Bearing with seal

Abstract

To supply a lubricant to an internal space of a ball bearing in a large size and prevent invasion of a foreign substance, which may exert adverse influences upon a lifetime of the ball bearing, in a compatible manner in a bearing with a seal for protecting the ball bearing by means of a seal member capable of separating a seal lip and a seal slide surface by forming an oil film between the seal lip and the seal slide surface.SOLUTION: A plurality of projections 23 of a seal lip 16 of a seal member 2 generates an oil passage 24, which communicates an internal space 7 of a ball bearing 1 and the outside, between the seal lip 16 and a seal slide surface 18. An oil film is formed between the seal lip 16 and the seal slide surface 18 by an oil film of a lubricant which is slid from the oil passage 24 into between the projections 23 and the seal slide surface 18 with rotation of the ball bearing 1, thereby separating the seal lip and the seal slide surface. An inner diameter d of the ball bearing 1, a height h of the projections 23, a coefficient A, and a constant B satisfy 0.08≤h≤A×d+B, 0.002≤A≤0.004, and -0.043≤B≤-0.002.SELECTED DRAWING: Figure 1

Description

[Technical Field]

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

[0002] Conventionally, seals have been used to prevent early failure of ball bearings. For example, the transmissions installed in vehicles such as automobiles and various construction machines contain foreign matter such as gear wear debris, and seals are used to prevent this debris from entering the bearing. Seals can be broadly divided into contact seals and non-contact seals.

[0003] A typical contact seal has a seal lip made of an elastic material. This type of seal is typically attached to one of the inner and outer raceways of a ball bearing. The mating components, such as the raceway and slinger, rotate circumferentially relative to the seal as the ball bearing rotates. A seal sliding surface is formed on the mating components. The seal lip and the seal sliding surface are in sliding contact around the entire circumference, creating a microscopic solid contact area. The drag resistance (seal torque) of the seal lip due to this sliding contact increases bearing torque and contributes to an increase in the temperature of the ball bearing. Furthermore, because the entire interior space of the ball bearing is sealed off from the outside by the seal, air in the bearing expands thermally as the bearing temperature increases during operation and is expelled to the outside of the bearing. However, when the bearing temperature drops and the air in the bearing contracts during operation, the pressure difference between the inside and outside of the ball bearing causes the seal lip to be pressed against the seal sliding surface, resulting in an adsorption effect and an increase in seal torque. For these reasons, general contact seals have limitations on the high speed rotation of ball bearings.

[0004] In contrast, the sealed bearing disclosed in Patent Document 1 employs a seal lip including multiple protrusions arranged circumferentially. These protrusions are formed in a manner that allows an oil film to form between the seal lip and the seal sliding surface. That is, the shape and spacing of these protrusions are designed to create an oil passage between the seal lip and the seal sliding surface that connects the internal space of the ball bearing to the outside. The flow of lubricating oil between the internal space of the ball bearing and the outside through this oil passage ensures ample lubrication on the seal sliding surface. As the ball bearing rotates, the lubricating oil is drawn from the oil passage into the gap between the protrusions and the seal sliding surface, creating a wedge effect that forms an oil film between each protrusion and the seal sliding surface, separating them. Therefore, the sealed bearing disclosed in Patent Document 1 prevents the aforementioned suction effect and significantly reduces seal torque by eliminating solid contact between the seal lip and the seal sliding surface at speeds above a certain level.

[0005] Furthermore, in the sealed bearing disclosed in Patent Document 1, noting that the intrusion of foreign matter with a particle size of 0.05 mm or less does not adversely affect the life of the ball bearing, the height of the protrusion of the seal lip is set to 0.07 mm or less to prevent foreign matter with a particle size exceeding 0.05 mm from entering the interior of the ball bearing through the aforementioned oil passage. As a result, the sealed bearing disclosed in Patent Document 1 can achieve a ball bearing life that is comparable to that of the aforementioned general contact seal, regardless of the size of the contact ellipse or indentation that occurs at the contact area between the ball and raceway. [Prior art documents] [Patent documents]

[0006] [Patent Document 1] Japanese Patent Application Laid-Open No. 2017-161069 Summary of the Invention [Problem to be solved by the invention]

[0007] However, demand for higher rotational speeds for ball bearings continues to grow. Ball bearings with large inner diameters are sometimes used for applications such as supporting the rotating parts of drive motors and reducers for commercial vehicles. The sealed bearing disclosed in Patent Document 1 has a protrusion height of 0.07 mm or less, regardless of bearing size. However, in the case of ball bearings with larger bearing sizes, the protrusion height is small (0.07 mm or less) relative to the bearing size, resulting in a narrow opening for the oil passage. This raises concerns that in the future, when large-sized ball bearings rotate at high speeds, the supply of lubricating oil to the internal space of the ball bearing will be insufficient.

[0008] On the other hand, if the protrusions are made higher than 0.07 mm to avoid the aforementioned lack of lubrication and the need for high-precision dies, it becomes possible for large particles to pass through the aforementioned gap and enter the inside of the ball bearing. This causes large-diameter indentations where the raceway and ball contact, which can lead to early flaking.

[0009] In view of the above background, the problem that this invention aims to solve is to provide a sealed bearing that protects a ball bearing with a seal member that can separate the seal lip and the seal sliding surface by an oil film formed between the seal lip and the seal sliding surface, and that can simultaneously supply sufficient lubricating oil to the internal space of a large ball bearing and prevent the intrusion of foreign matter that would adversely affect the life of the ball bearing. [Means for solving the problem]

[0010] In order to achieve the above object, the present invention provides a ball bearing including an inner raceway ring, an outer raceway ring, a plurality of balls arranged between the raceways rings, and a cage that holds the balls; seal members that are located on one axial side and the other axial side of the balls and the cage and are arranged between the inner raceway ring and the outer raceway ring; and a seal sliding surface that is formed on the inner raceway ring and slides circumferentially against the seal member, wherein the seal member is a seal member that includes a plurality of protrusions arranged in the circumferential direction. The sealed bearing has a seal lip, and the plurality of protrusions form an oil passage between the seal lip and the seal sliding surface, communicating between the internal space of the ball bearing and the outside. The sealed bearing is characterized in that, where d is the inner diameter (mm) of the ball bearing, h is the height (mm) of the protrusions, A is a coefficient for d, and B is a constant, the following relationships are satisfied: 0.08≦h≦A×d+B, 0.002≦A≦0.004, and −0.043≦B≦−0.002. The aforementioned internal space refers to the space enclosed by the inner and outer raceways, between both axial end faces of the inner and outer raceways. The outside refers to any space other than the internal space.

[0011] There is a relationship between the particle size of foreign matter that can enter the internal space of the ball bearing through the oil passage and the diameter of the indentation that occurs at the contact point between the ball and the raceway. Furthermore, even if the indentation diameter is the same, a larger contact ellipse diameter increases the ball bearing's lifespan. Based on these relationships, by satisfying the conditions 0.08≦h≦A×d+B, 0.002≦A≦0.004, and −0.043≦B≦−0.002, as in Configuration 1 above, the relationship between the maximum particle size of foreign matter that can enter the internal space of the ball bearing through the oil passage and the contact ellipse diameter can be limited to a range that does not adversely affect the ball bearing's lifespan. Furthermore, by setting the protrusion height h to 0.08 mm or more, the opening of the oil passage between the seal lip and the seal sliding surface can be widened, improving the supply of lubricating oil to the internal space, even in ball bearings with a large bearing size (ball bearing bore diameter d).

[0012] Specifically, in the above configuration 1, when a load that is 15% of the basic dynamic load rating of the ball bearing is applied to the ball bearing, the diameter (mm) of the contact ellipse between the raceway and the balls is defined as X, and the height (mm) of the protrusion is defined as h, the relationship 0.08≦h≦{(0.044X−0.15) / 2}+0.02 may be adopted.

[0013] In high-speed applications such as vehicle traction motors and reducers, the load imposed on ball bearings during steady operation is approximately 15% of their dynamic load rating. Based on this load and the aforementioned relationship, the maximum particle size (mm) that can be tolerated without adversely affecting the bearing's life was calculated. The maximum value was found to be {(0.044X - 0.15) / 2}. It has also been found that the maximum particle size that can penetrate the bearing's internal space through the oil passage is approximately the protrusion height h - 0.02 mm. In other words, as in Configuration 2, satisfying 0.08 ≤ h ≤ {(0.044X - 0.15) / 2} + 0.02 ensures a better supply of lubricating oil to the bearing's internal space compared to a protrusion height h ≤ 0.07 mm when a large-sized ball bearing rotates at high speed under typical load conditions, while achieving a ball bearing life comparable to that of a protrusion height h ≤ 0.07 mm.

[0014] In the above configuration 1 or 2, configuration 3 can be adopted in which the inner diameter d of the ball bearing is 40 mm or more.

[0015] In any one of the above configurations 1 to 3, configuration 4 can be adopted in which the hardness of the elastic material is 60 HS or more and 80 HS or less.

[0016] In the above configuration 4, configuration 5 can be adopted in which either a thermosetting elastomer or a thermoplastic elastomer is used as the elastic material.

[0017] In any one of the above configurations 1 to 5, a configuration 6 can be adopted in which the rotating part of either one of the vehicle's traction motor and reduction gear is supported. [Effects of the Invention]

[0018] As described above, by adopting the above-mentioned configuration 1, the present invention is able to simultaneously supply lubricating oil to the internal space of a large-sized ball bearing and prevent the intrusion of foreign matter that would adversely affect the life of the ball bearing, in a sealed bearing that protects a ball bearing with a seal member that can separate the seal lip and the seal sliding surface by an oil film formed between the seal lip and the seal sliding surface. [Brief explanation of the drawings]

[0019] [Figure 1] 1 is a cross-sectional view showing a sealed bearing according to an embodiment of the present invention; [Figure 2] 2 is a cross-sectional view showing the natural state of the sealing member of FIG. 1; [Figure 3] 2 is a side view showing the lip head and its vicinity of the seal member of FIG. 1 from the inside of the bearing. [Figure 4] Enlarged view of the seal lip area in Figure 1 [Figure 5] Cross section of line VV in Figure 4 [Figure 6] Graph showing the relationship between the contact ellipse diameter of the ball and raceway and the life ratio of a ball bearing according to Lorosch HK [Figure 7] A graph showing the relationship when the life ratio of the contact ellipse diameter of 5.6 mm and the indentation diameter of 0.1 mm in Figure 6 is changed to 1 [Figure 8] Graph showing an example of the relationship between the height of the protrusion of the seal lip and the inner diameter of the ball bearing, which does not adversely affect the life of the ball bearing. DETAILED DESCRIPTION OF THE INVENTION

[0020] A sealed bearing according to an embodiment of the present invention will be described with reference to the accompanying drawings, FIGS. 1 to 6. FIG.

[0021] The sealed bearing shown in FIG. 1 comprises a ball bearing 1 and two seal members 2 arranged on both sides of the ball bearing 1.

[0022] Ball bearing 1 is composed of an inner raceway 3, an outer raceway 4, a number of balls 5 arranged between these raceways 3, 4, and a cage 6 that holds these balls 5. Ball bearing 1 is a deep groove ball bearing. The inner raceway 3 is an annular bearing component with a raceway groove on its outer periphery. The outer raceway 4 is an annular bearing component with a raceway groove on its inner periphery. Balls 5 are spheres that roll in the raceway grooves of the inner and outer raceways 3, 4. These raceways 3, 4 and balls 5 are each made of steel. The cage 6 is an annular bearing component that distributes the balls 5 evenly circumferentially.

[0023] 1, the direction along the central axis of ball bearing 1 (not shown, same below) is defined as the axial direction, the direction perpendicular to the central axis is defined as the radial direction, and the direction along the circumference of a circle centered on the central axis is defined as the circumferential direction. The axial direction corresponds to the left-right direction in FIG. 1, and the radial direction corresponds to the up-down direction in FIG. 1.

[0024] The inner diameter d of the ball bearing 1 is the same as the inner diameter of the inner raceway 3. The inner raceway 3 is attached to a shaft S. The outer raceway 4 is attached to a member that surrounds the ball bearing 1 radially, such as a housing or a gear.

[0025] The seal member 2 is disposed between an internal space 7 of the ball bearing 1 and the outside. The internal space 7 is formed around the entire circumference between the outer periphery of the inner raceway 3 and the inner periphery of the outer raceway 4. A seal groove 8 that holds the seal member 2 is formed at the end of the inner periphery of the outer raceway 4.

[0026] The internal space 7 is lubricated by lubricating oil (not shown; the same applies below) supplied from the outside. Examples of lubrication methods include a splash method in which lubricating oil is poured onto the sealed bearing, or an oil bath method in which the lower part of the sealed bearing is immersed in an oil bath. An appropriate amount of grease may be sealed in the internal space 7 as an initial lubricant.

[0027] The shaft S is provided as a rotating part of either the vehicle's traction motor or reduction gear. Here, the vehicle's traction motor refers to an electric motor provided as a power source in an electric vehicle (EV), hybrid electric vehicle (HEV), etc. The vehicle's traction motor and reduction gear may also constitute an eAxle. The eAxle refers to a unit that integrates the traction motor, inverter, and reduction gear (transaxle).

[0028] In large vehicles such as commercial vehicles and vehicles that require durability, the shaft diameter of the motor shaft of the driving motor, the first stage rotating shaft of the reducer, etc. may be 40 mm or more. The inner diameter d of the ball bearing 1 that supports the rotating part of such a shaft S is 40 mm or more.

[0029] Depending on the installation location of the sealed bearing, foreign matter such as gear wear powder, clutch wear powder, and small crushed stones may exist outside the ball bearing 1. Such powdery foreign matter can reach the vicinity of the seal member 2 due to the flow of lubricating oil or atmosphere. The seal member 2 seals the internal space 7 from the outside. The purpose of this seal is to prevent external foreign matter from entering the internal space 7 and thereby prevent early damage to the ball bearing 1, but it is not to seal the internal space 7 liquid-tight.

[0030] 2 shows a half cross section of the seal member 2 in its natural state where no external force other than gravity acts on it. The seal member 2 is made up of a core metal 9 and an elastic material 10.

[0031] 2, the core metal 9 is composed of a cylindrical plate portion 11 extending in the circumferential and axial directions, an annular plate portion 12 extending radially from the cylindrical plate portion 11, and a conical plate portion 13 extending in a direction inclined relative to the radial direction from the annular plate portion 12. The cylindrical plate portion 11 to the conical plate portion 13 are formed from metal plates such as steel plates.

[0032] The outer periphery of the cylindrical plate portion 11 defines the outer diameter of the core 9. The inner peripheral edge 14, which defines the inner diameter of the core 9, is formed from the tip of the conical plate portion 13 and is formed as a flat surface along the circumferential and axial directions around the entire circumference.

[0033] The seal member 2 has a fitting portion 15 bonded to the cylindrical plate portion 11 of the core metal 9, a seal lip 16 located radially inward of the inner peripheral edge 14 of the core metal 9, and a side surface portion 17 connecting the fitting portion 15 and the seal lip 16. The fitting portion 15 to the side surface portion 17 are formed from an elastic material 10.

[0034] The hardness of the elastic material 10 is 60 HS or more and 80 HS or less. This Shore hardness (HS) is a value measured by a Shore hardness test method in accordance with JIS K 6301 "Physical testing methods for vulcanized rubber."

[0035] Either a thermosetting elastomer or a thermoplastic elastomer can be used as the elastic material 10. Examples of thermosetting elastomers include nitrile rubber (NBR), acrylic rubber (ACM), and fluororubber (FKM). Examples of thermoplastic elastomers include polystyrene-based (TPS), olefin / alkene-based (TPO), and polyamide-based (TPAE).

[0036] As shown in Fig. 1, a seal sliding surface 18 that slides in the circumferential direction against the seal lip 16 is formed on the outer periphery of the inner bearing ring 3. The seal sliding surface 18 is a cylindrical surface that extends along the circumferential direction all around.

[0037] The outer peripheral edge of the fitting portion 15 of the seal member 2 shown in Figure 2 defines the outer diameter of the seal member 2. The seal member 2 is attached to the outer raceway 4 by press-fitting the fitting portion 15 into the seal groove 8 shown in Figure 1.

[0038] As shown in Figure 2, the seal lip 16 consists of a waist 19 extending radially, a lip head 20 extending radially inward from the waist 19 in a direction inclined axially outward, and a heel 21 continuing between the waist 19 and the inner peripheral edge 14 of the core metal 9.

[0039] The waist 19 is formed in a circular ring shape along the circumferential and radial directions around the entire circumference. The heel 21 is bonded to the inner peripheral edge 14 of the core 9, and becomes thicker as it approaches the inner peripheral edge 14 of the core 9, and is continuous with the side portion 17.

[0040] Here, Fig. 3 shows the side of the seal lip 16 in Fig. 1 facing the internal space 7. Fig. 4 shows an enlarged view of the vicinity of the sliding contact portion between the member 2 in Fig. 1 and the seal sliding surface 18. As shown in Figs. 3 and 4, the lip head 20 is made up of a solid portion 22 formed around the entire circumferential direction, and a plurality of protrusions 23 protruding from the solid portion 22 toward the seal sliding surface 18.

[0041] The multiple protrusions 23 are aligned in the circumferential direction. The multiple protrusions 23 are spaced at a constant circumferential pitch. The protrusions 23 extend in a direction perpendicular to the circumferential direction over their entire length. The overall shape of the seal lip 16 is rotationally symmetrical corresponding to the pitch of the protrusions 23.

[0042] A radial interference is set between the seal lip 16 and the seal sliding surface 18. When the seal member 2 is attached to the outer raceway 4 as shown in Figures 1 and 4, the seal lip 16 is pressed against the seal sliding surface 18 by the multiple protrusions 23, causing the lip head 20 to bend from the waist 19 as shown in Figure 4. Installation errors, manufacturing errors, etc. of the seal member 2 are absorbed by changes in the degree of bending of the seal lip 16.

[0043] FIG. 5 shows a cross section taken along line VV in FIG. 4. This cross section is a cross section on an imaginary plane perpendicular to the axial direction. As shown in FIG. 5, the cross section of the protrusions 23 is semicircular with a curvature radius r. Note that in FIG. 5, the cross section of the protrusions 23 shows the shape in the natural state shown in FIG. 2. In reality, when the seal member 2 is installed, the multiple protrusions 23 contact the seal sliding surface 18. At this time, due to the action of the elastic restoring force of the deflection deformation of the entire periphery of the seal lip 16, the protrusions 23 are pushed toward the seal sliding surface 18, and the tops of the protrusions 23 are slightly compressed. The solid portions 22 are slightly deflected toward the seal sliding surface 18 but are not in contact with the seal sliding surface 18. Oil passages 24 communicating with the internal space 7 and the outside are formed between adjacent protrusions 23 in the circumferential direction and between the seal sliding surface 18 and the solid portions 22. When the ball bearing 1 rotates, the seal lip 16 slides against the seal sliding surface 18 only on the plurality of protrusions 23 .

[0044] The protrusions 23 form a wedge-shaped gap between them and the seal sliding surface 18, which is larger on the oil passage 24 side and smaller on the protrusion 23 side. The elastic contact area between the protrusions 23 and the seal sliding surface 18 is formed over a finite length L in the axial direction (however, L is exaggerated) due to the elastic deformation of the protrusions 23, as shown in FIG. 4. When the ball bearing 1 rotates, the seal lip 16 and the seal sliding surface 18 rotate relative to each other in the circumferential direction. In the elastic contact area between the protrusions 23 and the seal sliding surface 18, the protrusions 23 drag the lubricating oil in the oil passage 24 (the flow of lubricating oil when the seal sliding surface 18 rotates clockwise in FIG. 5 is schematically illustrated by arrows) into the circumferential direction between the protrusions 23 and the seal sliding surface 18, creating a wedge effect that promotes oil film formation, and the oil film thickness between the protrusions 23 and the seal sliding surface 18 increases.

[0045] When the ball bearing 1 is stopped or its rotational speed is below a certain level, microscopically, the elastic contact area between the protrusions 23 and the seal sliding surface 18 includes a solid contact area, and the lubrication mode is boundary lubrication or mixed lubrication. When the peripheral speed of the relative rotation between the protrusions 23 and the seal sliding surface 18 exceeds a certain level, the oil film thickness between the protrusions 23 and the seal sliding surface 18 easily exceeds the composite roughness σ between the protrusions 23 and the seal sliding surface 18, and each protrusion 23 and the seal sliding surface 18 are separated by an oil film. This allows the seal lip 16 and the seal sliding surface 18 to be separated by an oil film. In this state, the seal torque of the seal member 2 is reduced to the same level as a non-contact seal, which in turn suppresses temperature rise in the sealed bearing and prevents suction of the seal lip 16.

[0046] The cross-sectional shape and pitch of the protrusions 23 may be appropriately set so that the seal lip 16 and the seal sliding surface 18 are separated by an oil film formed between them under predetermined conditions of peripheral speed and oil temperature.

[0047] The amount of lubricating oil circulating between the internal space 7 and the outside through the oil passage 24 depends on the height h of the protrusion 23 from the solid portion 22. Here, the height h of the protrusion 23 is the height between the seal lip 16 and the seal sliding surface 18 in the direction normal to the solid portion 22. The height h of the protrusion 23 is determined based on the relationships shown in FIGS. 6 to 8.

[0048] Figure 6 here is a transcription of a graph published in the following source: Source: Lorosch HK.”Research on Longer Life for Rolling-Element Bearings.” ASLE Lubrication Engineering, Vol. 41 (1985), pp. 37

[0049] Figure 6 shows the relationship between the contact ellipse diameter and the life ratio for indentation diameters of 0.1 mm and 0.3 mm, with the contact ellipse diameter between the ball and raceway on the vertical axis and the life ratio of the ball bearing on the horizontal axis. Figure 6 shows that even if the indentation diameter is the same, there is a relationship in which the life ratio increases as the contact ellipse diameter increases.

[0050] As shown in Patent Document 1, it is believed that the lifespan will not be reduced if the particle size of the foreign matter that penetrates into the inside of the bearing is 0.05 mm or less. For example, the lifespan of a ball bearing was not reduced in the following lifespan test. Life test conditions Ball bearing model number: 6206 Basic load rating of ball bearing: 21.6kN Load applied to ball bearing: 6.9kN Contact ellipse diameter between ball and inner ring: 5.6 mm The above model numbers and basic load ratings are values ​​that comply with the latest corresponding JIS standards at the time of filing this application.

[0051] Generally, it is known from experience that the diameter of the indentation that occurs at the contact point between the ball and raceway of a ball bearing is about twice the particle size of the foreign matter that is caught between the ball and raceway. If the particle size of the foreign matter is 0.05 mm, the indentation diameter will be 0.1 mm.

[0052] On the horizontal axis (life ratio) of Fig. 6, if the life ratio is changed to 1 when the indentation diameter is 0.1 mm and the contact ellipse diameter is 5.6 mm using the life test results described above, the relationship shown in Fig. 7 is obtained. In Fig. 7, when the indentation diameter is 0.3 mm, the contact ellipse diameter at which the life ratio is 1 is 10.2 mm. From Fig. 7, the relationship in the following equation 1 holds between the two parameters, contact ellipse diameter X and indentation diameter Y, at which the life ratio is 1. Y=0.044X-0.15...Equation 1 Here, the indentation diameter Y is the diameter (mm) of the indentation formed at the contact point between the ball and the raceway groove of the inner ring, and the contact ellipse diameter X is the diameter (mm) of the contact ellipse at the contact point between the ball and the raceway groove of the inner ring, and this contact ellipse diameter is equivalent to twice the length of the semimajor axis of the contact ellipse calculated in the Hertz contact theory.

[0053] Based on the empirical rule for the indentation diameter and particle size mentioned above, the relationship between the upper limit of particle size (mm) of foreign matter that is thought to enter the internal space of a ball bearing without adversely affecting the service life of the ball bearing (hereinafter referred to as the allowable intrusion foreign matter diameter Z) and the contact ellipse diameter X is expressed by the following equation 2. 2Z=0.044X-0.15 Z=(0.044X-0.15) / 2...Equation 2

[0054] In high-speed rotation applications such as vehicle traction motors and reducers, the load imposed on a ball bearing during steady operation is approximately 15% of the ball bearing's dynamic load rating, taking into account the bearing's lifespan. Here, the dynamic load rating of a ball bearing refers to that specified in the Japanese Industrial Standards (JIS B 1518, "Rolling Bearings - Dynamic Load Rating and Rated Life"). For example, in the case of a deep groove ball bearing used as a radial bearing, this is the basic dynamic radial load rating Cr. If the contact ellipse diameter X is calculated based on an applied load of 15% of the typical dynamic load rating for ball bearings and the dimensional relationships of the ball diameter, raceway groove diameter, and raceway ring thickness, and then rearranged in terms of the ball bearing bore diameter d and the allowable foreign object diameter Z, the following relationships between the allowable foreign object diameter Z and the ball bearing bore diameter d are expressed by Equations 3 to 5: Z=A×d+B0...Equation 3 0.002≦A≦0.004...Equation 4 -0.063≦B0≦-0.022...Equation 5

[0055] In Patent Document 1, the height h of the protrusion is limited to 0.07 mm or less to prevent foreign matter with a particle size exceeding 0.05 mm from entering the internal space of the ball bearing through the oil passages. As can be seen from this, the allowable height h of the protrusion is the allowable diameter Z of the foreign matter entering the bearing + 0.02 mm. Therefore, the relationship between the two parameters, the allowable height h of the protrusion that is allowed to limit the height h to the allowable diameter Z of the foreign matter entering the bearing, and the inner diameter d of the ball bearing, is expressed by the following equations 6 to 8. Note that the height h of the protrusion is a dimensional value at room temperature (5°C to 35°C) as specified in JIS Z 8703. h≦A×d+B...Equation 6 0.002≦A≦0.004...Equation 7 -0.043≦B≦-0.002...Equation 8

[0056] For ball bearings that comply with the Japanese Industrial Standards (JIS B 1512, "Boundary Dimensions of Rolling Bearings"), their designation numbers are specified in the Japanese Industrial Standards (JIS B 1513, "Designation Numbers of Rolling Bearings"). If the designation number of a ball bearing corresponds to the 62 series, then in Equation 6, coefficient A for the ball bearing's bore diameter d is 0.004 and constant B is -0.043. If the designation number of a ball bearing corresponds to the 63 series, then coefficient A is 0.002 and constant B is -0.002. Figure 8 shows the relationship between the bore diameter d and the protrusion height h of these 63 and 62 series ball bearings, and Table 1 below shows examples of the bore diameter d, contact ellipse diameter X, indentation diameter, allowable foreign object diameter Z, and protrusion height h of ball bearings for various model numbers in the same series.

[0057] [Table 1]

[0058] Based on the relationships shown in Figures 6 to 8 and Equations 1 to 8, the ball bearing 1 and seal member 2 shown in Figures 1 and 5 are configured to satisfy the conditions 0.08≦h≦A×d+B, 0.002≦A≦0.004, and −0.043≦B≦−0.002 in order to increase the supply of lubricating oil to the internal space 7 while preventing foreign matter that could adversely affect the life of the ball bearing 1 from entering the internal space 7 through the oil passage 24. This configuration can be achieved by setting the height h of the protrusion 23 and the contact ellipse diameter X. The contact ellipse diameter X is determined by the magnitude of the applied load relative to the dynamic load rating of the ball bearing 1, the geometric relationship between the contact areas between the raceways 3 and 4 and the balls 5, and the Young's modulus and Poisson's ratio of the materials of the raceways 3 and 4 and the balls 5. The geometric relationship of the contact area between the raceways 3, 4 and the balls 5 is determined by the diameter of the balls 5, the groove radius of the raceway grooves of the raceways 3, 4 on the axial plane including the bearing center axis, and the radius of curvature of the raceway grooves of the raceways 3, 4 on the radial plane perpendicular to the bearing center axis.

[0059] In other words, when a load that is 15% of the basic dynamic load rating of ball bearing 1 is applied to ball bearing 1, the diameter X of each contact ellipse between each raceway 3, 4 and ball 5, and the height h of protrusion 23 are configured to satisfy the following formula 9. h≦{(0.044X-0.15) / 2}+0.02...Equation 9

[0060] When a large-sized ball bearing 1 is operated at a rotational speed of up to about 1.3 million dmn (inner diameter d × revolutions per minute n), it is sufficient to achieve the lubrication objective of preventing metal-to-metal contact between the balls 5 and the raceways 3 and 4 with an oil film. To achieve this, a height h of 0.07 mm or less of protrusions 23 can adequately supply lubricant to the internal space 7. However, when the bearing is operated at a rotational speed of 1.8 million dmn or more, heat generation inside the ball bearing 1 due to shearing of the lubricant between the balls 5 and the cage 6 becomes significant. Therefore, it may be necessary to supply even more lubricant to the internal space 7 to adequately cool the interior of the large-sized ball bearing 1 with a correspondingly large capacity. To meet this requirement, the upper limit of the height h of protrusions 23 can be optimized to a value of 0.08 mm or more based on the relationship between Equations 1 to 9, thereby improving the supply of lubricant to the internal space 7.

[0061] Furthermore, when the height h of the protrusions 23 is set to 0.08 mm or more, the circumferential width of the protrusions 23 is correspondingly increased, and as a result, the number of protrusions 23 formed on the seal lip 16 can be reduced compared to when the height h of the protrusions 23 is set to 0.07 mm or less. This is advantageous in that it reduces the precision of the mold for molding the seal lip 16, which is necessary to reduce dimensional variations between the protrusions 23.

[0062] The sealed bearing according to the embodiment (see Figures 1, 4 and 5) is as described above, and comprises a ball bearing 1 having an inner raceway 3, an outer raceway 4, a plurality of balls 5 arranged between these raceways 3, 4, and a cage 6 that holds these balls 5, a seal member 2 arranged between an internal space 7 of the ball bearing 1 and the outside, and a seal sliding surface 18 that slides circumferentially against the seal member 2, and the seal member 2 has a seal lip 16 that includes a plurality of protrusions 23 arranged in the circumferential direction, and the plurality of protrusions 23 slide against the internal space 7 and the outside. An oil passage 24 that communicates with the outside is created between the seal lip 16 and the seal sliding surface 18, and the seal lip 16 and the seal sliding surface 18 are formed in such a way that they can be separated by an oil film formed between them by the lubricating oil that is drawn from the oil passage 24 into between the protrusion 23 and the seal sliding surface 18 as the ball bearing 1 rotates, thereby preventing the intrusion of foreign matter of a predetermined particle size, enabling the ball bearing to operate at high speeds, and reducing the torque of the sealed bearing.

[0063] In particular, the sealed bearing of the embodiment satisfies 0.08≦h≦A×d+B, 0.002≦A≦0.004, and −0.043≦B≦−0.002, where d is the inner diameter (mm) of the ball bearing 1, h is the height (mm) of the protrusion 23, A is the coefficient for d, and B is a constant, thereby enabling the supply of lubricating oil to the internal space 7 of the large-sized ball bearing 1 and the prevention of the intrusion of foreign matter that would adversely affect the life of the ball bearing 1.

[0064] Furthermore, in the sealed bearing of the embodiment, when a load that is 15% of the basic dynamic load rating of ball bearing 1 is applied, the diameter (mm) of the contact ellipse between raceways 3, 4 and balls 5 is X, and the height (mm) of protrusion 23 is h, satisfies 0.08≦h≦{(0.044X−0.15) / 2}+0.02. By doing so, when a large-sized ball bearing 1 rotates at high speed under typical load conditions, the supply of lubricating oil to internal space 7 of ball bearing 1 is improved compared to when height of protrusion 23 is h≦0.07, and the life of ball bearing 1 can be obtained that is comparable to when height of protrusion 23 is h≦0.07.

[0065] In this embodiment, an example has been shown in which the seal lip 16 is a radial lip, but the present invention also allows for the seal lip 16 to be changed to an axial lip. Here, the radial lip refers to a seal lip that provides a sealing effect with a seal sliding surface along the axial direction or a seal sliding surface that has a gradient at an acute angle of 45° or less with respect to the axial direction, and the axial lip refers to a seal lip that provides a sealing effect with a seal sliding surface that has a gradient of more than 45° with respect to the axial direction.

[0066] Furthermore, in this embodiment, an inner ring rotating ball bearing is exemplified, but the present invention can also be applied to an outer ring rotating ball bearing.

[0067] Furthermore, in this embodiment, a ball bearing configured as a deep groove ball bearing is exemplified, but the present invention can also be applied to an angular contact ball bearing.

[0068] In addition, although this embodiment shows an example in which a seal lip is provided on the inner diameter side of the seal member and a seal sliding surface is formed on the inner raceway, this invention can also be applied to cases in which a seal lip is provided on the outer diameter side of the seal member and a seal sliding surface is formed on the outer raceway. Also, although an example in which the seal sliding surface is formed on the raceway has been shown, it may also be formed on another member that is integrated with the raceway, such as a slinger.

[0069] The embodiments disclosed herein should be considered to be illustrative in all respects and not restrictive. Therefore, the scope of the present invention is defined by the claims rather than the above description, and it is intended to include all modifications within the meaning and scope of the claims. [Explanation of symbols]

[0070] 1 ball bearing 2. Sealing material 3 Inner raceway 4 Outer raceway 5 balls 6 Cage 7. Interior Space 10 Elastic material 16 Sealing lip 18 Seal sliding surface 23 Protrusion 24 Oil passage d ball bearing bore diameter h Height of the protrusion X Contact ellipse diameter

Claims

1. a ball bearing having an inner raceway, an outer raceway, a plurality of balls disposed between the raceways, and a cage that holds the balls; seal members located on one axial side and the other axial side of the balls and the cage, and disposed between the inner raceway ring and the outer raceway ring; a seal sliding surface formed on the inner raceway ring and sliding in a circumferential direction relative to the seal member, the seal member has a seal lip including a plurality of protrusions arranged in a circumferential direction, In a sealed bearing, the plurality of protrusions form an oil passage between the seal lip and the seal sliding surface, which communicates between an internal space of the ball bearing and an outside thereof, A sealed bearing characterized in that, when the inner diameter (mm) of the ball bearing is d, the height (mm) of the protrusion is h, A is a coefficient for d, and B is a constant, the following relationships are satisfied: 0.08≦h≦A×d+B, 0.002≦A≦0.004, and −0.043≦B≦−0.

002.

2. 2. The sealed bearing according to claim 1, wherein X is the diameter (mm) of the contact ellipse between the raceway and the balls when a load that is 15% of the basic dynamic load rating of the ball bearing is applied to the ball bearing, and h is the height (mm) of the protrusion, and the relationship between X and h satisfies 0.08≦h≦{(0.044X−0.15) / 2}+0.

02.

3. 3. A sealed bearing according to claim 1, wherein the inner diameter d of the ball bearing is 40 mm or more.

4. 3. The sealed bearing according to claim 1, wherein the hardness of the elastic material forming the seal lip is 60 HS or more and 80 HS or less.

5. 5. A sealed bearing according to claim 4, wherein the elastic material is either a thermosetting elastomer or a thermoplastic elastomer.

6. 3. A sealed bearing according to claim 1, which supports a rotating part of either a traction motor or a reduction gear of a vehicle.

Images