Fatigue life analysis method of wind power main bearing based on thermal-mechanical coupling

Abstract

The invention belongs to the field of failure analysis of bearings, and discloses a fatigue life analysis method of a wind power main bearing based on thermal-mechanical coupling. In order to consider the influence of pre-tightening on the service life of the bearing, the method takes the pre-tightening force of the bearing as a starting point, and gradually establishes a relational expression between pre-tightening and a series of static characteristics such as contact deformation, contact load, fatigue life and the like by utilizing an analytical method; meanwhile, the friction torque is introduced, the finite element analysis method is used for solving the influence of thermal deformation on the static characteristics of the bearing, the influence is substituted into the static characteristic relational expression, and finally the service life analysis method of the bearing based on thermal-mechanical coupling is achieved. Compared with the conventional service life analysis method, the fatigue service life analysis method disclosed by the invention has the advantages that the thermal deformation of the bearing caused by friction is additionally considered, so that the analysis is more comprehensive, and the precision of the service life analysis of the bearing is further improved.

Description

Technical Field

[0001] The present invention belongs to the field of bearing failure analysis and relates to a fatigue life analysis method of a wind power main bearing based on thermal-mechanical coupling. Background Art

[0002] Wind turbine main bearings, as a crucial component of wind turbines, ensure safety and reliability, often crucial for ensuring the lifespan of the generator set. Wind turbine main bearings typically operate at low speeds and under heavy loads. Due to the harsh and changing operating environment and the inherent difficulty in repairing or replacing them, lifespan analysis has become even more critical. Furthermore, with the increasing global demand for energy, resulting in increased wind turbine capacity, premature bearing failure is becoming increasingly common. Therefore, a more comprehensive fatigue life analysis method for wind turbine main bearings is needed.

[0003] Currently, life analysis for wind turbine main bearings focuses on calculating the contact load and its distribution between the roller and raceway using external loads, and from this, calculating stress and life. At the same time, to extend fatigue life, bearings are also appropriately preloaded, so it is necessary to include bearing preload as a prerequisite for bearing fatigue life analysis. Furthermore, because the rollers, under the action of external loads, come into sliding contact with the bearing end faces and ribs, friction torque is generated. This friction torque often causes thermal deformation, which changes the original contact deformation and ultimately leads to changes in the bearing's contact load and life. Therefore, it is also important to include thermal deformation in bearing fatigue life analysis.

[0004] The thermal-mechanical coupling analysis method of the present invention first uses an analytical method to introduce preload as an initial condition on the basis of the existing life analysis. In the process, the friction torque is calculated according to the external load and the thermal deformation is obtained through the finite element method. The new contact load is then substituted into the analytical process and the life is calculated, making the life analysis more comprehensive on the original basis. Summary of the Invention

[0005] In order to make the fatigue life analysis of wind turbine main bearings more comprehensive, the present invention performs a thermal-mechanical coupling analysis of the bearings, which takes into account the effect of appropriate preload on bearing life and the effect of thermal deformation caused by bearing friction on bearing life, thereby making the bearing fatigue life calculation more accurate.

[0006] The technical solution of the present invention:

[0007] A fatigue life analysis method for wind turbine main bearings based on thermal-mechanical coupling, the steps are as follows:

[0008] Basic bearing fatigue life analysis model;

[0009] Step S1: determining the bearing and its structural parameters and the external load and lubricant-related parameters;

[0010] Determine the bearing type, obtain the preload or preload amount, axial and radial load, initial radial and axial clearance, and contact angle, number of rollers and raceway diameter and other parameters from the perspective of mechanical analysis; obtain the speed, viscosity and specific heat capacity of the lubricant and other parameters from the perspective of thermal analysis;

[0011] Step S2: calculating the bearing preload amount according to the type and condition of the preload;

[0012] For bearing preload, it is divided into constant pressure preload and positioning preload;

[0013] For positioning preload, determine the preload amount, calculate the change of the axial preload amount when the radial preload amount is applied, and vice versa, by the relationship between the radial preload amount and the axial preload amount, the relationship is as follows:

[0014]

[0015] Where P 0a is the initial axial clearance, P 0r is the initial radial clearance, P a is the axial preload amount, P r is the radial preload amount, and a is the contact angle;

[0016] For constant pressure preload, specific bearing types need to be measured on site, and the axial and radial preload amount data obtained after the axial and radial preload forces are applied are fitted to obtain the preload amount corresponding to the axial and radial preload forces, and then the influence of the radial preload amount on the axial preload amount and the influence of the axial preload amount on the radial preload amount are obtained;

[0017] Step S3: calculating the contact load and the original contact deformation;

[0018] First, according to the radial load and the axial load and the contact angle, the variable is obtained, where F a is the axial load and F r is the radial load;

[0019] Based on the variable , the radial load integral J r (ε), the axial load integral J a (ε), and the bearing load distribution coefficient ε data fitting, the following are obtained respectively:

[0020] (1) The relationship between the variable and the radial load integral J r (ε) is:

[0021]

[0022] (2) Variables and axial load integral J a (ε) is:

[0023]

[0024] (3) Variables The relationship between it and the load distribution coefficient ε is:

[0025]

[0026] Then calculate the maximum original contact load Q 0max , the relationship is as follows:

[0027]

[0028] Where Z is the number of rollers in the bearing;

[0029] In addition, according to the maximum original contact load Q 0max Obtain the maximum original contact deformation δ 0max , the relationship is as follows:

[0030] Q 0max =K l δ 0max n

[0031] Where n is a constant, which is 1.11; the load displacement coefficient K l Calculate according to the empirical formula, where l is the effective length of the roller;

[0032]

[0033] And through the simultaneous relationship, the original radial and original axial contact deformations δ are obtained: 0r , δ 0a , the relationship is as follows:

[0034] δ 0max =δ 0a sin(α)+δ 0r cos(α)

[0035]

[0036] Step S4: obtaining a new contact load according to the bearing preload and the change in contact deformation caused by thermal deformation;

[0037] After obtaining the bearing preload and the change in contact deformation caused by thermal deformation, they are combined with the original contact deformation obtained in step S3. The relationship is as follows:

[0038] delta a delta 0a P a delta G Ta

[0039] delta r delta 0r P r delta G T

[0040] delta r delta a delta 0r delta 0a P r P a delta G T delta G Ta

[0041] delta G T delta G Ta

[0042]

[0043] delta r delta a

[0044] delta max delta a sin(alpha) r cos(alpha)

[0045] Q max K l delta max n

[0046]

[0047] delta max Q max Q ψ psi

[0048] ​​​Step S5: Calculate fatigue life according to the contact load calculated in S4;

[0049] Solve the fatigue life empirical formula based on Lundberg and Palmgren, where the rated dynamic load Q is required. c , the relationship is as follows:

[0050]

[0051] Where λ is a constant coefficient selected according to the contact type; D is the roller diameter; l is the roller length; γ is obtained by calculation, and the relationship is as follows:

[0052]

[0053] Where D is the roller diameter; d m is the bearing pitch diameter;

[0054] Then the inner raceway dynamic load rating Q is calculated ci And outer raceway dynamic load rating Q co , the formula is as follows:

[0055]

[0056] Then, the equivalent dynamic load Q is obtained based on the fourth root mean square of the new contact load obtained in step S4. e , the relationship between contact load and equivalent dynamic load is as follows:

[0057]

[0058] Among them, Q j is the bearing load distribution Q ψ The contact load of each bearing in , j ranges from 1 to the number of rollers Z;

[0059] For the equivalent dynamic load Q of the inner and outer raceways ei , Q eo Since the contact loads of the inner and outer raceways are almost equal, Q is used. j As the contact load of the inner and outer raceways, used to calculate the equivalent dynamic load Q ei , Q eo ;

[0060] For non-rotating raceways, the square root of the equivalent dynamic load must be four times the Weibull slope e. The specific value must be calculated based on the contact angle α, as shown in the following formula:

[0061] e=1.5tan(α)

[0062] The equivalent dynamic load formula will also become:

[0063]

[0064] Among them, Q ewi With Q ewo That is, the equivalent dynamic load of the inner raceway and the equivalent dynamic load of the outer raceway in the non-rotating state;

[0065] Then use the equivalent dynamic load Q of the inner raceway Ei and rated dynamic load Q ci Calculate the inner raceway life L i , using the equivalent dynamic load Q of the outer raceway Eo and rated dynamic load Q co Calculate the outer raceway life L o , the relationship is as follows:

[0066]

[0067] Among them, the equivalent dynamic load Q of the inner raceway is Ei Including the equivalent dynamic load Q of the inner raceway under normal circumstances ei Equivalent dynamic load Q on the inner raceway in the non-rotating case ewi ;

[0068] Similarly, the equivalent dynamic load Q of the outer raceway is Eo Including the equivalent dynamic load Q of the outer raceway under normal circumstances eo Equivalent dynamic load Q on the outer raceway in the non-rotating case ewo ;

[0069] Finally, the bearing fatigue life L is calculated according to the following relationship:

[0070]

[0071] The fatigue life of the bearing is calculated from this.

[0072] Beneficial effects of the present invention: Compared with previous life analysis methods, the fatigue life analysis method of the present invention takes into account the thermal deformation of the bearing caused by friction, making the analysis more comprehensive and further improving the accuracy of the bearing life analysis. BRIEF DESCRIPTION OF THE DRAWINGS

[0073] Figure 1 This is the flow chart of bearing thermal-mechanical coupling fatigue life calculation. DETAILED DESCRIPTION

[0074] The specific implementation of the present invention is further described below in conjunction with the accompanying drawings and technical solutions.

[0075] like Figure 1 As shown in FIG, a fatigue life analysis method for wind turbine main bearings based on thermal-mechanical coupling is shown in FIG. The steps are as follows:

[0076] (1) Bearing basic fatigue life analysis model;

[0077] Step S1: determining the bearing and its structural parameters as well as external load and lubricant related parameters;

[0078] Determine the bearing type and obtain parameters such as preload or preload amount, axial and radial loads, initial radial and axial clearances, contact angles, number of rollers, and raceway diameter from a mechanical analysis perspective; obtain parameters such as speed, lubricant viscosity, and specific heat capacity from a thermal analysis perspective;

[0079] Step S2: Calculate the bearing preload amount according to the type and condition of preload;

[0080] For bearing preload, it is divided into constant pressure preload and positioning preload;

[0081] For positioning preload, determine the preload amount. By using the relationship between radial preload and axial preload, calculate the change in axial preload when radial preload is applied. Conversely, calculate the change in radial preload when axial preload is applied. The relationship is as follows:

[0082]

[0083] Among them, P 0a is the initial axial clearance, P 0r is the initial radial clearance, P a is the axial preload, P r is the radial preload, α is the contact angle;

[0084] For constant pressure preload, it is necessary to conduct field data measurement according to the specific bearing type, fit the axial and radial preload data obtained after the axial and radial preload forces are applied, and calculate the preload corresponding to the axial and radial preload forces. Then, the influence of the radial preload on the axial preload and the influence of the axial preload on the radial preload can be calculated.

[0085] Step S3: Calculate the contact load and original contact deformation;

[0086] First, calculate the variables based on radial load, axial load and contact angle Among them F a is the axial load, F r is the radial load;

[0087] Based on variables Radial load integral J r (ε), axial load integral J a (ε) and the bearing load distribution coefficient ε data fitting, respectively, obtained:

[0088] (1) Variables and radial load integral J r (ε) relation:

[0089]

[0090] (2) Variables and axial load integral J a (ε) is:

[0091]

[0092] (3) Variables The relationship between it and the load distribution coefficient ε is:

[0093]

[0094] Then calculate the maximum original contact load Q 0max , the relationship is as follows:

[0095]

[0096] Where Z is the number of rollers in the bearing;

[0097] In addition, according to the maximum original contact load Q 0max Obtain the maximum original contact deformation δ 0max , the relationship is as follows:

[0098] Q 0max =K l δ 0max n

[0099] Where n is a constant, which is 1.11; the load displacement coefficient K l Calculate according to the empirical formula, where l is the effective length of the roller;

[0100]

[0101] And through the simultaneous relationship, the original radial and original axial contact deformations δ are obtained: 0r , δ 0a , the relationship is as follows:

[0102] δ 0max =δ 0a sin(α)+δ 0r cos(α)

[0103]

[0104] Step S4: obtaining a new contact load according to the bearing preload and the change in contact deformation caused by thermal deformation;

[0105] After the bearing preload and the contact deformation change caused by thermal deformation are obtained, they are combined with the original contact deformation calculated in step S3, and the relationship is as follows:

[0106] δ a = δ 0a - P a - ΔG Ta

[0107] δ r = δ 0r - P r - ΔG T

[0108] wherein the radial and axial contact deformations δ r , δ a are obtained by combining the original radial and axial contact deformations δ 0r , δ 0a , the radial and axial preloads P r , P a of the bearing, and the radial and axial thermal deformations ΔG T , ΔG Ta ;

[0109] However, since the thermal deformation ΔG T only affects the radial contact deformation, the influence of the thermal deformation on the axial contact deformation, i.e. the axial thermal deformation ΔG Ta , needs to be calculated, and the relationship is as follows:

[0110]

[0111] Then, according to the radial and axial contact deformations δ r , δ a , the new maximum contact deformation, the new maximum contact load and the load distribution are calculated, and the relationship is as follows:

[0112] δ max = δ a sin (α) + δ r cos (α)

[0113] Q max = K l δ max n

[0114]

[0115] wherein δ max is the new maximum contact deformation calculated according to the new radial and axial contact deformations, Q max is the new maximum contact load calculated according to the new maximum contact deformation, and Q ψis the bearing load distribution, i.e. the bearing load at each azimuth angle, and ψ is the azimuth angle;

[0116] Step S5: Calculate fatigue life according to the contact load calculated in S4;

[0117] Solve the fatigue life empirical formula based on Lundberg and Palmgren, where the rated dynamic load Q is required. c , the relationship is as follows:

[0118]

[0119] Where λ is a constant coefficient selected according to the contact type; D is the roller diameter; l is the roller length; γ is obtained by calculation, and the relationship is as follows:

[0120]

[0121] Where D is the roller diameter; d m is the bearing pitch diameter;

[0122] Then the inner raceway dynamic load rating Q is calculated ci And outer raceway dynamic load rating Q co , the formula is as follows:

[0123]

[0124] Then, the equivalent dynamic load Q is obtained based on the fourth root mean square of the new contact load obtained in step S4. e , the relationship between contact load and equivalent dynamic load is as follows:

[0125]

[0126] Among them, Q j is the bearing load distribution Q ψ The contact load of each bearing in , j ranges from 1 to the number of rollers Z;

[0127] For the equivalent dynamic load Q of the inner and outer raceways ei , Q eo Since the contact loads of the inner and outer raceways are almost equal, Q is used. j As the contact load of the inner and outer raceways, used to calculate the equivalent dynamic load Q ei , Q eo ;

[0128] For non-rotating raceways, the square root of the equivalent dynamic load must be four times the Weibull slope e. The specific value must be calculated based on the contact angle α, as shown in the following formula:

[0129] e=1.5tan(α)

[0130] The equivalent dynamic load formula will also become:

[0131]

[0132] Among them, Q ewi With Q ewo That is, the equivalent dynamic load of the inner raceway and the equivalent dynamic load of the outer raceway in the non-rotating state;

[0133] Then use the equivalent dynamic load Q of the inner raceway Ei and rated dynamic load Q ci Calculate the inner raceway life L i , using the equivalent dynamic load Q of the outer raceway Eo and rated dynamic load Q co Calculate the outer raceway life L o , the relationship is as follows:

[0134]

[0135] Among them, the equivalent dynamic load Q of the inner raceway is Ei Including the equivalent dynamic load Q of the inner raceway under normal circumstances ei Equivalent dynamic load Q on the inner raceway in the non-rotating case ewi ;

[0136] Similarly, the equivalent dynamic load Q of the outer raceway is Eo Including the equivalent dynamic load Q of the outer raceway under normal circumstances eo Equivalent dynamic load Q on the outer raceway in the non-rotating case ewo ;

[0137] Finally, the bearing fatigue life L is calculated according to the following relationship:

[0138]

[0139] The fatigue life of the bearing is calculated from this;

[0140] (2) Bearing thermal deformation analysis model;

[0141] S2-2-1 Calculate frictional heat and convection heat transfer coefficient;

[0142] Obtain frictional heat and convection heat transfer coefficient to calculate thermal deformation;

[0143] Radial force friction torque M r Friction torque M with axial force a The calculation formula is as follows:

[0144]

[0145] Where n is the bearing speed; ν0 is the kinematic viscosity of the lubricant; K is a constant; f t is the radial load coefficient of the bearing, which is obtained by looking up the table based on K and the axial and radial loads; G is the geometric coefficient, which is obtained from the structural parameters of the bearing. The relationship is as follows:

[0146]

[0147] Among them, d m is the bearing pitch diameter; the radial force friction torque M is calculated separately r Friction torque M with axial force a Then, calculate the corresponding radial torque friction heat H r Friction heat H due to axial torque a , the relationship is the same, as shown below:

[0148] H r =1.047×10 -4 M r ·n

[0149] H a =1.047×10 -4 M a ·n

[0150] Among them, M r is the radial force friction torque; M a is the axial force friction torque;

[0151] For the convective heat transfer coefficient, you need to first calculate the Prandtl number P r and the Reynolds coefficient R e , both are related to the characteristics of the lubricant, and the relationship is as follows:

[0152]

[0153] Where c is the specific heat capacity of the lubricant, k is the thermal conductivity of the lubricant, ω is the angular velocity of the bearing, and ν0 is the kinematic viscosity of the lubricant. The convective heat transfer coefficient h is calculated according to the following formula: υ :

[0154]

[0155] Among them, D m is the average bearing diameter;

[0156] S2-2-2 Based on the results obtained in step S2-2-1, the temperature change is obtained through finite element analysis, and then the thermal deformation is calculated:

[0157] According to the input conditions obtained in step S2-2-1: radial torque friction heat H r Friction heat H due to axial torquea and the convective heat transfer coefficient h υ ; Perform thermal steady-state analysis on the bearing through finite element analysis to obtain the temperature change caused by friction and calculate the thermal deformation ΔG T , the relationship is as follows:

[0158] ΔG Ti =Γ1ΔT1D i

[0159] ΔG To =Γ2ΔT2D o

[0160] ΔG Te =Γ e ΔT e D

[0161] ΔG T =ΔG To -ΔG Ti -2ΔG Te

[0162] Among them, ΔG Ti is the thermal deformation of the inner raceway, ΔG To is the thermal deformation of the outer raceway, ΔG Te is the roller thermal deformation, Γ1 is the thermal expansion coefficient of the inner raceway, Γ2 is the thermal expansion coefficient of the outer raceway, Γ e is the roller expansion coefficient. The three expansion coefficients depend on their materials. ΔT1 is the temperature change of the inner raceway, ΔT2 is the temperature change of the outer raceway, and ΔT e is the roller temperature change, D i is the inner raceway diameter, D o is the outer raceway diameter;

[0163] The contact deformation change ΔG caused by thermal deformation is thus obtained T At the same time, it, the preload amount and the original contact deformation are substituted into step S4 to calculate the new contact deformation and the new contact load, and then the fatigue life of the bearing can be calculated according to step S5.

Claims

1. A fatigue life analysis method for wind turbine main bearings based on thermal-mechanical coupling, characterized in that: The steps are as follows: construct a basic bearing fatigue life analysis model; Step S1: determining the bearing and its structural parameters as well as external load and lubricant related parameters; Determine the bearing type and obtain the preload or preload amount, axial and radial loads, initial radial and axial clearances, contact angles, number of rollers, and raceway diameter from a mechanical analysis perspective; obtain the speed, lubricant viscosity, and specific heat capacity from a thermal analysis perspective; Step S2: Calculate the bearing preload amount according to the type and condition of preload; For bearing preload, it is divided into constant pressure preload and positioning preload; For positioning preload, determine the preload amount. By using the relationship between radial preload and axial preload, calculate the change in axial preload when radial preload is applied. Conversely, calculate the change in radial preload when axial preload is applied. The relationship is as follows: Among them, P 0a is the initial axial clearance, P 0r is the initial radial clearance, P a is the axial preload, P r is the radial preload, α is the contact angle; For constant pressure preload, it is necessary to conduct field data measurement according to the specific bearing type, fit the axial and radial preload data obtained after the axial and radial preload forces are applied, and calculate the preload corresponding to the axial and radial preload forces. Then, the influence of the radial preload on the axial preload and the influence of the axial preload on the radial preload can be calculated. Step S3: Calculate the contact load and original contact deformation; First, calculate the variables based on radial load, axial load and contact angle Among them F a is the axial load, F r is the radial load; Based on variables Radial load integral J r (ε), axial load integral J a (ε) and the bearing load distribution coefficient ε data fitting, respectively, obtained: (1) Variables and radial load integral J r (ε) relation: (2) Variables and axial load integral J a (ε) relation: (3) Variables The relationship between it and the load distribution coefficient ε is: Then calculate the maximum original contact load Q 0max , the relationship is as follows: Where Z is the number of rollers in the bearing; In addition, according to the maximum original contact load Q 0max Obtain the maximum original contact deformation δ 0max , the relationship is as follows: Q 0max =K l δ 0max n Where n is a constant, which is 1.11; the load displacement coefficient K l Calculate according to the empirical formula, where l is the effective length of the roller; And through the simultaneous relationship, the original radial and original axial contact deformations δ are obtained: 0r , δ 0a , the relationship is as follows: d 0max =d 0a sin(a)+d 0r cos(a) Step S4: obtaining a new contact load according to the bearing preload and the change in contact deformation caused by thermal deformation; After obtaining the bearing preload and the change in contact deformation caused by thermal deformation, they are combined with the original contact deformation obtained in step S3. The relationship is as follows: d a =d 0a -P a -ΔG Ta d r =d 0r -P r -ΔG T Among them, radial and axial contact deformation δ r , δ a is determined by: original radial and axial contact deformation δ 0r , δ 0a , radial and axial preload P of the bearing r 、P a , and radial and axial thermal deformation ΔG T , ΔG Ta Combination of However, due to thermal deformation ΔG T It will only affect the radial contact deformation, so it is necessary to calculate the effect of thermal deformation on axial contact deformation, that is, axial thermal deformation ΔG Ta , the relationship is as follows: Then according to the radial and axial contact deformation δ r , δ a Calculate the new maximum contact deformation, the new maximum contact load and the load distribution. The relationship is as follows: d max =d a sin(a)+d r cos(a) Q max =K l δ max n Among them, δ max is the new maximum contact deformation calculated from the new radial and axial contact deformations, Q max is the new maximum contact load calculated from the new maximum contact deformation, Q ψ is the bearing load distribution, i.e. the bearing load at each azimuth angle, and ψ is the azimuth angle; Step S5: Calculate fatigue life according to the contact load calculated in S4; Solve the fatigue life empirical formula based on Lundberg and Palmgren, where the rated dynamic load Q is required. c , the relationship is as follows: Where λ is a constant coefficient selected according to the contact type; D is the roller diameter; l is the roller length; γ is obtained by calculation, and the relationship is as follows: Where D is the roller diameter; d m is the bearing pitch diameter; Then the inner raceway dynamic load rating Q is calculated ci And outer raceway dynamic load rating Q co , the formula is as follows: Then, the equivalent dynamic load Q is obtained based on the fourth root mean square of the new contact load obtained in step S4. e , the relationship between contact load and equivalent dynamic load is as follows: Among them, Q j is the bearing load distribution Q ψ The contact load of each bearing in , j ranges from 1 to the number of rollers Z; For the equivalent dynamic load Q of the inner and outer raceways ei , Q eo Since the contact loads of the inner and outer raceways are almost equal, Q is used. j As the contact load of the inner and outer raceways, used to calculate the equivalent dynamic load Q ei , Q eo ; For non-rotating raceways, the square root of the equivalent dynamic load must be four times the Weibull slope e. The specific value must be calculated based on the contact angle α, as shown in the following formula: e=1.5tan(α) The equivalent dynamic load formula will also become: Among them, Q ewi With Q ewo That is, the equivalent dynamic load of the inner raceway and the equivalent dynamic load of the outer raceway in the non-rotating state; Then use the equivalent dynamic load Q of the inner raceway Ei and rated dynamic load Q ci Calculate the inner raceway life L i , using the equivalent dynamic load Q of the outer raceway Eo and rated dynamic load Q co Calculate the outer raceway life L o , the relationship is as follows: Among them, the equivalent dynamic load Q of the inner raceway is Ei Including the equivalent dynamic load Q of the inner raceway under normal circumstances ei Equivalent dynamic load Q on the inner raceway in the non-rotating case ewi ; Similarly, the equivalent dynamic load Q of the outer raceway is Eo Including the equivalent dynamic load Q of the outer raceway under normal circumstances eo Equivalent dynamic load Q on the outer raceway in the non-rotating case ewo ; Finally, the bearing fatigue life L is calculated according to the following relationship: The fatigue life of the bearing is calculated from this.

2. The fatigue life analysis method of a wind turbine main bearing based on thermal-mechanical coupling according to claim 1, characterized in that: Construct a bearing thermal deformation analysis model; S2-2-1 Calculate frictional heat and convection heat transfer coefficient; Obtain frictional heat and convection heat transfer coefficient to calculate thermal deformation; Radial force friction torque M r Friction torque M with axial force a The calculation formula is as follows: Where n is the bearing speed; ν0 is the kinematic viscosity of the lubricant; K is a constant; f t is the radial load coefficient of the bearing, which is obtained by looking up the table based on K and the axial and radial loads; G is the geometric coefficient, which is obtained from the structural parameters of the bearing. The relationship is as follows: Among them, d m is the bearing pitch diameter; the radial force friction torque M is calculated separately r Friction torque M with axial force a Then, calculate the corresponding radial torque friction heat H r Friction heat H due to axial torque a , the relationship is the same, as shown below: H r =1.047×10 -4 M r ·n H a =1.047×10 -4 M a ·n Among them, M r is the radial force friction torque; M a is the axial force friction torque; For the convective heat transfer coefficient, you need to first calculate the Prandtl number P r and the Reynolds coefficient R e , both are related to the characteristics of the lubricant, and the relationship is as follows: Where c is the specific heat capacity of the lubricant, k is the thermal conductivity of the lubricant, ω is the angular velocity of the bearing, and ν0 is the kinematic viscosity of the lubricant. The convective heat transfer coefficient h is calculated according to the following formula: υ : Among them, D m is the average bearing diameter; S2-2-2 Based on the results obtained in step S2-2-1, the temperature change is obtained through finite element analysis, and then the thermal deformation is calculated: According to the input conditions obtained in step S2-2-1: radial torque friction heat H r Friction heat H due to axial torque a and the convective heat transfer coefficient h υ ; Perform thermal steady-state analysis on the bearing through finite element analysis to obtain the temperature change caused by friction and calculate the thermal deformation ΔG T , the relationship is as follows: ΔG Ti =Γ1ΔT1D i ΔG To =Γ2ΔT2D o ΔG Te =C e ΔT e D ΔG T =ΔG To -ΔG Ti -2ΔG Te Among them, ΔG Ti is the thermal deformation of the inner raceway, ΔG To is the thermal deformation of the outer raceway, ΔG Te is the roller thermal deformation, Γ1 is the thermal expansion coefficient of the inner raceway, Γ2 is the thermal expansion coefficient of the outer raceway, Γ e is the roller expansion coefficient. The three expansion coefficients depend on their materials. ΔT1 is the temperature change of the inner raceway, ΔT2 is the temperature change of the outer raceway, and ΔT e is the roller temperature change, D i is the inner raceway diameter, D o is the outer raceway diameter; The contact deformation change ΔG caused by thermal deformation is thus obtained T At the same time, it, the preload amount and the original contact deformation are substituted into step S4 to calculate the new contact deformation and the new contact load, and then the fatigue life of the bearing can be calculated according to step S5.

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