Gear service life prediction method based on lubrication-vibration coupling model and related device

Abstract

The invention discloses a gear service life prediction method based on a lubrication-vibration coupling model and a related device, and relates to the technical field of gear service life prediction.The method comprises the steps that firstly, according to gear transmission input data, a three-dimensional mixed lubrication model containing a Reynolds equation coupling solving system is established, and meanwhile system excitation is calculated; then the oil film and the meshing rigidity are connected in parallel to form comprehensive rigidity, a bending-torsion coupling kinetic equation containing torsion and transverse freedom degrees is established, and a lubrication-vibration coupling model is constructed; finally, by means of coupling model output, the lubrication state is evaluated, the vibration characteristic is optimized, and the fatigue and abrasion life is predicted. According to the method, the three-dimensional mixed lubrication model and the system excitation calculation and coupling dynamics model are integrated, the lubrication and vibration coupling effect is considered, the limitation of independent analysis is overcome, the evaluation precision is improved, gear parameter optimization, lubrication scheme making and fault diagnosis can be guided, and the method has engineering application value.

Description

Technical Field

[0001] This application relates to the field of gear life prediction technology, and in particular to a gear life prediction method and related device based on a lubrication-vibration coupling model. Background Technology

[0002] The lubricating oil film acts on the gear meshing interface, not only reducing friction, wear, and providing cooling, but also directly influencing dynamic parameters such as tooth surface deformation and gear meshing stiffness. This further affects the vibration characteristics of the entire gear transmission system. The fluctuations caused by system vibration then feed back to the lubrication interface, affecting lubrication state parameters such as lubrication pressure and film thickness, generating noise and vibration. Therefore, establishing a three-dimensional lubrication-vibration coupled model of the gear transmission system, taking into account the influence of oil film condition and system vibration factors, is of great significance. Summary of the Invention

[0003] The purpose of this application is to provide a gear life prediction method and related device based on a lubrication-vibration coupling model, which can realize the synergistic analysis of lubrication and vibration and improve the accuracy of gear transmission system performance evaluation.

[0004] To achieve the above objectives, this application provides the following solution:

[0005] In a first aspect, this application provides a gear life prediction method based on a lubrication-vibration coupling model, comprising the following steps:

[0006] The input data for the gear transmission process is acquired; the input data includes geometric parameters, operating parameters, and material parameters.

[0007] A three-dimensional hybrid lubrication model for gears is established based on the input data. Numerical solutions are performed on the three-dimensional hybrid lubrication model to obtain the three-dimensional lubrication-related parameters of the gears. The three-dimensional lubrication-related parameters include the pressure distribution in the contact area, the oil film thickness, and the lubrication state. The three-dimensional hybrid lubrication model for gears is a coupled solution system based on the Reynolds equation, the film thickness equation, and the elastic deformation equation.

[0008] The excitation of the gear transmission system is calculated based on the input data; the excitation of the gear transmission system is used to determine the vibration excitation parameters of the gear; the vibration excitation parameters include time-varying meshing force, stiffness band, and dynamic load spectrum.

[0009] Based on the three-dimensional hybrid lubrication model of gears and the excitation of the gear transmission system, a lubrication and vibration coupling model is established. The lubrication and vibration coupling model is established by forming a comprehensive stiffness by connecting the oil film stiffness and gear meshing stiffness in parallel, and establishing a bending-torsional coupling dynamic equation that includes torsional and lateral degrees of freedom. By iteratively solving the lubrication and vibration coupling model, the coupling analysis of lubrication and vibration can be realized, and the dynamic oil film pressure, minimum film thickness and vibration displacement / velocity spectrum can be output.

[0010] By analyzing the output results of the lubrication and vibration coupling model, gear life prediction results are obtained; the analysis includes lubrication condition assessment, vibration characteristic optimization, fatigue life prediction, and wear life prediction.

[0011] Secondly, this application provides a gear life prediction system based on a lubrication-vibration coupling model, comprising:

[0012] The gear transmission data acquisition module is used to acquire input data during the gear transmission process; the input data includes geometric parameters, operating condition parameters, and material parameters.

[0013] The gear lubrication model construction module is used to build a three-dimensional hybrid lubrication model of gears based on input data. Numerical solutions are performed on the three-dimensional hybrid lubrication model of gears to obtain three-dimensional lubrication-related parameters of gears. The three-dimensional lubrication-related parameters include pressure distribution in the contact area, oil film thickness, and lubrication state. The three-dimensional hybrid lubrication model of gears is a coupled solution system based on the Reynolds equation, film thickness equation, and elastic deformation equation.

[0014] The gear transmission excitation calculation module is used to calculate the excitation of the gear transmission system based on the input data; the excitation of the gear transmission system is used to determine the vibration excitation parameters of the gear; the vibration excitation parameters include time-varying meshing force, stiffness band, and dynamic load spectrum.

[0015] The lubrication and vibration coupling module is used to establish a lubrication and vibration coupling model based on the three-dimensional hybrid lubrication model of gears and the excitation of the gear transmission system. The lubrication and vibration coupling model is established by forming a comprehensive stiffness by connecting the oil film stiffness and gear meshing stiffness in parallel, and establishing a bending-torsional coupling dynamic equation that includes torsional and lateral degrees of freedom. By iteratively solving the lubrication and vibration coupling model, the coupling analysis of lubrication and vibration can be realized, and the dynamic oil film pressure, minimum film thickness and vibration displacement / velocity spectrum can be output.

[0016] The gear life prediction module is used to obtain gear life prediction results by analyzing the output results of the lubrication and vibration coupling model; the analysis includes lubrication condition assessment, vibration characteristic optimization, fatigue life prediction, and wear life prediction.

[0017] Thirdly, this application provides a computer device, including: a memory, a processor, and a computer program stored in the memory and executable on the processor, wherein the processor executes the computer program to implement the steps of the gear life prediction method based on the lubrication-vibration coupling model described above.

[0018] Fourthly, this application provides a computer-readable storage medium having a computer program stored thereon, which, when executed by a processor, implements the steps of the gear life prediction method based on the lubrication-vibration coupling model described above.

[0019] Fifthly, this application provides a computer program product, including a computer program that, when executed by a processor, implements the steps of the gear life prediction method based on the lubrication-vibration coupling model described above.

[0020] According to the specific embodiments provided in this application, the following technical effects are disclosed:

[0021] This application provides a gear life prediction method and related apparatus based on a lubrication-vibration coupling model. The method includes: establishing a three-dimensional hybrid lubrication model of the gear based on input data from the gear transmission process, using a coupled solution system of Reynolds equations, film thickness equations, and elastic deformation equations; simultaneously calculating the excitation of the gear transmission system based on the input data; subsequently, based on the three-dimensional hybrid lubrication model and the excitation of the gear transmission system, forming a comprehensive stiffness by paralleling the oil film stiffness and gear meshing stiffness, and establishing a bending-torsional coupled dynamic equation including torsional and lateral degrees of freedom, thus establishing a lubrication-vibration coupling model; and using the output results of the lubrication-vibration coupling model, performing lubrication state assessment, vibration characteristic optimization, fatigue life prediction, and wear life prediction to obtain the gear life prediction result. This application integrates the three-dimensional hybrid lubrication model, system excitation calculation, and coupled dynamic model, considering the coupling effect of lubrication and vibration, overcoming the limitations of separate analysis in existing technologies, achieving synergistic analysis of lubrication and vibration, improving the accuracy of gear transmission system performance evaluation, and directly guiding gear parameter optimization (such as tooth profile modification), lubrication scheme formulation (viscosity selection), and fault diagnosis (vibration spectrum identification), demonstrating significant engineering application value. Attached Figure Description

[0022] To more clearly illustrate the technical solutions in the embodiments of this application or the prior art, the drawings used in the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of this application. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0023] Figure 1This is a flowchart illustrating a gear life prediction method based on a lubrication-vibration coupling model, provided as an embodiment of this application.

[0024] Figure 2 This is a schematic diagram of the equivalent solution domain and roughness mesh processing for gear line contact mixed lubrication in one embodiment of this application.

[0025] Figure 3 This is a force analysis diagram of the crank connecting rod in one embodiment of this application.

[0026] Figure 4 This is a diagram showing the structure and stress analysis of the gas distribution mechanism in one embodiment of this application.

[0027] Figure 5 This is a force analysis diagram of the oil supply cam in one embodiment of this application.

[0028] Figure 6 This is a force analysis diagram of the gear teeth in one embodiment of this application.

[0029] Figure 7 This is a schematic diagram of a gear dynamics model in one embodiment of this application.

[0030] Figure 8 This is a schematic diagram of the coupling analysis process of gear three-dimensional lubrication state and vibration characteristics in another embodiment of this application.

[0031] Figure 9 This is a schematic diagram of the equivalent model of a multi-branch gear shaft system constructed in another embodiment of this application.

[0032] Figure 10 This is a schematic diagram of the sensor arrangement in another embodiment of this application.

[0033] Figure 11 This is a schematic diagram comparing the instantaneous rotational speed of the flywheel end in another embodiment of this application.

[0034] Figure 12 This is a schematic diagram comparing the instantaneous rotational speed of the oil pump shaft gear end in another embodiment of this application.

[0035] Figure 13 This is a spectrum diagram of the flywheel end angular displacement in another embodiment of this application.

[0036] Figure 14 This is a spectrum diagram of the angular displacement of the oil pump shaft end in another embodiment of this application.

[0037] Figure 15(a) is a schematic diagram comparing the dynamic load and conventional load spectrum results in another embodiment of this application.

[0038] Figure 15(b) is a schematic diagram comparing the speed fluctuation results of the active gear in another embodiment of this application.

[0039] Figure 15(c) is a schematic diagram comparing the speed fluctuation results of the driven gear in another embodiment of this application.

[0040] Figure 15(d) is a schematic diagram comparing the minimum oil film thickness in another embodiment of this application.

[0041] Figure 16(a) is a schematic diagram of the distribution of special points along the meshing line during the meshing cycle in another embodiment of this application.

[0042] Figure 16(b) is a schematic diagram comparing the gear oil film pressure results under smooth surface conditions when considering vibration characteristics oil film pressure in another embodiment of this application.

[0043] Figure 16(c) is a schematic diagram comparing the gear oil film pressure results under smooth surface conditions when there is no vibration characteristic oil film pressure in another embodiment of this application.

[0044] Figure 17(a) shows the average oil film thickness along the meshing line of the hobbing-grinding surface in another embodiment of this application, indicating whether there is vibration or not.

[0045] Figure 17(b) shows the film thickness ratio of the hobbing-grinding surface along the meshing line with or without vibration characteristics in another embodiment of this application.

[0046] Figure 18 This is a schematic diagram of the functional modules of a gear life prediction system based on a lubrication-vibration coupling model, provided in an embodiment of this application.

[0047] Figure 19 This is a schematic diagram of the structure of a computer device provided in an embodiment of this application. Detailed Implementation

[0048] The technical solutions of the embodiments of this application will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of this application, and not all embodiments. Based on the embodiments of this application, all other embodiments obtained by those of ordinary skill in the art without creative effort are within the scope of protection of this application.

[0049] To make the above-mentioned objectives, features and advantages of this application more apparent and understandable, the application will be further described in detail below with reference to the accompanying drawings and specific embodiments.

[0050] The gear life prediction method based on the lubrication-vibration coupling model provided in this application, in an exemplary embodiment, is as follows: Figure 1 As shown, it includes the following steps:

[0051] Step 1: Acquire input data for the gear transmission process. Input data includes geometric parameters, operating parameters, and material parameters. Specifically, geometric parameters include: gear module, number of teeth, pressure angle, helix angle, tooth width, and surface roughness. Operating parameters include: speed, torque, load fluctuations (such as cylinder burst pressure, camshaft load), and lubricant properties (viscosity, viscosity-pressure coefficient). Material parameters include: elastic modulus, Poisson's ratio, and density. These input data can be acquired through experimental measurements (such as roughness testers and speed sensors), CAD model extraction, operating manual lookup, or simulation input.

[0052] Step 2: Establish a three-dimensional hybrid lubrication model for the gear based on the input data; numerically solve the three-dimensional hybrid lubrication model for the gear to obtain the three-dimensional lubrication-related parameters of the gear; the three-dimensional lubrication-related parameters include the pressure distribution in the contact area, the oil film thickness, and the lubrication state; the three-dimensional hybrid lubrication model for the gear is a coupled solution system based on the Reynolds equation, the film thickness equation, and the elastic deformation equation.

[0053] Specifically, solving the Reynolds equations requires considering the viscosity-compression effect (Roelands equations) and density variation (Dowson-Higginson equations), with inlet / outlet pressure as the boundary condition. Solving the film thickness equations requires combining the normal approximation, roughness magnitude, and elastic deformation. The elastic deformation calculation employs the Discrete Convolution-Fast Fourier Transform (DC-FFT) method to efficiently solve the pressure-deformation relationship, avoiding direct matrix integration. Quasi-system numerical solutions and successive mesh refinement (PMD) are used to handle nonlinear problems, ensuring convergence under heavy loads and low speeds.

[0054] In this embodiment, the Reynolds equation for the three-dimensional hybrid lubrication model of the gear is as follows:

[0055]

[0056] Where x and y are coordinate points, t is time, h(t) is the oil film thickness as a function of time, η and ρ are the viscosity and density of the lubricating oil, respectively, p(t) is the oil film pressure as a function of time, and U(y,t) is the velocity at the lubrication interface. This method can uniformly solve for the entire lubrication state, including dry contact, boundary lubrication, mixed lubrication, and full film lubrication, with the entrainment velocity direction coinciding with the x-axis.

[0057] Under the action of surface elastic deformation v(x,y,t) caused by oil film pressure, the film thickness equation of the three-dimensional hybrid lubrication model of gear is as follows:

[0058]

[0059] Where h0(t) is the initial film thickness, and R is the equivalent radius of curvature of the contact pair. δ1(x,y,t) is the geometric approximation term, v(x,y,t) is the elastic deformation displacement, δ1(x,y,t) is the surface roughness of the contact surface of the driving gear, and δ2(x,y,t) is the surface roughness of the contact surface of the driven gear.

[0060] The elastic deformation equation of the three-dimensional hybrid lubrication model of gears is shown below:

[0061]

[0062] Where p(ξ,ζ,t) represents the pressure at different positions (ξ,ζ) and at different times t, Ω is the integration domain of (ξ,ζ), and E' is the elastic modulus of the elastic body.

[0063] As shown in the above equation, the elastic deformation at any point is related to the pressure distribution at each point, requiring integration over the entire solution domain. Traditional calculation methods approximate the pressure distribution on the cell mesh using polynomials to obtain influence coefficients, and then use direct matrix multiplication to calculate the elastic deformation, resulting in extremely high computational complexity. To simplify the workload, a Discrete Convolution-Fast Fourier Transform (DC-FFT) method has been developed to obtain the influence coefficients (pressure-deformation coefficients) of each pressure node on the surface for elastic deformation. The pressure and pressure-deformation coefficients are transformed to the frequency domain, and the elastic deformation at any point on the entire surface under surface pressure is obtained by multiplying the corresponding terms. Thus, the elastic deformation at point (k,l) caused by the pressure distribution in the solution domain can be written as...

[0064]

[0065] Where D(ik,jl) is the influence coefficient of the unit load at coordinate (i,j) on the elastic deformation caused at coordinate (k,l), and p(i,j) is the pressure at coordinate (i,j).

[0066] Integrating the pressure distribution across the entire contact area yields the equation for the lubricating film's load-bearing relationship, as shown below:

[0067] w(t)=∫∫ Ω p(x,y,t)dxdy.

[0068] Where w(t) is the dynamic meshing force obtained in the gear dynamics model.

[0069] A schematic diagram of the equivalent solution domain and roughness mesh processing for gear line contact mixed lubrication is shown below. Figure 2As shown, for the rough surface being processed, the rough surface is first discretized to obtain the roughness amplitude at different discrete points, and then the derivative is calculated for each discrete point. Specifically, in the Reynolds solution domain mesh, the roughness amplitude and derivative values ​​at the solution nodes of the computational mesh are obtained by interpolation using the values ​​S11, S12, S13, and S14 of the neighboring roughness mesh points. If the rough surface is smaller than the computational domain, the rough surface is replicated. The relationship between the roughness mesh and the computational mesh is as follows: Figure 2 As shown, the boundary conditions for solving the Reynolds equations are as follows:

[0070]

[0071] Where the subscript x in and x out These represent the inlet and outlet boundaries of the oil film along the x-direction within the contact domain, respectively, where p is the pressure, and the second term indicates that the pressure derivative in the y-direction is 0.

[0072] Step 3: Calculate the excitation of the gear transmission system based on the input data. The excitation of the gear transmission system is used to determine the relevant parameters of gear vibration excitation. The relevant parameters of vibration excitation include time-varying meshing force, stiffness band, and dynamic load spectrum. The excitation of the gear transmission system includes: cylinder excitation, camshaft load excitation, and time-varying meshing excitation of the gear system. Camshaft load excitation includes cam load torque and oil supply cam normal force; the time-varying meshing excitation of gears is divided into spur gear time-varying meshing excitation and helical gear time-varying meshing excitation according to the gear type.

[0073] In this embodiment, cylinder excitation, camshaft load excitation, and time-varying meshing excitation of spur / helical gears in the gear system are included in the vibration analysis to improve the internal excitation conditions of the gear transmission system.

[0074] Specifically, the cylinder excitation force includes tangential pressure and reciprocating inertial force. For the working process of the crank-connecting rod mechanism, the force analysis of this mechanism is as follows: Figure 3 As shown. Figure 3 In the diagram, A is the equivalent position of the piston's center of mass, i.e., bottom dead center, A' is the top dead center, and α... 00 The centerline and crankshaft angle are given, PH is the gas pressure inside the cylinder, P is the pressure component, PN is the normal force, PC is the centrifugal force, C is the crankshaft rotation center, and β is the crankshaft rotation center. 00 Let L0 be the swing angle, L0 be the length, and s be the length. 00 O3B is the stroke, B is the crank radius, B is the equivalent position of the crank center of mass, x is the distance, and w1 is the direction of crank rotation.

[0075] The tangential excitation force of the crank pin is manifested as the equivalent torque of the tangential gas pressure, which can be calculated using the following formula:

[0076]

[0077] Among them, P T For the equivalent torque of the tangential pressure, α 00 β is the normal pressure angle, i.e., the angle between the normal force and the normal to the tooth surface; 00 The helix angle is the angle between the tooth direction and the axis; P g This represents the tangential pressure of the gas acting on the piston.

[0078] The tangential forces acting on each segment of the crankshaft will generate corresponding tangential moments, as shown in the following formula:

[0079]

[0080] Among them, M g R3 represents the tangential torque generated at each end of the crankshaft, D is the cylinder diameter, and R3 is the geometric parameter related to the lever arm.

[0081] During the rotation of the cylinder's moving parts, the reciprocating inertial force acts on the crankpin and varies periodically with the rotation of the crankshaft. The reciprocating inertial force of the moving parts can be expressed by the following formula:

[0082] P I =-ma.

[0083] Based on the force analysis of the crank and connecting rod, the equivalent torque of the reciprocating inertial force is calculated according to the following formula:

[0084]

[0085] Among them, M I P is the equivalent torque of the reciprocating inertial force. I R3 is the reciprocating inertial force, m is the mass of the reciprocating component, ω1 is the angular velocity of the rotating component, and λ is the link ratio.

[0086] By theoretically deriving the load on the valve train, the load excitation of the camshaft in the shaft system is supplemented, such as... Figure 4 As shown in the figure, this diagram illustrates the structure and stress of the gas distribution mechanism. Figure 4 In the diagram, P4 is the point of action of the cam tappet, O4 is the center of the circle, e is the distance between the center and the point of action, R4 is the base circle radius, h4 is the valve lift, w2 is the angular velocity, T is the torque, and F is the torque. s For friction, F z The force acting on the cam can be calculated using the principle of torque balance, as shown in the following equation:

[0087] TF z ·eF s ·(h4+R4)=0.

[0088] The fuel supply system employs an auxiliary pump injection structure, and the force analysis of its fuel supply camshaft section is as follows: Figure 5 As shown. Figure 5 In the diagram, F1 is the normal force, F2 is the lateral force, and F... z5 Let R6 be the pressure exerted by the plunger on the cam, R5 be the radius of the roller, O5 be the center of rotation of the cam, O6 be the center of rotation of the roller, C6 be the contact point between the roller and the cam, and ω3 be the angular velocity of the cam. From the principle of force balance, the following equation can be obtained:

[0089]

[0090] Where α5 is the cam pressure angle.

[0091] The load torque T6 on a single cam is, T6 = F z (R5+R6+h5)·tanα5, where h5 is the displacement of the cam follower. Decomposing the velocity of the roller, if the cam's rotation angle at that time point is θ6, we obtain the following formula:

[0092]

[0093] Therefore, the simplified result is: The load torque on the cam is:

[0094]

[0095] Where T is the cam load torque, F z5 h5 is the load force acting on the component in contact with the cam, θ6 is the displacement of the cam follower, and d is the derivative.

[0096] During gear meshing, the meshing stiffness exhibits a periodic function characteristic, caused by changes in the number of contact tooth pairs and the tooth contact positions. Based on the principles of mechanics of materials and the energy method, various stiffnesses of spur gears can be synthesized to obtain the time-varying meshing stiffness of external meshing spur gear pairs. The gear teeth are equivalent to non-uniform cantilever beams, and beam theory is used to analyze the forces acting on the teeth. The meshing forces on the gear teeth are as follows... Figure 6 α1 is the meshing angle, α is the meshing pressure angle, and F a and F b R represents the axial and radial forces borne by the gear. b Let M be the base circle radius, M be the input torque, ω be the gear speed, h be the distance from the load application point to the support point, ψ be the phase angle, o be the center of rotation, x be the displacement, and D be the pitch circle diameter. x and h y Let the distance of the force components in the x and y directions be... Figure 6 When the orange line of engagement moves, the engagement angle will also change accordingly, where α2 is the base circle half-tooth angle.

[0097] The meshing angle α0' of the driving gear is,

[0098]

[0099] In the formula, α0' is the pressure angle, θ is the rotation angle of the driving gear, N1 is the number of teeth of the driving gear, and N2 is the number of teeth of the driven gear.

[0100] The meshing angle α1” of the driven gear is,

[0101]

[0102] The stiffness produced by the Hertzian contact effect of gears is shown in the following formula:

[0103]

[0104] Where L1 is the effective contact tooth width and ν is the Poisson's ratio of the gear tooth material.

[0105] When the gear type is a spur gear, if the spur gear meshing passes through the single-tooth meshing zone... The time-varying meshing excitation of spur gears is calculated according to the following formula:

[0106]

[0107] Among them, K nw For helical gears, the time-varying meshing excitation is given, where θ is the rotation angle of the driving gear, and n is the gear tooth number. d = (c0-1)2π / N1 is the meshing start phase angle, c0 is the overlap ratio, N1 is the number of teeth of the driving gear, and k h For the Hertzian contact stiffness of the spur gear, k b1 k s1 k a1 and k f1 These represent the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the gear body, respectively, k. b2 k s2 k a2 and k f2 These are the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the driven gear teeth, respectively.

[0108] If the spur gear meshes through a multi-tooth meshing area The time-varying meshing excitation of spur gears is calculated according to the following formula:

[0109]

[0110] Where i represents the number of teeth of the spur gear meshing simultaneously, and the subscript n represents the nth gear.

[0111] The difference between helical gears and spur gears lies in the helix angle of the helical gear. A helical gear can be divided into several thin tooth segments along its tooth width, and equivalent treatment can be applied in the tooth width direction, so that each small thin tooth segment can be considered as a spur gear. Finally, by calculating the meshing stiffness of each thin tooth segment and performing integration, the meshing stiffness of the entire helical gear tooth surface can be obtained. When the gear type is helical, the stiffnesses of various types are superimposed in parallel, and the time-varying meshing stiffness of the end face, i.e., the time-varying meshing excitation of the helical gear, can be obtained according to the following formula:

[0112]

[0113] Among them, K tw For time-varying meshing excitation of helical gears, i0 is the number of teeth of the helical gear meshing simultaneously, and k hw For the Hertzian contact stiffness of the helical gear, k bw1 k sw1 k aw1 and k fw1 These represent the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the gear body of the driving gear when the gear type is helical. bw2 k sw2 k aw2 and k fw2 These are the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the driven gear when the gear type is helical gear.

[0114] Spur gear time-varying meshing excitation K nw Meshing stiffness K with end face tw The relationship between them is shown in the following formula:

[0115] 1 / K nw =(1 / K) tw ) / cos 2 (β b ).

[0116] Step 4: Based on the three-dimensional hybrid lubrication model of the gear and the excitation of the gear transmission system, establish a lubrication and vibration coupling model. The lubrication and vibration coupling model is established by connecting the oil film stiffness and the gear meshing stiffness in parallel to form a comprehensive stiffness, and establishing a bending-torsional coupling dynamic equation that includes torsion and lateral degrees of freedom. By iteratively solving the lubrication and vibration coupling model, the coupling analysis of lubrication and vibration can be realized, and the dynamic oil film pressure, minimum film thickness and vibration displacement / velocity spectrum can be output.

[0117] The lubricating oil film between gear teeth not only reduces friction, wear, and provides cooling, but also reflects the displacement state of the lubricating oil film within the contact region under transient forces. Furthermore, it directly affects the dynamic parameter of the overall meshing stiffness of the gear teeth, and consequently influences the vibration of the entire system. Researching the coupling relationship between the three-dimensional lubrication state of gear pairs and their dynamic characteristics and system vibration characteristics has significant theoretical value and engineering guiding significance for the design and research of gear-shaft systems.

[0118] Under elastohydrodynamic lubrication, the definition of oil film stiffness in the elastohydrodynamic formula is as follows:

[0119]

[0120] Where, k o For oil film stiffness, δ ∞ This represents the compressive deformation of the oil film. When a gear is subjected to transient loads, the oil film deformation under different loads can be obtained as follows:

[0121] δ ∞ =h(F)-h(F+ΔF).

[0122] If the oil film is deformed by δ ∞ Dimensionless transformation is performed to obtain the dimensionless oil film deformation Δ. ∞ It is defined as follows:

[0123]

[0124] Among them, c c For Hertzian contact deformation, the expression is: Therefore, by combining the above formulas, a new oil film stiffness k can be obtained. oi The formula is as follows:

[0125]

[0126] Where p(L) and q(L) are both functions of the dimensionless lubrication parameter L, and their specific expressions are as follows:

[0127] p(L)=[(4-0.2L) 7 +(3.5+0.1L)] 17 .

[0128]

[0129]

[0130] Where, α η Let be the viscosity-pressure coefficient. Furthermore, the oil film stiffness under the average film thickness method can be obtained as follows:

[0131]

[0132] Where ΔF is the load increment, The increment of the average oil film thickness within the Hertzian contact zone is represented by the following formula:

[0133]

[0134] Considering the effect of oil film lubrication, the meshing stiffness of the gears is coupled with the oil film stiffness to solve for the overall stiffness of the system. The calculation formula is shown below:

[0135]

[0136] Where, k t For overall stiffness, k o For oil film stiffness, k s This refers to the gear meshing stiffness.

[0137] After obtaining the overall meshing stiffness of the gears, a refined model of bending-torsional coupling is performed on the spur / helical gears of this model. In this model, the lateral degree of freedom is used as a condition for the input of the lateral load on the gears, which will have a coupling effect on the torsional vibration characteristics of the transmission shaft system. For shaft components other than gears, only the torsional degree of freedom is considered. Considering the combined effects of torsional θ, lateral x and y direction motion, and tooth surface meshing force, a model is established as follows: Figure 7 The gear pair bending-torsional coupling model shown is as follows. Figure 7 In the text, subscripts 1 and 2 represent component 1 and component 2 respectively, such as F dy1 F represents the dynamic load on component 1 in the y-direction. sy1 F is the static yield force of component 1 in the y direction. sx1 F is the static lateral force in the x-direction. dx1 For the dynamic load in the x-direction, F d12 F is the interaction force between component 1 and component 2 under dynamic coupling force. s12 k represents the interaction force between component 1 and component 2 under static coupling force. y1 k is the stiffness coefficient in the y-direction. x1 k is the stiffness coefficient in the x-direction. p For torsional stiffness, k t For contact stiffness, the elastic properties of the contact surface, c x1 c is the damping coefficient in the x-direction. p c is the torsional damping coefficient. g c is the gear damping coefficient. t R is the contact damping coefficient. b1 Let I be the radius of the base circle, and I1 be the moment of inertia. m I is the moment of inertia of the motor. b Let x1 be the displacement in the x-direction due to the inertia of the external load, and θ be the displacement in the x-direction.m Let θ be the motor rotation angle. b M1 is the load angle, M2 is the input torque, and M is the output torque. pk For peak torque, M pc For continuous operating torque, M gk M is the gear meshing torque. gc For the continuous torque of the gear, C y1 C is the dynamic load factor in the y-direction. v2 r is the velocity-dependent damping coefficient. sy1 This is the static yield force correction factor. The model has a fixed rigid boundary, where the contact boundary at the bearing support is connected by a linear spring k. x1 k x2 Equivalently, damping is achieved through c x1 c x2 Equivalent.

[0138] By setting the direction of the gear meshing line as the y-axis and the perpendicular direction of the meshing line as the x-axis, a local rectangular coordinate system xoy is established. The bending-torsional coupling dynamic equation of the gear pair with oil film stiffness is as follows:

[0139]

[0140] Where m1 is the mass of the driving gear, m2 is the mass of the driven gear, I1 is the moment of inertia of the driving gear, I2 is the moment of inertia of the driven gear, and I... m I is the moment of inertia of the internal shaft system. b Let R be the moment of inertia of the external shaft system, k and c be the stiffness and damping respectively, and R be the moment of inertia of the external shaft system. b1 R b2 Let θ be the radius of the base circle, and θ1 and θ2 be the rotation angles of the driving and driven wheels, respectively. m θ is the rotation angle output by the motor. b M1 and M2 are the external load rotation angles, and M1, M2, M3 are the load torques. x1, x2, y1, and y2 are the displacements of the driving wheel and the driven wheel in the x and y directions, respectively.

[0141] The dynamic relative displacement difference between the driving and driven gears is the dynamic transmission error, and the calculation formula is as follows:

[0142] δ=R b1 θ1-R b2 θ2+y1-y2+e(t).

[0143] Where e(t) represents the errors in tooth pitch and tooth profile during gear manufacturing, and the calculation formula is:

[0144] e(t) = e r sin(2πf m t+φ).

[0145] Among them, e r This is a comprehensive measure of manufacturing errors such as gear pressure angle error and pitch error, f m Let φ be the meshing frequency and φ be the initial phase. Then, the dynamic meshing force, or dynamic load F, during gear transmission is... dp Calculated using the following formula:

[0146] F dp =k t δ.

[0147] Step 5: By analyzing the output results of the lubrication and vibration coupling model, the gear life prediction results are obtained; the analysis includes lubrication condition assessment, vibration characteristic optimization, fatigue life prediction, and wear life prediction.

[0148] For example, lubrication condition assessment involves determining the mixed lubrication zone through film thickness ratio and predicting wear risk (e.g., increased micro-protrusion contact when film thickness ratio < 3). Vibration characteristic optimization involves analyzing speed fluctuations and dynamic load peaks to optimize gear parameters and reduce resonance risk. Fatigue life prediction is based on the number of contact stress cycles and the material's SN curve. Wear life prediction combines the Archard model with lubrication condition data to calculate the wear rate.

[0149] The main process in actual implementation includes: inputting actual working condition parameters → coupled model simulation → extracting stress / film thickness time history → applying damage accumulation theory → outputting life curve.

[0150] In another exemplary embodiment, the entire gear transmission shaft system is lumped parametrically modeled to obtain the inherent characteristics of the entire shaft system. Simultaneously, various internal and external excitations are integrated to predict the forced vibration response of the shaft system. Due to the harsh working conditions and complex interface micro-roughness, gears experience both lubrication and contact during meshing. This coexistence, combined with vibration fluctuation characteristics, elastic deformation, viscosity-pressure relationships, oil film shear effects, and actual surface roughness factors, leads to a strongly nonlinear problem in the fundamental equations of mixed lubrication. Traditional numerical methods, to compensate for this deficiency, primarily discretize the pressure flow on the left-hand side during iteration, treating the discretized nodal pressure as unknowns and the right-hand side as knowns. The coefficient matrix of the algebraic equations is entirely derived from the pressure flow terms on the left-hand side. However, when low-speed, heavy-load conditions occur, resulting in extremely thin oil films and very weak pressure flows, traditional numerical methods can cause the coefficient matrix to become ill-conditioned, losing its diagonal-dominant characteristic, leading to degenerate solution processes, convergence difficulties, and even overflow.

[0151] To address the shortcomings of traditional numerical methods, a quasi-system numerical method is introduced. Specifically, the film thickness in the right-hand convection flow is also considered a function of the unknown nodal pressure, thus constructing a coefficient matrix that includes the convection flow. Therefore, even under heavy load and low speed conditions, the coefficient matrix maintains diagonal dominance, significantly improving the convergence and stability of the solution. Simultaneously, a mesh refinement method (PMD) is employed to ensure rapid convergence of the computational solution. This involves replacing the approximately convergent solution obtained in a lower-level mesh with the solution in a higher-level mesh, and increasing the convergence accuracy requirement in the higher-level mesh to obtain a satisfactory final solution. This accelerates the solution speed while ensuring the accuracy of the numerical solution.

[0152] Figure 8 This is a flowchart of the coupled analysis and calculation of the three-dimensional lubrication state and vibration characteristics of a gear system. Using measured actual machined surfaces, gear shaft structure, operating conditions, and lubricating oil parameters as inputs, a lumped parametric model of the gear transmission system is performed to obtain the inherent characteristics and transient response of the entire system. After verifying the accuracy of the established coupled model using actual shaft test data, the internal excitation state of the entire system is obtained by combining the force analysis of each working element. Based on the improved gear dynamic characteristic model, dynamic load and speed fluctuations are obtained and substituted into the three-dimensional lubrication state analysis model to obtain the oil film stiffness state. The oil film stiffness excitation is fed back into the gear dynamic characteristic model to complete the model coupling. When the preset number of steps is reached, the required results are output, and the calculation terminates.

[0153] In another exemplary embodiment of this application, the Newmark stepwise integration algorithm is used to analyze the forced vibration of the gear transmission shaft system. The vibration differential equation of the lumped parameter matrix of the shaft system is:

[0154]

[0155] The Newmark stepwise integration master equation is:

[0156]

[0157]

[0158] Where, {x} n , Let denot n be the displacement, velocity, and acceleration of the nth iteration, and let γ and β be higher-order convergence constants. When γ ≥ 1 / 2 and β ≥ γ / 2, the results of the Newmark loop calculation converge unconditionally.

[0159] Based on the principle of equivalent simplification, a lumped parameter model is performed on the multi-branch transmission shaft system of this gear. Its equivalent model is as follows: Figure 9As shown. The entire shaft system consists of 67 moments of inertia and 66 torsional stiffnesses between these moments of inertia. The lateral vibration system consists of 22 gear masses and lateral stiffnesses. To verify the accuracy of the established coupled analysis model, experiments were conducted on the transmission shaft system at both the free vibration and forced vibration levels. Torsional vibration measuring points were set up at the flywheel and oil pump shaft. The approximate structure and sensor arrangement are shown below. Figure 10 As shown.

[0160] Steady-state response tests were conducted to examine the speed fluctuations at different locations, and the calculated and measured instantaneous speed signals at the flywheel end and the oil pump shaft gear end were then verified and analyzed. For example... Figures 11-12 The time-domain results are shown. From Figure 11 It can be seen that the calculated value of the flywheel end speed fluctuation is 21.599 r / min, which deviates from the measured speed fluctuation by 19.46%. From... Figure 12 It can be seen that the calculated value of the speed fluctuation at the oil pump shaft gear end is 242.817 r / min, which deviates from the measured speed fluctuation by 6.71%. The speed fluctuation deviation at the flywheel end is significantly greater than that at the oil pump shaft gear end. This is mainly because the flywheel is subjected to more external disturbances, making the vibration transmission more prominent at the flywheel. In addition, the flywheel connects to more transmission components, increasing the complexity and uncertainty of the system. At the same time, the load variation on the flywheel is greater than that at the oil pump shaft gear. These factors combined lead to a large deviation between the measured and actual speed fluctuation values ​​at the flywheel end. Furthermore, the combined experimental and simulation results also show that the speed fluctuation at the oil pump shaft gear end is significantly greater than that at the flywheel end, and its vibration state is relatively more severe.

[0161] The comparison charts of the frequency domain and measured data of the angular displacement at the flywheel end and the oil pump shaft gear end are shown below. Figure 13 and 14 As shown in the diagram, analysis of the frequency spectrum reveals that the vibration energy of the displacement spectrum is mainly concentrated in the low-frequency region. Regarding the flywheel end angular displacement frequency spectrum, due to the excessive external load on this four-stroke engine, a low-frequency disturbance with a high peak value exists at 5Hz, with an amplitude greater than the shaft frequency amplitude, and a peak value also exists at 0.5 times the shaft frequency harmonic frequency. Therefore, it can be concluded that the measured and calculated values ​​of the flywheel end and oil pump shaft gear end angular displacements are consistent in frequency composition, with only amplitude deviations at some shaft frequency harmonic peaks.

[0162] Comprehensive analysis shows that when the constructed coupled analysis model is used for forced vibration analysis, the calculated instantaneous rotational speed in the time domain is in excellent agreement with the measured data; in the frequency domain, the measured and calculated values ​​have the same frequency components. The spectral amplitude of the torsion angle of the oil pump shaft gear is greater than that of the torsion angle at the flywheel end, which means that there are problems with excessive speed fluctuations and energy distribution, resulting in poor vibration characteristics and indicating that the gear meshing environment at the oil pump shaft is extremely harsh.

[0163] In another exemplary embodiment of this application, in order to compare the influence of combining the three-dimensional lubrication model and the vibration model on the transient meshing characteristics of the gear, Figure 15(a) shows a comparison between the dynamic load caused by the coupled vibration characteristics and the traditional load spectrum results in the following formula. Figures 15(b)-15(c) Figure 15(d) shows the comparison of the effects of velocity fluctuations on the smooth surface of gear teeth caused by coupled vibration characteristics.

[0164]

[0165] Among them, t p t is the total engagement time. a =0, t d =t p .

[0166] As shown in Figure 15(a), the traditional load diagram utilizes a relatively ideal alternating single and double tooth meshing to change the transient load state. Compared to the load diagram, the dynamic meshing force during gear meshing differs significantly. Figures 15(b)-15(d) It can be seen that the torsional vibration present during gear transmission does cause velocity fluctuations on the surfaces of the two gears. The velocity fluctuations on the gear surfaces caused by vibration bring about new entrainment velocities, thereby changing the slippage state and meshing performance of the two gears. The minimum oil film thickness distribution also changes significantly compared to the state without the influence of vibration factors, with obvious changes in value and trend. There are many 0 film thickness positions along the meshing line, and the lubrication condition becomes worse.

[0167] Figure 16(a) shows the special positions of the gear meshing process within one cycle, where B b and B a It refers to the engagement point and the disengagement point, point A. b and point A a These are the transition points between single-tooth meshing and double-tooth meshing, respectively, and point P is a node. Figures 16(b)-16(c)To consider the impact of vibration characteristics on the smooth surface of the gear, a comparison was made between the two-dimensional center oil film pressure and the three-dimensional contact area oil film pressure distribution. As shown in the figure, after coupling vibration characteristics, the two-dimensional center oil film pressure increased significantly, while the three-dimensional oil film pressure distribution changed considerably compared to the condition without vibration characteristics. The numerical and trend changes were obvious, and the secondary pressure peak characteristics were more pronounced.

[0168] The film thickness ratio, the ratio of the average thickness of the lubricating oil film between contacting tooth surfaces to its root mean square roughness, is an important indicator for measuring lubrication status. Figure 17(a) shows the average oil film thickness, and Figure 17(b) shows the film thickness ratio distribution. The average oil film thickness after considering the coupling characteristics of three-dimensional lubrication and vibration is significantly smaller than the result without vibration. The maximum oil film thickness under vibration influence is only 81.93% of the result without vibration influence, and its average oil film thickness is smaller and fluctuates more drastically. Furthermore, the film thickness ratios affected by the coupling characteristics of three-dimensional lubrication and vibration are all less than 3.0, fluctuating in the range of 1.72-2.88. At this point, the oil film thickness is insufficient to completely separate the two surfaces, and micro-protrusions within the contact domain come into contact, resulting in a completely mixed lubrication state during meshing. In contrast, the film thickness ratios without vibration influence fluctuate in the range of 2.32-3.51, with 29.67% of the contacting tooth surfaces within the contact domain completely separated by the oil film, indicating a full film lubrication state. Based on the above analysis, the influence of vibration characteristic coupling effect on lubrication state characteristics is significant and cannot be ignored.

[0169] The method described in this application considers the influence of interfacial mechanics factors such as the transient working conditions of meshing gears and the lubrication-three-dimensional rough peak contact state. Combined with the force analysis of each working element, it obtains the internal excitation state of the entire system, and then establishes a lubrication and vibration coupling model under various types of comprehensive excitation, including oil film excitation. This enables the coupled analysis of gear vibration and three-dimensional lubrication. The correctness of the model is verified by combining actual machine test and simulation test data. This method is of great significance for establishing a three-dimensional lubrication and vibration coupling model of gear transmission system.

[0170] Based on the same inventive concept, this application also provides a system for implementing the gear life prediction method based on the lubrication-vibration coupling model described above. The solution provided by this system is similar to the solution described in the above method; therefore, the specific limitations in one or more system embodiments provided below can be found in the limitations of the gear life prediction method based on the lubrication-vibration coupling model described above, and will not be repeated here.

[0171] In one exemplary embodiment, such as Figure 18 As shown, a gear life prediction system based on a lubrication-vibration coupling model is provided, including:

[0172] The gear transmission data acquisition module is used to acquire input data during the gear transmission process; the input data includes geometric parameters, operating condition parameters, and material parameters.

[0173] The gear lubrication model construction module is used to build a three-dimensional hybrid lubrication model of gears based on input data. Numerical solutions are performed on the three-dimensional hybrid lubrication model of gears to obtain three-dimensional lubrication-related parameters of gears. The three-dimensional lubrication-related parameters include pressure distribution in the contact area, oil film thickness, and lubrication state. The three-dimensional hybrid lubrication model of gears is a coupled solution system based on the Reynolds equation, film thickness equation, and elastic deformation equation.

[0174] The gear transmission excitation calculation module is used to calculate the excitation of the gear transmission system based on the input data; the excitation of the gear transmission system is used to determine the vibration excitation parameters of the gear; the vibration excitation parameters include time-varying meshing force, stiffness band, and dynamic load spectrum.

[0175] The lubrication and vibration coupling module is used to establish a lubrication and vibration coupling model based on the three-dimensional hybrid lubrication model of gears and the excitation of the gear transmission system. The lubrication and vibration coupling model is established by forming a comprehensive stiffness by connecting the oil film stiffness and gear meshing stiffness in parallel, and establishing a bending-torsional coupling dynamic equation that includes torsional and lateral degrees of freedom. By iteratively solving the lubrication and vibration coupling model, the coupling analysis of lubrication and vibration can be realized, and the dynamic oil film pressure, minimum film thickness and vibration displacement / velocity spectrum can be output.

[0176] The gear life prediction module is used to obtain gear life prediction results by analyzing the output results of the lubrication and vibration coupling model; the analysis includes lubrication condition assessment, vibration characteristic optimization, fatigue life prediction, and wear life prediction.

[0177] certainly, Figure 18 The architecture shown is merely exemplary; it can be omitted as needed when implementing different functionalities. Figure 18 One or at least two components of the system shown.

[0178] In one exemplary embodiment, a computer device is provided, which may be a server or a terminal, and its internal structure diagram may be as follows. Figure 19As shown, the computer device includes a processor, memory, input / output interfaces (I / O), and a communication interface. The processor, memory, and I / O interfaces are connected via a system bus, and the communication interface is also connected to the system bus via the I / O interfaces. The processor provides computing and control capabilities. The memory includes non-volatile storage media and internal memory. The non-volatile storage media stores the operating system, computer programs, and a database. The internal memory provides an environment for the operation of the operating system and computer programs stored in the non-volatile storage media. The I / O interfaces are used for exchanging information between the processor and external devices. The communication interface is used for communicating with external terminals via a network connection. When the computer program is executed by the processor, it can implement the methods provided in the preceding embodiments.

[0179] Those skilled in the art will understand that Figure 19 The structure shown is merely a block diagram of a portion of the structure related to the present application and does not constitute a limitation on the computer device to which the present application is applied. Specific computer devices may include more or fewer components than those shown in the figure, or combine certain components, or have different component arrangements.

[0180] In one exemplary embodiment, a computer device is also provided, including a memory and a processor, wherein the memory stores a computer program, and the processor executes the computer program to implement the steps in the above-described method embodiments.

[0181] In one exemplary embodiment, a computer-readable storage medium is provided storing a computer program that, when executed by a processor, implements the steps in the above-described method embodiments.

[0182] In one exemplary embodiment, a computer program product is provided, including a computer program that, when executed by a processor, implements the steps in the above-described method embodiments.

[0183] It should be noted that the user information (including but not limited to user device information, user personal information, etc.) and data (including but not limited to data used for analysis, data stored, data displayed, etc.) involved in this application are all information and data authorized by the user or fully authorized by all parties, and the collection, use and processing of the relevant data must comply with relevant regulations.

[0184] Those skilled in the art will understand that all or part of the processes in the above embodiments can be implemented by a computer program instructing related hardware. The computer program can be stored in a non-volatile computer-readable storage medium. When executed, the computer program can include the processes of the embodiments described above. Any references to memory, databases, or other media used in the embodiments provided in this application can include at least one of non-volatile and volatile memory. Non-volatile memory can include read-only memory (ROM), magnetic tape, floppy disk, flash memory, optical memory, high-density embedded non-volatile memory, resistive random access memory (ReRAM), magnetic random access memory (MRAM), ferroelectric random access memory (FRAM), phase change memory (PCM), graphene memory, etc. Volatile memory can include random access memory (RAM) or external cache memory, etc. By way of illustration and not limitation, RAM can take many forms, such as Static Random Access Memory (SRAM) or Dynamic Random Access Memory (DRAM).

[0185] The databases involved in the embodiments provided in this application may include at least one type of relational database and non-relational database. Non-relational databases may include, but are not limited to, blockchain-based distributed databases. The processors involved in the embodiments provided in this application may be general-purpose processors, central processing units, graphics processing units, digital signal processors, programmable logic devices, quantum computing-based data processing logic devices, etc., and are not limited to these.

[0186] The technical features of the above embodiments can be combined in any way. For the sake of brevity, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0187] This document uses specific examples to illustrate the principles and implementation methods of this application. The descriptions of the above embodiments are only for the purpose of helping to understand the methods and core ideas of this application. Furthermore, those skilled in the art will recognize that, based on the ideas of this application, there will be changes in the specific implementation methods and application scope. Therefore, the content of this specification should not be construed as a limitation of this application.

Claims

1. A method for predicting gear life based on a lubrication-vibration coupling model, characterized in that, include: The input data during the gear transmission process is acquired; the input data includes geometric parameters, operating condition parameters, and material parameters. A three-dimensional hybrid lubrication model for gears is established based on the input data. Numerical solutions are then performed on the three-dimensional hybrid lubrication model to obtain the gear's three-dimensional lubrication-related parameters. These parameters include contact area pressure distribution, oil film thickness, and lubrication state. The three-dimensional hybrid lubrication model is a coupled solution system based on the Reynolds equation, film thickness equation, and elastic deformation equation. The excitation of the gear transmission system is calculated based on the input data; the excitation of the gear transmission system is used to determine the vibration excitation parameters of the gear; the vibration excitation parameters include time-varying meshing force, stiffness band, and dynamic load spectrum. Based on the aforementioned three-dimensional hybrid lubrication model of the gear and the excitation of the gear transmission system, a lubrication and vibration coupling model is established. The lubrication and vibration coupling model is established by forming a comprehensive stiffness by connecting the oil film stiffness and the gear meshing stiffness in parallel, and establishing a bending-torsional coupling dynamic equation that includes torsional and lateral degrees of freedom. By iteratively solving the lubrication and vibration coupling model, the coupling analysis of lubrication and vibration can be realized, and the dynamic oil film pressure, minimum film thickness, and vibration displacement / velocity spectrum can be output. By analyzing the output results of the lubrication and vibration coupling model, gear life prediction results are obtained; the analysis includes lubrication condition assessment, vibration characteristic optimization, fatigue life prediction, and wear life prediction.

2. The gear life prediction method based on the lubrication-vibration coupling model according to claim 1, characterized in that, The geometric parameters include: gear module, number of teeth, pressure angle, helix angle, tooth width, and surface roughness; the operating parameters include: speed, torque, load fluctuation, and lubricating oil characteristics; the material parameters include: elastic modulus, Poisson's ratio, and density; the input data is acquired through experimental measurement, CAD model extraction, operating manual lookup, or simulation input.

3. The gear life prediction method based on the lubrication-vibration coupling model according to claim 1, characterized in that, The Reynolds equation for the three-dimensional hybrid lubrication model of the gear is shown below: Where x and y are coordinate points, t is time, h(t) is the oil film thickness as a function of time, η and ρ are the viscosity and density of the lubricating oil, respectively, p(t) is the oil film pressure as a function of time, and U(y,t) is the velocity of movement on the lubrication interface. The film thickness equation for the three-dimensional hybrid lubrication model of the gear is shown in the following formula: Where h0(t) is the initial film thickness, and R is the equivalent radius of curvature of the contact pair. δ1(x,y,t) is the geometric approximation term, v(x,y,t) is the elastic deformation displacement, δ1(x,y,t) is the surface roughness of the contact surface of the driving gear, and δ2(x,y,t) is the surface roughness of the contact surface of the driven gear. The elastic deformation equation of the three-dimensional hybrid lubrication model of the gear is shown in the following equation: Where p(ξ,ζ,t) represents the pressure at different positions (ξ,ζ) and at different times t, Ω is the integration domain of (ξ,ζ), and E' is the elastic modulus of the elastic body; The boundary conditions for solving the Reynolds equations are shown in the following equation: Where the subscript x in and x out These represent the inlet and outlet boundaries of the oil film along the x-direction within the contact domain, respectively, where p is the pressure, and the second term indicates that the pressure derivative in the y-direction is 0.

4. The gear life prediction method based on the lubrication-vibration coupling model according to claim 1, characterized in that, The excitation of the gear transmission system includes: cylinder excitation, camshaft load excitation, and time-varying gear meshing excitation of the gear system; the time-varying gear meshing excitation is classified into spur gear time-varying meshing excitation and helical gear time-varying meshing excitation according to the gear type; The cylinder excitation force includes tangential pressure and reciprocating inertial force. The equivalent torque of the tangential pressure is calculated according to the following formula: Among them, P T For the equivalent torque of the tangential pressure, α 00 β is the normal pressure angle, i.e., the angle between the normal force and the normal to the tooth surface; 00 The helix angle is the angle between the tooth direction and the axis; P g This represents the tangential pressure of the gas acting on the piston. The equivalent torque of the reciprocating inertial force is calculated according to the following formula: Among them, M I P is the equivalent torque of the reciprocating inertial force. I R3 is the reciprocating inertial force, R3 is the geometric parameter related to the lever arm, m is the mass of the reciprocating motion component, ω1 is the angular velocity of the rotating component, and λ is the link ratio. The camshaft load excitation includes the cam load torque and the normal force of the oil supply cam; the cam load torque is calculated according to the following formula: Where T is the cam load torque, F z5 h5 is the load force acting on the component in contact with the cam, θ6 is the displacement of the cam follower, and d is the derivative.

5. The gear life prediction method based on the lubrication-vibration coupling model according to claim 4, characterized in that, When the gear type is a spur gear, if the spur gear meshes through a single-tooth meshing zone... The time-varying meshing excitation of the spur gear is calculated according to the following formula: Among them, K nw For helical gears, the time-varying meshing excitation is given, where θ is the rotation angle of the driving gear, and n is the gear tooth number. d = (c0-1)2π / N1 is the meshing start phase angle, c0 is the overlap ratio, N1 is the number of teeth of the driving gear, and k h For the Hertzian contact stiffness of the spur gear, k b1 k s1 k a1 and k f1 These represent the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the gear body, respectively, k. b2 k s2 k a2 and k f2 These are the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the driven gear teeth, respectively. If the spur gear meshes through a multi-tooth meshing area The time-varying meshing excitation of the spur gear is calculated according to the following formula: Where i is the number of teeth of the spur gear meshing simultaneously, and the subscript n is the nth gear; When the gear type is helical gear, the time-varying meshing excitation of the helical gear is calculated according to the following formula: Among them, K tw For time-varying meshing excitation of helical gears, i0 is the number of teeth of the helical gear meshing simultaneously, and k hw For the Hertzian contact stiffness of the helical gear, k bw1 k sw1 k aw1 and k fw1 These represent the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the gear body of the driving gear when the gear type is helical. bw2 k sw2 k aw2 and k fw2 These are the bending stiffness, shear stiffness, axial compressive stiffness, and additional stiffness of the driven gear when the gear type is helical gear.

6. The gear life prediction method based on the lubrication-vibration coupling model according to claim 1, characterized in that, The oil film stiffness and gear meshing stiffness are combined in parallel according to the following formula to form the overall stiffness: Where, k t For overall stiffness, k o For oil film stiffness, k s For gear meshing stiffness; The dynamic equations for bending-torsional coupling are shown below: Where m1 is the mass of the driving gear, m2 is the mass of the driven gear, I1 is the moment of inertia of the driving gear, I2 is the moment of inertia of the driven gear, and I... m I is the moment of inertia of the internal shaft system. b Let R be the moment of inertia of the external shaft system, k and c be the stiffness and damping respectively, and R be the moment of inertia of the external shaft system. b1 R b2 Let θ be the radius of the base circle, and θ1 and θ2 be the rotation angles of the driving and driven wheels, respectively. m θ is the rotation angle output by the motor. b M1 and M2 are the external load rotation angles, and M1, M2, M3 are the load torques. x1, x2, y1, and y2 are the displacements of the driving wheel and the driven wheel in the x and y directions, respectively.

7. A gear life prediction system based on a lubrication-vibration coupling model, characterized in that, include: The gear transmission data acquisition module is used to acquire input data during the gear transmission process; the input data includes geometric parameters, operating condition parameters, and material parameters. The gear lubrication model construction module is used to establish a three-dimensional hybrid lubrication model of the gear based on the input data; numerical solutions are performed on the three-dimensional hybrid lubrication model of the gear to obtain the three-dimensional lubrication-related parameters of the gear; the three-dimensional lubrication-related parameters include the pressure distribution in the contact area, the oil film thickness, and the lubrication state; the three-dimensional hybrid lubrication model of the gear is a coupled solution system based on the Reynolds equation, the film thickness equation, and the elastic deformation equation; The gear transmission excitation calculation module is used to calculate the excitation of the gear transmission system based on the input data; the excitation of the gear transmission system is used to determine the vibration excitation related parameters of the gear; the vibration excitation related parameters include time-varying meshing force, stiffness band, and dynamic load spectrum. The lubrication and vibration coupling module is used to establish a lubrication and vibration coupling model based on the three-dimensional hybrid lubrication model of the gear and the excitation of the gear transmission system. The lubrication and vibration coupling model is established by forming a comprehensive stiffness by connecting the oil film stiffness and the gear meshing stiffness in parallel, and establishing a bending-torsional coupling dynamic equation that includes torsional and lateral degrees of freedom. By iteratively solving the lubrication and vibration coupling model, the coupling analysis of lubrication and vibration can be realized, and the dynamic oil film pressure, minimum film thickness and vibration displacement / velocity spectrum can be output. The gear life prediction module is used to obtain gear life prediction results by analyzing the output results of the lubrication and vibration coupling model; the analysis includes lubrication condition assessment, vibration characteristic optimization, fatigue life prediction and wear life prediction.

8. A computer device, comprising: A memory, a processor, and a computer program stored in the memory and executable on the processor, characterized in that the processor executes the computer program to implement the gear life prediction method based on the lubrication-vibration coupling model according to any one of claims 1-6.

9. A computer-readable storage medium having a computer program stored thereon, characterized in that, When the computer program is executed by the processor, it implements the gear life prediction method based on the lubrication-vibration coupling model as described in any one of claims 1-6.

10. A computer program product, comprising a computer program, characterized in that, When the computer program is executed by the processor, it implements the gear life prediction method based on the lubrication-vibration coupling model as described in any one of claims 1-6.

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