Cylindrical roller bearing, retainer and bearing assembly for wind power gear box

Abstract

The utility model relates to a cylindrical roller bearing for a wind power gear box, a retainer and a bearing assembly. The cylindrical roller bearing comprises: an inner ring, the outer peripheral surface of which is provided with a row of raceways; the retainer is arranged outside the inner ring and provided with a plurality of pockets at equal intervals in the circumferential direction, the cylindrical rollers are located in the raceway, at least two rows of cylindrical rollers are arranged in the raceway side by side in the axial direction, at least two cylindrical rollers are arranged in each pocket side by side in the axial direction, and the cylindrical rollers are located in the raceway. The diameter Dw of each cylindrical roller is 25-35 mm, the ratio lambda of the axial length Lw to the diameter Dw of each cylindrical roller is 2.7-4.5, and the ratio lambda 1 of the raceway width sigma Le of the single inner ring to the diameter Dw of the single cylindrical roller is 5.4-9.

Description

TECHNICAL FIELD

[0001] The utility model relates to the technical field of wind power gear box, especially to a cylindrical roller bearing, retainer and bearing assembly for wind power gear box. BACKGROUND

[0002] Realizing the goal of wind power and thermal power with the same price and reducing the levelized cost of energy (LCoE) is the main pressure that the wind power industry continuously faces. The torque density of the wind power gear box directly affects the LCoE, therefore, it is particularly urgent to improve the torque density of the wind power gear box.

[0003] The mainstream wind power gear box is usually composed of two or more stages of planetary gears. Increasing the number of planetary gears can improve the torque density, because more number of planetary gears will reduce the contact stress level of the single gear contact surface, which can significantly reduce the diameter of the planetary gears and the gear ring. This has formed the trend of the development of the megawatt number of wind turbines to large-scale in recent years, but the design of the wind power gear box tends to miniaturization and light weight. This design trend significantly reduces the size of the gear box, and also reduces the bearing installation space.

[0004] In order to enable the bearing to meet the comprehensive requirements of load capacity, calculated life, etc., in the application field of wind power gear box, the current two bearing solutions are: cylindrical roller bearing and hydraulic sliding bearing.

[0005] The cylindrical roller bearing has the characteristics of maturity, stability and high reliability in the application field of wind power gear box, and has strong fault operation ability. Despite this, under the design trend of wind power gear box, the size of the gear box and the bearing installation space are further reduced, and the existing cylindrical roller bearing still faces the following main challenges: under the existing size, the performance of the cylindrical roller bearing (especially the contact stress on the bearing raceway) has reached the upper limit of the design standard requirements of the wind power gear box. If the bearing size is further reduced to adapt to smaller space, it will be difficult to meet the design standard requirements of the bearing contact stress, and the theoretical calculated life of the bearing will also be sharply reduced, which cannot meet the life requirements of the planetary gear bearing of the wind power gear box. In addition, the current design of the cylindrical roller bearing fails to fully utilize the axial space to bear the load, and also fails to fully utilize the radial space to bear the load. The effective load bearing entity of the bearing only accounts for about 75% of the volume of the bearing installation area.

[0006] Another solution for the trend of reducing the size of the planetary gear of the wind turbine gearbox is the hydraulic sliding bearing, which has a simple structure and can effectively reduce the size of the bearing, save installation space, and has a low cost. In addition, the hydraulic sliding bearing bears the load through the hydrodynamic effect. Since it adopts surface contact, it can ideally bear extremely high loads. However, it still has the following problems: It needs to be equipped with a complex lubrication system, the maintenance cost is high, and the service life of the hydraulic sliding bearing may be limited by the stability of the lubricating oil film. In harsh working conditions (such as temperature changes, pollution, and humidity), the stability of the lubricating oil film will be affected, thereby affecting the service life and reliability of the bearing. In some specific working conditions, such as idling or blade hoisting scenarios, the hydraulic sliding bearing has a risk of failure. For example, when a single blade is hoisted, the oil film has not yet formed at this time, and the bearing bears a huge force, which is very easy to cause the bearing to fail. Once the hydraulic sliding bearing fails, it may cause direct metal contact between the gear inner hole wall and the bearing, generate a large amount of heat, and the sliding bearing will be quickly damaged in a short time, causing the entire wind turbine gearbox to burn out and causing serious failure.

[0007] In other fields (such as mechanical equipment fields such as rolling mills and automobile transmissions), in order to improve the load capacity or service life of cylindrical roller bearings, relevant patent documents such as CN219139627U and CN210218408U have disclosed designs that cancel the middle rib and place two or more rows of rollers side by side in one raceway. However, in the wind turbine gearbox, there has never been a similar design and application case, and under the existing design structure parameters and load conditions of the wind turbine gearbox, the design of arranging two or more rows of conventional rollers side by side cannot make the bearing reach the required load level. Because, the wind turbine gearbox has special application requirements for the ratio of the outer diameter and the width of the planetary gear bearing in the design structure, which cannot be fully met in the design of other fields. Practical new type content

[0008] To overcome the problems in the related art, the present disclosure provides a cylindrical roller bearing, a retainer, and a bearing assembly for a wind turbine gearbox.

[0009] According to a first aspect of the embodiments of the present disclosure, the present disclosure provides a cylindrical roller bearing for a wind turbine gearbox, comprising: an inner ring, the outer circumferential surface of the inner ring is provided with a row of raceways; a retainer is arranged outside the inner ring and is provided with a plurality of pockets at equal intervals in the circumferential direction, and cylindrical rollers are located in the raceways, at least two rows of cylindrical rollers are arranged side by side in the axial direction in the raceway, at least two cylindrical rollers are arranged side by side in the axial direction in each pocket, and the diameter Dw of the cylindrical roller is 25-35 mm, and the ratio λ of the axial length Lw of the cylindrical roller to the diameter Dw is 2.7-4.5.

[0010] In some embodiments, the diameter Dw of the cylindrical roller is 25-35 mm, and the ratio λ of the axial length Lw of the cylindrical roller to the diameter Dw is 3.2-4.5.

[0011] According to a second aspect of the embodiments of the present disclosure, the present disclosure provides a cylindrical roller bearing for a wind turbine gearbox, comprising an inner ring, an outer periphery of the inner ring being provided with a row of raceways; a cage, the cage being provided outside the inner ring and being provided with a plurality of pockets at equal intervals in the circumferential direction, and cylindrical rollers, the cylindrical rollers being located in the raceways, at least two rows of the cylindrical rollers being arranged side by side in the axial direction in the raceways, at least two of the cylindrical rollers being arranged side by side in the axial direction in each of the pockets, and a ratio λ1 of a raceway width ∑Le of a single inner ring to a diameter Dw of a single cylindrical roller being 5.4-9.

[0012] In some embodiments, the ratio λ1 of the raceway width ∑Le of a single inner ring to the diameter Dw of a single cylindrical roller is 6.4-9.

[0013] In some embodiments, the rolling surface of the cylindrical roller is in direct rolling contact with the outer peripheral surface of the inner ring and is used for direct rolling contact with the inner peripheral surface of the inner hole of the planetary gear of the wind turbine gearbox.

[0014] In some embodiments, the end faces of the cylindrical rollers of adjacent rows located in the same row of the raceways abut against each other.

[0015] In some embodiments, the axial lengths Lw of the at least two cylindrical rollers located in the same pocket are not equal.

[0016] In some embodiments, the axial positions at which the end faces of the two rows of the cylindrical rollers abut against each other in the circumferential direction are different.

[0017] In some embodiments, the raceway is bounded by two radially outwardly arranged and axially spaced retaining edges of the inner ring.

[0018] In some embodiments, the retaining edges are fixed retaining edges and / or movable retaining rings, the fixed retaining edges being integrally formed with the inner ring, and the movable retaining rings being separately formed with the inner ring.

[0019] In some embodiments, the inner ring is integrally formed or is spliced by two sub-inner rings.

[0020] In some embodiments, the inner diameter of the cylindrical roller bearing is 200-350 mm, the diameter of the envelope circle of the cylindrical roller bearing is 260-500 mm, and the axial width of the inner ring is 180-440 mm.

[0021] According to a third aspect of the embodiments of the present disclosure, the present disclosure provides a cage applied to the cylindrical roller bearing of the first aspect and the second aspect, comprising: a plurality of cross beams, the cross beams being provided with cross beam inclined surfaces on both circumferential sides thereof; and two edge beams, respectively connected to both axial ends of the cross beams and enclosing the cross beams to form a plurality of closed-loop pockets, the pockets being used for arranging at least two cylindrical rollers side by side in an axial direction, wherein the cross beam inclined surfaces are provided with protrusions for contact guiding with rolling surfaces of the cylindrical rollers, the protrusions comprising end protrusions and intermediate protrusions, the intermediate protrusions being used for contact guiding with rolling surfaces of two adjacent cylindrical rollers, and the intermediate protrusions making the rolling surface of each cylindrical roller in the pocket have four contact positions with the cross beam.

[0022] According to a fourth aspect of the embodiments of the present disclosure, the present disclosure provides a cage applied to the cylindrical roller bearing of the first aspect and the second aspect, comprising: a plurality of cross beams, the cross beams being provided with cross beam inclined surfaces on both circumferential sides thereof; and two edge beams, respectively connected to both axial ends of the cross beams and enclosing the cross beams to form a plurality of closed-loop pockets, the pockets being used for arranging at least two cylindrical rollers side by side in an axial direction, wherein the cross beam inclined surfaces are provided with protrusions for contact guiding with rolling surfaces of the cylindrical rollers, the protrusions comprising end protrusions and intermediate protrusions, in a direction perpendicular to the cross beam inclined surfaces, the highest position of the end protrusions has a height difference h1 with the cross beam inclined surface, the highest position of the intermediate protrusions has a height difference h2 with the cross beam inclined surface, and h1 < h2, 0 < Δh = h2 - h1 < 0.2 mm.

[0023] In some embodiments, the axial spacing L1 between the starting positions of the two end protrusions on the inner side in the axial direction satisfies the relationship: 0.50Lwn < L1 < 0.95Lwn, and the axial length L2 of the intermediate protrusion satisfies the relationship: 0.05Lwn < L2 < 0.50Lwn, where Lwn is the sum of the axial lengths Lw of n cylindrical rollers in a single pocket.

[0024] In some embodiments, the surface of the protrusion comprises a curved surface extending in the axial direction, and the intermediate protrusion comprises two sub-protrusions, so that the intermediate protrusion has two highest positions for contact guiding with rolling surfaces of two adjacent cylindrical rollers, respectively.

[0025] In some embodiments, the curved surface is a cylindrical surface, and the radius R of the curved surface satisfies the relationship: R > 0.09 * Lw 2 .

[0026] In some embodiments, the two sub-protrusions are spaced apart in the axial direction.

[0027] In some embodiments, the surface of the protrusion is planar.

[0028] In some embodiments, the intermediate protrusion comprises two axially spaced sub-protrusions.

[0029] In some embodiments, the relationship between the height difference H between the highest position of the protrusion and the slope of the cross beam and the diameter Dw of the cylindrical roller is: 0.1mm<H<0.05*Dw, where H is h1 and / or h2.

[0030] According to a fifth aspect of the embodiments of the present disclosure, the present disclosure provides a bearing assembly, comprising: a mandrel; a planetary gear sleeved outside the mandrel; a cylindrical roller bearing for a wind power gear box as described in the first aspect and the second aspect, the cylindrical roller bearing being arranged between the mandrel and the planetary gear for rotating the planetary gear relative to the mandrel, wherein at least two groups of the cylindrical roller bearings are arranged along the axial direction of the mandrel, and the inner circumferential surface of the planetary gear is provided with a radially protruding intermediate rib, which respectively abuts against the end surface of the cylindrical roller of the adjacent cylindrical roller bearing.

[0031] The technical solutions provided by the embodiments of the present disclosure can have the following beneficial effects:

[0032] 1. The outer circumferential surface of a single inner ring only has one raceway, eliminating the stress generated by the size deviation of multiple raceway surfaces. By removing the intermediate rib and arranging at least two rows of cylindrical rollers in the single raceway, the axial space can be more effectively utilized, and the effective load-bearing entity of the bearing accounts for about 90% of the volume of the bearing installation area, thereby enhancing the load-bearing capacity of the bearing.

[0033] 2. The diameter of the cylindrical roller, the ratio of the axial length of the cylindrical roller to the diameter, and the ratio of the raceway width of the inner ring to the diameter of the cylindrical roller are optimized. The diameter of the cylindrical roller is smaller, so that the cylindrical roller is small and the number of cylindrical rollers in the circumferential direction is large, and the bearing has stronger load-bearing capacity, which is improved by 20% compared to the existing cylindrical roller bearing of the wind power gear box. In addition, the outer shape size of the cylindrical roller bearing of the present disclosure is reduced by 25% compared to the existing cylindrical roller bearing of the wind power gear box, and has a bearing installation space size basically equivalent to a hydraulic sliding bearing. The cylindrical roller bearing is more compact and smaller in size, and has a smaller linear speed at the same rotational speed. The friction torque and friction power consumption are small during the start-up stage and operation of the wind power gear box. In addition, the weight is reduced by 50% compared to the conventional cylindrical roller bearing of the wind power gear box, and the acceleration and deceleration working conditions are very good, and the impact inertia is small. The reduction in weight reduces the production cost, and the bearing cost is basically equivalent to that of a hydraulic sliding bearing.

[0034] 3. The cylindrical roller bearing without outer ring design reduces the overall cylindrical roller bearing and even the radial dimension of the wind turbine gearbox, which is beneficial for the pursuit of the ultimate radial space of the wind turbine gearbox.

[0035] 4. The end protrusion and the middle protrusion of the cage not only reduce the contact area of the cage and the cylindrical roller, but also avoid the main load bearing area of the cylindrical roller. Each cylindrical roller is supported by four support points, making the support more stable and the force on the cylindrical roller more uniform. BRIEF DESCRIPTION OF DRAWINGS

[0036] The accompanying drawings, which are incorporated herein and form a part of the specification, illustrate embodiments consistent with the present disclosure and, together with the description, further serve to explain the principles of the disclosure.

[0037] Figure 1 is a cross-sectional view of a cylindrical roller bearing for a wind turbine gearbox according to an exemplary embodiment;

[0038] Figure 2 is Figure 1 is a perspective structural schematic diagram of a cylindrical roller bearing for a wind turbine gearbox in

[0039] Figure 3 is a cross-sectional view of a cylindrical roller bearing for a wind turbine gearbox according to another exemplary embodiment.

[0040] Figure 4 is a perspective schematic diagram of the cage cooperating with the cylindrical roller according to an exemplary embodiment;

[0041] Figure 5 is a partial perspective view of the cage according to an exemplary embodiment;

[0042] Figure 6 is a transverse cross-sectional view of the cage according to an exemplary embodiment;

[0043] Figure 7 is a structural schematic diagram of a single cross beam and its partial G and H enlarged views according to an exemplary embodiment;

[0044] Figure 8 is Figure 6 is a schematic diagram of the cage and its pockets in the B direction view in

[0045] Figure 9 is Figure 6 is a partial schematic diagram of the cage and its pockets in the B direction view in

[0046] Figure 10 is a cross-sectional view of a bearing assembly according to an exemplary embodiment. DETAILED DESCRIPTION

[0047] The exemplary embodiments will be described in detail herein with reference to the accompanying drawings. In the following description, like reference numerals refer to like elements unless indicated otherwise. The following exemplary embodiments described herein represent the best understanding of the present disclosure, but are not the only ones consistent with it. Instead, they are only examples of devices and methods consistent with some aspects of the present disclosure, as detailed in the appended claims.

[0048] In the present application, unless otherwise specified, the axial direction A, the radial direction R and the circumferential direction W refer to the axial direction A, the radial direction R and the circumferential direction W of the cylindrical roller bearing 100, respectively.

[0049] To solve the above technical problems, the present disclosure provides a cylindrical roller bearing 100 which can be used in the bearing assembly (such as shown in Figure 10

[0050] As shown in Figure 1 and Figure 2 Each independent cylindrical roller bearing 100 can include an inner ring 10, a cage 20 and cylindrical rollers 30. The outer periphery of the inner ring 10 is provided with a raceway 11 and two ribs 12, the two ribs 12 protrude radially from the outer surface of the inner ring 10, and the two ribs 12 are located at the axial ends of the inner ring 10, respectively, and the two ribs 12 are axially spaced to form the raceway 11. It can be seen that the inner ring 10 is a single raceway 11 inner ring 10 with double ribs 12. The cage 20 is sleeved on the outside of the inner ring 10 and is provided with a plurality of pockets at equal intervals along the circumferential direction. At least two rows of cylindrical rollers 30 are arranged side by side along the axial direction A in each row of raceways 11, and at least two cylindrical rollers 30 are arranged side by side along the axial direction A in each pocket. The axial end faces of the cylindrical rollers 30 of adjacent rows can also abut each other.

[0051] The inner ring of the existing cylindrical roller bearing has two or more rows of raceways, and each row of raceways is provided with only one row of rollers. However, during the manufacturing process, it is difficult to control the size of multiple rows of raceways, and the parallelism deviation of two or more rows of raceways may occur. Such deviation will cause uneven stress on the rollers during operation, thereby increasing the internal stress of the bearing and reducing its load capacity and service life.

[0052] Compared with the existing cylindrical roller bearing, the outer periphery of the single inner ring 10 of the cylindrical roller bearing 100 of the present disclosure has only one raceway 11, which eliminates the parallelism deviation problem between multiple rows of raceways 11 and ensures the consistency of the raceway 11. Since there is only one row of raceways 11, the size control during the production process is also easier, and the stress problem caused by the deviation of the production process is reduced. By eliminating the parallelism and size deviation, the cylindrical rollers 30 are more evenly stressed during operation, thereby improving the load capacity and service life of the bearing. ​

[0053] In the embodiment, as shown in Figure 10 The rolling surface of the cylindrical roller 30 is in direct rolling contact with the outer peripheral surface of the inner ring 10 (i.e. the raceway 11) and the inner peripheral surface of the inner hole of the planetary gear 200 of the wind turbine gearbox, and the envelope circle diameter, i.e. the maximum outer diameter of the cylindrical roller bearing 100, cancels the traditional outer ring, reduces the radial dimension of the entire cylindrical roller bearing 100 and even the wind turbine gearbox, and is conducive to the pursuit of the ultimate radial space of the wind turbine gearbox.

[0054] In the embodiment, at least two rows of cylindrical rollers 30 are arranged axially side by side in the raceway 11, and at least two cylindrical rollers 30 are arranged axially A side by side in each pocket.

[0055] Therefore, under the same bearing width, the axial length Lw of the cylindrical roller 30 is increased, the axial space can be used more efficiently, the entity of the bearing effectively bearing load accounts for about 90% of the volume of the bearing installation area, not only meets the more compact design requirement, but also enhances the bearing load bearing capacity and improves the bearing efficiency. At the same time, under the same bearing demand, the axial length Lw of the cylindrical roller 30 is increased, which not only improves the bearing load capacity and promotes the miniaturization and light weight of the equipment using the cylindrical roller bearing 100, which is particularly important for the wind power industry pursuing high torque density.

[0056] In the embodiment, the retainer 20 can be an integrated retainer formed by stamping a steel plate. The pockets of the steel retainer are in a state of being wrapped by a very small part of the cylindrical roller, so the radial dimension of the cross section of the retainer is small, the weight of the retainer is light, thereby the overall weight of the bearing can be reduced, and the retainer is formed by stamping a steel plate, so the material cost and manufacturing cost are low, thereby the overall cost of the bearing can be reduced.

[0057] The diameter Dw of the cylindrical roller 30 of the present disclosure is 25-35 mm, and the ratio λ (i.e. the aspect ratio) of the axial length Lw of the cylindrical roller 30 to the diameter Dw is 2.7-4.5;

[0058] The ratio λ1 of the raceway width ∑Le of the single inner ring 10 to the diameter Dw of the single cylindrical roller 30 is 5.4-9. The diameter range of the cylindrical roller 30 of the present disclosure, the ratio range of the axial length Lw of the cylindrical roller 30 to the diameter Dw, and the ratio range of the raceway width ∑Le of the inner ring 10 to the diameter Dw of the cylindrical roller 30, make the radial occupation size of the cylindrical roller 30 small, so that the diameter of the envelope circle of the cylindrical roller bearing 100 is smaller, and in the present embodiment, specifically, the diameter of the envelope circle of the cylindrical roller bearing can be 260-500 mm, so that the cylindrical roller bearing 100 of the present disclosure occupies a small radial space as a whole, has a bearing radial installation space size basically equivalent to a hydraulic sliding bearing, and the inner diameter size of the inner hole of the planetary gear 200 matched therewith can be smaller, thereby greatly reducing the diameter of the planetary gear, allowing more planetary gears to be arranged, and thereby improving the torque density of the wind turbine gearbox.

[0059] In addition, the diameter Dw of the cylindrical roller 30 of the present disclosure is 25-35 mm, the length-diameter ratio λ of the cylindrical roller 30 is 2.7-4.5, and the ratio λ1 of the raceway width ∑Le of the inner ring 10 to the diameter Dw of the cylindrical roller 30 is 5.4-9, so that the outer diameter of the cylindrical roller 30 is reduced, and under the condition of the same bearing pitch circle diameter, more cylindrical rollers 30 can be arranged in the circumferential direction, thereby improving the stiffness of the overall cylindrical roller bearing 100, and more cylindrical rollers 30 means that there are more contact areas between the raceway 11 of the inner ring 10 and the planetary gear 200, and more contact areas can disperse the load, thereby reducing the pressure (also known as contact stress) borne by the single cylindrical roller 30 and the raceway 11, thereby improving the static load capacity of the bearing, the overall load capacity, and the theoretical calculation life under the application working condition.

[0060] Preferably, the ratio λ (i.e., the length-diameter ratio) of the axial length Lw of the cylindrical roller 30 to the diameter Dw is 3.2-4.5, and the ratio λ1 of the raceway width ∑Le of the inner ring 10 to the diameter Dw of the cylindrical roller 30 is 6.4-9, so that the cylindrical roller bearing 100 can ensure high bearing performance while also taking into account the demand for lightweight, miniaturization, and cost control of the wind turbine gearbox in this special application scenario.

[0061] Compared with the size of the cylindrical roller bearing of the traditional design logic, the carrying capacity of the cylindrical roller bearing 100 of the present disclosure is increased by 20%; the overall size of the cylindrical roller bearing 100 is reduced by 25%, the same linear speed is small, the friction torque is small in the starting stage and the running stage of the wind turbine gearbox, and the friction power consumption is reduced; the weight of the cylindrical roller bearing 100 is also reduced by 50%, not only reducing the material cost, but also the acceleration and deceleration working condition is very good, and the impact inertia is small. Compared with the traditional cylindrical roller bearing, the cylindrical roller bearing 100 of the present disclosure is more slender, but still belongs to the category of cylindrical roller bearings, and all the advantages of rolling bearings are gathered. When it is applied in the wind turbine gearbox, it still has the characteristics of maturity, stability and reliability. It has the same installation, disassembly and maintenance as the current cylindrical roller bearing, without changing any assembly process and assembly technology of the existing wind turbine gearbox, and still has the advantages of strong and long running ability with fault. Compared with the hydraulic sliding bearing, the cylindrical roller bearing 100 of the present disclosure has higher reliability, and does not need to additionally configure the complex lubrication system required by the hydraulic sliding bearing. Therefore, the cylindrical roller bearing 100 of the present disclosure not only has all the advantages of rolling bearings, but also has the same advantages of small size, low cost and high carrying capacity as hydraulic sliding bearings.

[0062] In some embodiments, the end faces of the cylindrical rollers 30 of adjacent rows located in the same row of pockets abut against each other. In this way, by allowing the end faces of the adjacent cylindrical rollers 30 to directly contact and abut against each other, the assembly process is simplified because the cylindrical rollers 30 are naturally aligned and axially support each other during assembly, reducing the need for adjustment and improving assembly efficiency and accuracy.

[0063] In addition, the end faces of the cylindrical rollers 30 abutting against each other enhance the overall rigidity and stability of the cylindrical roller bearing 100, especially when subjected to high axial load, and can better resist deformation.

[0064] In some embodiments, the axial lengths Lw of the at least two rows of cylindrical rollers 30 located in the same row of raceways 11 can be equal. Where two independent cylindrical roller bearings 100 are arranged axially to form a four-row integrated cylindrical roller bearing 100, the axial lengths of the cylindrical rollers 30 of the two independent cylindrical roller bearings 100 can be different.

[0065] When the axial lengths Lw of the cylindrical rollers 30 are equal, the cylindrical rollers 30 of uniform length simplify the manufacturing process, reduce the types of cylindrical rollers 30, facilitate mass production, and reduce costs.

[0066] In some embodiments, the axial lengths Lw of the at least two rows of cylindrical rollers 30 located in the same row of raceways 11 can be different, and the abutment points of the end faces of the at least two cylindrical rollers 30 in the circumferential direction are different.

[0067] When the axial length Lw of the cylindrical rollers 30 is not equal, it can be customized according to the actual application requirements to adapt to specific load distribution or space limitations, improving the overall adaptability and performance of the bearing. In addition, the abutting points or positions of the axial end faces of the two cylindrical rollers 30 in different pockets 23 are different in the circumferential direction W, so that the abutting points in each pocket 23 in the circumferential direction are staggered, which can reduce and disperse the impact and vibration during rolling, thereby reducing the operating noise and improving the smoothness of operation.

[0068] In some embodiments, as shown in FIG. 1, the retaining edge 12 can be a fixed retaining edge 12, and the two fixed retaining edges 12 at the axial ends of the inner ring 10 are integrally formed with the inner ring 10. The fixed retaining edge 12 is integrally formed with the inner ring 10, which simplifies the structure of the bearing and reduces the number of parts, which is not only conducive to machining and quality control, but also makes the bearing form a solid overall structure, improves the impact resistance and stability of the bearing, and is especially suitable for working conditions with high load or severe vibration. In addition, since the fixed retaining edge 12 is integrally formed with the inner ring 10, the precision of the raceway 11 is easier to control, ensuring the smoothness of the cylindrical rollers 30 when rolling in the raceway 11, and improving the operating efficiency and precision of the bearing. Figures 1 to 3 In some embodiments, the retaining edge 12 can be a movable retainer, and the two retaining edges 12 at the axial ends of the inner ring 10 are separately provided with the inner ring 10 and fixed with the inner ring 10 by interference fit or clamping. The movable retainer can be independently installed or replaced, without the need to replace the entire inner ring 10, reducing maintenance cost and complexity. At the same time, the movable retainer can adjust the axial position according to the actual application requirements, thereby adjusting the axial length of the raceway 11, improving the adaptability and flexibility of the cylindrical roller bearing 100 as a whole. In addition, the movable retainer allows a small amount of movement between the retainer and the inner ring 10, reducing the risk of deformation caused by temperature difference or mechanical stress, and prolonging the service life of the cylindrical roller bearing 100.

[0069] In some embodiments, the retaining edge 12 can be a floating retaining edge, i.e., the retaining edge 12 at one axial end of the inner ring 10 can be a fixed retaining edge, and the retaining edge 12 at the other axial end of the inner ring 10 can be a movable retainer. Similarly, the movable retainer of the floating retaining edge simplifies the assembly process, especially when it is necessary to accurately adjust the gap between the abutting end faces of the two cylindrical rollers 30, the movable retainer can provide the necessary axial adjustment space to ensure that the contact end faces of the adjacent two cylindrical rollers 30 have no gap, thereby reducing vibration and noise during operation and improving the smoothness of operation.

[0070]

[0071] ​In addition, the axial position of the movable retainer ring in the inner ring 10 can be adjusted according to different working conditions or load requirements of the bearing, so as to change the axial length of the raceway 11 on the outer periphery of the inner ring 10, and install cylindrical rollers 30 of different lengths or numbers.

[0072] In addition, when it is necessary to check or replace the damaged retainer 12, the movable retainer ring of the floating retainer facilitates partial repair or replacement, without the need to disassemble the entire inner ring 10, thereby reducing maintenance costs and prolonging the service life of the entire inner ring 10.

[0073] In some embodiments, the radially inner wall of the retainer 20 is in abutment with the radially outer wall of the retainer 12, and the retainer 20 is a retainer 20 guided by the retainer 12 of the inner ring 10, which corrects the operation of the cylindrical roller 30 through contact with the inner ring 10 in operation.

[0074] In some embodiments, the radially inner wall of the retainer 20 is in abutment with the radially outer wall of the retainer 12, and the retainer 20 is a retainer 20 guided by the retainer 12 of the inner ring 10, which corrects the operation of the cylindrical roller 30 through contact with the inner ring 10 in operation.

[0075] In some embodiments, the radially inner wall of the retainer 20 is radially spaced apart from the radially outer wall of the retainer 12 (not shown in the figure), and the spacing is large, so that the radially outer wall of the retainer 12 cannot be in abutment with the radially inner wall of the retainer 20 during operation of the bearing, and the side wall of the cylindrical roller 30 is used to abut against the retainer 20. Therefore, the retainer 20 is a retainer 20 guided by the cylindrical roller 30.

[0076] The direct contact of the retainer 20 with the cylindrical roller 30 without abutment with the retainer 12 provides greater freedom for the cylindrical roller 30, improves the flexibility and adaptability of the cylindrical roller bearing 100, avoids continuous friction between the retainer 20 and the retainer 12, reduces wear, helps to reduce noise, prolongs the service life of the retainer 20, and is particularly suitable for products that need to be continuously operated for a long time and have low noise.

[0077] The inner ring 10 of each independent cylindrical roller bearing 100 can be an integrally formed inner ring, as shown in Figure 1 The raceway 11 on the outer periphery of the inner ring 10 is formed on the outer wall of the independent inner ring 10. The integrally formed inner ring 10 simplifies the manufacturing process, reduces the assembly steps, and can improve the overall strength and stability of the inner ring 10.

[0078] In other embodiments, the inner ring 10 can also be a split formed inner ring, as shown in Figure 3As shown, the inner ring 10 includes two sub-inner rings 13 in the axial direction, each sub-inner ring 13 is provided with a rib 12 at one end in the axial direction, and the other end of each sub-inner ring 13 away from the rib 12 is open without a rib, and the other ends of the two sub-inner rings 13 away from the rib are in contact with each other, and the rib 12 of each sub-inner ring 13 and the outer wall of each sub-inner ring 13 jointly form the raceway 11 of the outer periphery of the inner ring 10, and the raceway 11 has a joint. In the environment where the installation space is limited, the inner ring 10 formed by splicing the sub-inner rings 13 not only ensures the integrity of the raceway 11, but also enhances the flexibility of assembly, which is convenient for installation and maintenance.

[0079] Based on the above, the inventors found that when the load capacity of the cylindrical roller bearing 100 is improved or two or more rows of cylindrical rollers 30 need to be installed in the pocket 23 of the retainer 20 due to size limitations, one side or both sides of the cylindrical roller 30 is not guided, which seriously affects the posture of the cylindrical roller 30 during operation, which can cause an increase in friction, and an increase in friction can cause damage to the coating and scratches on the rolling surface of the cylindrical roller 30 at the corresponding position, and more seriously, can cause early peeling of the surface of the cylindrical roller 30 and other failure phenomena. In the extreme case, the cylindrical roller 30 without correct guidance may be stuck in the pocket due to tilting, causing the bearing to be locked and other more serious consequences.

[0080] Therefore, based on the same inventive concept, the present disclosure also provides an improved retainer 20 applied in the cylindrical roller bearing described above, such as Figure 4 As shown, the retainer 20 can include a plurality of cross beams 21 and two edge beams 22. The circumferential two sides of the cross beam 21 are respectively provided with a cross beam inclined surface 211, such as Figure 6 As shown, the cross section of the cross beam 21 is hexagonal, and the two edge beams 22 are respectively connected to the axial two ends of the cross beam 21 and jointly enclose the cross beam 21 to form a plurality of closed-loop pockets 23, and the cross beam 21 and the edge beam 22 of the retainer 20 can be integrally formed.

[0081] In the embodiment of the present disclosure, two rows of cylindrical rollers 30 are arranged side by side in a single pocket 23. As shown in Figure 5 As shown, the same cross beam 21 can be provided with a plurality of protrusions in the axial direction for contacting and guiding the rolling surface of the cylindrical roller 30.

[0082] The protrusions include end protrusions 212 and intermediate protrusions 213. The end protrusion 212 is the protrusion at the end of the cross beam 21. On the same cross beam 21, two end protrusions 212 are provided, which are respectively located at the positions close to the two edge beams 22, but in one pocket 23, four end protrusions 212 are provided. The intermediate protrusion 213 is located in the middle of the end protrusion 212 in the axial direction A, and the intermediate protrusion 212 is axially spaced from the end protrusion 212, that is, the intermediate protrusion 212 and the end protrusion 212 are not connected.

[0083] In the present embodiment, as shown in Figure 7 and Figure 8 the end protrusion 212 is in connection with the edge beam 22 in the axial direction A, and the axial ends of the two cylindrical rollers 30 adjacent to each other in the pocket 23 are in abutment with the intermediate protrusion 213, and the end protrusion 212 and the intermediate protrusion 213 together provide stable axial support for the cylindrical rollers 30.

[0084] The number of intermediate protrusions 213 can be determined according to the number of cylindrical rollers 30 arranged side by side in the axial direction in a single pocket 23. Therefore, the intermediate protrusion 213 makes the rolling surface of each cylindrical roller 30 in a single pocket 23 have four contact positions with the cross beam 21. In other words, a single pocket 23 can form four support points for the rolling surface of each cylindrical roller 30, so that the cage 20 forms a stable support for each cylindrical roller 30, and the cylindrical roller 30 is more uniform in force during operation.

[0085] In addition, the end protrusion 212 and the intermediate protrusion 213 increase the material entity of the cross beam 21. Under the same structure, the cross beam 21 with more material entity can achieve greater structural strength and rigidity.

[0086] In addition, the inventor found that due to the side-by-side arrangement of two or more rows of cylindrical rollers 30 in a single raceway 11, the axial length of a single pocket 23 of the cage 20 is longer, and when the cage 20 is subjected to load during operation, the deformation displacement of the intermediate protrusion of the cross beam 21 is greater than that of the end protrusion. This deformation difference causes the intermediate protrusion of the cross beam 21 to easily produce a gap with the rolling surface of the cylindrical roller, so that the cage cannot correctly guide the cylindrical roller 30.

[0087] To solve the above technical problems, the present disclosure further provides an improved cage 20 applied to the above cylindrical roller bearing, in the direction perpendicular to the cross beam slope surface (i.e. the direction perpendicular to Figure 6 E), as shown in Figure 7 the highest position of the end protrusion 212 and the height difference of the cross beam slope surface 211 is h1, and the highest position of the intermediate protrusion 213 and the height difference of the cross beam slope surface 211 is h2, in the present embodiment, h1

[0088] Specifically, according to the empirical analysis of the contact stress of the cylindrical roller 30 acting on the cross beam 21, as shown in Figure 7 in the direction perpendicular to Figure 6The theoretical deformation displacement of the middle protrusion 213 is a2, which is greater than the theoretical deformation displacement of the end protrusion 212, which is a1. As the contact stress increases and the length of the beam 21 increases, the difference Δa between the theoretical deformation displacements of the middle protrusion 213 and the end protrusion 212 gradually increases.

[0089] As a result, the circumferential gap between the rolling surface of the cylindrical roller 30 and the middle protrusion 213 increases. When Δa increases to a certain value, the middle protrusion 213 cannot guide the cylindrical roller 30. Simulation results show that when the force between the beam 21 of the retainer 20 and the cylindrical roller 30 is in the range of hundreds to thousands of Newton, the difference Δa between the theoretical deformation displacements of the middle protrusion 213 and the end protrusion 212 satisfies 0 < Δa < 0.2 mm.

[0090] To improve the adverse consequences caused by the difference between the theoretical deformation displacements of the middle protrusion 213 and the end protrusion 212, the difference Δa between the theoretical deformation displacements is compensated by design optimization, that is, adjusting Δh. Meanwhile, Δh is not the larger the better. One reason is that simulation results show that the compensation effect does not increase linearly with Δh. The second reason is that the increase of Δh reduces the circumferential gap of the cylindrical roller 30 inside the pocket 23. When Δh increases to a certain value, the cumulative superposition of various related size tolerances will reduce the circumferential gap between the middle protrusion 213 of the beam 21 and the cylindrical roller 30 to a negative value, which will further cause the cylindrical roller 30 to be stuck at this position.

[0091] Therefore, h1 < h2 and 0 < Δh = h2 - h1 < 0.2 mm are designed. Simulation results show that, under the size limitation, when the force between the beam 21 of the retainer 20 and the cylindrical roller 30 is in the range of hundreds to thousands of Newton, the difference Δa between the theoretical deformation displacements of the middle protrusion 213 and the end protrusion 212 tends to 0, which ensures the function of the middle protrusion 213 guiding the cylindrical roller 30.

[0092] Meanwhile, to achieve the size limitation of h1 < h2 and 0 < Δh = h2 - h1 < 0.2 mm of the present disclosure, only a slight adjustment of the punching window process is required relative to the existing production process. At this stage, the difference Δh between the single-sided heights of the middle protrusion 213 and the end protrusion 212 is ensured, and the subsequent slope forming process does not need to be adjusted additionally, which can be quickly realized, and the production cost of the retainer 20 almost does not change, but a larger bearing performance benefit can be brought.

[0093] The beam 21 of the retainer 20 of the present disclosure is provided with the middle protrusion 213 and the end protrusion 212 in the axial direction, which compensates for the difference Δa between the theoretical deformation displacements of the two contact guiding regions, ensures that the circumferential gap between the middle protrusion 213 and the end protrusion 212 and the rolling surface of the cylindrical roller 30 is consistent when the force is borne during operation, and further ensures that the retainer 20 still has an effective guiding function under long-time operation.

[0094] As shown in Figure 9 , the cylindrical roller 30 can be divided into a modified region 31 and a straight segment region 32 along the axial direction A. Among them, the straight segment region 32 is the main load bearing region of the cylindrical roller 30 when bearing a larger radial load Fr (such as the direction shown in B in the figure). Figure 6 With the gradual increase of the radial load Fr, the wear and tear of the raised portion of the crossbeam 21 of the retainer 20 to the coating of the rolling surface of the cylindrical roller 30 also increases.

[0095] The end protrusions 212 and the intermediate protrusion 213 abut the axial ends of the cylindrical roller 30 as much as possible, so that in the axial A direction, the actual contact position of the raised portion (i.e. the contact guide region) of the crossbeam 21 of the retainer 20 with the rolling surface of the cylindrical roller 30 falls in the modified region 31 of the cylindrical roller 30, avoiding the main load bearing region, i.e. the straight segment region 32.

[0096] In this way, even in the case of a larger radial load Fr, the wear and tear of the raised portion to the coating of the rolling surface of the cylindrical roller 30 increases, but the wear and tear of the coating is located in the modified region 31, not in the straight segment region 32 of the cylindrical roller 30, i.e. the main load bearing region. In other words, it can maximize the guarantee that the coating of the main load bearing region outside the rolling surface of the cylindrical roller 30 is preserved on the surface of the cylindrical roller 30 for as long as possible during operation, improves the friction and wear of the rolling surface of the cylindrical roller 30 in the main load bearing region, effectively reduces the probability of white corrosion cracks in the straight segment region 32 of the cylindrical roller 30, and prolongs the service life of the cylindrical roller 30.

[0097] In some embodiments, as shown in Figure 8 , the axial spacing L1 between the starting positions of the two end protrusions 212 on the inside in the axial direction is related as follows: 0.50Lwn < L1 < 0.95Lwn; the axial length L2 of the intermediate protrusion 213 is: 0.05Lwn < L2 < 0.50Lwn, where Lwn is the sum of the axial lengths Lw of the n cylindrical rollers in a single pocket.

[0098] In some embodiments, the height difference h1 and / or h2 between the highest position of the raised portion and the height of the crossbeam slope 211 and the diameter Dw of the cylindrical roller 30 is related as follows: 0.1mm < h1 and / or h2 < 0.05*Dw.

[0099] In this embodiment, the surface of the raised portion can include a curved surface extending in the axial direction. In some embodiments, the surface of the raised portion can be a plane.

[0100] By Figure 6As shown, since the rolling surface of the cylindrical roller 30 is also a curved surface, and the cross section of the rolling surface of the cylindrical roller 30 is circular, when the convex part of the curved surface abuts against the rolling surface of the cylindrical roller 30, it is equivalent to the abutment of two convex curved surfaces, thereby reducing the contact area between the cylindrical roller 30 and the pocket 23 of the retainer 20, so that the friction in the contact area between the cylindrical roller 30 and the convex part is small, the wear of the coating on the rolling surface of the cylindrical roller 30 is also small, and scratches on the rolling surface are also avoided, thereby prolonging the service life of the cylindrical roller 30 and reducing the replacement frequency.

[0101] In the present embodiment, when the surface of the convex part includes a curved surface extending in the axial direction, then the intermediate convex part 213 includes two sub-convex parts, i.e. the intermediate convex part 213 actually has two highest positions for respectively contacting and guiding the rolling surfaces of the two adjacent cylindrical rollers 30. Preferably, the intermediate convex part 213 is an integrated guiding area, i.e. the two sub-convex parts are combined in the axial direction, and the surface of the intermediate convex part 213 has two curved surfaces, so that the production difficulty can be reduced and the production efficiency can be improved. In other embodiments, the two sub-convex parts can be arranged in the axial direction with a spacing.

[0102] In some embodiments, the curved surface can be a cylindrical surface, and the relationship between the radius R of the cylindrical surface and the axial length Lw of the cylindrical roller 30 is: R>0.09*Lw 2 .

[0103] When the surface of the convex part is a plane, the intermediate convex part 213 can also be an integral or divided into two sub-convex parts spaced in the axial direction. The height difference h2 of the surface of the intermediate convex part 213 relative to the surface of the cross beam 21 is consistent.

[0104] Based on the same inventive concept, the present disclosure also provides a bearing assembly, such as Figure 10 As shown, the bearing assembly is used in a wind power gear box. The bearing assembly includes a mandrel 300, a planetary gear 200, a planet carrier 400, and the cylindrical roller bearing 100 described above, the planetary gear 200 is sleeved on the outside of the mandrel 300, and the cylindrical roller bearing 100 is rollingly arranged between the mandrel 300 and the planetary gear 200, for enabling the planetary gear 200 to rotate relative to the mandrel 300, wherein at least two groups of cylindrical roller bearings 100 are arranged in the axial direction A, and the inner circumferential surface of the planetary gear 200 is provided with a radially convex intermediate rib 201.

[0105] Wherein, as Figure 10As shown, the middle retaining edge 201 abuts against the end faces of two adjacent cylindrical rollers 30 of two adjacent cylindrical roller bearings 100, respectively, and the retaining edge 12 of the inner ring 10 of the cylindrical roller bearing 100 abuts against the end face of the end cylindrical roller 30, so it can be known that the axial force Fa of the planetary gear 200 on the cylindrical roller 30 can be transmitted to the axial end face of the cylindrical roller 30 through the middle retaining edge 201, and the axial force Fa is transmitted through a plurality of cylindrical rollers 30 and finally transmitted to the retaining edge 12 of the inner ring 10. It can be known that the middle retaining edge 201 can be used to transmit the axial force Fa of the cylindrical roller 30 from the planetary gear 200, thereby improving the overall stability and carrying capacity of the bearing assembly.

[0106] The specific manner of the bearing assembly in the above embodiments and the functions achieved in the wind power gear box have been described in detail in the embodiments related to the cylindrical roller bearing 100 and the cage 20, and will not be described in detail here.

[0107] It can be understood that "a plurality of" in the present disclosure means two or more, and other quantifiers are similar. The association relationship of "and / or" describing the associated objects means that there can be three relationships, for example, A and / or B can represent: A exists alone, A and B exist simultaneously, and B exists alone. The character " / " generally represents that the associated objects before and after are in an "or" relationship. The singular form "a", "said" and "the" are also intended to include the plural form, unless the context clearly indicates otherwise.

[0108] Other embodiments of the present disclosure will be readily apparent to those skilled in the art upon considering the specification and practice of the utility model disclosed herein. The present application is intended to cover any variations, uses, or adaptive changes to the present disclosure that follow the general principles of the present disclosure and include common knowledge or conventional technical means in the technical field of the present disclosure not disclosed by the present disclosure. The specification and examples are only considered as exemplary, and the true scope and spirit of the present disclosure are indicated by the following right range.

[0109] It should be understood that the present disclosure is not limited to the precise structures described above and shown in the drawings, and various modifications and changes can be made without departing from the scope thereof. The scope of the present disclosure is limited only by the appended right range.

Claims

1. A cylindrical roller bearing (100) for a wind turbine gearbox, comprising: an inner ring (10) provided with a row of raceways (11) on an outer periphery thereof; a cage (20) provided outside the inner ring (10) and provided with a plurality of pockets (23) equi-spaced in a circumferential direction (W), and cylindrical rollers (30) located in the raceways (11), characterized in that at least two rows of the cylindrical rollers (30) are arranged side by side in the raceways (11) in an axial direction (A), at least two of the cylindrical rollers (30) are arranged side by side in each of the pockets (23) in the axial direction (A), and a diameter Dw of the cylindrical rollers (30) is 25-35 mm, and a ratio λ of an axial length Lw of the cylindrical rollers (30) to the diameter Dw is 2.7-4.

5.

2. The cylindrical roller bearing (100) for a wind power gear box according to claim 1, characterized in that a diameter Dw of the cylindrical rollers (30) is 25-35 mm, and a ratio λ of an axial length Lw of the cylindrical rollers (30) to the diameter Dw is 3.2-4.

5.

3. A cylindrical roller bearing (100) for a wind turbine gearbox, comprising: an inner ring (10) provided with a row of raceways (11) on an outer periphery thereof; a cage (20) provided outside the inner ring (10) and provided with a plurality of pockets (23) equi-spaced in a circumferential direction (W), and cylindrical rollers (30) located in the raceways (11), characterized in that at least two rows of the cylindrical rollers (30) are arranged side by side in the raceways (11) in an axial direction (A), at least two of the cylindrical rollers (30) are arranged side by side in each of the pockets (23) in the axial direction (A), and a ratio λ1 of a raceway width ∑Le of the inner ring (10) to a diameter Dw of the cylindrical rollers (30) is 5.4-9.

4. The cylindrical roller bearing (100) for a wind power gear box according to claim 3, characterized in that a ratio λ1 of a raceway width ∑Le of the inner ring (10) to a diameter Dw of the cylindrical rollers (30) is 6.4-9.

5. The cylindrical roller bearing (100) for a wind power gear box according to any one of claims 1 to 4, characterized in that Rolling surfaces of the cylindrical rollers (30) are in direct rolling contact with an outer peripheral surface of the inner ring (10) and are configured to be in direct rolling contact with an inner peripheral surface of a bore of a planetary gear (200) of the wind turbine gearbox.

6. The cylindrical roller bearing (100) for a wind power gear box according to any one of claims 1 to 4, characterized in that End faces of the cylindrical rollers (30) of adjacent rows located in the same row of the raceways (11) abut against each other.

7. The cylindrical roller bearing (100) for a wind turbine gearbox according to any one of claims 1 to 4, characterized in that: axial lengths Lw of the at least two cylindrical rollers (30) located in the same pocket (23) are not equal.

8. The cylindrical roller bearing (100) for a wind power gear box according to claim 7, characterized in that End faces of the two rows of the cylindrical rollers (30) abut against each other at different axial positions in the circumferential direction.

9. The cylindrical roller bearing (100) for a wind turbine gearbox according to any one of claims 1 to 4, characterized in that: the raceways (11) are delimited by two radially outwardly provided, radially projecting and axially spaced flanges (12) of the inner ring (10).

10. The cylindrical roller bearing (100) for a wind power gear box according to claim 9, characterized in that The retaining flange (12) is a fixed retaining flange (12) integrally formed with the inner ring (10) and / or a movable retaining ring separately formed with the inner ring (10).

11. The cylindrical roller bearing (100) for a wind turbine gearbox according to claim 9, characterized in that The inner ring (10) is integrally formed or spliced by two sub-inner rings (13).

12. The cylindrical roller bearing (100) for a wind turbine gearbox according to claim 9, characterized in that, The inner diameter of the cylindrical roller bearing (100) is 200-350 mm, and the diameter of the envelope circle of the cylindrical roller bearing (100) is 260-500 mm. The axial width of the inner ring (10) is 180-440 mm.

13. Cage (20) for application in a cylindrical roller bearing (100) according to any one of claims 1 to 12, characterized in that Comprising: A plurality of crossbeams (21) are provided with crossbeam inclined surfaces on the circumferential sides thereof; Two edge beams (22) are connected to the axial ends of the crossbeams (21) and enclose the crossbeams (21) to form a plurality of closed-loop pockets, and at least two cylindrical rollers (30) are arranged side by side in the pockets in the axial direction, Wherein, the crossbeam inclined surfaces are provided with protrusions for contact guidance with the rolling surfaces of the cylindrical rollers (30), the protrusions include end protrusions (212) and intermediate protrusions (213), the intermediate protrusions (213) are used for contact guidance with the rolling surfaces of two adjacent cylindrical rollers (30), and the intermediate protrusions (213) make the rolling surface of each cylindrical roller (30) in the pocket (23) have four contact positions with the crossbeam (21).

14. Cage (20) for application in a cylindrical roller bearing (100) according to any one of claims 1 to 12, characterized in that Comprising: A plurality of crossbeams (21) are provided with crossbeam inclined surfaces on the circumferential sides thereof; Two edge beams (22) are connected to the axial ends of the crossbeams (21) and enclose the crossbeams (21) to form a plurality of closed-loop pockets, and at least two cylindrical rollers (30) are arranged side by side in the pockets in the axial direction, Wherein, the crossbeam inclined surfaces are provided with protrusions for contact guidance with the rolling surfaces of the cylindrical rollers (30), The protrusions include end protrusions (212) and intermediate protrusions (213), in the direction perpendicular to the crossbeam inclined surfaces, the height difference between the highest position of the end protrusions (212) and the crossbeam inclined surfaces is h1, the height difference between the highest position of the intermediate protrusions (213) and the crossbeam inclined surfaces is h2, and h1 < h2, 0 < Δh = h2 - h1 < 0.2 mm.

15. The retainer (20) according to claim 13 or 14, characterized in that, The axial spacing L1 between the starting positions of the two end protrusions (212) on the inner side in the axial direction satisfies the relationship: 0.50Lwn < L1 < 0.95Lwn; The axial length L2 of the intermediate protrusions satisfies the relationship: 0.05Lwn < L2 < 0.50Lwn, Wherein, Lwn is the sum of the axial lengths Lw of n cylindrical rollers (30) in a single pocket.

16. The retainer (20) according to claim 13 or 14, characterized in that, The surface of the protrusion comprises a curved surface extending in axial bending, and the intermediate protrusion (213) comprises two sub-protrusions, which make the intermediate protrusion (213) have two highest positions for respectively contacting and guiding the rolling surfaces of two adjacent cylindrical rollers (30).

17. The cage (20) according to claim 16, characterized in that, The radius R of the curved surface is in the relationship R > 0.09 * Lw with the axial length Lw of the cylindrical roller (30) 2 .

18. The cage (20) according to claim 16, characterized in that, The two sub-protrusions are spaced apart in the axial direction.

19. Cage (20) according to claim 13 or 14, characterized in that The surface of the protrusion is a flat surface.

20. The cage (20) according to claim 19, characterized in that, The intermediate protrusion (213) comprises two axially spaced sub-protrusions.

21. The cage (20) according to claim 13 or 14, characterized in that, The highest position of the protrusion has a height difference H from the inclined surface of the cross beam (21) and the diameter Dw of the cylindrical roller (30) in the following relationship: 0.1mm<H<0.05*Dw, wherein H is h1 and / or h2.

22. A bearing assembly, characterized in that comprising: a mandrel (300); a planetary gear (200) sleeved outside the mandrel (300); the cylindrical roller bearing (100) for a wind power gear box as claimed in any one of claims 1 to 12, which is arranged between the mandrel (300) and the planetary gear (200) for rotating the planetary gear (200) relative to the mandrel (300), wherein, at least two sets of the cylindrical roller bearing (100) are arranged in the axial direction of the mandrel (300), and the inner circumferential surface of the planetary gear (200) is provided with a radially protruding intermediate rib (201), which respectively abuts the end surface of the cylindrical roller (30) of the adjacent cylindrical roller bearing (100).

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