The fluid domain medium studied in this paper is an incompressible Newtonian fluid, encompassing the fundamental principles of mass conservation and momentum conservation, which can be expressed as:

Law of conservation of mass: (1) ∂ u ∂ x + ∂ v ∂ y + ∂ w ∂ z = 0 , here, u, v, and w denote the velocity components along the x, y, and z coordinate axes, respectively.

Momentum conservation equations:

The momentum equations along the X, Y, and Z directions are as follows:

X-axis:

Y-axis:

Z-axis: (4) ρ ∂ w ∂ t + u ∂ w ∂ x + v ∂ w ∂ y + w ∂ w ∂ z = − ∂ p ∂ z + μ ∂ 2 w ∂ x 2 + ∂ 2 w ∂ y 2 + ∂ 2 w ∂ z 2 where ρ, μ, and p correspond to the working fluid density, dynamic viscosity, and pressure, respectively, and t denotes time.

Energy conservation equation:

For the steady-state, incompressible flow with negligible viscous dissipation and internal heat sources, the energy conservation equation is expressed as: (5) ρ c p u ∂ T ∂ x + v ∂ T ∂ y + w ∂ T ∂ z = λ ∂ 2 T ∂ x 2 + ∂ 2 T ∂ y 2 + ∂ 2 T ∂ z 2 in this equation, c_p stands for the specific heat at constant pressure, λ indicates the fluid thermal conductivity, and T corresponds to the temperature of the fluid. The left-hand side of Eq. (5) depicts convective thermal transport induced by fluid motion, whereas the right-hand side stands for conductive heat transfer caused by temperature gradients.

Fig. 1 illustrates the three-dimensional structural schematic of the microchannel heat sink, primarily comprising a heat source and microchannels. The microchannel section primarily consists of an inlet buffer zone, an outlet buffer zone, and ten microchannels in between. The fluid medium enters the buffer zone through the inlet, undergoes buffering, and then flows uniformly into each microchannel for heat exchange. Finally, it converges in the outlet buffer zone before exiting through the outlet.

The specific parameters of the geometric model are shown in Table 1.

Generate a surface mesh using Fluent with the meshing component, then create the fluid domain. Subsequently, a boundary layer is arranged at the junction of fluid and solid computational domains. The simulation model employs a polyhedral mesh, generating 224,863 solid-domain and 199,486 fluid-domain meshes, totaling 424,349. The mesh orthogonality quality factor is 0.2, exceeding the 0.15 threshold. That indicates stable mesh quality, ensuring the numerical simulation remains stable. The overall and local mesh configurations of the simulation model after mesh partitioning are shown in Fig. 2, ((a) Local import grid partitioning of the microchannel region; (b) Overall mesh partitioning of the entire radiator; (c) Detailed mesh generation within the microchannels).

The computational model was imported into the ANSYS Meshing platform for grid generation. To minimize numerical discrepancies and enhance computational precision, local mesh refinement within the fluid domain was implemented via the “Refine” function. Nevertheless, an elevated refinement level leads to a substantial increase in the number of mesh elements, thereby imposing more rigorous computational requirements. To verify the mesh independence of the solution, a specific case characterized by 5 grooves, a spacing of 2.4 mm, and an inlet velocity of 0.6 m/s was selected as the reference scenario; the grid convergence was systematically examined by adjusting the refinement levels ranging from 0 to 1. The test results are shown in Table 2.

As shown in Table 2, when the refinement level is set to 2, the relative error is 2.66% and 1.57%, respectively. Therefore, in this study, refinement level 2 is sufficient to meet the accuracy requirements.

For the numerical simulation of groove-type microchannel heat sinks, the following assumptions are made: (1)

Assume single-phase flow heat transfer;

The finite volume method is adopted to simulate microchannel convective heat transfer. The Coupled implicit pressure-velocity coupling scheme is applied to solve continuity and momentum equations simultaneously, matching the pressure drop and flow-thermal coupling characteristics. For spatial discretization, cell-based least squares, second-order scheme and second-order upwind scheme are applied for gradient, pressure, momentum and energy terms respectively.

The cooling medium is set as liquid water with an inlet temperature of 283 K. Several distinct flow velocity values were adopted in the simulation, namely 0.6 m/s, 0.8 m/s, 1.0 m/s, 1.2 m/s, 1.4 m/s and 1.6 m/s. The outlet gauge pressure is 0 kPa, and backflow suppression is selected. Both the heat source and heat sink are made of copper. The heat flux at the heat source wall is 130 W/cm². Coupled boundary conditions are applied between internal flow channels, with a convective heat transfer coefficient set to 1800 W/(m²·K).

The specific parameters for the boundary conditions and physical quantities mentioned above are listed in Table 3.

To assess the fluid flow behaviors and heat transfer properties of microchannels, a number of dimensionless parameters have been adopted. The formula for the Reynolds number (Re) is presented as follows: (6) Re = ρ f u m D h μ f (7) D h = 2 W H ( W + H ) (8) u m = v W 1 H 1 n W H where ρ_f denotes the density of the working fluid; u_m is the average inlet velocity for each individual microchannel; D_h represents the hydraulic diameter of the microchannel; μ_f stands for the dynamic viscosity of the working fluid.

The thermophysical properties of both water and copper are presented in Table 4. The calculated value of the Reynolds number are presented in Table 5.

To comprehensively assess the flow and heat transfer performance of microchannel structures, the performance evaluation criteria (PEC) are introduced. This criterion quantifies the synergistic relationship between heat transfer enhancement (characterized by Nu) and hydraulic penalty (represented by pressure drop), and is expressed as: (9) P E C = N u N u 0 / f f 0 1 / 3 where Nu represents the Nusselt number of the target case, Nu₀ is the Nusselt number of the baseline reference case, f defines the friction factor of the target microchannel, and f₀ is the friction factor associated with the reference configuration.

The equation for the mean Nusselt number is presented below: (10) N u = h D h λ f (11) h = q T w , a v g − T f , a v g where h denotes the mean convective heat transfer coefficient, q stands for the average heat flux across the fluid-solid interface, while T_w,avg and T_f,avg refer to the bulk-averaged wall temperature and fluid temperature, respectively. All the parameters discussed above all pertain to conditions within the microchannel region.

The simulated value of the friction coefficient (f) can be calculated using the following formula, which is derived from the Darcy–Weisbach formula: (12) f = 2 Δ p D h ρ f L u m 2 where, ΔP indicates the pressure drop, while L is the overall length of the microchannel structure.

Regarding the laminar flow in a rectangular cross-section microchannel, Shah proposed an expression for the average friction coefficient in the undeveloped flow regime: (13) f R e = 96 ( 1 − 1.3553 α c + 1.9467 α c 2 − 1.7012 α c 3 + 0.9563 α c 4 − 0.2537 α c 5 ) where, α c represents the aspect ratio of the channel, which is set to 1.33 in this paper; f  represents the friction coefficient.

The temperature non-uniformity over the heating surface domain is quantitatively evaluated using the following metric: (14) Δ T non-uni = 1 2 T max − T avg + T min − T avg where, T max refers to the maximum temperature of the heating surface, T min represents the minimum temperature, and T avg stands for the average temperature of the heat source. A lower value of Δ T non-uni indicates a more uniform thermal profile within the heat source region.

A grooved microchannel with five grooves and a 2.4 mm interval was adopted. Six inlet velocities from 0.6 m/s to 1.6 m/s were set for comparison. The temperature variation patterns of the heat source, microchannel heat sink, and cooling working fluid were analyzed. The results are shown in Fig. 3 and Fig. 4 respectively.

Fig. 3 depicts the thermal distribution characteristics of the heat source and microchannel heat exchanger across different flow conditions. As the working fluid velocity increases, the maximum temperature of the heat source decreases significantly, and the high-temperature region of the microchannel heat exchanger shrinks.

Fig. 4 illustrates the temperature field characteristics of the working fluid across different flow conditions. As the working fluid flow rate increases, the outlet temperature decreases significantly.

As shown in Fig. 5, both the heat source temperature and the working fluid outlet temperature decrease with increasing flow velocity. Analysis reveals the following mechanism: the high-temperature heat source transfers heat to the microchannel wall by conduction, and the cooling working fluid subsequently removes heat from the wall by convective heat transfer. Increased fluid velocity within the microchannel elevates the flow Reynolds number. Increased turbulence intensifies the disruption of the thermal boundary layer, which further reduces thermal resistance and thereby enhances the intensity of convective heat transfer. Consequently, heat generated by the heat source is rapidly transferred through the microchannel wall into the working fluid, preventing heat accumulation at the source and thereby lowering its temperature.

Meanwhile, under conditions where the structural dimensions of the microchannel remain unchanged, higher inlet velocity directly decreases the residence time of the working fluid in the microchannel, reducing heat accumulation. This prevents the working fluid from being sufficiently heated within the channel, consequently lowering the outlet temperature. This effect helps maintain a larger temperature difference across the heat transfer surface, thereby enhancing the heat transfer rate. Furthermore, an increase in flow velocity directly results in a higher mass flow rate of the working fluid through the microchannel, thereby increasing the total heat absorption. Research indicates that increased flow velocity enhances convective heat transfer within microchannels while reducing thermal resistance. Concurrently, the increased working fluid flow rate and reduced residence time significantly increase the heat extraction rate from the heat source, mitigating the temperature rise in the working fluid. That ultimately achieves simultaneous reductions in both the heat source temperature and the working fluid outlet temperature.

To more accurately investigate the effects of flow velocity on flow and heat transfer in grooved microchannels, the following analysis will focus solely on how the pressure drop, friction factor, temperature non-uniformity, Nusselt number, and PEC within the microchannel region vary with changes in the Reynolds number.

Fig. 6 illustrates the variation of pressure drop (Δp) within the microchannel region as a function of the Reynolds number (Re). As observed from the plot, the pressure difference between the microchannel inlets and outlets increases approximately linearly with the rise of the Reynolds number. The Reynolds number is directly proportional to the inlet flow velocity; higher velocity of the working medium significantly increases the inertial forces and wall shear stresses acting on the fluid within the microchannel. At the same time, the grooved structure exacerbates secondary flow and boundary layer separation, further increasing flow resistance, which ultimately causes the pressure drop to increase in tandem with the Reynolds number.

Fig. 6 presents a comparison between the theoretical laminar flow results for rectangular microchannels (calculated via Shah’s equation) and the measured friction coefficient (f) of the groove-type microchannels. With an increase in the Reynolds number, the friction factor exhibits an initial sharp decline followed by a gradual decrease. Specifically, as the Reynolds number rises from 547 to 1459, the friction coefficient of the grooved microchannel decreased from 0.121 to 0.065, a reduction of 46.2%. In terms of the patterns of change, the overall trend of the simulated values is in complete agreement with the theoretical results for laminar flow in rectangular microchannels: within the laminar range, the friction coefficient decreases as the Reynolds number increases, and the rate of decrease gradually slows down at high Reynolds numbers.

It is worth noting that the friction coefficients of grooved microchannels are all higher than the theoretical values for smooth rectangular microchannels. This is because the grooved structure disrupts the stability of the boundary layer on the channel walls, inducing vortical disturbances and secondary flow within the fluid, which significantly increases local flow resistance and energy loss. Although the rate of increase in pressure drop slows as the Reynolds number rises, the additional resistance caused by the grooves persists, resulting in a friction coefficient that remains higher than that of smooth rectangular microchannels. At the same time, the discrepancy between the simulation results and Shah’s equation confirms the significant influence of the grooved structure on flow characteristics, providing data support for subsequent structural optimization.

Fig. 7a shows how the average Nu changes with Re for the grooved microchannel. As observed in the figure, the average Nu increases with the Re throughout the investigated range of Re = 547 to 1459. This trend is consistent with the results reported in Reference [23], confirming the reliability of the numerical simulation. The Reynolds number is directly proportional to the inlet flow velocity; an increase in flow velocity significantly enhances the intensity of fluid turbulence within the microchannel, effectively disrupting the thermal boundary layer attached to the wall and reducing its thickness. This substantially lowers the convective heat transfer resistance and improves heat transfer performance, ultimately resulting in a continuous increase in the average Nusselt number.

The variation of heat source temperature non-uniformity ( Δ T non-uni ) with Re is plotted in Fig. 7b. The results indicate that heat source temperature non-uniformity decreases as the Reynolds number increases. Within the investigated range of Reynolds numbers (547–1459), Δ T non-uni decreased from approximately 26.2°C to approximately 19.3°C, a reduction of 26.3%. Based on the trend shown in Fig. 7a, it can be concluded that increasing the Reynolds number offers a twofold benefit: the average Nusselt number increases significantly, thereby improving overall heat transfer performance, while at the same time the temperature uniformity of the heat source is significantly enhanced. This is because the convective heat transfer capacity of the fluid increases at high flow rates, allowing heat to be rapidly removed from all areas of the heat source. This prevents localized heat buildup, thereby reducing the difference between the maximum and average temperatures of the heat source and resulting in a more uniform temperature distribution.

In summary, increasing the Reynolds number (i.e., increasing the inlet flow velocity) can simultaneously achieve the dual effects of enhancing heat transfer performance and optimizing temperature uniformity, thereby providing a clear theoretical basis for optimizing the operating conditions of grooved microchannel heat sinks.

To quantitatively evaluate the overall heat dissipation performance of grooved microchannels under different inlet flow velocities, this paper employs the Performance Evaluation Criterion (PEC) to perform a trade-off analysis between heat transfer capacity and flow resistance. The calculation results as a function of the Reynolds number (Re) are shown in Fig. 8.

As can be seen from the Fig. 8, within the Reynolds number range studied (Re = 547~1459), the PEC value exhibits an increasing trend as the Reynolds number rises. As the Reynolds number increased from 547 to 1459, the PEC value rose from 1.22 to 2.05, representing an increase of 68.0%. This principle indicates that, although increasing the flow velocity leads to a significant increase in pressure drop and friction coefficient within the microchannel (resulting in greater flow resistance), the improvement in heat transfer performance far outweighs the increase in resistance.

Based on the previous analysis of the Nusselt number and pressure drop, it can be seen that the enhanced convective heat transfer resulting from increased flow velocity (a significant increase in the Nusselt number) effectively offsets the energy loss caused by the rise in pressure drop, ultimately leading to continuous optimization of the overall heat dissipation performance of the grooved microchannels. These results further confirm that, within the operating conditions considered in this study, increasing the inlet flow velocity is an effective strategy for enhancing the overall performance of grooved microchannel heat sinks, providing important guidance for selecting operating conditions in engineering applications involving high heat flux scenarios.

Set groove count N = 0/1/3/5/7/9/11. Under conditions of working fluid inlet velocity v = 0.8 m/s and groove spacing S = 2.4 mm, simulations were conducted to investigate the effect of disturbance unit count on microchannel heat dissipation performance. Temperature variation patterns for the heat source, the microchannel heat sink, and the cooling working fluid are shown in Fig. 9, Fig. 10 and Fig. 11 respectively.

Fig. 9 displays the temperature field characteristics of the heat source and heat sink across different groove configurations. The results show that a gradual increase in the number of grooves leads to both a contraction of the heat sink’s high-temperature region and a notable drop in the heat source’s peak temperature. When N = 0 (smooth channel), the high-temperature zone is the largest, and the center temperature is the highest; When N = 1, the high-temperature zone shrinks significantly, and the temperature decreases. As N increases from 3 to 11, the high-temperature zone gradually shrinks.

As presented in Fig. 10, the thermal distribution of the working fluid is shown for different groove numbers. Even with a gradual increase in the number of grooves, the working fluid outlet temperature remains nearly constant.

Fig. 11 shows how the heat source and fluid temperatures vary with the number of grooves. Grooves, as typical passive heat transfer enhancement structures within microchannels, disrupt the working fluid’s laminar boundary layer (see Fig. 12 ((a) Overall flow distribution across the full radiator; (b) Partial flow distribution in the microchannel array; (c) Detailed flow distribution within a single microchannel); Fig. 13 ((a) Overall flow distribution across the full radiator; (b) Partial flow distribution in the microchannel array; (c) Detailed flow distribution within a single grooved microchannel), inducing recirculation and vortex motion in the groove region. That significantly enhances mixing between the near-wall fluid and the main flow zone. Moreover, a higher number of grooves increases the density of disturbance points within the channel, leading to a marked improvement in the local convective heat transfer coefficient. On the other hand, the groove structure increases the actual effective heat exchange area of the microchannel, with the heat exchange area expanding in tandem as the number of grooves increases. Research indicates that increasing the number of grooves within microchannels can effectively enhance convective heat transfer rates between heat sources and working fluids, thereby reducing heat source temperatures and improving thermal reliability. However, as the number of grooves increases, the marginal effect diminishes. The marginal effect is greatest when increasing from N = 0 to N = 1, significantly declines when increasing from N = 1 to N = 5, and approaches zero when increasing from N = 7 to N = 11.

Fig. 14 presents the velocity vector plots for different groove numbers, which visualize the flow structures and vortex formation within the microchannel. As the groove number increases, recirculation vortices appear within the grooves, as indicated by the local eddy patterns in the vector field. These vortices disrupt the laminar flow and enhance transverse mixing, which breaks the thermal boundary layer and promotes convective heat transfer.

The fluid outlet temperature remains nearly constant with increasing groove number, which can be rigorously explained from the first law of thermodynamics (global energy conservation). For the present system, the total heat input rate q w is fixed, and the working fluid mass flow rate m ˙ , specific heat capacity c p , and inlet temperature T in are all prescribed as constant boundary conditions. The steady-flow energy balance yields:

Since q w , m ˙ , c p , and T in are independent of the groove configuration, the outlet temperature T out is uniquely determined and must remain constant regardless of the number of grooves. This conclusion follows directly from the first law and holds irrespective of internal flow details, local thermal non-uniformities, or the degree of convective enhancement.

While the outlet temperature is invariant, increasing the groove number significantly alters the local temperature field within the solid and fluid domains. The additional grooves disrupt the thermal boundary layer, promote flow mixing, and enhance local convective heat transfer coefficients. These effects do not change the total energy absorbed by the fluid but redistribute the thermal resistance, resulting in a significantly more homogeneous temperature field on the heat source and a reduction in its maximum temperature. From a thermodynamic perspective, the increased groove number increases internal irreversibilities (entropy generation) due to enhanced mixing and flow disturbance, yet the global energy balance remains unchanged. Thus, the observed invariance of outlet temperature is a direct consequence of global energy conservation, while the improved thermal performance manifests in local temperature uniformity and reduced peak temperatures.

To more accurately elucidate the influence of the number of grooves on the flow and heat transfer characteristics of microchannels, this section systematically analyzes the variation of pressure drop (Δp), friction factor (f), temperature non-uniformity, Nusselt number (Nu), and the comprehensive performance evaluation factor (PEC) with respect to the number of grooves under a fixed Reynolds number (Re = 730).

The variation in the pressure drop between the microchannel inlet and outlet with different groove configurations is plotted in Fig. 15a. As shown in the figure, within the range of groove numbers studied (N = 0~11), the pressure drop in the microchannel exhibits a monotonically increasing trend as the number of grooves increases, and the rate of increase gradually slows down as the number of grooves increases. Under steady-state conditions with Re = 730, as the number of grooves increased from 0 (smooth rectangular microchannel) to 11, the pressure drop rose from 416.16 Pa to 483.00 Pa, representing an increase of 16.05%. This is because the grooved structure disrupts the stability of the boundary layer on the channel walls, inducing vortical disturbances and secondary flow within the fluid, thereby increasing local flow resistance and energy loss; the greater the number of grooves, the larger the scope and intensity of the fluid disturbances, and the more pronounced the cumulative effect on flow resistance, ultimately leading to a continuous increase in pressure drop.

Fig. 15b shows the variation of the microchannel friction coefficient as a function of the number of grooves. The results indicate that the trend in the friction coefficient aligns with that of the pressure drop: it increases as the number of grooves increases, and the rate of increase gradually slows down at higher groove counts. As the number of grooves increased from 0 to 11, the coefficient of friction rose from 0.0911 to 0.1056, representing an increase of 15.92%. This finding further confirms the role of grooved structures in enhancing flow resistance.

To further clarify the effect of the number of grooves on the heat transfer performance and temperature uniformity of microchannels, Fig. 15a,b shows the variation of the average Nusselt number and heat source temperature non-uniformity with respect to the number of grooves, respectively.

As shown in Fig. 16a, the average Nusselt number increases as the number of grooves increases, and the rate of increase gradually slows down as the number of grooves increases. Smooth microchannels without grooves have the lowest average Nusselt number, at just 31.8; when the number of grooves is increased to 11, the Nusselt number rises to 46.9, representing a 47.5% increase compared to smooth channels. This indicates that the grooved structure enhances heat transfer in microchannels; however, once the number of grooves exceeds a certain threshold, its effect on disrupting the thermal boundary layer tends to plateau, and the marginal benefit of heat transfer enhancement gradually decreases. Fig. 16b shows the curve of the temperature non-uniformity of the heat source as a function of the number of grooves. It can be concluded that increasing the number of grooves contributes to a rapid reduction in the heat source’s temperature unevenness in the initial stage, followed by a subsequent upward trend.

Fig. 17 shows the variation of the comprehensive performance evaluation factor PEC with the number of grooves; all calculations are nondimensionalized using a smooth rectangular microchannel as a reference. As shown in the figure, PEC exhibits an upward trend as the number of grooves increases, but the rate of increase gradually slows down as the number of grooves rises. For a smooth channel without grooves, the PEC is 1.0; when the number of grooves is increased to 8, the PEC rises to 1.39, a 39% increase over the smooth channel; when the number of grooves is further increased to 11, the PEC rises only slightly to 1.405, an increase of less than 1.1%. These results indicate that grooved structures can effectively enhance the overall flow and heat transfer performance of microchannels; however, as the number of grooves increases, the marginal benefit of this performance improvement gradually decreases. Combining the preceding analysis of the Nusselt number and pressure drop reveals that when the number of grooves is small, the heat transfer enhancement resulting from adding grooves (a significant increase in the Nusselt number) far outweighs the increase in flow resistance, causing the PEC to rise rapidly. However, when the number of grooves exceeds 8, the marginal benefit of heat transfer enhancement diminishes, while flow resistance continues to accumulate, leading to a significant slowdown in the rate of increase of the PEC. This finding also supports the conclusion reached earlier: under operating conditions of Re = 730, setting the number of grooves to 8 achieves an optimal balance between heat transfer enhancement, temperature uniformity, and flow resistance. Further increasing the number of grooves has only a limited effect on improving overall performance, while actually increasing manufacturing costs and flow energy consumption.

As discussed earlier, incorporating grooves within the microchannel can enhance the heat dissipation performance of microchannel heat sinks. Under fixed boundary conditions (groove radius R = 0.8 mm, groove height H = 1.5 mm, heat flux density q = 130 W/cm², working fluid inlet temperature T = 10°C, flow velocity v = 0.8 m/s), microchannel heat dissipation performance simulations were conducted by varying only the groove spacing S across six conditions: 1 mm, 2 mm, 3 mm, 4 mm, 5 mm, and 6 mm.

Fig. 18 shows the temperature distribution of the heat source and heat sink at different groove spacings. When the groove spacing is 1 mm, 2 mm, 3 mm, 4 mm, 5 mm, and 6 mm, respectively, the high-temperature zones are extremely concentrated, slightly dispersed, noticeably dispersed, further dispersed, and most extensive.

Fig. 19 shows the fluid temperature distribution at different groove spacings. When the groove spacing is small (S = 1 mm), temperature stratification occurs, with the high-temperature zone concentrated near the wall surface. At a moderate groove spacing (S = 3 mm), the temperature distribution becomes uniform, and the high-temperature zone extends toward the center of the flow channel. When the groove spacing is large (S = 6 mm), temperature stratification reappears.

As illustrated in Fig. 20, with the gradual enlargement of groove spacing, the highest temperature of the heat source presents a trend of initial reduction followed by elevation. The optimal condition with the lowest central temperature appears at S = 3 mm. When groove spacing is small (1–2 mm), flow obstruction reduces the heat transfer coefficient; simultaneously, vortex interference creates localized hot spots. Among the groove spacings considered in this study, the configuration with 3 mm spacing yields the lowest heat source temperature. At this spacing, eddy currents develop fully without excessive interference, resulting in minimal thermal resistance and maximum heat transfer coefficient under the present geometric and operating conditions. When the groove spacing is large (5–6 mm), the boundary layer fully develops, leading to a decrease in the heat transfer coefficient.

To further clarify the influence of groove spacing on the flow and heat transfer characteristics of microchannels, we will systematically analyze how the pressure drop, friction factor, temperature non-uniformity, Nusselt number, and the comprehensive performance evaluation factor PEC vary with changes in groove spacing within the microchannel region.

Fig. 21a shows the variation of the pressure drop at the microchannel inlet and outlet as a function of groove spacing. As shown in the figure, the pressure drop first increases and then decreases as the groove spacing increases. Fig. 21b shows the variation of the microchannel friction coefficient as a function of groove spacing. As can be seen, the trend in the friction coefficient is consistent with that of the pressure drop.

The influence of the groove spacing on the heat transfer performance of the microchannel and the temperature uniformity of the heat source is presented in Fig. 22a,b. As can be seen from Fig. 22a, as the groove spacing increases, the average Nusselt number presents an initial upward trend followed by a gradual reduction. Fig. 22b shows the variation curve of the heat source temperature non-uniformity with the groove spacing. Overall, the fluctuation range of is relatively small, varying only within the range of 28.0–29.2°C, indicating that the influence of the groove spacing on the temperature uniformity of the heat source is relatively limited.

To quantitatively evaluate the comprehensive flow and heat transfer performance of microchannels with different groove spacings, this paper conducts non-dimensionalization based on a smooth rectangular microchannel and calculates the comprehensive performance evaluation factor PEC. As shown in Fig. 23, the PEC value first increases, then decreases, and then slightly rebounds as the groove spacing increases. In summary, within the parameter range studied in this paper, a groove spacing of 3 mm can achieve an optimal balance between heat transfer enhancement and flow resistance, and it is the optimal structural parameter for the grooved microchannel under the current working conditions.