A prediction method for gas turbine bearing scuffing failure considering true surface roughness

Abstract

The present invention provides a method for predicting the gluing failure of a gas turbine bearing taking into account the real surface roughness. The method considers the typical working conditions and geometric characteristics of the gas turbine bearing, couples a three-dimensional mixed lubrication analysis model and a second-kind Volterra integral equation, and establishes a method for predicting the gluing failure of a gas turbine bearing pair. The method has good practicability and can realize the prediction of the transient temperature rise and friction coefficient of the bearing pair under any processing technology and extreme working conditions. Among them, the method for solving the mixed lubrication equation taking into account the comprehensive effect of multiple factors takes into account the real surface roughness, the non-Newtonian characteristics of the lubricating oil, the specific structure and service conditions of the ship bearing, and can realize the prediction of the entire lubrication state from full film lubrication, mixed lubrication, boundary lubrication to dry contact, and can guide the optimal design of the bearing structure from the aspect of the degree of failure (wear and gluing, etc.).

Description

Technical Field

[0001] The invention belongs to the field of tribology of ship power plants, and in particular relates to a method for predicting the gluing failure of a combustion engine bearing taking into account the actual surface roughness. Background Art

[0002] Harsh working conditions such as high speed, heavy load and high temperature cause gas turbine ball bearings to be in a mixed lubrication state. Under mixed lubrication conditions, the rolling element and raceway of the gas turbine ball bearing will have local rough peak contact and relative slip, which will aggravate the friction phenomenon. Friction generates heat, affects the temperature distribution, and even causes high-temperature bonding failure of the bearing. The temperature will affect the shear modulus, react to the friction, and produce a friction-temperature mutual coupling effect. At present, there are still few studies on friction-flash temperature coupling under mixed lubrication conditions for ship gas turbine rolling bearings. Based on the Block bonding failure judgment criterion, the flash temperature change law under the lubrication-contact state between the bearing rolling element and the inner raceway can be used to predict the bonding failure position of the bearing pair, reveal the formation mechanism of bonding failure, and provide theoretical guidance for effectively preventing bonding failure of bearings during service. Summary of the invention

[0003] In view of this, the purpose of the present invention is to provide a method for predicting the bonding failure of gas turbine bearings taking into account the actual surface roughness, which can accurately and efficiently simulate the typical harsh working conditions of marine bearings such as complex load excitation, high-frequency variable working conditions and extremely long service time, and the resulting lubrication characteristics and friction flash temperature changes, thereby accurately predicting the bonding failure caused by excessive flash temperature at the bearing contact interface.

[0004] A method for predicting the abrasion failure of a combustion engine bearing taking into account the actual surface roughness comprises the following steps:

[0005] Step 1: Establish the geometric model:

[0006] Take the circumferential direction of the inner ring raceway as x 2j Axis, along the inner ring raceway radial direction z 2j Axis, perpendicular to the circumference and radial direction is y 2j The center of mass of the shaft and bearing is the origin of the coordinate system o 2j , establish a rectangular coordinate system; the x obtained after the rolling element completes the above geometric equivalence 2j o 2j z 2j Plane curvature radius With y 2j o 2j z 2j Plane curvature radius same:

[0007]

[0008] The inner raceway is analyzed along the x 2j o 2j z 2j Plane curvature radius With y 2j o 2j z 2j Plane curvature radius They are:

[0009]

[0010]

[0011] Where D W is the rolling element diameter, D m is the pitch diameter of the ball bearing, α 2j is the initial contact angle of the bearing, which is equal to the average diameter of the inner and outer raceways, i.e. D m =0.5(D r1 +D r2 ), f i is the curvature coefficient of the inner and outer channels;

[0012] Solve the rolling element and the inner raceway along x 2j With y 2j The equivalent radius of curvature R in the direction x2j With R y2j :

[0013]

[0014]

[0015] Step 2: Consider the Reynolds equation for instantaneous velocity:

[0016] Considering the entrainment velocity during the operation of the bearing pair, the following three-dimensional point contact mixed lubrication Reynolds equation is used to solve the pressure distribution:

[0017]

[0018] In the formula, t represents the time variable, p 2j (x 2j ,y 2j ) is the oil film pressure, h is the oil film thickness, η is the lubricating oil viscosity, ρ is the lubricating oil density, u x2j The rolling and inner raceway along x 2j The entrainment velocity in the axial direction;

[0019] Step 3: Establish the oil film thickness equation:

[0020] The specific form of the oil film thickness equation considering the actual surface roughness is as follows:

[0021]

[0022]

[0023] In the formula, v e (x 2j ,y 2j ,t) is the elastic deformation between the contact pairs, E' is the equivalent elastic modulus between the bearing pairs, δ 1 (x 2j ,y 2j ,t) and δ 2 (x 2j ,y 2j , t) are the actual roughness of the rolling element and inner raceway surface, h 0 (t) represents the normal approach between the rolling element and the inner raceway, Describes the equivalent ellipsoidal geometry of the rolling element and the inner raceway, R x2j With R y2j Through step 2, we can obtain that ξ and The computation node is at x 2j Axis and y 2j The coordinates of the axes, Ω represents the solution area;

[0024] Step 4: Establish the basic lubrication equation:

[0025] Consider the oil film viscosity and density as pressure related equations:

[0026]

[0027]

[0028] η 0 and ρ 0 are the ambient viscosity and ambient density respectively, α is the viscosity-pressure coefficient of the lubricating oil;

[0029] Step 5: Friction-flash temperature equation considering non-Newtonian fluid effects:

[0030] Bearing friction under mixed lubrication conditions is mainly composed of fluid shear friction and rough peak contact friction. The friction force in the fluid lubrication zone is calculated using the viscoelastic Bair-Winer non-Newtonian fluid rheological model:

[0031]

[0032] In the formula, τ L is the limiting shear stress, is the shear stress derivative, G ∞ is the limiting shear modulus, both of which depend on the rheological properties of the lubricant as a function of pressure and temperature.

[0033] The shear rate is The rolling element linear velocity is The inner raceway linear velocity is expressed as Substitute it into formula (12) to obtain the shear stress τ at any node in the calculation domain: fluid Nonlinear equation for the distribution:

[0034]

[0035] The shear stress at dry friction is solved as follows:

[0036] τ contact =f contact ·p x2j (x 2j ,y 2j ) (15)

[0037] f contact represents the dry friction coefficient;

[0038] Final bearing overall shear stress τ total for:

[0039] τ total =τ contact +τ fluid (16)

[0040] The friction force distribution of the bearing contact pair is obtained by integrating the shear stress in the calculation domain, and the friction coefficient between the rolling element and the inner raceway contact surface is obtained by the ratio of friction force to contact load:

[0041]

[0042] The generated heat is divided into two parts according to the set heat distribution coefficient A:

[0043]

[0044] At the same time, according to Plint's research results, the local velocity distribution and temperature change curve of lubrication under sliding conditions are calculated, the heat distribution coefficient between the rolling element and the inner raceway is determined, and the following is obtained:

[0045]

[0046] Where, T 1 , T 2 are the surface temperatures of the rolling element and inner raceway, K f Indicates the thermal conductivity of lubricating oil;

[0047] The flash temperature calculation of the rolling element and inner raceway interface is based on the above-mentioned fast moving heat source theory on a semi-infinite solid. Combining this theory, a rolling element-inner raceway surface temperature calculation model is established. Based on the above theory, the second kind of Volterra integral equation is expressed as:

[0048]

[0049] Where, T b1 , T b2 are the initial temperatures of the rolling element and inner raceway surface, C 1 and C 2 is the specific heat capacity of the solid, k 1 , k 2 is the heat transfer coefficient between the rolling element and the inner raceway, ρ 1 ,ρ 2 are the density of rolling element and inner raceway respectively. 2j The axis is divided into grids, λ represents any grid, T 1 (λ) and T 2 (λ) represents the temperature of the rolling element and the inner raceway surface at the grid λ; q(λ) represents the heat at the grid λ; λ∈(-x 2j ,ξ), ξ is the x corresponding to the final temperature rise 2j Axis coordinates;

[0050] The formula for judging bonding failure is as follows:

[0051] T b +T fm ≤T sc (twenty one)

[0052] Among them, T sc is the critical bonding temperature, and the contact interface temperature rise T fm =T 1 / T 2 ;

[0053] When judging whether the rolling element has failed due to bonding, T in formula (21) b Replaced by the initial rolling element temperature T 1b , according to this formula, determine whether it is failed; when judging whether the inner raceway has failed due to bonding, T in formula (21) b Replaced by the initial temperature of the inner raceway T 2b , and determine whether it is invalid according to this formula.

[0054] Preferable:

[0055] G ∞ =1.2p 2j (x 2j ,y 2j) / (2.52+0.024T 2 )-10 -9 (12)

[0056] τ L =0.25G ∞ (13)

[0057] Better: Dry friction coefficient f contact The value range is 0.07-0.15.

[0058] The present invention has the following beneficial effects:

[0059] The present invention provides a method for predicting the gluing failure of a gas turbine bearing taking into account the real surface roughness. The method considers the typical working conditions and geometric characteristics of the gas turbine bearing, couples a three-dimensional mixed lubrication analysis model and a second-kind Volterra integral equation, and establishes a method for predicting the gluing failure of a gas turbine bearing pair. The method has good practicability and can realize the prediction of the transient temperature rise and friction coefficient of the bearing pair under any processing technology and extreme working conditions. Among them, the method for solving the mixed lubrication equation taking into account the comprehensive effect of multiple factors takes into account the real surface roughness, the non-Newtonian characteristics of the lubricating oil, the specific structure and service conditions of the ship bearing, and can realize the prediction of the entire lubrication state from full film lubrication, mixed lubrication, boundary lubrication to dry contact, and can guide the optimal design of the bearing structure from the aspect of the degree of failure (wear and gluing, etc.). BRIEF DESCRIPTION OF THE DRAWINGS

[0060] Figure 1 It is an equivalent diagram of the bearing rolling element and inner raceway lubrication model adopted by the present invention;

[0061] Figure 2(a) is a three-dimensional inner raceway temperature rise diagram of the grinding surface, Figure 2(b) is a three-dimensional inner raceway temperature rise diagram of the honing surface, Figure 2(c) is a three-dimensional inner raceway temperature rise diagram of the shaved surface, and Figure 2(d) is a three-dimensional inner raceway temperature rise diagram of the polished surface;

[0062] Figure 3 It is a trend diagram of the friction coefficient and lubrication solution under the real surface roughness adopted by the present invention;

[0063] Figure 4 (a) and Figure 4 (b) are the maximum temperature rise diagrams of the rolling element surface and the maximum temperature rise diagram of the inner raceway surface under different load conditions;

[0064] Figure 5(a) and Figure 5(b) are the changing trend diagrams of oil film pressure and oil film thickness at different entrainment velocities, respectively. DETAILED DESCRIPTION

[0065] The present invention is described in detail below with reference to the accompanying drawings and embodiments.

[0066] The present invention provides a method for analyzing the interface lubrication-flash temperature coupling of a marine gas turbine bearing. The method is based on the actual measured surface roughness and the point contact mixed lubrication model, and takes into account the high-speed and heavy-load, large size, and three-dimensional morphological parameters of the surface roughness of the marine bearing pair. The second-class Volterra integral equation and the theory of fast-moving heat sources are used to obtain the transient temperature rise distribution between the contact pairs. The Blok transient temperature criterion is used as the bonding criterion. The interface temperature rise affects the friction-lubrication characteristics of the bearing pair by affecting the limiting shear modulus and limiting shear stress of the lubricating oil. The frictional heat is used to reversely affect the temperature between the contact pairs, so that friction and temperature rise affect each other. Finally, a bearing lubrication-flash temperature coupling analysis method is formed, which can also reveal the influence of working conditions, geometry, morphology texture and other parameters on the friction coefficient and flash temperature distribution, and provide theoretical guidance for the prediction of transient temperature rise of marine gas turbine bearing pairs and the design of bearing structures. The specific technical scheme is as follows:

[0067] Step 1: Create a geometric model

[0068] Figure 1 The established ball bearing inner ring-rolling element contact geometric model takes the contact between the rolling element and the inner raceway as an example, which can be converted into an ellipsoid and a semi-infinite plane contact model, with the inner ring raceway circumferential direction as x. 2j Axis, along the inner ring raceway radial direction z 2j Axis, perpendicular to the circumference and radial direction is y 2j The center of mass of the shaft and bearing is the origin of the coordinate system o 2j , establish a rectangular coordinate system; the x obtained after the rolling element completes the above geometric equivalence 2j o 2j z 2j Plane curvature radius With y 2j o 2j z 2j Plane curvature radius same:

[0069]

[0070] The inner raceway is analyzed along the x 2j o 2j z 2j Plane curvature radius With y 2j o 2j z 2j Plane curvature radius They are:

[0071]

[0072]

[0073] Where DW is the rolling element diameter, D m is the pitch diameter of the ball bearing, α 2j is the initial contact angle of the bearing, which is equal to the average diameter of the inner and outer raceways, i.e. D m =0.5(D r1 +D r2 ), f i is the curvature coefficient of the inner and outer channels.

[0074] Solve the rolling element and the inner raceway along x 2j With y 2j The equivalent radius of curvature R in the direction x2j With R y2j , so that the rolling element and the inner raceway are transformed into the contact between the ellipsoid and the semi-infinite plane as shown in Figure 2:

[0075]

[0076]

[0077] Step 2: Reynolds equation considering instantaneous velocity

[0078] Considering the entrainment velocity during the operation of the bearing pair, the following three-dimensional point contact mixed lubrication Reynolds equation is used to solve the pressure distribution:

[0079]

[0080] In the formula, t represents the time variable, p 2j (x 2j ,y 2j ) is the oil film pressure, h is the oil film thickness, η is the lubricating oil viscosity, ρ is the lubricating oil density, u x2j The rolling and inner raceway along x 2j The suction speed in the axial direction.

[0081] Step 3: Establish the oil film thickness equation

[0082] For bearing pairs, the transient real surface roughness is an important factor affecting the contact film thickness. The specific form of the oil film thickness equation considering the real surface roughness is as follows:

[0083]

[0084]

[0085] In the formula, v e (x 2j ,y 2j ,t) is the elastic deformation between the contact pairs, E' is the equivalent elastic modulus between the bearing pairs, δ 1 (x 2j ,y2j ,t) and δ 2 (x 2j ,y 2j , t) are the actual roughness of the rolling element and inner raceway surface, h 0 (t) represents the normal approach between the rolling element and the inner raceway, Describes the equivalent ellipsoidal geometry of the rolling element and the inner raceway, R x2j With R y2j Through step 2, we can obtain that ξ and The computing node is at x 2j Axis and y 2j The coordinates of the axis, Ω represents the solution area.

[0086] Step 4: Basic Lubrication Equations

[0087] When the pressure on the lubricating oil film changes, the intermolecular forces and distances between the oil film molecules change, and thus the viscosity and density of the oil film change accordingly. The calculation model of the present invention regards the viscosity and density of the oil film as pressure-related equations, and the specific solution is as follows:

[0088]

[0089]

[0090] η 0 and ρ 0 are the ambient viscosity and ambient density respectively, and α is the viscosity-pressure coefficient of the lubricating oil.

[0091] Step 5: Friction-flash temperature equation considering non-Newtonian fluid effects

[0092] Bearing friction under mixed lubrication conditions is mainly composed of fluid shear friction and rough peak contact friction. The friction force in the fluid lubrication zone is calculated using the viscoelastic Bair-Winer non-Newtonian fluid rheological model.

[0093]

[0094] In the formula, τ L is the limiting shear stress, is the shear stress derivative, G ∞ is the limiting shear modulus, both of which depend on the rheological properties of the lubricant oil as a function of pressure and temperature. The lubricant oil used in the calculation of the present invention is a typical mineral oil, so the Dyson empirical formula can be used to estimate G ∞ and τ L :

[0095] G ∞ =1.2p 2j (x 2j ,y 2j) / (2.52+0.024T 2 )-10 -9 (12)

[0096] τ L =0.25G ∞ (13)

[0097] Typical shear rates for mineral oils are The rolling element linear velocity is The inner raceway linear velocity is expressed as Substituting it into formula (12), we can obtain the shear stress τ at any node in the calculation domain: fluid Nonlinear equations for distributions.

[0098]

[0099] The friction coefficient generated by rough peak contact is relatively easy to measure. In engineering practice, the dry friction coefficient generally fluctuates within a smaller range. Usually, the dry friction coefficient f contact It floats within the range of 0.07-0.15. The material contact of the specific bearing can simulate the actual working conditions on the friction and wear testing machine to measure the dry friction coefficient. Based on this, the shear stress at the dry friction point is solved as follows:

[0100] τ contact =f contact ·p x2j (x 2j ,y 2j ) (15)

[0101] Final bearing overall shear stress τ total for:

[0102] τ total =τ contact +τ fluid (16)

[0103] The friction force distribution of the bearing contact pair is obtained by integrating the shear stress in the calculation domain, and the friction coefficient between the rolling element and the inner raceway contact surface is obtained by the ratio of friction force to contact load.

[0104]

[0105] Friction and heat are interdependent. The heat generated during the operation of ball bearings is mainly transferred through the flow of lubricating oil and heat conduction between the rolling elements and the inner raceway and the environment. Based on Francis' research results, the generated heat can be divided into two parts according to the heat distribution coefficient A:

[0106]

[0107] At the same time, according to Plint's research results, the local velocity distribution and temperature change curve of lubrication under sliding conditions are calculated, the heat distribution coefficient between the rolling element and the inner raceway is determined, and the following is obtained:

[0108]

[0109] Where, T 1 , T 2 are the surface temperatures of the rolling element and inner raceway, K f Indicates the thermal conductivity of lubricating oil.

[0110] The flash temperature calculation of the rolling element and inner raceway interface is based on the above-mentioned fast moving heat source theory on a semi-infinite solid. Combining this theory, a rolling element-inner raceway surface temperature calculation model is established. Based on the above theory, the second kind of Volterra integral equation is expressed as:

[0111]

[0112] Where, T b1 , T b2 are the initial temperatures of the rolling element and inner raceway surface, C 1 and C 2 is the specific heat capacity of the solid, k 1 , k 2 is the heat transfer coefficient between the rolling element and the inner raceway, ρ 1 ,ρ 2 are the density of the rolling element and the inner raceway respectively, q is the heat generated in the contact area between the rolling element and the inner raceway due to the friction of the rough peak or the shear effect of the lubricant. 2j The axis is divided into grids, λ represents any grid, T 1 (λ) and T 2 (λ) represents the temperature of the rolling element and the inner raceway surface at the grid λ; q(λ) represents the heat at the grid λ; λ∈(-x 2j ,ξ), ξ is the x corresponding to the final temperature rise 2j axis coordinate, h is the oil film thickness obtained through step 2.

[0113] According to the transient temperature criterion proposed by Blok, bonding is caused by the local transient temperature of the surface reaching a critical value. This is currently the only ISO-certified principle for judging bonding failure.

[0114] T b +T fm ≤T sc (twenty one)

[0115] Among them, T sc is the critical bonding temperature, and the contact interface temperature rise T fm =T1 / T 2 .

[0116] When judging whether the rolling element has failed due to bonding, T in formula (21) b Replaced by the initial rolling element temperature T 1b , according to this formula, determine whether it is failed; when judging whether the inner raceway has failed due to bonding, T in formula (21) b Replaced by the initial temperature of the inner raceway T 2b , determine whether it is invalid according to this formula.

[0117] Example:

[0118] This embodiment takes a certain type of gas turbine ball bearing as the research object. Table 1 gives the operating parameters of the bearing pair. Based on the three-dimensional point contact mixed lubrication model, a bearing lubrication-flash temperature prediction method is developed that comprehensively considers the influencing factors such as the flash temperature of the contact interface, elastic deformation and measured mechanical surface roughness. It can realize the mixed lubrication of the bearing pair during the service cycle and the prediction of the transient temperature rise state of the rolling element and the inner raceway, and provide theoretical guidance for the prediction of the bonding failure of the gas turbine bearing pair and the optimal design of the bearing structure.

[0119] Table 1 Bearing pair operating parameters

[0120]

[0121]

[0122] Figure 2 shows the influence of the real rough surface on the interface temperature rise of the gas turbine ball bearing under four different processing technologies. Under the same working conditions, the high temperature area and temperature rise peak of the inner raceway of the shaved rough surface are much larger than those of the other three rough surfaces. The maximum temperature rise of the inner raceway under the polished rough condition is 63.6℃, ​​which is lower than the flash point temperature rise. The maximum temperature rise of the inner raceway under the three processing methods of honing, grinding and shaving is 183%, 247% and 793% higher than that of the polished surface, respectively. In the bearing temperature rise analysis, the temperature rise of the inner raceway of the polished rough surface is far from the critical value of the interface bonding temperature, and it performs best in anti-bonding performance. Therefore, in order to avoid bonding failure in the service of gas turbine high-speed bearings, the bearing contact surface should be polished as much as possible.

[0123] Figure 3 A comparison chart of lubrication performance under different rough surfaces was drawn. By comparison, it can be seen that the ball bearing is in full film lubrication state in the case of polishing rough surface only, and the other three surfaces are in direct contact with rough peaks to varying degrees, resulting in an increase in friction coefficient, aggravated friction loss process, and dry friction heat generation cannot transfer heat energy to the external environment in time, which is also the main reason why the interface temperature rise of the other three contact surfaces is much greater than that of the polished surface.

[0124] Figures 4(a) and 4(b) show the effect of load size on the maximum temperature rise distribution of the rolling element and inner raceway of a gas turbine bearing at different instants. It can be seen from the figure that the maximum temperature rise of the inner raceway surface is smaller than the maximum temperature rise of the rolling element surface. As the load gradually increases, the maximum temperature rise of the interface changes significantly. When the contact load between the interfaces reaches 1.4 GPa, the maximum temperature rise of the interface increases by 103% compared to 0.9 GPa.

[0125] Figure 5(a) records the oil film thickness distribution at different entrainment speeds. In the figure, the oil film thickness of 0 indicates the dry friction area. It can be seen from the figure that when the entrainment speed is high, reaching 30m / s, the two contact surfaces are completely separated by the oil film. At this time, the interface is in a full film lubrication state. As the entrainment speed decreases, the oil film thickness gradually decreases. When the speed drops to 3m / s, a local dry friction area has appeared on the interface. As the entrainment speed continues to decrease, the dry friction area further expands. For the oil film pressure distribution in Figure 5(b), since the oil film thickness of part of the outlet area is 0 when the speed is 3m / s, the rough peak contact is caused, resulting in a higher pressure peak at 3m / s than at 30m / s in the outlet area.

[0126] In summary, the above are only preferred embodiments of the present invention and are not intended to limit the protection scope of the present invention. Any modifications, equivalent substitutions, improvements, etc. made within the spirit and principles of the present invention should be included in the protection scope of the present invention.

Claims

1. A method for predicting the abrasion failure of a combustion engine bearing taking into account the actual surface roughness, characterized in that: The steps include: Step 1: Establish the geometric model: Take the circumferential direction of the inner ring raceway as x 2j Axis, along the inner ring raceway radial direction z 2j Axis, perpendicular to the circumference and radial direction is y 2j The center of mass of the shaft and bearing is the origin of the coordinate system o 2j , establish a rectangular coordinate system; the x obtained after the rolling element completes the above geometric equivalence 2j o 2j z 2j Plane curvature radius With y 2j o 2j z 2j Plane curvature radius same: The inner raceway is analyzed along the x 2j o 2j z 2j Plane curvature radius With y 2j o 2j z 2j Plane curvature radius They are: Where D W is the rolling element diameter, D m is the pitch diameter of the ball bearing, α 2j is the initial contact angle of the bearing, which is equal to the average diameter of the inner and outer raceways, i.e. D m =0.5(D r1 +D r2 ), f i is the curvature coefficient of the inner and outer channels; Solve the rolling element and the inner raceway along x 2j With y 2j The equivalent radius of curvature R in the direction x2j With R y2j : Step 2: Consider the Reynolds equation for instantaneous velocity: Considering the entrainment velocity during the operation of the bearing pair, the following three-dimensional point contact mixed lubrication Reynolds equation is used to solve the pressure distribution: In the formula, t represents the time variable, p 2j (x 2j ,y 2j ) is the oil film pressure, h is the oil film thickness, η is the lubricating oil viscosity, ρ is the lubricating oil density, u x2j The rolling and inner raceway along x 2j The entrainment velocity in the axial direction; Step 3: Establish the oil film thickness equation: The specific form of the oil film thickness equation considering the actual surface roughness is as follows: In the formula, v e (x 2j ,y 2j ,t) is the elastic deformation between the contact pairs, E' is the equivalent elastic modulus between the bearing pairs, δ1(x 2j ,y 2j ,t) and δ2(x 2j ,y 2j ,t) are the actual roughness of the rolling element and the inner raceway surface, h0(t) represents the normal approximation between the rolling element and the inner raceway, Describes the equivalent ellipsoidal geometry of the rolling element and the inner raceway, R x2j With R y2j Through step 2, we can obtain that ξ and The computation node is at x 2j Axis and y 2j The coordinates of the axes, Ω represents the solution area; Step 4: Establish the basic lubrication equation: Consider the oil film viscosity and density as pressure related equations: η0 and ρ0 are the ambient viscosity and ambient density respectively, α is the viscosity-pressure coefficient of the lubricating oil; Step 5: Friction-flash temperature equation considering non-Newtonian fluid effects: Bearing friction under mixed lubrication conditions is mainly composed of fluid shear friction and rough peak contact friction. The friction force in the fluid lubrication zone is calculated using the viscoelastic Bair-Winer non-Newtonian fluid rheological model: In the formula, τ L is the limiting shear stress, is the shear stress derivative, G ∞ is the limiting shear modulus, both of which depend on the rheological properties of the lubricant as a function of pressure and temperature. The shear rate is The rolling element linear velocity is The inner raceway linear velocity is expressed as Substitute it into formula (12) to obtain the shear stress τ at any node in the calculation domain: fluid Nonlinear equation for the distribution: The shear stress at dry friction is solved as follows: t contact =f contact ·p x2j (x 2j ,y 2j ) (15) f contact represents the dry friction coefficient; Final bearing overall shear stress τ total for: t total =t contact +t fluid (16) The friction force distribution of the bearing contact pair is obtained by integrating the shear stress in the calculation domain, and the friction coefficient between the rolling element and the inner raceway contact surface is obtained by the ratio of friction force to contact load: The generated heat is divided into two parts according to the set heat distribution coefficient A: At the same time, according to Plint's research results, the local velocity distribution and temperature change curve of lubrication under sliding conditions are calculated, the heat distribution coefficient between the rolling element and the inner raceway is determined, and the following is obtained: Where, T1 and T2 are the surface temperatures of the rolling element and inner raceway respectively, K f Indicates the thermal conductivity of lubricating oil; The flash temperature calculation of the rolling element and inner raceway interface is based on the above-mentioned fast moving heat source theory on a semi-infinite solid. Combining this theory, a rolling element-inner raceway surface temperature calculation model is established. Based on the above theory, the second kind of Volterra integral equation is expressed as: Where, T b1 , T b2 are the initial temperatures of the rolling element and the inner raceway surface, C1 and C2 are the solid specific heat capacities, k1 and k2 are the heat transfer coefficients of the rolling element and the inner raceway, ρ1 and ρ2 are the densities of the rolling element and the inner raceway, and the rolling element and the inner raceway surface are respectively moved along the x 2j The axis is meshed, λ represents any one of the meshes, T1(λ) and T2(λ) represent the temperature of the rolling element and the inner raceway surface at the mesh λ respectively; q(λ) represents the heat at the mesh λ; λ∈(-x 2j ,ξ), ξ is the x corresponding to the final temperature rise 2j Axis coordinates; The formula for judging bonding failure is as follows: T b +T fm ≤T sc (21) Among them, T sc is the critical bonding temperature, and the contact interface temperature rise T fm =T1 / T2; When judging whether the rolling element has failed due to bonding, T in formula (21) b Replaced by the initial rolling element temperature T 1b , according to this formula, determine whether it is failed; when judging whether the inner raceway has failed due to bonding, T in formula (21) b Replaced by the initial temperature of the inner raceway T 2b , determine whether it is invalid according to this formula.

2. A method for predicting the scuffing failure of a combustion engine bearing taking into account the real surface roughness according to claim 1, characterized in that: G ∞ =1.2p 2j (x 2j ,y 2j ) / (2.52+0.024T2)-10 -9 (12) t L =0.25G ∞ (13)。 3. The method for predicting the scuffing failure of a combustion engine bearing taking into account the real surface roughness according to claim 1, characterized in that: Dry friction coefficient f contact The value range is 0.07-0.15.

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