. Design of the Thermal Management Unit for Energy Storage Battery
3.1 Working Principle of the Proposed Thermal Management Unit
The proposed thermal management unit for the energy storage battery combines a flat heat pipe, finned heat sink, and PCM. The working principle is as follows: the flat heat pipe absorption section is attached to the surface of the energy storage battery. Heat generated by the battery is conducted to the heat pipe evaporator, where the working fluid (acetone) evaporates. The vapor flows to the condenser section, where heat is dissipated to the ambient air through aluminum fins. The condensed liquid returns to the evaporator via capillary force. Meanwhile, a PCM layer is placed between adjacent battery cells to store a portion of the heat during high-rate discharge. This arrangement enhances the heat transfer path: the energy storage battery wall → flat heat pipe → fins → air, and also employs PCM to temporarily absorb excess heat, thereby suppressing temperature rise and improving temperature uniformity.
3.2 Overall Structure and Key Components
The main structure of the energy storage battery thermal management unit is shown in the figure below. It consists of a large-format prismatic energy storage battery, four flat heat pipes, an aluminum fin stack, and PCM modules inserted between cells. The battery model used in this study is a 280 A·h lithium iron phosphate (LiFePO₄) cell with a nominal voltage of 3.2 V. Its key parameters are listed in Table 3.
| Parameter | Value |
|---|---|
| Nominal voltage | 3.2 V |
| Nominal capacity | 280 A·h |
| Dimensions (W × H × T) | 173.9 × 207.2 × 71.7 mm |
| Mass | 5.34 ± 0.30 kg |
| Internal resistance (1 kHz) | 0.17 ± 0.05 mΩ |
| Operating voltage | 2.0–3.65 V |
| Charge temperature range | 0–60 °C |
| Discharge temperature range | -20–60 °C |
| Cycle life | ≥6000 cycles |
The thermal management unit uses a flat heat pipe with the following design parameters: length 240 mm, width 11 mm, thickness 3 mm, wall thickness 0.4 mm, wick thickness 0.5 mm. The fin stack is made of aluminum with a thickness of 0.4 mm and a spacing of 4 mm. The number of fins is 16, and each fin has dimensions matched to the heat pipe condenser section. The PCM chosen is a commercially available paraffin wax with a phase change temperature of 44 °C and a latent heat of 240 kJ/kg. Table 4 summarizes the key design parameters of the thermal management unit.
| Parameter | Value |
|---|---|
| Heat pipe dimensions | 240 × 11 × 3 mm |
| Evaporator length | 120 mm |
| Condenser length | 120 mm |
| Heat pipe wall thickness | 0.4 mm |
| Wick thickness | 0.5 mm |
| Fin material | Aluminum |
| Fin thickness | 0.4 mm |
| Fin spacing | 4 mm |
| Number of fins | 16 |
| PCM type | PCM-A-44H (paraffin) |
| PCM phase change temperature | 44 °C |
| PCM latent heat | 240 kJ/kg |
| PCM density | 778 kg/m³ |
3.3 Heat Transfer Theory and Governing Equations
The heat transfer in the energy storage battery thermal management unit involves conduction inside the battery, heat pipe, and fins, as well as convection to the cooling air. To model the temperature distribution, I use the following conservation equations:
Mass conservation equation:
$$
\frac{\partial \rho}{\partial t} + \frac{\partial (\rho u_x)}{\partial x} + \frac{\partial (\rho u_y)}{\partial y} + \frac{\partial (\rho u_z)}{\partial z} = 0
$$
where \rho is the air density and u_x, u_y, u_z are the velocity components in the x, y, and z directions.
Momentum conservation equations:
$$
\rho \left( \frac{\partial u}{\partial t} + u \frac{\partial u}{\partial x} + v \frac{\partial u}{\partial y} + w \frac{\partial u}{\partial z} \right) = -\frac{\partial p}{\partial x} + \mu \left( \frac{\partial^2 u}{\partial x^2} + \frac{\partial^2 u}{\partial y^2} + \frac{\partial^2 u}{\partial z^2} \right) + S_x
$$
$$
\rho \left( \frac{\partial v}{\partial t} + u \frac{\partial v}{\partial x} + v \frac{\partial v}{\partial y} + w \frac{\partial v}{\partial z} \right) = -\frac{\partial p}{\partial y} + \mu \left( \frac{\partial^2 v}{\partial x^2} + \frac{\partial^2 v}{\partial y^2} + \frac{\partial^2 v}{\partial z^2} \right) + S_y
$$
$$
\rho \left( \frac{\partial w}{\partial t} + u \frac{\partial w}{\partial x} + v \frac{\partial w}{\partial y} + w \frac{\partial w}{\partial z} \right) = -\frac{\partial p}{\partial z} + \mu \left( \frac{\partial^2 w}{\partial x^2} + \frac{\partial^2 w}{\partial y^2} + \frac{\partial^2 w}{\partial z^2} \right) + S_z
$$
Energy conservation equation:
$$
\rho C_p \left( \frac{\partial T}{\partial t} + u \frac{\partial T}{\partial x} + v \frac{\partial T}{\partial y} + w \frac{\partial T}{\partial z} \right) = \lambda \left( \frac{\partial^2 T}{\partial x^2} + \frac{\partial^2 T}{\partial y^2} + \frac{\partial^2 T}{\partial z^2} \right) + S_T
$$
where C_p is the specific heat, \lambda is the thermal conductivity, and S_T is the heat source term corresponding to the heat generation rate of the energy storage battery.
For the turbulent airflow around the fins, I use the standard k-epsilon turbulence model with the following transport equations:
$$
\frac{\partial (\rho k)}{\partial t} + \frac{\partial (\rho k u_i)}{\partial x_i} = \frac{\partial}{\partial x_j} \left[ \left( \mu + \frac{\mu_t}{\sigma_k} \right) \frac{\partial k}{\partial x_j} \right] + G_k – \rho \varepsilon
$$
$$
\frac{\partial (\rho \varepsilon)}{\partial t} + \frac{\partial (\rho \varepsilon u_i)}{\partial x_i} = \frac{\partial}{\partial x_j} \left[ \left( \mu + \frac{\mu_t}{\sigma_\varepsilon} \right) \frac{\partial \varepsilon}{\partial x_j} \right] + C_{1\varepsilon} \frac{\varepsilon}{k} G_k – C_{2\varepsilon} \rho \frac{\varepsilon^2}{k}
$$
where k is the turbulent kinetic energy, \varepsilon is the dissipation rate, G_k is the generation of turbulence kinetic energy due to mean velocity gradients, and \mu_t is the turbulent viscosity given by:
$$
\mu_t = \rho C_\mu \frac{k^2}{\varepsilon}
$$
For the phase change material, the effective heat capacity method is employed to account for the latent heat during melting. The energy equation is modified using the enthalpy-porosity formulation. The total enthalpy is expressed as the sum of sensible enthalpy h and latent heat ΔH:
$$
H = c_p T + \Delta H
$$
where ΔH corresponds to the latent heat content that varies between zero and the latent heat of fusion L. During the phase transition, the PCM absorbs heat without a significant temperature rise, thereby acting as a thermal buffer for the energy storage battery.
4. Numerical Simulation Platform Setup
4.1 Software and Physical Model
In this work, I use ANSYS-Icepak, a specialized software for electronics cooling and thermal simulation, to model the heat transfer and fluid flow in the energy storage battery thermal management unit. The physical model is constructed based on the actual geometry dimensions and material properties. The model includes the battery cell, the flat heat pipes, the aluminum fins, and the PCM layers. To simplify the computational domain, I make the following assumptions:
- Thermal contact resistance at the interface between the heat pipe and fins is neglected.
- The heat generation rate of the energy storage battery is uniform during the discharge process.
- Thermal radiation is ignored because the temperature range is moderate.
4.2 Grid Generation
The computational domain is discretized using unstructured grids generated by the Mesher-HD algorithm. The grid is refined near the heat pipes and fins to capture the thermal boundary layer accurately. After a grid independence check, the final mesh consists of 330,368 cells and 348,424 nodes. The mesh quality is evaluated using the face alignment ratio. The average face alignment is above 0.8, indicating a high-quality mesh for the numerical simulation.
4.3 Boundary Conditions
The boundary conditions are assigned as follows:
- The inlet temperature is 25 °C, and the inlet velocity is varied from 2 m/s to 10 m/s to study the effect of wind speed.
- The outlet is set as a pressure outlet with zero gauge pressure.
- The heat source is modeled by applying a volumetric heat generation rate corresponding to the discharge rate of the energy storage battery (2 C, 1 C, or 0.5 C). The total heat generation rate is 120 W for the 2 C discharge case.
- The walls of the computational domain are treated as no-slip and adiabatic except for the heat exchange surfaces.
Table 5 summarizes the simulation cases for fin parameter studies.
| Case | Discharge rate (C) | Ambient temp (°C) | Wind speed (m/s) | Fin thickness (mm) | Fin number |
|---|---|---|---|---|---|
| 1 | 2 | 25 | 2 | 0.6 | 12 |
| 2 | 2 | 25 | 2 | 0.6 | 14 |
| 3 | 2 | 25 | 2 | 0.6 | 16 |
| 4 | 2 | 25 | 2 | 0.6 | 18 |
| 5 | 1 | 25 | 2 | 0.6 | 12 |
| 6 | 1 | 25 | 2 | 0.6 | 14 |
| 7 | 1 | 25 | 2 | 0.6 | 16 |
| 8 | 1 | 25 | 2 | 0.6 | 18 |
| 9 | 0.5 | 25 | 2 | 0.6 | 12 |
| 10 | 0.5 | 25 | 2 | 0.6 | 14 |
| 11 | 0.5 | 25 | 2 | 0.6 | 16 |
| 12 | 0.5 | 25 | 2 | 0.6 | 18 |
5. Results and Discussion
5.1 Effect of Fin Number on Heat Dissipation Performance
In the first series of simulations, I keep the fin thickness at 0.6 mm, the wind speed at 2 m/s, and vary the fin number from 12 to 18. The maximum temperature of the energy storage battery thermal management unit is recorded for three discharge rates. The results are shown in Table 6 and Figure (not shown here for clarity).
| Discharge rate (C) | Fin number | Maximum temperature (°C) |
|---|---|---|
| 0.5 | 12 | 33.06 |
| 14 | 32.44 | |
| 16 | 31.91 | |
| 18 | 32.74 | |
| 1 | 12 | 41.06 |
| 14 | 39.83 | |
| 16 | 38.77 | |
| 18 | 40.42 | |
| 2 | 12 | 56.84 |
| 14 | 53.74 | |
| 16 | 52.37 | |
| 18 | 55.58 |
It is observed that the maximum temperature decreases as the fin number increases from 12 to 16, because a larger fin number provides a larger heat transfer area. However, when the fin number reaches 18, the airflow may be obstructed, and the thermal boundary layers merge, leading to a reduction in the local heat transfer coefficient. Consequently, the temperature rises again. Therefore, the optimal fin number is 16 regardless of the discharge rate.
5.2 Effect of Fin Thickness on Heat Dissipation Performance
By setting the fin number to 16, wind speed to 2 m/s, and varying the fin thickness from 0.4 mm to 1.4 mm, I obtain the maximum temperature values presented in Table 7.
| Discharge rate (C) | Fin thickness (mm) | Maximum temperature (°C) |
|---|---|---|
| 0.5 | 0.4 | 31.62 |
| 0.6 | 31.86 | |
| 0.8 | 31.91 | |
| 1.0 | 32.53 | |
| 1.2 | 32.98 | |
| 1.4 | 33.60 | |
| 1 | 0.4 | 38.20 |
| 0.6 | 38.68 | |
| 0.8 | 38.78 | |
| 1.0 | 40.01 | |
| 1.2 | 40.90 | |
| 1.4 | 42.11 | |
| 2 | 0.4 | 51.25 |
| 0.6 | 52.19 | |
| 0.8 | 52.38 | |
| 1.0 | 54.79 | |
| 1.2 | 56.51 | |
| 1.4 | 58.87 |
From Table 7, the fin thickness of 0.4 mm gives the lowest temperature. As the fin thickness increases, the temperature rises due to increased conduction resistance and reduced effective air gap between adjacent fins, which impedes airflow. Thus, for the present energy storage battery thermal management unit, a thin fin of 0.4 mm is preferred.
5.3 Effect of Fin Spacing
Next, I fix the fin thickness at 0.4 mm and the fin number at 16, and vary the fin spacing from 1 mm to 5 mm. The maximum temperature results are shown in Table 8.
| Discharge rate (C) | Fin spacing (mm) | Maximum temperature (°C) |
|---|---|---|
| 0.5 | 1 | 32.80 |
| 2 | 31.33 | |
| 3 | 31.62 | |
| 4 | 31.25 | |
| 5 | 31.41 | |
| 1 | 1 | 40.52 |
| 2 | 37.62 | |
| 3 | 38.12 | |
| 4 | 37.45 | |
| 5 | 37.80 | |
| 2 | 1 | 55.82 |
| 2 | 50.10 | |
| 3 | 51.25 | |
| 4 | 49.77 | |
| 5 | 50.46 |
The optimal fin spacing is 4 mm for all discharge rates. At a spacing of 1 mm, the airflow is almost blocked; at a spacing of 5 mm, the heat transfer area is reduced because fewer fins can fit in the same space. Thus, a moderate spacing of 4 mm provides a balance between heat transfer area and airflow resistance.
5.4 Effect of Heat Pipe Distribution
I also investigate the effect of the number of heat pipes uniformly distributed along the surface of the energy storage battery. The simulations are performed with the optimized fin configuration (fin number 16, fin thickness 0.4 mm, fin spacing 4 mm) and an inlet velocity of 2 m/s. The results are summarized in Table 9.
| Discharge rate (C) | Number of heat pipes | Maximum temperature (°C) |
|---|---|---|
| 0.5 | 3 | 32.21 |
| 4 | 31.91 | |
| 5 | 31.75 | |
| 6 | 31.66 | |
| 7 | 31.62 | |
| 1 | 3 | 39.38 |
| 4 | 38.77 | |
| 5 | 38.46 | |
| 6 | 38.29 | |
| 7 | 38.20 | |
| 2 | 3 | 53.56 |
| 4 | 52.37 | |
| 5 | 51.76 | |
| 6 | 51.42 | |
| 7 | 51.25 |
As expected, increasing the number of heat pipes reduces the maximum temperature because the heat extraction path is enhanced. However, the improvement becomes marginal beyond four heat pipes. Considering the added weight and cost, I select four heat pipes as the optimal configuration. In addition, a non-uniform arrangement with the four heat pipes shifted toward the center leads to a higher maximum temperature (32.47 °C) than the uniform arrangement (31.91 °C) for the 0.5 C discharge case. Therefore, the uniform distribution is preferred.
5.5 Combined Effect of Discharge Rate and Wind Speed
To evaluate the combined influence of discharge rate and wind speed, I use the optimized fin and heat pipe parameters (16 fins, 0.4 mm thickness, 4 mm spacing, 4 heat pipes) and simulate the thermal response for three discharge rates (0.5 C, 1 C, 2 C) and five wind speeds (2, 4, 6, 8, 10 m/s). The maximum temperatures are listed in Table 10.
| Discharge rate (C) | Wind speed (m/s) | Maximum temperature (°C) |
|---|---|---|
| 0.5 | 2 | 31.62 |
| 4 | 29.62 | |
| 6 | 28.94 | |
| 8 | 28.54 | |
| 10 | 28.27 | |
| 1 | 2 | 38.20 |
| 4 | 34.23 | |
| 6 | 32.87 | |
| 8 | 32.08 | |
| 10 | 31.54 | |
| 2 | 2 | 51.25 |
| 4 | 43.41 | |
| 6 | 40.70 | |
| 8 | 39.14 | |
| 10 | 38.07 |
The results reveal that the temperature decreases significantly as the wind speed increases from 2 m/s to 4 m/s. Further increases in wind speed yield diminishing returns. The temperature difference between 8 m/s and 10 m/s is very small, indicating that the convective heat transfer has nearly reached its limit. For energy saving, a wind speed of 4 m/s is recommended because it provides a good trade-off between cooling performance and fan power consumption.
5.6 Effect of Phase Change Material
To assess the contribution of PCM to the thermal management of the energy storage battery, I compare the temperature fields with and without PCM for discharge rates of 1 C, 2 C, and 3 C. The results are summarized in Table 11.
| Discharge rate (C) | With PCM (max temp. °C) | Without PCM (max temp. °C) | Temperature difference reduction (°C) |
|---|---|---|---|
| 1 | 38.2 | 39.0 | 0.8 |
| 2 | 51.2 | 52.4 | 1.2 |
| 3 | 65.8 | 67.5 | 1.7 |
The inclusion of PCM reduces the maximum temperature by approximately 1 °C and also reduces the temperature gradient inside the module. When the discharge rate is low (e.g., 1 C), the PCM does not reach its phase change temperature, so it mainly acts as a thermal mass and improves the temperature uniformity. At higher discharge rates, the PCM absorbs a substantial amount of latent heat, which suppresses the temperature rise effectively. This confirms that the PCM-enhanced heat pipe system is beneficial for the thermal management of the energy storage battery, especially under high discharge conditions.
5.7 Overall Optimal Configuration
Based on the comprehensive parametric study, I determine the optimal configuration of the energy storage battery thermal management unit as follows:
- Battery discharge rate: 0.5 C (for the given operating condition; the design is also suitable for 1 C and 2 C if the ambient temperature is lower or wind speed is increased)
- Wind speed: 4 m/s
- Number of heat pipes: 4 (uniformly distributed)
- Number of fins: 16
- Fin thickness: 0.4 mm
- Fin spacing: 4 mm
- PCM layer: 44 °C phase change paraffin with a thickness of 2 mm between adjacent cells
This configuration ensures that the maximum temperature of the energy storage battery remains below 30 °C under 0.5 C discharge conditions, and the temperature difference across the module is minimized. The thermal management unit is lightweight, compact, and reliable, making it suitable for practical energy storage applications.
6. Summary and Outlook
In this paper, I have designed and evaluated a thermal management unit for a large-format energy storage battery. The unit integrates flat heat pipes, aluminum fins, and phase change material to achieve efficient heat transfer and temperature uniformity. Through a series of ANSYS-Icepak simulations, I analyzed the effects of fin parameters, heat pipe number, discharge rate, wind speed, and PCM. The main conclusions are as follows:
- The optimal fin number is 16, fin thickness is 0.4 mm, and fin spacing is 4 mm, as this combination provides the largest effective heat transfer area without excessively impeding airflow.
- Four uniformly distributed heat pipes are sufficient to maintain the battery temperature within acceptable limits; increasing the number beyond four yields only marginal improvement.
- A wind speed of 4 m/s offers a good balance between cooling performance and energy consumption, because the temperature reduction from 4 to 10 m/s is relatively small.
- The addition of PCM reduces the maximum temperature by about 1 °C and improves the temperature uniformity, especially at high discharge rates.
- The theoretical calculations and numerical simulation results are in good agreement, confirming the feasibility of the proposed design method.
Future work will focus on:
(1) developing a more detailed electro-thermal model of the energy storage battery that accounts for the internal layered structure;
(2) performing experimental validation on a prototype module;
(3) optimizing the layout of heat pipes and fins using multi-objective methods to further reduce the temperature difference;
(4) extending the study to battery packs with multiple cells and different cooling strategies.
In summary, the proposed thermal management unit for the energy storage battery offers a simple, lightweight, and efficient solution for temperature control. It is expected to enhance the safety and longevity of the energy storage battery in practical applications.