A train axle box bearing life evaluation method considering transient impact load influence

Abstract

The application provides a train axle box bearing life evaluation method considering the influence of transient impact load, comprising the following steps: 1) obtaining physical information contained in bearing quasi-dynamics modeling; 2) constructing an axle box bearing quasi-dynamics model; 3) introducing transient impact load in the axle box bearing external load, and solving the above constructed axle box bearing quasi-dynamics model by using the Runge-Kutta method, and verifying the rationality of the constructed quasi-dynamics model; 4) calculating the real internal load fluctuation of the axle box bearing in each direction under the action of the transient impact load and the corresponding equivalent dynamic load fluctuation; 5) constructing a damage superposition model suitable for time-varying load, and realizing the axle box bearing life evaluation considering the influence of the transient impact load; by introducing the transient impact load in the axle box bearing life evaluation, the actual running process of the train is simulated, and the accuracy of the axle box bearing life evaluation under the actual service condition is improved.

Description

Technical Field

[0001] This invention relates to the field of rolling bearing technology, and in particular to a method for evaluating the life of train axle box bearings that takes into account the effects of transient impact loads. Background Technology

[0002] Axle box bearings are core supporting components in train bogie mechanisms. Their reliability is the foundation for ensuring the high-density, long-term safe operation of train sets. Once axle box bearings fail due to fatigue, it will not only affect the normal operation of the train set and cause huge economic losses, but may also lead to major safety accidents and casualties. Therefore, accurately assessing the bearing life under actual service conditions is of great significance for carrying out targeted operation, maintenance and optimization design of axle box bearings.

[0003] Currently, the rolling bearing industry mainly relies on bearing bench test results to evaluate the life characteristics of axle box bearings. However, such tests can only accurately assess the life characteristics of axle box bearings under steady-state load conditions and cannot reflect the impact of transient impact loads on the life characteristics of axle box bearings. For example, during actual train operation, factors such as track irregularities, wheel flats, and debris on the track surface can cause axle box bearings to be subjected to transient impact loads, thereby reducing the actual service life of axle box bearings and threatening the safety of train operation. Therefore, there is an urgent need for a reasonable method for evaluating the life of train axle box bearings that takes into account the impact of transient impact loads. Summary of the Invention

[0004] The purpose of this invention is to address the shortcomings of existing technologies by providing a method for assessing the life of train axle box bearings that takes into account the influence of transient impact loads. By introducing transient impact loads into the life assessment of axle box bearings, the actual operation process of the train is simulated, thereby improving the accuracy of axle box bearing life assessment under actual service conditions.

[0005] To achieve the above objectives, the present invention provides the following technical solution:

[0006] A method for assessing the life of train axle box bearings that considers the effects of transient impact loads includes the following steps:

[0007] 1) Obtain the physical information contained in the bearing pseudo-dynamic modeling through physical measurement and consultation of sample manuals, including the geometric parameters of rollers and races, the material parameters of rollers and races, and the lubricant material parameters;

[0008] 2) Taking into account the roller type of the bearing, the contact angle between the roller and the inner and outer raceways, the angular position of the roller inside the bearing, the material properties of the roller and raceway, and the material properties of the lubricant, a three-degree-of-freedom vibration differential equation system is established.

[0009] 3) Calculate the combined stiffness and combined damping of a single roller in contact with the inner and outer rings, construct the mechanical expressions for the internal stiffness and internal damping of the bearing, and construct the pseudo-dynamic model of the bearing box by combining them with the three-degree-of-freedom vibration differential equations.

[0010] 4) Introduce transient impact loads into the external loads of the axle box bearings, and use the Runge-Kutta method to solve the pseudo-dynamic model of the axle box bearings constructed above, and extract the bearing time-domain vibration acceleration signal and its vibration signal envelope spectrum from it;

[0011] 5) Based on the vibration displacement of the bearing ring at each time step in the vibration acceleration signal of the axle box bearing, calculate the actual internal load fluctuation of the axle box bearing in each direction under transient impact load and the corresponding equivalent dynamic load fluctuation.

[0012] 6) Construct a damage superposition model for axle box bearings suitable for time-varying loads, and combine it with time-varying equivalent dynamic load data of axle box bearings to realize the life assessment of axle box bearings considering the influence of transient impact loads.

[0013] As a preferred embodiment, in step 4), the spectral components of the vibration acceleration signal of the axle box bearing affected by the transient impact load are obtained by envelope spectrum analysis; if the excitation response frequency in the vibration acceleration signal of the axle box bearing is the same as or a multiple of the frequency of the transient impact load, the rationality of the constructed pseudo-dynamic model can be verified.

[0014] As a preferred embodiment, in step 2), the contact between the roller and the inner and outer raceways is considered as a spring-damping system, and a set of three-degree-of-freedom vibration differential equations for the outer ring of the bearing box is established:

[0015]

[0016] In the formula:

[0017] m o For the mass of the bearing outer ring; k x ,k y ,k z c is the internal stiffness coefficient of the bearing. x ,c y ,c z F is the internal damping coefficient of the bearing. x ,F y ,F z x, y, z represent the external load on the outer ring of the roller; x, y, z represent the vibration displacement of the outer ring of the bearing. This represents the vibration velocity of the outer ring of the bearing. This represents the vibration acceleration of the outer ring of the bearing.

[0018] As a preferred embodiment, in step 2), based on the Hertz linear contact elasticity theory model, the elastic deformation-contact load mapping relationship between the roller and the inner and outer raceways, as shown in equation (2), is obtained, that is, the nonlinear stiffness of the contact between the roller and the inner and outer raceways:

[0019]

[0020] In the formula:

[0021] δ i,o E represents the deformation caused by the contact between the roller and the inner and outer rings. i,o E represents the elastic modulus of the inner and outer rings. r μ is the elastic modulus of the roller. i,o Poisson's ratio for the inner and outer loops; μ r Q is the Poisson's ratio for rollers; i,o The contact load between the roller and the inner and outer rings is l; the effective length of the roller is R. r R is the average radius of the roller; i R is the average radius of the inner raceway. o K is the average radius of the outer raceway. i,o The nonlinear stiffness of the contact between the roller and the inner and outer rings.

[0022] As a preferred embodiment, in step 2), based on the two-dimensional Reynolds equation, the mapping relationship between the normal phase oil film extrusion rate and the contact load between the roller and the inner and outer raceways, as shown in equation (3), is obtained, that is, the oil film damping of the roller in contact with the inner and outer raceways:

[0023]

[0024] In the formula:

[0025] R r R is the average radius of the roller; i R is the average radius of the inner raceway. o η0 is the average radius of the outer raceway; l is the effective length of the roller; η0 is the viscosity of the lubricant under atmospheric conditions; h c The oil film thickness at the center of the roller-raceway contact pair follows the Dowson-Higginson oil film thickness expression; v in,on C represents the rolling-raceway phase oil film extrusion rate. i,o This refers to the oil film damping in contact between the rollers and the inner and outer raceways.

[0026] As a preferred embodiment, in step 3), the contact stiffness between the roller and the inner and outer raceways and the oil film damping are considered to be in series, thus obtaining the bearing stiffness-damping system model.

[0027] As a preferred embodiment, in step 3), based on the contact angle between the roller and the inner and outer raceways and the distribution of the roller inside the bearing, the mechanical expression for the internal stiffness of the bearing as shown in equation (4) and the mechanical expression for the internal damping of the bearing as shown in equation (5) are obtained:

[0028]

[0029]

[0030] By combining equations (1), (4), and (5), a pseudo-dynamic model of the axle box bearing is constructed.

[0031] In the formula:

[0032] m is the number of roller rows in the bearing housing; n r This represents the total number of rollers in each row of the bearing housing. δ represents the position angle of the j-th roller in the m-th column; mjn (x,y,z) represents the component of the vibration displacement of the bearing outer ring in the normal phase of contact between the j-th roller in the m-th column and the outer raceway; v mjn (x,y,z) represents the component of the vibration velocity of the outer ring of the bearing in the normal phase of contact between the j-th roller in the m-th column and the outer raceway.

[0033] As a preferred embodiment, in step 4), based on the solution results of step 3), the vibration displacements x, y, and z of the bearing rings at each time step are extracted, and combined with the angular positions of each roller. Calculate the contact deformation between each roller and the raceway; substitute the above contact deformation into equation (2) to calculate the contact load between each roller and the raceway, and sum them to obtain the true internal load f of the bearing box under transient impact load in each direction. x (t),f y (t),f z (t) Fluctuation:

[0034] The contact deformation between each roller and the raceway is calculated using the following formula:

[0035]

[0036] In the formula:

[0037] P0 is the initial clearance of the axle box bearing;

[0038] The equivalent dynamic load fluctuation corresponding to the actual internal load fluctuation of the axle box bearing is obtained by the following formula:

[0039]

[0040] In the formula:

[0041] P(t) is the time-varying equivalent dynamic load of the axle box bearing; fx (t),f y (t),f z (t) represents the time-varying actual load of the bearing box in all directions; X and Y are the radial dynamic load coefficient and the axial dynamic load coefficient, respectively.

[0042] As a preferred option, in step 6), based on the International Organization for Standardization (ISO) rolling bearing fatigue life model and Palmgren-Miner damage superposition theory, a damage superposition model for axle box bearing suitable for time-varying loads is constructed. Then, combined with the time-varying equivalent dynamic load data of axle box bearings, the life assessment of axle box bearings considering the influence of transient impact loads is realized.

[0043] As a preferred option, the time-varying equivalent dynamic load of the axle box bearing calculated in step 5) is used as the input condition to extract the equivalent dynamic load of the axle box bearing in each time step, and the bearing rotation number corresponding to each time step is calculated according to the operating speed of the axle box bearing.

[0044] Based on the ISO rolling bearing fatigue life model, the basic rated life corresponding to the equivalent dynamic load of the bearing box in each time step is calculated. Then, it is combined with the corresponding bearing speed to calculate the cumulative damage of the bearing box. As the operating speed of the bearing box increases, when the cumulative damage of the bearing box exceeds the damage threshold, the bearing is considered to have reached its life limit.

[0045]

[0046] In the formula:

[0047] a1 is the reliability lifetime correction factor; a ISO [1] is the life correction factor based on the life calculation system method; C is the basic fixed dynamic load / N; [D] is the damage threshold, the value of which is determined by the bearing bench test results or taken as 1 based on experience;

[0048] As the operating speed of the axle box bearing increases, when the accumulated damage to the axle box bearing exceeds the damage threshold, the bearing is considered to have reached its life limit.

[0049] The beneficial effects of this invention are as follows:

[0050] This invention analyzes the influence of transient impact loads on the vibration characteristics of axle box bearings by establishing a pseudo-dynamic model. Then, by combining the Palmgren-Miner damage superposition theory, it evaluates the impact of time-varying equivalent dynamic loads caused by transient impact loads on the life of axle box bearings. Compared with traditional axle box bearing life assessment models that only consider constant static load conditions, the model proposed in this invention can more accurately assess the life characteristics of axle box bearings under actual service conditions, thus providing a theoretical basis for targeted operation, maintenance and optimization design of train axle box bearings. Attached Figure Description

[0051] Figure 1 This is a schematic diagram of the process of the present invention;

[0052] Figure 2 This is a schematic diagram of the rolling bearing spring-damping system model of the present invention;

[0053] Figure 3 This is a time-domain diagram of the vibration signal of the bearing box of the present invention;

[0054] Figure 4 This is the envelope spectrum of the bearing vibration signal of the axle box of the present invention;

[0055] Figure 5 This is a diagram of the time-varying equivalent dynamic load under operating condition 2 in this embodiment of the invention. Detailed Implementation

[0056] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0057] In the description of this invention, it should be understood that the terms "center," "longitudinal," "lateral," "length," "width," "thickness," "upper," "lower," "front," "rear," "left," "right," "vertical," "horizontal," "top," "bottom," "inner," "outer," "clockwise," and "counterclockwise," etc., indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings. They are only for the convenience of describing this invention and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation. Therefore, they should not be construed as limitations on this invention.

[0058] Furthermore, the terms "first" and "second" are used for descriptive purposes only and should not be construed as indicating or implying relative importance or implicitly specifying the number of technical features indicated. Thus, a feature defined as "first" or "second" may explicitly or implicitly include one or more of that feature. In the description of this invention, "a plurality of" means two or more, unless otherwise explicitly specified.

[0059] Example

[0060] like Figure 1-5 As shown, the following example further illustrates the train axle box bearing life assessment method considering the influence of transient impact loads according to the present invention. This example is not intended to limit the present invention. The specific steps are as follows:

[0061] 1) Obtain the physical information included in the bearing pseudo-dynamic modeling through physical measurement and consultation of sample manuals, including the geometric parameters of the rollers and races, the material parameters of the rollers and races, and the lubricant material parameters, as shown in Table 1:

[0062] Table 1. Parameter table of a shaft box bearing for an example.

[0063]

[0064]

[0065] 2) Taking into account the roller type of the bearing housing, the contact angle between the roller and the inner and outer raceways, the angular position of the roller inside the bearing, the material properties of the roller and raceway, and the material properties of the lubricant, a three-degree-of-freedom vibration differential equation system is established; among which,

[0066] 2-1) Consider the contact between the rollers and the inner and outer raceways as Figure 2 The spring-damped system shown is used as an example. A set of three-degree-of-freedom vibration differential equations for the outer ring of the bearing is established:

[0067]

[0068] In the formula:

[0069] m o For the mass of the bearing outer ring; k x ,k y ,k z c is the internal stiffness coefficient of the bearing. x ,c y ,c z F is the internal damping coefficient of the bearing. x ,F y ,F z x, y, z represent the external load on the outer ring of the roller; x, y, z represent the vibration displacement of the outer ring of the bearing. This represents the vibration velocity of the outer ring of the bearing. This refers to the vibration acceleration of the outer ring of the bearing.

[0070] 2-2) Based on the Hertz line contact elasticity theory model, the elastic deformation-contact load mapping relationship between the roller and the inner and outer raceways shown in equation (2) is obtained, that is, the nonlinear stiffness of the contact between the roller and the inner and outer raceways:

[0071]

[0072] In the formula:

[0073] δ i,o E represents the deformation caused by the contact between the roller and the inner and outer rings. i,o E represents the elastic modulus of the inner and outer rings. r μ is the elastic modulus of the roller. i,o Poisson's ratio for the inner and outer loops; μ r Q is the Poisson's ratio for rollers; i,o The contact load between the roller and the inner and outer rings is l; the effective length of the roller is R. r R is the average radius of the roller; i R is the average radius of the inner raceway. o K is the average radius of the outer raceway. i,o The nonlinear stiffness of the contact between the roller and the inner and outer rings;

[0074] In this study, the contact stiffness between the rollers and the inner and outer raceways is considered to be in series, resulting in... Figure 2 The combined stiffness of a single roller in contact with the inner and outer races is shown.

[0075] 2-3) Based on the two-dimensional Reynolds equation, the mapping relationship between the normal phase oil film extrusion rate and the contact load between the roller and the inner and outer raceways, as shown in equation (3), is obtained, that is, the oil film damping of the roller in contact with the inner and outer raceways:

[0076]

[0077] In the formula:

[0078] R r R is the average radius of the roller; i R is the average radius of the inner raceway. o η0 is the average radius of the outer raceway; l is the effective length of the roller; η0 is the viscosity of the lubricant under atmospheric conditions; h c The oil film thickness at the center of the roller-raceway contact pair follows the Dowson-Higginson oil film thickness expression; v in,on C represents the rolling-raceway phase oil film extrusion rate. i,o Oil film damping for contact between the rollers and the inner and outer raceways;

[0079] In this case, the oil film between the roller and the inner and outer raceways is considered to be in series, resulting in... Figure 2 The combined damping shown is the contact between the roller and the inner and outer rings.

[0080] 3) Calculate the combined stiffness and combined damping of a single roller in contact with the inner and outer rings, construct the mechanical expressions for the internal stiffness and internal damping of the bearing, and construct the pseudo-dynamic model of the bearing box by combining them with the three-degree-of-freedom vibration differential equations.

[0081] 3-1) The contact stiffness between the roller and the inner and outer raceways and the oil film damping are considered to be in series, so as to obtain the stiffness-damping system model of the axle box bearing.

[0082] 3-2) Based on the contact angle between the roller and the inner and outer raceways, as well as the angular position of the roller inside the bearing, the mechanical expressions related to the internal stiffness coefficient and damping coefficient of the bearing are obtained, and the results are shown in equations (4) and (5):

[0083]

[0084]

[0085] By combining equations (1), (4), and (5), a pseudo-dynamic model of the axle box bearing is constructed.

[0086] In the formula:

[0087] m is the number of roller rows in the bearing housing; n r This represents the total number of rollers in each row of the bearing housing. δ represents the position angle of the j-th roller in the m-th column; mjn (x,y,z) represents the component of the vibration displacement of the bearing outer ring in the normal phase of contact between the j-th roller in the m-th column and the outer raceway; v mjn (x,y,z) represents the component of the vibration velocity of the outer ring of the bearing in the normal phase of contact between the j-th roller in the m-th column and the outer raceway.

[0088] 4) Introduce transient impact load into the external load of the axle box bearing (i.e., the term on the right side of equation (1)), and solve the aforementioned pseudo-dynamic model of the axle box bearing using the Runge-Kutta method, and extract the bearing's time-domain vibration acceleration signal and its vibration acceleration signal envelope spectrum from it, such as Figure 3 , 4 As shown. By identifying the excitation response component in the time-domain vibration acceleration signal that suddenly increases in amplitude and then oscillates and decays (as shown). Figure 3 ), and the excitation response frequency in the envelope spectrum is the same frequency as the transient impact load, or is a harmonic of the transient impact load frequency ( ), and the frequency of the excitation response in the envelope spectrum is the same as the transient impact load frequency. Figure 4 Then the rationality of the constructed pseudo-dynamic model can be verified.

[0089] 5) Based on the vibration displacement of the bearing ring at each time step in the vibration acceleration signal of the axle box bearing, calculate the actual internal load fluctuation of the axle box bearing in each direction under transient impact load and the corresponding equivalent dynamic load fluctuation.

[0090] 5-1) Based on the solution results of step 4), extract the vibration displacement x, y, z of the bearing ring at each time step, and combine them with the angular position of each roller. Calculate the contact deformation between each roller and the raceway, as shown in equation (6); substitute the above contact deformation into equation (2) to calculate the contact load between each roller and the raceway, and sum them to obtain the fluctuation of the actual internal load fx(t), fy(t), and fz(t) of the bearing box under transient impact load in each direction. Use the fluctuation of the actual internal load of the bearing box to obtain the corresponding equivalent dynamic load fluctuation, as shown in equation (6). Figure 5 As shown;

[0091] The contact deformation between the roller and the raceway is calculated using the following formula:

[0092] In the formula:

[0093] P0 is the initial clearance of the axle box bearing;

[0094] The equivalent dynamic load displacement is calculated using the following formula:

[0095]

[0096] In the formula:

[0097] P(t) is the time-varying equivalent dynamic load of the axle box bearing; f x (t),f y (t),f z (t) represents the time-varying actual load of the bearing box in all directions; X and Y are the radial dynamic load coefficient and the axial dynamic load coefficient, respectively.

[0098] 6) Based on the International Organization for Standardization (ISO) rolling bearing fatigue life model and Palmgren-Miner damage superposition theory, a damage superposition model for axle box bearings suitable for time-varying loads is constructed. Then, combined with the time-varying equivalent dynamic load data of axle box bearings, the life assessment of axle box bearings considering the influence of transient impact loads is realized.

[0099] 6-1) Using the time-varying equivalent dynamic load of the axle box bearing calculated in step 5) as the input condition, extract t1~t n The equivalent dynamic load on the axle box bearing within each time step is calculated, and the bearing rotational speed N(t1) to N(t2) corresponding to each time step is calculated based on the axle box bearing operating speed.n );

[0100] 6-2) Based on the ISO rolling bearing fatigue life model, calculate the basic rated life of the bearing housing under the equivalent dynamic load in each time step, and then correlate it with the corresponding bearing speeds N(t1) to N(t2). n Combined with other methods, calculate the cumulative damage to the axle box bearings:

[0101]

[0102] In the formula:

[0103] a1 is the reliability lifetime correction factor; a ISO [ ] represents the life correction factor based on the life calculation system method; C is the basic fixed dynamic load / N; [D] is the damage threshold, the value of which can be determined by the bearing bench test results or taken as 1 based on experience;

[0104] As the operating speed of the axle box bearing increases, when the accumulated damage to the axle box bearing exceeds the damage threshold, the bearing can be considered to have reached its life limit.

[0105] To verify the accuracy of the method proposed in this patent, three working conditions with and without transient impact loads were selected for case analysis. The specific working conditions and the axle box bearing life results obtained by using the method proposed in this invention under these working conditions are shown in Table 2.

[0106] Table 2. Axle box bearing life assessment results under different working conditions

[0107]

[0108]

[0109] As shown in Table 2 above, compared to operating condition 1, the constant load components of operating conditions 2 and 3 are consistent. However, due to the transient impact loads included in operating conditions 2 and 3, the influence of transient impact loads cannot be accounted for using traditional rolling bearing life assessment methods. Therefore, the bearing life of the axle box under the three operating conditions would be considered to be the same. However, using the method proposed in this invention, it is possible to effectively assess that transient impact loads reduce the fatigue life of the axle box bearing under operating conditions 2 and 3 by 17.7% and 46.2%, respectively. Therefore, the method proposed in this invention is indispensable for accurately assessing the life of axle box bearings. The time-varying equivalent dynamic load diagram under operating condition 2 is shown below. Figure 5 As shown.

[0110] The above description is merely a preferred embodiment of the present invention and is not intended to limit the present invention. Any modifications, equivalent substitutions, and improvements made within the spirit and principles of the present invention should be included within the protection scope of the present invention.

Claims

1. A method for evaluating the life of train axle box bearings considering the influence of transient impact loads, characterized in that, Includes the following steps: 1) Obtain the physical information contained in the bearing pseudo-dynamic modeling, including the geometric parameters of the rollers and races, the material parameters of the rollers and races, and the lubricant material parameters; 2) Taking into account the roller type of the bearing, the contact angle between the roller and the inner and outer raceways, the angular position of the roller inside the bearing, the material properties of the roller and raceway, and the material properties of the lubricant, a three-degree-of-freedom vibration differential equation system is established. 3) Calculate the combined stiffness and combined damping of a single roller in contact with the inner and outer rings, construct the mechanical expressions for the internal stiffness and internal damping of the bearing, and construct the pseudo-dynamic model of the bearing box by combining them with the three-degree-of-freedom vibration differential equations. 4) Introduce transient impact loads into the external loads of the axle box bearings, and use the Runge-Kutta method to solve the pseudo-dynamic model of the axle box bearings constructed above, and extract the bearing time-domain vibration acceleration signal and its vibration signal envelope spectrum from it; 5) Based on the vibration displacement of the bearing ring at each time step in the vibration acceleration signal of the axle box bearing, calculate the actual internal load fluctuation of the axle box bearing in each direction under transient impact load and the corresponding equivalent dynamic load fluctuation. 6) Construct a damage superposition model for axle box bearings suitable for time-varying loads, and combine it with time-varying equivalent dynamic load data of axle box bearings to realize the life assessment of axle box bearings considering the influence of transient impact loads.

2. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 1, characterized in that, In step 4), the rationality of the constructed pseudo-dynamic model can be verified by identifying the excitation response component in the time-domain vibration acceleration signal that suddenly increases in amplitude and then oscillates and decays, as well as the excitation response frequency in the envelope spectrum that is the same frequency as the transient impact load or is a harmonic of the transient impact load frequency.

3. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 1, characterized in that, In step 2), the contact between the roller and the inner and outer raceways is considered as a spring-damping system. Taking the outer ring of the bearing housing as the object, a set of three-degree-of-freedom vibration differential equations for the outer ring of the bearing housing is established: In the formula: m o For the mass of the bearing outer ring; k x ,k y ,k z c is the internal stiffness coefficient of the bearing. x ,c y ,c z F is the internal damping coefficient of the bearing. x ,F y ,F z x, y, z represent the external load on the outer ring of the roller; x, y, z represent the vibration displacement of the outer ring of the bearing. This represents the vibration velocity of the outer ring of the bearing. This represents the vibration acceleration of the outer ring of the bearing.

4. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 3, characterized in that, In step 2), based on the Hertz linear contact elasticity theory model, the elastic deformation-contact load mapping relationship between the roller and the inner and outer raceways, as shown in equation (2), is obtained, which represents the nonlinear stiffness of the contact between the roller and the inner and outer raceways: In the formula: δ i,o E represents the deformation caused by the contact between the roller and the inner and outer rings. i,o E represents the elastic modulus of the inner and outer rings. r μ is the elastic modulus of the roller. i,o Poisson's ratio for the inner and outer loops; μ r Q is the Poisson's ratio for rollers; i,o The contact load between the roller and the inner and outer rings is l; the effective length of the roller is R. r R is the average radius of the roller; i R is the average radius of the inner raceway. o K is the average radius of the outer raceway. i,o The nonlinear stiffness of the contact between the roller and the inner and outer rings.

5. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 4, characterized in that, In step 2), based on the two-dimensional Reynolds equation, the mapping relationship between the normal phase oil film extrusion rate and the contact load between the roller and the inner and outer raceways, as shown in equation (3), is obtained, that is, the oil film damping of the roller in contact with the inner and outer raceways: In the formula: R r R is the average radius of the roller; i R is the average radius of the inner raceway. o η0 is the average radius of the outer raceway; l is the effective length of the roller; η0 is the viscosity of the lubricant under atmospheric conditions; h c The oil film thickness at the center of the roller-raceway contact pair follows the Dowson-Higginson oil film thickness expression; v in,on C represents the rolling-raceway phase oil film extrusion rate. i,o This refers to the oil film damping in contact between the rollers and the inner and outer raceways.

6. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 1, characterized in that, In step 3), the contact stiffness between the roller and the inner and outer raceways and the oil film damping are considered to be in series, thus obtaining the bearing stiffness-damping system model.

7. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 5, characterized in that, In step 3), based on the contact angle between the roller and the inner and outer raceways and the angular position of the roller inside the bearing, mechanical expressions related to the internal stiffness coefficient and damping coefficient of the bearing are obtained, and the results are shown in equations (4) and (5): By combining equations (1), (4), and (5), a pseudo-dynamic model of the axle box bearing is constructed. In the formula: m is the number of roller rows in the bearing housing; n r This represents the total number of rollers in each row of the bearing housing. δ represents the position angle of the j-th roller in the m-th column; mjn (x,y,z) represents the component of the vibration displacement of the bearing outer ring in the normal phase of contact between the j-th roller in the m-th column and the outer raceway; v mjn (x,y,z) represents the component of the vibration velocity of the outer ring of the bearing in the normal phase of contact between the j-th roller in the m-th column and the outer raceway.

8. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 3, characterized in that, In step 4), based on the solution results of step 3), the vibration displacement x, y, z of the bearing ring at each time step are extracted, and combined with the angular position of each roller. Calculate the contact deformation between each roller and the raceway; substitute the above contact deformation into equation (2) to calculate the contact load between each roller and the raceway, and sum them to obtain the true internal load f of the bearing box under transient impact load in each direction. x (t),f y (t),f z (t) Fluctuation; The contact deformation between each roller and the raceway is calculated using the following formula: In the formula: P0 is the initial clearance of the axle box bearing; The equivalent dynamic load fluctuation corresponding to the actual internal load fluctuation of the axle box bearing is obtained by the following formula: In the formula: P(t) is the time-varying equivalent dynamic load of the axle box bearing; f x (t),f y (t),f z (t) represents the time-varying actual load of the bearing box in all directions; X and Y are the radial dynamic load coefficient and the axial dynamic load coefficient, respectively.

9. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 8, characterized in that, In step 6), based on the International Organization for Standardization (ISO) rolling bearing fatigue life model and Palmgren-Miner damage superposition theory, a damage superposition model for axle box bearing suitable for time-varying loads is constructed. Then, combined with the time-varying equivalent dynamic load data of axle box bearings, the life assessment of axle box bearings considering the influence of transient impact loads is realized.

10. The method for evaluating the life of train axle box bearings considering the influence of transient impact loads according to claim 9, characterized in that, Using the time-varying equivalent dynamic load of the axle box bearing calculated in step 5) as input conditions, extract the equivalent dynamic load of the axle box bearing in each time step, and calculate the bearing rotation number corresponding to each time step based on the operating speed of the axle box bearing. Based on the ISO rolling bearing fatigue life model, the basic rated life corresponding to the equivalent dynamic load of the bearing box in each time step is calculated. Then, it is combined with the corresponding bearing speed to calculate the cumulative damage of the bearing box. As the operating speed of the bearing box increases, when the cumulative damage of the bearing box exceeds the damage threshold, the bearing is considered to have reached its life limit.

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