Experimental Investigation of Flash Spray Cooling for Power Electronics

Abstract

Power electronics convert and control electrical power in applications ranging from electric motors to telecommunications and computing. Ongoing efforts to miniaturize these systems and boost power density demand advanced thermal management solutions to maintain optimal cooling and temperature control. Spray cooling offers an effective means of removing high heat fluxes and keeping power electronics within safe operating temperatures. This study presents an experimental investigation of flash spray cooling in a closed-loop system using R410A refrigerant. In particular, two nozzles with different spraying angles are used to study the effects of the distance between the spray nozzle and a heated flat surface, as well as the mass flow rate of the coolant. Results indicate that three key flow-pattern factors—surface coverage, impingement intensity, and liquid film dynamics—govern the heat transfer mechanisms and determine cooling efficiency. Flash spray cooling using refrigerants like R410A demonstrates strong potential as a high-performance thermal management strategy for next-generation power electronics.

1. Introduction

In recent years the development of the power electronics industry in technological areas such as communication technologies, renewable energy systems, electric vehicles, electricity distribution, etc., has led to the need to develop innovative cooling techniques to dissipate significant amounts of heat. Power electronics are characterized by intense thermal energy generation, which leads to high temperature fluctuations, resulting in the thermal fatigue of devices. For example, heat fluxes in the order of 1000 W/cm² have been observed in laser diode arrays. Undermining efficient heat dissipation causes the appearance of overheating points, which lead electronic components to failure. Increasing the temperature of a component from 25 to 75 °C increases the failure rate by five times, while the reliability of power electronics is halved each time their temperature increases by 10 °C [1,2]. Innovative cooling techniques are required to support the continuous reduction in the volume of electronic devices and the increase in power density, keeping component temperatures below the permissible limits set by operational requirements, construction materials, and the desired service life.

The new techniques developed utilize two-phase refrigerants, aiming to exploit both sensible and latent heat for more efficient heat dissipation. Among these techniques, cooling by refrigerant spray provides the ability to dissipate significant amounts of heat with relatively low refrigerant flows, while achieving uniform distribution of low temperatures on the surfaces of electronic components. Such solutions are very promising, especially when a refrigeration system is already in place or easily implemented, as in the case of data centers facing the high-density cooling requirements of AI workloads or cooling power electronics of electric vehicles.

Yan et al. [3] elaborate on four heat transfer mechanisms proposed previously [4,5,6,7], that predominate during spray cooling: (a) evaporation at the surface of the liquid film, (b) forced convection due to the impact of droplets on the heated surface, (c) enhancement of nucleation sites on the heated surface, and (d) creation of secondary nucleation sites on the surface of spray droplets. Cheng et al. [8] report that secondary bubbles activated by secondary nucleation points are due to the impact and penetration of droplets into the liquid film. The incoming droplets carry along vapors created due to evaporation during their flight, which act as nucleation points for the formation of new small bubbles [3,9]. At the same time, the impact of droplets on the surface assists the release of bubbles [8]. In general, many researchers converge on the view that the dominant heat transfer mechanism during spray cooling is the creation of secondary nucleation points [4,6,7,10]. The flow phenomena that develop near the surface as well as the use of modified surfaces or additives in fluid operation appear to have a significant effect on spray cooling performance [9]. In this context, the present study aims to experimentally investigate the influence of spray dynamics—specifically mass flow rate and nozzle-to-surface distance—on the cooling performance of refrigerant spray systems, without employing surface modifications or additives. Using R410A as the working fluid, the objective is to analyze how variations in spray impingement characteristics affect Critical Heat Flux (CHF) and Heat Transfer Coefficient (HTC) in a controlled experimental setup. The findings are further contextualized through a quantitative comparison with the existing literature to identify the key factors governing spray-cooling efficiency and potential pathways for performance enhancement.

Table 1. Spray-cooling results from previous studies.

Reference	Refrigerant	Surface Modification	CHF (W/cm2)	HTC (W/cm2 °C)	Surface Temperature (°C)	ΔCHF (%)	ΔHTC (%)
Carneiro and Barbosa, 2025 [11]	R-134a R-1234yf R-600a	-	47.1 43.4 27.6	4.29 4.34 2.58	<50	−80.4% −81.9% −88.5%	−68.0% −67.6% −80.7%
Wang et al., 2023 [12]	R410A	-	175.0	9.04	<35	–27.1%	–32.5%
Zhou et al., 2019 [13]	R410A	YES	330	30.0	<10	+37.5%	+123.9%
Lin et al., 2019 [14]	R410A	-	264	21.0	<30	+10.0%	+56.7%
Bostanci et al., 2018 [15]	R-134a HF0-1234yf	YES	370/ 300	8.0/ 6.4	40/ 43.4	+54.2%	–40.3%
Liu et al., 2018 [16]	R-134a	YES	102	3.1	-	–57.5%	–76.9%
Chen et al., 2015 [17]	R-134a R22	-	117.2 /276.1			–51.2%	-
Yata and Bostanci, 2017 [18]	HFE-7100	YES	190	9.26	≈91.6	–20.8%	–30.9%
Yaddanapudi and Bostanci, 2015 [19]	R-134a HF0-1234yf	YES (macro-structures)	270/ 190	14.1/7.6	≈12/ ≈28	+12.5%	+5.2%
Xu et al., 2014 [20]	R600a	-	110	3.0	31.5	–54.2%	–77.6%

provides a comparative overview of the spray-cooling performance metrics reported in previous experimental studies, specifically focusing on CHF, HTC, and surface temperature. To facilitate a direct comparison with the present investigation, two additional metrics are incorporated, presenting the percentage differences in CHF (ΔCHF) and HTC (ΔHTC) relative to the peak values obtained in this study (CHF ≈ 240 W/cm², HTC ≈ 13.4 W/cm²·K).

These comparative metrics elucidate the influence of various factors—such as surface modifications, refrigerant properties, spray distribution characteristics, and experimental configurations—on the overall cooling performance. Notably, studies employing micro-structured surface enhancements, such as the research by Zhou et al. [13], demonstrated significant improvements, achieving a 37.5% increase in CHF and a 123.9% enhancement in HTC compared to the present results. Conversely, configurations without surface modifications, such as those examined by Lin et al. [14], Wang et al. [12], and Carneiro and Barbosa [11], exhibited either moderate performance gains or reductions, predominantly governed by the degree of optimization in spray dynamics and nozzle-to-surface interactions. The disparities observed in Table 1 underscore the pivotal role of surface engineering and precise control over spray parameters in maximizing spray-cooling efficiency, particularly in high-power electronic applications where thermal management is critical. These comparisons illustrate that, although surface engineering can markedly enhance heat transfer performance, the precise tuning of spray dynamics—such as nozzle positioning, mass flow rate, and droplet impingement intensity—can also yield significant improvements, even on untreated surfaces. Accordingly, the present study systematically investigates these spray parameters to optimize cooling efficiency without relying on surface modifications or additives.

2. Experimental Device and Procedure

2.1. Spray Chamber and Heating System

A key element of the experimental device is the spray chamber (Figure 1a and Figure 2). A spray system for cooling power electronic devices is expected to have a modular and compact design for production applications. Actual configurations will be optimized once the critical spray and geometry influence have been identified. For the time being, in the present experiment, a larger cavity is used to facilitate parametric investigation of the effect of several control parameters, such as the use of different spray nozzles and the variation of the distance between the spray outlet and the heated surface. A rather compact cylindrical spray chamber with diameter 30 mm is used, which allows optical access to the spray through Plexiglas windows on two adjacent sides. The spray is formed by either of two Danfoss solid cone nozzles (OD 5 USgal/h or 18.93 L/h) with nominal spray angles for oil of 60° or 45°, located at the top of the chamber, at a controlled distance from the heated surface. A round smooth copper surface of 17 mm diameter, at the base of the chamber, substitutes for the surface of an electronic power device from which heat is dissipated. The diameter was selected to provide a similar cooling surface area to typical power electronic devices (e.g., AUIRGPS4067D1 in TO-274 package, Infineon Technologies AG, Neubiberg, Germany). This area is the upper surface of the heating system, the base of which incorporates four resistors with a total electrical power of 4 × 400 W (Figure 1b and Figure 2), enclosed in an aluminum casing. The upper surface of the casing is tightly connected to a copper block, with thermally conducting paste in between. Heat is conducted through the solid copper block, whose upper section is cylindrical (neck, Figure 1); it is thermally insulated around the perimeter with Peek insulating material, to ensure one-dimensional heat flow in its longitudinal direction. The Peek material reaches to the upper edge of the copper cylinder and descends radially to avoid flooding of the heated surface. To determine the heat flux along the copper neck, two K-type thermocouples are fixed on its central axis, at a distance of 7 mm from each other. The spray chamber and heating system are covered with thick layers of insulation material to prevent heat losses to the environment. During the operation of the experimental device, the heat flux, conducted from the electrical resistors through the copper neck to the refrigerant, is regulated with a PID controller reading the output of a thermocouple located at the base of the copper block.

2.2. Refrigeration Loop

The experimental device consists of a closed, modified refrigeration circuit (Figure 3), which utilizes elements of a domestic air-conditioning unit, with a thermal power of 3.5 kW (12,000 Btu/h). The circuit is equipped with appropriate devices and sensors to control and record the operation of the system as shown in Figure 3. The design of the circuit allows for a comprehensive investigation of the effect of several parameters on the performance of the technique. The implementation on an end operating system will be significantly simpler based on optimized control strategies, depending on the particular requirements of each application. The choice of refrigerant (R410A) was based on the appropriate fluid thermal and environmentally friendly characteristics [12,14].

The operation of the closed refrigeration loop can be described with the help of the indicative thermodynamic cycle presented in Figure 4 (capital letters correspond to the components in Figure 3). Point 1 represents the inlet to the compressor, A, which compresses the refrigerant vapor to high pressure and temperature (point 2). The overheated vapor from point 2 enters the condenser, B, exchanging heat with the ambient, and the heat exchanger, C, exchanging heat with the return loop to the compressor inlet (point 1). These two apparatuses are connected in parallel, with corresponding outlets 3 and 4, which are mixed at this location. The liquid mixture undergoes a pressure reduction through the expansion valve, D, and point 5 corresponds to the two-phase mixture in the separator, E. Liquid from the separator is driven to the inlet of the spray nozzle, 6, whereas some of the vapor in the separator, 10, may flow through suitable valves and a pressure regulator to the outlet of the spray chamber, 11. The liquid from point 6 undergoes expansion to a two-phase mixture in the spray nozzle at point 7 and then adjusts its pressure to the pressure of the spray chamber at point 8. From point 8 to point 9, the refrigerant mixture is heated under constant pressure and temperature, cooling the heated surface. Point 9’ represents the two-phase mixture resulting from mixing the two-phase mixture of point 9 and the superheated vapor of point 11. From this point to the compressor inlet, 1, the mixture is heated to reach outside the two-phase area, exchanging heat with the high-temperature refrigerant in the heat exchanger, C.

2.3. Main Variable Estimation and Uncertainty

Two K-type thermocouples, as depicted in Figure 1b, measured the temperature differential on the copper cylinder; these data were used to compute both the heat flux through the cylinder and the heat transfer coefficient at the cylinder–coolant interface. According to the Fourier law for one-dimensional conduction, the heat flux, q, is calculated using Equation (1):

$$\mathrm{q}=\mathrm{k}{\displaystyle \frac{\mathsf{\Delta }\mathrm{T}}{\mathsf{\Delta }\mathrm{x}}}$$

(1)

where ΔT and Δx are the temperature difference and distance between the two thermocouples T₂ and T₁, respectively, in Figure 1b. The vertical distance between the thermocouples is 7 mm. The surface temperature T_w and the heat transfer coefficient h can be calculated using Equations (2) and (3):

$${\mathrm{T}}_{\mathrm{w}}={\mathrm{T}}_1-{\displaystyle \frac{\mathrm{q}{\mathsf{\Delta }\mathrm{x}}_{\mathrm{w}-1}}{\mathrm{k}}}$$

(2)

$$\mathrm{h}={\displaystyle \frac{\mathrm{q}}{{\mathrm{T}}_{\mathrm{w}}-{\mathrm{T}}_{\mathrm{s}\mathrm{a}\mathrm{t}}}}={\displaystyle \frac{\mathrm{q}}{\mathsf{\Delta }{\mathrm{T}}_{\mathrm{s}\mathrm{a}\mathrm{t}}}}$$

(3)

where Δx_w−1 is the distance between the thermocouple T₁ and the upper surface of the copper cylinder. The temperature of the upper surface of the copper cylinder is T_w, while T_sat is the saturation temperature in the spray chamber, corresponding to the saturation pressure of the coolant.

Uncertainty of the thermocouple reading is approximately ±2.5 °C, across the temperature range of −30 °C to 250 °C. Pressure was measured with two Carel SPKT0 sensors: the E3P model (0.10–1.28 MPa range) has a full-scale uncertainty of ±0.01656 MPa, and the E6P model (0.00–4.50 MPa range) has a full-scale uncertainty of ±0.054 MPa. The uncertainty of the distance between the thermocouples T₁ and T₂ is ±0.1 mm, while the flow meter sensor has a full uncertainty range of ±5% across the range of 0.3 to 3 L/min. According to the uncertainty estimation theory [21], the heat flux measurement uncertainty for an average value of 200 W/cm² is ±7.00%, while the uncertainties of surface temperature T_w and heat transfer coefficient h are ±8.16% and ±10.8%, respectively. Monitoring the heat flux variations for specific temperature settings of the PID controller (SRS1, Shimaden Co., Ltd., Kitamachi, Nerima-ku, Tokyo, Japan) showed standard deviations of less than 1% in all cases. To provide a visual indication of the uncertainty, without cluttering the data, error bars are included only for one data set in the following diagrams.

3. Experimental Results

3.1. Results for 60° Spray Nozzle

Diagrams presented in Figure 5 depict the results for the 60° spray nozzle. Data curves on each diagram correspond to three different mass flow rates of the coolant, m, (4.5 g/s—blue continuous line; 5.5 g/s—green dashed line; and 6.0 g/s—red dotted/dashed line). The pair of diagrams in each row refer a specific distance between the nozzle outlet and the surface: H = 15, 20, 25, 30, and 35 mm. The first column, (a), shows the correlation of heat flux, q, to overheating, i.e., the difference between the surface temperature and the saturation temperature. The second column, (b), includes diagrams of the heat transfer coefficient, h, relative to the heat flux, q. Symbols of each measurement curve, as indicated in the legend, refer to the distance between the nozzle outlet and the heated surface, H (mm), to the nominal spray angle of the nozzle, θ (degrees), and the mass flow rate of the coolant, m (g/s).

Diagrams in column (a) of Figure 5, illustrate the evolution of heat flux, q, with regard to overheating, ΔT_sat. At low degrees of overheat, the progression is effectively linear. At small nozzle–surface distance levels, all mass flow rates produce curves with a nearly constant slope that terminates at progressively higher overheat values as the gap widens (see successive diagrams). Beyond this linear region, the slope between adjacent points falls off sharply. In contrast, the largest nozzle–surface distances yield an extended linear regime—reaching higher overheat values—but with a reduced, flow-rate-dependent slope. These cases also attain substantially greater overheat and heat flux at the curve endpoints. According to the previously presented equations, the slope of the line from the origin to each data point directly quantifies the heat transfer coefficient, h.

This quantitative relationship is shown in column (b) of the Figure 5, where the heat transfer coefficient, h, is plotted versus heat flux, q. Initially, h rises from about 8 to 11 W/cm²·K as q increases. Beyond a certain q threshold, however, h falls off sharply at small nozzle–surface distance levels, while at larger gaps the decline is more gradual or even imperceptible. The peak h is then followed by a steep drop, indicating a severe breakdown of heat transfer performance and likely the onset of surface drying and critical heat flux (CHF). Surface drying refers to the presence of vapor on the heated surface, which leads to severe degradation of the heat transfer coefficient due to significantly lower conductivity of the vapor. Surface drying may occur either due to intense vapor formation at the nucleation sites on the surface, within the liquid film formed over it (film boiling), or due to the disruption of the thin liquid film due to the evaporation at the surface of the liquid and droplet impingent. Consequently, it should be noted that the critical heat flux in spray-cooling, despite the sharp drop in coefficient, h, for small variations of q, does not have the characteristics of an instantaneous burnout point, as in pool boiling, since the constant impact of droplets delays the immediate dry out of the surface with a stable film of vapor. In any case, the occurrence of surface drying at random locations leads to the reduction of the average heat transfer coefficient, abruptly at short distances of the nozzle and more gradually at larger ones. It is interesting that for successive distances of the nozzle, the taxonomy of curves for each flow rate does not show a consistent trend, probably due to the different contribution of the heat transfer mechanisms mentioned in the introduction [3], a feature which will be further discussed later. However, it seems that initially, increasing the distance of the nozzle from the surface results in a shift in the value of the critical heat flux to larger values, with a parallel small increase in the maximum heat transfer coefficient, although at the greatest distance studied, the values of the coefficient show small variations at much smaller values.

At a mass flux of 5.5 g/s and a distance of 30 mm, the heat transfer coefficient peaks at 13.4 W/cm²·K for a heat flux of q = 173.6 W/cm². At 4.5 g/s and 20 mm, h reaches 13.0 W/cm²·K at q = 174.9 W/cm². Increasing the distance to 35 mm (maximum distance level) delays CHF to approximately 240 W/cm² but reduces h to about 8 W/cm²·K.

3.2. Results for 45° Spray Nozzle

Corresponding diagrams for the 45° spray nozzle are presented in Figure 6. Excluding the nozzle spray angle, all other controlled experimental conditions are the same as for the 60° nozzle. On the left column of the figure, the development trends are rather similar to those of the 60° nozzle, although the slope in the first part of the curves is always weaker. Change in the linear trend is, in general, less obvious and is prolonged to higher overheat values. The curves terminate at higher overheat values and lower heat flux levels. Similar trends are also observed in the diagrams of the heat transfer coefficient, in the right column of the figure, although its values are in all cases lower. Differences in the evolution between the three curves for the different mass flow rates are even smaller than for the 60° nozzle in most cases. The heat transfer coefficient initially increases almost imperceptibly as the heat flux is increased. Degradation of the heat transfer mechanisms is reached at lower heat fluxes, although the decline of the heat transfer coefficient afterwards is more gradual. The maximum value of the heat transfer coefficient, h = 8.9 W/cm²K, is obtained for a flow rate of 4.5 g/s, a distance of 20 mm, and a heat flux q = 109.0 W/cm². Again, a delay of the critical heat flux (q = ~220 W/cm²), at the greater distance of 35 mm, seems to be achieved, at the cost of a rather low heat transfer coefficient (h = ~5.5 W/cm² K).

4. Discussion

4.1. Effect of PID Tuning on Heat Flux

A notable decline in heat flux was observed in the final stages of some test cases, particularly when conditions approached or exceeded the degradation of the heat transfer and the critical heat flux (CHF) threshold. This behavior is illustrated in Figure 5a and Figure 6a, where heat flux initially increases with surface overheating but then peaks and subsequently declines. This phenomenon is closely linked to the thermal control strategy utilizing a PID controller, which regulates power input to the heating resistors based on a setpoint temperature, while the saturation temperature in the spray chamber remains constant at a specific pressure of the two-phase refrigerant mixture.

The heat flux transferred from the heated surface to the coolant is governed by the temperature difference between the PID setpoint and saturation temperature, in conjunction with the total thermal resistance of the system, as shown in the simplified network in Figure 7. The latter is the sum of the thermal resistance due to the conduction within the copper block and stem and that due to the convection coefficient formed by the two-phase heat transfer mechanisms between the surface and the coolant, connected in series (Equation (4)).

$$\mathrm{q}={\displaystyle \frac{{\mathrm{T}}_{\mathrm{P}\mathrm{I}\mathrm{D}}-{\mathrm{T}}_{\mathrm{s}\mathrm{a}\mathrm{t}}}{{\mathrm{R}}_{\mathrm{C}\mathrm{o}\mathrm{p}\mathrm{p}\mathrm{e}\mathrm{r}}+{\mathrm{R}}_{\mathrm{c}\mathrm{o}\mathrm{n}\mathrm{v}\mathrm{e}\mathrm{c}\mathrm{t}\mathrm{i}\mathrm{o}\mathrm{n}}}}$$

(4)

As the heat flux approaches CHF, a degradation of the heat transfer mechanism or even localized surface drying starts to develop, creating vapor patches that hinder heat transfer due to the significantly lower thermal conductivity of vapor in comparison with that of the surrounding liquid. This leads to an increase in total thermal resistance, throttling heat transfer.

Due to the low heat transfer coefficient, close to and beyond the critical heat flux conditions, the total thermal resistance increases. Therefore, with the same PID setting and the same saturation temperature, the higher overall heat resistance results in lower heat flux. The total thermal resistance sometimes becomes so large that despite the increase in the control temperature of the PID, the overall heat flux decreases, as observed in the latter portions of the curves in Figure 5a and Figure 6a.

4.2. Effect of Nozzle-to-Surface Distance and Spray Angle

The distance between the spray nozzle and the heated surface (denoted as H) significantly influences both the critical heat flux and the heat transfer coefficient. As shown in Figure 5 and Figure 6, increasing the nozzle–surface distance results in a general shift of the CHF toward higher heat flux values.

For instance, with the larger angle nozzle (Figure 5), increasing H from 15 mm to 35 mm raises the CHF from approximately 175 W/cm² to around 240 W/cm². However, this comes at the cost of a reduced maximum heat transfer coefficient, which drops from about 13.4 W/cm²·K (at H = 30 mm) to ~8 W/cm²·K (at H = 35 mm), as seen in Figure 5b. A similar effect is depicted for the 45° nozzle (Figure 6), although both h and CHF are generally lower across all distances due to the narrower spray cone.

Assuming a uniform spray distribution and using Danfoss SPKT0 solid-cone nozzles—whose nominal spray angles are specified with oil as the working fluid—the 60° nozzle would theoretically cover nearly 100% of the target surface at 15 mm and expand to about 450% of that area at 35 mm. In contrast, the 45° nozzle would cover only 53% at 15 mm and reach about 190% coverage at 35 mm. However, optical observations show that refrigerant spray behaves more like a cylindrical jet (Figure 3), producing a smaller wetted area at each distance. Besides the use of different liquid and inlet conditions, this is due to the evaporation of droplets before impact, especially at the periphery of the spray, narrowing of the jet by the chamber back-pressure [22], and recirculating flows near the surface, which further distort the pattern. Consequently, the actual coverage is consistently lower than the nozzles’ oil-based specifications would imply.

These phenomena profoundly affect the four heat transfer modes described in the introduction. At small nozzle–surface distances, the 45° nozzle’s spray strikes only the central zone—where droplet impact enhances heat transfer—while the peripheral region supports a radially outward liquid film that transfers heat primarily by surface evaporation. This film flow not only yields a lower local heat transfer coefficient but also sweeps away developing nucleation sites. Consequently, the 45° nozzle’s average coefficient is lower than that of the 60° nozzle, whose wider spray covers a greater area despite slower droplets. The peripheral film under the 45° nozzle also dries out more rapidly, precipitating CHF at lower heat fluxes. As the nozzle–surface distance increases, both sprays broaden to envelop the entire surface, suppressing dry-patch formation and delaying CHF, but the reduced droplet velocity limits impact-enhanced heat transfer, so average heat transfer coefficients decline for both nozzle angles.

4.3. Effect of Mass Flow Rate

The mass flow rate of the refrigerant spray (denoted as m) is also critical for spray thermal performance. For the 60° nozzle (Figure 5), the highest heat transfer coefficient of 13.4 W/cm²·K is achieved at m = 5.5 g/s and H = 30 mm. Similarly, at m = 4.5 g/s and H = 20 mm, a nearly equivalent h value of 13.0 W/cm²·K is observed. These peaks coincide with heat fluxes near 175 W/cm². At even higher mass flow (6.0 g/s), however, performance gains are limited or reversed. In contrast, for the 45° nozzle (Figure 6), the influence of mass flow rate is more subdued. The peak h of 8.9 W/cm²·K is observed at m = 4.5 g/s and H = 20 mm, with diminishing improvements at higher flow rates.

Overall, the results suggest that the flow patterns developing over the heated surface, and their effect on the activation of the four heat transfer mechanisms associated with spray cooling, play a decisive role in the achieved heat transfer coefficient and the delay of CHF effects. Three performance factors can be identified in this respect.

1. 

Uniform spray coverage is essential to engage all four heat transfer mechanisms. When droplet impact is confined to the central region, the surrounding liquid film relies solely on surface evaporation and is prone to dry-spot formation—explaining the lower efficiency of the narrower-angle nozzle.

2. 

High-velocity droplet impingement enhances heat transfer by breaching the liquid film, activating all four mechanisms, and suppressing local dry-out.

3. 

A thin liquid film maximizes the efficacy of droplet-induced mechanisms, delivering high heat transfer coefficients up to near-CHF conditions, beyond which random dry spots emerge.

These combined effects account for the peak coefficients observed with the high-angle nozzle at m = 5.5 g/s, H = 30 mm and m = 4.5 g/s, H = 25 mm and their subsequent decline. Conversely, at larger nozzle–surface distances, low-velocity droplets promote a thick film that both reduces h and delays CHF.

In summary, optimal spray cooling demands uniform, high-intensity coverage at tuned mass flows to prevent thick-film formation. Practical implementations must also address spatial variations in heat transfer modes to minimize surface temperature non-uniformities.

4.4. Comparison of the Present Results with Those of Lin et al. [14]

A comparison of results between similar experimental investigations can provide significant insight into the effect of the different experimental realizations. A comparison between the current experimental data and the results reported by Lin et al. [14] was undertaken to evaluate the consistency and performance of the flash spray-cooling system using R410A in a closed-loop configuration. Lin et al. [14] utilized a similar spray-cooling setup, also employing a flat heated surface and the same refrigerant. A full-cone spray nozzle (Spraying Systems Co. TG-XX) with a 60° nominal spray angle for water was installed within a somewhat larger cylindrical chamber of 50 mm diameter. The heated copper surface was located 8 mm over the bottom of the chamber to reduce the influence of reflected droplets and avoid liquid accumulation on the surface. Figure 8 shows the present results for a coolant mass flow rate of m = 5.5 g/s, in comparison with those of Lin et al. [14] for m = 6.7 g/s, for different nozzle-to-surface distances (H = 15, 20, 25, 30, 35 mm). The nominal spray angle for both experiments is 60°.

Figure 8a compares the variation of heat flux, q, as a function of surface overheating, ΔTsat, in the two investigations. Both sets of data exhibit similar trends: an almost initial linear increase in heat flux with overheating, followed by a gradual deviation from linearity as the CHF conditions are approached. The present results show slightly lower maximum heat flux values than those of Lin et al. [14], and the slope of the curves is generally different.

Similarly, Figure 8b illustrates the heat transfer coefficient, h, as a function of heat flux, q, for the same experimental conditions. An almost linear increase is observed for low q values, which after reaching a peak, is followed by a decline, indicating a degradation of heat transfer efficiency.

Comparing Lin et al.’s [14] findings to the present results underscores three key factors—surface coverage, droplet-impingement intensity, and film thickness—and their influence on the four spray-cooling heat transfer mechanisms. Despite sharing a nominal spray angle, Lin et al.’s [14] water nozzle, operating with R410 refrigerant, exhibits a larger effective spray angle (though still under 60°) than the Danfoss oil nozzle used herein. Consequently, it delivers wider and more uniform surface coverage at comparable nozzle–surface distances with higher-velocity droplets.

This combination yields optimal performance at H = 25 mm, where a thin, continuously refreshed liquid film engages all four heat transfer modes. In Figure 8a, the q-versus-ΔT curves become vertical—or even slope negatively—showing that a small overheat drives ever-increasing heat fluxes. These intense conditions translate into the markedly higher heat transfer coefficients seen in Figure 8b, significantly exceeding our own results. As q rises further, however, film thinning and random dry-spot formation trigger CHF onset.

Similar but slightly less optimal behavior occurs at H = 20 mm. At H = 30 mm, reduced droplet count and lower velocities limit mechanism activation, resulting in lower h and earlier CHF at smaller q values. Overall, our trends and magnitudes align closely with Lin et al. [14], validating both methodologies.

This comparison highlights the pivotal roles of spray coverage, impingement intensity, and film thickness in maximizing h and delaying CHF, reinforcing the broader applicability of flash spray cooling for power-electronics thermal management.

5. Concluding Remarks and Future Directions

5.1. Main Contributions

This experimental study investigated the thermal performance of a flash spray-cooling system using R410A refrigerant in a closed-loop configuration, aiming to enhance heat dissipation for power electronics applications. By systematically varying key parameters such as nozzle spray angle, mass flow rate, and nozzle-to-surface distance, we identified the critical interplay between spray dynamics and cooling efficiency. The findings highlighted three dominant factors influencing spray-cooling performance: (a) uniform surface coverage by the spray, (b) intensity of droplet impingement, and (c) thickness and stability of the liquid film over the heated surface. The 60° nozzle configuration consistently outperformed the 45° nozzle, particularly at intermediate mass flow rates (≈5.5 g/s) and nozzle distances between 20 and 30 mm, achieving peak HTC values of 13.4 W/cm²·K and CHF values up to 240 W/cm². Furthermore, the results align well with similar studies (e.g., Lin et al. [14]), validating the experimental methodology and reinforcing the potential of flash spray cooling as a high-performance thermal management solution.

5.2. Limitations of the Present Study

Despite the promising results, the following limitations of the current work should be acknowledged:

(a)

The investigation was limited to smooth, unmodified surfaces; the influence of advanced surface engineering techniques was not explored.

(b)

All experiments were conducted under steady-state thermal loads, which do not fully capture the dynamic conditions faced in real-world electronic applications.

(c)

Optical visualization was limited to qualitative observations without quantitative droplet size or velocity measurements.

(d)

The PID control strategy, while effective, was based on static tuning and may not adequately respond to rapid thermal fluctuations near CHF conditions.

(e)

Environmental considerations, such as refrigerant GWP (Global Warming Potential) impact, were not addressed through alternative refrigerant testing.

5.3. Prioritized Future Research Directions

To address these limitations and further enhance spray-cooling performance, the following future research directions are proposed, prioritized by feasibility and impact:

Surface Enhancement: Preliminary results on microstructured surfaces (e.g., hexagonal pyramids) suggest substantial improvements in both CHF and heat transfer coefficients. Further investigations could explore different surface geometries, wettability modifications, or nano-coatings to optimize droplet dynamics and nucleation site density.

Dynamic Load Conditions: Most current tests were conducted under steady-state conditions. Studying transient thermal loads would offer more realistic insights into spray-cooling behavior in practical electronic devices, particularly those subjected to pulsed or rapidly varying power demands.

Droplet Characterization and Visualization: High-speed imaging and laser diagnostics (e.g., PIV and LIF) could provide deeper insights into the droplet impact dynamics, surface wetting behavior, and film formation, leading to better physical models of spray–surface interactions.

Closed-loop Control Optimization: The PID control strategy could be further tuned or even replaced with model-predictive or adaptive control algorithms to respond more intelligently to the onset of CHF and mitigate surface drying conditions.

Alternative Refrigerants: While R410A performed well in this study, exploration of low-GWP refrigerants or environmentally safer alternatives (e.g., R1234yf, CO₂) would align the system with future regulatory and sustainability targets.

In conclusion, flash spray cooling using refrigerants like R410A demonstrates strong potential as a high-performance thermal management strategy for next-generation power electronics. Optimizing nozzle configuration, refrigerant delivery, and surface design can yield significant gains in cooling efficiency and system reliability, paving the way for compact and robust electronic devices capable of handling higher thermal loads.