JP7201752B2 - Sealed bearing - Google Patents

Description

The present invention relates to a sealed bearing comprising a rolling bearing and a sealing member.

For example, in a transmission mounted on a vehicle such as an automobile and various construction machines, foreign matters such as wear powder of gears are mixed. For this reason, sealing members are used to prevent foreign matter from entering the inner space of the bearing and prevent early damage to the rolling bearing.

A typical sealing member has a sealing lip made of a rubber-like material or the like. Bearing parts such as bearing rings, slingers, and the like that rotate in the circumferential direction with respect to the seal member are formed with a seal sliding surface that makes sliding contact with the seal lip. Since the seal lip and the seal sliding surface are in sliding contact over the entire circumference, the drag resistance (seal torque) of the seal lip causes an increase in bearing torque. Moreover, the friction of the sliding contact accelerates the temperature rise of the rolling bearing. As this temperature rise progresses, the pressure difference between the inner space of the bearing and the outside invites an adsorption effect, increasing the friction.

In order to suppress the seal torque of such a contact seal member, shot peening is applied to the seal sliding surface to make the seal sliding surface having fine unevenness with a maximum roughness Ry of 2.5 μm or less, and lubricating oil stored in the concave portion (Patent Document 1).

JP 2007-107588 A

However, the reduction in torque by shot peening as in Patent Document 1 is achieved by reducing the sliding area between the seal lip and the seal sliding surface, but there is a limit to the reduction, so the reduction that can be achieved is Torque was limited.

In addition, since the temperature of the lubricating oil is relatively low at the beginning of the bearing operation, the viscosity of the lubricating oil is relatively high, and the lubricating conditions make it easy to form an oil film. Then, the lubricating condition becomes such that the oil film is easily broken. As the bearing is operated at higher speeds, the peripheral speed of the seal sliding surface relative to the seal lip increases, and the heat generated due to the friction between the seal lip and the seal sliding surface increases. Abrasion progresses easily. Therefore, there is a limit to the high-speed operation and allowable rotational speed of the sealed bearing due to lubrication conditions. In electric vehicles (EV), there is a strong demand for high-speed operation of sealed bearings that support the rotating parts of the drive system, but the demand cannot be met by reducing torque by shot peening.

If a non-contact seal member is used, it is possible to eliminate the seal torque, but it becomes difficult to manage various errors in the size of the gap between the seal member and the bearing parts so as to prevent the intrusion of foreign matter of a predetermined particle size.

In view of the above-described background, the problem to be solved by the present invention is to reduce the torque of a sealed bearing while preventing the intrusion of foreign matter of a predetermined particle size and enabling the bearing to operate at high speed.

In order to achieve the above object, the present invention provides a seal member that separates a space between a bearing interior and an exterior, a seal lip provided on the seal member, and a seal slide that slides in the circumferential direction with respect to the seal lip. and a plurality of projections that are formed on the seal lip and create between the seal sliding surface and the seal lip an oil passage communicating between the bearing inner space and the outside, wherein the projection and the The gap between the seal sliding surfaces is formed in a wedge shape with a large gap on the oil passage side and a small gap on the projection side. It is configured in a bearing.

According to the above configuration, an oil passage is formed by the protrusion between the seal sliding surface and the seal lip, and the lubricating oil in the oil passage is dragged by a wedge effect between the seal sliding surface and the seal lip as the bearing rotates. It promotes the formation of an oil film between them. Therefore, the coefficient of friction between the seal lip and the seal sliding surface is lowered, and the seal torque is reduced. Furthermore, since the oil passage improves the oil passage between the inner space of the bearing and the outside, the temperature rise of the rolling bearing is suppressed, and the adsorption effect is also prevented.
The smaller the interval between the protrusions adjacent to each other in the circumferential direction, that is, the greater the number of protrusions, the more the number of protrusions passing per rotation when the seal sliding surface rotates in the circumferential direction relative to the seal lip. become more. When the protrusions are arranged in the circumferential direction at intervals of 2.6 mm or less, the oil film is maintained in a continuous state over the entire circumferential direction of the seal sliding surface, and the wedge effect between each protrusion is maintained. Thus, the bearing can be operated in a state where the seal lip and the seal sliding surface are completely separated by the oil film and do not come into direct contact (that is, in a fluid lubrication state). In the fluid lubrication state, the seal torque is substantially reduced to zero, the seal lip is not substantially worn, and heat generation due to sliding between the seal lip and the seal sliding surface can be suppressed. Therefore, the permissible peripheral speed of the seal sliding surface relative to the seal lip also increases, making it possible to meet the demand for high-speed operation of bearings with seals, which could not be achieved in the past.
If the distance between the protrusions is set to less than 0.3 mm, it becomes difficult to manufacture a mold for molding the seal lip, so it is preferable to set the distance to 0.3 mm or more.
Also, the particle size of the foreign matter that can pass through the oil passage can be determined based on the projection height of the projection. Therefore, it is possible to arbitrarily determine the particle size of foreign matter to be prevented from entering, and to prevent the foreign matter having the predetermined particle size from entering from the oil passage.
Thus, by adopting the above-described configuration, the present invention can reduce the torque of the sealed bearing while preventing the intrusion of foreign matter of a predetermined particle size and enabling the bearing to operate at high speed.

1 is a sectional view showing a sealed bearing according to a first embodiment of the present invention; FIG. Enlarged view of the vicinity of the seal lip of the seal member on the right side of the drawing in FIG. Enlarged cross-sectional view of line III-III in FIG. FIG. 2 is a partial front view showing the seal lip according to the first embodiment from the axial direction; Diagram showing particle size distribution and number of foreign matter contained in lubricating oil in vehicle transmission (AT/MT) Pie chart showing percentage of particle size distribution in FIG. Diagram showing particle size distribution and number of foreign matter contained in lubricating oil in vehicle transmission (CVT) Pie chart showing percentage of particle size distribution in FIG. Graph showing the test results of measuring the torque of the bearing for each R dimension of the projection Lubrication area diagram showing fluid lubrication mode in the first embodiment A diagram showing the relationship between the distance between projections, the theoretical oil film thickness, and the bearing rotational torque in the first embodiment. A diagram showing the relationship between the R dimension of the projection and the theoretical oil film thickness in the first embodiment. FIG. 4 is a partial perspective view showing a seal lip according to a second embodiment of the invention; (a) is a schematic diagram showing vulcanization molding of the seal lip of FIG. 13, and (b) is a schematic diagram of the seal lip molded in (a). (a) is a schematic diagram showing vulcanization molding of the seal lip of the virtual model, and (b) is a schematic diagram of the seal lip molded in (a). FIG. 3 is a cross-sectional view showing a sealed bearing according to a third embodiment of the present invention; Enlarged view of the vicinity of the seal lip of the seal member on the right side of the drawing in FIG. FIG. 4 is a cross-sectional view showing a sealed bearing according to a fourth embodiment of the present invention; FIG. 5 is a cross-sectional view showing a sealed bearing according to a fifth embodiment of the present invention; A diagram showing the relationship between the distance between projections and the theoretical oil film thickness in the fifth embodiment. A sectional view showing an example of a transmission provided with a sealed bearing according to the present invention.

A preferred embodiment of the invention will now be described.
In the first embodiment, the projection extends in a direction orthogonal to the circumferential direction, and the projection gradually approaches the seal sliding surface from both ends of the circumferential width toward the center of the circumferential width. It has a shape. According to the first embodiment, the projection extends in a direction orthogonal to the circumferential direction, which is the sliding direction with the seal sliding surface, and has an R shape that reduces the area where the projection may come into sliding contact. The region of sliding contact between the projection and the seal sliding surface can be linear. In addition, with such an R shape, the wedge angle of the wedge-shaped gap described above gradually decreases from the wide side to the narrow side, so that the wedge effect is effectively generated to increase the hydraulic pressure in the linear region. It is easy to make the lubricating state between the projection and the seal sliding surface into a fluid lubricating state. In addition, even if the projection rubs against the seal sliding surface during installation of the seal member, there is no concern that the R-shaped projection will be bent in the circumferential direction, and there is no risk of impairing the seal torque reduction performance during installation.

In the first embodiment, the height of the protrusion is set to 0.07 mm or less. Foreign matter with a grain size exceeding 0.05 mm (50 μm) adversely affects bearing life. It has been found that setting the height of the protrusion to 0.07 mm or less can create a narrow gap (including the oil passage) between the seal lip and the seal sliding surface through which such foreign matter cannot easily pass. rice field.

In the first embodiment, the R dimension of the projection is 0.15 mm or more and less than 2.0 mm, and the circumferential width of the projection is 0.2 mm or more and 1.0 mm or less. The larger the R dimension of the projection, the more easily the wedge effect occurs. When the R dimension is 0.15 mm, it is possible to form an oil film sufficient to completely separate the projection and the seal sliding surface. When the height of the projection is set to 0.07 mm or less, it is preferable to set the R dimension to 0.15 mm or more and less than 2.0 mm so as not to make the mold manufacturing of the seal lip difficult. Since the circumferential width of the protrusion depends on the R dimension, it is preferably set to 0.2 mm or more and 1.0 mm or less.

In the second embodiment, the protrusions are arranged at regular intervals over the entire circumference. According to the second embodiment, it is possible to uniformly promote the formation of an oil film over the entire circumference of the seal sliding surface.

A first embodiment according to the present invention will be described below with reference to FIGS. 1 to 12. FIG. As shown in FIG. 1, the first embodiment includes an inner ring 110, an outer ring 120, a plurality of rolling elements 140 held by a retainer 130, and a bearing internal space formed between the inner ring 110 and the outer ring 120. The sealed bearing 100 is provided with two seal members 150 that seal both ends. In addition, below, the direction along the bearing central axis of the bearing 100 with seal is called "axial direction." A direction orthogonal to the axial direction is called a “radial direction”. The circumferential direction around the bearing center axis is called the “circumferential direction”.

An annular bearing inner space 160 is formed by the inner ring 110 and the outer ring 120 . The plurality of rolling elements 140 revolves while being interposed between the inner ring 110 and the outer ring 120 within the bearing inner space 160 . Lubricating oil is supplied to the bearing inner space 160 by appropriate means such as grease or an oil bath.

The inner ring 110 is attached to a rotating shaft (not shown) and rotates integrally with the rotating shaft. The rotating shaft is provided, for example, as a rotating part of a vehicle transmission or differential. The outer ring 120 is attached to a member such as a housing, a gear, or the like that applies the load from the rotating shaft.

This sealed bearing 100 is a deep groove ball bearing. A ball is adopted as the rolling element 140 . The inner ring 110 and the outer ring 120 respectively have raceway grooves 111 and 121 which correspond to about 1/3 of the circumference of the rolling element 140 in cross section (corresponding to the cross section shown in the drawing) and which are continuous over the entire circumferential direction. .

A seal groove 122 for holding a seal member 150 is formed at the end of the inner circumference of the outer ring 120 . The seal member 150 is attached to the outer ring 120 by press-fitting its outer peripheral edge into the seal groove 122 .

The seal member 150 separates the bearing interior space 160 and the exterior. On the outside of the seal member 150 as a boundary, there are foreign matter depending on where the bearing with seal 100 is installed, such as wear powder from gears, wear powder from clutches, fine crushed stones, and the like. Such powdery foreign matter can reach the vicinity of the seal member 150 due to the flow of lubricating oil or atmosphere. The seal member 150 prevents foreign matter from entering the bearing internal space 160 from the outside.

The seal member 150 has a tongue-shaped seal lip 151 protruding from its inner peripheral side. A seal sliding surface 112 that slides on the seal lip 151 in the circumferential direction is formed on the outer periphery of the inner ring 110 . The seal sliding surface 112 is in the shape of a cylindrical surface covering the entire circumferential direction. The seal lip 151 is a radial lip. Here, the radial lip is a seal lip that exerts a sealing action with a seal sliding surface along the axial direction or a seal sliding surface having an acute angle gradient of 45° or less with respect to the axial direction. It has a radial interference between it and the moving surface.

FIG. 2 shows an enlarged view of the vicinity of the seal lip 151 of the seal member 150 of FIG. FIG. 3 shows a cross-sectional view taken along line III--III in FIG. This cross section is designed to define a gap between the seal lip 151 and the seal sliding surface 112 in the direction perpendicular to the seal sliding surface 112 (including the oil passage 170) in the direction perpendicular to the seal sliding surface 112. It shows the state at the narrowest point. 4 shows the appearance of the seal lip 151 when viewed in the axial direction from the side of the inner space of the bearing. FIG. 4 depicts the outline of the seal lip 151 in the single and natural state of the seal member 150 shown in FIG. Here, the natural state refers to a state in which no external force acts on the seal member in a single state, that is, the state in which the seal member is not deformed by an external force (hereinafter, this state is simply referred to as "natural state"). call it a state”.).

As shown in FIGS. 2 to 4, the seal lip 151 has a protrusion 152 having a protruding height in a direction perpendicular to the seal sliding surface 112, that is, in a normal direction perpendicular to a tangential line in contact with the seal sliding surface 112. have. Since the seal sliding surface 112 has a cylindrical surface shape centering on the bearing center axis, the direction orthogonal to this corresponds to the radial direction.

A radial interference is set between the seal lip 151 and the seal sliding surface 112 . Due to this interference, the seal lip 151 radially pressed against the seal sliding surface 112 undergoes rubber-like elastic deformation in which the seal lip 151 bends outward, and the seal lip 151 exerts a straining force. Mounting errors, manufacturing errors, etc. of the seal member 150 are absorbed by changes in the degree of curvature of the seal lip 151 .

As shown in FIG. 4, the seal lip 151 has a tip 153 that defines the inner diameter of the seal lip 151 in its natural state.

As shown in FIGS. 2 to 4, the protrusions 152 extend in a direction orthogonal to the circumferential direction. The protrusion 152 extends to the tip 153 of the seal lip 151 and is formed over substantially the entire range having a radial interference with the seal sliding surface 112 .

The protrusions 152 are arranged at regular intervals d in the circumferential direction. Considering the appearance of the seal lip 151 viewed from the axial direction, a plurality of projections 152 appear radially arranged in the circumferential direction at a constant pitch angle θ corresponding to the interval d. The radial center is on the central axis of the seal member 150 (not shown) (coincident with the bearing central axis).

The distance d between the circumferentially adjacent protrusions 152 and the circumferential width w of the protrusions 152, together with the presence of each radially arranged protrusion 152 near the tip 153 of the seal lip 151, provide a seal. The lip 151 is set so that it can slide against the seal sliding surface 112 only on each projection 152 , and an oil passage 170 is always created between each projection 152 . That is, when the seal member 150 is attached, the projection 152 in contact with the seal sliding surface 112 stretches against the tension force of the seal lip 151 , so that the inner space 160 and the outer space of the bearing are formed on both sides of the projection 152 in the circumferential direction. An oil passage 170 is created which communicates therebetween. Lubricating oil reaches the bearing inner space 160 from the outside through the oil passage 170 . The lubricating oil that has entered the bearing internal space 160 and the base oil in the case where grease is enclosed reach the outside through the oil passage 170 from the bearing internal space 160 .

The particle size that can pass through the oil passage 170 can be determined based on the protrusion height h of the protrusion 152 perpendicular to the seal sliding surface 112 . Therefore, in the first embodiment, it is possible to arbitrarily determine the particle size of foreign matter to be prevented from entering, and to prevent foreign matter of the predetermined particle size from entering the bearing internal space 160 from the oil passage 170 .

Abrasion powder that causes early failure of rolling bearings is foreign matter having a particle size of more than 50 μm. It has been found that setting the projection height h of the projection 152 to 0.07 mm or less can create an oil passage 170 through which such abrasion powder cannot pass. The protrusion height h is preferably set to 0.04 mm or less, and is set to 0.07 mm or less in consideration of tolerance. On the other hand, by setting the protrusion height h of the protrusion 152 to 0.07 mm, the oil passage of the oil passage 70 can be improved.

The gap generated between the protrusion 152 and the seal sliding surface 112 is formed in a wedge shape, the larger the side closer to the oil passage 170 in the circumferential direction and the smaller the side closer to the protrusion 152 in the circumferential direction. A surface 154 of the projection 152 is curved corresponding to the aforementioned wedge shape. As shown in FIG. 3, when the seal sliding surface 112 rotates in the circumferential direction with respect to the seal lip 151 as the inner ring 110 rotates (the direction of rotation is indicated by arrow A in the figure), the oil passage As the seal sliding surface 112 rotates, the lubricating oil in 170 (indicated by dots) is drawn between the seal sliding surface 112 and the protrusion 152 of the seal lip 151, forming an oil film therebetween. promote Therefore, the coefficient of friction (μ) between the seal lip 151 and the seal sliding surface 112 is lowered, and the seal torque is reduced. Furthermore, the oil passage between the bearing inner space 160 and the outside is improved by the oil passage 170 . Therefore, the temperature rise of the sealed bearing 100 is suppressed, and the adsorption action of the seal lip 151 is also prevented.

As shown in FIG. 1, the sealing member 150 is formed of a cored bar 155 made of a metal plate and a vulcanized rubber material 156 attached to at least the inner diameter portion of the cored bar 155 . The seal lip 151 is formed in a tongue shape from a vulcanized rubber material 156 . The cored bar 155 is a press-formed part that is annularly formed over the entire circumference. The vulcanized rubber material 156 is a vulcanized rubber portion. The sealing member 150 is manufactured as an integral part by, for example, putting the core metal 155 into a mold and vulcanizing and molding the vulcanized rubber material 156 . The vulcanized rubber material 156 may be attached to the entire cored bar 155 or may be attached only to the inner diameter portion of the cored bar 155 .

Thus, in the first embodiment, the protrusion 152 can be formed on the seal lip 151 during vulcanization molding of the seal lip 151, and the seal sliding surface 112 can be easily processed into a cylindrical shape, It is easy to directly form the same cross-sectional shape, such as a groove shape, on the bearing ring over the entire circumference.

Due to the straining force of the seal lip 151 and the oil pressure of the lubricating oil, substantial deformation (deformation that affects the lubrication performance between the protrusion 152 and the seal sliding surface 112) does not occur in the solid projection 152. there is Therefore, the shape of the protrusion 152 during bearing operation may be considered to be the same as the shape transferred during vulcanization molding of the seal lip 151 .

This sealed bearing 100 is intended for use in supporting rotating parts in vehicle transmissions. Oiling of sealed bearings present in vehicle transmissions is generally accomplished by any suitable method, such as splash, oil bath, or nozzle injection. Therefore, lubricating oil exists around the seal lip of the seal member fixed to the inner ring or outer ring of the sealed bearing. The lubricating oil supplied is commonly used for other lubricating parts such as gears in the transmission. The lubricating oil is circulated by an oil pump and filtered by an oil filter provided in the circulation path.

The inventors of the present application collected lubricating oils actually used in the market according to the travel distance of the vehicle, and investigated the number of foreign substances mixed in the used lubricating oil, the particle size distribution of the foreign substances, and the materials of the foreign substances. . FIG. 5 shows the number of foreign matter and the particle size distribution of the lubricating oil collected from eight automatic transmission (AT) or manual transmission (MT) vehicles. The vertical axis in FIG. 5 has a logarithmic scale, the horizontal axis represents the distance traveled by the vehicle, and the vertical axis represents the number of foreign matter (fine particles) per 100 ml. Contaminants to be counted had a particle size of 5 μm or more. Counting was performed for each particle size category. The classification is as follows: particle diameter 5 μm or more and less than 15 μm, particle diameter 15 μm or more and less than 25 μm, particle diameter 20 μm or more and less than 50 μm, particle diameter 50 μm or more and less than 100 μm, and particle diameter 100 μm or more. The measurement here was carried out using a measuring instrument of model number 8000A manufactured by Hyacroico Co., Ltd., using a fine particle amount of contaminants. The particle size distribution of FIG. 5 is shown in a pie chart in FIG.

FIG. 7 shows, in the same manner as FIG. 5, the number and particle size distribution of foreign matter examined for lubricating oil collected from 10 continuously variable transmission (CVT) vehicles. The particle size distribution of FIG. 7 is shown in a pie chart in FIG. The vehicle manufacturers, vehicle models, and mileage covered are different, but as is clear from a comparison of Figures 5 and 6 with Figures 7 and 8, AT/MT, which uses a lot of gears, is better than CVT. Also, there was a tendency for both the particle size and the number of foreign particles to be large. Moreover, 99.9% or more of the particle size distribution was 50 μm or less, regardless of the type of transmission. The number of foreign matter with a particle size of more than 50 μm was less than 1000 for AT/MT and less than 200 for CVT even when the traveling distance was increased. This indicates that in recent years, the performance of oil filters has improved, and foreign matter in lubricating oil has become finer (that is, foreign matter with a large particle size is removed by the oil filter).

On the other hand, when the lubricating oil inside the bearing contains foreign matter, we investigated the relationship between the particle diameter of the foreign matter and the bearing life. However, if there is a small amount of foreign matter with a particle size of 50 μm or more, such as the environment inside the transmission in recent years, even if the foreign matter enters the bearing without a seal, the life ratio of the rolling bearing (actual life) will be reduced. It was found that the ratio to the calculated service life shows a value (for example, about 7 to 10 times) that can withstand practical use in automobile transmissions.

Based on the above results, when lubricating oil that is filtered by an oil filter is supplied to sealed bearings used to support rotating parts of drive systems such as vehicle transmissions and differential gears, it is found that large particles with a particle size exceeding 50 μm As long as the seal member prevents foreign matter from entering the bearing, it can be said that allowing foreign matter with a particle size of 50 μm or less contained in the lubricating oil to enter the bearing does not cause a problem in bearing life. If this is allowed, the flow of lubricating oil between the seal lip and the seal sliding surface should be sufficiently ensured, and coupled with the aforementioned wedge effect, fluid lubrication between the seal lip and the seal sliding surface should be achieved. It is feasible to make the state

Therefore, as shown in FIGS. 2 and 3, the height h of the protrusion 152 is set to 0.07 mm. The height h of the protrusion 152 is the highest value within the range where it can be in sliding contact with the seal sliding surface 112 in terms of design. This position is also where the interference set between each projection 152 and the seal sliding surface 112 is maximized. Since the amount of deformation of the projection 152 during bearing operation is negligible, the clearance (including the oil passage 170) between the seal lip 151 and the seal sliding surface 112 in the orthogonal direction to the seal sliding surface 112 is At its narrowest point in the direction perpendicular to the surface 112, it is as wide as the height h of the protrusion 152, and does not substantially exceed 0.07 mm. Therefore, even if the external lubricating oil contains foreign matter having a particle size of more than 50 μm, it is considered that the foreign matter does not pass through the oil passage 170 .

The protrusion 152 has an R shape that gradually approaches the seal sliding surface 112 from both ends of the circumferential width w toward the center of the circumferential width. This R shape is given over the entire length of the projection 152 in the radial direction. Therefore, the region where the protrusion 152 and the seal sliding surface 112 can slide and contact each other exists linearly on the imaginary axial plane Pax passing through the center of the circumferential width of the protrusion 152 . The curvature center of the R shape of the protrusion 152 is on the virtual axial plane Pax.

When the seal sliding surface 112 rotates relative to the seal lip 151 in the direction of arrow A in the drawing, the lubricating oil in the oil passage 170 is dragged into the wedge-shaped gap between the projection 152 and the seal sliding surface 112. be taken in. Since the wedge angle in the wedge-shaped gap described above gradually decreases from the wide side of the oil passage 170 where lubricating oil is drawn in toward the narrow side, the projection 152 and the seal sliding surface 112 are in sliding contact. The closer to the obtained linear region (on the imaginary axial plane Pax), the stronger the wedge effect. Therefore, the hydraulic pressure of the oil film in the linear region can be increased more effectively, the projection 152 can be completely separated from the seal sliding surface 112, and the oil film in the linear region can be thickened. The lubrication state between 152 and the seal sliding surface 112 can be easily brought into a fluid lubrication state.

Here, if there is an oil film that completely separates the protrusion 152 and the seal sliding surface 112 , the seal sliding surface 112 slides against the protrusion 152 without direct contact, resulting in fluid lubrication. By maintaining such an oil film between each projection 152 and the seal sliding surface 112, the seal lip 151 and the seal sliding surface 112 can be in a state of fluid lubrication.

In order to easily realize the fluid lubrication state, it is better to set the tightening force of the seal lip 151 based on the interference between the seal lip 151 and the seal sliding surface 112 as weak as possible. For this reason, the waist portion of the seal lip 151, which is subjected to bending deformation toward the outside, is formed as thin as possible.

Also, the smaller the maximum height roughness Rz, the smaller the thickness of the oil film required for fluid lubrication. Therefore, the seal sliding surface 112 is not subjected to shot peening treatment, and the maximum height roughness Rz of the seal sliding surface 112 is less than 1 μm. Here, the maximum height roughness Rz means the maximum height roughness defined in JIS B0601:2013.

Also, the smaller the distance d between the projections 152 adjacent in the circumferential direction, that is, the greater the number of projections 152, the more the seal sliding surface 112 rotates in the circumferential direction relative to the seal lip 151. The number of passages of the contacting projections 152 increases, and the oil film is kept continuous over the entire circumferential direction of the seal sliding surface 112, and the wedge effect between the projections 152 tends to occur continuously. , it becomes easier to maintain the fluid lubrication state.

Also, the larger the R dimension of the protrusion 152 (the radius of curvature of the surface 154 of the protrusion 152), the more easily the wedge effect occurs.

In addition, when the torque measurement test of the bearing was performed for each R dimension of the projection 152 (see FIG. 9), when the R dimension of the projection 152 was 1.5 mm, it was compared with when the R dimension was 0.2 mm or 0.8 mm. As a result, the torque at a temperature of 80° C., which is close to the actual use environment of the bearing 100, has the minimum value. When the R dimension of the projection 152 was 1.8 mm, a result equivalent to 1.5 mm was obtained. In order to easily generate the wedge effect and achieve low torque, the R dimension of the projection 152 is preferably 1.5 mm or more and 2.0 mm or less in consideration of tolerance. As a reference, FIG. 9 shows the theoretical torque of a bearing that is not equipped with a seal and is used in an oil bath lubrication environment where the center of the rolling element of the bearing (lowest end of the bearing) is the oil surface. Also shown are calculated values. By setting the R dimension of the projection 152 to 1.5 mm or more and 2.0 mm or less (more specifically, 1.5 mm or more and 1.8 mm or less), there is no seal torque because no seal is provided, and , a low torque equivalent to that of a bearing (calculated values shown in FIG. 9), which is considered to have low agitation resistance, can be realized.

Assuming that the height h of the protrusions 152 is 0.07 mm, the inventors of the present application determined that the distance d between the protrusions 152 should be 0.3 mm or more and 2.6 mm or less, and the circumferential width w of the protrusions 152 should be 0.2 mm or more and 1.0 mm. Below, the theoretical oil film thickness between the protrusion 152 and the seal sliding surface 112 was calculated in the range of the R dimension of the protrusion 152 from 0.15 mm to less than 2.0 mm. The calculation was performed with the speed (peripheral speed) at which the seal sliding surface 112 rotates in the circumferential direction relative to the seal lip 151 in the range of 0.02 to 20.2 m/s. In addition, the calculation was performed at oil temperatures of 30°C and 120°C, assuming that the lubricating oil is a CVTF that lubricates the pulleys and belts of the CVT. As a result, under these conditions of use, in the lubrication region diagram in the case of line contact based on the viscosity parameter gv and the elastic parameter ge, which are dimensionless numbers determined by Greenwood-Johnson, the isoviscosity-rigid region ( RI mode) or an equiviscosity-elastic region (EI mode, soft EHL), that is, a fluid lubrication state. FIG. 10 shows the distribution of the calculation results in the above lubrication area diagram.

Further, the inventor of the present application calculated the relationship between the distance d between the projections 152 shown in FIGS. Further, the inventor of the present application calculated the relationship between the R dimension of the protrusion 152 and the theoretical oil film thickness between the protrusion 152 and the seal sliding surface 112 . For the theoretical oil film thickness, Martin's minimum film thickness calculation formula was used in the RI mode, and Herrebrough's minimum film thickness calculation formula was used in the EI mode. Furthermore, the inventor of the present application conducted an experiment to investigate the relationship between the distance d between the projections 152 and the rotational torque of the bearing 100 with seal. FIG. 11 shows the relationship between the distance d between the projections 152, the theoretical oil film thickness, and the bearing rotation torque. FIG. 12 shows the relationship between the R dimension of the projection 152 and the theoretical oil film thickness. If the oil film between the projection 152 and the seal sliding surface 112 is too thin, the coefficient of friction μ will increase. An optimum oil film thickness may be set on the premise that the oil film thickness exceeds Rz.

From FIG. 11, when the distance d between the projections 152 is 2.6 mm, the oil film between the projections 152 and the seal sliding surface 112 is calculated to be about 3 μm (maximum height roughness Rz of less than 1 μm). ) is formed, and it can be seen that the oil film tends to become thicker when it is smaller than 2.6 mm, and when it is 2.6 mm or less, the bearing rotational torque tends to decrease (that is, the seal torque tends to decrease). I understand. Therefore, the distance d between the protrusions 152 can be set to 2.6 mm or less. If the distance d between the protrusions 152 is less than 0.3 mm, it becomes difficult to mold the seal lip 151, so it is preferable to set the distance d to 0.3 mm or more.

As can be seen from FIG. 12, when the R dimension of the protrusion 152 is 0.15 mm, an oil film of 3 μm or more is formed by calculation. Therefore, when the height h of the protrusion 152 is set to 0.07 mm, it is preferable to set the R dimension to 0.15 mm or more and less than 2.0 mm in consideration of molding with a mold. Further, when the height h of the projection 152 is set to 0.07 mm, the circumferential width w of the projection 152 depends on the R dimension, so the circumferential width w of the projection 152 is set to 0.2 mm or more and 1.0 mm or less. preferably.

Based on the calculation results so far, etc., in the first embodiment, the height h of the projections 152 as shown in FIGS. 2.6 mm or less, the circumferential width w of the protrusion 152 is set to 0.2 mm or more and 1.0 mm or less, and the R dimension of the protrusion 152 is set to a range of 0.15 mm or more and less than 2.0 mm. . That is, the seal lip 151 is provided with the projection 152 in such a manner that the seal lip 151 and the seal sliding surface 112 can be fluidly lubricated at the peripheral speed of the seal sliding surface 112 of 0.2 m/s or higher. formed. When the peripheral speed is less than 0.2 m/s, boundary lubrication occurs between the protrusion 152 and the seal sliding surface 112 . Since the circumferential speed of 0.2 m/s is a speed that the vehicle transmission reaches quickly after it stops, the fluid lubrication state can be maintained between the seal lip 151 and the seal sliding surface 112 for substantially the entire bearing operating time. can.

As described above, the sealed bearing 100 according to the first embodiment prevents foreign matter having a particle size that adversely affects the service life of the bearing from entering the inner space of the bearing, while the seal lip 151 and the seal slide are prevented. The coefficient of friction μ of sliding between the moving surfaces 112 can be reduced to the limit by fluid lubrication, and the seal torque can be significantly reduced to significantly reduce the bearing rotation torque (see FIGS. 1 and 3). ).

Furthermore, the bearing 100 with seal according to the first embodiment has a peripheral speed of the seal sliding surface that would otherwise cause problems such as wear of the seal lip and heat generation due to sliding between the seal lip and the seal sliding surface. When the seal lip 151 and the seal sliding surface 112 are in a fluid lubricating state without direct contact when the seal lip 151 and the seal sliding surface 112 are operated at a speed (for example, 30 m/s or more), wear of the seal lip 151 is substantially eliminated. At the same time, the aforementioned heat generation can be suppressed. Therefore, the sealed bearing 100 according to the first embodiment can meet the demand for high-speed operation of the sealed bearing, which could not be achieved conventionally.

Furthermore, in the sealed bearing 100 according to the first embodiment, the protrusion 152 is formed in an R shape, so even if the protrusion 152 is rubbed against the seal sliding surface 112 when the seal member 150 is attached to the outer ring 120, , there is no fear that the protrusion 152 will be bent in the circumferential direction, and there is no fear that the seal torque reduction performance will be impaired during installation. For example, if the protrusions are sharpened, it is not possible to know which direction in the circumferential direction the many protrusions rubbed against the seal sliding surface will bend when the seal member is installed, and the rotation direction relative to the seal sliding surface cannot be determined. It is extremely difficult to install so that all the projections are bent towards a suitable wedge gap. The wedge effect cannot be satisfactorily obtained at the protrusions bent in an inappropriate direction, and the seal torque reduction performance is impaired.

A second embodiment will be described with reference to FIGS. 13 to 15. FIG. The second embodiment is different from the first embodiment only in the shape of the projection. As shown in FIG. It has a shape that gradually becomes lower. 13 is an enlarged perspective view of the vicinity of the projection 201 of the seal lip 202 in its natural state. Regarding the R dimension (curvature radius of the surface 204 of the projection 201) and the center of curvature of the projection 202, since the projection 201 is gradually lowered toward the tip 203 of the seal lip 202, the R dimension is gradually increased toward the tip 203 of the seal lip 202. is enlarged, and the center of curvature is moved to the outer side.

The height of the protrusion 201 is substantially zero above the tip 203 of the seal lip 202 . Therefore, the protrusion 201 does not reach the tip 203 of the seal lip 202, and a flat surface 205 exists between the protrusion 201 and the tip 203 of the seal lip 202. For this reason, the tip 203 of the seal lip 202 serves as a boundary line where a surface 205 of the seal lip 202 on the inner space side of the seal 202 extending in the circumferential direction and a surface of the seal lip 202 on the outer side of the seal lip 202 extending in the circumferential direction intersect. ing.

FIG. 14 shows how the seal lip 202 is vulcanized. It should be noted that FIG. 14 is drawn schematically for easy understanding, and also roughly shows the shape of the seal lip 202 . Vulcanization molding of the seal lip 202 is performed by vulcanization molding of a rubber sheet on the metal core 206 . At this time, the rubber sheet is sandwiched between the upper mold Mp1 and the lower mold Mp2, and the rubber portion such as the seal lip 202 of the seal member is molded. The vertical direction in which the upper mold Mp1 and the lower mold Mp2 are aligned corresponds to the axial direction. Therefore, the tip 203 that defines the inner diameter of the seal lip 202 in the natural state is the boundary line between the upper surface of the seal lip 202 in contact with the transfer surface of the upper mold Mp1 and the lower surface of the seal lip 202 in contact with the transfer surface of the lower mold Mp2. Therefore, it is positioned on the parting line Pl, which is the joining portion of the upper mold Mp1 and the lower mold Mp2.

Fig. 15 shows a hypothetical model in which the protrusion reaches the tip of the seal lip. In this virtual model, since the parting line Pl has unevenness for forming the projection 201' on the tip 203' of the seal lip 202', the burr 207 as shown in the figure is likely to occur after vulcanization. If the burr 207 is generated and separated from the seal lip 202' during operation of the bearing, it may cause clogging of the oil filter and lubricating oil circulation path.

On the other hand, as shown in FIG. 14, when the seal lip 202 has a shape having a flat surface 205 between the projection 201 and the tip 203 of the seal lip 202, the projections 201 are formed on the parting line Pl. There is no burr 207 as shown in FIG. Thus, the seal member according to the second embodiment can prevent burrs from being generated on the tip 203 of the seal lip 202 when the seal lip 202 is vulcanized.

The second embodiment shows an example in which the protrusion 201 does not have a height above the tip 203 of the seal lip 202, and a flat surface 205 exists between the protrusion 201 and the tip 203 of the seal lip 202. However, even if the protrusion has a height above the tip of the seal lip, if the protrusion gradually lowers toward the tip of the seal lip, the unevenness for forming the protrusion on the parting line is gentle. Therefore, burrs are less likely to occur on the tip of the seal lip.

A third embodiment will be described with reference to FIGS. 16 and 17. FIG. The inner ring 310 of the sealed bearing 300 according to the third embodiment has a seal groove 311 formed along the entire circumference. Seal members 330 and 340 for sealing both ends of the bearing inner space formed between the inner ring 310 and the outer ring 320 are held in the seal grooves 321 and 322 of the outer ring 320 .

The seal members 330 and 340 have seal lips 331 and 341 provided as axial lips and outer lips 332 and 342 located outside the seal lips 331 and 341 . Sealing lips 331 , 341 and outer lips 332 , 342 branch from waists that attach to cores 333 , 343 .

Here, the axial lip is a seal lip that exerts a sealing action with a seal sliding surface along the radial direction or a seal sliding surface having an acute angle gradient of less than 45° with respect to the radial direction. It has axial interference with the moving surface.

The seal sliding surfaces 312 and 313 that slide against the seal lips 331 and 341 of the seal members 330 and 340 are present on the side surfaces of the groove expanding from the groove bottom of the seal groove 311 toward the raceway groove 314 side. , has an acute angle α of less than 45° with respect to the radial direction.

Of the pair of shoulders 315 and 316 formed on both sides of the raceway groove 314 of the inner ring 310, the shoulder 315 on the load side (right side in the drawing) that receives the axial load is on the opposite non-load side (left side in the drawing). It is formed higher than the shoulder portion 316 . The shoulder portion 316 on the non-load side has a shoulder height corresponding to that of a deep groove ball bearing. Therefore, the sealed bearing 300 exhibits the good low-torque property of a deep groove ball bearing, and also has an axial load bearing capacity superior to that of a deep groove ball bearing.

Seal sliding surfaces 312 and 313 are formed on the shoulder portions 315 and 316, respectively. Since there is a large diameter difference between the outer diameter of the load-side shoulder 315 and the non-load-side shoulder 316, the diameter of the seal sliding surface 312 formed on the load-side shoulder 315 and The diameter of the seal sliding surface 313 formed on the shoulder portion 316 on the non-load side is different. That is, the diameter of the seal sliding surface 312 on the right side of the drawing is larger than the diameter of the seal sliding surface 313 on the left side of the drawing. Therefore, during operation of the bearing, the peripheral speed of the seal sliding surface 312 on the right side in the drawing becomes higher than that of the seal sliding surface 313 on the left side in the drawing.

Further, during operation of the bearing, a pump action is generated in the inner space of the bearing due to the difference in diameter between the shoulders 315 and 316, which sends the lubricating oil from the left side to the right side in the drawing.

The projections 334, 344 of the seal lips 331, 341 are formed in the radial direction during vulcanization molding. In the drawing, the shape of the seal lips 331, 341 corresponding to the natural state is drawn in order to show the interference between the seal sliding surfaces 312, 313 and the protrusions 334, 344. The protrusions 334, 344 are pressed against the seal sliding surfaces 312, 313 from the axial direction, so that the seal lips 331, 341 incline substantially along the seal sliding surfaces 312, 313, and the protrusions 334, 344 and the seal sliding surfaces 312 and 313, an oil passage and a wedge-shaped gap are formed (see FIG. 3).

The outer lips 332 , 342 shown in FIGS. 16 and 17 form a labyrinth clearance 350 with the outer groove wall portion of the seal groove 311 . Therefore, foreign matter with a particle size exceeding 50 μm cannot easily enter the seal groove 311 from the outside.

The number of protrusions 334 on the sealing member 330 on the right side of the drawing is different from the number of protrusions 344 on the sealing member 340 on the left side of the drawing. Also, the circumferential pitch angle of the projections 334 of the seal member 330 on the right side of the figure differs from the circumferential pitch angle of the projections 344 of the seal member 340 on the left side of the figure. These differences are due to the difference in lubrication conditions between the seal member 330 and the seal sliding surface 312 on the right side of the drawing and between the seal member 340 and the seal sliding surface 313 on the left side of the drawing due to the aforementioned peripheral speed difference and pump action. Therefore, it is necessary to optimize the oil film formed between the left and right sides to be equal and to prevent foreign matter exceeding 50 μm in diameter from easily entering due to the formation of an excessively thick oil film. set as a purpose. Therefore, in the sealed bearing 300 according to the third embodiment, the seal member 330 on the right side of the drawing and the seal sliding surface 312 and the seal member 340 on the left side of the drawing and the seal sliding surface 313 are appropriately In the fluid lubricating state, it is possible to achieve both low torque and suppression of foreign matter intrusion.

Further, in the sealed bearing 300 according to the third embodiment, the formation of the labyrinth clearance 350 makes it difficult for foreign matter to reach the seal lips 331 and 341. Therefore, it is possible to prevent foreign matter from entering the sealing lips 331 and 341 so as not to hinder torque reduction. can be suppressed. In general, a seal lip provided as an axial lip has a larger amount of movement in the axial direction during bearing operation than a seal lip provided as a radial lip. There may be large gaps in the Therefore, providing a seal lip as an axial lip is disadvantageous against intrusion of foreign matter. In the third embodiment, the disadvantage of the seal lips 331 and 341, which are axial lips, can be compensated for by the sealing effect of the labyrinth clearance 350. In contrast, there is no concern that the bearing life will be greatly deteriorated.

A fourth embodiment will be described with reference to FIG. A sealed bearing 400 according to the fourth embodiment includes an inner ring 410, an outer ring 420, a plurality of balls 430 interposed between the raceway grooves 411 and 421 of the inner ring 410 and the outer ring 420, and formed between the inner ring 410 and the outer ring 420. and two seal members 440 and 450 for sealing both ends of the inner space of the inner ring 410 . In common with the third embodiment in that the shoulder 412 is formed higher than the shoulder 413 on the opposite non-load side, the sealing members 440 and 450 are provided as radial lips. It is common to the first embodiment in that it has the seal lips 441 and 451 that are formed on the bottom.

The inner ring 410 has no seal groove and has a cylindrical seal sliding surface 414 that defines the outer diameter of the shoulder portion 412 on the load side and a cylindrical seal sliding surface 414 that defines the outer diameter of the shoulder portion 413 on the non-load side. and a seal sliding surface 415 of . The diameter difference between the seal sliding surface 414 on the right side of the drawing and the seal sliding surface 415 on the left side of the drawing is about the same as in the third embodiment, and the peripheral speed difference during bearing operation is also about the same. . Therefore, the difference in the number of protrusions 442 and 452 and the pitch angle in the circumferential direction between the sealing member 440 on the right side of the drawing and the sealing member 450 on the left side of the drawing is about the same as in the third embodiment. As a result, even in the sealed bearing 400 according to the fourth embodiment, the oil film between the seal member 440 and the seal sliding surface 414 and between the seal member 450 and the seal sliding surface 415 is optimally and equally formed. It achieves both low torque by forming a fluid lubrication state and suppression of foreign matter intrusion.

A fifth embodiment will be described with reference to FIG. A fifth embodiment is the application of the present invention to a tapered roller bearing. As shown in FIG. 19, a sealed bearing 500 according to the fifth embodiment includes an inner ring 510 having a raceway surface 511, a large rib 512 and a small rib 513, an outer ring 520 having a raceway surface 521, an inner ring 510 and an outer ring. A tapered roller bearing comprising a plurality of tapered rollers 530 interposed between raceway surfaces 511 and 521 of 520 and two seal members 540 and 550 for sealing both ends of a bearing inner space formed between an inner ring 510 and an outer ring 520. It's becoming

The large rib 512 guides the large end faces of the tapered rollers 530 and receives axial loads during bearing operation. The small flange 513 has an outer diameter smaller than that of the large flange 512 and receives the small end face of the tapered roller 530 to prevent the tapered roller 530 from falling off from the inner ring 510 .

The outer ring 520 does not have a seal groove, and the metal cores of the seal members 540 and 550 are press-fitted to the inner peripheral end portion of the outer ring 520 .

The sealing members 540, 550 have sealing lips 541, 551 provided as radial lips. A seal sliding surface 514 that slides in the circumferential direction against the seal lip 541 of the seal member 540 on the right side of the drawing is formed in a cylindrical surface that defines the outer diameter of the large flange 512 of the inner ring 510 . A seal sliding surface 515 that slides in the circumferential direction against the seal lip 551 of the seal member 550 on the left side in the drawing is formed in a cylindrical shape that defines the outer diameter of the small flange 513 . Since there is a large difference in diameter between the seal sliding surface 514 formed on the large rib 512 and the seal sliding surface 515 formed on the small rib 513, during operation of the bearing, the seal sliding surface 514 on the right side in the figure is is higher than that of the seal sliding surface 515 on the left side in the drawing. Further, during operation of the bearing, a pumping action is generated in the inner space of the bearing due to the above-mentioned difference in diameter to send the lubricating oil from the left side to the right side in the drawing.

The protrusions 542 and 552 formed on the seal lips 541 and 551 create oil passages and wedge-shaped gaps as shown in FIG. It is possible to put the fluid lubrication state between the moving surfaces 515 respectively.

Here, the number of projections 542 on the sealing member 540 on the right side in the drawing differs from the number of projections 552 on the sealing member 550 on the left side in the drawing. Also, the circumferential pitch angle of the protrusions 542 of the sealing member 540 on the right side of the drawing differs from the circumferential pitch angle of the protrusions 552 of the sealing member 550 on the left side of the drawing. These differences are due to the difference in lubrication conditions between the seal member 540 and the seal sliding surface 514 on the right side of the figure and between the seal member 550 and the seal sliding surface 515 on the left side of the figure due to the above-mentioned peripheral speed difference and pump action. Therefore, it is set for the purpose of equalizing and optimizing the oil films formed between the left and right sides.

FIG. 20 shows the relationship between the distance d between the protrusions 542 and 552 and the theoretical oil film thickness. In this calculation result, the theoretical oil film thickness when the interval d (indicated as "pitch interval" in the figure) between the projections 552 on the side of the small flange 513 (indicated as "small end face side" in the figure) is 6.4 mm. is 4.0 μm, and the theoretical oil film thickness when the interval d (indicated as “pitch interval” in the figure) between the protrusions 542 on the side of the large flange 512 (indicated as “large end face side” in the figure) is 11.5 mm. is 4.0 μm. As is clear from this calculation result, the number of projections and the pitch angle in the circumferential direction, which are parameters related to the distance d between the projections, are varied to allow the distance between the seal lip 541 and the seal sliding surface 514 shown in FIG. An oil film having the same thickness can be formed between 551 and seal sliding surface 515 .

In addition, if the gap between the seal lip 541 and the seal sliding surface 514 and the gap between the seal lip 551 and the seal sliding surface 515 have a spread exceeding 0.07 mm due to the formation of an excessively thick oil film, the grain size will be 50 μm. foreign matter exceeding By determining the oil film thickness that does not cause such a problem, and by varying the number of projections and the circumferential pitch angle, which are parameters related to the distance d between projections, the seal member 540 on the right side of the figure and the seal sliding An oil film with an optimum thickness can be formed between the surfaces 514 and between the seal member 550 and the seal sliding surface 515 on the left side of the drawing.

In this way, the bearing 500 with seal according to the fifth embodiment has a gap between the seal member 540 on the right side in the drawing and the seal sliding surface 514 and between the seal member 540 and the seal sliding surface 514 on the left side in the drawing, as in the third and fourth embodiments. The fluid lubricating state between the seal member 550 and the seal sliding surface 515 can be appropriately set to achieve both low torque and suppression of foreign matter intrusion.

Moreover, conventional tapered roller bearings have a problem of larger torque than ball bearings, and for this reason, in some cases, seal members were not provided. In this case, the pumping action derived from the shape of the lubricant causes the lubricant to flow into the bearing, increasing the stirring resistance of the lubricant. In some cases, hardening treatment such as special heat treatment was required for the inner and outer rings so that the moving surfaces would not be damaged. On the other hand, in the bearing with seal 500, seal members 540 and 550 can be provided in the tapered roller bearing due to the low torque achieved by the fluid lubrication described above. Further, when the seal members 540 and 550 are provided, the inflow of the lubricant into the bearing due to the pumping action is suppressed, and by suppressing the stirring resistance of the lubricant, the torque of the tapered roller bearing itself can be reduced. Furthermore, by preventing foreign matter from entering together with the lubricant by means of the seal members 540, 550, special treatment of the inner and outer rings 510, 520 becomes unnecessary, and cost reduction can be realized.

FIG. 21 shows an example in which the sealed bearing according to the present invention is used as a rolling bearing that supports a rotating portion of a vehicle transmission. The illustrated transmission is a multi-stage transmission that changes the gear ratio step by step, and the sealed bearing B that rotatably supports the rotating parts (for example, the input shaft S1 and the output shaft S2) of the above-described embodiment is used. Equipped with sealed bearings such as The illustrated transmission includes an input shaft S1 to which the rotation of the engine is input, an output shaft S2 provided parallel to the input shaft S1, and a plurality of gear trains G1 to G4 for transmitting rotation from the input shaft S1 to the output shaft S2. and a clutch (not shown) incorporated between each gear train G1-G4 and the input shaft S1 or the output shaft S2. By selectively engaging the clutch, the gear train G1-G4 is used. By switching, the gear ratio of the rotation transmitted from the input shaft S1 to the output shaft S2 is changed. The rotation of the output shaft S2 is output to the output gear G5, and the rotation of the output gear G5 is transmitted to the differential gear and the like. The input shaft S1 and the output shaft S2 are rotatably supported by sealed bearings B, respectively. In addition, this transmission has lubricating oil splashed or sprayed by the splashing of lubricating oil as the gear rotates, or by the injection of lubricating oil from a nozzle (not shown) provided inside the housing H. is applied to the side surface of each bearing B with a seal.

In each of the above-described embodiments, the projections are R-shaped, but the projections have a wedge effect that allows fluid lubrication when the peripheral speed relative to the seal sliding surface is above a certain level. An appropriate shape may be used so as to obtain the desired shape, and for example, chamfering shapes such as R chamfering and C chamfering can be adopted.

Further, in each of the above-described embodiments, an example in which the protrusions are evenly arranged in the circumferential direction has been shown, but it is also possible to arrange them non-uniformly.

Further, in each of the above-described embodiments, the seal member is composed of a metal core and a vulcanized rubber material, but the present invention can also be applied to a seal member formed of a single material. . In this case, it suffices if a required interference can be set for the seal lip, and for example, a rubber material or a resin material can be used as the material of the seal member.

In addition, in the above-described embodiments, inner ring rotating bearings and radial bearings have been exemplified, but the present invention can also be applied to outer ring rotating bearings and thrust bearings.

It should be considered that the embodiments and examples disclosed this time are illustrative in all respects and not restrictive. Therefore, the scope of the present invention is indicated by the scope of the claims rather than the above description, and is intended to include all modifications within the meaning and scope of equivalents of the scope of the claims.

100, 300, 400, 500, B sealed bearings 110, 310, 410, 510 inner rings 111, 121, 314, 411, 421 raceway grooves 112, 312, 313, 414, 415, 514, 515 seal sliding surfaces 120, 320, 420, 520 outer ring 122, 311, 321, 322 seal groove 140 rolling element 150, 330, 340, 440, 450, 540, 550 seal member 151, 202, 331, 341, 441, 451, 541, 551 seal lip 152, 201, 334, 344, 442, 452, 542, 552 Protrusions 153, 203 Tips 154, 204 Surfaces 155, 206, 333, 343 Core metal 156 Vulcanized rubber material 160 Bearing inner space 170 Oil passage 205 Surfaces 315, 316 , 412, 413 Shoulders 332, 342 Outer lip 350 Labyrinth clearance 430 Balls 511, 521 Raceway surface 512 Large rib 513 Small rib 530 Tapered roller S1 Input shaft (rotating part)
S2 Output shaft (rotating part)

Claims (4)

a sealing member that separates the bearing interior space and the exterior;
a seal lip provided on the seal member;
a seal sliding surface that slides in the circumferential direction on the seal lip;
a plurality of projections that are formed on the seal lip and create an oil passage between the seal sliding surface and the seal lip that communicates between the bearing inner space and the outside,
The gap between the projection and the seal sliding surface is formed in a wedge shape with a large gap on the oil passage side and a small gap on the protrusion side,
The protrusions are arranged in the circumferential direction at intervals of 0.3 to 2.6 mm,
The projection extends in a direction orthogonal to the circumferential direction, and has an R shape that gradually approaches the seal sliding surface from both ends of the circumferential width of the projection toward the center of the circumferential width,
The cross-sectional shape of the protrusion is a part of the circumference of a circle with a radius R,
The height of the protrusion is set to 0.07 mm or less,
The R dimension of the projection is 0.15 mm or more and less than 2.0 mm,
The circumferential width of the protrusion is 0.2 mm or more and 1.0 mm or less,
The space between the seal lip and the seal sliding surface is completely separated by an oil film of lubricating oil that is dragged from the oil passage between the projection and the seal sliding surface as the bearing rotates, thereby establishing a fluid lubricating state. A bearing with a seal, wherein the projection is formed in such a manner that it is possible to 2. The sealed bearing according to claim 1 , wherein said protrusions are arranged at uniform intervals along the entire circumference. 3. A bearing with seal according to claim 1, wherein said seal lip is formed of a vulcanized rubber material adhered to at least an inner diameter portion of said core metal. 4. A sealed bearing according to any one of claims 1 to 3 for supporting at least one rotating part in a vehicle transmission or differential.

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