JP2017155818A - Ball bearing - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a ball bearing which has an improved loading capacity in an axial load than a standard deep groove ball bearing, prevents the intrusion of foreign matters having a specified particle diameter, and achieves lower torque and higher speed of the ball bearing.SOLUTION: A projection 152 is formed on a seal lip 151 of a seal member 150. An oil passage 180 communicating between a bearing interior space 170 and the exterior and not allowing the passing of foreign matters having a specified particle diameter, is formed between a seal slide face 114 sliding in a circumferential direction with respect to the projection 152, and the seal lip 151. With the rotation of the bearing, lubrication oil in the oil passage 180 is drawn into an interspace between the seal lip 151 and the seal slide face 114, and oil film formation in the interspace is promoted.SELECTED DRAWING: Figure 1

Description

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

  A ball bearing with improved axial load capacity from a standard deep groove ball bearing as a rolling bearing suitable for applications that support rotating parts such as transmissions and differentials mounted on vehicles such as automobiles and various construction machines. There is a bearing (Patent Document 1).

  The ball bearing disclosed in Patent Document 1 includes an inner ring, an outer ring, and a plurality of balls interposed between the race grooves of the inner ring and the outer ring, and a pair of shoulder portions formed on both sides of the race groove of the inner ring. Of these, the shoulder on the load side that receives the axial load is formed higher than the shoulder on the opposite non-load side. As the shoulder on the load side is raised, the contact angle on the load side between the ball and raceway groove can be allowed to increase, so the low-torque property of the bearing operating torque that the deep groove ball bearing has Moreover, it is possible to achieve both the load capacity improvement of the axial load.

  When the ball bearing disclosed in Patent Document 1 is applied to a transmission, foreign matters such as gear wear powder are mixed in the transmission housing. It is done to prevent damage.

  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 seal member, by performing shot peening on the seal sliding surface, a seal sliding surface having minute irregularities with a maximum roughness Ry of 2.5 μm or less is obtained, and the lubricating oil stored in the concave portion is used. It has been proposed to promote oil film formation between the seal lip and the seal sliding surface (Patent Document 2).

JP 2000-145795 A JP 2007-107588 A

  However, the torque reduction by shot peening as in Patent Document 2 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 ball bearings due to lubrication conditions. In an electric vehicle (EV), there is a strong demand for high-speed operation of a ball bearing that supports a rotating portion of a 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 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-mentioned background, the problem to be solved by the present invention is to improve the bearing operation torque while preventing foreign matter intrusion with a predetermined particle size in a ball bearing with improved axial load capacity from a standard deep groove ball bearing. The aim is to reduce torque and speed.

  To achieve the above object, the present invention comprises an inner ring, an outer ring, and a plurality of balls interposed between the inner and outer raceway grooves, and a pair of shoulders formed on both sides of the inner ring raceway groove. In the ball bearing in which the shoulder on the load side receiving the axial load is formed higher than the shoulder on the opposite non-load side, both ends of the bearing inner space formed between the inner ring and the outer ring are sealed. Two seal members, a seal lip provided on the seal member, a seal formed on the load-side shoulder and the non-load-side shoulder, respectively, and sliding in the circumferential direction with respect to the seal lip A sliding surface, and a protrusion formed at least in one circumferential direction of the seal lip, and an oil passage communicating between the bearing internal space and the outside is formed between the seal sliding surface and the seal lip. More Wherein the sealing lip is obtained by the configuration in which the protrusion in the possible embodiments to between the sealing lip and the sealing sliding surface in a fluid lubrication state is formed.

According to the above configuration, since the shoulder portion on the load side is formed higher than the shoulder portion on the non-load side, the ball bearing has improved the load capacity of the axial load from the standard deep groove ball bearing.
An oil passage is formed between the seal sliding surface and the seal lip that seals the bearing internal space, and the lubricating oil in the oil passage is dragged between the seal sliding surface and the seal lip by the wedge effect as the bearing rotates. , Promote oil film formation during this period. For this reason, since 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), the seal torque is brought close to substantially zero, and the seal lip is substantially reduced. 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. Furthermore, the adsorption action is also prevented.
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, the present invention is a ball bearing in which the load capacity of the axial load is improved from the standard deep groove ball bearing by adopting the above-described configuration, and it is possible to reduce the bearing operation torque while preventing the intrusion of a predetermined particle size. And speeding up can be achieved.

Sectional drawing which shows the ball bearing which concerns on 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. Lubrication area diagram explaining fluid lubrication mode The partial perspective view which shows the seal lip which concerns on 2nd Example of this invention (A) is a schematic diagram showing a state of vulcanization molding of the seal lip of FIG. 10, and (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 13 is an enlarged view of the vicinity of the seal lip of the right seal member in FIG. Sectional drawing which shows an example of a transmission provided with the ball bearing which concerns on this invention

A preferred embodiment of the present invention will be described.
In the first embodiment, the diameter of the seal sliding surface formed on the shoulder portion on the load side is different from the diameter of the seal sliding surface formed on the shoulder portion on the non-load side. And the number of the projections on the seal member arranged on the same side as the seal sliding surface on the load side, and the projection on the seal member on the same side as the seal sliding surface on the non-load side. The number is different. The difference in outer diameter is different between the shoulder on the load side and the shoulder on the non-load side. Since it is difficult to make the diameter of the load-side seal sliding surface and the diameter of the non-load-side seal sliding surface the same, it is preferable to make the diameters of these seal sliding surfaces different. As a result, the seal sliding surface formed on the shoulder on the load side has a larger diameter than the seal sliding surface formed on the shoulder on the non-load side. The relative peripheral speed of the seal sliding surface with respect to the seal lip is higher than the relative peripheral speed of the seal sliding surface with respect to the non-load side seal lip. Also, during the relative rotation of the inner ring and the outer ring, the above-mentioned difference in the outer diameter between the shoulder portions causes a pumping action to feed the lubricating oil from the shoulder portion on the non-load side toward the shoulder portion on the load side in the bearing internal space. . Due to the effect of these peripheral speed differences and pumping action, the lubrication conditions differ between the load-side seal sliding surface and the seal lip and between the non-load-side seal slide surface and the seal lip. According to the first embodiment, since the number of protrusions on the load-side seal member and the number of protrusions on the non-load-side seal member are different, one rotation of the relative rotation between the inner ring and the outer ring is made. It is possible to make the number of passes of the hitting projections appropriate on the load side and the non-load side, and to form an equivalent oil film on both sides, or to optimize the thickness of the oil film.

  In the second embodiment, the diameter of the seal sliding surface formed on the shoulder portion on the load side is different from the diameter of the seal sliding surface formed on the shoulder portion on the non-load side. And the circumferential pitch angle of the protrusions at the seal member arranged on the same side as the seal sliding surface on the load side, and the side arranged on the same side as the seal sliding surface on the non-load side The circumferential pitch angle of the protrusions on the seal member is different. According to the second embodiment, since the circumferential pitch angle of the protrusions on the load-side seal member and the circumferential pitch angle of the protrusions on the non-load-side seal member are different, the inner ring and the outer ring The number of passages of protrusions per rotation relative to each other can be made appropriate on both the load side and the non-load side, and an equivalent oil film can be formed on both sides, and the thickness of the oil film can be optimized.

  In 3rd Embodiment, the said protrusion is arrange | positioned at equal intervals over the circumferential direction perimeter. According to the third embodiment, oil film formation can be promoted uniformly over the entire circumference of the seal sliding surface.

  In the fourth embodiment, 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. According to the fourth embodiment, as the bearing rotates, the lubricating oil in the oil passage is easily dragged to the protrusion side due to the wedge effect generated in the gap between the protrusion and the seal sliding surface, and the protrusion and the seal slide. An oil film is easily formed between the surfaces. Further, the protrusion shape corresponding to the wedge shape can reduce the area of the sliding contact that has a great influence on the seal torque.

  In the fifth 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 fifth embodiment, the protrusion extends in a direction orthogonal 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 with each other 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 from the tip, and there is no fear that the sealing torque reduction performance will be impaired during attachment.

  A first embodiment according to the present invention will be described below with reference to FIGS. As shown in FIG. 1, the first embodiment is a ball bearing 100 including an inner ring 110, an outer ring 120, a plurality of balls 140 held by a cage 130, and two seal members 150 and 160. ing. Hereinafter, a direction along the bearing central axis of the ball bearing 100 is referred to as an “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 170 is formed by the inner ring 110 and the outer ring 120. The plurality of balls 140 revolve in the bearing inner space 170 while being interposed between the inner ring 110 and the outer ring 120. Lubricating oil is supplied to the bearing internal space 170 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.

  The inner ring 110 and the outer ring 120 respectively have track grooves 111 and 121 that correspond to about one third of the circumference of the ball 140 in a transverse cross section (corresponding to the cross section shown in the figure) and are uninterrupted over the entire circumference in the circumferential direction.

  The inner ring 110 has a pair of shoulder portions 112 and 113 formed on both sides of the raceway groove 111. Among these shoulder portions 112 and 113, the shoulder portion 112 on the load side (right side in the drawing) that receives an axial load is formed higher than the shoulder portion 113 on the opposite non-load side (left side in the drawing). The shoulder 113 on the non-load side has a shoulder height equivalent to a deep groove ball bearing. Therefore, the ball bearing 100 has a load capacity of an axial load superior to that of the deep groove ball bearing while exhibiting the good low torque property of the deep groove ball bearing. The outer ring 120 also has a shoulder height difference between the shoulder on the load side and the shoulder on the non-load side.

  Seal grooves 122 and 122 for holding the seal members 150 and 160 are formed at the inner peripheral end of the outer ring 120. The seal members 150 and 160 are held in the seal grooves 122 and 122 of the outer ring 120. The seal members 150 and 160 are attached to the outer ring 120 by press-fitting their outer peripheral edges into the seal grooves 122 and 122.

  The seal members 150 and 160 separate the bearing inner space 170 from the outside, and seal both ends of the bearing inner space 170. On the outer side with the seal members 150 and 160 as a boundary, there are foreign matters such as gear wear powder, clutch wear powder, fine crushed stone, and the like depending on where the ball bearing 100 is assembled. Such powdery foreign matter can reach the vicinity of the seal members 150 and 160 by the flow of lubricating oil or atmosphere. Seal members 150 and 160 prevent foreign matter from entering the bearing internal space 170 from the outside.

  The seal members 150 and 160 have seal lips 151 and 161 protruding in a tongue shape on the inner peripheral side thereof. The seal lips 151 and 161 are radial lips. 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.

  Seal sliding surfaces 114 and 115 that slide in the circumferential direction with respect to the seal lips 151 and 161 are formed on the outer periphery of the inner ring 110. The seal sliding surface 114 formed on the shoulder portion 112 on the load side (right side in the figure) has a cylindrical surface extending over the entire circumference in the circumferential direction so as to define the outer diameter of the shoulder portion 112. The seal sliding surface 115 formed on the shoulder portion 113 on the non-load side (left side in the figure) has a cylindrical surface over the entire circumference in the circumferential direction so as to define the outer diameter of the shoulder portion 113 on the non-load side. Yes.

  Since there is a large diameter difference between the outer diameter of the load-side shoulder 112 and the outer diameter of the non-load-side shoulder 113, the diameter of the seal sliding surface 114 formed on the load-side shoulder 112, The diameter of the seal sliding surface 115 formed on the non-load side shoulder 113 is different. That is, the diameter of the seal sliding surface 114 on the right side in the drawing is larger than the diameter of the seal sliding surface 115 on the left side in the drawing. For this reason, during the bearing operation, the peripheral speed of the seal sliding surface 114 on the right side in the drawing becomes higher than that of the seal sliding surface 115 on the left side in the drawing.

  Further, during the operation of the bearing, in the bearing inner space, 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 difference in diameter between the shoulder portions 112 and 113.

  Between the seal member 150 on the right side in the drawing and the seal member 160 on the left side in the drawing, the outer diameter difference corresponding to the diameter difference between the seal groove 122 on the right side in the drawing and the seal groove 122 on the left side in the drawing of the outer ring 120. And an inner diameter difference corresponding to the difference in diameter between the seal sliding surface 114 on the right side of the inner ring 110 and the seal sliding surface 115 on the left side in the drawing is set. Yes. Therefore, for further details of the seal members 150 and 160, the seal member 150 on the right side in the drawing will be described as a representative example, and the seal member 160 will be given a number corresponding to the seal member 150 in FIG. Keep on.

  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 orthogonal to the seal sliding surface 114 in terms of design with respect to a gap in a direction orthogonal to the seal sliding surface 114 (including an oil passage 180 described later) between the seal lip 151 and the seal sliding surface 114. It shows the situation in the narrowest direction. 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 114, that is, a normal direction perpendicular to a tangent line contacting the seal sliding surface 114. Have. Since the seal sliding surface 114 has a cylindrical surface centered on the bearing central axis, the direction orthogonal thereto corresponds to the radial direction.

  The seal lip 151 has a tip 153 that defines the inner diameter of the seal lip 151 in the natural state of the seal member 150.

  A margin in the radial direction is set between the seal lip 151 and the seal sliding surface 114. Due to this tightening allowance, the seal lip 151 pressed in the radial direction against the seal sliding surface 114 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.

  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 114.

  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 114 only on each protrusion 152, and the oil passage 180 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 114 is stretched against the tight force of the seal lip 151, so that the bearing inner space 170 and the outer side are formed on both sides in the circumferential direction with the protrusion 152 as a boundary. An oil passage 180 communicating between them is formed. The lubricating oil passes from the outside through the oil passage 180 to the bearing internal space 170. Lubricating oil that has entered the bearing internal space 170 and base oil in the case where grease is sealed reach the outside through the oil passage 180 from the bearing internal space 170.

  The particle diameter that can pass through the oil passage 180 can be determined based on the protrusion height h of the protrusion 152 in the direction orthogonal to the seal sliding surface 114. Therefore, in the first embodiment, it is possible to arbitrarily determine the particle diameter to prevent intrusion and prevent foreign matters having the predetermined particle diameter from entering the bearing internal space 170 from the oil passage 180.

  The wear powder that causes early breakage of the rolling bearing is a foreign matter having a particle diameter exceeding 50 μm. If the protrusion height h of the protrusion 152 is set to 0.05 mm or less, an oil passage 180 through which such wear powder cannot pass can be generated. On the other hand, in order to improve the oil permeability of the oil passage 180, the protrusion height h of the protrusion 152 is preferably set to 0.05 mm or more.

  The gap formed between the protrusion 152 and the seal sliding surface 114 is formed in a wedge shape in which the side close to the circumferential direction of the oil passage 180 is large and the side close to the circumferential direction of the protrusion 152 is small. As shown in FIG. 3, when the seal sliding surface 114 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 180 (indicated by a dot pattern in the drawing) is dragged between the seal sliding surface 114 and the projection 152 of the seal lip 151 as the seal sliding surface 114 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 114 is lowered, and the seal torque is reduced. Further, oil permeability between the bearing internal space 170 and the outside is improved by the oil passage 180. For this reason, the temperature rise of the ball bearing 100 is suppressed, and consequently the adsorption 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 114 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 tightening 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 114). 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 ball bearing 100 is assumed to be used for supporting a rotating part in a transmission of a vehicle. In general, oil supply to a ball bearing existing in a transmission of a vehicle is performed by an appropriate method such as splashing, oil bath, nozzle injection, or the like. Therefore, lubricating oil exists around the seal fixed to the inner ring or outer ring of the ball bearing. 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 machine of model number 8000A manufactured by Hiac 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 with a particle size of 50 μm or more as in recent transmission environments, the life ratio of the rolling bearing (actual life even if foreign matter enters the bearing without seal) It was found that the ratio to the calculated life) shows a value (for example, about 7 to 10 times) that can be sufficiently put into practical use in an automobile transmission.

  Based on the above results, when the lubricating oil filtered by the oil filter is supplied to the ball bearing used to support the rotating part of the drive system such as the transmission of the vehicle and the differential gear, a large foreign matter having a particle diameter exceeding 50 μm As long as the seal member prevents the intrusion from entering the inside of the bearing, it can be said that there is no problem in the bearing life even if a foreign matter having a particle size of 50 μm or less contained in the lubricating oil is allowed to enter the inside of 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 0.05 mm. The height h of the protrusion 152 is a value at the highest position in the range where the seal sliding surface 114 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 114 is maximized. Since the deformation amount of the protrusion 152 during the operation of the bearing is negligible, a gap (including the oil passage 180) in the direction perpendicular to the seal sliding surface 114 between the seal lip 151 and the seal sliding surface 114 is the seal sliding. The narrowest portion in the direction orthogonal to the surface 114 has a width corresponding to the height h of the protrusion 152 and does not substantially exceed 0.05 mm. For this reason, even if a foreign matter having a particle size of more than 50 μm is included in the external lubricating oil, it is considered that the foreign matter hardly passes through the oil passage 180.

  When the seal sliding surface 114 rotates relative to the seal lip 151 in the direction of the arrow in the figure, the lubricating oil in the oil passage 180 is dragged into the wedge-shaped gap between the projection 152 and the seal sliding surface 114. It is. The wedge angle in the above-mentioned wedge-shaped gap gradually decreases from the wide oil passage 180 where the drawn lubricating oil exists to the narrow side, so that the protrusion 152 and the seal sliding surface 114 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. Therefore, 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 114, 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 114 a fluid lubrication state.

  Here, if there is an oil film that completely separates the protrusion 152 and the seal sliding surface 114, a fluid lubrication state is reached in which the seal sliding surface 114 slides against the protrusion 152 without being in direct contact. . By maintaining such an oil film between each protrusion 152 and the seal sliding surface 114, the space between the seal lip 151 and the seal sliding surface 114 can be in a fluid lubrication state.

The fluid lubrication condition, theoretically calculated, (see FIG. 9) in the lubricating area diagram in the case of a line contact which is based on the viscosity parameter _{g v} and elastic parameter _{g e} is a dimensionless number decided in Greenwood-Johnson, etc viscosity - This corresponds to a lubrication mode corresponding to one of a rigid body region (RI mode) or an isoviscous-elastic body region (EI mode, soft EHL). In addition, the plot shown in FIG. 9 illustrates the case corresponding to the RI mode or EI mode.

  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 114 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 treatment is not performed on the seal sliding surface 114, and the maximum height roughness Rz of the seal sliding surface 114 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.

  The protrusion 152 may be formed on the seal lip 151 in such a manner that the height h is 0.05 mm or less and the space between the seal lip 151 and the seal sliding surface 114 can be in a fluid lubrication state. The aspect can be determined by the distance d between the protrusions 152 adjacent in the circumferential direction, the circumferential width w of the protrusions 152, the pitch angle θ of the protrusions 152 arranged at a constant interval in the circumferential direction, and the shape of the protrusions 152. The smaller the distance d between the protrusions 152, that is, the greater the number of protrusions 152, the more the number of times the protrusion 152 passes per rotation when the seal sliding surface 114 rotates in the circumferential direction relative to the seal lip 151. Since the oil film is kept in a continuous state over the entire circumference in the circumferential direction of the seal sliding surface 114 and the wedge effect between the protrusions 152 is easily generated without interruption, it is easy to maintain the fluid lubrication state. Become.

  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.

  If the oil film thickness between the protrusion 152 and the seal sliding surface 114 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.

  Further, the number of protrusions 152 on the right seal member 150 in FIG. 1 is different from the number of protrusions 162 on the left seal member 160 in FIG. Further, the circumferential pitch angle of the protrusion 152 in the right seal member 150 in FIG. 1 is different from the circumferential pitch angle of the protrusion 162 in the left seal member 160 in FIG. The difference between the seal member 150 on the right side in FIG. 1 and the seal sliding surface 114 and the seal member 160 on the left side in FIG. Because there is a difference, the oil film formed between each of these left and right should be equalized and optimized so that foreign matter exceeding 50 μm in diameter is not likely to occur due to the formation of an oil film that is too thick. It is set for the purpose.

  As described above, the ball bearing 100 according to the first embodiment has the seal lips 151 and 161 while the seal members 150 and 160 prevent entry of foreign matters having a particle size that adversely affects the bearing life. In addition, the friction coefficient μ of the sliding between the seal sliding surfaces 114 and 115 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 FIG. 1, see FIG.

  Furthermore, this ball bearing 100 has a peripheral speed (for example, 30 m / s or more) of the seal sliding surface that would cause a problem of heat generation due to wear of the seal lip or sliding between the seal lip and the seal sliding surface. In the case of operating with the seal lip 151, 161 and the seal sliding surfaces 114, 115 can be brought into a fluid lubrication state without direct contact, so that the wear of the seal lips 151, 161 is substantially eliminated. The aforementioned heat generation can also be suppressed. For this reason, this ball bearing 100 can respond to the demands for downsizing and high-speed operation of ball bearings that could not be achieved conventionally.

  Further, the ball bearing 100 has the number of protrusions 152 and circumferential pitch angle θ on the right seal member 150 in FIG. 1 and the number of protrusions 162 and circumferential pitch angle on the left seal member 160 in FIG. Because of the difference, the number of passages of the protrusions 152 and 162 per rotation of the relative rotation between the inner ring 110 and the outer ring 120 is determined between the seal member 150 on the load side (right side in the figure) and the seal sliding surface 114, and no load It is appropriate between the seal member 160 on the side (left side in the figure) and the seal sliding surface 313, and an equivalent oil film can be formed on both sides, and the thickness of the oil film can be optimized. As a state, it is possible to achieve both a reduction in torque and suppression of entry of foreign matter.

  Further, since the protrusion 152 is formed in an R shape in the ball bearing 100, even when the protrusion 152 is rubbed against the seal sliding surface 114 when the seal member 150 is attached to the outer ring 120, the protrusion 152 does not circulate from the tip. There is no fear of bending in the direction, and there is no risk of impairing the sealing torque reduction performance during installation. For example, when the protrusions are pointed, it is not possible to know which side of the circumferential direction the ends of the many protrusions rubbed against the seal sliding surface when the seal member is attached, and in the direction of relative rotation with the seal sliding surface. On the other hand, it is extremely difficult to attach all the protrusions so as to bend in a direction in which a proper wedge-shaped gap is formed. 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 projection shape is changed from the first embodiment. As shown in FIG. 10, the protrusion 201 according to the second embodiment has a shape that gradually decreases toward the tip 203 of the seal lip 202. FIG. 10 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 201 (the radius of curvature at the surface 204 of the protrusion 201) and the center of curvature are gradually reduced 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. That is, the tip 203 of the seal lip 202 is substantially an edge where two conical surfaces intersect, and the surface 205 is substantially a part of one of the conical surfaces.

  FIG. 11 shows a state where the seal lip 202 is vulcanized. Note that FIG. 11 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. Accordingly, the tip 153 that defines the inner diameter of the seal lip 151 in the natural state is a boundary line between the upper surface portion of the seal lip 151 in contact with the transfer surface of the upper mold Mp1 and the lower surface portion of the seal lip 151 in contact with 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. 12 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. 11, 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. 12 is generated. Thus, according to the second embodiment, it is possible to 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 ball 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.

  As in the first embodiment, the inner ring 310 is formed with a shoulder portion 315 on the load side (right side in the drawing) and a shoulder portion 316 on the non-load side (left side in the drawing). Seal grooves 311 and 311 and seal sliding surfaces 312 and 313 are formed in the shoulder portions 315 and 316, respectively.

  The diameter difference between the seal sliding surface 312 on the right side in the drawing and the seal sliding surface 313 on the left side in the drawing is about the same as that in the first embodiment, and the difference in peripheral speed during the bearing operation is also about the same. . For this reason, the number of protrusions 334 and 344 and the difference in the circumferential pitch angle between the seal member 330 on the right side in the drawing and the seal member 340 on the left side in the drawing are the same as those in the first embodiment. Thereby, also in the ball bearing 300 according to the third embodiment, an oil film is optimally and equally formed between the seal member 330 and the seal sliding surface 312 and between the seal member 340 and the seal sliding surface 313. Thus, both lowering the torque to a fluid lubrication state and suppressing the entry of foreign matter are achieved.

  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).

  As shown in FIGS. 13 and 14, the outer lips 332 and 342 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.

  Since the ball bearing 300 according to the third embodiment makes it difficult for the foreign matter to reach the seal lips 331 and 341 due to the formation of the labyrinth clearance 350, the foreign matter intrusion is further suppressed so as not to inhibit the torque reduction. Can do. 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 first seal lips provided as radial lips are employed. There is no concern that the bearing life is inferior to that of the second embodiment.

  FIG. 15 shows an example in which a ball bearing 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 in which the gear ratio is changed stepwise, and the ball bearing B that rotatably supports the rotating portion (for example, the input shaft S1 and the output shaft S2) is used in the above-described embodiment. Such a ball bearing 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 ball bearings B, 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. However, it is applied to the side surface of each ball bearing B.

  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.

  Further, 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 arranged unevenly or only at one place in the circumferential direction. The oil passage can be formed by the protrusion even at one place, and the effect of reducing the seal torque can be expected.

  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, B Ball bearing 110, 310 Inner ring 111, 121, 314 Race groove 112, 113, 315, 316 Shoulder 114, 115, 312, 313 Seal sliding surface 122, 311, 321, 322 Seal groove 120, 320 Outer ring 140 Ball 150, 160, 330, 340 Seal member 151, 161, 202, 331, 341 Seal lip 152, 162, 201, 334, 344 Projection 153, 203 Tip 154, 204 Surface 155, 206, 333, 343 Core 156 Vulcanized rubber material 170 Bearing inner space 180 Oil passages 332 and 342 Outer lip 350 Labyrinth clearance S1 Input shaft (rotating part)
S2 Output shaft (rotating part)

Claims (10)

An inner ring, an outer ring, and a plurality of balls interposed between the race grooves of the inner ring and the outer ring, and of the pair of shoulder portions formed on both sides of the race groove of the inner ring, the shoulder on the load side that receives an axial load In the ball bearing in which the portion is formed higher than the shoulder on the opposite non-load side,
Two sealing members for sealing both ends of a bearing internal space formed between the inner ring and the outer ring;
A seal lip provided on the seal member;
A seal sliding surface formed on each of the load-side shoulder and the non-load-side shoulder and sliding in the circumferential direction with respect to the seal lip;
A protrusion that is formed at least in one circumferential direction of the seal lip and that causes an oil passage communicating between the bearing internal space and the outside between the seal sliding surface and the seal lip;
The ball bearing according to claim 1, wherein the protrusion is formed on the seal lip in such a manner that a fluid lubrication state can be formed between the seal lip and the seal sliding surface. The diameter of the seal sliding surface formed on the shoulder on the load side is different from the diameter of the seal sliding surface formed on the shoulder on the non-load side,
The number of the protrusions in the seal member disposed on the same side as the seal sliding surface on the load side, and the seal member disposed on the same side as the seal sliding surface on the non-load side. The ball bearing according to claim 1, wherein the number of protrusions is different. The diameter of the seal sliding surface formed on the shoulder on the load side is different from the diameter of the seal sliding surface formed on the shoulder on the non-load side,
The circumferential pitch angle of the protrusions at the seal member disposed on the same side as the seal sliding surface on the load side, and the seal member disposed on the same side as the seal sliding surface on the non-load side The ball bearing according to claim 1, wherein a circumferential pitch angle of the protrusion is different from each other.   The ball bearing according to any one of claims 1 to 3, wherein the protrusions are disposed at uniform intervals over the entire circumference in the circumferential direction.   The ball bearing according to any one of claims 1 to 4, wherein a 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.   The protrusion extends in a direction orthogonal to the circumferential direction, and the protrusion has an R shape that gradually approaches the seal sliding surface from both ends of the circumferential width toward the center of the circumferential width. 5. A ball bearing according to 5.   The ball bearing according to claim 1, wherein the seal lip is a radial lip.   The ball bearing according to claim 1, wherein the seal lip is an axial lip.   The ball bearing according to any one of claims 1 to 8, wherein the seal lip is formed of a vulcanized rubber material attached to at least an inner diameter portion of a cored bar.   The ball bearing according to claim 1, wherein the ball bearing supports at least one rotating part in a transmission and a differential of the vehicle.

Images