Method for heat transfer reduction order prediction of natural gas engine turbocharger bearing body

Abstract

The application relates to a heat transfer reduction prediction method for a turbocharger bearing body of a natural gas engine, and relates to the technical field of supercharging devices. The turbocharger bearing body of the natural gas engine comprises a heat protection structure, the heat protection structure comprises a heat shield, a heat insulation cavity and a ventilation hole, the heat transfer reduction prediction method is realized based on the heat protection structure, the heat transfer of a turbine bearing coupling system is disassembled into three independent heat transfer paths, the three independent heat transfer paths comprise a volute bearing body flange solid conduction path, a heat shield cavity gas convection heat exchange path and a non-shielded bearing body exposed area convection heat exchange path, and a reduction calculation model is constructed. The application solves the problems in the prior art, such as long three-dimensional simulation cycle, difficulty in realizing multi-parameter rapid iteration, unknown bearing body heat transfer path, unable to quantify heat contribution proportion, lack of theoretical support for heat protection design, gas leakage problem of a single heat shield, poor heat insulation effect caused by poor ventilation hole design, and difficulty in weakening the dominant solid flange heat conduction problem of a single heat insulation structure.

Description

Technical Field

[0001] This invention relates to the field of turbocharger technology, specifically to a method for predicting the heat transfer degradation of the bearing housing of a natural gas engine turbocharger. Background Technology

[0002] Natural gas engines, with their core advantages of lower fuel operating costs and significantly lower carbon and particulate emissions compared to diesel engines, are seeing a continuous increase in market penetration in areas such as long-haul logistics trucks, urban sanitation vehicles, mining machinery, port transfer equipment, and distributed gas generator sets, forming a structural replacement trend for traditional diesel power. However, the combustion characteristics of gas engines and diesel engines differ fundamentally, directly resulting in a significant difference in exhaust temperature: the steady-state exhaust temperature of conventional diesel engines is only 600-750℃, while the normal exhaust temperature of natural gas engines reaches 760-850℃, and the instantaneous exhaust temperature can exceed 900℃ during heavy-load climbing, rapid acceleration, and long-distance high-load driving, which is more than 100℃ higher than that of diesel engines. Radiant and conductive heat from the turbine end continuously accumulates in the bearing housing, subjecting it to extreme thermal load conditions for extended periods. Coupled with the periodic hot and cold shocks caused by frequent start-stop cycles in commercial vehicles and the alternating shocks of ultra-high rotor speeds of 100,000-200,000 r / min, the traditional bearing housing cooling and insulation structures designed for diesel engine operating conditions, as well as the turbine reliability design, all fail.

[0003] Industry statistics show that heat-induced failures account for as much as 62% of turbocharger failures in gas engines, with the core failure modes concentrated in four categories: First, high-temperature creep and thermal deformation failure of the turbine end seal ring, leading to exhaust gas entering the lubrication oil circuit, causing oil dilution and oil pressure instability; Second, high temperature of the bearing housing causes the lubricating oil to decompose and carbonize, resulting in carbon accumulation on the shaft surface, disruption of dynamic balance, and causing vibration and abnormal noise of the entire machine; Third, high-temperature lubrication failure of the floating bearing, causing a sharp drop in oil film load-bearing capacity, direct dry friction between the bearing and the shaft, resulting in rapid wear and a significant reduction in turbocharger life; Fourth, turbine "thermal-mechanical" low-cycle fatigue failure caused by large-scale rapid alternating operating conditions. Market research shows that the average trouble-free operating time of a conventional diesel engine turbocharger after being equipped with a gas engine is only 8,000-12,000 hours, less than 60% of that of diesel engine-equipped models. The annual replacement rate of turbochargers for gas-powered heavy-duty trucks in long-haul logistics fleets exceeds 35%, with a replacement cost of 5,000-12,000 yuan per unit. The total annual maintenance and spare parts replacement expenditure for gas engine turbochargers nationwide exceeds 18 billion yuan, causing huge operating cost losses for logistics companies and equipment manufacturers.

[0004] The shortcomings of current mainstream industry solutions are obvious: First, relying solely on high-temperature resistant special alloy materials to increase the heat resistance limit of the bearing body results in a 120%-200% increase in the cost of special stainless steel and nickel-based alloy materials compared to ordinary cast iron, which significantly compresses the profit margins of main engine and component companies and makes it difficult to commercialize in large quantities. Secondly, simply increasing the cooling water flow and adding an external heat shield without clarifying the multi-field coupled heat transfer mechanism of turbine radiation, oil cooling convection, and solid heat conduction within the bearing body increases cooling energy consumption by more than 15% while the thermal protection effect is minimal, failing to fundamentally suppress core problems such as seal failure, carbon buildup, and bearing wear.

[0005] In predicting the heat transfer characteristics of bearing housings, existing technologies mainly rely on two methods: whole-machine bench testing and three-dimensional fluid-structure-thermal coupling simulation. However, these methods also have obvious limitations. The bench testing equipment requires a large investment and the test point layout is limited. It is impossible to accurately dissect the heat ratio of the three heat transfer paths: solid conduction through the flange, convection through the insulation cavity, and heat exchange through the exposed turbine back plate. Multi-condition cyclic testing will also bring extremely high time and prototype wear and tear costs. In three-dimensional multiphysics simulation, the multi-scale coupled calculation of turbine high-speed rotating fluid, bearing lubricating oil Couette laminar flow, and solid high-temperature heat conduction has a large degree of freedom. The convergence time for a single constant working condition can be several hours. If multi-variable optimization design such as heat shield ventilation hole size, flange thickness, and cooling oil flow rate is carried out, the massive working condition simulation requirements will seriously slow down the research and development cycle.

[0006] In summary, as the existing technology has been used, its shortcomings have gradually become apparent, mainly in the following aspects: First, traditional 3D simulation has a long calculation cycle and high bench testing costs, which cannot support rapid iterative optimization of multiple parameters. Second, the heat transfer path of the bearing body is unclear, and the heat contribution ratio of each path cannot be quantified, resulting in a lack of theoretical guidance for the design of thermal protection structures. Third, the problem of air leakage through the gaps in the inner edge of the single-layer heat shield and the problem of poor design of ventilation holes leading to the deterioration of the heat insulation effect; Fourth, a single insulation structure can only regulate the convective heat transfer path and cannot effectively reduce the problem of the dominant solid flange heat conduction path.

[0007] In conclusion, the existing technology obviously has inconveniences and defects in practical use, so it is necessary to improve it. Summary of the Invention

[0008] To address the shortcomings of existing technologies, this invention solves the problems of long three-dimensional simulation cycles, high bench testing costs, and difficulty in achieving rapid iteration of multiple parameters in traditional technologies; unclear heat transfer paths of bearing bodies, inability to quantify the proportion of heat contribution, and lack of theoretical support for thermal protection design; gas leakage problems in single-layer thermal shields, poor ventilation hole design leading to poor insulation effect, and difficulty in weakening the dominant heat conduction of solid flanges by a single insulation structure.

[0009] To address the above problems, the present invention provides the following technical solution: A method for predicting heat transfer order reduction in a natural gas engine turbocharger bearing housing, wherein the natural gas engine turbocharger bearing housing includes a thermal protection structure, which includes a heat shield, a heat insulation cavity, and ventilation holes. The method is based on the thermal protection structure and decomposes the heat transfer of the turbine bearing coupling system into three independent heat transfer paths: a solid conduction path through the turbine housing bearing flange, a gas convection heat transfer path through the heat shield cavity, and a convection heat transfer path through the exposed area of ​​the unshielded bearing housing. A reduction calculation model is constructed, including the following steps: Step S1: Establish the heat balance equation and the total heat superposition formula; perform volume integral on all solid control volumes of the bearing housing, and combine the Gaussian divergence theorem to decompose the flange heat conduction boundary, the heat shield cavity convection boundary, the turbine exposed area convection boundary, and the lubricating oil cooling boundary under steady-state heat balance conditions to obtain the total heat relationship of the superposition of heat flow of the three heat transfer paths, and define the ratio of heat flow of each path to the total input heat as the heat proportion coefficient. Step S2: Solid heat conduction path of the volute bearing flange; the volute flange, gasket, heat shield ring, and bearing housing flange are considered as a multi-layer series solid heat conduction structure. Based on Fourier's law of thermal conductivity, combined with the thermal conductivity, geometric thickness, and interlayer contact thermal resistance of each layer, the total thermal resistance of the multi-layer series structure is calculated, thereby obtaining the solid heat conduction from the flange to the bearing body; the interlayer contact thermal resistance is determined by the bolt preload and the surface roughness of the mating surfaces. Step S3: Gas convection heat transfer path in the heat shield cavity; the heat insulation cavity enclosed by the turbine and bearing housing is treated as an independent control body. Based on the characteristic dimensions of the heat insulation cavity, the temperature difference of the cavity wall, and the gas physical properties, the Rayleigh number is calculated. The convection heat transfer coefficient of the heat insulation cavity is solved using the natural convection Nusselt number correlation to obtain the convection heat transfer from the heat insulation cavity to the bearing housing. At the same time, the compressible subsonic throttling flow rate of the ventilation hole is calculated, and the heat brought into the heat insulation cavity by the gas leaking through the ventilation hole is added to the convection heat transfer. The alternating bending stress of the thin-walled heat shield under alternating hot and cold conditions in the closed heat insulation cavity is calculated. The fatigue damage of the heat shield is calculated using the fatigue damage accumulation criterion to determine the minimum flow area of ​​the ventilation hole used for pressure equalization, taking into account both the fatigue reliability and heat insulation performance of the heat shield. Step S4: Convective heat transfer path in the exposed area of ​​the unshielded bearing housing; Calculate the rotating Reynolds number and Prandtl number based on the turbine rotor angular velocity, characteristic radius and gas properties, solve the convective heat transfer coefficient of the exposed wall using the rotating disk Nusselt number correlation, and calculate the convective heat transfer brought by the high-temperature gas in the back disk gap to the exposed wall of the bearing housing. Step S5: Lubricating oil cooling and heat transfer calculation; the total heat carried away by the lubricating oil is divided into two parts: the heat carried away by the bearing clearance oil film and the heat carried away by the oil supply pipeline; the bearing clearance shear Reynolds number is calculated, and when the shear Reynolds number is less than the preset laminar flow judgment threshold, the oil film is determined to be Cueter laminar flow, and the shear laminar flow heat transfer correlation formula is used to solve the bearing clearance oil film heat transfer coefficient; the lubricating oil heat transfer coefficient in the oil supply pipeline is taken as a fixed value, and the heat transfer of the two parts of lubricating oil is calculated separately and summed; Step S6: Calculation of heat transfer in the vortex shell; Based on the free vortex isocirculation condition, the Euler equation of the turbine and the one-dimensional compressible gas flow conservation equations, solve for the gas temperature distribution and airflow velocity distribution in the vortex shell flow channel; establish a one-dimensional forced convection heat transfer model of the inner wall of the vortex shell and solve for the total heat transfer from the gas to the vortex shell. Step S7: Calculation of natural convection heat dissipation from the vortex shell to the environment; Discretize the vortex shell into several one-dimensional micro-element segments along the circumference, calculate the Grashof number and Nusselt number of natural convection on the outer wall of each micro-element segment, solve for the heat dissipation of each micro-element segment to the outside and sum them to obtain the total heat dissipated from the outer wall of the vortex shell to the environment, and use the heat dissipation result as the convergence criterion for the shell heat conduction iteration. Step S8: Iterative solution of the model; perform iterative calculations on all temperature nodes of the reduced-order model, and use the maximum relative deviation between two adjacent iterations of the temperature of each node as the iterative residual. Stop the iteration when the iterative residual meets the preset convergence criterion, and output the bearing body temperature field, the heat flow of the three heat transfer paths, and the heat contribution ratio of each path.

[0010] As an optimized approach, in step S1, the total heat accumulation formula is as follows: , Therefore, the heat transfer in the bearing housing can be mainly divided into three heat transfer paths, namely: Path: Solid heat conduction path of the volute bearing flange; Path: Gas convection heat transfer path in the heat shield cavity; Path: Convection heat transfer path in the exposed area of ​​the unshielded bearing housing.

[0011] In the formula, The lubricating oil carries away the total heat. The total heat transferred to the bearing housing. Thermal conductivity of solid materials such as bearing housings, flanges, and heat shields, expressed in W / (m⋅K); The thermodynamic temperature at any location within the solid domain, in K; The proportions of the three heat transfer paths are as follows: To identify the contribution of the three heat transfer paths, each path has a corresponding percentage coefficient: , .

[0012] As an optimized solution, the single-layer steady-state Fourier heat flux in step S2 is: , In the formula, For the thickness of a single layer of material, For interlayer temperature difference, denoted as the thermal conductivity of the material.

[0013] , In the formula, and The constant is determined by the surface roughness of the material. Therefore, when the bolt preload is fixed, It is a constant. Effective contact area of ​​the flange. Total heat transfer flow: , In the formula, The steady-state wall temperature at the stop of the vortex shell. This refers to the steady-state wall temperature of the bearing housing stop.

[0014] Take the partial derivative with respect to the structural design variables in the formula: , Flange dimensions, Bolt preload, The thickness of the parts is a constraint parameter of the overall machine structure and assembly process. Passive thermal insulation structures on the bearing side, such as heat shields, ventilation holes, and double-layer insulation, only act on the thermal boundary of the insulation cavity and do not change the flange series thermal resistance system. , For the area covered by the heat shield, This refers to the area of ​​the bearing housing exposed within the backplate clearance. Ventilation opening flow area; Therefore, insulation optimization methods cannot significantly change the situation. , Naturally, it becomes the dominant heat source for heat transfer to the bearing housing.

[0015] As an optimized solution, in step S3, the convective heat transfer at the interface of the insulation cavity satisfies the time-averaged energy boundary coupling: , Total convective heat flow: , , In the formula, The projected area of ​​the bearing housing covered by the heat shield. The average convective heat transfer coefficient of the insulation cavity is given. The average gas temperature inside the insulation cavity. The turbulent heat transfer coefficient of the insulation cavity is expressed using the dimensionless Nusselt correlation: , In the formula, The characteristic length of the cavity; The Reynolds number for the cavity airflow; For Grashof numbers; , , In the formula, The average air intake velocity of the ventilation holes. The coefficient of thermal expansion of the gas is _____. Kinematic viscosity; Therefore, The airflow rate is directly controlled by the ventilation holes and is the core variable for regulation. The above analysis treats the insulation cavity as a controlled volume rather than a closed cavity. If it is treated as a closed cavity, under alternating hot and cold conditions, the gas inside the insulation cavity will exert alternating loads on the stamped thin-walled thermal shield, which can easily lead to thermal shield fatigue failure. Assuming the insulation cavity is considered a closed body, the cavity volume is... The internal gas satisfies the ideal gas law: , Turbine temperature cycling range under varying operating conditions Intracavitary pressure fluctuation amplitude: , Thin-walled stainless steel thermal shields experience membrane bending stress due to cyclic pressure differentials. Theoretical explanation of thin-plate bending stress: , It is alternating compressive stress. Alternating thermal stress, The coefficient of thermal expansion of the material. The elastic modulus of the material. The inner diameter of the heat shield. For the thickness of the heat shield; Under the continuous action of this alternating stress, the Miner linear damage criterion is used to calculate the cumulative fatigue damage of the material: , In the formula, The number of alternations between hot and cold operating conditions. This represents the fatigue life corresponding to alternating stress. Inside the sealed insulated cavity, It can reach over 0.2 MPa, while the inner diameter of the heat shield... Much larger than its thickness When specific geometric parameters are introduced, the pressure drops to the tens of megapascals, resulting in a thermal shield fatigue life far lower than the required life of the turbocharger. Therefore, in engineering practice, ventilation holes must be installed to achieve pressure equalization. The presence of a ventilation opening inevitably leads to the exchange of compressible flow mass. Since the ventilation opening is a thin-walled throttling orifice, the compressible subsonic throttling flow rate formula is used: , The throttling flow coefficient; This represents the total flow area of ​​the ventilation holes; Specific heat ratio of gases; Mass flow rate of high-temperature gas flowing into the cavity Directly increases the average temperature of the cavity gas At the same time, increase the cavity ; Simultaneous formula: , Therefore, the requirements for structural reliability necessitate the inclusion of ventilation holes in the heat shield to balance the pressure inside and outside the insulation chamber. This increases mass transfer within the insulation chamber and turbine flow channels, thereby promoting… The addition of heat shields weakens their thermal protection effectiveness to some extent.

[0016] As an optimized solution, in step S4, the area of ​​the turbine backplate not covered by the heat shield directly contacts the high-temperature mainstream combustion gas, and the convective heat transfer flux is: , Total heat flow along this path: , The formula requires calculation of the heat transfer coefficient under rotating conditions. , Define the Reynolds number for the rotating state: , In the formula, The angular velocity of the turbine rotor. The radius of the pivot is 1. For high-temperature gas density, The dynamic viscosity of the gas; Backplate clearance gas Prandtl number: , In the formula, For the specific heat capacity of the gas at constant pressure, is the thermal conductivity of the gas.

[0017] Nusselt number correlation for rotating disk convection: , In the formula, , Therefore, we can obtain: .

[0018] As an optimized solution, in step S5, the heat carried by the lubricating oil is divided into two parts: , In the formula, The heat is carried away from the bearing clearance. This refers to the heat carried away from the oil supply line.

[0019] Regarding heat transfer within the bearing clearance: The shear Reynolds number is , In the formula, For the density of lubricating oil, The viscosity of the lubricating oil. The linear velocity of the floating bearing. This refers to the clearance of the outer oil film in the floating bearing. Lubricating oil Prandtl number , The specific heat capacity of lubricating oil at constant pressure. This refers to the thermal conductivity of the lubricating oil.

[0020] Due to the external oil film state Much less than 1000, belonging to the laminar flow range, narrow-gap rotating shear laminar flow, classic Couette flow heat transfer correlation: , The heat transfer coefficient of the bearing housing at the position of the outer oil film is: , Because the oil supply management is relatively coarse and the flow rate is slow, the heat transfer coefficient of the lubricating oil in the oil supply pipeline is treated as a constant: , In the formula, The average temperature of the oil supply pipe wall. The oil inlet temperature, This refers to the surface area of ​​the inner wall of the oil supply pipe.

[0021] As an optimized solution, in step S6, After establishing the bearing housing heat transfer model, assuming that the combustion gas satisfies the free vortex equal circulation condition in the vortex flow channel, the heat transfer calculation models of the vortex and turbine for the bearing housing heat input need to be established simultaneously: , In the formula, The radial radius of the flow channel. The tangential velocity of the gas. The cross-sectional area of ​​the vortex casing flow channel increases linearly along the circumference to ensure uniform and stable gas supply to the turbine inlet. The work done by turbine expansion follows Euler's equations: , In the formula, 1 represents the turbine inlet linear velocity, 1 represents the turbine inlet, and 2 represents the turbine outlet.

[0022] The one-dimensional conservation equations for compressible gas flow, including the continuity equation: , In the formula, This refers to the mass flow rate of the gas. For Zhou Xiang The cross-sectional area of ​​the vortex shell during flow; This refers to the speed of the meridional airflow. The density of the gas.

[0023] Momentum equation: , In the formula, For the centrifugal inertial force term of the free vortex, Energy equation: , In the formula, Total enthalpy of the gas; For static enthalpy, Ideal gas law: , Static temperature of combustion gas at various locations on the inner wall of the vortex casing airflow speed , as input to the inner wall convection heat transfer model.

[0024] in, , The above model establishes the fluid heat transfer boundary for vortex heat transfer. Next, a one-dimensional forced convection heat transfer model along the vortex flow channel is established to solve for the heat flux density of the inner wall. The total heat absorbed by the vortex shell is obtained. : vortex shell Local gas Reynolds number on cross section: , In the formula, The hydraulic diameter of the vortex flow channel; For high-temperature gas dynamic viscosity, with Interpolation, Gas Prandtl number: , In the formula, The thermal conductivity of the exhaust gas; Nusselt number correlation for forced convection in annular channels: , Suitable for turbulent flow with Re > 2300, and matched for high-flow, high-temperature exhaust gas conditions in natural gas engines; Forced convection heat transfer coefficient of annular channel: , One-dimensional convective heat transfer distribution of fluid in a vortex shell towards the vortex shell wall: , The solid temperature inside the vortex shell. Total heat exchange of gas on the inner wall of the vortex shell: The vortex shell is discretized circumferentially into N one-dimensional infinitesimal segments, where b is the axial width of the vortex shell and the heat transfer area of ​​each infinitesimal segment. : , That is, the heat transferred from the turbine flow channel to the volute casing.

[0025] As an optimized solution, in step S7, On one hand, the heat is transferred to the bearing body through path 1; on the other hand, some of it is lost through heat dissipation from the outer wall of the volute shell towards the environment. This portion of the lost heat is denoted as: , This formula serves as the convergence criterion for the heat conduction iteration of the vortex shell, achieving the coupling of heat transfer on the gas side, internal heat conduction of the shell, and external heat dissipation of the shell. The vortex shell dissipates heat to the environment. Assuming the vortex shell is a rotating shell, along the circumferential direction... Divide into one-dimensional infinitesimal elements, and the outer surface area of ​​the infinitesimal element is: , The axial width of the vortex shell; The outer diameter of the vortex shell at the micro-element; ambient air temperature. The engine compartment temperature is typically between 25-80℃. Calculation of dimensionless number of natural convection on external surface, Grashof number: , It is the acceleration due to gravity; The coefficient of thermal expansion of air; This refers to the temperature of the outer wall of the vortex shell. The characteristic length of the outer surface of the vortex shell; The viscosity is the kinematic viscosity of air.

[0026] Natural convection Nusselt number correlation: , In the formula, Take 0.703, applicable scope; The local heat transfer coefficient for dissipating heat to the environment is: , In the formula, The thermal conductivity of air. The total heat lost to the environment is: .

[0027] As an optimized solution, in step S8, the above turbocharger turbine-bearing housing heat transfer reduced-order prediction model is used in... Iteration between nodes, convergence criterion: , In the formula, , Based on the results of the above order reduction prediction method, it is known that: reducing the temperature inside the insulation cavity... That is, reducing the mass transfer between the heat insulation cavity and the turbine flow channel is to reduce This further reduces An important way; by increasing , thereby increasing Thus reducing This reduces the main heat transfer path of the bearing housing. An important way of transferring heat.

[0028] The present invention also discloses a thermal protection structure for a natural gas engine turbocharger bearing housing, including a compressor, a bearing housing, and a turbine. A heat shield is fixed between the bearing housing and the turbine back of the turbine, and a heat insulation cavity is formed between the heat shield and the bearing housing. The inner edge of the heat shield is fitted and sealed with the mating surface of the bearing housing to prevent high-temperature gas from entering the heat insulation cavity from the back plate gap. The heat shield is provided with ventilation holes to balance the pressure inside and outside the heat insulation cavity; The turbine includes a finned heat dissipation vortex housing, and heat dissipation fins are evenly distributed circumferentially on the outer wall of the vortex housing.

[0029] Compared with the prior art, the beneficial effects of the present invention are: (1) The reduced-order model has high computational efficiency and supports rapid iteration of multiple parameters. The bearing body heat transfer reduced-order prediction model constructed in this invention completely preserves the solid thermal conductivity of the flange. Convection in insulation cavity Exposed bearing housing heat exchange Based on the physical mechanisms of the three core heat transfer pathways, lubricating oil cooling, and natural heat dissipation from the outer wall of the volute, the solution degree of freedom is greatly reduced, compressing the calculation time of traditional three-dimensional CFD simulation from several hours to seconds. It can support rapid optimization design of multiple variables such as heat shield ventilation hole size, flange thickness, and cooling oil flow rate, significantly shortening the R&D cycle. (2) High prediction accuracy, with a maximum deviation of 27K and a relative error of <5%. The reduced-order prediction model was used to calculate the turbochargers of multiple product platforms and compared with the experimental results. Under the respective turbine design conditions, the trend of the prediction results and the experimental results was basically consistent with the change of the turbine inlet temperature. The maximum prediction deviation was 27K and the relative deviation was less than 5%, which met the engineering prediction accuracy requirements. The calculation results can be used for rapid structural iteration and guide the design of efficient thermal management structure for turbochargers. (3) For the first time, the proportion of the three heat transfer paths was quantitatively decomposed, and the optimization direction was clarified. This invention quantifies the heat contribution of the three heat transfer paths of the turbocharger bearing housing for the first time through a reduced-order model: Solid thermal conductivity accounts for 72%-82% of the total heat load, making it the core heat source; Convection within the insulation cavity accounts for 16%-25%, making it a secondary control target. The fact that the heat exchange rate in the exposed area of ​​the bearing housing is less than 3% clarifies that a single insulation structure can only optimize convection and cannot fundamentally solve the underlying contradiction of the dominant heat conduction load. For the first time, a passive thermal management design logic of "cooling on the heat source side and flow control in the insulation cavity" is established, providing clear theoretical guidance for the design of thermal protection structures. (4) The inner edge sealing design significantly improves the heat insulation effect. The inner edge sealing heat shield proposed in this invention blocks the high-temperature gas from entering the heat insulation cavity along the inner edge gap, and limits the heat exchange mode in the heat insulation cavity to inefficient natural convection. Experimental data show that under the high temperature condition of 1023K at the front of the vortex, the fully sealed configuration achieves a maximum temperature difference of 23.1K at the sealing ring and 7.8K at the floating bearing compared with the fully open configuration. The heat protection advantage of the sealed inner edge is significant in the entire operating range. (5) The pressure difference balanced ventilation hole takes into account both reliability and heat insulation performance. This invention reveals the contradiction of "pressure relief-heat insulation" in the design of ventilation holes and proposes the design principle of pressure difference balanced ventilation hole: the single hole scheme should be given priority to obtain robust thermal protection that is not sensitive to changes in operating conditions; the double hole scheme must arrange the ventilation holes at a position where the circumferential pressure is approximately equal to avoid the pressure difference driving the high temperature gas to flow into the inner cavity. (6) The double-layer thermal shield effectively suppresses natural convection. The double-layer thermal shield structure proposed in this invention divides a single insulated cavity into two sealed chambers by adding an auxiliary thermal shield. This segmentally compresses the temperature difference between the inside and outside of the cavity, weakening the convective heat transfer driven by buoyancy from the source. The temperature difference in the outer cavity is about 100K and in the inner cavity is about 200K. The buoyancy driving force is greatly reduced, and the fluid in the cavity is almost stagnant, thus inhibiting the convective heat transfer branch. The heat exchange rate decreased significantly; (7) The coupling optimization scheme has a significant cooling effect. The present invention combines the optimal scheme of coupling the double-layer heat shield with the finned heat dissipation vortex shell, and achieves a significant temperature drop at the three key monitoring points of the bearing body: the average temperature drop of the vortex end sealing ring is 10.5K, the average temperature drop of the floating bearing is 4.7K, and the total heat transferred to the bearing body by the whole machine is reduced by 0.22kW. The cooling effect of this scheme is stable in the wide turbine inlet temperature range of 873K-1023K. It is expected that under the same bearing body heat load, the turbine inlet exhaust temperature can be increased from 750℃ to 855℃, which can basically meet the needs of conventional automotive gas engines. (8) Low cost and easy to be promoted in engineering. The thermal protection structure of the present invention is a passive optimized design. There is no need to add a cooling system or use expensive special alloy materials. High-efficiency thermal protection can be achieved through structural optimization. The cost increase is limited. It is easy to mass-produce and promote in engineering. It has significant economic benefits and market application value. Attached Figure Description

[0030] To more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the accompanying drawings used in the description of the specific embodiments or the prior art will be briefly introduced below. In all the drawings, similar elements or parts are generally identified by similar reference numerals. In the drawings, the elements or parts are not necessarily drawn to scale.

[0031] Figure 1 This is a schematic diagram of the calculation steps of the reduced-order prediction model of the present invention; Figure 2 This is a schematic diagram comparing the predicted temperature of the reduced-order model of this invention with experimental results; Figure 3 This is a schematic diagram showing the selection of the pressure balance position for the ventilation holes in this invention; Figure 4 This is a schematic diagram of the thermal protection structure of the present invention; Figure 5This is a schematic diagram of the structure of the heat shield of the present invention; Figure 6 This is a schematic diagram of the ventilation hole structure of the present invention; Figure 7 This is a schematic diagram of the vortex shell flow channel of the present invention; Figure 8 This is a schematic diagram of the back plate clearance structure of the present invention; Figure 9 This is a schematic diagram of the structure of the heat sink fins of the present invention; Figure 10 This is a schematic diagram of the main heat shield and auxiliary heat shield of the present invention.

[0032] In the diagram: 1-compressor, 2-bearing housing, 201-bearing housing mating surface, 3-turbine, 301-volute flow channel, 302-finned volute, 303-volute outer wall, 304-heat dissipation fins, 4-thermal protection structure, 401-inner edge of the heat shield, 402-back plate clearance, 403-heat insulation cavity, 404-ventilation hole, 405-heat shield, 406-main heat shield, 407-auxiliary heat shield. Detailed Implementation

[0033] The embodiments of the technical solution of the present invention will now be described in detail with reference to the accompanying drawings. These embodiments are merely illustrative of the technical solution of the present invention and are therefore intended to limit the scope of protection of the present invention.

[0034] like Figures 1 to 10 As shown, a method for predicting the heat transfer order reduction of a natural gas engine turbocharger bearing housing 2 is described. The natural gas engine turbocharger bearing housing 2 includes a thermal protection structure 4, which comprises a heat shield 405, a heat insulation cavity 403, and a ventilation hole 404. The heat transfer order reduction prediction method is based on the thermal protection structure 4 and decomposes the heat transfer of the turbine bearing coupling system into three independent heat transfer paths: a solid conduction path through the turbine housing bearing flange, a gas convection heat transfer path through the heat shield 405 cavity, and a convection heat transfer path through the exposed area of ​​the unshielded bearing housing. A reduction calculation model is constructed, including the following steps: Step S1: Establish the heat balance equation and the total heat superposition formula; perform volume integral on all solid control volumes of the bearing housing, and combine the Gaussian divergence theorem to decompose the flange heat conduction boundary, the heat shield cavity convection boundary, the turbine exposed area convection boundary, and the lubricating oil cooling boundary under steady-state heat balance conditions to obtain the total heat relationship of the superposition of heat flow of the three heat transfer paths, and define the ratio of heat flow of each path to the total input heat as the heat proportion coefficient. Step S2: Solid heat conduction path of the volute bearing flange; the volute flange, gasket, heat shield ring, and bearing housing flange are considered as a multi-layer series solid heat conduction structure. Based on Fourier's law of thermal conductivity, combined with the thermal conductivity, geometric thickness, and interlayer contact thermal resistance of each layer, the total thermal resistance of the multi-layer series structure is calculated, and thus the solid heat conduction transferred from the flange to the bearing housing 2 is obtained; the interlayer contact thermal resistance is determined by the bolt preload and the surface roughness of the mating surfaces. Step S3: Gas convection heat transfer path in the heat shield cavity; The heat insulation cavity 403 enclosed by the turbine and bearing 2 is treated as an independent control body. Based on the characteristic dimensions of the heat insulation cavity 403, the temperature difference of the cavity wall, and the gas physical properties, the Rayleigh number is calculated. The convection heat transfer coefficient of the heat insulation cavity 403 is solved using the natural convection Nusselt number correlation to obtain the convection heat transfer from the heat insulation cavity 403 to the bearing 2. At the same time, the compressible subsonic throttling flow rate of the ventilation hole 404 is calculated, and the heat brought into the heat insulation cavity 403 by the gas passing through the ventilation hole 404 is added to the convection heat transfer. The alternating bending stress of the thin-walled heat shield 405 under alternating hot and cold conditions is calculated. The fatigue damage of the heat shield 405 is calculated using the fatigue damage accumulation criterion to determine the minimum flow area of ​​the ventilation hole 404 used for pressure equalization, taking into account both the fatigue reliability and heat insulation performance of the heat shield 405. Step S4: Convective heat transfer path in the exposed area of ​​the unshielded bearing housing; Calculate the rotating Reynolds number and Prandtl number based on the turbine rotor angular velocity, characteristic radius and gas properties, solve the convective heat transfer coefficient of the exposed wall using the rotating disk Nusselt number correlation, and calculate the convective heat transfer brought by the high-temperature gas in the back disk gap 402 on the exposed wall of bearing housing 2. Step S5: Lubricating oil cooling and heat transfer calculation; the total heat carried away by the lubricating oil is divided into two parts: the heat carried away by the bearing clearance oil film and the heat carried away by the oil supply pipeline; the bearing clearance shear Reynolds number is calculated, and when the shear Reynolds number is less than the preset laminar flow judgment threshold, the oil film is determined to be Cueter laminar flow, and the shear laminar flow heat transfer correlation formula is used to solve the bearing clearance oil film heat transfer coefficient; the lubricating oil heat transfer coefficient in the oil supply pipeline is taken as a fixed value, and the heat transfer of the two parts of lubricating oil is calculated separately and summed; Step S6: Calculation of heat transfer in the vortex shell; Based on the free vortex isocirculation condition, the Euler equation of the turbine and the one-dimensional compressible gas flow conservation equations, solve for the gas temperature distribution and airflow velocity distribution in the vortex shell flow channel 301; establish a one-dimensional forced convection heat transfer model of the inner wall of the vortex shell and solve for the total heat transfer from the gas to the vortex shell. Step S7: Calculation of natural convection heat dissipation from the vortex shell to the environment; Discretize the vortex shell into several one-dimensional micro-element segments along the circumference, calculate the Grashof number and Nusselt number of natural convection on the outer wall of each micro-element segment, solve for the heat dissipation of each micro-element segment to the outside and sum them to obtain the total heat dissipated from the outer wall 303 of the vortex shell to the environment, and use the heat dissipation result as the convergence criterion for the shell heat conduction iteration. Step S8: Iterative solution of the model; perform iterative calculations on all temperature nodes of the reduced-order model, and use the maximum relative deviation between two adjacent iterations of the temperature of each node as the iterative residual. Stop the iteration when the iterative residual meets the preset convergence criterion, and output the temperature field of bearing body 2, the heat flow of the three heat transfer paths, and the heat contribution ratio of each path.

[0035] In step S1, the formula for total heat accumulation is: , Therefore, the heat transfer in the bearing housing can be mainly divided into three heat transfer paths, namely: Path: Solid heat conduction path of the volute bearing flange; Path: Gas convection heat transfer path in the heat shield cavity; Path: Convection heat transfer path in the exposed area of ​​the unshielded bearing housing.

[0036] In the formula, The lubricating oil carries away the total heat. The total heat transferred to the bearing housing. Thermal conductivity of solid materials such as bearing housings, flanges, and heat shields, expressed in W / (m⋅K); The thermodynamic temperature at any location within the solid domain, in K; The proportions of the three heat transfer paths are as follows: To identify the contribution of the three heat transfer paths, each path has a corresponding percentage coefficient: , .

[0037] In step S2, the monolayer steady-state Fourier heat flux is: , In the formula, For the thickness of a single layer of material, For interlayer temperature difference, denoted as the thermal conductivity of the material.

[0038] , In the formula, and The constant is determined by the surface roughness of the material. Therefore, when the bolt preload is fixed, It is a constant. Effective contact area of ​​the flange. Total heat transfer flow: , In the formula, The steady-state wall temperature at the stop of the vortex shell. This refers to the steady-state wall temperature of the bearing housing stop.

[0039] Take the partial derivative with respect to the structural design variables in the formula: , Flange dimensions, Bolt preload, The thickness of the parts is a constraint parameter of the overall machine structure and assembly process. Passive thermal insulation structures on the bearing side, such as heat shields, ventilation holes, and double-layer insulation, only act on the thermal boundary of the insulation cavity 403 and do not change the flange series thermal resistance system, that is: , For the area covered by the heat shield, This refers to the area of ​​the bearing housing exposed within the backplate clearance. Ventilation opening flow area; Therefore, insulation optimization methods cannot significantly change the situation. , Naturally, it becomes the dominant heat source for heat transfer to the bearing housing.

[0040] In step S3, the convective heat transfer at the interface of the insulation cavity 403 satisfies the time-averaged energy boundary coupling: , Total convective heat flow: , , In the formula, The projected area of ​​the bearing housing covered by the heat shield. The average convective heat transfer coefficient of the insulation cavity is given. The average gas temperature inside the insulation cavity. The turbulent heat transfer coefficient of the insulation cavity is expressed using the dimensionless Nusselt correlation: , In the formula, The characteristic length of the cavity; The Reynolds number for the cavity airflow; For Grashof numbers; , , In the formula, The average air intake velocity of the ventilation holes. The coefficient of thermal expansion of the gas is _____. Kinematic viscosity; Therefore, The airflow rate is directly controlled by the ventilation holes and is the core variable for regulation. The above analysis treats the insulation cavity as a controlled volume rather than a closed cavity. If it is treated as a closed cavity, under alternating hot and cold conditions, the gas inside the insulation cavity will exert alternating loads on the stamped thin-walled thermal shield, which can easily lead to thermal shield fatigue failure. Assuming the insulation cavity is considered a closed body, the cavity volume is... The internal gas satisfies the ideal gas law: , Turbine temperature cycling range under varying operating conditions Intracavitary pressure fluctuation amplitude: , Thin-walled stainless steel thermal shields experience membrane bending stress due to cyclic pressure differentials. The theory of thin-plate bending stress is as follows: , It is alternating compressive stress. Alternating thermal stress, The coefficient of thermal expansion of the material. The elastic modulus of the material. The inner diameter of the heat shield. For the thickness of the heat shield; Under the continuous action of this alternating stress, the Miner linear damage criterion is used to calculate the cumulative fatigue damage of the material: , In the formula, The number of alternations between hot and cold operating conditions. This represents the fatigue life corresponding to alternating stress. Inside the sealed insulated cavity, It can reach over 0.2 MPa, while the inner diameter of the heat shield Much larger than its thickness When specific geometric parameters are introduced, the pressure drops to the tens of megapascals, resulting in a thermal shield fatigue life far lower than the required life of the turbocharger. Therefore, in engineering practice, ventilation holes must be installed to achieve pressure equalization. The presence of a ventilation opening inevitably leads to the exchange of compressible flow mass. Since the ventilation opening is a thin-walled throttling orifice, the compressible subsonic throttling flow rate formula is used: , The throttling flow coefficient; This represents the total flow area of ​​the ventilation holes; Specific heat ratio of gases; Mass flow rate of high-temperature gas flowing into the cavity Directly increases the average temperature of the cavity gas At the same time, increase the cavity ; Simultaneous formula: , Therefore, the requirements for structural reliability necessitate the inclusion of ventilation holes in the heat shield to balance the pressure inside and outside the insulation chamber. This increases mass transfer within the insulation chamber and turbine flow channels, thereby promoting… The addition of heat shields weakens their thermal protection effectiveness to some extent.

[0041] In step S4, the area of ​​the turbine backplate not covered by the heat shield directly contacts the high-temperature mainstream combustion gas, and the convective heat transfer flux is: , Total heat flow along this path: , The formula requires calculation of the heat transfer coefficient under rotating conditions. , Define the Reynolds number for the rotating state: , In the formula, The angular velocity of the turbine rotor. The radius of the pivot is 1. For high-temperature gas density, The dynamic viscosity of the gas; Backplate clearance gas Prandtl number: , In the formula, For the specific heat capacity of the gas at constant pressure, is the thermal conductivity of the gas.

[0042] Nusselt number correlation for rotating disk convection: , In the formula, , Therefore, we can obtain: .

[0043] In step S5, the heat carried by the lubricating oil is divided into two parts: , In the formula, The heat is carried away from the bearing clearance. This refers to the heat carried away from the oil supply line.

[0044] Regarding heat transfer within the bearing clearance: The shear Reynolds number is , In the formula, For the density of lubricating oil, The viscosity of the lubricating oil. The linear velocity of the floating bearing. This refers to the clearance of the outer oil film in the floating bearing. Lubricating oil Prandtl number , The specific heat capacity of lubricating oil at constant pressure. This refers to the thermal conductivity of the lubricating oil.

[0045] Due to the external oil film state Much less than 1000, belonging to the laminar flow range, narrow-gap rotating shear laminar flow, classic Couette flow heat transfer correlation: , The heat transfer coefficient of the bearing housing at the position of the outer oil film is: , Because the oil supply management is relatively coarse and the flow rate is slow, the heat transfer coefficient of the lubricating oil in the oil supply pipeline is treated as a constant: , In the formula, The average temperature of the oil supply pipe wall. The oil inlet temperature, This refers to the surface area of ​​the inner wall of the oil supply pipe.

[0046] In step S6, After establishing the bearing housing heat transfer model, assuming that the combustion gas satisfies the free vortex equal circulation condition in the vortex flow channel, the heat transfer calculation models of the vortex and turbine for the bearing housing heat input need to be established simultaneously: , In the formula, The radial radius of the flow channel. The tangential velocity of the gas. The cross-sectional area of ​​the vortex casing flow channel increases linearly along the circumference to ensure uniform and stable gas supply to the turbine inlet. The work done by turbine expansion follows Euler's equations: , In the formula, 1 represents the turbine inlet linear velocity, 1 represents the turbine inlet, and 2 represents the turbine outlet.

[0047] The one-dimensional conservation equations for compressible gas flow, including the continuity equation: , In the formula, This refers to the mass flow rate of the gas. For Zhou Xiang The cross-sectional area of ​​the vortex shell during flow; This refers to the speed of the meridional airflow. The density of the gas.

[0048] Momentum equation: , In the formula, For the centrifugal inertial force term of the free vortex, Energy equation: , In the formula, Total enthalpy of the gas; For static enthalpy, Ideal gas law: , Static temperature of combustion gas at various locations on the inner wall of the vortex casing airflow speed , as input to the inner wall convection heat transfer model.

[0049] in, , The above model establishes the fluid heat transfer boundary for heat transfer in the vortex shell. Next, a one-dimensional forced convection heat transfer model along the vortex shell flow channel is established to solve for the heat flux density on the inner wall. The total heat absorbed by the vortex shell is obtained. : vortex shell Local gas Reynolds number on cross section: , In the formula, The hydraulic diameter of the vortex flow channel; For high-temperature gas dynamic viscosity, with Interpolation, Gas Prandtl number: , In the formula, The thermal conductivity of the exhaust gas; Nusselt number correlation for forced convection in annular channels: , Suitable for turbulent flow with Re > 2300, and matched for high-flow, high-temperature exhaust gas conditions in natural gas engines; Forced convection heat transfer coefficient of annular channel: , One-dimensional convective heat transfer distribution of fluid in a vortex shell towards the vortex shell wall: , The solid temperature inside the vortex shell. Total heat exchange of gas on the inner wall of the vortex shell: The vortex shell is discretized circumferentially into N one-dimensional infinitesimal segments, where b is the axial width of the vortex shell and the heat transfer area of ​​each infinitesimal segment. : , That is, the heat transferred from the turbine flow channel to the volute casing.

[0050] In step S7, On one hand, the heat is transferred to the bearing body through path 1; on the other hand, some of it is lost through heat dissipation from the outer wall of the volute shell towards the environment. This portion of the lost heat is denoted as: , This formula serves as the convergence criterion for the heat conduction iteration of the vortex shell, achieving the coupling of heat transfer on the gas side, internal heat conduction of the shell, and external heat dissipation of the shell. The vortex shell dissipates heat to the environment. Assuming the vortex shell is a rotating shell, along the circumferential direction... Divide into one-dimensional infinitesimal elements, and the outer surface area of ​​the infinitesimal element is: , The axial width of the vortex shell; The outer diameter of the vortex shell at the micro-element; ambient air temperature. The engine compartment temperature is typically between 25-80℃. Calculation of dimensionless number of natural convection on external surface, Grashof number: , It is the acceleration due to gravity; The coefficient of thermal expansion of air; This refers to the temperature of the outer wall of the vortex shell. The characteristic length of the outer surface of the vortex shell; The viscosity is the kinematic viscosity of air.

[0051] Natural convection Nusselt number correlation: , In the formula, Take 0.703, applicable scope; The local heat transfer coefficient for dissipating heat to the environment is: , In the formula, The thermal conductivity of air. The total heat lost to the environment is: .

[0052] In step S8, the above turbocharger turbine-bearing housing heat transfer reduced-order prediction model is used in... Iteration between nodes, convergence criterion: , In the formula, , Based on the results of the above order reduction prediction method, it is known that: reducing the temperature inside the insulation cavity... That is, reducing the mass transfer between the heat insulation cavity and the turbine flow channel is to reduce This further reduces An important way; by increasing , thereby increasing Thus reducing This reduces the main heat transfer path of the bearing housing. An important way of transferring heat.

[0053] It also includes compressor 1, bearing housing 2, and turbine 3; A heat shield 405 is fixed between the bearing body 2 and the turbine back of the turbine 3, and a heat insulation cavity 403 is formed between the heat shield 405 and the bearing body 2. The inner edge 401 of the heat shield 405 is fitted and sealed with the mating surface 201 of the bearing body; this is used to prevent high-temperature gas from entering the heat insulation cavity 403 from the back plate gap 402. The heat shield 405 is provided with ventilation holes 404 to balance the pressure inside and outside the heat insulation cavity 403; The turbine 3 includes a finned heat dissipation volute 302, and heat dissipation fins 304 are evenly distributed circumferentially on the outer wall 303 of the volute of the finned heat dissipation volute 302.

[0054] The heat shield 405 can be a double-layer heat shield 405, including a main heat shield 406 and an auxiliary heat shield 407. An auxiliary heat shield 407 is added between the main heat shield 406 and the bearing body 2. The auxiliary heat shield 407 divides the heat insulation cavity 403 into two sealed chambers, inner and outer. The thickness of the main heat shield 406 and the auxiliary heat shield 405 is (0.8-1.5) mm. The heat transfer between the inside and outside of the cavity is segmented, which weakens the convective heat transfer driven by buoyancy from the source. The fluid in the cavity is almost stagnant, and a convective heat transfer path is realized for the heat insulation cavity 403. The blockade.

[0055] There can be two ventilation holes 404, and the two ventilation holes 404 are located in different phases. Located at the position of minimum pressure difference in the pressure field of vortex flow channel 301 The pressure difference between the two ventilation holes 404 is close to zero, avoiding forced intake driven by pressure difference.

[0056] The working principle of this device is as follows: The reduced-order model boasts high computational efficiency and supports rapid iteration across multiple parameters. The bearing housing heat transfer reduced-order prediction model constructed in this invention fully preserves the heat conduction of the flange solid. Convection in insulation cavity Exposed back plate for heat exchange Based on the physical mechanisms of the three core heat transfer pathways, lubricating oil cooling, and natural heat dissipation from the outer wall of the volute, the solution degree of freedom is greatly reduced, compressing the calculation time of traditional three-dimensional CFD simulation from several hours to seconds. It can support rapid optimization design of multiple variables such as heat shield ventilation hole size, flange thickness, and cooling oil flow rate, significantly shortening the R&D cycle. The prediction accuracy is high, with a maximum deviation of 27K and a relative error of <5%. The reduced-order prediction model was used to calculate the turbochargers of multiple product platforms and compared with the experimental results. Under the respective turbine design conditions, the trend of the prediction results and the experimental results is basically consistent with the change of turbine inlet temperature. The maximum prediction deviation is 27K and the relative deviation is less than 5%, which meets the engineering prediction accuracy requirements. The calculation results can be used for rapid structural iteration and guide the design of efficient thermal management structures for turbochargers. For the first time, this invention quantifies the proportion of heat transfer paths in a turbocharger bearing housing by breaking down the heat contribution of the three heat transfer paths into quantitative values, thus clarifying the direction for optimization. Solid thermal conductivity accounts for 72%-82% of the total heat load, making it the core heat source; Convection within the insulation cavity accounts for 16%-25%, making it a secondary control target. The fact that the heat exchange rate in the exposed area of ​​the bearing housing is less than 3% clarifies that a single insulation structure can only optimize convection and cannot fundamentally solve the underlying contradiction of the dominant heat conduction load. For the first time, a passive thermal management design logic of "cooling on the heat source side and flow control in the insulation cavity" is established, providing clear theoretical guidance for the design of thermal protection structures. The inner edge sealing design significantly improves the heat insulation effect. The inner edge sealed heat shield proposed in this invention blocks the intrusion of high-temperature combustion gas along the inner edge gap into the heat insulation cavity, limiting the heat exchange mode in the heat insulation cavity to inefficient natural convection. Experimental data shows that under high-temperature conditions with a vortex inlet temperature of 1023K, the fully sealed configuration achieves a maximum temperature difference of 23.1K at the sealing ring and 7.8K at the floating bearing compared to the fully open configuration. The thermal protection advantage of the sealed inner edge is significant in all operating conditions. The differential pressure balanced ventilation hole takes into account both reliability and thermal insulation performance. This invention reveals the contradiction between "pressure relief and thermal insulation" in ventilation hole design and proposes the design principles of differential pressure balanced ventilation hole: single hole scheme is preferred to obtain robust thermal protection that is not sensitive to changes in operating conditions; double hole scheme must arrange the ventilation holes at a position where the circumferential pressure is approximately equal to avoid pressure difference driving high temperature gas to flow into the inner cavity. The double-layer thermal shield effectively suppresses natural convection. The double-layer thermal shield structure proposed in this invention splits a single insulated cavity into two sealed chambers, inner and outer, by adding an auxiliary thermal shield. This segmentally compresses the temperature difference between the inside and outside of the chamber, weakening convective heat transfer driven by buoyancy at its source. The temperature difference in the outer chamber is approximately 100K, and in the inner chamber, approximately 200K. The buoyancy driving force is significantly reduced, and the fluid inside the chamber almost stagnates, effectively suppressing the convective heat transfer pathways. The heat exchange rate decreased significantly; The optimized coupling scheme achieves significant cooling effects. This invention combines a double-layer thermal shield with a finned heat dissipation turbine housing, resulting in a comprehensive and optimal solution that significantly reduces the temperature at three key monitoring points of the bearing housing: an average temperature reduction of 10.5K for the turbine end sealing ring, an average temperature reduction of 4.7K for the floating bearing, and a reduction of 0.22kW in the total heat transferred to the bearing housing from the entire engine. This scheme demonstrates stable cooling performance within a wide turbine inlet temperature range of 873K-1023K. It is estimated that, under the same bearing housing heat load, the turbine inlet exhaust temperature can be increased from 750℃ to 855℃, which can basically meet the requirements of conventional automotive gas engines. With low cost and easy engineering promotion, the thermal protection structure of this invention is a passive optimized design, which does not require additional cooling system or expensive special alloy materials. High-efficiency thermal protection can be achieved through structural optimization with limited cost increase. It is easy to mass-produce and promote in engineering, and has significant economic benefits and market application value.

[0057] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, and not to limit them. Although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some or all of the technical features therein. Such modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the scope of the technical solutions of the embodiments of the present invention, and they should all be covered within the scope of the claims and specification of the present invention.

Claims

1. A method for predicting the heat transfer order reduction of a natural gas engine turbocharger bearing housing, characterized in that: The natural gas engine turbocharger bearing housing (2) includes a thermal protection structure (4), which includes a heat shield (405), a heat insulation cavity (403), and a ventilation hole (404). The heat transfer order reduction prediction method is based on the thermal protection structure (4), and the heat transfer of the turbine bearing coupling system is decomposed into three independent heat transfer paths. The three independent heat transfer paths include: the solid conduction path of the turbine bearing housing flange, the gas convection heat transfer path of the heat shield cavity, and the convection heat transfer path of the exposed area of ​​the unshielded bearing housing. The order reduction calculation model is constructed, including the following steps: Step S1: Establish the heat balance equation and the total heat superposition formula; perform volume integral on all solid control volumes of the bearing housing, and combine the Gaussian divergence theorem to decompose the flange heat conduction boundary, the heat shield cavity convection boundary, the bearing housing exposed area convection boundary, and the lubricating oil cooling boundary under steady-state heat balance conditions to obtain the total heat relationship of the superposition of heat flow of the three heat transfer paths, and define the ratio of heat flow of each path to the total input heat as the heat proportion coefficient; Step S2: Solid conduction path of the volute bearing flange; the volute flange, gasket, heat shield ring, and bearing seat flange are regarded as a multi-layer series solid heat conduction structure. Based on Fourier's law of thermal conductivity, combined with the thermal conductivity, geometric thickness and interlayer contact thermal resistance of each layer material, the total conduction thermal resistance of the multi-layer series structure is calculated, and then the solid heat conduction of the flange to the bearing body (2) is obtained; the interlayer contact thermal resistance is determined by the bolt preload and the surface roughness of the mating surface; Step S3: Gas convection heat transfer path in the heat shield cavity; the heat insulation cavity (403) formed by the turbine (3) and the bearing body (2) is taken as an independent control body. Based on the characteristic size of the heat insulation cavity (403), the temperature difference of the cavity wall and the gas physical parameters, the Rayleigh number is calculated. The convection heat transfer coefficient of the heat insulation cavity (403) is solved by the natural convection Nusselt number correlation to obtain the convection heat transfer from the heat insulation cavity (403) to the bearing body (2); at the same time, the compressibility of the ventilation hole (404) is calculated. The subsonic throttling flow rate is used to add the heat brought into the insulation cavity (403) by the air leakage through the ventilation hole (404) to the convective heat transfer; the alternating bending stress of the thin-walled heat shield (405) under the alternating hot and cold conditions of the closed insulation cavity (403) is calculated, and the fatigue damage of the heat shield (405) is calculated by using the fatigue damage accumulation criterion, so as to determine the minimum flow area of ​​the ventilation hole (404) used for pressure equalization, taking into account both the fatigue reliability and heat insulation performance of the heat shield (405); Step S4: Convective heat transfer path of the exposed area of ​​the unshielded bearing body; Calculate the rotating Reynolds number and Prandtl number based on the turbine rotor angular velocity, characteristic radius and gas properties, solve the convective heat transfer coefficient of the exposed wall using the rotating disk Nusselt number correlation, and calculate the convective heat transfer brought by the high temperature gas in the back disk gap (402) of the exposed wall of the bearing body (2); Step S5: Lubricating oil cooling and heat transfer calculation; the total heat carried away by the lubricating oil is divided into two parts: the heat carried away by the bearing clearance oil film and the heat carried away by the oil supply pipeline; the bearing clearance shear Reynolds number is calculated, and when the shear Reynolds number is less than the preset laminar flow judgment threshold, the oil film is determined to be Cueter laminar flow, and the shear laminar flow heat transfer correlation formula is used to solve the bearing clearance oil film heat transfer coefficient; the lubricating oil heat transfer coefficient in the oil supply pipeline is taken as a fixed value, and the heat transfer of the two parts of lubricating oil is calculated separately and summed; Step S6: Calculation of heat transfer in the vortex shell; Based on the free vortex isocirculation condition, the Euler equation of the turbine and the one-dimensional compressible gas flow conservation equations, solve the gas temperature distribution and airflow velocity distribution in the vortex shell flow channel (301); establish a one-dimensional forced convection heat transfer model of the inner wall of the vortex shell and solve the total heat transfer from the gas to the vortex shell shell. Step S7: Calculation of natural convection heat dissipation from the vortex shell to the environment; Discretize the vortex shell into several one-dimensional micro-element segments along the circumference, calculate the Grashof number and Nussel number of natural convection on the outer wall of each micro-element segment, solve the heat dissipation of each micro-element segment to the outside and sum them to obtain the total heat dissipation from the outer wall (303) of the vortex shell to the environment, and use the heat dissipation result as the convergence criterion for the shell heat conduction iteration. Step S8: Iterative solution of the model; perform iterative calculations on all temperature nodes of the reduced-order model, and use the maximum relative deviation between two adjacent iterations of the temperature of each node as the iterative residual. Stop the iteration when the iterative residual meets the preset convergence criterion, and output the bearing body temperature field, the heat flow of the three heat transfer paths, and the heat contribution ratio of each path.

2. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S1, the formula for total heat accumulation is as follows: ， Therefore, the heat transfer in the bearing housing can be mainly divided into three heat transfer paths, namely: Path: Solid heat conduction path of the volute bearing flange; Path: Gas convection heat transfer path in the heat shield cavity; Path: Convection heat transfer path in the exposed area of ​​the unshielded bearing housing; In the formula, The lubricating oil carries away the total heat. The total heat transferred to the bearing housing, Thermal conductivity of solid materials such as bearing housings, flanges, and heat shields, expressed in W / (m⋅K); The thermodynamic temperature at any location within the solid domain, in K; The proportions of the three heat transfer paths are as follows: To identify the contribution of the three heat transfer paths, each path has a corresponding percentage coefficient: ， 。 3. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S2, the single-layer steady-state Fourier heat flux is: ， In the formula, For the thickness of a single layer of material, For interlayer temperature difference, The thermal conductivity of the material; ， In the formula, and The constant is determined by the surface roughness of the material; therefore, when the bolt preload is fixed, It is a constant; the effective contact area of ​​the flange Total heat transfer flow: ， In the formula, The steady-state wall temperature at the stop of the vortex shell. This refers to the steady-state wall temperature of the bearing housing stop. Take the partial derivative with respect to the structural design variables in the formula: ， Flange dimensions, Bolt preload, The thickness of the parts is a constraint parameter of the overall machine structure and assembly process. Passive thermal insulation structures on the bearing side, such as heat shields, ventilation holes, and double-layer insulation, only act on the thermal boundary of the insulation cavity (403) and do not change the flange series thermal resistance system, i.e.: ， For the area covered by the heat shield, This refers to the area of ​​the bearing housing exposed within the backplate clearance. Ventilation opening flow area; Therefore, insulation optimization methods cannot significantly change the situation. , Naturally, it becomes the dominant heat source for heat transfer to the bearing housing.

4. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S3, the convective heat transfer at the interface of the insulation cavity (403) satisfies the time-averaged energy boundary coupling: ， Total convective heat flow: ， ， In the formula, The projected area of ​​the bearing housing covered by the heat shield. The average convective heat transfer coefficient of the insulation cavity is given. The average gas temperature inside the insulation cavity. The turbulent heat transfer coefficient of the insulation cavity is expressed using the dimensionless Nusselt correlation: ， In the formula, The characteristic length of the cavity; The Reynolds number for the cavity airflow; For Grashof numbers; ， ， In the formula, The average air intake velocity of the ventilation holes. The coefficient of thermal expansion of the gas is _____. Kinematic viscosity; Therefore, The airflow rate is directly controlled by the ventilation holes and is the core variable for regulation. The above analysis treats the insulation cavity as a controlled volume rather than a closed cavity. If it is treated as a closed cavity, under alternating hot and cold conditions, the gas inside the insulation cavity will exert alternating loads on the stamped thin-walled thermal shield, which can easily lead to thermal shield fatigue failure. Assuming the insulation cavity is considered a closed body, the cavity volume is... The internal gas satisfies the ideal gas law: ， Turbine temperature cycling range under varying operating conditions Intracavitary pressure fluctuation amplitude: ， Thin-walled stainless steel thermal shields experience membrane bending stress due to cyclic pressure differentials. The theory of thin-plate bending stress is as follows: ， It is alternating compressive stress. Alternating thermal stress, The coefficient of thermal expansion of the material. The elastic modulus of the material. The inner diameter of the heat shield, For the thickness of the heat shield; Under the continuous action of this alternating stress, the Miner linear damage criterion is used to calculate the cumulative fatigue damage of the material: ， In the formula, The number of alternations between hot and cold operating conditions. This represents the fatigue life corresponding to alternating stress. Inside the sealed insulated cavity, It can reach over 0.2 MPa, while the inner diameter of the heat shield... Much larger than its thickness When specific geometric parameters are introduced, the pressure drops to the tens of megapascals, resulting in a thermal shield fatigue life far lower than the required life of the turbocharger. Therefore, in engineering practice, ventilation holes must be installed to achieve pressure equalization. The presence of a ventilation opening inevitably leads to the exchange of compressible flow mass. Since the ventilation opening is a thin-walled throttling orifice, the compressible subsonic throttling flow rate formula is used: ， The throttling flow coefficient; This represents the total flow area of ​​the ventilation holes; Specific heat ratio of gases; Mass flow rate of high-temperature gas flowing into the cavity Directly increases the average temperature of the cavity gas At the same time, increase the cavity ; Simultaneous formula: ， Therefore, the requirements for structural reliability necessitate the inclusion of ventilation holes in the heat shield to balance the pressure inside and outside the insulation chamber. This increases mass transfer within the insulation chamber and turbine flow channels, thereby promoting… The addition of heat shields weakens their thermal protection effectiveness to some extent.

5. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S4, the area of ​​the turbine backplate not covered by the heat shield directly contacts the high-temperature mainstream combustion gas, and the convective heat transfer flux is: ， Total heat flow along this path: ， The formula requires calculation of the heat transfer coefficient under rotating conditions. , Define the Reynolds number for the rotating state: ， In the formula, The angular velocity of the turbine rotor. The radius of the pivot is 1. For high-temperature gas density, The dynamic viscosity of the gas; Backplate clearance gas Prandtl number: ， In the formula, For the specific heat capacity of the gas at constant pressure, The thermal conductivity of the gas; Nusselt number correlation for rotating disk convection: ， In the formula, , Therefore, we can obtain: .

6. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S5, the heat carried by the lubricating oil is divided into two parts: ， In the formula, The heat is carried away from the bearing clearance. This refers to the heat carried away from the oil supply line; Regarding heat transfer within the bearing clearance: The shear Reynolds number is , In the formula, For the density of lubricating oil, The viscosity of the lubricating oil. The linear velocity of the floating bearing. This refers to the clearance of the outer oil film in the floating bearing. Lubricating oil Prandtl number , The specific heat capacity of lubricating oil at constant pressure. The thermal conductivity of the lubricating oil; Due to the external oil film state Much less than 1000, belonging to the laminar flow range, narrow-gap rotating shear laminar flow, classic Couette flow heat transfer correlation: ， The heat transfer coefficient of the bearing housing at the position of the outer oil film is: ， Because the oil supply management is relatively coarse and the flow rate is slow, the heat transfer coefficient of the lubricating oil in the oil supply pipeline is treated as a constant: ， In the formula, The average temperature of the oil supply pipe wall. The oil inlet temperature, This refers to the surface area of ​​the inner wall of the oil supply pipe.

7. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S6 After establishing the bearing housing heat transfer model, assuming that the combustion gas satisfies the free vortex equal circulation condition in the vortex flow channel, the heat transfer calculation models of the vortex and turbine for the bearing housing heat input need to be established simultaneously: ， In the formula, The radial radius of the flow channel. The tangential velocity of the gas. The cross-sectional area of ​​the vortex casing flow channel increases linearly along the circumference to ensure uniform and stable gas supply to the turbine inlet. The work done by turbine expansion follows Euler's equations: ， In the formula, 1 represents the turbine inlet linear velocity, 1 represents the turbine inlet, and 2 represents the turbine outlet. The one-dimensional conservation equations for compressible gas flow, including the continuity equation: ， In the formula, This refers to the mass flow rate of the gas. For Zhou Xiang The cross-sectional area of ​​the vortex shell during flow; This refers to the speed of the meridional airflow. The density of the gas; Momentum equation: ， In the formula, For the centrifugal inertial force term of the free vortex, Energy equation: ， In the formula, Total enthalpy of the gas; For static enthalpy, Ideal gas law: ， Static temperature of combustion gas at various locations on the inner wall of the vortex casing airflow speed , as input to the inner wall convection heat transfer model; in, , The above model establishes the fluid heat transfer boundary for vortex heat transfer. Next, a one-dimensional forced convection heat transfer model along the vortex flow channel is established to solve for the heat flux density of the inner wall. The total heat absorbed by the vortex shell is obtained. : vortex shell Local gas Reynolds number on cross section: ， In the formula, The hydraulic diameter of the vortex flow channel; For high-temperature gas dynamic viscosity, with Interpolation, Gas Prandtl number: ， In the formula, The thermal conductivity of the exhaust gas; Nusselt number correlation for forced convection in annular channels: ， Suitable for turbulent flow with Re > 2300, and matched for high-flow, high-temperature exhaust gas conditions in natural gas engines; Forced convection heat transfer coefficient of annular channel: ， One-dimensional convective heat transfer distribution of fluid in a vortex shell towards the vortex shell wall: ， The solid temperature inside the vortex shell. Total heat exchange of gas on the inner wall of the vortex shell: The vortex shell is discretized circumferentially into N one-dimensional infinitesimal segments, where b is the axial width of the vortex shell and the heat transfer area of ​​each infinitesimal segment. : ， That is, the heat transferred from the turbine flow channel to the volute casing.

8. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S7 On one hand, the heat is transferred to the bearing body through path 1; on the other hand, some of it is lost through heat dissipation from the outer wall of the volute shell towards the environment. This portion of the lost heat is denoted as: ， This formula serves as the convergence criterion for the heat conduction iteration of the vortex shell, achieving the coupling of heat transfer on the gas side, internal heat conduction of the shell, and external heat dissipation of the shell. The vortex shell dissipates heat to the environment. Assuming the vortex shell is a rotating shell, along the circumferential direction... Divide into one-dimensional infinitesimal elements, and the outer surface area of ​​the infinitesimal element is: ， The axial width of the vortex shell; The outer diameter of the vortex shell at the micro-element; ambient air temperature. The engine compartment temperature is typically between 25-80℃. Calculation of dimensionless number of natural convection on external surface, Grashof number: ， It is the acceleration due to gravity; The coefficient of thermal expansion of air; This refers to the temperature of the outer wall of the vortex shell. The characteristic length of the outer surface of the vortex shell; The viscosity of air in motion; Natural convection Nusselt number correlation: ， In the formula, Take 0.703, applicable scope; The local heat transfer coefficient for dissipating heat to the environment is: ， In the formula, The thermal conductivity of air. The total heat lost to the environment is: 。 9. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: In step S8, the above turbocharger turbine-bearing housing heat transfer order reduction prediction model is in Iteration between nodes, convergence criterion: ， In the formula, , Based on the results of the above order reduction prediction method, it is known that: reducing the temperature inside the insulation cavity... That is, reducing the mass transfer between the heat insulation cavity and the turbine flow channel is to reduce This further reduces An important way; by increasing , thereby increasing Thus reducing This reduces the main heat transfer path of the bearing housing. An important way of transferring heat.

10. The method for predicting heat transfer order reduction of a natural gas engine turbocharger bearing housing according to claim 1, characterized in that: It also includes compressors, bearing housings, and turbines; The heat shield is fixed between the bearing body and the turbine back of the turbine, and forms the heat insulation cavity between the heat shield and the bearing body. The inner edge of the heat shield is fitted and sealed with the mating surface of the bearing body to prevent high-temperature gas from entering the heat insulation cavity from the back plate gap. The ventilation holes are located on the heat shield. The turbine includes a finned heat dissipation vortex housing, and heat dissipation fins are evenly distributed circumferentially on the outer wall of the vortex housing.

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