Experimental Study of a Bionic Porous Media Evaporative Radiator Inspired by Leaf Transpiration: Exploring Energy Change Processes

Abstract

The performance of photovoltaic (PV) cells is significantly influenced by their operating temperature. While conventional active cooling methods are limited by economic feasibility, passive cooling strategies often face challenges related to insufficient heat dissipation capacity. This study presents a bio-inspired evaporative heat sink, modeled on the transpiration and water transport mechanisms of plant leaves, which leverages porous media flow and heat transfer. The device uses capillary pressure, generated through the evaporation of the cooling medium under sunlight, to maintain continuous coolant flow, thereby achieving effective cooling. An experimental setup was developed to validate the device’s performance under a heat flux density of 1200 W/m², resulting in a maximum temperature reduction of 5 °C. This study also investigated the effects of porous medium thickness and porosity on thermal performance. The results showed that increasing the thickness of the porous medium reduces cooling efficiency due to reduced fluid flow. In contrast, the effect of porosity was temperature-dependent: at evaporation temperatures below 67 °C, a porosity of 0.4 provided better cooling, while at higher temperatures, a porosity of 0.6 was more effective. These findings confirm the feasibility of the proposed device and provide valuable insights into optimizing porous media properties to enhance the passive cooling of photovoltaic cells.

1. Introduction

Energy is a fundamental resource for driving sustainable societal and economic development. In recent years, solar energy has emerged as a clean and sustainable energy source with significant applications spanning energy, power, aerospace, and everyday uses. It has become a critical component of renewable energy development strategies worldwide, aligning with the United Nations’ Sustainable Development Goal 7 (SDG-7), which aims to ensure access to affordable, reliable, sustainable, and modern energy for all. Solar energy can be utilized through several methods, including solar thermal, photovoltaic, and photochemical processes [1]. Among these, solar photovoltaic (PV) power generation plays a pivotal role in renewable energy adoption due to its ability to directly convert solar energy into electricity. This technology is instrumental in achieving SDG-7 by providing clean and sustainable energy solutions to meet the growing global demand. However, one of the significant challenges associated with PV technology is its low efficiency in converting sunlight into electricity. Only a small fraction of incident photons is successfully converted into electrical energy, while the majority are dissipated as heat. This heat accumulation leads to a rise in the temperature of the solar cells, reducing their efficiency and reliability [2,3,4]. To address the adverse effects of elevated temperatures on PV cells, cooling devices are commonly employed. These devices are typically categorized into three types: active cooling [5,6,7,8], passive cooling [9,10,11], and hybrid cooling techniques [12,13,14,15]. Active cooling systems utilize additional electrical energy to drive the cooling medium—such as water, air, or nanofluids—but are often considered economically unviable, as demonstrated by economic analyses conducted by Nižetić et al. [16] and Li et al. [17]. In contrast, passive cooling techniques, which do not require external energy input, are more economically viable and widely implemented [18]. However, the efficiency of passive cooling is constrained by the heat dissipation area and the convection heat transfer coefficient of air [19]. Recent studies by Alktranee et al. [20] and Haidar et al. [21] have shown that the performance of passive cooling systems can be significantly improved by leveraging liquid evaporation, which utilizes the latent heat released during the phase transition. This presents a promising opportunity to develop efficient passive cooling technologies that ensure the stable and efficient operation of photovoltaic cells.

Current cooling technologies for photovoltaic (PV) cells are primarily categorized into active and passive cooling. Active cooling involves the use of pumps, fans, and similar devices to remove heat from the PV panels, while passive cooling relies on natural convection, radiation, and phase changes in materials to lower panel temperatures. Bahaidarah et al. [22] introduced a water-cooling plate behind the PV panels, using pump-driven water for cell cooling. Their experimental results showed that under 1000 W/m² irradiation, the temperature of the PV cells was reduced by 20%, leading to a 9% increase in electrical efficiency. Similarly, Mojumder et al. [23], Shahsavar et al. [24], Hussain et al. [25], and Kaiser et al. [26] proposed various heat sink configurations placed behind PV cells, combined with fans for forced air convection. This approach reduced PV panel temperatures by up to 87% and improved electrical efficiency by up to 14%. Ebaid et al. [27] and Feng et al. [28] further enhanced water-cooling efficiency by adding nanoparticles, such as TiO₂, Al₂O₃, and PCM nanoparticles, to the cooling fluid, requiring at least 50% of the pumping power to achieve effective cooling. While active cooling methods can significantly lower PV panel temperatures and enhance electrical efficiency, the overall system efficiency is limited due to the energy consumption of the pumping systems. Passive cooling, which does not require additional energy input, is a widely explored alternative for PV cooling. Bayrak et al. [29] and Hernandez-Perez et al. [30] proposed the use of heat-dissipating aluminum sheets installed behind the PV panels to enhance natural convection. This approach reduced panel temperatures by up to 10 °C and improved electrical efficiency by 4%. Hussein et al. [31] investigated the use of fins in passive cooling systems, achieving 8–10% efficiency improvements and identifying potential design optimizations. Harsh et al. [32] developed a multiphysics model for concentrated photovoltaic (CPV) systems integrated with microchannel heat sinks, addressing localized heating and thermal uniformity issues. Similarly, Refaey et al. [33] demonstrated significant performance enhancement using nano-enhanced phase change materials (PCMs) combined with advanced heat sink designs, achieving a reduction in PV surface temperatures of up to 57.7%. Hassan et al. [34] explored the use of PCMs for cooling PV cells during phase change processes. By employing different PCM shell shapes and materials, their study showed that under 1000 W/m² light intensity, PV cell temperatures could be reduced by 18 °C within a short time, while a 10 °C reduction could be maintained for up to 5 h. Evaporative cooling is another promising technique that leverages the heat-absorbing properties of phase changes. Agyekum et al. [9] employed water-soaked materials on the front and back of PV panels to enable cooling through evaporation. Their results demonstrated that this method could reduce PV cell temperatures by up to 20 °C. However, water coverage on the panel surface was found to adversely affect PV efficiency. Seo et al. [35] introduced an evaporative heat sink inspired by the transpiration mechanism in plant leaves, achieving a maximum effective heat transfer coefficient of 0.058 W/cm²·K. Additionally, Helin et al. [36] and Nicola et al. [37,38] conducted topology optimization analyses on various microchannel void structures to evaluate heat transfer performance, providing valuable insights into improving thermal management for PV systems. Rajput et al. [39] analyzed multi-crystalline silicon (mc-Si), amorphous silicon (a-Si), and hetero-junction with an intrinsic thin layer (HIT) PV technology in hot semi-arid climates. Their findings showed that while HIT modules demonstrated superior energy metrics, such as lower energy payback time (EPBT) and a higher energy production factor (EPF), mc-Si and a-Si had lower annualized uniform costs (UAC). Including degradation rate (DR) in these assessments provides valuable insights into selecting suitable PV technologies for specific regions.

Building upon current research, this study proposes an evaporative heat sink design that leverages the operating characteristics of photovoltaic (PV) cells to achieve efficient cooling. Compared to existing cooling methods, the proposed device offers high cooling efficiency without requiring external energy input. Additionally, its simple structure, absence of moving parts, and low maintenance cost make it an attractive solution for PV cooling applications. The schematic diagram of the heat sink is presented in Figure 1. The design incorporates siphon flow within a hydrophilic porous medium, as depicted in the diagram. This porous medium enables the transport of cooling water from the suction end to the evaporation end, facilitating heat extraction from the hot end and promoting effective cooling and heat dissipation. To validate the feasibility of this cooling device and evaluate its performance under varying conditions, an experimental platform was constructed. The platform enabled performance testing and analysis of the cooling efficiency across different evaporation temperatures, porous media thicknesses, and porosities. This study introduces a novel approach to cooling photovoltaic cells, providing a cost-effective, energy-efficient, and sustainable solution.

2. Materials and Methods

2.1. System Description

The evaporative radiator developed in this study consists of four main components: the porous medium, the soaking plate, the thermally conductive silicone, and the cooled device (e.g., photovoltaic cells). The porous medium plays a critical role as the transport channel for the cooling medium, which is typically water. To ensure efficient flow and the absorption of the coolant, the porous medium must exhibit hydrophilic properties. Hydrophilic porous materials, such as polyvinyl alcohol sponges, are flexible but necessitate rigid structural support to preserve functionality. In the radiator design, the suction end of the porous medium is immersed in the coolant, while the evaporation section is suspended. This configuration facilitates the movement of the cooling medium through capillary action and enables effective energy exchange. The soaking plate is positioned horizontally on the surface of the porous medium, serving to evenly distribute the heat generated by the cooled device and thereby increasing the heat transfer area. However, when the rigid cooled device and the soaking plate come into contact, thermal contact resistance may arise. To minimize this resistance and enhance heat transfer, the gap between the two components is filled with thermally conductive silicone to enhance thermal conductivity and overall system efficiency.

As illustrated in Figure 2, the working principle of this cooling device for photovoltaic cells involves exposing the photovoltaic cell and porous material to sunlight. When exposed to sunlight, the photovoltaic cell converts solar energy into both electrical and thermal energy, while the porous material absorbs and converts solar energy solely into thermal energy. The heat generated by the photovoltaic cell is transferred to the soaking plate through a layer of thermally conductive silicone. From the soaking plate, the heat is transferred to the porous medium and subsequently to the cooling medium within the pores. The cooling medium, driven by continuous evaporation and capillary action, flows through the porous medium, effectively dissipating the heat generated by the photovoltaic cell and achieving efficient cooling.

2.2. Experimental Setup

Building on the outlined fundamental principles, an experimental setup was constructed, as shown in Figure 2. The cooling process of the evaporative radiator relies on water absorption at the suction end and subsequent evaporation at the evaporation end. To evaluate the radiator’s effectiveness, the experimental setup was modified as depicted in Figure 3. The setup was divided into two sections: one designated for water absorption and the other for evaporation, with the cooled module positioned adjacent to the radiator. To simulate the cooled heat source and the evaporative heat source, two heaters were employed in the setup.

The cooled unit consists of a layered structure comprising acrylic sheets, glass wool, porous media, and manually processed copper sheets. The acrylic sheet, made of polymethyl methacrylate (PMMA), serves as the outer shell of the unit, providing structural support and insulation. Glass wool, with its low thermal conductivity, is placed between the porous medium and the acrylic sheet to enhance thermal insulation. The porous medium selected for this experiment is a polyvinyl alcohol (PVA) sponge, chosen for its excellent water absorption properties, making it ideal for this application. The module to be cooled is enclosed within a 2 mm thick acrylic plate. Inside the acrylic shell, a 10 mm thick layer of glass wool serves as an insulating layer, positioned beneath the porous medium. A heater with an area of 50 mm × 50 mm and a thickness of 3 mm is positioned above the porous medium to simulate the device requiring thermal management. This heater is insulated with an additional 10 mm thick layer of glass wool. In the evaporative section of the porous medium, the lowermost layer is heated to represent the heat source for evaporation. A copper block is placed above the heater to ensure uniform heat distribution. The small gap between the copper block and the heater is filled with a thin layer of thermally conductive silicone with a thermal conductivity of 4 W/m·K. The copper block is enclosed within the evaporative section of the porous medium. As the heater warms the porous medium, the liquid within it evaporates. Simultaneously, the porous medium absorbs liquid from the reservoir, which is continuously circulated through the cooled device via the capillary effect, thereby achieving the cooling effect.

Figure 3. The diagram of the experimental setup: (a) the schematic diagram of the structure of the experimental setup; (b) the physical diagram of the experimental setup.

Figure 3. The diagram of the experimental setup: (a) the schematic diagram of the structure of the experimental setup; (b) the physical diagram of the experimental setup.

The physical diagram of the experimental setup is presented in Figure 3. The high-precision analytical balance used in this study is the JCS-W model, manufactured by Harbin Zhonghui Weighing Instrument Co (Haebin, China). It has a maximum range of 500 g, an accuracy of 0.001 g, and an error rate of ±0.003 g. Both heaters employed in the setup are polyimide electric heating films. The simulated cooled heater is powered by a 24 V DC source, with a maximum power output of 24 W and a maximum temperature of 120 °C. In contrast, the evaporation heater operates on a 220 V AC power supply, generating a maximum power of 100 W and reaching a maximum temperature of 150 °C. The cooled heater is connected to an external DC adjustable power supply for precise voltage control, ensuring consistent power output during operation. Meanwhile, the evaporation heater is equipped with an external temperature control switch, enabling precise regulation of the evaporation temperature. Temperature measurements were conducted using a three-wire Resistance Temperature Detector (Pt100), with a temperature range of −50 °C to 200 °C and an error rate of ± (0.15 + 0.002T). Four temperature sensors were strategically positioned within the setup:

• T₁: located in the porous medium in front of the cooled module.
• T₂: located in the porous medium behind the cooled module.
• T₃: positioned between the evaporative heater and the copper block.
• T₄: positioned between the cooled device and the porous medium.

Additionally, two temperature sensors monitored the ambient temperature and the temperature of the liquid in the reservoir. To maintain stable liquid absorption and evaporation, various super-absorbent polyvinyl alcohol sponges with differing porosities were selected as the porous media. These sponges facilitated effective liquid transport and evaporation during the cooling experiments.

2.3. Theoretical Model

The cooled module is simplified according to the described scenario, and the corresponding physical model is presented in Figure 4. In this model, the heater of the cooled device applies constant power to the upper surface of the porous medium, while capillary forces drive the flow of the cooling medium from the left side to the right side. The lower surface of the porous medium is assumed to be an adiabatic wall.

To analyze the system, conservation equations for momentum and energy are established using the representative elementary volume (REV) approach. The following assumptions are made, drawing on analyses from previous studies [40,41,42]:

1. The porous skeleton is homogeneous, and the particles within it are modeled as rigid spheres of uniform size.

2. The experimental process is considered quasi-static, with the heating, ambient, and cooling fluid inlet temperatures remaining constant over time.

3. The physical properties of the porous skeleton and the cooling fluid are assumed to be constant.

4. The fluid enters the porous medium at the inlet with a uniform flow rate.

5. The skeleton and the fluid are treated as continuous and homogeneous media, with an equal convective heat transfer coefficient throughout.

(1) Continuity equation

A continuity equation is an expression of the conservation of mass, and it is as follows:

$${\displaystyle \frac{\partial u}{\partial x}}+{\displaystyle \frac{\partial v}{\partial y}}=0 ,$$

(1)

(2) Momentum equation

The flow process of a fluid in porous medium is described using a modified Darcy’s law averaged over the REV [43].

$${\displaystyle \frac{{\rho }_f}{\varphi }}\left(\overrightarrow{V}\cdot \nabla \overrightarrow{V}\right)=-\nabla p+{\displaystyle \frac{{\mu }_f}{\varphi }}{\nabla }^2\overrightarrow{V}-{\displaystyle \frac{{\mu }_f}{K}}\overrightarrow{V}-{\rho }_f{\displaystyle \frac{F\varphi }{\sqrt{K}}}\left|\overrightarrow{V}\right|\overrightarrow{V} ,$$

(2)

where K denotes the permeability of porous media and F denotes the Brinkman term correction factor [44].

$$K={\displaystyle \frac{{\varphi }^2{d}_p^2}{150{(1-\varphi )}^2}} ,$$

(3)

$$F={\displaystyle \frac{1.75}{\sqrt{150}{\varphi }^{3/2}}} ,$$

(4)

(3) Energy equation [45]

The solid skeleton of the porous medium and the flowing medium within it are treated as two distinct continuous media, with heat transfer within the porous structure regarded as interphase heat transfer.

The equation for the conservation of energy in the solid phase is as follows:

$${\lambda }_{s,eff}\left({\displaystyle \frac{{\partial }^2{T}_s}{\partial x^2}}+{\displaystyle \frac{{\partial }^2{T}_s}{\partial y^2}}\right)-h_v\left(T_s-T_f\right)=0 ,$$

(5)

The equation for the conservation of energy in the fluid phase is the following:

$${\displaystyle \frac{\partial \left(C_f{\rho }_f{\mu }_f{T}_f\right)}{\partial x}}+{\displaystyle \frac{\partial \left(C_f{\rho }_f{v}_f{T}_f\right)}{\partial y}}-{\lambda }_{f,eff}\left({\displaystyle \frac{{\partial }^2{T}_s}{\partial x^2}}+{\displaystyle \frac{{\partial }^2{T}_s}{\partial y^2}}\right)+h_v\left(T_s-T_f\right)=0 ,$$

(6)

where

$${\lambda }_{s,eff}=\left(1-\varphi \right){\lambda }_s ,$$

(7)

and

$${\lambda }_{f,eff}=\varphi {\lambda }_f ,$$

(8)

$h_v$ is the volume convection heat transfer coefficient between the flid and the solid skeleton of the porous medium [40]:

$$h_v={\displaystyle \frac{6h\left(1-\varphi \right)}{d_p}},$$

(9)

where the Nusselt number can be expressed as follows [46]:

$$Nu={\displaystyle \frac{h{d}_p}{{\lambda }_f}}={\left[{\left(1.18R{e}^{0.58}\right)}^4+10.23R{e_h}^{0.75}\right]}^{0.25} ,$$

(10)

$$Re=\varphi \rho u_p{d}_p R{e}_h={\displaystyle \frac{Re}{1-\varphi }} ,$$

(11)

The boundary conditions for the equation are express as

$$\begin{gathered} x=0 : u=u_{in} , v=0 , T_s=T_f=T_{in} , \\ x=L : {\displaystyle \frac{\partial u}{\partial x}}=0 , {\displaystyle \frac{\partial v}{\partial y}}=0 , {\displaystyle \frac{\partial T_s}{\partial y}}=0,{\displaystyle \frac{\partial T_f}{\partial y}}=0 \\ y=0 : {\displaystyle \frac{\partial u}{\partial x}}=0 , v=0 , {\displaystyle \frac{\partial T_s}{\partial y}}=0,{\displaystyle \frac{\partial T_f}{\partial y}}=0 \\ y=\delta : {\displaystyle \frac{\partial u}{\partial x}}=0 , v=0 ,-{\lambda }_{s,eff}{\displaystyle \frac{\partial T_s}{\partial y}}=q_w \end{gathered}$$

(12)

2.4. Numerical Calculation and Comparison of Results

Prior to investigating the cooling performance of the evaporative heat exchanger, the results of numerical calculations were compared and verified using experimental data. To evaluate the agreement between the numerical calculations and the experimental measurements, the root mean square deviation (RMSD) [22] was used:

$$\mathrm{RMSD}=\sqrt{{\displaystyle \frac{\sum {\left[\left(X_{sim,i}-X_{exp,i}\right)/X_{exp,i}\right]}^2}{n}}} ,$$

(13)

where $X_{sim,i}$ and $X_{exp,i}$ are the simulated and experimental values, respectively, and n is the number of experimental data.

Given the actual solar radiation on the ground, approximately 1300 W/m², and the physical properties of the porous media used in the experiment, the parameters used for the calculations are as follows:

• The thermal conductivity of the solid phase in the porous media (λ_s) is 0.2 W/(m∙K).
• The heating power applied to the upper surface (q_w) is 1200 W/m².
• The porosity of the material (ϕ) is 0.4.
• The chosen mass flow rates correspond to experiments conducted with 1 mm porous media.

The numerical calculations divide the x-direction into 100 nodes and the y-direction into 10 nodes.

As depicted in Figure 5, both the experimental and the simulation results demonstrate a decrease in the temperature of the heating surface as the mass flow rate increases; however, there is a maximum deviation of 9.38% between the experimental and simulation results. Meanwhile, the RMSD of the mean temperature of the upper surface is 8.01%. Considering the numerous simplifications and assumptions made in the numerical simulation, as well as potential errors in the experimental measurement process, the experimental results can reasonably be considered plausible.

2.5. Course of the Experiment and Processing of Data

Experimental studies were conducted to evaluate the cooling performance under various conditions by varying the evaporation section temperatures, porous medium thicknesses, and porosities. The specific steps of the experiment are outlined as follows:

Step 1:

The experiment began by turning on the cooled heating plate while keeping the evaporation heating plate turned off. The voltage for the cooled heating plate was set to 10 V using a constant voltage DC power supply, which indicated a current of 0.3 A, resulting in a heating power of 3 W for the cooled heating plate.

Step 2:

The initial temperature, denoted as T_4,0, was recorded when the lower surface temperature of the cooled heating plate stabilized at a constant value.

Step 3:

The evaporation heating plate was activated, and its temperature controller was adjusted to maintain temperatures of 60 °C, 62.5 °C, 65 °C, 67.5 °C, and 70 °C.

Step 4:

The following data were recorded once the lower surface temperature of the cooled heating plate stabilized:

• Lower surface temperature of the cooled heating plate (T_4,60, T_4,62.5, T_4,65, T_4,67.5, and T_4,70).
• Inlet temperature of the cooling medium (T_1,60, T_1,62.5, T_1,65, T_1,67.5, and T_1,70).
• Outlet temperature of the cooling medium (T_2,60, T_2,62.5, T_2,65, T_2,67.5, and T_2,70).
• Evaporation temperature (T_3,60, T_3,62.5, T_3,65, T_3,67.5, and T_3,70).
• Mass of the cooling medium recorded by the analytical balance (M₆₀, M_62.5, M₆₅, M_67.5, and M₇₀).

Data collection was performed using a data logger at intervals of 5 s, and each evaporation temperature condition was recorded continuously for 30 min.

Step 5:

The above experimental procedure was repeated by replacing the porous medium with materials of different thicknesses and porosities.

To investigate the effect of different porous medium thicknesses and porosities on the cooling performance, the experiments were conducted using porous media with thicknesses of 1 mm, 2 mm, and 3 mm, and porosities of 0.2, 0.4, and 0.6, respectively. Throughout the experiments, the heating power of the cooled heating plate was kept constant, while temperatures at various points and changes in the mass of the cooling medium were recorded under different evaporation temperatures.

The recorded temperatures at points T₁, T₂, T₃, and T₄ were averaged over a 30 min period for each experimental condition. The change in the mass of the cooling medium was measured as the difference in its value over the same 30 min duration.

The experimental results were primarily evaluated based on the cooling performance. The key parameters of interest were the temperature change at T₄ after activating the evaporation heater and the cooling power of the cooled device.

The temperature change at point T₄ is denoted as $\Delta T_4$:

$$\Delta T_{4,i}=\left|T_{4,i}-T_{4,0}\right| ,$$

(14)

Here, $T_{4,0}$ represents the stabilized temperature at point T₄, when the cooled heating plate is turned on, but the evaporation heating plate remains off. The variable i denotes the evaporation temperature when the evaporation heating plate is activated, with the values set at 60 °C, 62.5 °C, 65 °C, 67.5 °C, and 70 °C.

The cooling power ($P_i$) represents the amount of heat removed from the cooled plate per unit of time by the radiator. It is expressed as follows:

$$P_i={\displaystyle \frac{\left|T_{2,i}-T_{1,i}\right|\times C_p}{30\times 60}}\times M_i ,$$

(15)

where

• $M_i$ is the mass flow rate of the cooling medium (g/h);
• $C_p$ is the constant-pressure specific heat capacity of water, taking the value of $4.1785 \mathrm{J}/\mathrm{g}·\mathrm{K}$;
• $T_{2,i}$ and $T_{1,i}$ are the inlet and outlet temperatures of the cooling medium, respectively.

The unit of cooling power Pᵢ is watts (W).

3. Results and Discussion

The experiments were conducted to evaluate the cooling performance of the device by varying the evaporation temperature, the thickness, and the porosity of the porous media. The mass flow rate of the cooling medium and the evaporation performance of the porous medium were assessed by monitoring changes in the cooling medium’s mass with an analytical balance. The cooling performance of the device was primarily evaluated based on the temperature difference between the front and back of the cooled heating plate and the cooling power.

3.1. Different Evaporation Temperatures

Experimental data for a porous medium with a thickness of 1 mm and a porosity of 0.4 were used to examine the relationships among the evaporation temperature, the mass flow rate of the cooling medium, the temperature drop across the cooled heating plate, and the cooling power. The trends are illustrated in Figure 6. The results show that increasing the evaporation temperature results in increases in the mass flow rate of the cooling medium, the temperature drop across the cooled heating plate, and the cooling power. When the evaporation temperature was increased from 60 °C to 70 °C, the mass flow rate increased from 40.86 g/h to 53.68 g/h, the temperature drop across the cooled heating plate rose from 5.33 °C to 10.57 °C, and the cooling power improved from 1.10 W to 1.44 W. These findings demonstrate that using a 1 mm thick porous medium in this device results in a significant surface temperature reduction exceeding 5 °C in areas with a heat flow density of 1200 W/m², without the need for additional pumping power.

The analysis focuses on the mechanisms behind the cooling effect, as well as the trends observed in the graph. In the porous medium on the right side of the evaporation heating plate, the cooling medium undergoes heat-induced evaporation, generating capillary pressure within the porous medium. This capillary pressure drives the cooling medium to flow continuously from the left side of the device to the right, carrying heat away from the cooled heater. As the evaporation temperature increases, the cooling medium’s evaporation rate increases, resulting in a higher mass flow rate within the porous medium. This increased flow removes more heat per unit of time from the cooled heating plate, leading to a reduction in its temperature and an overall enhancement of the device’s cooling power.

In theory, the module to be cooled should ideally be surrounded by thermal insulation material, acting as an adiabatic boundary. Under such ideal conditions, the cooling power would match the heating power of the cooled heating plate after achieving thermal equilibrium. However, in practice, the thermal insulation material cannot provide perfect adiabatic conditions. Additionally, heat dissipation occurs through the porous medium running through the center of the module. As a result, the cooling power is lower than the heating power of the cooled heating plate. Despite this discrepancy, it remains valuable to analyze the variations in cooling power under different conditions, as they provide meaningful insights into the performance of the cooling device.

3.2. Different Porous Medium Thicknesses

The experimental results indicate that the thickness of the porous medium significantly affects the cooling performance of the device, as shown in Figure 7. In general, the cooling effect decreases as the thickness of the porous medium increases. The mass flow rate, temperature drop, and cooling power exhibit similar trends. For a porous medium thickness of 1 mm, these parameters exhibit a linear relationship with temperature. However, at thicknesses of 2 mm and 3 mm, the variation curves for these parameters level off as the evaporation temperature increases, indicating that the rate of change decreases with increasing temperature. Additionally, as the thickness of the porous medium increases, the cooling performance of the device exhibits reduced sensitivity to evaporation temperature changes.

This trend is evident in Figure 7, where it can be observed that for thicknesses of 2 mm and 3 mm, the cooling power eventually reaches its maximum value as the evaporation temperature continues to rise. This maximum cooling power provides a valuable reference for the future design and construction of a complete cooling system. In contrast, a maximum cooling power is not clearly observed for a thickness of 1 mm. This is attributed to the more efficient evaporation facilitated by a 1 mm porous medium, requiring higher evaporation temperatures for the cooling power to reach its maximum value.

The mass flow rate of the cooling medium decreases as the thickness of the porous medium increases. The maximum reduction in mass flow rate was approximately 14% and 19% for porous medium thicknesses of 2 mm and 3 mm, respectively, compared to 1 mm. This implies that a thinner porous medium at the same evaporation temperature results in a higher evaporation rate in the evaporation section. This phenomenon occurs because a thinner porous medium more effectively transfers heat from the heating plate to the upper surface, thereby accelerating the rate of evaporation.

For the same cross-sectional flow area, a higher mass flow rate of the fluid results in a higher flow velocity. However, assuming that the voids in the porous medium are evenly distributed, increasing the thickness of the porous medium leads to an increased cross-sectional flow area due to the additional pores. With a lower mass flow rate and a larger cross-sectional flow area, the flow velocity of the fluid within a thicker, porous medium is reduced. The calculations reveal that the maximum reduction in fluid flow rate is approximately 28% and 57% for thicknesses of 2 mm and 3 mm, respectively, compared to 1 mm. Since heat transfer within porous media is primarily dominated by convective heat transfer between the fluid and the porous skeleton, lower flow velocities result in less effective convective heat transfer. This is the primary reason for the observed reduction in temperature drop and cooling power.

The experimental data show that increasing the thickness of the porous medium causes the cooling efficiency to level off more quickly. Therefore, to achieve optimal cooling performance, the evaporation temperature should be maximized while minimizing the thickness of the porous medium. When the porous medium thickness is 1 mm and the evaporation temperature is 70 °C, the maximum temperature drop of the cooling radiator is 10.57 °C, and the maximum cooling power is 1.45 W. However, it is important to note that as the evaporation temperature increases and the porous medium thickness decreases, the flow channels of the cooling medium may be affected, which could weaken the overall cooling performance. Thus, selecting an appropriate combination of evaporation temperature and porous medium thickness is crucial for achieving the best cooling performance.

3.3. Porous Media with Different Porosities

Figure 8 illustrates the variations in mass flow rate, temperature drop, and cooling power as functions of porous media porosity and evaporation temperature. However, the extent of these changes varies significantly across the different parameters. As shown in Figure 8a, the mass flow rate increases with the evaporation temperature for all porosities. At a porosity of 0.2, the mass flow rate exhibits minimal changes and eventually levels off. For a porosity of 0.4, the mass flow rate shows a nearly linear correlation with evaporation temperature. In contrast, for a porosity of 0.6, the mass flow rate increases more rapidly as the evaporation temperature rises. Overall, the mass flow rate increases with both evaporation temperature and porosity.

Figure 8b,c reveal similar trends for temperature drop and cooling power as those observed for mass flow rate. For a porosity of 0.2, the rate of change for these parameters decreases with increasing evaporation temperature. At a porosity of 0.4, the rate of change remains relatively constant, while at a porosity of 0.6, the rate of change intensifies with rising evaporation temperature. However, the relationship between porosity and temperature drop or cooling power is not strictly incremental. At evaporation temperatures below 67 °C, porous media with a porosity of 0.4 exhibit greater temperature drops. Conversely, at evaporation temperatures above 67 °C, porous media with a porosity of 0.6 exhibit greater temperature drops.

From the perspective of cooling power, an increase in porosity generally results in higher cooling power at the same evaporation temperature. However, at lower evaporation temperatures, the difference in cooling power between porosities of 0.4 and 0.6 is negligible. As the evaporation temperature increases, this difference becomes more pronounced, indicating that higher porosities yield improved cooling performance.

The reasons behind the observed phenomenon are analyzed as follows:

First, the variation in mass flow rate is examined. In the experiments, the thickness of the porous media remained constant, with only the porosity being varied. In porous materials with lower porosity, the fluid flow channels are narrower, leading to higher flow resistance and lower mass flow rates. Additionally, the rate of change in the mass flow rate is directly related to the maximum flow rate achievable within the channels. Smaller flow channels cause greater resistance, which limits the maximum flow rate. This relationship is evident in Figure 8a, where the mass flow rate for a porosity of 0.2 levels off at an evaporation temperature of 70 °C, indicating that the maximum flow rate has been reached. Conversely, for porosity values of 0.4 and 0.6, the maximum flow rate has not yet been reached, even at an evaporation temperature of 70 °C.

Second, the relationship between temperature drop and mass flow rate is not strictly linear. As discussed in Section 3.2 regarding the impact of increased thickness, the flow cross-sectional area per unit volume increases with higher porosity. As a result, an increase in mass flow rate does not necessarily translate to a proportional increase in the flow velocity of the cooling medium within the pores. Consequently, the cooling effect cannot be determined solely by porosity.

This is further illustrated in Figure 8b,c. At lower evaporation temperatures, porous media with lower porosity may exhibit better cooling performance because narrower flow channels offer greater thermal resistance. However, at higher evaporation temperatures, porous media with higher porosity show better performance. This is because higher porosity allows for a greater maximum mass flow rate, enhancing the cooling effect under elevated temperature conditions.

4. Conclusions

This study details the design and experimental validation of a bio-inspired evaporative cooling device. The device’s cooling performance was evaluated under varying evaporation temperatures, porous media thicknesses, and porosities. The key conclusions are as follows:

1. The bio-inspired structure of the cooling device demonstrates significant cooling performance. Specifically, when the porous medium has a thickness of 1 mm and a porosity of 0.4, the device can reduce the temperature of an object with a heat flux density of 1200 W/m² by up to 5 °C.

2. The cooling effect improves as the evaporation temperature increases; however, it eventually plateaus as a result of resistance within the void channels of the porous media, which limits the mass flow rate and, consequently, the cooling efficiency.

3. Increasing the thickness of the porous medium adversely affects the cooling performance. Thicker porous media slow down fluid flow within the voids, reducing heat transfer efficiency. To optimize the cooling effect, selecting a thinner porous medium as the transport layer for the cooling fluid is critical.

4. The cooling performance is influenced by both porosity and evaporation temperature. At evaporation temperatures below 67 °C, a porosity of 0.4 provides better cooling performance. Conversely, at temperatures above 67 °C, a porosity of 0.6 is more effective. This behavior is attributed to porosity’s impact on the maximum mass flow rate of the fluid, with higher mass flow rates improving cooling performance at elevated temperatures.

In conclusion, this study demonstrates the feasibility and effectiveness of the radiator experimentally. The cooling performance is closely related to the evaporation temperature, porous media thickness, and porosity. This research provides a novel approach to passive cooling of photovoltaic panels.

However, this study has some limitations that were not addressed in practical engineering scenarios, including system-wide thermal efficiency, the influence of the salt content, and creep thermal expansion, which require further investigation. The evaporative cooler device should be integrated with photovoltaic panels to evaluate its cooling performance under actual working conditions. Future numerical simulations should be conducted to explore the influence of additional parameters and determine the optimal cooling conditions to improve device performance.