Radial compact wind power bearing and checking method thereof

Abstract

The application discloses a radial structure compact wind power bearing and a checking method thereof, and relates to the technical field of wind power bearings. An annular boss is arranged on the outer periphery of the inner ring of the wind power bearing, and two axial sides of the annular boss and the inner side of the outer ring form two axial raceways. A main thrust roller with an axis parallel to the axis of the bearing and a radial roller with an axis perpendicular to the axis of the bearing are arranged in the raceways, and the two rollers are alternately arranged along the circumference and are distributed in an axial symmetry. An isolation block is arranged between the main thrust roller and the radial roller, and two arcuate contact surfaces of the isolation block are respectively matched with corresponding rollers. The checking method comprises the following steps: calculating a critical value of the working length of the roller for avoiding chord arc interference, determining the number ratio of the main thrust roller and the radial roller, calculating the contact stress and the equivalent dynamic load, checking the static force balance and verifying the service life and the strength, and finally determining the design parameters through iterative optimization. The application reduces the radial size and the overall weight of the bearing, improves the bearing efficiency and the lubrication effect, and avoids the interference risk through checking.

Description

Technical Field

[0001] This invention relates to the field of wind turbine bearing technology, and in particular to a radially compact wind turbine bearing and its verification method. Background Technology

[0002] The wind turbine pitch bearing is a key component connecting the blades and hub in a wind turbine generator set. It is mainly used to realize the pitch function of the blades. It must withstand the huge radial and axial loads and overturning moments from the blades, and also ensure that the blades can achieve precise angle adjustment during operation. It has an important impact on the reliability, efficiency and life of the unit.

[0003] Currently, wind turbine pitch bearings generally adopt a design approach of large diameter and ultra-thick cross-section to obtain sufficient radial and axial combined load-bearing capacity. However, this design results in a large overall bearing size, leading to high processing and manufacturing costs, and excessive weight for both the bearing and the turbine unit. This not only increases the difficulty and cost of transportation and installation but also has a certain impact on the bearing's service life, sealing performance, and lubrication effect. These problems are clearly inconsistent with the current trend in the wind power industry towards lightweight and compact turbine designs.

[0004] Furthermore, to avoid scratches or interference between the roller end face and the raceway, the art currently typically designs the roller working length to be 0.5 mm to 1 mm shorter than its diameter, based on experience. However, this empirical value is rather rough and fails to establish a precise mathematical relationship with the specific structural parameters of the bearing. It cannot accurately quantify the risk boundary of chord-arc interference, which may lead to a conservative or potentially risky design of the roller working length. Summary of the Invention

[0005] The technical problem to be solved by the present invention is to provide a radially compact wind turbine bearing and its verification method, so as to reduce the radial dimension of the bearing, reduce the overall weight and reduce the manufacturing cost while ensuring load-bearing performance, thereby adapting to the development needs of lightweight and compact wind turbines.

[0006] To achieve the above objectives, the technical solution adopted by the present invention is as follows: In a first aspect, the present invention provides a radially compact wind turbine bearing, comprising a first outer ring, an inner ring, and a second outer ring. The outer periphery of the inner ring is provided with a radially outwardly extending annular boss. Two axial sides of the annular boss are respectively opposite to the axial inner sides of the first and second outer rings, forming a first axial raceway and a second axial raceway. Both the first and second axial raceways contain alternating main thrust rollers and radial rollers arranged circumferentially. Both the main thrust rollers and radial rollers are cylindrical rollers. The axis of the main thrust roller is parallel to the bearing axis and is used to bear axial loads; the axis of the radial roller is perpendicular to the bearing axis and is used to bear radial loads. The first axial raceway contains main thrust rollers and the second axial raceway. In a second aspect, the present invention provides a verification method for the above-mentioned radially compact wind turbine bearing, comprising the following steps: S1. Based on the roller diameter d0 and the raceway mean diameter D, calculate the minimum critical reduction value H0 of the roller length required to avoid interference between the roller end face and the raceway, and determine the working length L of the roller based on the critical reduction value H0. S2, based on the expected operating conditions of the wind turbine bearing, determine the radial load Fr, axial load Fa, and overturning moment M that it will bear, and accordingly determine the roller quantity configuration ratio of the main thrust roller and the radial roller. S3. Calculate roller contact stress and deformation: Based on the radial load Fr, axial load Fa, and roller configuration ratio determined in step S2, apply Hertzian contact theory to calculate the contact stress, elastic deformation, and equivalent dynamic load P between the principal roller and the radial roller under the radial load Fr and axial load Fa, respectively. The coefficients X and Y are determined by the ratio of the number of rollers; S4. Substitute the radial load Fr, axial load Fa, and overturning moment M into the static balance equations of the wind turbine bearing to check whether the load distribution between the rollers is in a balanced state. S5. The maximum contact stress calculated in step S3 and the expected bearing life calculated based on the equivalent dynamic load P are compared and verified with the safety factor and life index required by the design. S6. If the verification result of step S5 does not meet the design requirements, return to adjust the roller diameter d0, roller working length L, or roller quantity configuration, and repeat the calculation and verification process of steps S1 to S5 until all design indicators are met, thereby determining the final design parameters of the wind turbine bearing.

[0007] As a preferred embodiment, step S1 includes a sub-step for designing the working length L of the roller: S11, Calculate the critical reduction value H0: S12, determine the working length L of the roller and ensure that it satisfies: d0-L>H0.

[0008] As a preferred embodiment, the principle for determining the roller quantity configuration ratio in step S2 includes: when the radial load Fr is much greater than the axial load Fa in the expected working condition, increase the configuration ratio of radial rollers relative to main push rollers; when the axial load Fa is much greater than the radial load Fr in the expected working condition, increase the configuration ratio of main push rollers relative to radial rollers; when Fr and Fa are comparable, use an equal number of radial rollers and main push rollers.

[0009] As a preferred embodiment, in step S4, the static equilibrium equation set includes radial equilibrium equations, axial equilibrium equations, and overturning moment equilibrium equations: Radial equilibrium equations: The axial equilibrium equation is: Overturning moment equilibrium equation: in, P is the position angle of the roller on the circumference of the slewing bearing. r K represents the radial clearance of the bearing. r F is the load-deformation constant between the radial roller and the two raceway surfaces. a For axial load, Position angle The normal load of the lower row of main thrust rollers located in the second axial raceway. Position angle The normal load of the upper row of main push rollers located in the first axial raceway is M, which is the overturning moment.

[0010] The main push rollers are symmetrically distributed along the annular boss, and the radial rollers in the first axial raceway and the radial rollers in the second axial raceway are symmetrically distributed along the annular boss. An isolation block is provided between the main push rollers and the radial rollers. The isolation block has a first contact surface and a second contact surface with opposite sides. The first contact surface is a first arched surface adapted to the outer cylindrical surface of the main push roller, and the second contact surface is a second arched surface adapted to the outer cylindrical surface of the radial roller. The arched directions of the first arched surface and the second arched surface are perpendicular to each other. Both the first arched surface and the second arched surface are provided with microtextures to store lubricant and reduce friction.

[0011] As a preferred embodiment, the microtexture is a V-shaped groove array, and the opening direction of the V-shaped groove array is consistent with the rolling direction of the roller.

[0012] As a preferred embodiment, the main push roller and the radial roller are cylindrical rollers with the same roller diameter and the same axial length, and the roller diameter is greater than its axial length to avoid the roller end face contacting the adjacent outer ring when bearing load.

[0013] According to the above technical solution, the beneficial effects of the present invention are: 1. This invention integrates the axial and radial bearing interfaces by placing the main push roller and the radial roller perpendicularly to each other on the same end face. This changes the design method in the prior art where the radial roller must have a separate raceway, significantly reducing the overall weight of the bearing. This not only conforms to the trend of lightweighting, but also reduces material consumption and processing costs, and helps to extend the service life of the bearing.

[0014] 2. In this invention, the horizontally placed main push roller and the vertically placed radial roller provide corresponding load transmission paths for axial force and radial force, respectively, making the force flow transmission direction clear, the efficiency higher, and the bearing operation stability better.

[0015] 3. This invention establishes a quantitative relationship model between the ratio of radial to axial rollers and the overall load-bearing capacity of the bearing, based on bearing static analysis and Hertzian contact theory. During design, the optimal number and spatial layout of the two types of rollers can be calculated and optimized using this model based on the rated radial and axial loads under target operating conditions, thereby achieving the optimal configuration of the bearing's load-bearing capacity. For example, under conditions of high radial load and low axial load, the number of radial rollers can be increased to improve radial load-bearing capacity and lifespan; conversely, the configuration of the main thrust rollers can be increased. This design principle can not only accurately adapt to diverse wind power operating conditions but can also be extended to other mechanical fields that require bearing complex loads.

[0016] 4. In this invention, the surface of the isolation block is provided with a V-shaped micro-texture in a specific direction, the opening direction of which is consistent with the rotation direction of the roller. This can effectively store the lubricating medium, extend the effective lubrication time, and reduce the coefficient of friction. The pits of the V-shaped micro-texture can accommodate metal chips or contaminants of 1 to 5 µm, preventing them from being crushed into the raceway by the roller, thereby reducing the risk of damage caused by abrasive wear and ensuring the reliability of the bearing's long-term operation. Attached Figure Description

[0017] Figure 1 This is a cross-sectional view of the radially compact wind turbine bearing of the present invention; Figure 2 for Figure 1 Top view of the bearing shown; Figure 3 for Figure 1 The diagram shows the internal structure of the bearing after the first and second outer rings have been removed. Figure 4 for Figure 3 A schematic diagram of the three-dimensional structure; Figure 5 This is a schematic diagram of the three-dimensional structure of the isolation block; Figure 6 This is the right view of the isolation block.

[0018] The markings in the diagram are: 1. First outer ring, 2. Main push roller, 3. Inner ring, 4. Radial roller, 5. Second outer ring, 6. Spacer block. Detailed Implementation

[0019] The present invention will now be described in further detail with reference to the accompanying drawings and specific embodiments.

[0020] The structures, proportions, sizes, etc., shown in the accompanying drawings of this specification are only for the purpose of assisting those skilled in the art in understanding and reading the content disclosed in the specification, and are not intended to limit the conditions under which the present invention can be implemented. Therefore, they have no substantial technical significance. Any modifications to the structure, changes in the proportions, or adjustments to the size, without affecting the effects and objectives that the present invention can produce, should still fall within the scope of the technical content disclosed in the present invention.

[0021] Furthermore, it should be noted that, unless otherwise stated, the terms "upper," "lower," "left," "right," "front end," "rear end," etc., indicating the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings, are only for the convenience of describing the present 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, and therefore should not be construed as a limitation of the present invention. Changes or adjustments to their relative relationships, without substantially altering the technical content, should also be considered as within the scope of the present invention. In addition, the terms "first," "second," etc., are used for descriptive purposes only and should not be construed as indicating or implying relative importance.

[0022] Example 1

[0023] like Figure 1 and Figure 2 As shown, this embodiment provides a radially compact wind turbine bearing, including a first outer ring 1, a main thrust roller 2, an inner ring 3, radial rollers 4, a second outer ring 5, and a spacer block 6.

[0024] The outer periphery of the inner ring 3 is provided with an integral annular boss extending radially outward, and the two axial sides of the annular boss, such as Figure 1 The upper and lower axial sides shown in the diagram are opposite to the axial inner sides of the first outer ring 1 and the second outer ring 5, respectively, thereby forming the first axial raceway and the second axial raceway.

[0025] Both axial raceways are equipped with main push rollers 2 and radial rollers 4, both of which are cylindrical rollers and are arranged alternately along the circumference. Specifically, the axis of the main push roller 2 is parallel to the axis of rotation of the bearing and is placed transversely within the raceway to bear axial loads; the axis of the radial roller 4 is perpendicular to the axis of rotation of the bearing and is placed vertically within the raceway to bear radial loads. Figure 3 and Figure 4 As shown, the main push rollers 2 in the two axial raceways are symmetrically distributed about the annular boss, and the radial rollers 4 in the two axial raceways are symmetrically distributed about the annular boss.

[0026] In this embodiment, the diameter of the cylindrical roller is greater than its axial length to ensure that the end face of the main push roller 2 will not contact the flange of the adjacent outer ring when subjected to radial load; and the end face of the radial roller 4 will not contact the flange of the adjacent outer ring when subjected to axial load, thereby effectively avoiding interference and wear caused by end face contact.

[0027] A spacer block 6 is provided between the main push roller 2 and the radial roller 4. For example... Figure 5 and Figure 6 As shown, the isolation block 6 has a first contact surface and a second contact surface. The first contact surface is a first arched surface adapted to the outer cylindrical surface of the main push roller 2, and the second contact surface is a second arched surface adapted to the outer cylindrical surface of the radial roller 4, and the arched directions of the first arched surface and the second arched surface are perpendicular to each other.

[0028] Both the first and second arched surfaces are provided with microtextures. In this embodiment, the microtexture is an array of V-shaped grooves, the opening direction of which is consistent with the rolling direction of the roller. This microtexture can store grease and extend the effective lubrication time. At the same time, its micro-pits can accommodate 1-5µm metal shavings or contaminants generated during operation, preventing them from being repeatedly crushed into the raceway by the rollers, thereby significantly reducing the coefficient of friction and abrasive wear rate, and improving the long-term reliability of the bearing.

[0029] Example 2

[0030] To avoid chordal interference between the roller end face and the raceway, which could lead to bearing seizure, the working length L of the roller must be precisely designed. This embodiment provides a design method for a radially compact wind turbine bearing as described in Embodiment 1. The specific steps of the design method include: S1: In this embodiment, the first axial raceway is taken as the object of analysis. A spatial rectangular coordinate system is established with the vertex of its theoretical conical surface as the origin and the bearing rotation centerline as the x-axis. Since the first and second axial raceways are structurally symmetrical and have the same parameters, this method and conclusions are also applicable to the second axial raceway.

[0031] Let d0 be the roller diameter, D be the raceway pitch diameter of the wind turbine pitch bearing, and H0 be the critical reduction value of the roller length relative to the roller diameter to avoid chordal interference. Establish a spatial rectangular coordinate system, and the equation of the outer raceway conical surface is: (1) (2) Substitute equation (2) into equation (1) to find z. 2 Then, the solution for x is derived by working backwards, namely: (3) The critical reduction value H0 is obtained by solving using the vector formula. H0 is a function of the roller diameter d0 and the raceway pitch diameter D, and its expression is: (4) S2: To ensure that the rollers do not interfere during operation, their working length L must meet the following constraints: d0 - L>H0(5) In this embodiment, the roller diameter d0 = 30mm and D = 1220mm are taken. Substituting into equation (4), the critical reduction value H0≈ 0.128mm is calculated. Therefore, the maximum allowable working length L of the roller is... max = d0 - H0 = 30 - 0.128 = 29.872mm. Therefore, as long as the actual designed roller working length L satisfies L < 29.872mm, interference can be avoided.

[0032] Example 3

[0033] In addition to determining the working length L of the roller in Example 2, the deformation of the roller is also a factor that must be considered to ensure that when subjected to axial load, the main push roller 2 bears most of the axial force, and when subjected to radial force, the radial roller 4 bears the force. The roller deformation verification method includes the following steps: S1: Preliminary Design and Proportioning In this invention, the main bearing roller 2 and radial roller 4 are arranged in a cross pattern, and their load-bearing characteristics are determined by the ratio of the number of rollers in each configuration. When the bearing is subjected to radial load Fr and axial load Fa, the load-bearing capacity is checked based on the equivalent dynamic load P, and the general calculation formula is as follows: (6) Where X is the radial coefficient and Y is the axial coefficient.

[0034] In this embodiment, the values ​​of coefficients X and Y are directly determined by the ratio of the number of main thrust rollers to radial rollers. By adjusting the ratio of the number of main thrust rollers to radial rollers, the bearing's adaptability to different load conditions can be optimized. When the external radial load is much greater than the axial load (Fr >> Fa): if a conventional uniformly distributed design is used, the roller rows that mainly bear the radial force will be overloaded, while the roller rows that bear the axial force will hardly play a role. This results in the overall life of the bearing being limited by the radial roller rows, and the structural space utilization is low. To address this, increasing the proportion of radial rollers 4 can be achieved by using a layout of two rows of radial rollers 4 with one row of main push rollers 2 in a 2:1 ratio, or a layout of three rows of radial rollers 4 with one row of main push rollers 2 in a 3:1 ratio. This can effectively reduce the equivalent value of the radial coefficient X, thereby reducing the equivalent dynamic load P and improving the bearing life under heavy radial loads.

[0035] When the external axial load is much greater than the radial load (Fa >> Fr), the roller row that mainly bears the axial force is prone to premature failure, reducing bearing life. To address this, the proportion of the main push rollers 2 can be increased. A layout of one row of radial rollers 4 and two rows of main push rollers 2 in a ratio of 1:2 can be adopted, which can reduce the equivalent value of the axial coefficient Y, thereby improving bearing life.

[0036] When the radial load and axial load are approximately equal (Fr ≈ Fa): using an equal number of radial rollers 4 and main push rollers 2 can achieve optimal load distribution among the roller rows, improving material and space utilization. In this configuration, the wear and fatigue processes of the radial rollers 4 and main push rollers 2 are essentially the same, thus optimizing the overall system lifespan.

[0037] S2: The formula for calculating line contact is as follows: Contact half width: Maximum compressive stress at the center of the contact width: Compressive stress at any point on the contact surface: Contact elastic deformation is: For rolling bearings made of bearing steel, substituting the material constant into the formula simplifies to: S3: Due to the crossed arrangement of the rollers, the bearing must maintain static equilibrium under combined loads to ensure a reasonable load distribution. The equilibrium state is described by the following set of equations: Radial equilibrium equations: (11) Among them, F r For radial load, Kr δ is the load-deformation constant between the radial roller and the two raceway surfaces. r This represents the axial displacement of the inner ring under external load. P is the position angle of the roller on the circumference of the slewing bearing. r This refers to the radial clearance of the bearing.

[0038] The axial equilibrium equation is: (12) Among them, F a For axial load, Position angle The normal load of the lower row of main thrust rollers located in the second axial raceway. Position angle The normal load on the upper row of main push rollers located in the first axial raceway.

[0039] The equilibrium equation for the overturning moment M is: (13) S4: Based on the above calculation results, perform the final verification: Strength safety factor: The maximum contact stress P0 calculated in step S2 of this embodiment is compared with the allowable contact stress of the bearing material to obtain the safety factor, which should meet the design requirements.

[0040] Operating life: Based on the equivalent dynamic load P determined in step S1 of this embodiment, the basic rated life L of the bearing is calculated according to the internationally accepted bearing life calculation standard. 10 The calculated expected lifespan L 10 It should not be less than the minimum working life required by the design.

[0041] In this embodiment, a turntable bearing used in a wind power pitch system is checked. Its operating conditions and requirements are as follows: radial load Fr = 2316 KN, axial load Fa = 68.9 KN, overturning moment M = 1622 KN·m. It operates for 8 hours per day, 200 days per year, at a speed of 0.8° per minute, with a safety factor S ≥ 2 and a service life of not less than 15 years.

[0042] According to the method of this invention, after the system calculation and optimization design of the above steps, the key verification results are shown in the table below: Table 1. Example of bearing verification results The calculated safety factors are all much greater than 2, and the calculated lifespan meets the requirement of more than 15 years, proving that the design fully meets the working conditions.

[0043] S5: Iterative Optimization Design If the verification results in step S4 do not fully meet the design requirements, the process must return to the design starting point, adjust variables such as roller diameter d0, working length L, and roller quantity configuration ratio, and repeat the verification process of steps S1-S4 in this embodiment until all performance indicators are qualified. This iterative process will ultimately output a set of optimal bearing design parameters, ensuring that it is compact and lightweight while possessing high reliability and long service life.

[0044] Matters not covered in this invention are common knowledge.

[0045] The above embodiments are only for illustrating the technical concept and features of the present invention, and are intended to enable those skilled in the art to understand the content of the present invention and implement it accordingly. They should not be construed as limiting the scope of protection of the present invention. All equivalent changes or modifications made in accordance with the spirit and essence of the present invention should be covered within the scope of protection of the present invention.

Claims

1. A radially compact wind turbine bearing, comprising a first outer ring (1), an inner ring (3), and a second outer ring (5), characterized in that: The outer periphery of the inner ring (3) is provided with an annular boss extending radially outward. The two axial sides of the annular boss are respectively opposite to the axial inner side of the first outer ring (1) and the axial inner side of the second outer ring (5), forming a first axial raceway and a second axial raceway. Both the first and second axial raceways are provided with main push rollers (2) and radial rollers (4) arranged alternately in the circumferential direction. The main push rollers (2) and radial rollers (4) are both cylindrical rollers. The axis of the main push rollers (2) is parallel to the bearing axis and is used to bear axial loads. The axis of the radial rollers (4) is perpendicular to the bearing axis and is used to bear radial loads. The main push rollers (2) in the first axial raceway and the main push rollers (2) in the second axial raceway are symmetrically distributed along the annular boss. The radial rollers (4) in the first axial raceway and the radial rollers (4) in the second axial raceway are symmetrically distributed along the annular boss. An isolation block (6) is provided between the main push roller (2) and the radial roller (4). The isolation block (6) has a first contact surface and a second contact surface. The first contact surface is a first arched surface adapted to the outer cylindrical surface of the main push roller (2). The second contact surface is a second arched surface adapted to the outer cylindrical surface of the radial roller (4). The arched directions of the first arched surface and the second arched surface are perpendicular to each other. Both the first arched surface and the second arched surface are provided with microtextures for storing lubricant and reducing friction.

2. The radially structured compact wind turbine bearing according to claim 1, characterized in that: The microtexture is a V-shaped groove array, and the opening direction of the V-shaped groove array is consistent with the rolling direction of the roller.

3. The radially compact wind turbine bearing according to claim 1, characterized in that: The main pusher roller (2) and the radial roller (4) are cylindrical rollers with the same roller diameter and the same axial length, and the roller diameter is greater than its axial length to avoid the roller end face contacting the adjacent outer ring when bearing load.

4. A verification method for the radially compact wind turbine bearing of claim 1, characterized in that: Includes the following steps: S1. Based on the roller diameter d0 and the raceway mean diameter D, calculate the minimum critical reduction value H0 of the roller length required to avoid interference between the roller end face and the raceway, and determine the working length L of the roller based on the critical reduction value H0. S2, based on the expected working conditions of the wind turbine bearing, determine the radial load Fr, axial load Fa and overturning moment M it bears, and determine the roller quantity configuration ratio of the main thrust roller (2) and the radial roller (4) accordingly; S3. Calculate roller contact stress and deformation: Based on the radial load Fr, axial load Fa, and roller configuration ratio determined in step S2, apply Hertzian contact theory to calculate the contact stress, elastic deformation, and equivalent dynamic load P between the main thrust roller (2) and the radial roller (4) under the radial load Fr and axial load Fa, respectively. The coefficients X and Y are determined by the ratio of the number of rollers; S4. Substitute the radial load Fr, axial load Fa, and overturning moment M into the static balance equations of the wind turbine bearing to check whether the load distribution between the rollers is in a balanced state. S5. The maximum contact stress calculated in step S3 and the expected bearing life calculated based on the equivalent dynamic load P are compared and verified with the safety factor and life index required by the design. S6. If the verification result of step S5 does not meet the design requirements, return to adjust the roller diameter d0, roller working length L, or roller quantity configuration, and repeat the calculation and verification process of steps S1 to S5 until all design indicators are met, thereby determining the final design parameters of the wind turbine bearing.

5. The verification method according to claim 4, characterized in that: Step S1 includes the sub-step of designing the working length L of the roller: S11, Calculate the critical reduction value H0: S12, determine the working length L of the roller and ensure that it satisfies: d0-L>H0.

6. The verification method according to claim 4, characterized in that: In step S2, the principle for determining the configuration ratio of the number of rollers includes: when the radial load Fr is much greater than the axial load Fa in the expected working condition, the configuration ratio of radial rollers (4) to main push rollers (2) is increased; when the axial load Fa is much greater than the radial load Fr in the expected working condition, the configuration ratio of main push rollers (2) to radial rollers (4) is increased; when Fr is comparable to Fa, the radial rollers (4) and main push rollers (2) of equal number are used.

7. The verification method according to claim 4, characterized in that: In step S4, the static equilibrium equation set includes radial equilibrium equations, axial equilibrium equations, and overturning moment equilibrium equations: Radial equilibrium equations: The axial equilibrium equation is: Overturning moment equilibrium equation: in, P is the position angle of the roller on the circumference of the slewing bearing. r K represents the radial clearance of the bearing. r F is the load-deformation constant between the radial roller and the two raceway surfaces. a For axial load, Position angle The normal load of the lower row of main thrust rollers located in the second axial raceway. Position angle The normal load of the upper row of main push rollers located in the first axial raceway is M, which is the overturning moment.

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