Sealed ball bearing and bearing device

Abstract

To suppress deformation of a cage while preventing intrusion of foreign matter into an internal space of a bearing.SOLUTION: A sealed ball bearing which is used under an environment at a dmn value of 700,000 or more, comprises a seal member 20 capable of being brought into a fluid lubrication state, and a cage 10 is a crown-shaped cage containing engineering plastic. In the bearing: a dimensional relation between a diameter φx of a ball 5, a radius Ry of a radial-direction inner peripheral surface D of a pocket 11 in a cross-section which passes through a pocket center C of the pocket 11 and includes an axial center O of the bearing, and a radius Rz of a circumferential-direction inner peripheral surface E of the pocket 11 in a cross-section which, at the pocket center C, is orthogonal to a straight line in the bearing radius direction, which passes through the pocket center C, is φx<2Ry<2Rz; the circumferential-direction inner peripheral surface E comprises a first inner peripheral surface E1 on a base 12 side and a second inner peripheral surface E2 on a tip side; and a center C2 of the second inner peripheral surface E2 is further eccentric to the base 12 side than a center C1 of the first inner peripheral surface E1.SELECTED DRAWING: Figure 4

Description

[Technical field]

[0001] The present invention relates to a sealed ball bearing and a bearing device. [Background technology]

[0002] For example, in various vehicles such as automobiles and construction machines, and various industrial machines, many bearings are used in the motor shafts of transmissions (speed-up / down gears) equipped with drive motors. Bearings used in these devices are generally used under high-speed conditions compared to bearings used to support shafts in general devices. The lubricating oil used in these devices contains foreign matter such as gear wear powder. For this reason, a seal member is provided at the end of the bearing to prevent foreign matter from entering the inside of the bearing, thereby suppressing the reduction in the bearing life, and by suppressing the amount of lubricating oil entering the bearing, the stirring resistance is reduced.

[0003] For example, in Patent Document 1, a seal lip provided on a seal member is provided with a protrusion to create an oil passage that connects the inside and outside of the bearing, thereby creating a fluid lubrication state between the seal lip and the seal sliding surface. This suppresses the reduction in bearing life caused by gear wear debris and the like, while realizing fluid lubrication of the seal portion and reducing seal torque.

[0004] Patent Document 2 discloses a technology in which a crown cage has an annular base and pillars protruding from the base, the radial thickness of the crown cage gradually decreases from the base to the tips of the pillars, and the radial gap between the outer peripheral surface of the base and the inner peripheral surface of the shoulder of the outer ring is larger than the radial gap between the inner peripheral surface of the base and the outer peripheral surface of the shoulder of the inner ring.

[0005] According to Patent Document 2, even if the lubricant is scattered by centrifugal force during high-speed rotation, the lubricant is stored between the outer peripheral surface of the base and the inner peripheral surface of the shoulder of the outer ring, preventing the lubricant from flowing into the raceway surface of the outer ring. This reduces the stirring resistance, making it possible to reduce torque and heat generation, and also prevents deformation of the cage due to centrifugal force during high-speed rotation. [Prior art documents] [Patent documents]

[0006] [Patent Document 1] International Publication No. 2016 / 143786 [Patent Document 2] JP 2021-195973 A Summary of the Invention [Problem to be solved by the invention]

[0007] In Patent Document 1, it is possible to suppress the decrease in bearing life. However, in the case of bearings used under high-speed conditions such as supporting the motor shaft of a drive motor, there is a concern that the raceway ring may heat up due to interference with the cage and the cage may be damaged under harsher operating conditions than expected.

[0008] In addition, in Patent Document 2, measures are taken to prevent damage to the cage under high-speed conditions. However, although this bearing has seal members at both axial ends of the bearing, the seal members are non-contact seals, and therefore cannot prevent the intrusion of foreign matter large enough to affect the bearing life. For this reason, when this bearing is used in a device such as a transmission or a speed increase / decrease gear, there is concern that the bearing life will be reduced due to the intrusion of foreign matter into the bearing. In addition, in recent years, devices that integrate a drive motor, an inverter, and a gear (transmission, etc.) are becoming more widespread. In such devices, oil lubrication is mainstream, and the bearings that support the motor shaft of the drive motor also tend to be lubricated with the same oil as the inside of the transmission, etc. For this reason, there is a demand for more reliable prevention of the intrusion of foreign matter into the bearing.

[0009] Therefore, an object of the present invention is to suppress deformation of the cage under high speed conditions while preventing intrusion of foreign matter into the internal space of the bearing. [Means for solving the problem]

[0010] In order to solve the above problems, the present invention provides a bearing comprising an inner ring, an outer ring, balls arranged between the inner ring and the outer ring, a cage having pockets for holding the balls in a circumferential direction, and a seal member for closing an opening at an axial end of an internal space of a bearing formed between the inner ring and the outer ring, dmn={(D+d) / 2}×n D: Bearing outer diameter (mm) d: bearing inner diameter (mm) n: Rotation speed (min -1 ) For ball bearings used in environments where the dmn value specified in is 700,000 or more, the seal member is provided with a seal lip that is fixed to one of the inner ring and the outer ring and that slides against a seal sliding surface set on the other of the inner ring and the seal lip has a plurality of protrusions arranged in a circumferential direction, the plurality of protrusions being formed in a manner that generates gaps between adjacent protrusions in the circumferential direction and that enables a fluid lubrication state between the seal lip and the seal sliding surface by an oil film of lubricating oil that is drawn into between the protrusions and the seal sliding surface from the gaps as the bearing rotates, the cage is a crown cage including an engineering plastic and including an annular base and a plurality of pillars protruding in one direction from the base, wherein a dimensional relationship between a diameter φx of the balls, a radius Ry of a radial inner peripheral surface of the pocket in a cross section passing through a center of the pocket and including an axis of the bearing, and a radius Rz of a circumferential inner peripheral surface of the pocket in a cross section intersecting at the center of the pocket at right angles to a radial line of the bearing passing through the center of the pocket is φx<2Ry<2Rz, The circumferential inner surface includes a first inner surface on the base side and a second inner surface closer to the tip of the column portion than the first inner surface, and a sealed ball bearing is used in which the center of the second inner surface is eccentric toward the base side relative to the center of the first inner surface (Configuration 1).

[0011] In configuration 1, a configuration can be adopted in which, with respect to the radius Rz of the circumferentially inner peripheral surface, a radius Rz1 of the first inner peripheral surface and a radius Rz2 of the second inner peripheral surface are set equal to each other (configuration 2).

[0012] In addition, in the configuration 1, a configuration can be adopted in which a recess is provided as an oil reservoir between the first inner circumferential surface and the second inner circumferential surface (configuration 3).

[0013] In addition, in configuration 1, a configuration can be adopted in which the seal lip is made of one or more materials selected from nitrile rubber, acrylic rubber, and fluororubber (configuration 4).

[0014] Furthermore, in configuration 1, when the supply of lubricant to the internal space of the bearing is in one direction from one axial end side to the other axial end side, a configuration can be adopted in which the sealing member is provided only at the opening on the one axial end side (configuration 5).

[0015] Furthermore, a plurality of elements selected from among the above-mentioned elements 2 to 5 can be added to element 1. That is, elements 2 and 3, elements 2 and 4, elements 2 and 5, elements 2 and 3 and 4, elements 2 and 3 and 5, elements 2 and 4 and 5, elements 2 and 3 and 4 and 5, and elements 2 and 3 and 4 and 5, etc. can be added to element 1.

[0016] A bearing device can be employed in which a rotating shaft of a drive motor, reducer, or speed increaser for electric transport equipment is supported by the sealed ball bearing according to any of these embodiments. Effect of the Invention

[0017] This invention uses a seal member that can maintain a fluid lubrication state between the seal lip and the seal sliding surface, thereby reducing the seal torque and controlling the amount of lubricant flowing into the bearing internal space. This reduces the agitation resistance of the lubricant in the bearing internal space, making it possible to use the bearing in a high-speed rotation range with a dmn value of 700,000 or more. At the same time, it also prevents the intrusion of foreign matter large enough to affect the bearing life.

[0018] In addition, by using engineering plastic as the material for the crown-shaped cage, the dimensional relationship between the diameter φx of the balls in the ball bearing, the radius Ry of the radial inner surface of the cage pocket, and the radius Rz of the circumferential inner surface has been optimized. Furthermore, the center of the second inner surface of the circumferential inner surface of the pocket on the side farther from the base is offset toward the base more than the center of the first inner surface on the base side, thereby suppressing deformation of the cage during high-speed operation and preventing foreign matter from becoming caught between the cage and the sealing member. [Brief description of the drawings]

[0019] [Figure 1] FIG. 1 is a vertical cross-sectional view showing an embodiment of the present invention. [Diagram 2] Enlarged view of the main part of Figure 1 [Diagram 3] A vertical cross-sectional view showing the main part of the cage [Figure 4] A cross-sectional view showing the main part of the cage, perpendicular to the radial direction of the bearing (cross-sectional view taken along the line IV-IV in FIG. 3) [Diagram 5] Schematic diagram showing the relationship between the cage and balls [Figure 6] Enlarged view of the main area near the seal ship [Figure 7A] Right side view of Figure 6 [Figure 7B] Enlarged view of the main part of Figure 7A [Figure 8] FIG. 11 is a vertical cross-sectional view showing a modified example. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0020] DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described with reference to the drawings. This embodiment is a rolling bearing 1 equipped with a seal member 20 at an opening at an axial end of the internal space of the bearing.

[0021] As shown in Figs. 1 to 3, the rolling bearing 1 comprises an inner ring 3, an outer ring 4, a plurality of rolling elements 5 arranged between the inner ring 3 and the outer ring 4, and a cage 10 having pockets 11 that hold the rolling elements 5 in the circumferential direction. Since spheres (steel balls) are used as the rolling elements 5, they will be referred to as balls 5 hereinafter. Furthermore, this rolling bearing 1 will be referred to as a sealed ball bearing 1 or simply as a bearing 1 hereinafter. Furthermore, the direction along the bearing central axis of the bearing 1 will be referred to as the "bearing axial direction" or simply as the "axial direction", the direction perpendicular to the axial direction will be referred to as the "bearing radial direction" or simply as the "radial direction", and the circumferential direction around the bearing central axis will be referred to as the "bearing circumferential direction" or simply as the "circumferential direction".

[0022] The cage 10 is a crown-shaped cage molded from engineering plastic. The cage 10 has an annular base 12 and a number of pillars 13 protruding from the base 12 in the axial direction. A pair of pillars 13, 13 are arranged in parallel along the circumferential direction and are provided at regular intervals along the circumferential direction. A concave pocket 11 is formed between the pair of pillars 13. The outer diameter surface of the cage 10 is a curved surface (cylindrical surface) without any steps. The outer diameter surface and inner diameter surface of the cage 10 are connected at the pocket 11.

[0023] The tip of the pillar portion 13 is a retaining claw 14. The retaining claws 14, 14 on both sides sandwiching the pocket 11 are curved in a direction approaching each other. Adjacent pillar portions 13, 13 may be connected to each other between circumferentially adjacent pockets 11. The balls 5 held by the pocket 11 revolve between the raceway surface 3a of the inner ring 3 and the raceway surface 4a of the outer ring 4 while being held by the pocket 11.

[0024] A rotating shaft (not shown) is fixed to the inner diameter portion 3b of the inner ring 3, and the inner ring 3 rotates in the circumferential direction together with the rotating shaft. The outer ring 4 is a member such as a housing or gear (not shown) and is attached to a fixed member that bears the load from the rotating shaft. In this way, the bearing 1 supports the rotating shaft rotatably relative to the fixed member. Examples of the rotating shaft include the rotating shaft of a drive motor equipped in an electric transport device such as an electric vehicle, or the rotating shaft of a reducer or speed increaser equipped in such an electric transport device. The bearing center axis of the bearing 1 and the rotation center axis of the rotating shaft are set coaxially.

[0025] When the bearing 1 is assembled, an appropriate lubricant such as grease is sealed in the bearing internal space A. During use, a lubricant (lubricating oil) is supplied from the outside into the bearing internal space A through an opening at one axial end of the bearing internal space A. The lubricant (lubricating oil) in the bearing internal space A flows out of the bearing internal space A from an opening at the other axial end of the bearing internal space A. The bearing 1 is basically used under oil lubrication, and the grease sealed in as initial lubrication is then replaced with lubricating oil supplied from outside the bearing. When sealing in grease for initial lubrication, it is desirable to seal in 5 to 20% of the total spatial volume of the bearing internal space A.

[0026] The lubricating oil that lubricates devices such as the drive motor and the transmission contains wear powder from gears, wear powder from clutches, and other foreign matter depending on the device into which the bearing 1 is installed. It is desirable that these foreign matter are captured somewhere rather than floating in the lubricating oil. In addition, the bearing 1 is lubricated with the same lubricating oil as the devices such as the drive motor and the transmission. For this reason, a seal member 20 is attached to the opening at the axial end of the bearing internal space A. In this embodiment, a seal member 20 is provided at each of both axial ends of the bearing internal space A.

[0027] The seal member 20 is an annular member that covers the opening at the axial end of the bearing internal space A. The seal member 20 separates the bearing internal space A from the outside. Since foreign matter is present in the lubricating oil on the outside of the bearing side across the seal member 20, the seal member 20 prevents this foreign matter from entering the bearing internal space A from the outside of the bearing.

[0028] As shown in Fig. 2, the seal member 20 includes a metal core 23 and an elastic portion 24 fixed integrally to the core 23. The core 23 is an annular member formed to have an L-shaped cross section over the entire circumference in the circumferential direction. In this embodiment, the core 23 is formed as a pressed product. The elastic portion 24 is made of rubber and is vulcanization bonded to the core 23. The vulcanization bonding can be performed, for example, by placing the core 23 in a mold and vulcanizing a vulcanized rubber material.

[0029] The elastic portion 24 of the seal member 20 includes a fitting portion 25 that protrudes radially outward on the outer diameter side, a main body portion 26 that covers the core metal 23, a protruding portion 22 that protrudes toward the inner diameter side from the main body portion 26, and a seal lip 21 that protrudes radially inward like a tongue at the tip of the protruding portion 22. A seal groove 8 is formed around the entire circumference of the inner periphery at the axial end of the outer ring 4. The seal member 20 is fixed to the outer ring 4 by fitting the fitting portion 25 into the seal groove 8.

[0030] A seal sliding surface B that slides in the circumferential direction against the seal lip 21 is formed on the outer periphery of the inner ring 3. The seal sliding surface B has a cylindrical surface shape that extends over the entire circumferential direction.

[0031] As shown in Fig. 6, the seal lip 21 has a waist portion 22 formed in a radially continuous annular shape with a constant width in the axial direction, and a head portion formed in a protruding piece shape bending outward from the waist portion 22. An interference is set between the head portion of the seal lip 21 and the seal sliding surface B. When the seal member 20 is attached in the predetermined arrangement shown in Figs. 1 and 2, the seal lip 21 is pressed against the seal sliding surface B by the interference, causing a rubber-like elastic deformation bending axially outward, generating a tension force of the seal lip 21. An attachment error, manufacturing error, etc. of the seal member 20 is absorbed by a change in the degree of bending of the seal lip 21.

[0032] As shown in Fig. 7A, the seal lip 21 has a plurality of protrusions 27 arranged in parallel along the circumferential direction. The protrusions 27 extend in a direction perpendicular to the circumferential direction over their entire length, and are arranged at uniform intervals over the entire circumference in the circumferential direction. Therefore, the gaps 28 formed between the protrusions 27 adjacent in the circumferential direction are also formed at uniform intervals over the entire circumference in the circumferential direction. The gaps 28 form an oil passage between the seal lip 21 and the seal sliding surface B, which communicates between the bearing internal space A and the outside of the bearing.

[0033] 7B, the projections 27 have circumferential ends 29 formed such that the gap between them and the seal sliding surface B increases from the center of the circumferential width toward both sides in the circumferential direction. That is, the projections 27 form a wedge-shaped gap between them and the seal sliding surface B, which is larger on the gap 28 side and smaller on the side closer to the center p of the circumferential width of the projections 27.

[0034] As shown in FIG. 6, the projection 27 has an area that is generally along the seal sliding surface B on a virtual plane including the bearing center axis. This area exists with a width in the direction along the seal sliding surface B (corresponding to the left-right direction (axial direction) in FIG. 6). For this reason, the oil film formation is promoted by the sliding part between the projection 27 and the seal sliding surface B accompanying the rotation of the bearing, that is, the wedge effect when the projection 27 drags the lubricating oil in the gap 28 in the circumferential direction toward the wedge-shaped gap between the projection 27 and the seal sliding surface B. The area where the oil film is interposed between the projection 27 and the seal sliding surface B occurs with a finite length of a predetermined length or more in the direction along the seal sliding surface B (axial direction) on the above-mentioned virtual plane. Since such a sliding part between the projection 27 and the seal sliding surface B is considered to occur in the shape of a contact ellipse based on Hertz's elastic contact theory, the major axis of the contact ellipse corresponds to the above-mentioned finite length.

[0035] When the circumferential speed of the relative rotation between the seal lip 21 and the seal sliding surface B is less than a certain value, microscopically, a boundary lubrication state or a mixed lubrication state including a solid contact region is formed. When the bearing rotation speed increases and the circumferential speed of the relative rotation between the protrusions 27 and the seal sliding surface B exceeds a certain value, the oil film thickness between the protrusions 27 and the seal sliding surface B exceeds the composite roughness σ between the protrusions 27 and the seal sliding surface B by a large margin, and each protrusion 27 and the seal sliding surface B are completely separated by an oil film, resulting in a fluid lubrication state. This allows the seal lip 21 and the seal sliding surface B to be completely separated by an oil film, resulting in a fluid lubrication state. When such a fluid lubrication state is formed, the seal torque by the seal member 20 can be reduced to the same level as that of a non-contact seal, which in turn suppresses the temperature rise of the sealed bearing and prevents the adsorption of the seal lip 21.

[0036] Here, if the oil film parameter Λ is 3 or more, the lubrication mode of the sliding part is considered to be in a hydrodynamic lubrication state. The oil film parameter Λ is the ratio of the composite roughness σ to the minimum oil film thickness h0 in the sliding part, and Λ = h0 / σ. The minimum oil film thickness h0 is calculated based on the theory of elastohydrodynamic lubrication. Composite roughness σ = √(Rq1 2 +Rq2 2) Rq1 is the root mean square roughness of the seal sliding surface B that constitutes the sliding portion described above. Rq2 is the root mean square roughness of the surface of the projection 27, where the root mean square roughness is the value (μm) of the root mean square roughness Rq defined in JIS (B0601:2013).

[0037] The oil film parameter Λ depends on the composite roughness σ, and the smaller the composite roughness σ, the thicker the oil film can be. In order to make the sliding part between the protrusion 27 and the seal sliding surface B in a fluid lubrication state even when the circumferential speed is extremely low, it is preferable to make the composite roughness σ at the sliding part 0.9 μm or less. For example, when the oil lubrication mode was determined using the Johnson chart under the calculation conditions of a composite roughness σ of 0.9 μm, a lubricant oil of transmission oil (30 cst, 40°C), an ambient temperature of 20°C, and a circumferential speed of 0.2 m / s, the minimum oil film thickness h0 was 2.8 μm, the oil film parameter Λ was 3 or more, and the lubrication mode was EI mode. Therefore, if the composite roughness σ of the protrusion 27 and the seal sliding surface B is 0.9 μm or less, it is expected that the lubrication state will be reliably fluid lubricated in the actual use range of the bearing.

[0038] For example, in applications that support rotating parts in a vehicle transmission, transmission oil is generally supplied as lubricant to a sealed bearing by a suitable method such as splashing or oil bath. The lubricant is circulated by an oil pump and filtered by an oil filter installed in the circulation path. It is considered that if large foreign matter with a particle size exceeding 0.05 mm enters the bearing internal space A, it will adversely affect the bearing life. If the protruding height h of the protrusion 27 (see FIG. 7B) is set to 0.07 mm or less, a gap 28 through which such large foreign matter cannot easily pass can be generated. Note that in order to improve the oil permeability of the gap 28, it is desirable to set the protruding height h of the protrusion 27 to 0.05 mm or more.

[0039] When the protrusion height h of the protrusion 27 is 0.07 mm or less, for example, the interval between adjacent protrusions 27 in the circumferential direction can be set to 0.3 mm or more and 2.6 mm or less, the circumferential width of the protrusion 27 can be set to 0.2 mm or more and 1.0 mm or less, and the radius of curvature of the surface of the protrusion 27 can be set to 0.15 mm or more and less than 2.0 mm. In this example, when the oil temperature is 30 to 120°C and the relative circumferential speed of the seal lip 21 and the seal sliding surface B is 0.2 m / s or more, it is considered that the lubrication mode is either the isoviscosity-rigid body region (RI mode) or the isoviscosity-elastic body region (EI mode, soft EHL) in the lubrication region diagram (Johnson chart) based on the viscosity parameter gv and the elastic parameter ge, which are dimensionless numbers determined by Greenwood-Johnson, that is, the above-mentioned fluid lubrication state. When the interval between adjacent projections 27 in the circumferential direction is 2.6 mm, an oil film of about 3 μm is formed between the projections 27 and the seal sliding surface B, and when the interval is less than 2.6 mm, the oil film tends to become thicker. When the interval is 2.6 mm or less, the bearing rotation torque tends to decrease (i.e., the seal torque tends to decrease). When the interval is less than 0.3 mm, it becomes difficult to form a transfer surface for molding the projections 27 in the mold by end milling.

[0040] As described above, by maintaining a fluid lubrication state between the seal lip 21 and the seal sliding surface B, the friction (seal torque) caused by rubbing between the seal lip 21 and the seal sliding surface B can be substantially reduced to zero, the seal lip 21 is substantially not worn, and heat generation due to sliding between the seal lip 21 and the seal sliding surface B can be suppressed. Furthermore, the peripheral speed of the relative rotation allowed between the seal lip 21 and the seal sliding surface B is increased, making it possible to operate under higher speed conditions than before.

[0041] The seal lip 21 may be made of one or more materials selected from the group consisting of nitrile rubber, acrylic rubber, and fluororubber. These materials may be used for only the seal lip 21, or may be used for the entire elastic portion 24 including the seal lip 21.

[0042] The pocket 11 of the cage 10 corresponds to the balls 5 with a diameter of φx. As shown in FIG. 3, the inner surface of the pocket 11 is an inner surface (spherical surface) that draws an arc along the bearing radial direction, and its radius in the bearing radial direction is set to Ry. 2Ry is twice Ry and corresponds to the diameter in the bearing radial direction of the inner surface of the pocket 11. That is, when the center of the pocket 11 is the pocket center C, the radius of the inner circumferential surface D (hereinafter referred to as the radial inner circumferential surface D) of the pocket 11 in a cross section (hereinafter referred to as the radial cross section) that passes through the pocket center C and includes the axis O of the bearing is Ry, and the diameter is 2Ry (hereinafter referred to as the radial pocket diameter 2Ry). The pocket center C and the center of the balls 5 are coincident by design. The center of the arc of the radial inner circumferential surface D is indicated by the symbol C' in FIG. 3, and is eccentric toward the radial inner circumferential surface D from the pocket center C. In Fig. 3, it is shown that the radial inner peripheral surface D continues from the outer diameter side end d2 to the inner diameter side end d1 at the bottom F of the pocket 11. Although Fig. 3 shows one cross section passing through the center line of the bearing, this radial pocket diameter 2Ry is set not only in this cross section but also in any cross section including a radial line of the bearing connecting the pocket center C and the axis O of the bearing (however, in cases where a recess H such as an oil reservoir is provided on the inner surface of the pocket 11, the location of the recess H is excluded).

[0043] As shown in Fig. 4, the inner surface of the pocket 11 is an inner surface (spherical surface) that describes a circular arc along the circumferential direction of the bearing, and its radius in the circumferential direction of the bearing is set to Rz. That is, the radius of the inner surface E of the pocket 11 (hereinafter referred to as the circumferential inner surface E) in a cross section (hereinafter referred to as the circumferential cross section) that is perpendicular to the radial line of the bearing passing through the pocket center C at the pocket center C is Rz, and the diameter is 2Rz (hereinafter referred to as the circumferential pocket diameter 2Rz). In Fig. 4, the circumferential inner surface E continues from the tip G of the retaining claw 14 through the bottom F of the pocket 11 to the tip G of the retaining claw 14 on the opposite side. FIG. 4 shows a cross section that passes through the pocket center C and is perpendicular to a radial line of the bearing connecting the pocket center C and the axis O of the bearing, but this circumferential pocket diameter 2Rz is set not only on this cross section, but also on any cross section that passes through the pocket center C and intersects with a radial line of the bearing that passes through the pocket center C (however, if a recess H such as an oil reservoir is provided on the inner surface of the pocket 11, the location of the recess H is excluded).

[0044] In addition, in Figs. 3 and 4, the difference in diameter (radius) of each circular arc and the positional relationship of the centers of the circular arcs are exaggerated.

[0045] Here, the dimensional relationship between the diameter φx of the ball 5, the radial pocket diameter 2Ry on the inner surface of the pocket 11, and the circumferential pocket diameter 2Rz is as follows: φx<2Ry<2Rz It is set so that:

[0046] In this regard, if the radial pocket diameter 2Ry (see FIG. 3) and the circumferential pocket diameter 2Rz (see FIG. 4) of the pockets 11 are not larger than the diameter φx of the balls 5, the pockets 11 will tightly embrace the balls 5, and the cage 10 will not function. For this reason, the requirements φx<2Ry and φx<2Rz are required.

[0047] Furthermore, by setting the circumferential pocket diameter 2Rz (see FIG. 4) of the pocket 11 larger than the radial pocket diameter 2Ry (see FIG. 3), it becomes easier to avoid interference between the cage 10 and the surrounding members (the inner ring 3, the outer ring 4, etc.) that occurs when the balls 5 advance or lag during high-speed rotation, that is, interference caused by deformation of the cage 10 caused by the balls 5 being pulled by the cage 10 due to the difference in circumferential speed as they rotate in the circumferential direction of the bearing. This is the effect of ensuring that the second gap w2 (see FIG. 4) between the balls 5 and the circumferential inner surface E of the pocket 11 is larger than the first gap w1 (see FIG. 3) between the balls 5 and the radial inner surface D of the pocket 11 by setting 2Ry<2Rz.

[0048] In the cross section of Fig. 3, the radially inner circumferential surface D is an arc having a diameter of 2Ry, with reference numeral d1 indicating its inner diameter end d1 and reference numeral d2 indicating its outer diameter end d2. In the cross section of Fig. 4, the circumferentially inner circumferential surface E is an arc having a diameter of 2Rz, with reference numeral F indicating its bottom portion F. Reference numeral G indicates the tip G of the retaining claw 14.

[0049] On the other hand, if 2Ry of the pocket 11 is too large, the radial play between the balls 5 and the cage 10 becomes large, and interference between the cage 10 and the surrounding components (the inner ring 3, the outer ring 4, etc.) becomes more likely to occur. In addition, during high-speed rotation, the above-mentioned interference becomes more likely to occur due to the influence of centrifugal force deformation. For this reason, it is not preferable to make 2Ry too large. The inventors of the present application have confirmed that the requirement of 2Ry<2Rz is desirable. However, in order to suppress such interference, the cage 10 must be made of engineering plastic.

[0050] 4, the circumferential inner surface E includes a first inner surface E1 on the base 12 side and a second inner surface E2 on the retaining claw 14 side, i.e., closer to the tip of the column 13 than the first inner surface E1, and the center C2 of the arc of the second inner surface E2 is eccentric toward the base 12 than the center C1 of the arc of the first inner surface E1. This makes it possible to ensure a large second gap w2 between the balls 5 and the circumferential inner surface E of the pocket 11 while maintaining the function of retaining the balls 5.

[0051] In Fig. 4, a line connecting the center C1 of the arc of the first inner circumferential surface E1 and the center C2 of the arc of the second inner circumferential surface E2 in the circumferential direction of the bearing is parallel to the axial direction of the bearing. The pocket center C is preferably the midpoint between the center C1 of the arc of the first inner circumferential surface E1 and the center C2 of the arc of the second inner circumferential surface E2. The distance w between the center C1 of the arc of the first inner circumferential surface E1 and the center C2 of the arc of the second inner circumferential surface E2 is appropriately set according to the diameter φx of the balls 5 and the specifications of the bearing, and can be set to a value between 0.1 and 0.2 mm, for example.

[0052] A disadvantage of making the circumferential pocket diameter 2Rz too large is that the axial play of the cage 10 increases. If the axial play of the cage 10 becomes large, there is a concern that it may interfere with the seal member 20 depending on the conditions of use. For this reason, it is desirable to ensure an axial clearance between the seal member 20 and the cage 10 of 0.1 mm or more, taking into consideration the axial play and deformation of the cage 10 due to centrifugal force.

[0053] According to the above configuration, dmn={(D+d) / 2}×n D: Bearing outer diameter (mm) d: bearing inner diameter (mm) n: Rotation speed (min -1 ) This allows the bearing 1 to be applied under high-speed conditions where the dmn value, as defined by the formula (1), is 700,000 or more at the maximum rotational speed. At the same time, the intrusion of foreign matter of a size that may affect the bearing life can be suppressed. Furthermore, deformation of the cage 10 during high-speed operation can be suppressed, and foreign matter can be prevented from becoming caught between the cage 10 and the seal member 20.

[0054] Regarding the radius Rz of the circumferential inner surface E, it is preferable that the radius Rz1 of the first inner surface E1 and the radius Rz2 of the second inner surface E2 are set to the same value. However, as long as there is no problem in retaining the balls 5 and ensuring the second gap w2, the radius Rz1 of the first inner surface E1 and the radius Rz2 of the second inner surface E2 may be different values, for example, as shown in Fig. 5. In this case, the radius Rz1 of the first inner surface E1 may be set to a value larger than the radius Rz2 of the second inner surface E2, or the radius Rz1 of the first inner surface E1 may be set to a value smaller than the radius Rz2 of the second inner surface E2.

[0055] In FIG. 5, the recess H for oil storage formed on the inner surface of the pocket 11 is omitted. Here, the theoretical connection point of the first inner circumferential surface E1 and the second inner circumferential surface E2 is on a plane that passes through the pocket center C and is perpendicular to the axial direction of the bearing. This theoretical intersection point is indicated by the symbol J in FIG. 5. When the theoretical intersection point J is viewed macroscopically, if the first inner circumferential surface E1 and the second inner circumferential surface E2 are directly connected, a bend point will be interposed between them. For this reason, it is preferable that the theoretical intersection point J is smoothly attached so that the curvature is as continuous as possible so that there is no bend point or the like where the curvature is discontinuous near the theoretical intersection point J. In this case, for example, if a recess H is provided as an oil storage between the first inner circumferential surface E1 and the second inner circumferential surface E2 as shown in FIG. 4, the theoretical intersection point J is located within the recess H, and the problem of the bend point between the first inner circumferential surface E1 and the second inner circumferential surface E2 can be solved.

[0056] In the above embodiment, the lubricant (lubricating oil) is supplied to the bearing internal space A from both the axial one end side and the axial other end side, so the contents of this invention have been described using a double-sided seal configuration in which the seal member 20 is provided at the openings at both axial ends as an example. As another example, when the lubricant (lubricating oil) is supplied to the bearing internal space A in one direction from the axial one end side to the axial other end side, a single-sided seal configuration in which the seal member 20 is provided only at the opening at the axial one end side may be adopted. That is, in the case of a single-sided seal, it is preferable that the seal member 20 is set at the opening on the side where the lubricant (lubricating oil) is supplied. In addition, in the case of a single-sided seal, it is preferable that the cage 10 is inserted into the bearing internal space A from the opposite seal side where the seal member 20 is not provided.

[0057] As described above, the bearing 1 is used under high-speed conditions where the dmn value at the maximum rotation speed is 700,000 or more, and it is required to suppress deformation of the cage 10 due to high-speed rotation by applying a crown cage made of engineering plastic. However, when applying this resin crown cage, for example, if the amount of lubricant (lubricating oil) flowing into the bearing internal space A is large, there is a concern that the bearing life will be reduced. In addition, depending on the usage conditions of the bearing 1, there is a concern that foreign matter such as wear powder of gears may enter the bearing internal space A and reduce the bearing life. For this reason, in this invention, by adopting a seal member 20 that can maintain the contact portion of the seal lip 21 in a fluid lubricated state, it is possible to suppress the intrusion of foreign matter that affects the bearing life and to control the inflow amount of lubricant (lubricating oil). In other words, the synergistic effect of the resin crown cage having a predetermined shape and performance and the seal member 20 that can maintain the contact portion of the seal lip 21 in a fluid lubricated state has been achieved to extend the bearing life more than before.

[0058] Table 1 below shows the results of rapid acceleration / deceleration tests based on the amount of internal oil in a plastic crown cage.

[0059] [Table 1]

[0060] From the above results, the bearing life is longer when the amount of lubricant (lubricating oil) flowing into the bearing internal space A is smaller. This is thought to be because the stirring resistance is smaller when the amount of lubricant (lubricating oil) flowing in is smaller. As described above, the seal member 20 that can maintain the contact part of the seal lip 21 in a fluid lubricated state can prevent the intrusion of foreign matter of 0.050 mm or more, which affects the fluid lubrication and bearing life. In addition, although the seal member 20 is a contact seal, it is fluid lubricated during operation, so there is no need to ensure an inflow of lubricant (lubricating oil) more than necessary, and it is understood that there is no problem even if the lubricant (lubricating oil) is prevented from excessively entering the bearing internal space A, as in Tests No. 3 and No. 4 in Table 1 above.

[0061] In the above embodiment, the cage 10 is made of engineering plastic, but the material of the cage 10 may contain at least engineering plastic.

[0062] In the above embodiment, the rotating shaft is exemplified as the rotating shaft of a drive motor equipped in an electric transport device such as an electric vehicle, or the rotating shaft of a reducer or speed increaser equipped in such electric transport device, but the bearing 1 of the present invention and a bearing device using the bearing 1 can also be applied to supporting parts of rotating shafts in various other transport devices, industrial machines, etc. For example, the bearing 1 can be applied to rotating parts of shafts, constant velocity joints, propeller shafts, turbochargers, transmissions, and wheel bearings in power transmission paths in various transport devices, or supporting parts of rotating shafts in various machine tools, generators, etc.

[0063] In the above embodiment, the configuration of the present invention has been described using an example of a sealed ball bearing 1 in which the seal lip 21 of the seal member 20 is arranged on the inner ring 3 side. However, it is also possible to reverse this and adopt a configuration in which the seal lip 21 is arranged on the outer ring 4 side, for example, as shown in Figure 8.

[0064] In the above embodiment, the elastic portion 24 of the seal member 20 is rubber, and the rubber is vulcanization bonded to the core metal 23, but the rubber may be integrated with the core metal 23 by a bonding method other than vulcanization bonding. As long as the strength and durability of the seal member 20 are ensured, the seal member 20 may be composed only of the elastic portion 24 without using the core metal 23. Furthermore, the material of the elastic portion 24 may be, other than rubber, for example, a synthetic resin.

[0065] The embodiments disclosed herein should be considered to be illustrative and not restrictive in all respects. The scope of the present invention is defined by the claims, not the above description, and is intended to include all modifications within the meaning and scope of the claims. [Explanation of symbols]

[0066] 1 Bearing (sealed ball bearing) 3. Inner Circle 4 Outer ring 5 balls (rolling elements) 10 Cage 11 Pocket 12 Base 13 Pillar section 20 Sealing material 21 Seal lip 27 Protrusion 28 Gap A Bearing internal space B Seal sliding surface C Pocket Center D Radial inner surface E Circumferential inner surface E1 1st inner surface E2 2nd inner surface

Claims

1. The bearing comprises an inner ring (3) and an outer ring (4), balls (5) arranged between the inner ring (3) and the outer ring (4), a cage (10) having pockets (11) for holding the balls (5) along a circumferential direction, and a seal member (20) for closing an opening at an axial end of a bearing internal space (A) formed between the inner ring (3) and the outer ring (4), dmn={(D+d) / 2}×n D: Bearing outer diameter (mm) d: bearing inner diameter (mm) n: Rotation speed (min －１ ) In a ball bearing used in an environment where the dmn value specified in is 700,000 or more, the seal member (20) is provided with a seal lip (21) that is fixed to one of the inner ring (3) and the outer ring (4) and that is in sliding contact with a seal sliding surface (B) set on the other of the inner ring (3) and the outer ring (4), the seal lip (21) has a plurality of protrusions (27) arranged in a circumferential direction, the plurality of protrusions (27) are formed in such a manner that gaps (28) are generated between the protrusions (27) adjacent to each other in the circumferential direction, and a fluid lubrication state can be established between the seal lip (21) and the seal sliding surface (B) by an oil film of lubricating oil that is dragged between the protrusions (27) and the seal sliding surface (B) from the gaps (28) as the bearing rotates, The cage (10) is a crown cage including an engineering plastic and includes an annular base (12) and a plurality of pillar portions (13) protruding in one direction from the base (12), wherein a dimensional relationship between a diameter φx of the balls (5), a radius Ry of a radial inner peripheral surface (D) of the pocket (11) in a cross section passing through a pocket center (C) of the pocket (11) and including an axis (O) of the bearing, and a radius Rz of a circumferential inner peripheral surface (E) of the pocket (11) in a cross section perpendicular to a radial line passing through the pocket center (C) at the pocket center (C) is φx<2Ry<2Rz, The circumferential inner surface (E) is a first inner surface (E １ ) and the first inner circumferential surface (E １ The second inner peripheral surface (E ２), and the second inner circumferential surface (E ２ ) center (C ２ ) is the first inner circumferential surface (E １ ) center (C １ ) toward the base portion (12).

2. The radius Rz of the circumferential inner peripheral surface (E) is １ ) radius Rz １ and the second inner circumferential surface (E ２ ) radius Rz ２ 2. The sealed ball bearing according to claim 1, wherein said and said bearing members are set equal to each other.

3. The first inner circumferential surface (E １ ) and the second inner circumferential surface (E ２ 2. A sealed ball bearing as claimed in claim 1, wherein a recess (H) is provided between the bearing and the outer circumferential surface of the bearing as an oil reservoir.

4. 2. The sealed ball bearing according to claim 1, wherein the seal lip (21) is made of one or more materials selected from the group consisting of nitrile rubber, acrylic rubber, and fluororubber.

5. 2. A sealed ball bearing as described in claim 1, wherein when the supply of lubricant to the internal space of the bearing is in one direction from one axial end side to the other axial end side, the seal member (20) is provided only at the opening on the one axial end side.

6. 6. A bearing device using the sealed ball bearing according to claim 1, in which a rotating shaft of a drive motor, a reducer or a speed increaser for an electric transport device is supported by the sealed ball bearing.

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