Method for evaluating assembly interference magnitude of pin shaft and planet carrier of planetary gear train of wind power gear box

Abstract

The invention provides a method for evaluating the assembly interference magnitude of a pin shaft and a planet carrier of a planetary gear train of a wind power gear box. According to the wind power gear box planetary gear train dynamic model provided by the method, dynamic coupling between system vibration and gear pair meshing rigidity is considered, and the calculation accuracy of structural deformation and dynamic response is improved. The invention provides an efficient and feasible method for evaluating the assembly interference magnitude between the pin shaft and the planet carrier of the planetary gear train of the wind power gear box, the assembly interference magnitude can be effectively and accurately evaluated quickly, frequent disassembly of the gear box is not needed, and the test cost is reduced.

Description

Technical Field

[0001] The present invention relates to the technical field of wind power gearboxes, and particularly to a method for evaluating the interference fit between the pin shaft and the planet carrier of a planetary gear train in a wind power gearbox. Background Art

[0002] With the rapid development of the wind power industry, wind turbines are gradually developing towards ultra-large sizes. In a wind turbine, the gearbox plays an important role in speed increase and torque reduction. The tooth surface load of the planetary gear train in the gearbox is significantly affected by factors such as system assembly accuracy and structural deformation. In order to further improve the power density of the wind power gearbox, an interference fit is often used between the pin shaft of the planetary gear train and the planet carrier.

[0003] The pin shaft is the only component supporting the planet gear. If the interference fit between the pin shaft and the planet carrier is not reasonably selected, it will cause uneven load on the tooth surface of the planetary gear train, reduce the fatigue life, and it is difficult to ensure the long-term stable operation of the wind power gearbox. If the interference fit between the pin shaft and the planet carrier is too large, it will cause obvious assembly deformation of the pin shaft - planet carrier, resulting in misalignment of the pin shaft axis, causing uneven load on the tooth surface of the planetary gear train and accelerating fatigue failure. If the interference fit is too small, the connection strength between the pin shaft and the planet carrier will be reduced, the pin shaft is likely to slip, the load-bearing capacity is insufficient, and it is difficult to ensure the stable operation of the wind power gearbox. Accurately evaluating the interference fit between the pin shaft and the planet carrier of the planetary gear train in a wind power gearbox is of great significance. However, the existing technology has at least the following disadvantages:

[0004] A. To evaluate the load sharing performance of the planetary gear train in a wind power gearbox, the existing method often pastes strain gauges at the tooth root of the internal gear ring to obtain the tooth root strain and calculate the load sharing coefficient of the planetary gear train. This method can indirectly obtain the comprehensive influence of factors such as the interference fit between the pin shaft and the planet carrier and structural deformation on the load sharing performance of the planetary gear train, but it cannot directly and quantitatively evaluate the interference fit between each planet gear pin shaft and the planet carrier, and the gearbox needs to be disassembled, resulting in high test costs.

[0005] B. Since the interference fit between the pin shaft and the planet carrier is a state that exists once the assembly is completed, that is, the strain will always be maintained. If the test is carried out by detecting the strain increment later (testing with strain gauges pasted), the interference fit between the pin shaft and the planet carrier cannot be obtained.

[0006] Therefore, accurately evaluating the interference fit between the pin shaft and the planet carrier is an urgent problem to be solved. Summary of the Invention

[0007] The purpose of the present invention is to provide a method for evaluating the interference fit between the pin shaft and the planet carrier of a planetary gear train in a wind power gearbox to solve the problems existing in the prior art.

[0008] The technical solution adopted to achieve the purpose of the present invention is as follows. An evaluation method for the interference fit of the pin shaft and the planet carrier of the planetary gear train of a wind power gearbox. The planetary gear train includes a planet carrier, an internal gear ring, a sun gear, and planetary gears. The planet carrier includes side plates, planetary gear pin shafts, bushings, and bearings. A plurality of mounting holes are circumferentially arranged on the plate surface of the side plates. Each mounting hole is inserted with a planetary gear pin shaft that is in interference fit with the mounting hole. A bushing, a bearing, and a planetary gear are sequentially sleeved on each planetary gear pin shaft from the inside to the outside to form a planetary gear set. An internal gear ring that meshes with the planetary gear set is provided on the outer ring of the planetary gear set. The internal gear ring is fixed on the gearbox housing. The sun gear is arranged in the central area of the planetary gear set and meshes with the planetary gears. The evaluation method for the interference fit of the pin shaft and the planet carrier of the planetary gear train of a wind power gearbox includes the following steps:

[0009] 1) Considering the structural flexibility, establish a dynamic load-bearing contact model of the gear pair affected by vibration displacement. Among them, the dynamic load-bearing contact model of the gear pair includes the meshing model of the sun gear-planetary gear pair and the meshing model of the internal gear ring-planetary gear pair.

[0010] 2) Using the finite element condensation theory, establish a flexible body of the internal gear ring. By establishing a finite element model of the pin shaft-planet carrier considering the interference fit between the pin shaft and the planet carrier, extract the misalignment amount of the pin shaft axis, and generate a flexible body at the same time.

[0011] 3) According to the misalignment amount of the pin shaft axis, re-derive the dynamic meshing deformation expressions of the sun gear-planetary gear and the internal gear ring-planetary gear.

[0012] 4) Establish a multi-flexible body dynamics model of the planetary gear train of the wind power gearbox.

[0013] 5) Analyze the influence laws of the pin shaft misalignment caused by different interference fit amounts between each planetary gear pin shaft and the planet carrier on the tooth surface partial load coefficient of the planetary gear, the load sharing coefficient of multiple planetary gears, and the vibration response of the outer surface of the internal gear ring.

[0014] 6) Construct the mapping relationships between the "pin shaft-planet carrier interference fit amount" corresponding to different planetary gears and the tooth surface partial load coefficient, the load sharing coefficient of multiple planetary gears, and the vibration response of the outer surface of the internal gear ring respectively.

[0015] 7) Conduct a bench test on the wind power gearbox to obtain the vibration response of the outer surface of the internal gear ring. Use a strain gauge bridge measurement method to form a sensor, and paste the strain gauges on the compression side and the tension side of the tooth root of the internal gear ring of the high-speed heavy-load planetary gear train.

[0016] 8) Substitute the vibration response of the outer surface of the internal gear ring obtained in step 7) into the surrogate model of the "pin shaft-planet carrier interference fit amount"-"vibration response of the outer surface of the internal gear ring" corresponding to different planetary gears, and inversely calculate the "pin shaft-planet carrier interference fit amount" corresponding to different planetary gears.

[0017] 9) By using the "pin - planet carrier interference amount" - "tooth surface partial load coefficient" surrogate model and the "pin - planet carrier interference amount" - "multi - planet load sharing coefficient" surrogate model, the tooth surface partial load coefficient and the multi - planet load sharing coefficient are obtained, realizing the rapid evaluation of the dynamic performance of the planetary gear train in the wind power gearbox.

[0018] Further, step 1) specifically includes the following sub - steps:

[0019] 1.1) Establish a complete finite - element model of the internal gear ring, planet gears, and sun gear. Then, apply unit normal forces to the selected tooth surface nodes in sequence. By assembling the corresponding tooth surface node deformations, construct the flexibility coefficient matrix λ of the internal gear ring - planet gear pair meshing and the sun gear - planet gear pair meshing. G 。

[0020]

[0021] In the formula, represents the deformation of node κ1 in the normal direction when a unit normal force is applied to node κ2. n represents the number of tooth surface nodes.

[0022] 1.2) Establish a partial finite - element model of a gear with only one tooth, and fully constrain all interfaces except the contacting tooth surfaces. Apply reverse unit normal forces to each surface node in sequence and construct the flexibility matrix λ L 。

[0023]

[0024] In the formula, represents the deformation of node ε1 in the normal direction when a unit normal force is applied to node ε2.

[0025] 1.3) The global flexibility matrix of the tooth surface nodes can be obtained: λ bf =λ G +λ L 。According to the elastic contact theory, the deformation compatibility relationship of potential contact points is expressed as the following formula (4)

[0026]

[0027] In the formula, [λ bfξ represents the global flexibility matrix of potential contact points on the ξ - th contact line, interpolated from the global flexibility matrix of tooth surface nodes according to the rotation angle θ zpi of the i - th planet gear. {F ξ} represents the load vector of potential contact points on the ξ - th contact line. {Y ξ} represents the residual clearance vector of potential contact points on the ξ - th contact line. {δ ξ} represents the DTE vector of the potential contact point on the ξ-th contact line. {ε ξ} represents the initial clearance vector of the potential contact point on the ξ-th contact line, including gear manufacturing errors, gear modification, backlash, and pin deviation. [λ cξ represents the local flexibility matrix of the potential contact point on the ξ-th contact line. F l ,Y l ,δ l , and ε l represent the contact force, remaining clearance, relative mesh deformation, and initial clearance of the l-th potential contact point, respectively. The initial value of the load distribution is F l = F / η, where F represents the total quasi-static meshing force and η represents the number of potential contact points.

[0028] 1.4) Establish a finite element model of the pin - planet carrier considering the interference fit between the pin and the planet carrier. By setting load boundaries, constraint boundaries, interference fit of the pin - planet carrier, etc., extract the misalignment of the pin axis, including translational deviation {u epi} and angular deviation {θ epi}. The planet gear and the sun gear are modeled as rigid bodies with 6 degrees of freedom. The interference fit between the pin and the planet carrier will cause the planet gear shaft (equivalent to the pin) to deviate from the ideal position. Finally, through the finite element condensation theory, obtain the flexible bodies of the pin - planet carrier and the internal gear ring.

[0029] The planet gear and the sun gear are modeled as rigid solids with 6 degrees of freedom, and the interference fit between the planet carrier and the flexible pin will cause the planet axis to deviate from the ideal position

[0030]

[0031] In the formula, i pi , j pi and k pi are the unit vectors along the xpi axis, ypi axis, and zpi of the i-th planetary coordinate system, respectively. u epix (θ epix ), u epiy (θ epiy ), u epiz (θ epiz ) represent the deformation displacements (angles) of the i-th pin relative to the xpi axis, ypi axis, and zpi, respectively.

[0032] For the i-th internal gear ring / sun gear - planet gear meshing, the relative contact deformation at any potential contact point (M l ) on the meshing plane is derived by considering the pin deviation as a combination of infinitesimal displacements and angles.

[0033]

[0034] In the formula, δ spi (M l )| ideal and δ rpi (M l )| ideal respectively represent the dynamic transmission errors of the i-th sun gear - planet gear and the i-th internal gear ring - planet gear pairs when there is no misalignment of the pin shaft axis.

[0035] Furthermore, in step 5), set the interference fit amount combinations of the pin shafts - planet carriers of different planet gears and different amplitudes of the interference fit amount of the pin shafts - planet carriers, repeat the solution of the planetary gear train dynamic models under different combinations and different amplitudes of the interference fit amount of the pin shafts - planet carriers, and calculate the time series response of the vibration response of the internal gear ring tooth surface.

[0036] Furthermore, in step 6, use the Kriging surrogate model method to construct the mapping relationship between the vibration response of the outer surface of the internal gear ring and the interference fit amount of the pin shafts - planet carriers.

[0037] The technical effects of the present invention are beyond doubt:

[0038] A. The proposed dynamic model of the planetary gear train of the wind power gearbox takes into account the dynamic coupling between the system vibration and the meshing stiffness of the gear pairs, and improves the calculation accuracy of the structural deformation and dynamic response;

[0039] B. It can quickly and effectively and accurately evaluate the assembly interference amount, without the need to frequently disassemble the gearbox, reducing the test cost. Description of the Drawings

[0040] Figure 1 is a flowchart of an evaluation method for the interference fit amount of the pin shafts and planet carriers of the planetary gear train of a wind power gearbox;

[0041] Figure 2 is the test structure diagram of the planetary gear train bench;

[0042] Figure 3 is the interference amount prediction effect diagram based on the mapping relationship between the vibration response of the internal gear ring and the input torque, input speed, and the vibration response of the outer surface of the internal gear ring;

[0043] Figure 4 is a schematic diagram of the planetary gear train;

[0044] Figure 5 is a schematic diagram of the planet carrier.

[0045] In the figure: planet carrier 1, side plate 101, planet gear pin shaft 102, pin sleeve 103, bearing 104, internal gear ring 2, sun gear 3, planet gear 4. Detailed Embodiments

[0046] The present invention will be further described below in conjunction with embodiments, but it should not be understood that the above-mentioned subject scope of the present invention is limited to the following embodiments. Without departing from the above-mentioned technical idea of the present invention, various substitutions and changes made according to common general knowledge and customary means in the art should be included within the protection scope of the present invention.

[0047] Embodiment 1:

[0048] Refer to Figure 1 , this embodiment provides a method for evaluating the interference amount of the pin shaft and the planet carrier assembly of the planetary gear train of a wind power gearbox. Refer to Figure 4 and Figure 5 , the planetary gear train includes a planet carrier 1, an internal gear ring 2, a sun gear 3 and planet gears 4. The planet carrier 1 includes side plates 101, planetary gear pins 102, pin sleeves 103 and bearings 104. A plurality of mounting holes are circumferentially arranged on the plate surface of the side plate 101. A planetary gear pin 102 that is in interference fit with the mounting hole is inserted into each mounting hole. A pin sleeve 103, a bearing 104 and a planet gear 4 are sequentially sleeved on each planetary gear pin 102 from the inside to the outside to form a planetary gear set. An internal gear ring 2 that meshes with the planetary gear set is provided on the outer ring of the planetary gear set. The internal gear ring 2 is fixed on the gearbox housing. The sun gear 3 is arranged in the central area of the planetary gear set and meshes with the planet gears 4. The input torque drives the planet carrier 1 to rotate, and then drives the planet gears 4 to rotate through the flexible pins. The rotation of the planet gears ultimately drives the sun gear 3. The method for evaluating the interference amount of the pin shaft and the planet carrier assembly of the planetary gear train of a wind power gearbox includes the following steps:

[0049] 1) Considering the structural flexibility, establish a dynamic load-bearing contact model of the gear pair that affects the vibration displacement. Among them, the dynamic load-bearing contact model of the gear pair includes a sun gear-planet gear pair meshing model and an internal gear ring-planet gear pair meshing model. Step 1) specifically includes the following sub-steps:

[0050] 1.1) Establish a complete finite element model of the internal gear ring, planet gears, and sun gear, and then sequentially apply unit normal forces to the selected tooth surface nodes. By assembling the corresponding tooth surface node deformations, construct the flexibility coefficient matrix λ of the internal gear ring-planet gear pair meshing and the sun gear-planet gear pair meshing G .

[0051]

[0052] In the formula, represents the deformation of node κ1 in the normal direction when a unit normal force is applied to node κ2. n represents the number of tooth surface nodes.

[0053] 1.2) Establish a partial finite element model of a gear with only one tooth retained. All interfaces are fully constrained except for the contacting tooth surfaces. Apply reverse unit normal forces sequentially at each surface node and construct the flexibility matrix λ L .

[0054]

[0055] Wherein, represents the deformation of node ε1 in the normal direction when a unit normal force is applied at node ε2.

[0056] 1.3) The global flexibility matrix of the tooth surface nodes can be obtained: λ bf = λ G + λ L . According to the elastic contact theory, the deformation compatibility relationship of potential contact points is expressed as the following formula (4)

[0057]

[0058] Wherein, [λ bfξ represents the global flexibility matrix of potential contact points on the ξ-th contact line, which is interpolated from the global flexibility matrix of tooth surface nodes according to the rotation angle θ zpi of the i-th planetary gear. {F ξ} represents the load vector of potential contact points on the ξ-th contact line. {Y ξ} represents the residual clearance vector of potential contact points on the ξ-th contact line. {δ ξ} represents the DTE vector of potential contact points on the ξ-th contact line. {ε ξ} represents the initial clearance vector of potential contact points on the ξ-th contact line, including gear manufacturing errors, gear modification, backlash, and pin shaft deviation. [λ cξ represents the local flexibility matrix of potential contact points on the ξ-th contact line. F l , Y l , δ l , and ε l respectively represent the contact force, residual clearance, relative mesh deformation, and initial clearance of the l-th potential contact point. The initial value of the load distribution is F l = F / η, where F represents the total quasi-static meshing force and η represents the number of potential contact points.

[0059] 1.4) Establish a pin shaft - planet carrier finite element model considering the interference fit between the pin shaft and the planet carrier. By setting load boundaries, constraint boundaries, pin shaft - planet carrier interference fit, etc., extract the misalignment amount of the pin shaft axis, including translational deviation {u epi} and angular deviation {θ epi}. The planet gear and the sun gear are modeled as rigid bodies with six degrees of freedom. The interference fit between the pin shaft and the planet carrier will cause the planet gear shaft (equivalent to the pin shaft) to deviate from the ideal position. Finally, through the finite element condensation theory, the flexible bodies of the pin shaft - planet carrier and the internal gear ring are obtained.

[0060] The planet gear and the sun gear are modeled as rigid solids with six degrees of freedom. The interference fit between the planet carrier and the flexible pin will cause the planet shaft to deviate from the ideal position

[0061]

[0062] In the formula, i pi , j pi and k pi are the unit vectors of the i - th planet coordinate system along the xpi - axis, ypi - axis and zpi - axis respectively. u epix (θ epix ), u epiy (θ epiy ), u epiz (θ epiz ) represent the deformation displacements (angles) of the i - th pin shaft relative to the xpi - axis, ypi - axis and zpi - axis respectively.

[0063] For the i - th internal gear ring / sun gear - planet gear meshing, the relative contact deformation at any potential contact point (M l ) on the meshing plane is derived by considering the pin shaft deviation as a combination of infinitesimal displacements and angles.

[0064]

[0065] In the formula, δ spi (M l )| ideal and δ rpi (M l )| ideal represent the dynamic transmission errors of the i - th sun gear - planet gear and the i - th internal gear ring - planet gear pairs during meshing when there is no misalignment of the pin shaft axis.

[0066] 2) Using the finite element condensation theory, establish the flexible body of the internal gear ring. By establishing a finite element model of the pin shaft - planet carrier considering the interference fit between the pin shaft and the planet carrier, extract the misalignment amount of the pin shaft axis and generate the flexible body at the same time. Using the finite element sub - structure condensation method to model the gear ring, planet carrier and pin shaft, their free vibration equations can be obtained.

[0067]

[0068] In the formula, [M cp = [χ T [Mcp_FEM [χ] and [M r = [χ T [M r_FEM [χ] represent the mass matrices of the planet carrier - pin and the internal gear ring respectively; [K cp = [χ T [K cp_FEM [χ] and [K r = [χ T [K r_FEM [χ] represent the stiffness matrices of the planet carrier - pin and the internal gear ring respectively; {q cp} = [χ]{q cp_EFM} and {q r} = [χ]{q r_FEM} represent the generalized displacement vectors of the planet carrier - pin and the internal gear ring respectively, including the master nodes and the internal retained orders; the subscript 'FEM' represents the corresponding global matrices, generated by the finite element commercial software ABAQUS.

[0069] 3) Rededuce the dynamic meshing deformation expressions of the sun gear - planet gear and the internal gear ring - planet gear according to the misalignment amount of the pin axis.

[0070] 4) Establish a multi - flexible body dynamics model of the planetary gear train of the wind power gearbox. Based on the degrees of freedom and their coupling relationships, the generalized displacement vector of the system nodes in their reference coordinates can be defined. This vector includes the displacements of the sun shaft, the gear ring, the planet gear carrier pin assembly, and the planet gears.

[0071]

[0072] In the formula, {X c}, {X r}, {X s}, {X p} represent the generalized displacements of the pin - planet carrier, the internal gear ring, the sun shaft, and the planet gears respectively. N p is the number of planets. According to the above formula (7), the mass matrices, stiffness matrices, and damping matrices of each component are combined to form the overall mass matrix, stiffness matrix, and damping matrix.

[0073]

[0074] Among them, [M sys , [K sys , [C sys represent the overall mass matrix, stiffness matrix, and damping matrix of the system respectively. {F sys} represents the force vector including the input torque and the load.

[0075] 5) Analyze the influence laws of pin misalignment caused by different interference fit amounts between each planetary gear pin shaft and the planetary carrier on the tooth surface partial load coefficient of the planetary gear, the load sharing coefficient of multiple planetary gears, and the vibration response of the outer surface of the internal gear ring. Step 5) Set different combinations of interference fit amounts between the pin shaft and the planetary carrier of different planetary gears, as well as different amplitudes of the interference fit amount between the pin shaft and the planetary carrier, and repeatedly solve the dynamic model of the planetary gear train under different combinations and different amplitudes of the interference fit amount between the pin shaft and the planetary carrier, and calculate the time series response of the vibration response of the internal gear ring tooth surface.

[0076] 6) Construct the mapping relationships between the "interference amount between the pin shaft and the planetary carrier" corresponding to different planetary gears and the tooth surface partial load coefficient, the multi-planetary load sharing coefficient, and the vibration response of the outer surface of the internal gear ring respectively. Step 6) Use the Kriging surrogate model method to construct the mapping relationship between the vibration response of the outer surface of the internal gear ring and the interference fit amount between the pin shaft and the planetary carrier. Step 6) Specifically includes the following sub-steps:

[0077] 6.1) Divide the time series vibration response of the outer surface of the internal gear ring into several time periods, that is, obtain the snapshot set x = [x1, x2,..., x d T . Adopt the POD proper orthogonal decomposition method for dimensionality reduction, and obtain the reduced-dimensional time series vibration response of the internal gear ring as

[0078]

[0079] In the formula, φ is the basis modal matrix, A is the coefficient matrix, and K is the truncation order of the modal matrix.

[0080] 6.2) Adopt the Kriging surrogate model algorithm, and through data training, establish the mapping relationship between the k-th order coefficient A k in the coefficient matrix of the vibration response of the internal gear ring and the torque T, the rotational speed n, and the interference fit amount (u, θ) between the pin shaft and the planetary carrier.

[0081]

[0082] 6.3) Combining formula (9) and formula (10), the reconstructed response of any snapshot set of the time series vibration response of the internal gear ring can be obtained as:

[0083]

[0084] 6.4) Repeat the steps to obtain all the time series vibration responses of the internal gear ring.

[0085] 7) Conduct a bench test on the wind power gearbox to obtain the vibration response of the outer surface of the internal gear ring. Figure 2 ​It is a schematic diagram of the bench test for the planetary gear train of a wind power gearbox. The test equipment mainly consists of a strain gauge, a dynamic tester, and a computer. A sensor is formed by using the strain gauge bridge measurement method. The strain gauges are arranged at intervals of 120° on the compression side and the tension side of the tooth root of the internal gear ring of the planetary gear train. Through the on-site sensor arrangement, it is convenient to test and obtain the input torque time-series data, input speed time-series data, and the vibration response signal time-series data of the outer surface of the internal gear ring as shown in Equation (12).

[0086]

[0087] In the formula, R(t) is the set of sensor test data. T(t) is the time-series data of the input torque, and ω(t) is the time-series data of the input speed, is the time-series data of the vibration response signal.

[0088] 8) Substitute the vibration response of the outer surface of the internal gear ring obtained in step 7) into the surrogate model of "pin - planet carrier interference amount" - "vibration response of the outer surface of the internal gear ring" corresponding to different planet gears, and inversely calculate the "pin - planet carrier interference amount" corresponding to different planet gears. Combining the mapping relationship between the vibration response of the internal gear ring, the input torque, the input speed, and the vibration response of the outer surface of the internal gear ring obtained in Equation (11), the interference fit amount Δd between the pin and the planet carrier can be quickly obtained according to Equation (13), realizing the rapid inverse calculation of the interference fit amount based on testable experimental data, and the effect prediction is as Figure 3 shown.

[0089] Δd = g(f(x), R(t)) (13)

[0090] In the formula, Δd is the interference fit amount between the pin and the planet carrier, and f(x) is the mapping relationship between the vibration response of the internal gear ring, the input torque, the input speed, and the vibration response of the outer surface of the internal gear ring.

[0091] 9) Use the surrogate models of "pin - planet carrier interference amount" - "tooth surface partial load coefficient" and "pin - planet carrier interference amount" - "multi - planet equal load coefficient" to obtain the tooth surface partial load coefficient and the multi - planet equal load coefficient, and realize the rapid evaluation of the dynamic performance of the planetary gear train of the wind power gearbox.

[0092] It should be noted that this embodiment can more comprehensively capture the complex deformations and dynamic responses of various components in the system, especially showing more accuracy in the cases of complex loads and interference fits. It can accurately simulate the contact behavior, local stress distribution, and the coupling effect between flexible components, providing more reliable data support for fatigue life prediction and design optimization.

Claims

1. An evaluation method for the interference fit of the pin shaft and the planet carrier of a planetary gear train in a wind power gearbox, characterized in that: The planetary gear train includes a planetary carrier (1), an internal gear ring (2), a sun gear (3) and planetary gears (4); the planetary carrier (1) includes side plates (101), planetary gear pins (102), pin sleeves (103) and bearings (104); a plurality of mounting holes are circumferentially arranged on the plate surface of the side plates (101); a planetary gear pin (102) in interference fit with the mounting hole is inserted into each mounting hole; a pin sleeve (103), a bearing (104) and a planetary gear (4) are sequentially sleeved on each planetary gear pin (102) from inside to outside to form a planetary gear set; an internal gear ring (2) meshing with the planetary gear set is arranged on the outer ring of the planetary gear set; the internal gear ring (2) is fixed on the gearbox housing; the sun gear (3) is arranged in the central area of the planetary gear set and meshes with the planetary gear (4); the evaluation method for the interference amount of the planetary gear train pin shaft and the planetary carrier of the wind power gearbox includes the following steps: 1) Considering the structural flexibility, establish a dynamic load-bearing contact model of the gear pair affected by vibration displacement; among them, the dynamic load-bearing contact model of the gear pair includes a sun gear-planetary gear pair meshing model and an internal gear ring-planetary gear pair meshing model; 2) Using the finite element condensation theory, establish a flexible body of the internal gear ring; by establishing a finite element model of the pin shaft-planetary carrier considering the interference amount between the pin shaft and the planetary carrier, extract the misalignment amount of the pin shaft axis, and generate a flexible body at the same time; 3) According to the misalignment amount of the pin shaft axis, re-derive the dynamic meshing deformation expressions of the sun gear-planetary gear and the internal gear ring-planetary gear; 4) Establish a multi-flexible body dynamics model of the planetary gear train of the wind power gearbox; 5) Analyze the influence laws of the pin shaft misalignment caused by different interference fit amounts between each planetary gear pin shaft and the planetary carrier on the tooth surface partial load coefficient, multi-planetary gear equal load coefficient and the vibration response of the outer surface of the internal gear ring; 6) Construct the mapping relationships between the "pin shaft-planetary carrier interference amount" corresponding to different planetary gears and the tooth surface partial load coefficient, multi-planetary equal load coefficient and the vibration response of the outer surface of the internal gear ring respectively; 7) Conduct a bench test on the wind power gearbox to obtain the vibration response of the outer surface of the internal gear ring; use a strain gauge bridge measurement method to form a sensor, and paste the strain gauges on the compression side and the tension side of the tooth root of the internal gear ring gear of the high-speed heavy-duty planetary gear train; 8) Substitute the vibration response of the outer surface of the internal gear ring obtained in step 7) into the surrogate model of the "pin shaft-planetary carrier interference amount"-"vibration response of the outer surface of the internal gear ring" corresponding to different planetary gears, and inversely calculate the "pin shaft-planetary carrier interference amount" corresponding to different planetary gears; 9) Use the surrogate models of the "pin shaft-planetary carrier interference amount"-"tooth surface partial load coefficient" and the "pin shaft-planetary carrier interference amount"-"multi-planetary equal load coefficient" to obtain the tooth surface partial load coefficient and the multi-planetary gear equal load coefficient, and realize the rapid evaluation of the dynamic performance of the planetary gear train of the wind power gearbox.

2. The evaluation method for the interference fit amount between the pin shaft and the planet carrier of the planetary gear train of a wind power gearbox according to claim 1, characterized in that, Step 1) specifically includes the following sub-steps: 1.1) Establish a complete finite element model of the internal gear ring, planet gears, and sun gear. Then, sequentially apply unit normal forces to the selected tooth surface nodes. By assembling the deformations of the corresponding tooth surface nodes, construct the flexibility coefficient matrices Ω for the meshing of the internal gear ring-planet gear pair and the sun gear-planet gear pair. G ; In the formula, represents the deformation of node κ1 in the normal direction when a unit normal force is applied to node κ2; n represents the number of tooth surface nodes; 1.2) Establish a partial finite element model of a gear with only one tooth retained. All interfaces are fully constrained except for the contacting tooth surfaces. Apply reverse unit normal forces sequentially at each surface node and construct the flexibility matrix Ω L ; In the formula, represents the deformation of node ε1 in the normal direction when a unit normal force is applied to node ε2; 1.3) The global flexibility matrix of the tooth surface node can be obtained: λ bf = λ G + λ L ; According to the elastic contact theory, the deformation compatibility relationship of potential contact points is expressed as the following formula (4) where, [λ bfξ represents the global flexibility matrix of the potential contact points on the ξ-th contact line, interpolated from the global flexibility matrix of the tooth surface nodes according to the rotation angle θ zpi of the i-th planetary gear; {F ξ} represents the load vector of the potential contact points on the ξ-th contact line; {Y ξ} represents the residual clearance vector of the potential contact points on the ξ-th contact line; {δ ξ} represents the DTE vector of the potential contact points on the ξ-th contact line; {ε ξ} represents the initial clearance vector of the potential contact points on the ξ-th contact line, including gear manufacturing errors, gear modifications, backlash, and pin deviation; [λ cξ represents the local flexibility matrix of the potential contact points on the ξ-th contact line; F l , Y l , δ l , and ε l respectively represent the contact force, residual clearance, relative mesh deformation, and initial clearance of the l-th potential contact point; the initial value of the load distribution is F l = F / η, where F represents the total quasi-static meshing force and η represents the number of potential contact points; 1.4) Establish a pin - planet carrier finite element model considering the assembly interference between the pin and the planet carrier. By setting load boundaries, constraint boundaries, the assembly interference between the pin and the planet carrier, etc., extract the misalignment of the pin axis, including the translational deviation {u epi} and the angular deviation {θ epi}; The planet gear and the sun gear are modeled as rigid bodies with six degrees of freedom. The assembly interference between the pin and the planet carrier will cause the planet gear shaft (equivalent to the pin) to deviate from the ideal position; Finally, through the finite element condensation theory, obtain the flexible bodies of the pin - planet carrier and the internal gear ring; The planetary gear and the sun gear are modeled as rigid solids with six degrees of freedom. The interference fit between the planetary carrier and the flexible pin will cause the planetary shaft to deviate from the ideal position where i pi , j pi and k pi are the unit vectors of the i-th planetary coordinate system along the xpi axis, ypi axis and zpi, respectively; u epix (θ epix ), u epiy (θ epiy ), u epiz (θ epiz ) represent the deformation displacements (angles) of the i-th pin shaft relative to the xpi axis, ypi axis and zpi, respectively; For the i-th internal gear ring / sun gear - planet gear mesh, the relative contact deformation at any potential contact point (M l ) on the meshing plane is derived by considering the pin deviation as a combination of infinitesimal displacements and angles; ∫δ spi (M l )=δ spi (M l )| ideal +[{u epi}+{θ (6) where δ spi (M l )| ideal and δ rpi (M l )| ideal represent the dynamic transmission errors of the i-th sun gear - planet gear and the i-th internal gear ring - planet gear pairs when there is no misalignment of the pin shaft axis, respectively.

3. The evaluation method for the interference fit of the pin shaft and the planet carrier of the planetary gear train of a wind power gearbox according to claim 1, wherein: In step 5), set the interference fit amount combinations between the pin shafts and the planet carrier of different planet gears, as well as different amplitudes of the interference fit amount between the pin shafts and the planet carrier, repeat the solution of the dynamic model of the planetary gear train under different combinations and different amplitudes of the interference fit amount between the pin shafts and the planet carrier, and calculate the time series response of the vibration response of the internal gear ring tooth surface.

4. The evaluation method for the interference fit amount between the pin shaft and the planet carrier of the planetary gear train of a wind power gearbox according to claim 1, characterized in that: In step 6), use the Kriging surrogate model method to construct the mapping relationship between the vibration response of the outer surface of the internal gear ring and the interference fit amount between the pin shafts and the planet carrier.

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