The rapid advancement of gasoline turbochargers has allowed these systems to operate in increasingly extreme environments, such as high temperatures, high speeds, and high pressures. These conditions can lead to the degradation of lubricants, causing them to carbonize, while the seal rings may accumulate carbon deposits and lose their elasticity. To ensure the proper functioning of the floating bearing and the sealing ring, it is generally recommended that the temperature at the turbine end's floating bearing area does not exceed 150°C, and the temperature at the sealing ring should remain below 230°C. This paper conducts a fluid-solid coupling analysis and steady-state thermal analysis of a water-cooled bearing body to determine the temperature distribution within the bearing, evaluate the design of the coolant channels, and verify whether the size of the coolant inlet flow is appropriate. This research provides a method for the development of bearing body designs.
**Fundamental**
1. **Calculation Model Analysis**
The fluids involved in the calculation include engine coolant and lubricating oil, while the solid component is the bearing body. The two fluids do not mix or come into direct contact, and their movement does not cause significant deformation of the bearing body. However, it does influence the flow path of the fluid. This is a one-way coupling problem where heat exchange occurs between the fluid and the bearing body. There is mutual constraint between the fluid and the solid wall, and the thermal boundary condition cannot be predetermined. Instead, it is a typical weak coupling problem, where heat exchange only occurs at the boundary, and both the temperature and heat transfer coefficient on the boundary are part of the solution. A more effective approach is to perform a fluid-solid coupling analysis by combining the fluid in the cavity with the bearing body to achieve accurate heat transfer between the fluid and the solid wall.
2. **Heat Transfer Theory**
When a fluid flows inside the bearing body cavity, a boundary layer forms near the wall surface, where the fluid velocity is very low, almost zero. Therefore, heat exchange between the fluid and the wall is primarily conducted through conduction. Outside this region, heat transfer occurs mainly through convection. The heat transfer due to conduction can be calculated using Fourier’s Law:
$$ q = -\lambda \frac{dT}{dx} $$
Where $ \lambda $ is the thermal conductivity (W/m·K), and $ \frac{dT}{dx} $ is the temperature gradient.
For convective heat transfer, Newton’s Law of Cooling is used:
$$ q = h(T_w - T_f) $$
Where $ h $ is the heat transfer coefficient (W/m²·K), $ T_w $ is the wall temperature (K), and $ T_f $ is the fluid temperature (K).
According to the principle of energy conservation, the heat released by the fluid at the boundary must equal the heat absorbed by the solid. This relationship is expressed as:
$$ \lambda \frac{dT}{dx} = h(T_w - T_f) $$
This equation ensures that the heat transfer at the interface is accurately modeled.
**Establishment of a Computational Model**
This study focuses on the temperature field distribution of the bearing body under the influence of coolant and lubricating oil. To reduce computational time and resource usage, the actual fluid domains (coolant and lubricating oil) are not included in the model. Instead, the temperature at the interface between these regions and the bearing body is obtained through experimental testing and simulation, and then applied as a boundary condition. The heat exchange between the turbine casing and the volute with the bearing body occurs via conduction, and the outer surface of the bearing body also exchanges heat with the surrounding air.
1. **Geometric Model Processing**
The research uses a gasoline engine supercharger developed by our company, which features a water-cooled bearing body. In addition to lubricating the floating bearing, the lubricating oil also helps dissipate heat from the bearing body. The three-dimensional model includes the bearing body, cooling water channel, and lubricating oil channel. First, a 3D model of the bearing body is created in UG, and non-critical features like chamfers and positioning holes are simplified. The 3D models of the cooling water and lubricating oil channels are derived from the bearing body model, and the fluid domains are simplified accordingly.
2. **Meshing of the Computational Model**
For the bearing body mesh, a tetrahedral mesh is used, processed using ANSYS structural meshing technology, with refined areas in critical zones. The cooling water and lubricating oil fluid domains are meshed using CFD grids, with special attention given to the boundary layers. The resulting grid models are shown in Figures 1, 2, and 3.
**Calculation Results**
Using CFX, a fluid-solid coupling analysis is performed, separating the fluid parts from the bearing body to calculate the surface temperature and convective heat transfer coefficients. These coefficients are then applied as loads on the corresponding surfaces of the bearing body, followed by a steady-state thermal analysis to obtain the full temperature distribution of the bearing body. The calculated thermal convection coefficients for the cooling water and lubricating oil bodies are shown in Figures 4 and 5, respectively. The final temperature distribution of the bearing body is presented in Figure 6.
**Conclusion**
The calculated temperature distribution aligns well with experimental results, confirming the feasibility of the analytical method used in this study. From the cloud data, the temperatures near the floating bearing and the seal ring are below the allowable limits, proving that the cooling design of the bearing body is effective and suitable for real-world applications.
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