CN116451521A - Impact fatigue life calculation method for electric automobile differential shell - Google Patents

Abstract

The invention discloses a method for calculating the impact fatigue life of an electric automobile differential shell, which comprises the following steps: constructing an assembled finite element model of the differential case; setting boundary conditions equipped with a finite element model; defining finite element model material properties; determining the positions and the number of loading points according to the structural symmetry of the differential mechanism shell; applying finite element model calculation load comprising bolt pretightening force and meshing force load at each loading position; setting working conditions of forward driving impact and reverse driving impact, wherein different load steps under the working conditions correspond to loading positions one by one; performing differential case finite element model calculation analysis; preliminary analysis of the stress field calculation result of the differential case; adding the plastic attribute of the differential case, and carrying out elastoplastic finite element analysis on an assembly model of the differential case; determining wave crests and wave troughs required by impact fatigue damage calculation according to the adopted calculation results; calculating an impact fatigue damage value when the loading position is I, J, K; impact fatigue life calculations were performed.

Description

Impact fatigue life calculation method for electric automobile differential shell

Technical Field

The invention belongs to the technical field of automobile part performance test, and particularly relates to an impact fatigue life calculation method of an electric automobile differential shell.

Background

As a representative of a cleaner new energy vehicle, the electric automobile has significant differences from the traditional fuel automobile in principle and part of construction, and the research and development test of a transmission system of the electric automobile has become an important point of industry. The driving motor with the characteristics of low speed and constant torque has rapid output response, but brings larger impact load to the transmission system due to the characteristics of electromagnetic induction principle and the application of the braking energy recovery system. Compared with a fuel oil automobile, the transmission system has the advantages that the transmission system torque response speed is accelerated, the braking energy is recovered to bring the reverse torque impact load to the transmission system and other factors caused by the performance requirements such as spline direct connection structural form, whole automobile acceleration performance and the like are eliminated, so that the impact problem of the transmission system of the electric automobile is particularly remarkable. The differential mechanism is used as the last ring of speed reduction and torque increase, the stress condition is worse, and the impact load can be caused to occur in different positions due to the symmetry difference of the structural design of the differential mechanism shell, the impact fatigue performance difference is larger, and the differential mechanism shell is used as the supporting structure of the differential mechanism assembly, so that the fatigue performance of the differential mechanism against the impact load is necessary to be brought into performance design, analysis and evaluation.

In recent years, fatigue life prediction of the differential case of the electric automobile is still mainly based on multi-axis fatigue life calculation and fatigue test along the traditional load spectrum, and the related study on the impact fatigue life caused by impact load is very little. For example, in "a High-Accuracy Fatigue Analysis of Vehicle Differential Case Based on Cumulative Damage Theory under Approximate Actual Load Conditions", stress analysis is performed on the differential case at 18 circumferential loading positions, and fatigue damage and life of the differential case are calculated, but only High-cycle fatigue damage under normal load is calculated, and impact fatigue condition under large impact load is not involved; similarly, in the section of the simulation analysis method for fatigue life of differential case, the author acquires the elastic stress of the differential case line by adopting a line elastic material, calculates the high cycle fatigue damage of the differential case after correcting the stress by applying a Neuber formula, and simulates the high cycle fatigue risk in the actual operation process of the differential by adopting a three-plus-one-minus conventional load type, so that the fatigue condition caused by positive and negative impact loads caused by acceleration-braking is not considered, and the low cycle fatigue condition caused by the impact of the load and in the elastoplastic state of the material is not involved. In the fatigue life analysis of the differential mechanism shell of the electric drive assembly, an author considers the influence of impact load on the service life of the differential mechanism shell based on actually measured load spectrum of the differential mechanism shell, but the acceptable precision is obtained only by calibrating and mutually verifying a finite element model and a bench test of the differential mechanism shell, the judging period for judging whether the performance of the differential mechanism shell meets the design requirement is overlong, the fatigue life of the shell is predicted only by adopting a high-cycle fatigue damage calculation method, and the low-cycle fatigue condition of the differential mechanism shell under the condition of yielding caused by the impact load is not analyzed. In the patent 'impact strength and fatigue analysis method of a transmission differential mechanism', the impact torque of the whole automobile is taken as an input load, the quasi-static impact strength is calculated, the stress analysis result is linearly scaled according to the load ratio of the load to the stress analysis, the fatigue damage of each torque level is obtained and then superimposed to calculate the fatigue damage, but the impact action mode of the torque load is still from a traditional fuel automobile, the fatigue caused by the stress process of one revolution of the differential mechanism is mainly considered, the impact fatigue occurrence scene of the differential mechanism is completely different from that of the differential mechanism of the electric automobile, the stress calculation is carried out on the differential mechanism at 20 different meshing positions, the process is complex, the data volume is huge, and the calculation redundancy exists.

In summary, in the prior art, the fatigue problem of the differential housing of the electric automobile under the action of the forward drive impact load and the reverse drive impact load and the life prediction of the differential housing by low cycle fatigue are not considered.

Disclosure of Invention

In order to solve the problems in the prior art, the invention provides a method for calculating the impact fatigue life of an electric automobile differential shell, which is used for realizing the fatigue problem of the electric automobile differential shell under the action of forward drive and reverse drive impact load and the life prediction of the differential shell by low cycle fatigue.

The invention aims at realizing the following technical scheme:

the method for calculating the impact fatigue life of the differential case of the electric automobile comprises the following steps:

s1, building an assembled finite element model of a differential shell:

the differential assembly comprises a differential gear, a differential housing, a straight axle, a planetary gear, a half axle gear, a pin shaft, a gasket, a bolt group and a bearing, wherein the differential gear, the differential housing, the straight axle, the planetary gear, the half axle gear, the pin shaft, the gasket, the bolt group and the bearing are respectively subjected to solid grid division, and interaction surfaces among parts with contact relation are subjected to grid division according to one-to-one correspondence of contact area grids; the contact conditions are defined in terms of the relationship of the mutual contact between the parts and they are assembled together.

S2, setting boundary conditions of the finite element model:

the boundary condition of the finite element model comprises two parts, namely fixing the outer sides of the outer rings of the bearings at two ends of the differential so as to simulate the supporting effect of the speed reducer shell on the differential bearing; secondly, the rotation freedom degree of the half shaft gear is restrained so as to simulate the action of the half shaft on the supporting reaction force of the half shaft gear;

s3, defining finite element model material properties:

defining the elastic modulus E and Poisson's ratio mu of the finite element model material of each part;

s4, determining the positions and the number of loading points according to the structural symmetry of the differential shell:

the differential shell is used as a rotary part, symmetrically designed in the circumferential direction, and the number of loading positions is determined to be N/(n+1) according to the number N of symmetry axes, wherein N is the number of uniformly distributed loading positions in the circumferential direction; establishing RBE3 units at corresponding differential gear pitch circles according to the number of loading positions;

s5, applying finite element model calculation load comprising bolt pretightening force and engaging force load at each loading position:

the finite element model load comprises two types, wherein the first type is bolt pretightening force, the bolt pretightening force is obtained by calculation through a formula (1), and the acting direction is along the axial direction of the bolt;

wherein F is the bolt pretightening force, T is the bolt tightening torque, k is the bolt tightening torque coefficient, and D is the bolt diameter;

The second type is the maximum impact torque M born by the differential mechanism, comprising the positive drive impact M _f And a counter-drive impact torque M _r The maximum impact torque is decomposed into circumferential force, radial force and axial force by adopting a formula (2), the circumferential force, the radial force and the axial force are applied by means of a partial cylindrical coordinate system defined on the rotation axis of the differential, the Z axis of the coordinate system is along the axis direction of the differential, R is along the radial direction of the differential, and t is determined by Z, R according to the right-hand rule; the impact torque is converted into gear meshing force according to gear meshing parameters and a gear load calculation formula, and then the gear meshing force is applied to each loading position determined in the step S4;

wherein F is _t 、F _r 、F _a Respectively, circumferential force, radial force and axial force of differential gear, d is the pitch diameter of the gear, a _n Is the normal pressure angle of the gear, beta is the helix angle of the pitch circle of the gear, M is the impact torque born by the differential gear, and comprises positive drive impact M _f And a counter-drive impact torque M _r ；

S6, setting working conditions of positive driving impact and negative driving impact, wherein different load steps under the working conditions correspond to loading positions one by one:

the positive driving impact working conditions comprise: all boundary conditions in step S2; all bolts pretension in step S5 and impact torque M by positive drive _f Calculating gear meshing force;

the back-driving impact working conditions comprise: all boundary conditions in step S2; all bolt pretension in step S5 and impact torque M by counter drive _r Calculating gear meshing force;

s7, calculating and analyzing a differential case finite element model:

outputting a stress field, a displacement field and a contact pressure after the calculation of each load step set in the step S6 is completed, so as to obtain a differential case stress field when the gear engagement force acts on different loading positions; the displacement field calculation result and the contact pressure calculation result are used for judging and verifying the validity of the finite element model;

s8, preliminary analysis of stress field calculation results of the differential case:

extracting the maximum stress calculation results of the nodes of the differential case under the forward driving impact working condition and the reverse driving impact working condition in the step S6, comparing the maximum stress calculation results of the nodes of each load step, determining whether the stress field in the corresponding load step exceeds the material yield limit when the differential case generates the maximum stress, executing the step S9 if the stress field exceeds the material yield limit, marking the execution result of the step S9 as the adopted calculation result, otherwise, directly jumping to the step S10, and marking the existing result in the step S8 as the adopted calculation result;

s9, adding plastic properties of the differential case, and carrying out elastoplastic finite element analysis on an assembly model of the differential case:

and modifying the differential shell material property data and the result output setting in the finite element calculation model corresponding to the impact working condition which is determined in the step S8 and needs to be executed in the step S9: material property data based on the material properties described in step S3, increasing material plasticity data; in the result output setting, a strain output option is newly added; after finishing modification, carrying out elastoplastic finite element analysis on the differential housing assembly model;

S10, determining peaks and troughs required by impact fatigue damage calculation according to the adopted calculation result:

determining peaks and troughs when impact fatigue occurs at load location I, J, K: when the loading position is I, J, K, the wave crests are all calculation results adopted by the positive driving impact working condition, and the wave troughs are all calculation results adopted by the negative driving impact working condition;

s11, calculating an impact fatigue damage value when the loading position is I, J, K:

determining a fatigue damage value calculation method according to the adopted calculation result, and if the adopted calculation result comes from the step S9, adopting a low-cycle fatigue calculation method in the fatigue damage calculation method; if the adopted calculation result comes from the step S8, the fatigue damage calculation method adopts a high-cycle fatigue calculation method; according to the corresponding fatigue damage calculation method, calculating impact fatigue damage values at I, J, K positions, and respectively marking as Di, dj and Dk, wherein wave crests and wave troughs in fatigue damage calculation at each loading position are determined in the step S10;

s12, impact fatigue life calculation:

the fatigue life L of the impact load at the loading position I, J, K is calculated by adopting the formula (11) _i 、L _j 、L _k When min { L _i 、L _j 、L _k When the impact fatigue performance of the differential case structure is more than the allowable design life of the impact fatigue corresponding to the design mileage of the whole vehicle, the impact fatigue performance of the differential case structure meets the design requirement, otherwise, the differential case design should be optimized;

in the formula (11), L is the whole vehicle life mileage corresponding to the impact fatigue life of the differential, S is the alternating load cycle number n in the step S11 ₁ Or n ₂ Corresponding whole vehicle life mileage D is the impact fatigue damage determined in the step S11, and subscripts i, j and k respectively represent that the impact load occurrence position is a loading position I, J, K.

Further, in the step S2, the outer rings of the bearings at the two ends of the differential mechanism are required to be fixed by means of the RBE3 unit, the main points of the RBE3 unit select the outer side surface nodes of the outer rings of the two ends of the differential mechanism, and the corresponding points in the middle positions of the widths of the outer rings on the axes of the outer rings of the bearings are selected from the points; the rotation of the side gear is fixed by means of a RBE3 unit, the principal point of the RBE3 unit selecting a node on the face of the internal spline teeth of the side gear, and the point from which to select a point on the centerline of the differential housing.

Further, in the step S4, N/(n+1) loading positions are loading points uniformly distributed on a pitch circle in the circumferential direction of the differential gear corresponding to the minimum asymmetric differential housing structure divided by N symmetry axes; the pitch circle corresponding to the loading position is determined by the differential gear and the gear position meshed with the differential gear; in the step S4, N is an even number not less than 12 and not more than 20; the RBE3 unit established at the pitch circle of the differential gear has a main point of selecting a tooth surface node of not less than 2 teeth near the loading position, and a corresponding loading position is selected from the points.

Further, in the two calculation conditions in the step S6, gear meshing forces are applied to the 1 st to N/(n+1) th loading positions determined in the step S4, respectively; in the two calculation conditions in step S6, the gear meshing forces at the loading positions are respectively defined to different loading steps.

Further, in the step S8, a rule for determining whether the stress field has exceeded the material yield limit is: extracting the stress values of 3 nodes of the maximum stress, the second maximum stress and the third maximum stress in the load step, and comparing the stress values with the yield limit respectively, wherein if the stress values of 3 nodes are all largeAt the material yield limit sigma _s It is determined that the stress field has exceeded the material yield limit.

Further, in step S8, the nodes corresponding to the maximum stress, the second maximum stress and the third maximum stress of the differential case are not nodes included in the contact area under the positive driving impact working condition and the negative driving impact working condition.

Further, in the step S9, the material plastic property data is composed of true stress σ and plastic strain ε ^pl The composition is formed by converting the nominal stress and the nominal strain of the differential housing material according to the formulas (3) to (6) and giving the differential housing finite element model;

σ＝σ _nom (1+ε _nom ) (3)

ε＝ln(1+ε _nom ) (4)

ε ^pl ＝ε-ε ^el (5)

In formulas (3) to (6): sigma is true stress, epsilon is true strain, sigma _nom For nominal stress, ε _nom For nominal strain, ε ^pl For plastic strain, epsilon ^el Is elastic strain;

when the true stress sigma of the material is equal to the limit of the yield strength of the material, the plastic strain epsilon ^pl Is not 0; when plastic strain epsilon of material ^pl Less than 1X 10 ^-5 When the value is directly taken as 0; when plastic strain epsilon ^pl When the material elongation is exceeded, the true stress sigma is equal to the true stress corresponding to the material elongation.

Further, in the step S10, the calculation result adopted is defined as the stress field calculation result in the step S8 if the step S9 is not executed, and the stress field and strain field calculation result after the step S9 is executed if the step S9 is executed.

Further, in the step S10, the method for determining the loading position of I, J, K is as follows: the loading step where the maximum stress value is located in the normal driving impact working condition is taken as the loading position I; the loading step where the stress maximum value is located in the counter-drive impact working condition is taken as a loading position J; and calculating stress amplitude values under each load step by taking each load step stress field under the positive driving impact working condition as a wave crest stress field and taking each load step stress field under the negative driving impact working condition as a wave trough stress field, and recording a loading position corresponding to the load step where the maximum value of the stress amplitude is positioned as a loading position K.

Further, in the step S11, the low cycle fatigue calculating method includes: calculating by adopting a Coffin-Manson equation (7) and a linear Palmgren-Miner rule formula (8);

in formulas (7) to (8): delta epsilon is the total strain range; e is the elastic modulus, which is the same as that in step S2; sigma'. _f Is the fatigue strength coefficient; b is the fatigue strength index; epsilon' _f Is the fatigue ductility coefficient; c is the fatigue ductility index; d is low-axis fatigue damage; n is n ₁ Is the number of alternating load cycles; 2N _f Cycle number on the material strain-life curve;

the high cycle fatigue calculation method comprises the following steps: calculating by adopting a power function formula (9) and a linear Palmgren-Miner rule formula (10);

Δσ ^m N ₂ ＝C (9)

in formulas (9) to (10): Δσ is the stress amplitude; m and C are constants related to material, stress ratio; d is high-cycle fatigue damage; n is n ₂ The number of cycles of the stress amplitude delta sigma; n (N) ₂ Is the number of cycles on the stress-life curve.

The invention has the following beneficial effects:

(1) the load loading position is determined according to the practical condition of symmetrical structure of the differential mechanism shell, in the calculation and analysis process, the influence of different engagement positions on the stress condition of the differential mechanism shell in the rotation process of the differential mechanism is considered, and the redundancy condition of rapid increase of calculated amount and data amount caused by uniformly distributing a large number of loading positions in the circumferential direction is avoided, so that the differential mechanism has the characteristics of simplicity and high efficiency.

(2) The finite element model required by the calculation of the differential mechanism assembly is built by establishing the finite element model, whether elastoplastic calculation is required or not and whether the damage adopts a high-cycle fatigue or low-cycle fatigue calculation method are judged and selected by means of the linear elastic calculation stress result, so that the problems of difficult convergence of the finite element and long calculation period caused by directly adopting elastoplastic calculation are avoided, the damage calculation method is determined, the damage calculation is carried out on the differential mechanism shell by reasonably selecting the high-cycle fatigue calculation method and the low-cycle fatigue calculation method, the rationality of performance prediction is ensured, and the problem of performance prediction difference caused by different engineers with different method selection standards is avoided.

(3) The method comprises the steps of establishing a finite element model of the differential housing assembly, carrying out line elasticity or elastoplastic stress calculation, determining to adopt a high-cycle or low-cycle fatigue damage calculation method, carrying out impact fatigue damage calculation according to stress amplitude values of the differential housing under each loading position when loads such as forward drive impact, reverse drive impact and the like act, further predicting impact fatigue life, and carrying out calculation method standardization and unification.

Drawings

The invention is described in detail below with reference to the drawings and the detailed description.

FIG. 1 is a schematic diagram of a differential assembly;

FIG. 2 is a schematic illustration of RBE3 units defined for applying boundary conditions at bearings at both ends of a differential;

FIG. 3 is a schematic representation of a RBE3 unit defined for applying boundary conditions at a differential side gear;

FIG. 4 is a schematic view of RBE3 established at each load position and determined load position based on differential housing structural symmetry

FIG. 5 is a schematic illustration of a differential finite element model applying loads at various loading locations;

FIG. 6 is a schematic diagram of Mises stress field, displacement field, contact pressure distribution at a positive drive impact condition loading location L1;

FIG. 7 is a schematic diagram of the Mises stress field of the differential housing in an elastoplastic state corresponding to the forward drive impact condition loading position L1;

FIG. 8 is a schematic diagram of the differential case impact fatigue damage distribution corresponding to load position L1 and load position L4;

fig. 9 is a schematic flow chart of an impact fatigue life calculating method of an electric automobile differential case according to an example of the invention.

Detailed Description

The invention is described in further detail below with reference to the attached drawings and examples:

the method for calculating the impact fatigue life of the differential shell of the electric automobile comprises the following steps:

S1, building an assembled finite element model of a differential case

The differential mechanism assembly comprises a differential mechanism gear, a differential mechanism shell, a straight axle, a planetary gear, a half axle gear, a pin shaft, a gasket, a bolt group and a bearing, wherein the differential mechanism gear, the differential mechanism shell, the straight axle, the planetary gear, the half axle gear, the pin shaft, the gasket, the bolt group and the bearing are respectively subjected to solid meshing, an interaction surface between parts with contact relation is meshed according to one-to-one correspondence of contact area meshes, so that model convergence debugging period caused by the contact relation is reduced, and the calculation efficiency is improved. The contact conditions are defined in terms of the relationship of the mutual contact between the parts and they are assembled together.

S2, setting boundary conditions of the finite element model

The boundary condition of the finite element model comprises two parts, namely fixing the outer sides of the outer rings of the bearings at two ends of the differential so as to simulate the supporting effect of the speed reducer shell on the differential bearing; secondly, the rotation freedom degree of the half shaft gear is restrained so as to simulate the action of the half shaft on the supporting reaction force of the half shaft gear;

further, in step S2, the outer rings of the bearings at the two ends of the differential mechanism are required to be fixed by means of RBE3 units, the main points of the RBE3 units select the outer surface nodes of the outer rings at the two ends of the differential mechanism, and the points at the corresponding middle positions of the widths of the outer rings on the axes of the outer rings of the bearings are selected from the points.

Further, in the step S2, the rotation of the side gear is fixed by means of the RBE3 unit, and the main point of the RBE3 unit selects a node on the surface of the internal spline teeth of the side gear, and a point on the center line of the differential case is selected from the points.

S3, defining finite element model material properties

Defining the elastic modulus E and Poisson's ratio mu of the finite element model material of each part;

s4, determining the positions and the number of loading points according to the structural symmetry of the differential mechanism shell

The differential shell is used as a rotary part, and is designed symmetrically in the circumferential direction, and the number of loading positions is determined to be N/(n+1) according to the number N of symmetry axes, wherein N is the number of uniformly distributed loading positions in the circumferential direction. Establishing RBE3 units at corresponding differential gear pitch circles by the number of loading positions to facilitate application of loads to the differential housing finite element model;

further, in the step S4, N/(n+1) loading positions are loading points uniformly distributed on the pitch circle in the circumferential direction of the differential gear corresponding to the minimum asymmetric differential housing structure divided by N symmetry axes.

Further, the pitch circle corresponding to the loaded position is determined by the differential gear, and the gear position meshed with the gear.

Further, in the step S4, N is an even number not less than 12 and not more than 20.

Further, a RBE3 unit established at the pitch circle of the differential gear has a main point selecting a tooth surface node of not less than 2 teeth near the loading position, and a corresponding loading position is selected from the points.

S5, applying finite element model calculation load comprising bolt pretightening force and engaging force load at each loading position

The finite element model load comprises two types, namely bolt pretightening force which is obtained by a relation between the diameter of a bolt and the tightening torque of the bolt and then is applied to the bolt; second, the maximum impact torque M borne by the differential mechanism comprises a positive drive impact M _f And a counter-drive impact torque M _r The impact torque is converted into a gear mesh force in accordance with the gear mesh parameter and the gear load calculation formula, and then applied to each loading position determined in S4.

In the step S5, the bolt pretightening force is calculated by adopting a formula (1), and the acting direction is along the axial direction of the bolt; the gear meshing force at each loading position is decomposed into circumferential force, radial force and axial force by the maximum impact torque by adopting the formula (2), the circumferential force, the radial force and the axial force are applied by means of a partial cylindrical coordinate system defined on the rotation axis of the differential, the Z axis of the coordinate system is along the direction of the axis of the differential, R is along the radial direction of the differential, and t is determined by Z, R according to the right-hand criterion;

Wherein F is the bolt pretightening force, T is the bolt tightening torque, k is the bolt tightening torque coefficient, and D is the bolt diameter.

Wherein F is _t 、F _r 、F _a Respectively, circumferential force, radial force and axial force of differential gear, d is the pitch diameter of the gear, a _n Is the normal pressure angle of the gear, beta is the helix angle of the pitch circle of the gear, M is the impact torque born by the differential gear, and comprises positive drive impact M _f And a counter-drive impact torque M _r ；

S6, setting working conditions of positive driving impact and negative driving impact, wherein different load steps under each working condition correspond to loading positions one by one

Positive drive impact condition: including all boundary conditions in step S2, all bolt pretension in step S5, and the impact torque M by the positive drive _f Calculating gear meshing force; reverse driving impact working condition: including all boundary conditions in step S2, all bolt pretension in step S5, and impact torque M by counter-drive _r Calculating gear meshing force;

further, in the two calculation conditions in the step S6, the gear meshing force needs to be applied to the 1 st to N/(n+1) th loading positions determined in the step S4 respectively;

further, in the two calculation conditions in step S6, the gear meshing force at each loading position needs to be defined to different loading steps, for example, the meshing force at the 1 st loading position is defined as loading step 1, and the gear meshing force at the 2 nd loading position is defined to loading step 2 … …, so that different stress states of the differential case corresponding to the gear meshing force at different loading positions can be obtained through one calculation and solution.

S7, calculating and analyzing the finite element model of the differential case

And (3) carrying out calculation and analysis on a finite element model of the differential casing, and outputting a stress field, a displacement field, contact pressure and the like after the calculation of each load step set in S6 is completed so as to obtain the stress field of the differential casing when the gear engagement force acts on different loading positions.

And S7, the displacement field calculation result and the contact pressure calculation result are used for judging and verifying the validity of the finite element model.

S8, preliminary analysis of stress field calculation results of differential case

And (3) extracting the maximum stress calculation results of the nodes of the differential case under the forward driving impact working condition and the reverse driving impact working condition in the step S6, comparing the maximum stress calculation results of the nodes of each load step, determining whether the stress field in the corresponding load step exceeds the material yield limit when the differential case generates the maximum stress, executing the step S9 if the stress field exceeds the material yield limit, marking the execution result of the step S9 as the adopted calculation result, otherwise, directly jumping to the step S10, and marking the existing result in the step S8 as the adopted calculation result.

Further, the rule of determination of whether the stress field has exceeded the yield limit of the material is: extracting the stress values of 3 nodes of the maximum stress, the second maximum stress and the third maximum stress in the load step, and comparing the stress values with the yield limit respectively, wherein if the stress values of 3 nodes are all larger than the material yield limit sigma _s It is determined that the stress field has exceeded the material yield limit.

Further, in S8, the nodes corresponding to the maximum stress, the second maximum stress, and the third maximum stress of the differential case are not nodes included in the contact area under the positive driving impact working condition and the negative driving impact working condition.

S9, adding plastic properties of the differential case, and carrying out elastoplastic finite element analysis on an assembly model of the differential case

And (3) modifying the finite element calculation model corresponding to the impact working condition required to execute the step S9 determined in the step S8, wherein the differential case material attribute data and the result output setting in the model are modified: the material property data is added with material plasticity data based on the material property in the step S3; in the result output setting, a strain output option is newly added. And after modification, carrying out elastoplastic finite element analysis on the differential housing assembly model.

Further, the plastic property data of the material consists of the true stress sigma and the plastic strain epsilon ^pl The composition is formed by converting the nominal stress and the nominal strain of the differential housing material according to the formulas (3) to (6) and giving the differential housing finite element model;

σ＝σ _nom (1+ε _nom ) (3)

ε＝ln(1+ε _nom ) (4)

ε ^pl ＝ε-ε ^el (5)

in formulas (3) to (6): sigma is true stress, epsilon is true strain, sigma _nom For nominal stress, ε _nom For nominal strain, ε ^pl For plastic strain, epsilon ^el Is an elastic strain.

Further, when the true stress sigma of the material is equal to the limit of the yield strength of the material, the plastic strain epsilon ^pl Is not 0; when plastic strain epsilon of material ^pl Less than 1X 10 ^-5 When the value is directly taken as 0; when plastic strain epsilon ^pl When the material elongation is exceeded, the true stress sigma is equal to the true stress corresponding to the material elongation.

Further, by completing the elastoplastic finite element analysis of the differential case assembly model performed after modification, the results of the differential case stress field, the strain field, the displacement field, the contact pressure, and the like in the elastoplastic analysis state can be recovered.

S10, determining wave crests and wave troughs required by impact fatigue damage calculation according to the adopted calculation result.

Determining peaks and troughs when impact fatigue occurs at load location I, J, K: when the loading position is I, J, K, the wave peaks are all calculation results adopted by the positive driving impact working condition, and the wave troughs are all calculation results adopted by the negative driving impact working condition.

Further, the calculation result adopted is defined as the stress field calculation result in S8 if S9 is not executed, and the stress field and strain field calculation result after S9 is executed if S9 is executed.

Further, the loading position I, J, K in S10 is determined by: the loading step where the maximum stress value is located in the normal driving impact working condition is taken as the loading position I; the loading step where the stress maximum value is located in the counter-drive impact working condition is taken as a loading position J; and calculating stress amplitude values under each load step by taking each load step stress field under the positive driving impact working condition as a wave crest stress field and taking each load step stress field under the negative driving impact working condition as a wave trough stress field, and recording a loading position corresponding to the load step where the maximum value of the stress amplitude is positioned as a loading position K.

And S11, calculating an impact fatigue damage value when the loading position is I, J, K.

And determining a fatigue damage value calculation method according to the adopted calculation result, namely, if the adopted calculation result is from S9 (namely, the elastoplastic finite element analysis result considering the plastic property of the differential case material), adopting a low-cycle fatigue calculation method for the fatigue damage calculation method, and if the adopted calculation result is from S8 (namely, the linear elastic finite element analysis result), adopting a high-cycle fatigue calculation method for the fatigue damage calculation method. According to the corresponding fatigue damage calculation method, impact fatigue damage values occurring at I, J, K positions are calculated and respectively marked as Di, dj and Dk, and wave crests and wave troughs during fatigue damage calculation at each loading position are determined in S10.

Further, the low cycle fatigue calculation method is as follows: calculating by adopting a Coffin-Manson equation (7) and a linear Palmgren-Miner rule formula (8);

in formulas (7) to (8): delta epsilon is the total strain range; e is the elastic modulus, which is the same as that in step S2; sigma'. _f Is the fatigue strength coefficient; b is the fatigue strength index; epsilon' _f Is the fatigue ductility coefficient; c is the fatigue ductility index; d is low-axis fatigue damage; n is n ₁ Is the number of alternating load cycles; 2N _f Cycle number on the material strain-life curve;

further, the high cycle fatigue calculation method is as follows: calculating by adopting a power function formula (9) and a linear Palmgren-Miner rule formula (10);

Δσ ^m N ₂ ＝C (9)

in formulas (9) to (10): Δσ is the stress amplitude; m and C are constants related to material, stress ratio; d is high-cycle fatigue damage; n is n ₂ The number of cycles of the stress amplitude delta sigma;N ₂ number of cycles on stress-life curve;

further, I, J, K may be the same or different loading locations, i.e. no more than 3 fatigue damage calculations need to be performed.

S12, performing impact fatigue life calculation

The fatigue life L of the impact load at the loading position I, J, K is calculated by adopting the formula (11) _i 、L _j 、L _k When min { L _i 、L _j 、L _k When the impact fatigue performance of the differential case structure is more than the allowable design life of the impact fatigue corresponding to the design mileage of the whole vehicle, the impact fatigue performance of the differential case structure meets the design requirement, otherwise, the differential case design should be optimized;

in the formula (11), L is the whole vehicle life mileage corresponding to the impact fatigue life of the differential, S is the alternating load cycle number n in S11 ₁ (lower week) or n ₂ The corresponding life mileage of the whole vehicle (high cycle), D is the impact fatigue damage determined in S11, and subscripts i, j, k respectively represent the impact load occurrence position as a loading position I, J, K.

Examples

As shown in fig. 9, a method for calculating the impact fatigue life of an electric automobile differential case includes the steps of:

s1, building an assembled finite element model of a differential case

As shown in fig. 1, the differential assembly includes a differential gear 1001, a differential case 2001, a first shaft 3001, a first planetary gear 4001, a second planetary gear 4002, a first half shaft gear 5001, a second half shaft gear 5002, a pin 6001, a first washer 7001, a second washer 7002, a third washer 7003, a fourth washer 7004, a bolt set 8001, a first bearing 9001, and a second bearing 9002, and the differential gear 1001, the differential case 2001, the first shaft 3001, the first planetary gear 4001, the second planetary gear 4002, the first half shaft gear 5001, the second half shaft gear 5002, the pin 6001, the first washer 7001, the second washer 7002, the third washer 7003, the fourth washer 7004, the bolt set 8001, the first bearing 9001, and the second bearing 9002 are subjected to solid meshing, and the interaction surfaces between the parts having contact relationships are meshed in a one-to-one correspondence by one mesh correspondence in terms of contact area meshes, so as to reduce a model convergence period due to the contact relationship, and improve calculation efficiency; defining contact conditions according to the mutual contact relation between parts, and assembling the parts together;

S2, setting boundary conditions of the finite element model

The boundary condition of the finite element model comprises two parts, namely, a first bearing 9001 and a second bearing 9002 at two ends of the differential are fixed on the outer sides of the outer rings, so that the supporting effect of a casing of the speed reducer on the differential bearing is simulated, as shown in fig. 2, a process that an RBE3 unit is required to be built on the outer ring is described by taking the first bearing 9001 as an example, a node of the outer ring of the first bearing 9001 is taken as a main point, a point 9003 corresponding to the middle point of the width of the outer ring of the first bearing 9001 is taken as a slave point, an RBE3 unit 9004 is built, and an RBE3 unit slave point 9005 can be found at the second bearing 9002 at the other end in the same way, and an RBE3 unit 9006 is built. All degrees of freedom are constrained for translation and rotation at the slave points 9003 and 9005 of RBE3 units 9004 and 9006.

Secondly, the rotational degrees of freedom of the first half-shaft gear 5001 and the second half-shaft gear 5002 are restrained, so as to simulate the action of the half shafts on the supporting reaction of the half-shaft gears, as shown in fig. 3, the process of establishing an RBE3 unit at the position of the half-shaft gears is described by taking the first half-shaft gear 5001 as an example, a node on the surface of an internal spline tooth of the first half-shaft gear 5001 is taken as a main point, and a point 5003 corresponding to the midpoint of the axial length of an internal spline of the first half-shaft gear 5001 on the axis of the differential casing 2001 is taken as a slave point of the RBE3 unit, so that an RBE3 unit 5004 is established; similarly, a slave 5005 of the RBE3 unit may be found at side gear number two 5002 and RBE3 unit 5006 is established. At the slave points 5003 and 5005 of the RBE3 units 5004 and 5006, rotational degrees of freedom about the axial direction of the differential case 2001 are imposed.

S3, defining finite element model material properties

Definition differential gear 1001, straight axle 3001, first planetary gear 4001, second planetary gear 4001Planetary gear 4002, side gear 5001, side gear 5002, pin 6001, gasket 7001, gasket 7002, gasket 7003, gasket 7004, bolt set 8001, bearing 9001, bearing 9002, and finite element model material elastic modulus e=210000 MPa, poisson ratio μ=0.3, all of which are wire elastic materials, differential case 2001 finite element model material elastic modulus e=175000 MPa, poisson ratio μ=0.3, yield limit σ _s ＝370MPa。

S4, determining the positions and the number of loading points according to the structural symmetry of the differential mechanism shell

The differential shell is used as a rotary part, and is designed symmetrically in the circumferential direction, and the number of loading positions is determined to be N/(n+1) according to the number N of symmetry axes, wherein N is the number of uniformly distributed loading positions in the circumferential direction. As shown in fig. 4, in the present embodiment, the differential case 2001 has a symmetry axis 2002 of n=1 in the circumferential direction, and the symmetry axis divides the differential case into two asymmetric structures. Taking an even number of N being not less than 12 and not more than 20 as 14, the number of loading positions can be determined to be 7.

Differential gear 1001 and the gears that mesh with gear 1001 together define pitch circle 1002. The 7 loading positions L1-L7 are evenly distributed on the pitch circle 1002 corresponding to the upper half of the differential case 2001, and the tooth width direction is the middle point of the tooth width, as shown in fig. 4. RBE3 units are respectively built at loading positions L1-L7 for simulating load loading during gear engagement, the main point of the RBE3 units selects tooth surface nodes of not less than 2 teeth near the loading positions, and the corresponding loading positions are selected from the points, as shown by RBE3 unit 1003 in FIG. 4, and the RBE3 units at the loading positions are built in the same way.

S5, applying finite element model calculation load comprising bolt pretightening force and engaging force load at each loading position

The finite element model load comprises two types, namely a bolt pretightening force F, which is calculated by a relation formula (1) between the bolt diameter D of a bolt group 8001 and a bolt tightening torque T, wherein the bolt tightening torque coefficient is 0.18, and the acting direction of the bolt pretightening force is applied to a bolt along the axial direction of the bolt group 8001;

wherein F is the bolt pretightening force, T is the bolt tightening torque, k is the bolt tightening torque coefficient which takes a value of 0.18, and D is the bolt diameter.

The maximum impact torque M born by the differential mechanism under the second type of load comprises a positive drive impact M _f =5000 Nm and counter-drive impact torque M _r The impact torque is decomposed into circumferential, radial and axial forces by the gear engagement parameters, according to equation (2), and then applied to RBE3 unit slave points at loading positions L1-L7 determined in S4, which are applied by means of a partial cylindrical coordinate system defined on the differential rotation axis, the Z axis of the coordinate system being along the differential axis direction, R being along the differential radial direction, t being determined by Z, R according to the right hand criterion, =5000 Nm;

wherein F is _t 、F _r 、F _a Respectively, circumferential force, radial force and axial force of differential gear, d is the pitch diameter of the gear, a _n Is the normal pressure angle of the gear, beta is the helix angle of the pitch circle of the gear, M is the impact torque born by the differential gear, and comprises positive drive impact M _f =5000 Nm and counter-drive impact torque M _r ＝5000Nm；

Further, for convenience of description and understanding, in this embodiment, the corresponding relationship between the loading position, the meshing force load and the impact torque is shown in table 1, and the loaded finite element model schematic diagram is shown in fig. 5.

TABLE 1 Gear mesh force relationship table

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S6, setting working conditions of positive driving impact and negative driving impact, wherein different load steps under each working condition correspond to loading positions one by one

Positive drive impact condition: including all boundary conditions in step S2, all bolt pretension in step S5, and the impact torque M by the positive drive _f Calculated gear mesh force F _1-f To F _7-f The method comprises the steps of carrying out a first treatment on the surface of the Reverse driving impact working condition: including all boundary conditions in step S2, all bolt pretension in step S5, and impact torque M by counter-drive _r Calculated gear mesh force F _1-r To F _7-r ；

Further, in the two calculation conditions in step S6, the gear meshing force at each loading position needs to be defined to different loading steps, for example, the meshing force at the 1 st loading position is defined as loading step 1, and the gear meshing force at the 2 nd loading position is defined to loading step 2 … …, so that different stress states of the differential case corresponding to the gear meshing force at different loading positions can be obtained through one calculation and solution. In this embodiment, the boundary conditions, pretightening force loads, impact loads and other conditions corresponding to the load steps of the forward driving impact working condition and the reverse driving impact working condition are shown in table 2 in detail.

TABLE 2 boundary and load conditions in Forward and reverse impact conditions

S7, calculating and analyzing the finite element model of the differential case

And (3) performing calculation and analysis on a finite element model of the differential casing, respectively calculating a forward driving working condition and a reverse driving working condition set in the step S6, and outputting a stress field, a displacement field, contact pressure and the like after each load step calculation is completed so as to obtain the stress field of the differential casing when the gear engagement force acts on different loading positions.

In the step S7, the displacement field calculation result and the contact pressure calculation result are used to determine and verify the validity of the finite element model, taking the calculation result of the loading position of the positive driving impact working condition at the L1 as an example, the Mises stress field, the total displacement field and the contact pressure CPRESS distribution are shown in fig. 6.

S8, preliminary analysis of stress field calculation results of differential case

And (3) extracting the maximum Mises stress of the nodes in the load steps of the differential shell under the forward driving impact working condition and the reverse driving impact working condition in S6, filling the calculated result into a table 3, and determining that the load steps corresponding to the maximum stress of the differential shell are the load step 1 of the forward driving impact working condition and the load step 4 of the reverse driving impact working condition respectively. Further aiming at whether the stress field of the load step 1 under the normal driving impact working condition and the load step 4 under the reverse driving impact working condition exceeds the material yield limit sigma _s Determination is performed=370 MPa: the maximum stress, the second maximum stress and the third maximum stress in the load step 1 of the positive drive impact working condition are all larger than the material yield limit sigma _s Judging that the stress field in the load step 1 of the positive drive impact working condition exceeds the yield limit of the material, and executing S9; similarly, the stress field in the load step 4 of the counter-drive impact working condition exceeds the yield limit of the material, and the S9 is executed and the execution result of the S9 is recorded as the adopted calculation result.

TABLE 3 differential case Mises stress results (unit MPa)

S9, adding plastic properties of the differential case, and carrying out elastoplastic finite element analysis on an assembly model of the differential case

And (3) modifying the material attribute data and the result output setting of the differential shell 2001 in the finite element calculation model corresponding to the impact working condition required to execute the step S9, namely the load step 1 of the positive driving impact working condition and the load step 4 of the reverse driving impact working condition, which are determined in the step S8: the material property data is added with the material plasticity data based on the material property described in S3, and the material plasticity data is formed by the actual stress sigma and the plastic strain epsilon ^pl A composition which converts the nominal stress, nominal strain of the material of the differential case 2001 according to formulas (3) to (6), and gives a finite element model of the differential case 2001; when the true stress sigma of the differential case 2001 material is equal to the material yield strength limit, the plastic strain epsilon ^pl Is not 0; when the material is plasticSexual strain epsilon ^pl Less than 1X 10 ^-5 When the value is 0, the value is directly taken as 0; when plastic strain epsilon ^pl When the elongation of the material is exceeded, the real stress sigma is equal to the real stress corresponding to the elongation of the material, and the plastic property of the stress sigma is shown in table 4; the method comprises the steps of carrying out a first treatment on the surface of the In the result output setting, a new strain output option is required, after modification is completed, elastoplastic finite element analysis is performed on the differential housing assembly model, and the results of the differential housing stress field, the strain field, the displacement field, the contact pressure and the like in an elastoplastic analysis state are obtained again, as shown in fig. 7, which is an illustration of calculation results of the elastoplastic stress field corresponding to the loading position 1 of the positive driving impact working condition in the embodiment.

σ＝σ _nom (1+ε _nom ) (3)

ε＝ln(1+ε _nom ) (4)

ε ^pl ＝ε-ε ^el (5)

In formulas (3) to (6): sigma is true stress, epsilon is true strain, sigma _nom For nominal stress, ε _nom For nominal strain, ε ^pl For plastic strain, epsilon ^el Is an elastic strain.

TABLE 4 Plastic mechanical Properties of materials

S10, determining wave crests and wave troughs required by impact fatigue damage calculation according to the adopted calculation result.

In this embodiment, the calculation result adopted is the calculation result after having been executed in S9. Determining peaks and troughs when impact fatigue occurs at load location I, J, K: when the loading position is I, J, K, the wave peaks are all calculation results adopted by the positive driving impact working condition, and the wave troughs are all calculation results adopted by the negative driving impact working condition.

Further, in this embodiment: the loading step where the stress maximum value is located in the positive driving impact working condition corresponds to the loading position of I=1; the loading step where the stress maximum value is located in the counter-drive impact working condition is taken, and the corresponding loading position is J=4; and (3) taking each load step stress field under the positive driving impact working condition in S6 as a wave crest stress field, taking each load step stress field under the negative driving impact working condition in S6 as a wave trough stress field, calculating the stress amplitude under each load step, and recording the loading position corresponding to the load step where the maximum stress amplitude is positioned as a loading position K, wherein in the embodiment, K=J=4.

And S11, calculating an impact fatigue damage value when the loading position is I, J, K.

And determining a fatigue damage value calculation method according to the adopted calculation result, namely, if the adopted calculation result is from S9 (namely, the elastoplastic finite element analysis result considering the plastic property of the differential case material), adopting a low-cycle fatigue calculation method for the fatigue damage calculation method, and if the adopted calculation result is from S8 (namely, the linear elastic finite element analysis result), adopting a high-cycle fatigue calculation method for the fatigue damage calculation method. In this embodiment, the adopted calculation result is from S9, a low cycle fatigue calculation method is fixedly adopted, and because j=k=4, only low cycle fatigue damage values at the loading position 1 and the loading position 4, that is, D1 and D4, are calculated, and the peaks corresponding to the low cycle fatigue damage values are respectively the stress field and the strain field at the loading position L1 and the loading position L4 of the positive driving impact working condition, and the troughs are respectively the stress field and the strain field at the loading position L1 and the loading position L4 of the negative driving impact working condition.

Further, the low cycle fatigue calculation method is as follows: calculating by adopting a Coffin-Manson formula (7) and a linear Palmgren-Miner rule formula (8);

in formulas (7) to (8): delta epsilon is the total strain range; sigma'. _f The fatigue strength coefficient is 816.47MPa; b is a fatigue strength index, and the value of b is-0.0909; epsilon' _f The fatigue ductility coefficient is 6.0273; c is a fatigue ductility index, and the value is-1.2987; d is low-axis fatigue damage; n is n ₁ The alternating load cycle number is 10000 times; 2N _f Cycle number on the material strain-life curve;

s12, performing impact fatigue life calculation

The fatigue life L of the impact load at the loading position I, J, K is calculated by adopting the formula (11) _i 、L _j 、L _k When min { L _i 、L _j 、L _k When the impact fatigue performance of the differential case structure is more than the allowable design life of the impact fatigue corresponding to the design mileage of the whole vehicle, the impact fatigue performance of the differential case structure meets the design requirement, otherwise, the differential case design should be optimized;

in the formula (11), L is the whole vehicle life mileage corresponding to the impact fatigue life of the differential, S is the alternating load cycle number n in the step (11) ₁ Corresponding whole vehicle life mileage is 24 ten thousand kilometers, D is the impact fatigue damage determined in S11, and subscripts I, J and K respectively represent the maximum damage values on the differential housing when impact load occurs and loading positions I=1 and J=K=4, and are respectively D ₁ ＝0.228,D ₄ =0.337. From this calculation, L ₁ ＝105,L ₄ 71 means that when the impact fatigue occurs at the loading position L1, the whole vehicle life mileage corresponding to the impact fatigue life of the differential case is 105 km, and when the impact fatigue occurs at the loading position L4, the whole vehicle life mileage corresponding to the impact fatigue life of the differential case is 71 km, the minimum value of the two is 71 km and is greater than 24 km, so that the impact fatigue performance of the differential case meets the design requirement 。

Although embodiments of the present invention have been shown and described, it will be understood by those skilled in the art that various changes, modifications, substitutions and alterations can be made therein without departing from the principles and spirit of the invention, the scope of which is defined in the appended claims and their equivalents.

Claims (10)

1. The impact fatigue life calculation method of the electric automobile differential shell is characterized by comprising the following steps of:

s1, building an assembled finite element model of a differential shell:

the differential assembly comprises a differential gear, a differential housing, a straight axle, a planetary gear, a half axle gear, a pin shaft, a gasket, a bolt group and a bearing, wherein the differential gear, the differential housing, the straight axle, the planetary gear, the half axle gear, the pin shaft, the gasket, the bolt group and the bearing are respectively subjected to solid grid division, and interaction surfaces among parts with contact relation are subjected to grid division according to one-to-one correspondence of contact area grids; the contact conditions are defined in terms of the relationship of the mutual contact between the parts and they are assembled together.

S2, setting boundary conditions of the finite element model:

the boundary condition of the finite element model comprises two parts, namely fixing the outer sides of the outer rings of the bearings at two ends of the differential so as to simulate the supporting effect of the speed reducer shell on the differential bearing; secondly, the rotation freedom degree of the half shaft gear is restrained so as to simulate the action of the half shaft on the supporting reaction force of the half shaft gear;

S3, defining finite element model material properties:

defining the elastic modulus E and Poisson's ratio mu of the finite element model material of each part;

s4, determining the positions and the number of loading points according to the structural symmetry of the differential shell:

the differential shell is used as a rotary part, symmetrically designed in the circumferential direction, and the number of loading positions is determined to be N/(n+1) according to the number N of symmetry axes, wherein N is the number of uniformly distributed loading positions in the circumferential direction; establishing RBE3 units at corresponding differential gear pitch circles according to the number of loading positions;

s5, applying finite element model calculation load comprising bolt pretightening force and engaging force load at each loading position:

the finite element model load comprises two types, wherein the first type is bolt pretightening force, the bolt pretightening force is obtained by calculation through a formula (1), and the acting direction is along the axial direction of the bolt;

wherein F is the bolt pretightening force, T is the bolt tightening torque, k is the bolt tightening torque coefficient, and D is the bolt diameter;

the second type is the maximum impact torque M born by the differential mechanism, comprising the positive drive impact M _f And a counter-drive impact torque M _r The maximum impact torque is decomposed into circumferential force, radial force and axial force by adopting a formula (2), the circumferential force, the radial force and the axial force are applied by means of a partial cylindrical coordinate system defined on the rotation axis of the differential, the Z axis of the coordinate system is along the axis direction of the differential, R is along the radial direction of the differential, and t is determined by Z, R according to the right-hand rule; the impact torque is converted into gear meshing force according to gear meshing parameters and a gear load calculation formula, and then the gear meshing force is applied to each loading position determined in the step S4;

Wherein F is _t 、F _r 、F _a Respectively, circumferential force, radial force and axial force of differential gear, d is the pitch diameter of the gear, a _n Is the normal pressure angle of the gear, beta is the helix angle of the pitch circle of the gear, M is the impact torque born by the differential gear, and comprises positive drive impact M _f And a counter-drive impact torque M _r ；

S6, setting working conditions of positive driving impact and negative driving impact, wherein different load steps under the working conditions correspond to loading positions one by one:

the positive driving impact working conditions comprise: full in step S2A section boundary condition; all bolts pretension in step S5 and impact torque M by positive drive _f Calculating gear meshing force;

the back-driving impact working conditions comprise: all boundary conditions in step S2; all bolt pretension in step S5 and impact torque M by counter drive _r Calculating gear meshing force;

s7, calculating and analyzing a differential case finite element model:

outputting a stress field, a displacement field and a contact pressure after the calculation of each load step set in the step S6 is completed, so as to obtain a differential case stress field when the gear engagement force acts on different loading positions; the displacement field calculation result and the contact pressure calculation result are used for judging and verifying the validity of the finite element model;

s8, preliminary analysis of stress field calculation results of the differential case:

Extracting the maximum stress calculation results of the nodes of the differential case under the forward driving impact working condition and the reverse driving impact working condition in the step S6, comparing the maximum stress calculation results of the nodes of each load step, determining whether the stress field in the corresponding load step exceeds the material yield limit when the differential case generates the maximum stress, executing the step S9 if the stress field exceeds the material yield limit, marking the execution result of the step S9 as the adopted calculation result, otherwise, directly jumping to the step S10, and marking the existing result in the step S8 as the adopted calculation result;

s9, adding plastic properties of the differential case, and carrying out elastoplastic finite element analysis on an assembly model of the differential case:

and modifying the differential shell material property data and the result output setting in the finite element calculation model corresponding to the impact working condition which is determined in the step S8 and needs to be executed in the step S9: material property data based on the material property in the step S3, increasing material plasticity data; in the result output setting, a strain output option is newly added; after finishing modification, carrying out elastoplastic finite element analysis on the differential housing assembly model;

s10, determining peaks and troughs required by impact fatigue damage calculation according to the adopted calculation result:

Determining peaks and troughs when impact fatigue occurs at load location I, J, K: when the loading position is I, J, K, the wave crests are all calculation results adopted by the positive driving impact working condition, and the wave troughs are all calculation results adopted by the negative driving impact working condition;

s11, calculating an impact fatigue damage value when the loading position is I, J, K:

determining a fatigue damage value calculation method according to the adopted calculation result, and if the adopted calculation result comes from the step S9, adopting a low-cycle fatigue calculation method in the fatigue damage calculation method; if the adopted calculation result comes from the step S8, the fatigue damage calculation method adopts a high-cycle fatigue calculation method; according to the corresponding fatigue damage calculation method, calculating impact fatigue damage values at I, J, K positions, and respectively marking as Di, dj and Dk, wherein wave crests and wave troughs in fatigue damage calculation at each loading position are determined in the step S10;

s12, impact fatigue life calculation:

the fatigue life L of the impact load at the loading position I, J, K is calculated by adopting the formula (11) _i 、L _j 、L _k When min { L _i 、L _j 、L _k When the impact fatigue performance of the differential case structure is more than the allowable design life of the impact fatigue corresponding to the design mileage of the whole vehicle, the impact fatigue performance of the differential case structure meets the design requirement, otherwise, the differential case design should be optimized;

In the formula (11), L is the whole vehicle life mileage corresponding to the impact fatigue life of the differential, S is the alternating load cycle number n in the step S11 ₁ Or n ₂ Corresponding whole vehicle life mileage D is the impact fatigue damage determined in the step S11, and subscripts i, j and k respectively represent that the impact load occurrence position is a loading position I, J, K.

2. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S2, bearing outer rings at two ends of the differential are fixed by means of RBE3 units, the main points of the RBE3 units select nodes on the outer side surfaces of the bearing outer rings at two ends of the differential, and the points at the middle positions of the widths of the corresponding outer rings on the axes of the bearing outer rings are selected from the points; the rotation of the side gear is fixed by means of a RBE3 unit, the principal point of the RBE3 unit selecting a node on the face of the internal spline teeth of the side gear, and the point from which to select a point on the centerline of the differential housing.

3. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S4, N/(n+1) loading positions are loading points uniformly distributed on a pitch circle in the circumferential direction of the differential gear corresponding to a minimum asymmetric differential case structure divided by N symmetry axes; the pitch circle corresponding to the loading position is determined by the differential gear and the gear position meshed with the differential gear; in the step S4, N is an even number not less than 12 and not more than 20; the RBE3 unit established at the pitch circle of the differential gear has a main point of selecting a tooth surface node of not less than 2 teeth near the loading position, and a corresponding loading position is selected from the points.

4. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein, in the two calculation conditions in the step S6, gear meshing forces are applied to the 1 st to N/(n+1) th loading positions determined in the step S4, respectively; in the two calculation conditions in step S6, the gear meshing forces at the loading positions are respectively defined to different loading steps.

5. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S8, a determination rule as to whether the stress field has exceeded the material yield limit is: extracting the stress values of 3 nodes of the maximum stress, the second maximum stress and the third maximum stress in the load step, and comparing the stress values with the yield limit respectively, wherein if the stress values of 3 nodes are all larger than the material yield limit sigma _s Then it is determined that the stress field has exceededMaterial yield limit.

6. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S8, the nodes corresponding to the maximum stress, the second maximum stress and the third maximum stress of the differential case are not nodes included in the contact area under the positive impact condition and the negative impact condition.

7. The method for calculating the impact fatigue life of an electric automobile differential case according to claim 1, wherein in the step S9, the material plasticity attribute data is composed of a true stress σ and a plastic strain ε ^pl The composition is formed by converting the nominal stress and the nominal strain of the differential housing material according to the formulas (3) to (6) and giving the differential housing finite element model;

σ＝σ _nom (1+ε _nom ) (3)

ε＝ln(1+ε _nom ) (4)

ε ^pl ＝ε-ε ^el (5)

in formulas (3) to (6): sigma is true stress, epsilon is true strain, sigma _nom For nominal stress, ε _nom For nominal strain, ε ^pl For plastic strain, epsilon ^el Is elastic strain;

when the true stress sigma of the material is equal to the limit of the yield strength of the material, the plastic strain epsilon ^pl Is not 0; when plastic strain epsilon of material ^pl Less than 1X 10 ^-5 When the value is directly taken as 0; when plastic strain epsilon ^pl When the material elongation is exceeded, the true stress sigma is equal to the true stress corresponding to the material elongation.

8. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S10, the calculation result adopted is defined as the stress field calculation result in the step S8 if the step S9 is not executed, and the stress field and strain field calculation result after the step S9 is executed if the step S9 is executed.

9. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S10, the method for determining the I, J, K loading position is as follows: the loading step where the maximum stress value is located in the normal driving impact working condition is taken as the loading position I; the loading step where the stress maximum value is located in the counter-drive impact working condition is taken as a loading position J; and calculating stress amplitude values under each load step by taking each load step stress field under the positive driving impact working condition as a wave crest stress field and taking each load step stress field under the negative driving impact working condition as a wave trough stress field, and recording a loading position corresponding to the load step where the maximum value of the stress amplitude is positioned as a loading position K.

10. The method for calculating the impact fatigue life of the differential case of the electric automobile according to claim 1, wherein in the step S11, the low cycle fatigue calculating method is as follows: calculating by adopting a Coffin-Manson equation (7) and a linear Palmgren-Miner rule formula (8);

in formulas (7) to (8): delta epsilon is the total strain range; e is the elastic modulus, which is the same as that in step S2; sigma'. _f Is the fatigue strength coefficient; b is the fatigue strength index; epsilon' _f Is the fatigue ductility coefficient; c is the fatigue ductility index; d is low-axis fatigue damage; n is n ₁ Is the number of alternating load cycles; 2N _f For strain-life of materialNumber of cycles on the curve;

the high cycle fatigue calculation method comprises the following steps: calculating by adopting a power function formula (9) and a linear Palmgren-Miner rule formula (10);

Δσ ^m N ₂ ＝C (9)

in formulas (9) to (10): Δσ is the stress amplitude; m and C are constants related to material, stress ratio; d is high-cycle fatigue damage; n is n ₂ The number of cycles of the stress amplitude delta sigma; n (N) ₂ Is the number of cycles on the stress-life curve.