JPH11210766A - Three point contact ball bearing - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce an eccentric wear and enlarge an axial rigidity by making the carvature radius of the track groove of a first track race having a single arc shape within the specific range of the ball diameter of a rolling body. SOLUTION: In a three point contact ball bearing B, the track groove surface 13a of an outer race 13 receiving a static load is made by a single arc section shape consisting of a radius Re, while the section shape of the track groove surface 12a of an inner race 12 receiving a rotary load is made by a Gothic arch shape contacted to a ball 14 by a rest angle β. The carvature (Re/Da) of the track groove of a first track race having the single arc shape is made into a range from 50.3% to 53.3% of a ball diameter Da. In this case, it is further favorable to make into the range from 50.5% to 50.0% of the ball diameter Da. Then, the eccentric wear by a spin can be reduced and the axial rigidity can be enlarged than a deep groove ball bearing.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a three-point contact ball bearing which is used under a radial load and has excellent axial positioning performance.

[0002]

2. Description of the Related Art In order to suppress axial displacement using a rolling bearing, an angular ball bearing or a tapered roller bearing is usually used. However, since these bearings can support only one axial load by themselves, it is necessary to use two or more of these bearings or adopt double-row bearings in order to suppress displacement due to axial loads in both directions. There is. Therefore, a certain space is required in the axial direction, and it is difficult to design a compact and lightweight mechanical device.

As shown in FIG. 1, a single row deep groove ball bearing A
Can support axial loads in both directions, but the axial displacements in the axial direction are large because the inner and outer rings 2 and 3 have a single circular cross section.

In order to suppress such axial displacement, a four-point contact ball bearing C as shown in FIG. 2 is sometimes used. In the four-point contact ball bearing C, the raceways of the inner ring 6 and the outer race 7 are opposed to each other, and when the ball 8 is pressed against the raceway, the line connecting the center of the ball 8 and the contact point is the center line of the bearing. Take an angle β (rest angle) with respect to it. In the four-point contact ball bearing C, the ball 8 always has the inner and outer rings 6, regardless of the load condition.
Since it is in contact with 7 at a certain angle, it is compact, has high axial rigidity, and has excellent axial positioning performance.

[0005]

However, when a radial load is applied to the four-point contact ball bearing C, the inner and outer rings 6, 7
The slip due to the spin motion at the contact point between the ball and the ball 8 increases. Since the position of the contact point that receives the maximum load changes every moment in the ball 8 and the bearing ring that receives the rotational load, it is considered that the possibility that a specific portion of the member surface is damaged by the slip is low. However, in a bearing ring which receives a static load, the position where the maximum load is received is constant, and therefore, there are portions repeatedly exposed to a large sliding motion on the raceway surface. In this portion, the risk of seizures increases due to a temperature rise due to heat generation at the contact point. Further, uneven running of the raceway surface at a portion where the slip is large may hinder the operation of the bearing.

[0006] In the case of a bearing for an electromagnetic clutch, a double-row ball bearing is used because a belt load acts on a portion displaced from the center of the shaft surface. The ball bearing and the pulley are displaced in the axial direction in order to make the electromagnetic clutch as small as possible. The radial load generated by the tension of the V-belt acts on the ball bearing as a moment load with inclination because the pulley and the ball bearing are displaced in the axial direction. For this reason, a relative inclination occurs between the outer ring and the inner ring of the ball bearing, and the rotor fitted to the outer ring undergoes an axial displacement with an inclination. When a double row ball bearing is used, the inclination of the rotor is reduced, and the rotor 35 and the armature 37 shown in FIG. 9 are brought into contact with each other to prevent the electromagnetic clutch from being damaged. However, double-row ball bearings have a large width (wide width), which makes it difficult to reduce the size of the electromagnetic clutch, and also makes it difficult to reduce costs associated with the reduction in size of the electromagnetic clutch.

SUMMARY OF THE INVENTION The present invention has been made to solve the above problems, and has a small slip due to spin motion at a contact point between a race and a ball, can reduce uneven wear, and has a high axial rigidity. 3 which can be increased
It is an object to provide a point contact ball bearing.

[0008]

According to the present invention, there is provided a three-point contact ball bearing comprising: a first race having a single arc-shaped raceway;
A first raceway having a single circular arc shape in a three-point contact ball bearing having an axially parallel cross section not having a single circular arc shape and having a second raceway having a raceway contacting the rolling element at two points. The raceway radius of curvature of the wheel is not less than 50.3% and not more than 53.3% of the ball diameter of the rolling element.

In the present invention, the raceway curvature (Re / Da) of the first raceway having a single arc shape is determined by calculating the diameter of the raceway of the ball having a diameter Da of five.
The range is from 0.3% to 53.3%. In this case, it is more preferable to set the range of 50.5% to 53.0% of the ball diameter Da. In this manner, uneven wear due to spin can be reduced, and axial rigidity can be increased as compared with a deep groove ball bearing.

It is preferable that the surface of at least one of the first and second races and the rolling elements is subjected to nitriding. This increases the wear resistance of the component.

The bearing preferably supports a radial load and is used for the purpose of positioning in the axial direction.

Further, it is desirable that the bearing is used in a position where the load of the pulley shaft on which the belt is applied is supported in a continuously variable transmission of an automobile.

Further, the bearing may be used for positioning the shaft in a toroidal type continuously variable transmission of an automobile.

In the case of a bearing operated under high-speed and high-load conditions, such as the above-described bearing for a continuously variable transmission of an automobile, a single circular arc shape is used in order to reduce uneven wear due to spin of a bearing ring which receives a static load. It is preferable that the first bearing ring has a static load and the second bearing ring that contacts the rolling element at two points receives a rotating load.

It is desirable that the bearing be used in a portion of the electromagnetic clutch device that rotatably supports a rotating rotor having a friction surface.

Further, it is desirable that the rolling elements and the second race (the inner race) are made of ceramics.

[0017]

In the three-point contact ball bearing according to the present invention, the first
Since the raceway cross-sectional shape of the raceway is a single arc,
As compared to point contact ball bearings, when a radial load is applied, spin sliding is reduced, and uneven wear of the raceway surface is prevented.

Also, in the three-point contact ball bearing of the present invention,
The orbital surfaces of the orbital rings face each other, and the ball always contacts the orbital surface at a large contact angle, so that the axial load in both directions exhibits higher axial rigidity than the deep groove ball bearing.

Further, if at least one of the first and second races and the rolling elements is made of ceramics, the ceramics have a smaller coefficient of linear expansion than a steel material, and when the temperature changes, the fluctuation of the bearing gap changes. Since the number of the images decreases, seizure hardly occurs. Further, when the rolling elements are made of ceramics, the axial displacement of the outer ring can be reduced with respect to the curvature of the same groove, so that the degree of freedom in design is increased.

[0020]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Various preferred embodiments of the present invention will be described below with reference to the accompanying drawings and tables. (Example 1) Each characteristic of a deep groove ball bearing A, a three-point contact ball bearing B, and a four-point contact ball bearing C having the same dimensions shown in Table 1 was simulated and analyzed by a computer. This analysis method includes "performance analysis of four-point contact ball bearings" (Taniguchi, Aramaki,
Masada: The method described in Tribology Conference of the Japan Society of Tribology, Tribology Conference, Spring 1996, Tokyo).

Here, it is assumed that the bearing to be analyzed applies a static load to the outer ring and is used for rotation of the inner ring. FIG.
As shown in the figure, in the three-point contact ball bearing B of the present invention, the raceway groove surface 13a of the outer ring 13 receiving a static load has a single arc-shaped cross-sectional shape having a radius Re, while the inner ring 12 receiving a rotational load. The cross-sectional shape of the raceway groove surface 12a is a Gothic arch shape that comes into contact with the ball 14 at a rest angle β. In this embodiment, the rest angle β of the ball 14 on the raceway groove surface 12a of the inner race is set to 30 °. On the other hand, the four-point contact ball bearing C of the comparative example had a gothic arch shape in which both the inner and outer rings 6 and 7 contact the ball 8 at a rest angle β. The surface roughness of the raceway grooves of the bearings A, B, and C was the same.

Table 2 shows the operating conditions used for each simulation analysis. Analysis 1 for each of the above bearings A, B, C
In Example 2, a pure radial load was applied, and in Analysis 2, a pure axial load was applied. The rotation speed of the inner ring is 1000 for both analysis 1 and 2.
0 rpm was set.

FIG. 6 shows the results of the simulation analysis.
The horizontal axis of the graph in FIG. 6 indicates the maximum value of the PV value at the contact point between the ball and the outer raceway surface according to the analysis 1 under the pure radial load. As shown in FIG. 5, the contact point between the ball and the bearing ring is actually represented by an elliptical area based on the elastic contact theory of Hertz due to the elastic deformation of the surface (the hatched area in the figure).
Becomes That is, in the case of point contact between quadratic surfaces, the mutual contact portion has an elliptical shape due to geometrical conditions, and when the rolling element load is Q, the maximum contact pressure Pmax is Q
Proportional to ^1/3, its elastic displacement δ is proportional to Q ^2/3.

The PV value is the product of the surface pressure P in the contact surface and the slip speed V. In the analyzes 1 and 2, the PV value is calculated in the contact surface between each ball and the outer ring, and the maximum value is shown here. The PV value is widely used as an index of heat generation and wear due to slip. The μPV value obtained by multiplying this PV value by the coefficient of sliding friction μ between surfaces is given by
Gives friction loss due to sliding per unit time. In the case of a four-point contact ball bearing C using bearing steel, if the PV value in this analysis exceeds 1.5 to 2.0 GPam / s, local wear may occur on the bearing ring that receives a static load. Is obtained from the experimental results.

The radius of curvature of the raceway grooves of the inner and outer rings is 50.
The analysis was performed for each of 5% and 52%. In each case, the outer ring maximum PV value of the three-point contact ball bearing B according to the present invention under a pure radial load was smaller than that of the four-point contact ball bearing C, and was substantially equal to that of the deep groove ball bearing A. The outer ring maximum PV value of the four-point contact ball bearing C is:
It exceeds 5 GPam / s, and there is a risk of causing local wear on the outer raceway surface. However, the PV value of the three-point contact ball bearing B is only about 1.1 GPam / s even when the radius of curvature of the raceway groove is 50.5%, so that the possibility of uneven wear of the outer ring 13 is small.

The vertical axis of the graphs shown in FIGS. 6 and 8 is the axial displacement of the bearing according to Analysis 2 under pure axial load. Here, the axial displacement includes an axial movement amount due to the internal clearance of the bearing, and a displacement due to elastic deformation of the ball or the track due to a load. As is clear from the figure, the axial displacement of the three-point contact ball bearing B of the present invention is 4
It is larger than the point contact ball bearing C, but 40 to 50% smaller than the deep groove ball bearing A. Further, the three-point contact ball bearing B of the present invention shows that the positioning effect is higher than that of the deep groove ball bearing A.

In the above analysis, the number of balls in each of the bearings A, B, and C is the same as that of a general deep groove ball bearing 6308. However, a three-point contact ball bearing B or a four-point contact ball bearing C of a type in which the ball 14 and the inner ring 12 contact at two points may have an inner ring divided into two by a plane perpendicular to the central axis. Such a three-point contact ball bearing B or a four-point contact ball bearing C is easier to assemble than a general deep groove ball bearing A of the same size, and therefore it is easy to increase the number of balls. As the number of balls increases, the load per ball decreases, and the surface pressure at the point of contact with the race and the amount of elastic deformation of the bearing decrease. Also in the three-point contact ball bearing B of the present invention, by increasing the number of balls, the PV value under the radial load and the axial displacement under the axial load are reduced as compared with those shown in FIG. Can be provided. (Embodiment 2) Using the same analysis method as in the first embodiment, the effect of the outer ring groove radius of curvature on the bearing characteristics of the above-mentioned three-point contact ball bearing B was examined. Groove radius of inner ring 12 is ball diameter Da
Of 52%. The specifications of the other bearings are as shown in Table 1.

FIG. 7 shows the results of analysis 1 in Table 2. The abscissa indicates the radius of curvature Re of the outer ring as a ratio to the ball diameter Da. The vertical axis indicates the maximum value of the PV value at the contact point between the ball and the outer raceway surface during operation under a pure radial load. The larger the radius of curvature of the groove, the greater the maximum PV of the outer ring
The values are small, indicating a low risk of excessive heat generation and wear. As described above, when the PV value exceeds 1.5 GPam / s, local wear of the outer ring 13 that receives a static load may occur. Therefore, as shown in FIG.
In order to prevent uneven wear of the outer ring, only the outer ring has a radius of curvature Re.
Must be 50.3% or more of the ball diameter Da. In order to suppress the PV value of the outer ring raceway to 1.0 GPam / s or less in consideration of manufacturing or operation errors, it is more preferable that the outer ring groove has a curvature radius Re of at least 50.5% of the ball diameter Da. Good.

FIG. 8 shows the results of analysis 2 in Table 2. Here, the vertical axis represents the axial displacement of the bearing under a pure axial load. The smaller the radius of curvature Re of the outer ring 13 is, the smaller the axial displacement is, indicating that the positioning performance is excellent. Here, the broken line on the graph indicates the axial displacement of the standard deep groove ball bearing 6308. As is apparent from this figure, in the three-point contact ball bearing B according to the present invention, when the outer ring groove radius of curvature Re is 53% or less of the ball diameter Da, the axial displacement is smaller than that of the standard deep groove ball bearing 6308. There is an advantage that it becomes.

In the above embodiment, the effect of the three-point contact ball bearing of the present invention has been shown in the case where the outer ring receives a static load and the inner ring rotates. In the case of an inner ring stationary load and an outer ring rotational load, the inner race has an inner race whose cross section is a single circular arc and an outer race having two races whose cross section is not a single circular arc.
By using the point contact ball bearing, there is obtained an effect that slip due to spin motion is small at a contact point between the bearing ring and the ball, uneven wear can be reduced, and axial rigidity can be increased.

In the three-point contact ball bearing B according to the present invention, the slip is large at the contact point between the bearing ring and the ball which receives the rotational load, but the position where the maximum load is applied changes on the surface of the raceway and the ball. However, the possibility of uneven wear of a specific part is low. However, it is conceivable that the surfaces of the balls and the raceway are worn out as a whole during the long-term operation. Therefore, preferably, the surface of the steel material is subjected to a nitriding treatment to enhance wear resistance. Also, for the bearing ring subjected to a static load, it is preferable to use the same wear-resistant material from the viewpoint of preventing uneven wear although the slip is small and from the viewpoint of preventing wear due to dust in the lubricating oil. .

Such a three-point contact ball bearing B is used for positioning in the axial direction, but can be used under a large radial load condition where it is difficult to apply the four-point contact ball bearing C. . The three-point contact ball bearing B of the present invention is effective not only for applications in which a radial load is supported during a steady operation but also applications in which a large radial load is temporarily applied at the time of starting or the like.

An example of such an application is a bearing for supporting a load of a pulley shaft on which a belt is applied in a continuously variable transmission using a belt made of metal or the like, which is used in an automobile or the like. Further, in a toroidal-type continuously variable transmission used for automobiles and the like, it can be used as a bearing for positioning a shaft while receiving a radial load.

[0034]

[Table 1]

[0035]

[Table 2]

(Embodiment 3) Next, a case where the three-point contact ball bearing of the present invention is used for an electromagnetic clutch will be described as a third embodiment with reference to FIGS.

The electromagnetic clutch 31 transmits or shuts off the rotational power generated in the vehicle running engine to the compressor of the refrigeration cycle. Since the belt load acts on the electromagnetic clutch bearing 36 at a position where the belt load is displaced from the center of the shaft surface, it is required to minimize the amount of axial displacement accompanied by the inclination.

The electromagnetic clutch 31 includes a stator 32 fixed to the compressor housing 40 and the stator 3
2, a pulley 34 around which a multi-stage V-belt 41 for transmitting the rotational power of the engine is wound, a rotor 35 around which the pulley 34 is mounted, and a rotatable rotor 35 Three-point contact ball bearing 36, an armature 37 attracted to the rotor 35 by a magnetic force generated by the electromagnetic coil 33, and a rotary power of the armature 37 (not shown).
And a compressor shaft 39 composed of a single or a plurality of members to be transmitted to the armature.

The rotor 35 is made of a magnetic metal (for example, iron), has a substantially U-shaped cross section, and has a recess 35 a facing the opposite side of the armature 37. This recess 35
The stator 32 is accommodated in a.

A single-row three-point contact ball bearing 36 is fitted between the rotor 35 and the compressor housing 40. The rotor 35 is rotatably supported by the ball bearing 36 with respect to the compressor housing 40. The outer ring 13 of the ball bearing 36 fits and contacts the inner circumference of the rotor 35, and the inner ring 12 of the ball bearing 36 fits and contacts the outer circumference of the hub of the compressor housing 40.

On the outer periphery of the rotor 35, a multi-stage V belt 4
The pulley 34 around which the belt 1 is wound is fixed by a joining technique such as welding. The pulley 34 always receives the rotation of the engine via the V-belt 41 and the fixed rotor 35
It rotates with it.

The ball bearing 36 and the pulley 34 are usually displaced in the axial direction in order to make the electromagnetic clutch 31 as compact as possible. The radial load generated by the tension of the V-belt 41 acts on the ball bearing 36 as a moment load with a tilt because the pulley 34 and the ball bearing 36 are displaced in the axial direction. Therefore, a relative inclination occurs between the outer ring 13 and the inner ring 12 of the ball bearing 36, and an axial displacement accompanied by an inclination occurs in the rotor 35 fitted to the outer ring 13.

Here, the design is such that a predetermined gap g (for example, 0.5 mm) exists between the rotor 35 and the armature 37 when the energization of the electromagnetic coil 33 is stopped. The gap g is determined in consideration of dimensional tolerances of various members or mounting errors. Here, if the gap g is too large, the magnetic force by the electromagnetic coil 33 must also be increased. Therefore, the gap g is set to a minimum value that allows for these.

In order to minimize the axial displacement caused by the inclination of the bearing 36 for the above-described reason, a double-row ball bearing is conventionally used as a bearing for supporting the rotor 35 of the electromagnetic clutch. When the double row ball bearings are used, the inclination of the rotor 35 is reduced, and the rotor 35 and the armature 37 are in contact with each other, thereby preventing the electromagnetic clutch from being damaged. However, double-row ball bearings have a large width (wide width), which makes it difficult to reduce the size of the electromagnetic clutch, and also makes it difficult to reduce costs associated with the reduction in size of the electromagnetic clutch.

On the other hand, the three-point contact ball bearing 3 of the present embodiment
In 6, the maximum displacement in the axial direction can be kept small, and the electromagnetic clutch 31 can be made compact and cost-effective. The results obtained by performing a computer simulation calculation of the maximum axial displacement of the outer diameter surface of the outer race 13 using the calculation conditions shown in Table 3 are shown.

In Table 3, the radius of curvature of the raceway ring 12 contacting at two points is fixed to 52% of the ball diameter, and the raceway ring 1 having a single arc is formed.
The radius of curvature of each of the three grooves was changed. The calculation result is shown in FIG.
Shown in FIG. 10 shows the result of examining the relationship between the outer diameter of the outer ring and the radius of curvature Re of the outer ring with respect to the ball diameter Da (percentage) on the horizontal axis, and the maximum axial displacement (mm) of the outer ring on the vertical axis. FIG. 6 is a characteristic diagram. In the figure, a characteristic line A represents a correlation between the two in the three-point contact ball bearing. Incidentally, in the conventional double-row ball bearing, the maximum axial displacement is about 0.09 mm. By improving the component accuracy and the assembly accuracy, the maximum axial displacement can be reduced to about 0.17 mm or less, which is about 90% larger than in the past. In order to satisfy this allowable value, a single arc is required. The radius of curvature Re of the race of the bearing ring must be 53.3% or less of the ball diameter Da. In addition, preferably, in order to realize compactness without changing the precision of parts and the precision of assembly, the radius of curvature of the groove of the single arc raceway is set to 51.9% or less of the ball diameter Da, and the maximum axial displacement is 0125 mm or less. And

More preferably, the maximum displacement in the axial direction is set to 0.
In order to suppress the diameter to 1 mm or less and to make it about the same as a conventional double row ball bearing, it is necessary to set the radius of curvature Re of the single circular raceway ring to 51.2% or less of the ball diameter. The minimum value of the groove curvature radius Re is theoretically 50% of the ball diameter Da. However, as described in FIG. 7, the PV value increases and the risk of abrasion is high. % Or more. Further, considering the manufacturing or operation error, the diameter of the ball
It is preferably at least 0.5%.

[0048]

[Table 3]

From the above results, the three-point contact ball bearing can suppress the amount of axial displacement even when a moment load acts like an electromagnetic clutch. In the case of a bearing for an electromagnetic clutch, the axial maximum displacement amount in the single arc shape outer ring groove even if the inner ring groove which is stationary ring into a single arc shape is rotating ring may be either because it does not change. However, in the case of a bearing for an electromagnetic clutch, since the surface pressure on the inner ring is high and separation occurs on the inner ring raceway surface, it is preferable to use a three-point contact ball bearing capable of dispersing the inner ring load in two places because the life is extended. Also, from the viewpoint of suppressing uneven wear in the electromagnetic clutch bearing, it is preferable to increase the wear resistance by performing nitriding treatment on the steel material.

From the viewpoint of increasing the rigidity and reducing the maximum axial displacement of the outer ring, it is preferable to use ceramics such as silicon nitride or silicon carbide having a higher Young's modulus than steel for the rolling elements. FIG. 11 shows the results of an examination of the relationship between the outer ring and the radius of curvature Re of the outer ring with respect to the ball diameter Da (percentage) on the horizontal axis and the maximum axial displacement (mm) of the outer ring on the vertical axis. FIG. 6 is a characteristic diagram. Characteristic line B in the figure
Shows the results using steel balls for the rolling elements, and the characteristic line C shows the results using ceramic balls (silicon nitride) for the rolling elements. As is clear from the figure, when the rolling elements are made of silicon nitride ceramic, the inclination rigidity of the bearing is further increased, so that the amount of axial displacement accompanying the inclination of the bearing 36 can be further reduced, thereby increasing the degree of freedom in design.

[0051]

According to the present invention, the first race has a single circular arc-parallel cross-section parallel to the axis, and the second race has a non-circular cross-section parallel to the axis. Touch at point 2
The first race has a groove having a radius of curvature of not less than 50.3% and not more than 53.3% of the diameter of the rolling element, thereby reducing uneven wear and reducing axial displacement. It is smaller than a ball bearing and can increase axial rigidity.

In the three-point contact ball bearing of the present invention, heat generation and wear are reduced even when a radial load is applied, while having the axial positioning function.

[Brief description of the drawings]

FIG. 1 is a sectional view showing an example of a conventional deep groove ball bearing.

FIG. 2 is a sectional view showing an example of a conventional four-point contact ball bearing.

FIG. 3 is a sectional view showing a three-point contact ball bearing according to the embodiment of the present invention.

FIG. 4 is an enlarged sectional view showing a part of the three-point contact ball bearing according to the embodiment of the present invention.

FIG. 5 is a schematic diagram for explaining elastic contact between a ball and a raceway groove surface.

FIG. 6 shows the relationship between the maximum PV value on the outer ring raceway surface under a pure radial load and the axial displacement under a pure axial load for each of a deep groove ball bearing, a three-point contact ball bearing, and a four-point contact ball bearing. FIG. 4 is a simulation analysis diagram showing a result of analyzing by using a computer.

FIG. 7 is a simulation analysis diagram showing a result of analyzing by a computer the relationship between the radius of curvature of the outer ring groove of the three-point contact ball bearing and the maximum value of the PV value on the outer ring raceway surface under pure radial load.

FIG. 8 is a simulation analysis diagram showing a result of analyzing, by a computer, a relationship between a radius of curvature of an outer ring groove of a three-point contact ball bearing and an axial displacement under a pure axial load.

FIG. 9 is a schematic sectional view showing an electromagnetic clutch using the three-point contact ball bearing according to the embodiment of the present invention.

FIG. 10 shows the results of analyzing the relationship between the maximum axial displacement of the outer surface of the outer ring and the radius of curvature of the single circular raceway ring by a computer when the three-point contact ball bearing of the present invention is applied to an electromagnetic clutch. FIG.

FIG. 11 shows steel balls and ceramic balls (silicon nitride) when the three-point contact ball bearing of the present invention is applied to an electromagnetic clutch.
The simulation analysis figure which showed the result of having analyzed by computer the difference of the relationship between the axial maximum displacement amount of the outer diameter surface of an outer ring, and the groove radius of a single circular-arc raceway.

[Explanation of symbols]

12 inner ring (second race), 12a raceway groove surface, 13
... outer ring (first race), 13a ... raceway surface, 14 ... ball (rolling element), A ... deep groove ball bearing, B ... three-point contact ball bearing,
C: 4-point contact ball bearing, β: rest angle, 31: electromagnetic clutch, 32: stator, 33: electromagnetic coil, 34 ...
Pulley, 35 ... rotor, 36 ... three-point contact ball bearing for electromagnetic clutch, 37 ... armature, 38 ... armature support part, 39 ... compressor shaft, 40 ... compressor housing (inner ring support part of bearing for electromagnetic clutch).

Continuation of the front page (72) Inventor Hiroshi Ishiguro 1-5-50 Kugenuma Shinmei, Fujisawa-shi, Kanagawa Nippon Seiko Co., Ltd.

Claims (1)

[Claims] A first orbital ring having a single arc-shaped orbit, and a second orbital ring having an orbital section whose axis-parallel cross section is not a single arc and which contacts the rolling element at two points. A three-point contact ball bearing having a single circular arc shape.
The raceway radius of curvature of the raceway of No. 50 is the ball diameter of the rolling element.
A three-point contact ball bearing having a ratio of 3% or more and 53.3% or less.