Design and Numerical Evaluation of a Forced-Air Duct Cooling System for the Thermal Management of Brake Discs in a Heavy-Duty Van

Efficient thermal management of braking systems remains a critical challenge in heavy-duty commercial vans operating under sustained load conditions [1]. During braking, a substantial portion of the vehicle’s kinetic energy is converted into thermal energy at the pad-disc interface [2]. During repeated or prolonged braking, such as extended driving on downhill slopes, high-frequency urban traffic, or fully loaded transport, this heat accumulates in the brake disc, generating high surface temperatures and significant thermal gradients [3, 4]. If heat dissipation is insufficient, the resulting thermal stress can lead to brake fade, reduced coefficient of friction, accelerated wear, surface cracking, and permanent deformation of the disc, ultimately compromising braking reliability and vehicle safety [5].

Commercial vans have a different operating profile than high-performance or racing vehicles. Instead of short-duration extreme braking, these platforms are subjected to sustained thermal loading under high gross weights and variable maneuvering conditions [6]. While ventilated discs promote internal airflow and improve heat transfer by convection, their effectiveness depends largely on the availability and directionality of external airflow within the wheel well. In practical driving situations, airflow distribution around the brake assembly is often irregular and can be significantly affected by vehicle speed, underbody flow structures, and steering-induced geometric variations [7, 8].

Traditional strategies to mitigate brake overheating include increasing the disc mass, modifying the internal ventilation geometry, or using high-temperature-resistant materials [9, 10]. While these approaches improve thermal capacity, they do not directly address the external mechanism of convective heat transfer that regulates heat dissipation during vehicle motion. A more direct strategy involves channeling high-pressure frontal airflow toward the brake assembly through dedicated ducting systems, thereby increasing local air velocity and convective heat dissipation at the disc surface [11].

Air duct cooling systems are commonly implemented in motorsports applications, where aerodynamic optimization and design flexibility facilitate their integration [12]. However, heavy-duty commercial vans operate under very different aerodynamic constraints and space limitations. The integration of a forced-air duct into these platforms must account for underbody flow behavior, wheel well confinement, steering angle variability, suspension movement, and pressure losses along the duct path [13]. Despite the practical relevance of these vehicles in logistics, emergency services, and passenger transport, research on the integrated aero-thermal performance of brake cooling duct systems specifically designed for commercial van geometries has been limited. In particular, the combined influence of variations in inlet velocity and steering angle on airflow distribution and cooling efficiency within the wheel assembly remains insufficiently explored [1, 14].

Recent studies have explored various strategies for improving the thermal performance of braking systems using numerical and experimental approaches. CFD-based research has focused primarily on optimizing the geometries of ventilated discs and their influence on internal airflow structures and heat dissipation, demonstrating that geometric modifications can significantly improve convective cooling performance [9, 15]. Complementarily, coupled aerothermal simulations have been used to evaluate the interaction between flow fields and temperature distribution under different operating conditions, highlighting the importance of accurately capturing airflow behavior to predict thermal response [10, 16, 17]. In parallel, experimental studies have examined the relationship between mass flow rate through ventilated discs and cooling efficiency, confirming that the intensity and distribution of the airflow are critical parameters governing heat removal [18]. Despite these advances, existing research has focused primarily on analysis at the rotor level or on isolated configurations, paying little attention to the complete path of the airflow from the external inlet zones to the brake assembly. In particular, the combined effects of varying inlet velocity and direction-induced geometric changes in airflow distribution within the wheel housing remain insufficiently addressed, especially in the context of heavy-duty commercial vehicle platforms.

In this context, the present study proposes the design and numerical evaluation of a forced-air duct cooling system dedicated to the thermal management of brake discs in a heavy-duty van. The system captures airflow from the lower front area of the vehicle, independent of the radiator core, and channels it through a dedicated duct toward the brake assembly. The numerical analysis evaluates the aerothermal performance of the configuration under different inlet velocities representative of realistic driving scenarios, also considering variations in the steering angle to assess the geometric effects on the wheel well. This study is limited to evaluating the performance of a forced-air duct cooling system for brake discs under different inlet velocities and steering conditions representative of real-world vehicle scenarios.

This section presents the numerical results obtained from CFD simulations of the brake cooling duct and the associated aerothermal behavior of the brake disc. The analysis focuses on three main aspects of cooling performance. First, the influence of the inlet air velocity on the internal flow structure and the thermal behavior of the duct is examined. Second, the resulting pressure losses and aerodynamic characteristics are evaluated to determine the impact of airflow acceleration within the duct geometry. Finally, the convective heat transfer performance at the brake disc surface is analyzed using the spatial distributions of the wall heat transfer coefficient and wall heat flux under different imposed disc temperatures. Taken together, these results provide a comprehensive understanding of the interaction between airflow dynamics and the heat transfer mechanisms that govern the cooling performance of the proposed duct configuration.

3.1 Effect of the inlet velocity into the duct

The inlet air velocity is one of the key parameters governing the aerodynamic and thermal performance of the proposed brake cooling duct. Figures 2 and 3 show the temperature and absolute pressure distributions inside the duct for three representative vehicle speeds: 30 km/h, 60 km/h, and 90 km/h. These results allow us to evaluate how variations in the inlet airflow affect both convective heat transfer within the duct and the associated aerodynamic losses.

Figure 2. Temperature distribution in the duct for three air inlet speeds: 30, 60, and 90 km/h

Figure 3. Absolute pressure distribution in the duct for three air inlet velocities: 30, 60, and 90 km/h

The temperature distribution within the duct shows a strong dependence on the inlet velocity. At the lowest speed (30 km/h), the airflow undergoes significant heating as it moves through the duct due to the heat flux imposed on the duct walls. Under these conditions, air temperatures exceed 430 K near the inlet and remain above approximately 380 K over much of the duct’s length. This behavior indicates a limited ability of the flow to remove heat from the surrounding surfaces, primarily due to the relatively low kinetic energy of the incoming air. The reduced velocity causes a thicker thermal boundary layer to form along the duct walls, which decreases the local temperature gradient at the wall and, consequently, reduces the convective heat transfer coefficient. In addition, the curved sections of the duct promote flow separation and localized recirculation, which increases the residence time of the hot air and further contributes to the high temperatures observed inside the duct.

A significant improvement in thermal performance is observed when the inlet velocity is increased to 60 km/h. In this configuration, the air temperature inside the duct decreases considerably, reaching temperatures between 330 and 370 K over most of the flow domain. Compared to the 30 km/h case, this represents a temperature reduction of between 50 and 70 K in large areas of the duct. The improvement is primarily associated with convective transport intensified by the higher air flow velocity. As the Reynolds number increases, the boundary layer thickness decreases, and the intensity of turbulence near the walls increases, which raises the local heat transfer coefficient and facilitates more efficient heat dissipation from the duct surfaces. Furthermore, the increased flow momentum reduces the extent of recirculation zones along the curved sections of the duct, promoting a more attached flow and a more uniform thermal field throughout the duct.

The highest inlet velocity considered in this study (90 km/h) produces the most favorable thermal behavior. Under these conditions, the air temperature remains near 300–320 K throughout most of the duct, indicating that the airflow is capable of effectively removing the heat load imposed on the duct walls. Compared to the 30 km/h case, the average air temperature inside the duct decreases by more than 100 K, demonstrating the strong influence of the inlet velocity on the system’s convective cooling capacity. It should be noted that this temperature reduction corresponds to the airflow temperature within the duct and does not represent a direct reduction in the brake disc surface temperature. In the present study, the disc temperature was prescribed as a boundary condition to evaluate the convective heat transfer behavior under different airflow conditions. The increase in flow velocity intensifies the convective heat transfer mechanism by reducing the thickness of the thermal boundary layer and increasing turbulent mixing within the flow. Furthermore, the increased momentum of the airflow shortens the residence time of the air within the duct, limiting the amount of heat absorbed by the fluid as it moves toward the outlet.

While higher inlet velocities significantly improve the duct’s thermal performance, they also introduce greater aerodynamic losses. The pressure distributions shown in Figure 3 indicate that the pressure drop along the duct increases progressively with inlet velocity. For the 30 km/h condition, the pressure gradient along the duct is relatively gentle, indicating relatively small aerodynamic losses associated with wall friction and geometric changes in the duct cross-section. However, as the inlet velocity increases to 60 km/h, the pressure drop increases substantially, reaching values close to 4.5–4.6 kPa between the inlet and outlet sections.

At maximum speed (90 km/h), pressure losses increase even further due to the quadratic relationship between dynamic pressure and flow velocity. The greater inertial forces associated with higher speeds amplify both the friction losses at the duct walls and the local losses generated by the curvature and cross-sectional transitions. The pressure field also reveals localized regions of high pressure near the inlet, where the airflow initially decelerates upon entering the duct, followed by a gradual reduction in pressure in the flow direction as the air accelerates through the curved channel.

These results highlight a clear trade-off between thermal performance and aerodynamic efficiency. Increasing the inlet velocity significantly improves the duct’s convective heat dissipation capacity, reducing the air temperature within the system and enhancing the cooling potential of the airflow supplied to the brake assembly. However, this improvement leads to higher pressure losses, which can affect the vehicle’s overall aerodynamic efficiency and the airflow available for cooling. Therefore, the duct’s optimal operating condition must balance these opposing effects to maximize cooling efficiency and minimize the associated aerodynamic penalty.

It should be noted that the pressure losses reported in the present study correspond primarily to internal aerodynamic losses within the brake cooling duct itself. The proposed duct configuration is positioned below the main radiator inlet and was designed to operate independently from the primary engine cooling airflow. Therefore, the predicted pressure drop is not expected to significantly affect the intake airflow supplied to other thermal management components such as the radiator, intercooler, or oil cooler. Nevertheless, the increase in pressure losses at higher vehicle speeds may contribute to a localized aerodynamic penalty and slightly influence the overall underbody airflow behavior of the vehicle. From an engineering perspective, this highlights the importance of optimizing the duct geometry in order to balance enhanced brake cooling performance with acceptable aerodynamic efficiency. Minimizing unnecessary flow resistance and reducing localized losses within the duct may improve cooling effectiveness while limiting potential impacts on overall vehicle aerodynamic performance.

3.2 Effect of the brake disc position

Figure 4 presents the airflow streamlines colored by velocity magnitude around the brake cooling duct and brake disc for different steering configurations. The flow field reveals the formation of accelerated jet-like regions at the duct outlet, together with recirculation structures and localized vortical motion around the brake assembly. These flow features strongly influence the spatial distribution of the convective heat transfer coefficient observed on the disc surface.

Figure 4. Airflow streamlines colored by velocity magnitude around the brake cooling duct and brake disc for different steering positions

Regions directly exposed to the accelerated airflow exhibit higher local velocities and enhanced turbulent mixing, which promote thinner thermal boundary layers and consequently higher convective heat transfer coefficients. In contrast, areas affected by flow separation and recirculation exhibit lower local momentum and reduced cooling effectiveness. The streamline distributions also show that changes in the relative alignment between the duct outlet and the brake disc modify the trajectory and impingement location of the incoming airflow, thereby affecting the local heat transfer behavior across the disc surface.

The relative orientation between the brake disc and the outlet of the forced-air duct significantly affects the aerodynamic interaction between the cooling airflow and the rotor surface. Figures 5 and 6 show the distributions of the convective heat transfer coefficient at the wall and the corresponding heat flux for three steering configurations: straight alignment and two steering positions corresponding to a 30° rotation of the wheel to the left and to the right.

The convective heat transfer coefficient shown in Figure 5 exhibits clear differences in both magnitude and spatial distribution depending on the disk’s orientation. In the straight configuration, the airflow emerging from the duct strikes the rotor relatively symmetrically, producing a moderately uniform heat transfer distribution across the disk’s surface. Most of the rotor area is characterized by heat transfer coefficient values ranging approximately between 90 and 150 W m⁻² K⁻¹, while regions located near the outer edge reach values exceeding 200 W m⁻² K⁻¹ due to flow acceleration and increased turbulence intensity. The spatially averaged convection coefficient over the rotor surface is estimated to be on the order of 120–140 W m⁻² K⁻¹, indicating stable but moderate cooling performance.

When the wheel rotates 30° to the left, the alignment between the airflow and the disk surface becomes less favorable. In this configuration, the incident jet interacts with the rotor at a more oblique angle, which reduces the effective momentum transfer to certain parts of the disk surface. Consequently, the distribution of the convective heat transfer coefficient becomes more heterogeneous. While maximum values close to 200–240 W m⁻² K⁻¹ are still observed in regions directly exposed to the incident airflow, larger areas of the rotor remain within the lower range of 80–130 W m⁻² K⁻¹. This results in a slightly lower average surface heat transfer coefficient compared to the straight configuration, suggesting a modest degradation in cooling efficiency.

In contrast, the 30° right-hand configuration provides the most favorable aerodynamic alignment between the duct outlet and the brake disc. Under these conditions, the airflow strikes the rotor surface more directly, increasing the magnitude of the local velocity and intensifying turbulence levels in the near-wall region. This interaction generates a broader distribution of high convective coefficients, with significant portions of the disc surface exhibiting values between 140 and 220 W m⁻² K⁻¹, and local maxima approaching 350–380 W m⁻² K⁻¹ near the outer radius. Consequently, the spatially averaged heat transfer coefficient in this configuration is significantly higher than in the other two cases, indicating a more efficient convective cooling process.

The influence of the disk orientation on cooling performance is further confirmed by the wall heat flux distributions shown in Figure 6. Since the heat flux at the wall is governed by the relationship between the convective heat transfer coefficient and the difference between the surface and ambient temperatures, the observed variations in the convective heat transfer coefficient are directly reflected in the heat flux patterns. In the straight configuration, most of the rotor surface exhibits heat flux values between approximately 2 and 4 kW m⁻², with localized peaks reaching 6–7 kW m⁻² near the outer perimeter.

Figure 5. Distribution of the heat transfer coefficient on the wall along the brake disc, considering three positions of the disc relative to the vehicle: turned 30 degrees to the left, turned 30 degrees to the right, and facing forward

Figure 6. Heat flux distribution across the brake disc, considering three positions of the disc relative to the vehicle: turned 30 degrees to the left, turned 30 degrees to the right, and facing forward

Figure 7. Distribution of the heat transfer coefficient in the wall along the brake disc, considering three disc surface temperatures: 600 K, 700 K, and 800 K

Figure 8. Heat flux distribution in the wall along the brake disc, considering three disc surface temperatures: 600 K, 700 K, and 800 K

In the left-hand configuration, the heat flux distribution becomes more irregular due to the asymmetric interaction between the incident airflow and the rotor surface. Regions directly exposed to the flow maintain relatively high heat dissipation rates, while areas affected by partial aerodynamic shielding show reduced heat flux values, close to 2-3 kW m⁻². This confirms that the altered flow orientation limits the overall effectiveness of the convective cooling mechanism.

On the other hand, the right-hand configuration produces the highest heat dissipation levels across the entire rotor. Large areas of the disk exhibit heat flux values between 4 and 7 kW m⁻², with localized maxima approaching 10–11 kW m⁻² near the outer radius. This represents the most effective cooling scenario among the three configurations, as a result of the greater flow impact and the greater disturbance of the boundary layer.

In summary, the results demonstrate that steering-induced variations in the relative alignment between the duct outlet and the brake disc can significantly influence the system’s convective cooling performance. The 30° right-hand steering position provides the most favorable aerodynamic interaction, resulting in higher average heat transfer coefficients and greater heat dissipation rates. Conversely, the left-hand steering configuration slightly reduces cooling efficiency due to the less-than-optimal alignment of the airflow with the rotor surface. These findings highlight the importance of considering the effects of wheel articulation during the design of forced-air brake cooling systems, particularly in commercial vehicles where steering angles during operation can alter the local flow field within the wheel assembly.

3.3 Effect of brake disc surface temperature

Figures 7 and 8 show the spatial distribution of the heat transfer coefficient at the wall and the heat flux on the surface of the brake disc for three specified temperatures: 600 K, 700 K, and 800 K. These results provide insight into how the thermal driving force and the aerodynamic field within the duct influence the convective heat dissipation of the brake disc.

The distribution of the wall heat transfer coefficient, shown in Figure 6, exhibits relatively similar spatial patterns for all three temperatures. Most of the disc surface exhibits values in the range of approximately 90–140 W m⁻² K⁻¹, while regions located near the internal ventilation channels reach values close to 160–180 W m⁻² K⁻¹. These higher coefficients are observed primarily along the inner annular region of the disc, where the airflow accelerates through the ventilation ducts. The increase in velocity in these regions intensifies turbulent mixing and reduces the thickness of the thermal boundary layer, resulting in higher local convective heat transfer coefficients.

In contrast, lower heat transfer coefficients—typically less than 60 W m⁻² K⁻¹—are observed near the outer edge of the disk and in regions where the airflow velocity decreases due to flow expansion or recirculation. These areas correspond to zones where the local flow momentum is reduced, resulting in thicker boundary layers and, consequently, weaker convective heat transfer. The general similarity in the heat transfer coefficient distributions for the three temperature conditions indicates that the convective heat transfer coefficient is determined primarily by the flow dynamics within the cooling duct, rather than by the absolute temperature of the disk surface. It should be noted that the present simulations assume constant air thermophysical properties. Under real operating conditions, variations in air viscosity and thermal conductivity at elevated temperatures may slightly modify the local convective heat transfer coefficient. However, within the range of operating conditions considered in this study, the spatial distribution of the heat transfer coefficient remains governed primarily by the airflow structure, turbulence intensity, and local velocity distribution generated by the duct configuration. Therefore, the relatively small variation observed in the convective heat transfer coefficient between the different disc temperatures is considered physically consistent with the assumptions adopted in the numerical model.

While the spatial distribution of the heat transfer coefficient remains relatively constant, the heat flux at the wall, illustrated in Figure 8, shows a clear increase with disc temperature. For the 600 K case, most regions of the disk exhibit heat flux values between approximately 2.5 and 4.5 kW m⁻², with localized peaks reaching around 6-7 kW m⁻² near the internal ventilation ducts. When the disc temperature increases to 700 K, the heat flux increases significantly across the entire surface, with typical values ranging between 4 and 7 kW m⁻², and maximum values approaching 8-9 kW m⁻² in the same high-velocity regions.

The highest thermal condition, corresponding to a disk temperature of 800 K, produces the highest heat transfer rates. Under these conditions, the heat flux at the wall across significant portions of the disk surface exceeds 6 kW m⁻², while regions located near the ventilation channels reach values close to 10 kW m⁻². Compared to the 600 K case, this represents an increase in heat flux of approximately 60–80%, reflecting the greater thermal driving force between the heated disc surface and the incoming cooling airflow.

This behavior can be explained by the classical convective heat transfer equation, considering the convective heat transfer coefficient, the disk surface temperature, and the incoming air temperature. Since the heat transfer coefficient remains relatively stable under different thermal conditions, the increase in heat flux is primarily due to the greater temperature difference between the disc surface and the cooling air as the disc temperature rises.

Another important observation is the non-uniform distribution of heat flux across the disc surface. The highest heat transfer rates occur near the internal ventilation ducts, where the airflow accelerates and turbulence increases. These flow structures promote more intense convective heat exchange and enhance the system’s local cooling capacity. In contrast, the outer regions of the disc experience lower heat transfer rates due to weaker interaction with the airflow and lower local velocities.

From a thermal management perspective, these results indicate that the cooling performance of the brake system is strongly influenced by the aerodynamic distribution of the airflow provided by the cooling duct. Areas exposed to higher-velocity flow benefit from significantly improved convective heat transfer, while areas with weaker airflow may experience reduced cooling efficiency. Therefore, optimizing the distribution of airflow within the duct and across the brake disc surface is essential for achieving a more uniform thermal field and preventing localized overheating during hard braking.

To assess the validity of the predicted convective heat transfer coefficients, the obtained values were compared with those reported in the literature for ventilated brake discs. Previous experimental and numerical studies have shown that the convective heat transfer coefficient on brake disc surfaces strongly depends on airflow conditions, disc geometry, and rotational effects. For instance, studies on ventilated brake discs have reported convective heat transfer coefficients in a similar order of magnitude, with values typically increasing with Reynolds number and airflow intensity [19, 20]. In addition, CFD investigations validated against experimental data have demonstrated good agreement between predicted and measured heat transfer coefficients for brake discs under realistic operating conditions [21]. Therefore, the range of 90-140 W m⁻² K⁻¹ obtained in the present study is consistent with previously reported values for forced convection over brake disc surfaces, supporting the physical reliability of the numerical model.