JP2017161069A - Sealed bearing - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce torque and increase speed of a bearing with a seal while preventing foreign matter having a predetermined particle size from entering while enabling high-speed operation of the bearing. SOLUTION: A large number of protrusions 152 are formed on a seal lip 151 of a seal member 150. Between the seal sliding surface 112 that slides in the circumferential direction with respect to these protrusions 152 and the seal lip 151, the bearing internal space 160 and the outside communicate with each other, and foreign matter having a predetermined particle size cannot pass through. An oil passage 170 is created. The protrusions 152 are arranged at intervals of 0.3 to 2.6 mm in the circumferential direction. A wedge effect is generated between the protrusion 152 and the seal sliding surface 112. As the bearing rotates, the lubricating oil in the oil passage 170 promotes the formation of an oil film between the seal lip 151 and the seal sliding surface 112, and the seal lip 151 and the seal sliding surface 112 are in a fluid lubricated state. .. [Selection diagram] Fig. 1

Description

  The present invention relates to a bearing with a seal including a rolling bearing and a seal member.

   For example, foreign substances such as gear wear powder are mixed in transmissions mounted on vehicles such as automobiles and various construction machines. For this reason, the seal member prevents foreign matter from entering the bearing internal space and prevents the rolling bearing from being damaged early.

  A general seal member has a seal lip formed of a rubber-like material or the like. A bearing sliding surface that rotates in the circumferential direction with respect to the seal member, such as a bearing ring and a slinger, is formed with a seal sliding surface that makes the seal lip slide. Since the seal lip and the seal sliding surface are in sliding contact over the entire circumference, the bearing torque increases due to the drag resistance (seal torque) of the seal lip. Moreover, the friction of the sliding contact promotes the temperature rise of the rolling bearing. As the temperature rises, an adsorption action due to a pressure difference between the bearing internal space and the outside is caused, and the friction increases.

  In order to suppress the seal torque of such a contact seal member, the seal sliding surface is subjected to shot peening to obtain a seal sliding surface having minute irregularities with a maximum roughness Ry of 2.5 μm or less, and the lubricating oil stored in the concave portion It has been proposed to promote the formation of an oil film between the seal lip and the seal sliding surface (Patent Document 1).

JP 2007-107588 A

  However, the torque reduction by shot peening as in Patent Document 1 is brought about by reducing the sliding area between the seal lip and the seal sliding surface. Torque was limited.

  Also, in the initial stage of bearing operation, the temperature of the lubricating oil is relatively low, so the viscosity of the lubricating oil is relatively high, and it is in a lubricating condition where it is easy to form an oil film. Then, it becomes the lubrication conditions which an oil film is easy to cut | disconnect. As the bearing is operated at a higher speed, the relative peripheral speed of the seal sliding surface with respect to the seal lip increases, and heat generated due to friction between the seal lip and the seal sliding surface increases. Wear tends to progress. For this reason, there are limits to the high speed operation and allowable rotational speed of the bearing with seal due to lubrication conditions. In an electric vehicle (EV), there is a strong demand for high-speed operation of a bearing with a seal that supports a rotating part of a drive system, but the reduction in torque by shot peening cannot meet the demand.

  If a non-contact seal member is used, it is possible to eliminate the seal torque, but it becomes difficult to manage various errors that can prevent the entry of a foreign substance having a predetermined particle size with respect to the size of the gap between the seal member and the bearing component.

  In view of the above-described background, the problem to be solved by the present invention is to reduce the torque of a bearing with a seal while preventing entry of a foreign substance having a predetermined particle size and being able to cope with high-speed operation of the bearing.

  In order to achieve the above object, the present invention provides a seal member that divides the bearing internal space and the outside, a seal lip provided on the seal member, and a seal slide that slides in the circumferential direction with respect to the seal lip. A plurality of protrusions formed between the seal sliding surface and the seal lip, and an oil passage formed in the seal lip and communicating between the bearing internal space and the outside. The gap between the seal sliding surfaces is formed in a wedge shape that is large on the oil passage side and small on the protrusion side, and the protrusions are arranged at intervals of 0.3 to 2.6 mm in the circumferential direction. This is a bearing.

According to the above configuration, an oil passage is formed between the seal sliding surface and the seal lip, and the lubricating oil in the oil passage is dragged between the seal sliding surface and the seal lip with a wedge effect as the bearing rotates. Oil film formation during this period is promoted. For this reason, the friction coefficient between the seal lip and the seal sliding surface is lowered, and the seal torque is reduced. Furthermore, since the oil permeability between the bearing internal space and the outside is improved by the oil passage, the temperature rise of the rolling bearing is suppressed, and the adsorption action is also prevented.
The smaller the interval between adjacent projections in the circumferential direction, that is, the greater the number of projections, the more the number of times the projections pass per rotation when the seal sliding surface rotates in the circumferential direction relative to the seal lip. Become more. When the protrusions are arranged at intervals of 2.6 mm or less in the circumferential direction, the oil film is maintained over the entire circumference of the seal sliding surface, and the wedge effect between the protrusions is not interrupted. As a result, the bearing operation can be performed in a state where the seal lip and the seal sliding surface are completely separated by the oil film and are not in direct contact with each other (that is, fluid lubrication state). In the fluid lubrication state, the seal torque is substantially brought 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. Accordingly, an allowable speed as a relative peripheral speed of the seal sliding surface with respect to the seal lip is increased, and it is possible to meet a demand for a high-speed operation of a bearing with a seal that cannot be achieved conventionally.
If the interval between the protrusions is set to less than 0.3 mm, it becomes difficult to manufacture a mold for molding the seal lip.
Further, the particle size of the foreign matter that can pass through the oil passage can be determined based on the protrusion height of the protrusion. Accordingly, it is possible to arbitrarily determine the particle diameter to prevent intrusion and prevent foreign matters having the predetermined particle diameter from entering from the oil passage.
As described above, according to the present invention, the adoption of the above configuration can reduce the torque of the sealed bearing while preventing the intrusion of a foreign substance having a predetermined particle diameter and being able to cope with the high speed operation of the bearing.

Sectional drawing which shows the bearing with a seal concerning the 1st example of this invention Enlarged view of the vicinity of the seal lip of the right seal member in FIG. Fig. 2 is an enlarged sectional view taken along line III-III in Fig. 2. The partial front view which shows the seal lip which concerns on 1st Example from an axial direction The figure which shows the particle size distribution and number of the foreign material contained in the lubricating oil in a transmission (AT / MT) of a vehicle Pie chart showing the ratio of particle size distribution in FIG. The figure which shows the particle size distribution and number of the foreign material contained in the lubricating oil in the transmission (CVT) of a vehicle Pie chart showing the ratio of particle size distribution in FIG. Graph showing test results of measuring bearing torque for each R dimension of protrusion Lubrication region diagram showing the fluid lubrication mode in the first embodiment The figure which shows the relationship between the space | interval between protrusions in 1st Example, a theoretical oil film thickness, and a bearing rotational torque. The figure which shows the relationship between R dimension of protrusion in 1st Example, and theoretical oil film thickness The partial perspective view which shows the seal lip which concerns on 2nd Example of this invention (A) is a schematic diagram showing the state of vulcanization molding of the seal lip of FIG. 13, (b) is a schematic diagram of the seal lip molded in (a). (A) is a schematic diagram showing the state of vulcanization molding of a virtual model seal lip, (b) is a schematic diagram of the seal lip molded in (a). Sectional drawing which shows the bearing with a seal concerning the 3rd example of this invention FIG. 15 is an enlarged view of the vicinity of the seal lip of the right seal member. Sectional drawing which shows the bearing with a seal concerning the 4th example of this invention Sectional drawing which shows the bearing with a seal concerning 5th Example of this invention The figure which shows the relationship between the space | interval between protrusions in 5th Example, and a theoretical oil film thickness. Sectional drawing which shows an example of a transmission provided with the bearing with a seal concerning this invention

A preferred embodiment of the present invention will be described.
In the first embodiment, the protrusion extends in a direction orthogonal to the circumferential direction, and the protrusion 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, since the protrusion extends in a direction perpendicular to the circumferential direction that is the sliding direction with the seal sliding surface, and has an R shape that reduces the area in which the protrusion can come into sliding contact, A region where the protrusion and the sliding surface of the seal are in sliding contact can be linear. In addition, with such an R shape, the wedge angle of the aforementioned wedge-shaped gap gradually decreases from the wide side toward the narrow side, so that the wedge effect is effectively generated and the hydraulic pressure in the linear region is increased. Therefore, the lubrication state between the protrusion and the seal sliding surface can be easily changed to the fluid lubrication state. Further, even when the protrusion is rubbed against the seal sliding surface when the seal member is attached, there is no concern that the R-shaped protrusion will bend in the circumferential direction, and there is no fear that the sealing torque reduction performance will be impaired during attachment.

  In the first embodiment, the height of the protrusion is set to 0.07 mm or less. Foreign matter having a particle size exceeding 0.05 mm (50 μm) adversely affects the bearing life. It can be seen that if the height of the protrusion is set to 0.07 mm or less, a narrow gap (including an oil passage) that cannot easily pass such foreign matter can be generated between the seal lip and the seal sliding surface. It was.

  In the first embodiment, the R dimension of the protrusion is 0.15 mm or more and less than 2.0 mm, and the circumferential width of the protrusion is 0.2 mm or more and 1.0 mm or less. The larger the R dimension of the protrusion, the easier 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 protrusion and the seal sliding surface. When the height of the protrusion 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 in order to prevent the manufacture of the seal lip mold. 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 uniform intervals over the entire circumference in the circumferential direction. According to the second embodiment, oil film formation can be promoted uniformly 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. 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 cage 130, and a bearing inner space formed between the inner ring 110 and the outer ring 120. The bearing 100 with a seal is provided with two seal members 150 that seal both ends. Hereinafter, the direction along the bearing central axis of the sealed bearing 100 is referred to as “axial direction”. The direction orthogonal to the axial direction is called “radial direction”. The circumferential direction around the bearing central axis is called “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 revolve while being interposed between the inner ring 110 and the outer ring 120 in the bearing inner space 160. Lubricating oil is supplied to the bearing internal space 160 by appropriate means such as grease and 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 vehicle transmission or a differential rotating portion. The outer ring 120 is attached to a member that applies a load from the rotating shaft, such as a housing or a gear.

  This sealed bearing 100 is a deep groove ball bearing. A ball is used as the rolling element 140. The inner ring 110 and the outer ring 120 each have a raceway groove 111, 121 that corresponds to about one third of the circumference of the rolling element 140 in a transverse cross section (corresponding to the cross section shown in the figure) and is uninterrupted over the entire circumference in the circumferential direction. .

  A seal groove 122 for holding the seal member 150 is formed at the inner peripheral end 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 internal space 160 from the outside. On the external side with the seal member 150 as a boundary, there are foreign matters according to the assembly destination of the bearing 100 with seal, such as gear wear powder, clutch wear powder, and fine crushed stone. Such powdery foreign matter can reach the vicinity of the seal member 150 by 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 seal lip 151 protruding like a tongue on the inner peripheral side. A seal sliding surface 112 that slides in the circumferential direction with respect to the seal lip 151 is formed on the outer periphery of the inner ring 110. The seal sliding surface 112 has a cylindrical surface shape over the entire circumference in the circumferential direction. The seal lip 151 is a radial lip. Here, the radial lip is a seal lip having a sealing action with a seal sliding surface along the axial direction or a seal sliding surface with an acute angle gradient of 45 ° or less with respect to the axial direction. This means that there is a radial allowance between the moving surface.

  FIG. 2 is an enlarged view of the vicinity of the seal lip 151 of the seal member 150 of FIG. 3 is a cross-sectional view taken along line III-III in FIG. This cross section is designed so that the gap (including the oil passage 170) in the direction orthogonal to the seal sliding surface 112 between the seal lip 151 and the seal sliding surface 112 is perpendicular to the seal sliding surface 112 by design. It shows the situation in the narrowest place. Moreover, the external appearance when the seal lip 151 is viewed in the axial direction from the bearing internal space side is shown in FIG. FIG. 4 depicts the outer shape of the seal lip 151 in the natural and natural state of the seal member 150 shown in FIG. Here, the natural state means a state where no external force is applied to the seal member in a single state, that is, the state where the seal member is not deformed by the external force (hereinafter, this state is simply referred to as “natural”). Call it a state. ")

  As shown in FIGS. 2 to 4, the seal lip 151 has a protrusion 152 having a protrusion height in a direction perpendicular to the seal sliding surface 112, that is, a normal direction perpendicular to a tangent line contacting the seal sliding surface 112. Have. Since the seal sliding surface 112 has a cylindrical surface centered on the bearing central axis, the direction orthogonal thereto corresponds to the radial direction.

  A tightening margin in the radial direction is set between the seal lip 151 and the seal sliding surface 112. Due to this tightening allowance, the seal lip 151 pressed in the radial direction against the seal sliding surface 112 is deformed into a rubber-like elastic shape that is bent to the outside, and generates a tightening force of the seal lip 151. A mounting error, a manufacturing error, and the like of the seal member 150 are absorbed by a change in the degree of bending 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 a natural state.

  As shown in FIGS. 2 to 4, the protrusion 152 extends 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 a constant interval d in the circumferential direction. Considering the appearance of the seal lip 151 viewed from the axial direction, the plurality of protrusions 152 appear radially in a circumferential direction with a constant pitch angle θ corresponding to the distance d. In addition, the radiation center is on the center axis (coincident with the bearing center axis) of the seal member 150 (not shown).

  The distance d between the protrusions 152 adjacent to each other in the circumferential direction and the circumferential width w of the protrusions 152 are coupled with the fact that the protrusions 152 arranged radially are present near the tip 153 of the seal lip 151. The lip 151 can come into sliding contact with the seal sliding surface 112 only on each protrusion 152, and the oil passage 170 is always generated between the protrusions 152. That is, when the seal member 150 is attached, the protrusion 152 that contacts the seal sliding surface 112 is stretched against the tight force of the seal lip 151, so that the bearing inner space 160 and the outer side are formed on both sides in the circumferential direction with the protrusion 152 as a boundary. An oil passage 170 is formed in communication therewith. The lubricating oil reaches the bearing internal space 160 from the outside through the oil passage 170. Lubricating oil that has entered the bearing internal space 160 and base oil in the case where grease is sealed are directed from the bearing internal space 160 to the outside through the oil passage 170.

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

  The wear powder that causes early breakage of the rolling bearing is a foreign matter having a particle diameter exceeding 50 μm. It has been found that if the protrusion height h of the protrusion 152 is set to 0.07 mm or less, an oil passage 170 through which such wear powder cannot pass can be generated. 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 tolerances. On the other hand, by setting the protrusion height h of the protrusion 152 to 0.07 mm, the oil permeability 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 in which the side close to the oil passage 170 in the circumferential direction is large and the side close to the protrusion 152 in the circumferential direction is small. The surface 154 of the protrusion 152 is a curved surface corresponding to the above-mentioned 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 an arrow A in the figure), the oil passage. The lubricating oil in 170 (indicated by a dot pattern in the drawing) is dragged between the seal sliding surface 112 and the projection 152 of the seal lip 151 as the seal sliding surface 112 rotates, and an oil film is formed between them. Promote. For this reason, the coefficient of friction (μ) between the seal lip 151 and the seal sliding surface 112 is lowered, and the seal torque is reduced. Further, oil permeability between the bearing internal space 160 and the outside is improved by the oil passage 170. For this reason, the temperature rise of the bearing 100 with a seal | sticker is suppressed and by extension, the adsorption | suction effect | action of the seal lip 151 is also prevented.

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

  As described above, in the first embodiment, the protrusion 152 can be formed on the seal lip 151 at the time of vulcanization molding of the seal lip 151, and the seal sliding surface 112 has a cylindrical surface shape that is easy to process, It is easy to directly form the raceway with the same cross-sectional shape over the entire circumference, such as a groove shape.

  Due to the tight force of the seal lip 151 and the oil pressure of the lubricating oil, the solid protrusion 152 is not substantially deformed (deformation that affects the lubricating performance between the protrusion 152 and the seal sliding surface 112). Yes. Therefore, the shape of the protrusion 152 during the bearing operation may be considered to be the same as the shape transferred when the seal lip 151 is vulcanized.

  This sealed bearing 100 is assumed to be used for supporting a rotating part in a transmission of a vehicle. Oil supply to a bearing with a seal present in a transmission of a vehicle is generally performed by an appropriate method such as splashing, an oil bath, nozzle injection, or the like. Therefore, lubricating oil exists around the seal lip of the seal member fixed to the inner ring or the outer ring of the bearing with seal. The lubricating oil to be supplied is commonly used in other lubricating parts such as gears existing in the transmission. The lubricating oil is circulated by an oil pump, and is filtered by an oil filter provided in the circulation path.

  The inventor of the present application collects the lubricating oil actually used in the market according to the mileage of the vehicle, and investigated the number of foreign matters mixed in the used lubricating oil, the particle size distribution of the foreign matters, and the foreign material. . FIG. 5 shows the number of foreign substances and the particle size distribution of the lubricating oil collected from the eight vehicles of the automatic transmission (AT) or the manual transmission (MT). The vertical axis in FIG. 5 is a logarithmic scale, the horizontal axis is the distance traveled by the vehicle, and the vertical axis is the number of foreign substances (fine particles) per 100 ml. The foreign objects to be counted were those having a particle size of 5 μm or more. Counting was performed for each particle size category. The particle size ranges from 5 μm to less than 15 μm, particle size from 15 μm to less than 25 μm, particle size from 20 μm to less than 50 μm, particle size from 50 μm to less than 100 μm, and particle size of 100 μm or more. In this measurement, a fine particle substance amount method was used with a measuring instrument of model number 8000A manufactured by Hyac Royco. The particle size distribution of FIG. 5 is shown as a pie chart in FIG.

  FIG. 7 shows the number of foreign matters and the particle size distribution of the lubricating oil collected from 10 vehicles of continuously variable transmissions (CVT), as in FIG. The particle size distribution of FIG. 7 is shown as a pie chart in FIG. Vehicle manufacturers, vehicle types, and travel distances that are subject to collection vary, but as is clear from a comparison between FIGS. 5, 6, 7, and 8, AT / MT that uses a lot of gear is better than CVT There was also a tendency for both the particle size of foreign matter and the number of foreign matters to be large. Regardless of the type of transmission, the particle size distribution accounted for 99.9% or more when the particle size was 50 μm or less. The number of foreign matters having a particle size exceeding 50 μm was less than 1000 in the case of AT / MT and less than 200 in the case of CVT even when the travel distance was increased. This indicates that in recent years, the performance of the oil filter has been improved, and foreign matters in the lubricating oil have become finer (that is, foreign matters having a large particle diameter are removed by the oil filter).

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

  Based on the above results, when lubricating oil filtered by an oil filter is supplied to a bearing with a seal that is used to support a rotating part of a drive system such as a transmission of a vehicle or a differential gear, a large particle size exceeding 50 μm is required. As long as the seal member prevents foreign matter from entering the inside of the bearing, it can be said that there is no problem in bearing life even if foreign matter having a particle size of 50 μm or less contained in the lubricating oil is allowed to enter the bearing. If this is allowed, fluid flow between the seal lip and the seal sliding surface should be ensured, and the lubrication between the seal lip and the seal sliding surface should be combined with the aforementioned wedge effect. It is feasible to get into 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 a value at the highest position in the range where the seal sliding surface 112 can be in sliding contact with the design. This position is also where the tightening margin set between each protrusion 152 and the seal sliding surface 112 is maximized. Since the deformation amount of the protrusion 152 during the operation of the bearing is negligible, the gap (including the oil passage 170) in the direction perpendicular to the seal sliding surface 112 between the seal lip 151 and the seal sliding surface 112 is the seal sliding. The narrowest portion in the direction orthogonal to the surface 112 has a width corresponding to the height h of the protrusion 152 and does not substantially exceed 0.07 mm. For this reason, even if a foreign matter having a particle size of 50 μm or more is included in the external lubricating oil, it is considered that the foreign matter hardly passes 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 protrusion 152 in the radial direction. For this reason, the region in which the protrusion 152 and the seal sliding surface 112 can come into sliding contact exists linearly on a virtual axial plane Pax passing through the center of the circumferential width of the protrusion 152. The center of curvature 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 figure, the lubricating oil in the oil passage 170 is dragged into the wedge-shaped gap between the protrusion 152 and the seal sliding surface 112. Is included. The wedge angle in the above-mentioned wedge-shaped gap gradually decreases from the wide oil passage 170 where the drawn lubricating oil is present toward the narrow side, so that the protrusion 152 and the seal sliding surface 112 are in sliding contact with each other. The closer to the linear region to be obtained (on the virtual axial plane Pax), the stronger the wedge effect occurs. Accordingly, the oil pressure of the oil film in the linear region can be increased more effectively, the protrusion 152 can be completely separated from the seal sliding surface 112, and the oil film in the linear region can be generated thickly. It becomes easy to make the lubrication state between 152 and the seal sliding surface 112 into a fluid lubrication state.

  Here, if there is an oil film that completely separates the protrusion 152 and the seal sliding surface 112, a fluid lubrication state in which the seal sliding surface 112 slides in a state where it does not directly contact the protrusion 152 is achieved. By maintaining such an oil film between each protrusion 152 and the seal sliding surface 112, the space between the seal lip 151 and the seal sliding surface 112 can be in a fluid lubrication state.

  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 tightening margin between the seal lip 151 and the seal sliding surface 112 as weak as possible. For this reason, the waist | hip | lumbar part which gives the bending deformation to the exterior side among the seal lips 151 is formed as thinly as possible.

  Further, the smaller the maximum height roughness Rz is, the smaller the thickness of the oil film necessary for achieving a fluid lubrication state. For this reason, shot peening processing is not performed on the seal sliding surface 112, and the maximum height roughness Rz of the seal sliding surface 112 is set to less than 1 μm. Here, the maximum height roughness Rz refers to the maximum height roughness defined in JIS standard B0601: 2013.

  Further, when the distance d between the protrusions 152 adjacent in the circumferential direction is smaller, that is, the number of the protrusions 152 is larger, the seal sliding surface 112 rotates relative to the seal lip 151 in the circumferential direction. Since the number of times that the projection 152 is passed is increased, the oil film is maintained over the entire circumference of the seal sliding surface 112, and the wedge effect between the projections 152 is easily generated without interruption. It becomes easy to maintain the fluid lubrication state.

  In addition, the wedge effect is more likely to occur when the R dimension of the protrusion 152 (the radius of curvature at the surface 154 of the protrusion 152) is larger.

  Further, when a bearing torque measurement test was performed for each R dimension of the protrusion 152 (see FIG. 9), when the R dimension of the protrusion 152 is 1.5 mm, it is compared with when the R dimension is 0.2 mm or 0.8 mm. Thus, the torque at a temperature of 80 ° C. close to the actual usage environment of the bearing 100 became a minimum value. When the R dimension of the protrusion 152 was 1.8 mm, a result equivalent to 1.5 mm was obtained. In order to easily generate a wedge effect and realize low torque, the R dimension of the protrusion 152 is preferably a value of 1.5 mm or more and 2.0 mm or less in consideration of tolerances. In FIG. 9, for reference, the torque of a bearing used in an oil bath lubrication environment without a seal and used in an oil bath lubrication environment in which the center of the rolling element of the bearing (the bottom end of the bearing) is the oil level is theoretically shown. The calculated values are also shown. By setting the R dimension of the protrusion 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 Further, low torque equivalent to that of the bearing (calculated value shown in FIG. 9) considered to have a small stirring resistance can be realized.

  The inventor of the present application assumes that the height h of the protrusion 152 is 0.07 mm, the distance d between the protrusions 152 is 0.3 mm or more and 2.6 mm or less, and the circumferential width w of the protrusion 152 is 0.2 mm or more and 1.0 mm. Hereinafter, the theoretical oil film thickness between the protrusion 152 and the seal sliding surface 112 was calculated in the range where the R dimension of the protrusion 152 was 0.15 mm or more and less than 2.0 mm. The calculation was performed at a speed (peripheral speed) at which the seal sliding surface 112 rotates in the circumferential direction relative to the seal lip 151 in a range of 0.02 to 20.2 m / s. In addition, the calculation was performed for the oil temperature of 30 ° C. and the oil temperature of 120 ° C., assuming CVTF for lubricating the CVT pulley and belt as the lubricating oil. As a result, under these usage conditions, in the lubrication region diagram in the case of line contact based on the viscosity parameter gv, which is a dimensionless number determined by Greenwood-Johnson, and the elastic parameter ge, in the calculation, the isoviscous-rigid region ( It was found that the oil was distributed in a lubrication mode in either the (R-I mode) or isoviscosity-elastic body region (EI mode, soft EHL), that is, a fluid lubrication state. FIG. 10 shows the distribution of calculation results in the above-mentioned lubrication region diagram.

  Further, the inventor of the present application calculated the relationship between the distance d between the protrusions 152 shown in FIGS. 3 and 4 and the theoretical oil film thickness between the protrusions 152 and the seal sliding surface 112 by calculation. In addition, the inventor of the present application obtained a 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 by calculation. 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. Further, the inventor of the present application examined the relationship between the distance d between the protrusions 152 and the rotational torque of the bearing 100 with seal. FIG. 11 shows the relationship among the distance d between the protrusions 152, the theoretical oil film thickness, and the bearing rotational torque. FIG. 12 shows the relationship between the R dimension of the protrusion 152 and the theoretical oil film thickness. If the oil film thickness between the protrusion 152 and the seal sliding surface 112 is too thin, the friction coefficient μ increases. Conversely, if the oil film thickness is too thick, there is a possibility that the effect of suppressing the intrusion of foreign matter may be deteriorated. An optimum oil film thickness may be set on the assumption that the oil film thickness exceeds Rz.

  From FIG. 11, when the distance d between the protrusions 152 is 2.6 mm, an oil film of about 3 μm (maximum height roughness Rz of less than 1 μm is allowed between the protrusion 152 and the seal sliding surface 112 in calculation. It is found that the oil film tends to be thicker when it is smaller than 2.6 mm, and the bearing rotational torque tends to decrease (ie, the seal torque tends to decrease) when it is 2.6 mm or less. I understand that. 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 is difficult to mold the seal lip 151 with a mold. Therefore, the distance d is preferably set to 0.3 mm or more.

  From FIG. 12, it is understood that when the R dimension of the protrusion 152 is 0.15 mm, an oil film of 3 μm or more is formed in 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 protrusion 152 is set to 0.07 mm, the circumferential width w of the protrusion 152 depends on the R dimension, so the circumferential width w of the protrusion 152 is set to 0.2 mm or more and 1.0 mm or less. It is preferable to do.

  Based on the calculation results so far, in the first embodiment, the height h of the protrusions 152 as shown in FIGS. 2 to 4 is set to 0.07 mm, and the distance d between the protrusions 152 is 0.3 mm or more. It is set to 2.6 mm or less, the circumferential width w of the protrusion 152 is set to 0.2 mm to 1.0 mm, and the R dimension of the protrusion 152 is set to a range of 0.15 mm to less than 2.0 mm. . That is, the protrusion 152 is provided on the seal lip 151 in such a manner that the space between the seal lip 151 and the seal sliding surface 112 can be fluidly lubricated at a peripheral speed of 0.2 m / s or more of the seal sliding surface 112 described above. Is formed. When the peripheral speed is less than 0.2 m / s, the boundary 152 is in a boundary lubrication state between the protrusion 152 and the seal sliding surface 112. Since the circumferential speed of 0.2 m / s is a speed that can be reached quickly from the stop in the transmission of the vehicle, the gap between the seal lip 151 and the seal sliding surface 112 may be brought into a fluid lubrication state during substantially the entire bearing operation time. it can.

  As described above, the bearing 100 with the seal according to the first embodiment has the seal lip 151 and the seal slide while preventing the foreign material having a particle size that adversely affects the bearing life from entering the bearing internal space by the seal member 150. The friction coefficient μ of sliding between the moving surfaces 112 can be reduced to the limit by fluid lubrication, and as a result, the seal torque can be significantly reduced to significantly reduce the bearing rotational torque (see FIGS. 1 and 3). ).

  Furthermore, the bearing 100 with the seal according to the first embodiment is a peripheral speed of the seal sliding surface that would cause problems of heat generation due to wear of the seal lip or sliding between the seal lip and the seal sliding surface. When operating at (for example, 30 m / s or more), the seal lip 151 and the seal sliding surface 112 can be brought into a fluid lubrication state without direct contact, so that the wear of the seal lip 151 is substantially eliminated. At the same time, the aforementioned heat generation can also be suppressed. For this reason, the bearing 100 with a seal according to the first embodiment can meet the demand for high-speed operation of the bearing with a seal that could not be achieved conventionally.

  Further, in the bearing 100 with seal according to the first embodiment, since the protrusion 152 is formed in an R shape, the protrusion 152 may be rubbed against the seal sliding surface 112 when the seal member 150 is attached to the outer ring 120. Further, there is no concern that the protrusion 152 bends in the circumferential direction, and there is no possibility of impairing the sealing torque reduction performance at the time of attachment. For example, when the protrusions are pointed, it is not possible to know which side of the circumferential direction the large number of protrusions rubbed against the seal sliding surface when the seal member is attached. It is extremely difficult to mount the projections so that all the projections bend toward the appropriate wedge-shaped gap. A wedge effect cannot be satisfactorily obtained at a protrusion bent in an inappropriate direction, and the sealing torque reduction performance is impaired.

  A second embodiment will be described with reference to FIGS. In the second embodiment, only the protrusion shape is changed from the first embodiment. As shown in FIG. 13, the protrusion 201 according to the second embodiment is directed toward the tip 203 of the seal lip 202. The shape becomes gradually lower. FIG. 13 is an enlarged perspective view of the vicinity of the protrusion 201 of the seal lip 202 in a natural state. The R dimension of the protrusion 202 (the radius of curvature at the surface 204 of the protrusion 201) and the center of curvature are gradually lowered toward the tip 203 of the seal lip 202 in order to make the protrusion 201 gradually lower toward the tip 203 of the seal lip 202. And the center of curvature is moved to the outside.

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

  FIG. 14 shows how the seal lip 202 is vulcanized. FIG. 14 is drawn schematically for easy understanding, and the shape of the seal lip 202 is also roughly shown. Vulcanization of the seal lip 202 is performed by vulcanizing and molding a rubber sheet on the core metal 206. At this time, a rubber sheet is sandwiched between the upper mold Mp1 and the lower mold Mp2, and a 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 combined corresponds to the axial direction. Therefore, the tip 203 that defines the inner diameter of the seal lip 202 in the natural state is a boundary line between the upper surface portion of the seal lip 202 that contacts the transfer surface of the upper mold Mp1 and the lower surface portion of the seal lip 202 that contacts the transfer surface of the lower mold Mp2. Therefore, it is located on the parting line Pl that is the joining portion of the upper mold Mp1 and the lower mold Mp2.

  Now, assuming a model in which a protrusion extends to the tip of the seal lip, FIG. 15 is obtained. In this virtual model, since the unevenness for forming the projection 201 ′ on the tip 203 ′ of the seal lip 202 ′ is present on the parting line Pl, a burr 207 as shown in the figure is likely to occur after vulcanization. When the burr 207 is generated, the burr 207 is separated from the seal lip 202 ′ during the operation of the bearing, which causes clogging of the oil filter and the lubricating oil circulation path.

  On the other hand, as shown in FIG. 14, when the seal lip 202 has a flat surface 205 between the protrusion 201 and the tip 203 of the seal lip 202, the unevenness for forming the protrusion 201 on the parting line Pl. No burr 207 as shown in FIG. 15 is generated. 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 on 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 when the protrusion has a height on the tip of the seal lip, if the protrusion is gradually lowered toward the tip of the seal lip, the uneven shape for forming the protrusion on the parting line is gentle. Therefore, it is possible to make it difficult to generate burrs on the tip of the seal lip.

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

  The seal members 330 and 340 include seal lips 331 and 341 provided as axial lips, and outer lips 332 and 342 located on the outer side of the seal lips 331 and 341. The seal lips 331 and 341 and the outer lips 332 and 342 are branched from the waist part attached to the core bars 333 and 343.

  Here, the axial lip is a seal lip that exhibits a sealing action with a seal sliding surface along the radial direction or a seal sliding surface with an acute angle gradient of less than 45 ° with respect to the radial direction. This means that there is an axial allowance between the moving surface.

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

  Of the pair of shoulder portions 315 and 316 formed on both sides of the raceway groove 314 of the inner ring 310, the shoulder portion 315 on the load side (right side in the figure) receiving the axial load is on the opposite non-load side (left side in the figure). It is formed higher than the shoulder 316. The non-load side shoulder 316 has a shoulder height equivalent to a deep groove ball bearing. Therefore, the bearing 300 with a seal has a load capacity of an axial load superior to that of the deep groove ball bearing, while exhibiting good low torque characteristics of the 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 difference in diameter between the outer diameter of the shoulder 315 on the load side and the outer diameter of the shoulder 316 on the non-load side, the diameter of the seal sliding surface 312 formed on the shoulder 315 on the load side, The diameter of the seal sliding surface 313 formed on the shoulder 316 on the non-load side is different. That is, the diameter of the seal sliding surface 312 on the right side in the drawing is larger than the diameter of the seal sliding surface 313 on the left side in the drawing. For this reason, during the bearing operation, the peripheral speed of the seal sliding surface 312 on the right side in the drawing becomes higher than the seal sliding surface 313 on the left side in the drawing.

  Further, during the operation of the bearing, in the bearing internal space, a pumping action for sending the lubricating oil from the left side to the right side in the drawing occurs due to the diameter difference between the shoulder portions 315 and 316 described above.

  The protrusions 334 and 344 of the seal lips 331 and 341 are formed in a direction along the radial direction during vulcanization molding. In the drawing, the shape of the seal lips 331 and 341 corresponding to the natural state is drawn in order to show the interference between the seal sliding surfaces 312 and 313 and the protrusions 334 and 344. When the protrusions 334 and 344 are pressed against the seal sliding surfaces 312 and 313 from the axial direction, the seal lips 331 and 341 are inclined so as to substantially follow the seal sliding surfaces 312 and 313, and the protrusions 334 and 344 and the seal sliding surfaces The oil passages and the wedge-shaped gaps as described above are generated between 312 and 313 (see FIG. 3).

  The outer lips 332 and 342 shown in FIGS. 16 and 17 form a labyrinth clearance 350 between the outer groove walls of the seal groove 311. For this reason, foreign matters having a particle size of more than 50 μm cannot easily enter the seal groove 311 from the outside.

  The number of protrusions 334 on the right seal member 330 in the figure is different from the number of protrusions 344 on the left seal member 340 in the figure. Further, the circumferential pitch angle of the protrusions 334 in the right seal member 330 in the figure is different from the circumferential pitch angle of the protrusions 344 in the left seal member 340 in the figure. These differences are the difference in the lubrication conditions due to the above-described difference in peripheral speed and pump action between the seal member 330 and the seal sliding surface 312 on the right side in the drawing and the seal member 340 and the seal sliding surface 313 on the left side in the drawing. Therefore, it is necessary to make the oil film formed between these left and right parts equal and to optimize the formation so that the entry of foreign matter having a particle diameter of more than 50 μm is not likely to occur due to the formation of an oil film that is too thick. It is set as a purpose. Therefore, in the bearing 300 with the seal according to the third embodiment, the seal member 330 and the seal sliding surface 312 on the right side in the drawing and the seal member 340 and the seal sliding surface 313 on the left side in the drawing are appropriately disposed. As the fluid lubrication state, it is possible to achieve both a reduction in torque and suppression of foreign matter intrusion.

  Further, the bearing 300 with a seal according to the third embodiment makes it difficult for foreign matter to reach the seal lips 331 and 341 due to the formation of the labyrinth clearance 350. Can be suppressed. In general, the seal lip provided as an axial lip has a larger axial movement during the bearing operation than the seal lip provided as a radial lip, and the gap between the seal lip and the corresponding seal sliding surface during the maximum movement is large. There may be large gaps. For this reason, providing a seal lip as an axial lip is disadvantageous for foreign matter intrusion. In the third embodiment, the disadvantages of the seal lips 331 and 341, which are such axial lips, can be compensated by the sealing effect by the labyrinth clearance 350, so that the first to second embodiments provided as radial lips are used. On the other hand, there is no concern that the bearing life will be inferior.

  A fourth embodiment will be described with reference to FIG. A bearing 400 with a seal according to the fourth embodiment is formed between an inner ring 410, an outer ring 420, a plurality of balls 430 interposed between the race grooves 411 and 421 of the inner ring 410 and the outer ring 420, and the inner ring 410 and the outer ring 420. Two seal members 440 and 450 for sealing both ends of the bearing inner space formed, and of the pair of shoulder portions 412 and 413 formed on both sides of the raceway groove 411 of the inner ring 410 on the load side that receives an axial load. The shoulder 412 is the same as that of the third embodiment in that the shoulder 412 is formed higher than the opposite shoulder 413 on the non-load side, and the seal members 440 and 450 are provided as radial lips. The second embodiment is common to the first embodiment in that the seal lips 441 and 451 are provided.

  The inner ring 410 does not have a seal groove, and has a cylindrical sliding surface 414 that defines the outer diameter of the load-side shoulder 412 and a cylindrical surface that defines the outer diameter of the non-load-side shoulder 413. And a seal sliding surface 415. The difference in diameter between the seal sliding surface 414 on the right side in the drawing and the seal sliding surface 415 on the left side in the drawing is about the same as in the third embodiment, and the difference in peripheral speed during the bearing operation is also about the same. . For this reason, the difference in the number of protrusions 442 and 452 and the circumferential pitch angle between the sealing member 440 on the right side in the drawing and the sealing member 450 on the left side in the drawing is about the same as in the third embodiment. Thereby, also in the bearing with seal 400 according to the fourth embodiment, the oil film is optimized and equalized between the seal member 440 and the seal sliding surface 414 and between the seal member 450 and the seal sliding surface 415. It achieves both lowering the torque to form a fluid lubrication state and suppressing entry of foreign matter.

  A fifth embodiment will be described with reference to FIG. In the fifth embodiment, the present invention is applied 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 brim 512 and a small brim 513, an outer ring 520 having a raceway surface 521, an inner ring 510 and an outer ring. A tapered roller bearing including a plurality of tapered rollers 530 interposed between the raceway surfaces 511 and 521 of the 520 and two seal members 540 and 550 for sealing both ends of a bearing inner space formed between the inner ring 510 and the outer ring 520; It has become.

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

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

  The seal members 540 and 550 have seal lips 541 and 551 provided as radial lips. A seal sliding surface 514 that slides in the circumferential direction with respect to the seal lip 541 of the seal member 540 on the right side in the drawing is formed in a cylindrical surface that defines the outer diameter of the large collar 512 of the inner ring 510. The seal sliding surface 515 that slides in the circumferential direction with respect to the seal lip 551 of the seal member 550 on the left side in the drawing is formed in a cylindrical surface that defines the outer diameter of the small brim 513. Since there is a large diameter difference between the seal sliding surface 514 formed on the large brim 512 and the seal sliding surface 515 formed on the small brim 513, the seal sliding surface 514 on the right side in the figure during the bearing operation. Is faster than the seal sliding surface 515 on the left side in the figure. Further, during the operation of the bearing, in the inner space of the bearing, a pumping action for sending the lubricating oil from the left side to the right side in the figure occurs due to the above-described diameter difference.

  The oil passage and the wedge-shaped gap as shown in FIG. 3 are generated by the protrusions 542 and 552 formed on the seal lips 541 and 551. Each of the moving surfaces 515 can be in a fluid lubrication state.

  Here, the number of protrusions 542 on the right seal member 540 in the drawing is different from the number of protrusions 552 on the left seal member 550 in the drawing. Further, the circumferential pitch angle of the protrusions 542 in the right seal member 540 in the figure is different from the circumferential pitch angle of the protrusions 552 in the left seal member 550 in the figure. The difference between the sealing member 540 and the seal sliding surface 514 on the right side in the drawing and the difference in the lubrication conditions due to the above-described peripheral speed difference and pump action between the sealing member 550 and the seal sliding surface 515 on the left side in the drawing. Therefore, it is set for the purpose of equalizing and optimizing the oil film formed between these left and right parts.

  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 distance d (shown as “pitch interval” in the figure) between the protrusions 552 on the small brim 513 side (indicated as “small end face side” in the figure) is 6.4 mm. The theoretical oil film thickness when the gap d (shown as “pitch interval” in the figure) between the protrusions 542 is 11.5 mm on the large brim 512 side (shown as “large end face side” in the figure). Becomes 4.0 μm. As is apparent from this calculation result, the difference between the number of protrusions and the circumferential pitch angle, which are parameters related to the distance d between the protrusions, allows the seal lip 541 and the seal sliding surface 514 shown in FIG. An oil film having an equivalent thickness can be formed between 551 and the 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 for the formation of an excessively thick oil film, the particle diameter is 50 μm. It is easy for foreign matter to enter. The oil film thickness is determined so that such a problem does not occur, and the number of protrusions and the circumferential pitch angle, which are parameters related to the distance d between the protrusions, are made different, so that the seal member 540 and the seal slide on the right side in the figure are changed. An oil film having 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 in the drawing.

  As described above, the sealed bearing 500 according to the fifth example is similar to the third example and the fourth example, between the seal member 540 on the right side in the figure and the seal sliding surface 514, and on the left side in the figure. Each of the seal member 550 and the seal sliding surface 515 can be appropriately fluid lubricated to achieve both a reduction in torque and suppression of foreign matter intrusion.

  Conventionally, a tapered roller bearing has a problem that the torque is larger than that of a ball bearing, and therefore, a seal member may not be provided. In this case, the lubricant flows into the bearing due to the pumping action derived from the shape and the stirring resistance of the lubricant increases, and foreign matter also flows along with the lubricant. In some cases, a special heat treatment or the like is required for the inner and outer rings so that the moving surface is not damaged. In contrast, the sealed bearing 500 can be provided with the seal members 540 and 550 in the tapered roller bearing due to the reduction in torque achieved by the fluid lubrication described above. Further, when the seal members 540 and 550 are provided, the torque of the tapered roller bearing itself can be reduced by suppressing the inflow of the lubricant into the bearing due to the pumping action and suppressing the stirring resistance of the lubricant. Further, by preventing foreign matter flowing in together with the lubricant by the seal members 540 and 550, special processing for the inner and outer rings 510 and 520 is not required, and cost can be reduced.

  FIG. 21 shows an example in which a bearing with a seal according to the present invention is used as a rolling bearing that supports a rotating portion of a transmission of a vehicle. The illustrated transmission is a multi-stage transmission that changes the gear ratio stepwise, and the above-described embodiment is used as a bearing B with a seal that rotatably supports a rotating portion (for example, the input shaft S1 and the output shaft S2). Such a bearing with a seal is provided. The illustrated transmission includes an input shaft S1 to which engine rotation is input, an output shaft S2 provided in parallel with the input shaft S1, and a plurality of gear trains G1 to G4 that transmit the rotation from the input shaft S1 to the output shaft S2. Each of the gear trains G1 to G4 and a clutch (not shown) incorporated between the input shaft S1 or the output shaft S2, and the gear trains G1 to G4 used by selectively engaging the clutches. This changes the speed ratio of the rotation transmitted from the input shaft S1 to the output shaft S2. 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 or the like. The input shaft S1 and the output shaft S2 are rotatably supported by a bearing B with a seal, respectively. Further, the transmission is splashed or sprayed by splashing of the lubricant accompanying the rotation of the gear or by jetting of lubricant 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 protrusion has an R shape. However, the protrusion cannot be in a fluid lubrication state when the relative peripheral speed with respect to the seal sliding surface is equal to or higher than a certain value, and has a rust effect. An appropriate shape may be used so as to be obtained. For example, a chamfered shape such as an R chamfer or a C chamfer can be employed.

  In each of the above-described embodiments, the example in which the protrusions are uniformly arranged in the circumferential direction has been described. However, the protrusions may be unevenly arranged.

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

  In each of the above-described embodiments, the inner ring rotation and the radial bearing are exemplified. However, the present invention can be applied to the outer ring rotation and the thrust bearing.

  It should be understood that the embodiments and examples disclosed herein are illustrative and non-restrictive in every respect. Accordingly, the scope of the present invention is defined by the terms of the claims, rather than the description above, and is intended to include any modifications within the scope and meaning equivalent to the terms of the claims.

100, 300, 400, 500, B Sealed bearing 110, 310, 410, 510 Inner ring 111, 121, 314, 411, 421 Track groove 112, 312, 313, 414, 415, 514, 515 Seal sliding surface 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 Protrusion 153, 203 Tip 154, 204 Surface 155, 206, 333, 343 Core metal 156 Vulcanized rubber material 160 Bearing internal space 170 Oil passage 205 Surface 315, 316 412, 413 Shoulder 332, 342 Outside Lip 350 Labyrinth clearance 430 Ball 511, 521 Raceway 512 Large brim 513 Small brim 530 Tapered roller S1 Input shaft (rotating part)
S2 Output shaft (rotating part)

Claims (7)

A seal member for separating the bearing internal space and the exterior;
A seal lip provided on the seal member;
A seal sliding surface sliding in a circumferential direction with respect to the seal lip;
A plurality of protrusions formed on the seal lip and creating an oil passage communicating between the bearing internal space and the outside between the seal sliding surface and the seal lip;
The gap between the protrusion and the seal sliding surface is formed in a wedge shape that is large on the oil passage side and small on the protrusion side,
A bearing with a seal in which the protrusions are arranged at intervals of 0.3 to 2.6 mm in the circumferential direction.   2. The protrusion extends in a direction perpendicular 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 protrusion toward the center of the circumferential width. A bearing with a seal as described in 1.   The bearing with seal according to claim 2, wherein the height of the protrusion is set to 0.07 mm or less. The R dimension of the protrusion is 0.15 mm or more and less than 2.0 mm,
The bearing with a seal according to claim 3, wherein a circumferential width of the protrusion is 0.2 mm or greater and 1.0 mm or less.   The bearing with a seal according to any one of claims 1 to 4, wherein the protrusions are arranged at uniform intervals over the entire circumference in the circumferential direction.   The bearing with a seal according to any one of claims 1 to 5, wherein the seal lip is formed of a vulcanized rubber material attached to at least an inner diameter portion of a cored bar.   The bearing with a seal according to any one of claims 1 to 6, which supports at least one rotating portion in a transmission and a differential of a vehicle.

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