JP7369511B2 - sealed ball bearings - Google Patents

Description

The present invention relates to a sealed ball bearing in which grease is sealed inside the bearing.

For example, foreign matter such as wear particles from gears is mixed in transmissions installed in vehicles such as automobiles and various construction machines. For this reason, a seal member is used to prevent foreign matter from entering the bearing internal space and to prevent early failure of the rolling bearing.

Furthermore, in applications where low torque performance of rolling bearings is important, ball bearings, which rotate with lower torque than roller bearings, are used.

In applications that require both prevention of early bearing damage and low torque performance, such as applications that support rotating parts such as vehicle transmissions and differentials, the annular bearing internal space formed by the inner and outer bearing rings and the outer A ball bearing with a seal is used, which is separated by a seal member. Grease is sealed in the bearing internal space as an initial lubricant.

A typical sealing member has a sealing lip made of a rubber-like material or the like. Bearing components such as bearing rings and slingers that rotate in the circumferential direction with respect to the seal member are formed with a seal sliding surface that brings the seal lip into sliding contact. The seal lip and the seal sliding surface are in sliding contact over the entire circumference.

Further, the sealing member has a peripheral edge portion that fits into a mating member. A slit communicating between the bearing internal space and the outside is formed at the peripheral edge thereof. While the bearing is rotating, the seal lip and the seal sliding surface come into sliding contact over the entire circumference, but lubricating oil supplied from the outside flows into the bearing internal space through the slit, and grease flows out to the outside.

When the seal lip and the seal sliding surface come into sliding contact around the entire circumference, drag resistance (seal torque) of the seal lip occurs, which contributes to bearing torque. The friction of the sliding contact also promotes an increase in the temperature of the bearing. As this temperature rise progresses, an adsorption effect occurs due to the pressure difference between the inner space and the outside of the bearing, and the resulting friction increases.

Conventionally, in order to suppress seal torque, shot peening is applied to the seal sliding surface to create a seal sliding surface with minute irregularities with a maximum roughness Ry of 2.5 μm or less, and the lubricating oil stored in the recesses is used to strengthen the seal lip and seal. It has been proposed to promote the formation of an oil film between sliding surfaces (Patent Document 1).

Japanese Patent Application Publication No. 2007-107588

However, the demand for energy saving in vehicles and the like continues to increase, and sealed ball bearings incorporated therein are also required to have even lower torque performance.

Low torque by shot peening as described in Patent Document 1 is achieved by reducing the sliding area between the seal lip and the seal sliding surface, but since there is a limit to this reduction, it is difficult to achieve the low torque that can be achieved. was limited.

In addition, at the beginning of bearing operation, the temperature of the lubricating oil is relatively low, so the viscosity of the lubricating oil is relatively high, making it easy to form an oil film, but as the bearing continues to operate, the oil temperature rises and the viscosity decreases. This creates a lubrication condition in which the oil film is likely to break. The higher the bearing is operated, the higher the circumferential speed of the seal sliding surface relative to the seal lip, and the more heat generated due to friction between the seal lip and the seal sliding surface, the higher the oil temperature and the seal lip. Wear tends to progress. Therefore, there are limits to the high-speed operation and allowable rotational speed of sealed ball bearings due to lubrication conditions. In electric vehicles (EVs), there is a strong demand for high-speed operation of the sealed ball bearings that support the rotating parts of the drive system, but reducing torque by shot peening cannot meet these demands.

It is possible to eliminate seal torque by not setting an interference between the seal member and the bearing parts that rotate in the circumferential direction relative to the seal member, and by making the seal member completely non-contact with the bearing parts. be. On the other hand, it becomes difficult to manage various errors in the size of the gap between the seal member and the bearing component so as to prevent foreign matter of a predetermined particle size from entering.

Furthermore, in the bearing internal space, grease is stirred by balls and a retainer. The agitation resistance of the grease also contributes to bearing torque. Therefore, in order to reduce the torque of the sealed ball bearing, it is preferable to suppress the stirring resistance of the grease.

In view of the above-mentioned background, the problem to be solved by this invention is to reduce the torque of a sealed ball bearing while preventing foreign matter of a predetermined particle size from entering the inside of the bearing and making the bearing compatible with high-speed operation. That's true.

In order to achieve the above object, the present invention provides an inner ring, an outer ring forming an annular bearing internal space between the inner ring, a plurality of balls interposed between the inner ring and the outer ring, and a plurality of balls interposed between the inner ring and the outer ring. a cage that holds the balls of the bearing, a seal member that separates the inner space of the bearing from 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 protrusion formed at at least one location in the circumferential direction of the seal lip to create an oil passage between the seal sliding surface and the seal lip that communicates between the bearing interior space and the exterior; and a grease sealed in a space, the protrusion is formed on the seal lip in a manner that allows fluid lubrication between the seal lip and the seal sliding surface, and the grease is sealed in the space. This is a ball bearing with a seal whose volume is 1 to 10% of the total space of the bearing. Here, the total bearing space volume refers to the volume obtained by subtracting the volumes of all the balls and the cage from the volume of the bearing internal space surrounded by the inner ring, the outer ring, and the seal member.

According to the above configuration, an oil passage is created by the protrusion 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 by a wedge effect as the bearing rotates. Promotes oil film formation during this time. Therefore, a state of fluid lubrication is created between the seal lip and the seal sliding surface. In the fluid lubrication state, the seal torque is substantially close to zero, the seal lip is not substantially worn, and heat generation due to sliding between the seal lip and the seal sliding surface can be suppressed. Therefore, the permissible relative circumferential speed of the seal sliding surface with respect to the seal lip is also increased, and it becomes possible to meet the demand for high-speed operation of sealed ball bearings, which was previously unachievable. Further, the particle size of foreign particles that can pass through the oil passage can be determined based on the height of the protrusion. Therefore, it is possible to arbitrarily determine the particle size to be prevented from entering and prevent foreign matter of the predetermined particle size from entering from the oil passage. Furthermore, since the oil passage improves oil permeability between the bearing internal space and the outside, temperature rise in the rolling bearing is suppressed, and as a result, adsorption is also prevented. Furthermore, if the oil permeability between the bearing internal space and the outside is improved, the amount of initially filled grease that should be filled into the bearing internal space as an initial lubricant will be reduced. By suppressing the amount of initially filled grease to 1 to 10% of the total space volume of the bearing, the stirring resistance of the grease can be suppressed.

Preferably, the arithmetic mean roughness Ra of the seal sliding surface is 1.0 μm or less. Here, the calculated average roughness Ra refers to the arithmetic average roughness Ra defined in JIS B 0601:2013 of the Japanese Industrial Standards. If the arithmetic mean roughness Ra of the seal sliding surface is 1.0 μm or less, a good fluid lubrication state can be achieved.

As described above, by adopting the above configuration, the present invention prevents foreign matter of a predetermined particle size from entering the internal space of the bearing, enables high-speed operation of the bearing, and suppresses sealing torque and grease agitation resistance. , the torque of the sealed ball bearing can be reduced.

A sectional view showing a sealed ball bearing according to an embodiment of the invention Enlarged view of the area around the seal lip of the seal member on the right side of the diagram in Figure 1 Enlarged sectional view taken along line III-III in Figure 2 A partial front view showing the seal lip according to the embodiment from the axial direction. Lubrication area diagram showing the study results of the fluid lubrication mode in the embodiment A sectional view showing an example of a transmission equipped with the sealed ball bearing of FIG. 1

An exemplary embodiment of the present invention will be described based on FIGS. 1 to 6.
This sealed ball bearing 100 shown in FIG. , a retainer 5 that holds the plurality of balls 4 , a pair of seal members 6 that separate the bearing internal space 2 from the outside, a seal lip 7 provided on the seal member 6 , and a seal lip 7 provided to the seal lip 7 . A seal sliding surface 8 that slides in the circumferential direction, a plurality of protrusions 9 formed on the seal lip 7, and grease 10 sealed in the bearing internal space 2 (indicated by a dot pattern in the figure). Be prepared.

Hereinafter, the direction along the bearing center axis, which is the designed rotation center of the sealed ball bearing 100, will be referred to as the "axial direction." The direction perpendicular to the axial direction is called the "radial direction." The circumferential direction around the bearing center axis is called the "circumferential direction."

The inner ring 1 and the outer ring 3 are each made of steel bearing rings. The inner ring 1 has raceway grooves 11 on the outside. The outer ring 3 has a raceway groove 12 inside. The bearing internal space 2 is an annular space defined by the outer circumference of the inner ring 1 and the inner circumference of the outer ring 3, which are coaxially arranged. The plurality of balls 4 revolve while being interposed between the inner raceway groove 11 and the outer raceway groove 12. The retainer 5 is an annular bearing component in which a plurality of balls 4 are evenly distributed in the circumferential direction. The inner ring 1, the outer ring 3, the plurality of balls 4, and the retainer 5 constitute a single row deep groove ball bearing.

The inner ring 1 is attached to the rotating shaft 200 and rotates together with the rotating shaft 200. The rotating shaft 200 is provided, for example, as a rotating part of a transmission or differential of a vehicle. The outer ring 3 is attached to a member, such as a housing or a gear, to which a load from the rotating shaft is applied.

A seal groove 13 for holding the seal member 6 is formed at the end of the inner circumference of the outer ring 3. The seal member 6 is attached to the outer ring 3 by press-fitting its outer periphery into the seal groove 13.

The seal lip 7 of the seal member 6 protrudes like a tongue from the inner peripheral side of the seal member 6. The seal lip 7 is a radial lip. Here, the radial lip is a seal lip that performs a sealing action with a seal sliding surface along the axial direction or a seal sliding surface having an acute angle of 45° or less with respect to the axial direction, and is Refers to something that has a radial interference between it and the moving surface.

The seal sliding surface 8 is formed on the outer periphery of the inner ring 1. The seal sliding surface 8 has a cylindrical surface shape that continues all around the circumferential direction.

FIG. 2 shows an enlarged view of the vicinity of the seal lip 7 of the seal member 6 in FIG. As shown in FIGS. 1 and 2, the protrusion 9 of the seal lip 7 creates an oil passage 14 between the seal sliding surface 8 and the seal lip 7 that communicates between the bearing interior space 2 and the outside.

By forming a plurality of protrusions 9 on the seal lip 7, the sliding area between the seal lip 7 and the seal sliding surface 8 is reduced. Here, since the friction torque T between the seal lip 7 and the seal sliding surface 8 is expressed as in Equation 1 below, it can be seen that the friction torque T decreases as the sliding area decreases.

The sealing member 6 is formed of a core metal 15 made of a metal plate and a vulcanized rubber material 16 attached to at least the inner diameter portion of the core metal. The core metal 15 is a pressed part formed in an annular shape that continues all around the circumferential direction. The vulcanized rubber material 16 is a rubber portion formed by vulcanization. The seal lip 7 is formed of a vulcanized rubber material 16. Examples of vulcanized rubber materials include ACM (polyacrylic rubber) with excellent heat resistance and chemical resistance, NBR (nitrile-butadiene rubber), HNBR (hydrogenated acrylonitrile-butadiene rubber) with excellent heat resistance, and EPDM ( (ethylene propylene rubber), etc. Furthermore, examples of other vulcanized rubber materials include FKM (fluororubber), silicone rubber, and the like.

The sealing member 6 is manufactured as an integral component by, for example, placing the core metal 15 in a mold and vulcanizing the vulcanized rubber material 16. A protrusion 9 is formed on the seal lip 7 during vulcanization. The vulcanized rubber material 16 may be attached to the entire core metal 15 or only to the inner diameter portion of the core metal 15.

On the outside of the bearing internal space 2 with the seal member 6 as a boundary, there are foreign substances depending on where the sealed ball bearing 100 is installed, such as gear wear powder, clutch wear powder, and micro-crushed stones. Such powdery foreign matter may reach the vicinity of the seal member 6 due to the flow of lubricating oil or atmosphere. The seal member 6 prevents foreign matter from entering the bearing internal space 2 from the outside.

The grease 10 sealed in the bearing internal space 2 is used to prevent the steel parts such as the inner ring 1 and the outer ring 3 from rusting in the bearing internal space 2 until the sealed ball bearing 100 starts to be used. This initial lubricant is intended to lubricate the inner ring 1, outer ring 3, balls 4, cage 5, and seal members 6 until the lubricating oil is supplied to the sealed ball bearing 100.

As the grease 10, it is preferable to use one that has good fretting resistance and excellent rust prevention performance. Specifically, the base oil is a mineral oil type, the thickener is a non-soap type such as urea, and the additives include a sulfonate type rust preventive agent and a phenolic type antioxidant, such as Pyroc Universal N6C. (Product name; manufactured by JX Energy). The base oil of Grease 10 has a kinematic viscosity of 100 mm ² /s or less at 40°C.

During rotation of the sealed ball bearing 100, lubricating oil is splashed from the outside and supplied by an appropriate means such as an oil bath. The lubricating oil enters the bearing internal space 2 from between the seal lip 7 and the seal sliding surface 8. Further, the lubricating oil and grease 10 in the bearing internal space 2 are discharged to the outside from between the seal lip 7 and the seal sliding surface 8.

A cross-sectional view taken along line III--III in FIG. 2 is shown in FIG. This cross section shows the gap between the seal lip 7 and the seal sliding surface 8 (including the oil passage 14) at the narrowest point in the design direction orthogonal to the seal sliding surface 8. . Further, FIG. 4 shows the appearance of the seal lip 7 when viewed in the axial direction from the bearing inner space side. FIG. 4 depicts the outer shape of the seal lip 7 of the seal member 6 shown in FIG. 1 in its independent and natural state. Here, the natural state refers to a state in which no external force is acting on the seal member in its independent state, that is, a state in which the seal member is not deformed by external force (hereinafter, this state is simply referred to as "natural state"). (called a state).

As shown in FIGS. 2 to 4, the protrusion 9 of the seal lip 7 has a height in the direction orthogonal to the seal sliding surface 8, that is, in the normal direction perpendicular to the tangent to the seal sliding surface 8. Since the seal sliding surface 8 has a cylindrical surface shape centered on the bearing center axis, the direction perpendicular thereto corresponds to the radial direction.

A radial interference is set between the seal lip 7 and the seal sliding surface 8. Due to this interference, the seal lip 7 pressed in the radial direction against the seal sliding surface 8 undergoes a rubber-like elastic deformation that bends outward, creating a tension force on the seal lip 7. Installation errors, manufacturing errors, etc. of the seal member 6 are absorbed by changes in the degree of bending of the seal lip 7.

As shown in FIG. 4, the sealing lip 7 has a tip 17 that defines the inner diameter of the sealing lip 7 in its natural state. As shown in FIGS. 2 to 4, the protrusion 9 extends in a direction perpendicular to the circumferential direction. The protrusion 9 extends to the tip 17 of the seal lip 7, and extends over almost the entire range with a radial interference between the protrusion 9 and the seal sliding surface 8.

The projection 9 has a shape that gradually becomes lower toward the tip 17 of the seal lip 7. In this way, if the protrusion 9 has a shape that gradually becomes lower toward the tip 17 of the seal lip 7, it is possible to prevent burrs from forming at the parting line on the tip 17.

The projection 9 has an R shape that gradually approaches the seal sliding surface 8 from both ends of the circumferential width w toward the center of the circumferential width w. This rounded shape is provided over the entire length of the protrusion 9 in the radial direction. Therefore, a region where the protrusion 9 and the seal sliding surface 8 can come into sliding contact exists linearly on the virtual axial plane Pax passing through the center of the circumferential width w of the protrusion 9. The center of curvature of the R shape of the protrusion 9 is on the virtual axial plane Pax. Regarding the radius of curvature _R1 and center of curvature of the protrusion 9, in order to make the protrusion 9 gradually lower toward the tip 17 of the seal lip 7, the radius of curvature _R1 is gradually increased toward the tip 17 of the seal lip 7, In addition, the center of curvature has been moved to the outside.

The protrusions 9 are arranged at regular intervals d in the circumferential direction. When considering the appearance of the seal lip 7 viewed from the axial direction, a plurality of protrusions 9 appear radially arranged in the circumferential direction at a constant pitch angle θ corresponding to the interval d. Note that the radiation center is on the central axis of the seal member 6 (not shown) (coinciding with the bearing central axis).

The spacing d between the radially arranged protrusions 9 and the circumferential width w of the protrusions 9 are such that the seal lip 7 can make sliding contact with the seal sliding surface 8 only on each protrusion 9, and there is no space between each protrusion 9. It is set so that the oil passage 14 is always formed. That is, when the seal member 6 is installed, the protrusion 9 that contacts the seal sliding surface 8 is stretched against the tension force of the seal lip 7, so that the bearing internal space 2 and the outside are An oil passage 14 is created that communicates between the two.

The diameter of the particles that can pass through the oil passage 14 can be determined based on the height h of the protrusion 9 in the direction orthogonal to the seal sliding surface 8 . Therefore, it is possible to arbitrarily set the particle size to be prevented from entering and prevent foreign matter of the predetermined particle size from entering the bearing internal space 2 from the oil passage 14.

Wear powder that causes early bearing failure is foreign matter with a particle size exceeding 50 μm (0.05 mm). By setting the height h of the protrusion 9 to 0.07 mm or less, it is possible to create an oil passage 14 through which such abrasion powder cannot pass. Note that in order to prevent the passage of foreign matter having a particle size of more than 50 μm, the height h of the protrusion 9 is preferably set to 0.04 mm or less, but in consideration of tolerances, it is set to 0.07 mm or less. Further, by setting the height h of the protrusion 9 to 0.07 mm, the oil permeability of the oil passage 14 can be improved.

The gap created between the protrusion 9 and the seal sliding surface 8 is formed in a wedge shape, with the gap closer to the oil passage 14 in the circumferential direction being larger and the gap closer to the protrusion 9 in the circumferential direction being smaller. The surface of the protrusion 9 has a curved surface corresponding to the wedge shape described above. As shown in FIG. 3, when the seal sliding surface 8 rotates in the circumferential direction with respect to the seal lip 7 as the inner ring 1 rotates (the direction of rotation is indicated by arrow A in the figure), the oil passage The lubricating oil in 14 (indicated by a dot pattern in the figure) is dragged between the seal sliding surface 8 and the protrusion 9 as the seal sliding surface 8 rotates, and promotes the formation of an oil film therebetween. Therefore, the coefficient of friction (μ) between the seal lip 7 and the seal sliding surface 8 decreases, and the seal torque decreases. Furthermore, oil permeability between the bearing internal space 2 and the outside is improved by the large number of oil passages 14. Therefore, the temperature rise of the sealed ball bearing 100 is suppressed, and the adsorption effect of the seal lip 7 is also prevented.

When the inventor of the present application conducted a test to confirm the above-mentioned oil permeability, the time required for 100 ml of lubricating oil to penetrate between the seal lip 7 and the seal sliding surface 8 was 500 seconds or less. On the other hand, similar tests were conducted on a conventional product (one in which the seal lip and seal sliding surface are in contact all around the circumference, and a small slit for oil passage is formed at one location on the outer diameter of the seal member). , the time required for 100 ml of lubricating oil to penetrate between the seal lip and the seal sliding surface exceeded 3,700 seconds. That is, the sealed ball bearing 100 can achieve oil permeability that is approximately eight times that of conventional products.

When there is such good oil permeability, the inside of the bearing internal space 2 can be quickly lubricated with the lubricating oil supplied to the side surface of the sealed ball bearing 100 from the outside. Therefore, it is possible to suppress the amount of grease 10 initially sealed in the bearing internal space 2 to a range of 1% to 10% of the total space volume of the bearing. The total bearing space volume corresponds to the volume obtained by subtracting the volumes of all the balls 4 and the retainer 5 from the volume of the bearing internal space 2 separated from the outside by the pair of seal members 6. Since use of the sealed ball bearing 100 is started with a small amount of grease 10, which is 10% or less of the total space volume of the bearing, sealed in the bearing internal space 2, while the sealed ball bearing 100 is rotating, the grease 10 is 4 and the retainer 5, and the resistance to stirring of the grease 10 is suppressed.

This sealed ball bearing 100 is intended for use in supporting a rotating part within a vehicle transmission. Oil supply to sealed ball bearings present in a vehicle transmission is generally carried out by an appropriate method such as splashing by rotation of a gear, an oil bath, or nozzle injection. Therefore, lubricating oil exists around the seal lip of the seal member fixed to the inner ring or outer ring of the sealed ball bearing. The lubricating oil supplied is also commonly used for other lubricated parts such as gears in the transmission. The lubricating oil is circulated by an oil pump and filtered by an oil filter provided in the circulation path.

The inventor of the present application collected lubricating oil used in vehicles according to the distance traveled by the vehicle, and investigated the number of foreign substances mixed in the used lubricating oil, the particle size distribution of the foreign particles, and the material of the foreign substances. Lubricating oil was collected from 8 vehicles with automatic transmission (AT) or manual transmission (MT) and 10 vehicles with continuously variable transmission (CVT). The manufacturers, models, and mileage of the vehicles targeted for recall vary widely. It was observed that AT/MT, which uses a lot of gears, had a tendency to have a larger particle size and number of foreign particles than CVT, but regardless of the transmission type, the particle size distribution of foreign particles was 50 μm or less. More than 99.9% of the cases were. The number of foreign particles with 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 running distance increased. This indicates that the performance of oil filters has improved in recent years, and foreign matter in lubricating oil has become finer (that is, foreign matter with large particle sizes are removed by oil filters). Note that this distribution was measured using a fine particle contaminant amount method using a measuring device model number 8000A manufactured by Hyac Leuco.

On the other hand, when the lubricating oil inside the bearing contains foreign matter, we investigated the relationship between the particle size of the foreign matter and the bearing life, and found that there is a tendency for the bearing life to decrease as the number of foreign matter with a large particle size increases. However, if there is a small amount of foreign matter with a particle size of 50 μm or more, such as in the environment in recent transmissions, even if foreign matter gets into the bearing without a seal, the life ratio of the rolling bearing (actual life It was found that the ratio (to the calculated lifespan) showed a value (for example, about 7 to 10 times) that could be used in practical use in automobile transmissions.

In other words, when lubricating oil that is filtered through an oil filter is supplied to sealed ball bearings used to support rotating parts of drive systems such as vehicle transmissions and differentials, large foreign particles with a particle size exceeding 50 μm may be inside the bearing. As long as the sealing member prevents foreign matter from entering the bearing, it can be said that even if foreign matter contained in the lubricating oil with a particle size of 50 μm or less is allowed to enter the inside of the bearing, it will not cause problems in the life of the bearing. If this is allowed, it is necessary to ensure sufficient flow of lubricating oil between the seal lip and the seal sliding surface, and in combination with the wedge effect mentioned above, fluid lubrication between the seal lip and the seal sliding surface is achieved. It is possible to achieve this state.

Therefore, the height h of the protrusion 9 shown in FIGS. 2 and 3 is set to 0.07 mm at the highest position within the range where it can come into sliding contact with the seal sliding surface 8. This position is also where the interference set between each protrusion 9 and the seal sliding surface 8 is maximum. During bearing operation, the radial size of the oil passage 14 does not become larger than the height h of the protrusion 9, and does not substantially exceed 0.07 mm. For this reason, even if foreign matter with a particle size exceeding 50 μm is contained in the external lubricating oil, it is considered that the foreign matter will almost never pass through the oil passage 14.

Since the wedge angle in the wedge-shaped gap between the protrusion 9 and the seal sliding surface 8 gradually becomes smaller toward the narrower side, a linear region ( The closer to the virtual axial plane (on the virtual axial plane Pax), the stronger the wedge effect occurs. Therefore, the oil pressure of the oil film in that linear region is more effectively increased, the protrusion 9 is completely separated from the seal sliding surface 8, the oil film in that linear region is thickened, and the protrusion 9 and the seal sliding surface are It becomes easy to bring the lubrication state between the surface 8 and the surface 8 into a fluid lubrication state.

Here, if there is an oil film that completely separates the protrusion 9 and the seal sliding surface 8, a fluid lubrication state will be created in which the seal sliding surface 8 slides on the protrusion 9 without directly contacting it. By maintaining such an oil film between each protrusion 9 and the seal sliding surface 8, a state of fluid lubrication can be achieved between the seal lip 7 and the seal sliding surface 8.

In order to easily realize the fluid lubrication state, it is better to set the tension force of the seal lip 7 based on the interference between the seal lip 7 and the seal sliding surface 8 as weak as possible. For this reason, the waist portion of the seal lip 7 that causes bending deformation toward the outside is formed as thin as possible.

Furthermore, the smaller the arithmetic mean roughness Ra of the seal sliding surface 8, the smaller the thickness of the oil film required to achieve fluid lubrication. Therefore, the arithmetic mean roughness Ra of the seal sliding surface 8 is set to be 1.0 μm or less. Specifically, the arithmetic mean roughness Ra of the seal sliding surface 8 is set to 0.01 μm or more and 1.0 μm or less.

Furthermore, the smaller the interval d between circumferentially adjacent protrusions 9, that is, the greater the number of protrusions 9, the more the seal sliding surface 8 rotates one rotation relative to the seal lip 7 in the circumferential direction. The number of times the protrusions 9 pass through increases, and the oil film is kept continuous over the entire circumference of the seal sliding surface 8, making it easier for the wedge effect between each protrusion 9 to occur without interruption. , it becomes easier to maintain fluid lubrication.

Furthermore, the larger the radius of curvature _R1 of the protrusion 9, the more likely the wedge effect will occur. Regarding torque at a temperature of 80° C., which is close to the actual usage environment of the sealed ball bearing 100, torque measurement tests were conducted on the sealed ball bearing 100 with various radius of curvature _R1 of the projections 9. When R ₁ was 1.5 mm, the torque was the minimum value compared to when the radius of curvature R ₁ was 0.2 mm or 0.8 mm. Also when the radius of curvature _R1 of the protrusion 9 was 1.8 mm, the torque was smaller than when it was 0.2 mm or 0.8 mm. In order to easily generate a wedge effect and realize low torque, the radius of curvature _R1 of the protrusion 9 is preferably 1.5 mm or more and 2.0 mm or less, taking into account tolerances.

If the oil film thickness between the protrusion 9 and the seal sliding surface 8 is too thin, the coefficient of friction μ will increase; The oil film thickness may be appropriately set to exceed the maximum height roughness Rz. Here, the maximum height roughness Rz refers to the maximum height roughness specified by JIS standard B0601:2013.

Usually, the value of the maximum height roughness Rz of the seal sliding surface 8 is approximately four times the arithmetic mean roughness Ra of the seal sliding surface 8. That is, when the arithmetic mean roughness Ra of the seal sliding surface 8 is 1.0 μm, it is considered that a fluid lubrication state occurs if the oil film thickness exceeds 4.0 μm.

The inventor of the present application has determined that the height h of the protrusions 9 is 0.07 mm, the distance d between the protrusions 9 is 0.3 mm or more and 2.6 mm or less, and the circumferential width w of the protrusions 9 is 0.2 mm or more and 1.0 mm. The theoretical oil film thickness between the protrusion 9 and the seal sliding surface 8 was calculated below and within the range of the radius of curvature R1 of the protrusion ₉ from 0.15 mm to less than 2.0 mm. The calculation was performed using a speed at which the seal sliding surface 8 rotates in the circumferential direction relative to the seal lip 7 (circumferential speed) in a range of 0.02 to 20.2 m/s. Further, the calculation was performed assuming that CVTF, which lubricates CVT pulleys and belts, is used as the lubricating oil, and the oil temperature is 30°C and the oil temperature is 120°C. As a result, under these conditions of use, in the lubrication region diagram for line contact based on the viscosity parameter gV and elastic parameter gE, which are dimensionless numbers determined by Greenwood-Johnson, it is calculated that the isoviscous-rigid region ( It was found that the lubrication mode was distributed in either the RI mode) or the isoviscosity-elastic body region (EI mode, soft EHL). Furthermore, the inventor calculated the relationship between the distance d between the protrusions 9 and the theoretical oil film thickness between the protrusions 9 and the seal sliding surface 8, and also determined the relationship between the radius of curvature _R1 of the protrusions 9 and the protrusions 9 and the seal sliding surface 8. The relationship with the theoretical oil film thickness between the seal sliding surfaces 8 was calculated. For the theoretical oil film thickness, Martin's minimum film thickness calculation formula was used in the RI mode, and Herrebrug's minimum film thickness calculation formula was used in the EI mode.

As a result of the above calculation, it was found that the oil film tends to become thicker when the distance d between the protrusions 9 is smaller than 2.6 mm. Therefore, the distance d between the protrusions 9 is preferably set to 2.6 mm or less. Note that if the distance d between the protrusions 9 is less than 0.3 mm, it will be difficult to mold the seal lip 7 with a mold, so it is preferable to set the distance d to 0.3 mm or more.

Furthermore, it has been found that as the radius of curvature _R1 of the protrusion 9 increases, the theoretical oil film thickness also tends to increase. It was calculated that when the radius of curvature _R1 of the protrusion 9 was 0.15 mm or more, it was possible to form an oil film of 5 μm or more. When setting the height h of the protrusion 9 to 0.07 mm, it is preferable to set the radius of curvature R ₁ of the protrusion 9 to 0.15 mm or more and less than 2.0 mm, considering that it will be molded with a mold. In addition, when setting the height h of the protrusion 9 to 0.07 mm, the circumferential width w of the protrusion 9 depends on the radius of curvature _R1 , so the circumferential width w of the protrusion 9 is set to 0.2 mm or more and 1.0 mm or less. It is preferable to set it to .

Based on the calculation results so far, in the sealed ball bearing 100, the height h of the projections 9 is set to 0.07 mm as shown in FIGS. 2 to 4, and the distance d between the projections 9 is set to 0.3 mm or more. 2.6 mm or less, the circumferential width w of the protrusion 9 is set to 0.2 mm or more and 1.0 mm or less, and the radius of curvature _R1 of the protrusion 9 is set to 0.15 mm or more and less than 2.0 mm. As a result, the plurality of protrusions 9 connect to the seal lip in a manner that allows fluid lubrication between the seal lip 7 and the seal sliding surface 8 at circumferential speeds of 0.2 m/s or more of the seal sliding surface 8. 7 is formed. Note that when the circumferential speed is less than 0.2 m/s, a state of boundary lubrication exists between the protrusion 9 and the seal sliding surface 8. Since the circumferential speed of 0.2 m/s is a speed quickly reached from a stop in the vehicle transmission, it is possible to maintain fluid lubrication between the seal lip 7 and the seal sliding surface 8 during almost the entire bearing operation time. can.

Further, whether or not a fluid lubrication state can be realized may be examined by theoretical calculation to determine whether the oil film parameter Λ>3 is satisfied. The oil film parameter Λ is determined by Equation 2 below.

The composite roughness R _rms [μm] is determined by the following equation 3.

The minimum film thickness h _min [μm] is determined by a formula depending on the lubrication mode. For example, in the EI mode, it is determined by Herrebrug's minimum film thickness calculation formula shown in Equation 4 below.

The viscosity η ₀ [Pa·s] of the fluid under atmospheric pressure is determined by Equation 5 below.

The average speed u [m/s] is determined by Equation 6 below.

In the sealed ball bearing 100, the seal member 6 is fixed to the outer ring 3 which is stationary with respect to the inner ring 1, so u ₂ =0. If the rotational speed of the seal sliding surface 8 per minute due to the rotation of the inner ring 1 relative to the outer ring 3 is N [min ^-1 ], then converting equation 6 to equation 7 below, u[m/ s] can be obtained.

The equivalent radius of curvature R[m] is determined by Equation 8 below.

The load W [N/m] per unit length is determined by the following equation 9.

The elasticity parameter gE is determined by Equation 10 below.

The equivalent modulus of longitudinal elasticity E' is determined by Equation 11 below.

The examination using the oil film parameter Λ will be illustrated under several conditions. The specifications of the inner ring 1 and lubricating oil assumed under each condition of this study are as follows.
・Radius of inner ring seal sliding surface: 32.135mm
・Young's modulus of inner ring: 210GPa
・Inner ring Poisson's ratio: 0.3
・Kinematic viscosity of lubricating oil (40°C): 26mm ² /s
・Kinematic viscosity of lubricating oil (100°C): 6.22mm ² /s
・Density of lubricating oil: 840kg/ ^m3
・Pressure coefficient of lubricating oil viscosity: 9.6×10 ^-9 Pa ^-1
・Protrusion pitch angle: 2°
・Tension force all around the seal lip: 1N
・Young's modulus of seal lip: 3.28× ^10-3 GPa
・Poisson's ratio of seal lip: 0.4999

The protrusion specifications, operating conditions, and minimum film thickness under each condition of this study are shown in Table 1 below, and the calculation results of the lubrication mode under each condition are shown in FIG.

If the oil film parameter Λ>3 is satisfied at the minimum film thickness of Conditions 1 to 7 above, it can be considered that fluid lubrication is present. Therefore, the composite roughness R _rms will be considered. In this study, the arithmetic mean roughness Ra of the projections 9 is assumed to be 0.11 μm, which is a typical value obtained by vulcanization molding. When the arithmetic mean roughness Ra of the seal sliding surface 8 is 1.0 μm, the composite roughness R _rms determined by Equation 3 is 1.26 μm. When the combined roughness R _rms is 1.26 μm under Conditions 1 to 5, the oil film parameter Λ determined by Equation 2 is 3.03 or more. When the combined roughness R _rms is 1.26 μm under Conditions 6 and 7, the oil film parameter Λ determined by Equation 2 is 0.78 under Condition 6 and 0.66 under Condition 7. That is, it can be seen that under conditions 1 to 5, the seal lip 7 and the seal sliding surface 8 are in a state of fluid lubrication, but under conditions 6 to 7, they are not in a state of fluid lubrication. For example, if the arithmetic mean roughness Ra of the seal sliding surface 8 is set to 0.238 μm or less, the oil film parameter Λ can be set to 3 or more even under condition 6.

This sealed ball bearing 100 is as described above, and the seal member 6 prevents foreign matter having a particle size that would adversely affect the bearing life from entering the bearing internal space 2, while the seal lip 7 and the seal slide The coefficient of friction μ of sliding between the moving surfaces 8 is reduced to the maximum by fluid lubrication, which in turn significantly reduces the sealing torque and significantly reduces the bearing rotational torque (see Figures 1 and 3). ).

Furthermore, the sealed ball bearing 100 has a circumferential speed of the seal sliding surface 8 (for example, 30 m/s (above), it is possible to create a fluid lubrication state without direct contact between the seal lip 7 and the seal sliding surface 8, which substantially eliminates wear of the seal lip 7 and also reduces the heat generation described above. It can be suppressed. Therefore, the sealed ball bearing 100 can meet demands for high-speed operation, which was previously unachievable.

Furthermore, in the sealed ball bearing 100, the oil permeability between the bearing internal space 2 and the outside is improved by the oil passage 14 between the seal lip 7 and the seal sliding surface 8, so that the oil passage 14 is added to the bearing internal space 2 as an initial lubricant. Since the amount of grease 10 to be sealed is reduced to 1 to 10% of the total space volume of the bearing, the agitation resistance of the grease 10 agitated by the balls 4 and retainer 5 during rotation of the bearing can be suppressed.

Furthermore, in the sealed ball bearing 100, since the arithmetic mean roughness Ra of the seal sliding surface 8 is 1.0 μm or less, it is possible to suppress the value of the composite roughness R _rms and realize a good fluid lubrication state. .

Therefore, the sealed ball bearing 100 can achieve low torque by minimizing the sealing torque due to a good fluid lubrication state and suppressing the stirring resistance of the grease 10. The results of evaluating the relationship between the amount of grease 10 sealed in the bearing internal space 2 and the arithmetic mean roughness Ra of the seal sliding surface 8 are shown in Table 2 below.

The evaluation in Table 2 is an evaluation of the low torque performance of the sealed ball bearing 100 divided into four stages. The best reduction performance was obtained when the amount of grease 10 was 1 to 10% of the total bearing volume and the arithmetic mean roughness Ra of seal sliding surface 8 was 1.0 μm or less. When the amount of grease 10 exceeded 10% of the total bearing volume, there was a limit to low torque performance even if the arithmetic mean roughness Ra of the seal sliding surface 8 was set to 1.0 μm or less. This is considered to be because the reduction in seal torque has reached its limit and the stirring resistance of the grease 10 becomes the main cause of torque.

In addition, in the sealed ball bearing 100, since the protrusion 9 is formed in an R shape, even if the protrusion 9 is rubbed against the seal sliding surface 8 when attaching the seal member 6 to the outer ring 3, the protrusion 9 will not move in the circumferential direction. There is no risk of it bending, and there is no risk of damaging seal torque reduction performance during installation. For example, if the protrusions are made into a pointed shape, it is difficult to know in which direction in the circumferential direction the many protrusions that rub against the seal sliding surface will bend when the seal member is installed, and in the direction of rotation relative to the seal sliding surface. It is extremely difficult to mount so that all the protrusions are bent towards creating a suitable wedge-shaped gap. If the protrusion is bent in an inappropriate direction, a satisfactory wedge effect cannot be obtained, which impairs sealing torque reduction performance.

In addition, in the sealed ball bearing 100, the oil passage 14 is formed by the protrusion 9 of the seal lip 7, so the seal sliding surface 8 can have the same cross-sectional shape all around the circumferential direction, and the bearing ring is sealed. Forming the sliding surface 8 is easy.

FIG. 6 shows an example in which the sealed ball bearing according to the present invention is used as a rolling bearing that supports a rotating part of a vehicle transmission. The illustrated transmission is a multi-stage transmission that changes the gear ratio in stages, and the above-mentioned sealed ball bearings rotatably support the rotating parts (for example, the input shaft S1 and the output shaft S2). A ball bearing 100 is provided.

The illustrated transmission includes an input shaft S1 to which engine rotation is input, an output shaft S2 provided parallel to the input shaft S1, and a plurality of gear trains G1 to G4 that transmit rotation from the input shaft S1 to the output shaft S2. and a clutch (not shown) installed between each gear train G1 to G4 and the input shaft S1 or the output shaft S2, and by selectively engaging the clutch, the gear train G1 to G4 to be used can be adjusted. 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 a differential or the like. The input shaft S1 and the output shaft S2 are each rotatably supported by a sealed ball bearing 100. In addition, the illustrated transmission is lubricated by splashing or injected lubricating oil as the gears rotate, or by injecting lubricating oil from a nozzle (not shown) provided inside the housing H. Oil is applied to the sides of each sealed ball bearing 100. In each sealed ball bearing 100, low torque is achieved by reducing seal torque and grease stirring resistance, so power loss can be reduced. Particularly in recent years, there has been a strong demand for fuel efficiency in automobiles. By using the sealed ball bearing 100 according to this embodiment to support the rotating part of a transmission and reducing the power loss of the transmission, it is possible to meet the demand for fuel efficiency of automobiles.

In the above-described embodiment, the protrusion has an R-shape, but the protrusion has a wedge effect that allows a state of fluid lubrication when the circumferential speed relative to the seal sliding surface is above a certain level. For example, a chamfered shape such as an R chamfer or a C chamfer can be adopted.

Further, in the above-described embodiment, an example was shown in which the protrusions were arranged uniformly in the circumferential direction, but it is also possible to arrange the protrusions non-uniformly.

Furthermore, in the above-described embodiments, the sealing member is made of a metal core and a vulcanized rubber material, but the present invention can also be applied to a sealing member made of a single material. In this case, it is only necessary that a required tightening margin can be set for the seal lip, and for example, a rubber material or a resin material can be used as the material of the seal member.

Further, in the above-described embodiment, the seal lip is a radial lip, but the present invention can also be applied to an axial lip. Here, the axial lip is a seal lip that performs a sealing action with a seal sliding surface along the radial direction or a seal sliding surface having an acute angle of less than 45° with respect to the radial direction, and is Refers to something that has an axial interference between it and the moving surface.

Further, in the above-described embodiments, an inner ring rotation and a radial bearing are illustrated, but the present invention can also be applied to an outer ring rotation and a thrust bearing.

The embodiments disclosed this time should be considered to be illustrative in all respects and not restrictive. Therefore, the scope of the present invention is indicated by the claims rather than the above description, and it is intended that all changes within the meaning and range equivalent to the claims are included.

1 Inner ring 2 Bearing internal space 3 Outer ring 4 Balls 5 Cage 6 Seal member 7 Seal lip 8 Seal sliding surface 9 Projection 10 Grease 14 Oil passage 100 Sealed ball bearing 200 Rotating shaft (rotating part)
S1 Input shaft (rotating part)
S2 Output shaft (rotating part)

Claims (2)

Inner circle and
an outer ring forming an annular bearing internal space between the inner ring;
a plurality of balls interposed between the inner ring and the outer ring;
a holder that holds the plurality of balls;
a sealing member that separates the bearing internal space from the outside;
a seal lip provided on the seal member;
a seal sliding surface that slides circumferentially relative to the seal lip;
a protrusion formed at at least one location in the circumferential direction of the seal lip and creating an oil passage between the seal sliding surface and the seal lip that communicates between the inner space and the outside of the bearing;
Grease sealed in the bearing internal space;
Equipped with
The protrusion is formed on the seal lip in a manner that allows fluid lubrication between the seal lip and the seal sliding surface,
The amount of the grease sealed as an initial lubricant before the bearing starts to be used is 1 to 10% of the total space volume of the bearing,
A ball bearing with a seal that is used to lubricate a transmission or differential of a vehicle, and in which lubricating oil, which is circulated by an oil pump and filtered by an oil filter provided in the circulation path, is supplied from the outside while the bearing is rotating. The sealed ball bearing according to claim 1, wherein the arithmetic mean roughness Ra of the seal sliding surface is 1.0 μm or less.

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