JPH07238938A - Full type ball bearing - Google Patents

Abstract

PURPOSE:To reduce a torque and elongate a life by setting a radius of curvature of an inner ring orbit larger than that of an outer ring orbit groove in a full type ball bearing in which a plurality of balls are arranged between the orbit grooves of inner and outer rings. CONSTITUTION:When heat expansion is generated and a rotary shaft 5, a ball 42, and an outer ring 41 are expanded in an axial and a radial directions, the outer ring 41 is relatively displaced to an inner ring orbit groove 51 in an axial direction so as to decrease a contact angle in respect to the ball 42 in between an orbit groove 41. The inner ring orbit groove 51 is set larger than the outer ring orbit groove 41a in its radius of curvature, so that the ball 42 is rolled while being guided by the outer ring orbit groove 41a along a sectional area shape of the inner ring orbit groove 51, brought into contact with the outer ring orbit groove 41a, and the inner ring orbit groove 51 without misalignment. Sliding between the ball 42 and the orbits 41a, 51 is suppressed, and abrasion of a luricating film on the surface of the ball is reduced. Life of a bearing is elongated. Surface pressure is reduced and torque is decreased, because a contact area is small.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a full ball bearing, and more particularly to a full ball bearing most suitable for rotatably supporting a rotary shaft under a vacuum or high temperature environment.

[0002]

2. Description of the Related Art Generally, a full ball bearing of this type is preferably used as a rolling bearing for high speed rotation such as a touchdown bearing of a magnetic bearing device, a bearing for a turbo molecular pump, and a bearing for a rotary anode X-ray tube. Is. FIG. 6 shows a configuration example in which such a full ball bearing is adopted in a general rotary anode X-ray tube. Referring to the figure, the rotary anode X-ray tube includes a cathode part 1 that emits electrons, a target plate 2 facing the cathode part 1, and a rotating shaft 5 that rotatably carries the target plate 2. ing.

The rotary shaft 5 is arranged laterally, and is rotatably supported by a bearing device composed of a pair of full ball bearings 4 and 4 arranged at intervals in the direction of the rotation center line. It is supported. This bearing device accommodates both full ball bearings 4 and 4 in a housing 3. As shown in FIG. 7, the full ball bearings 4 and 4 are composed of inner ring raceway grooves 43 and 43 formed on the outer periphery of the rotating shaft 5 and outer rings 41 and 41 concentrically arranged on the outer periphery of the rotating shaft 5. It is an integral type oblique contact ball bearing which is embodied by raceway grooves 41a, 41a and a large number of balls 42 interposed between the raceway grooves 43, 41a, 43, 41a. In addition, one outer ring 41 and the other outer ring 4
1 are axially juxtaposed to each other via collars 10 and 11.

The rotating shaft 5 and the outer rings 41, 41 are made of high speed steel (Japanese Industrial Standard: SKH4). The collars 10 and 11 are made of stainless steel (Japanese Industrial Standards: SUS440C, SUS403, SU).
It is formed in S410). In the full ball bearings 4 and 4, the outer ring raceway grooves 41a and 41a are back-to-back in shape. The balls 42, 42 are formed in a spherical shape, and on the surface thereof, ion plating of silver or other soft metal,
A solid lubricating film of gold, copper, lead or the like, or a solid lubricating film of molybdenum disulfide having a layered structure is applied.

Further, in the general full ball bearings 4 and 4,
The radius of curvature of the inner ring raceway grooves 43, 43 is determined by the outer ring raceway groove 41.
It was set the same as a and 41a. That is, balls 42, 4
When the diameter of 2 is Bd, each raceway groove 43, 43, 41
The curvature radii of a and 41a were set to (51 to 54)% × Bd. According to the above configuration, the electrons emitted from the cathode part 1 are accelerated by the high voltage and collide with the target plate 2. The colliding electrons generate X-rays due to their bremsstrahlung, or due to energy emission at the time of transition to fill the vacancy of electrons in the nucleus.

[0006]

When high-speed electrons collide with the target plate 2, high heat is generated and is conducted to the rotating shaft 5. Therefore, when the rotating anode X-ray tube is used, the temperature of the rotating shaft 5 reaches about 550 ° C. In contrast,
The outer ring 41 forming the bearings 4 and 4 has a temperature of about 350 ° C.
Is. For this reason, high-temperature heat conduction from the rotating shaft 5
The solid lubricating film applied to No. 2 is softened or melted to promote the early peeling and wear of the oil lubrication, which leads to the early life of the bearing.

Further, a temperature difference occurs between the rotary shaft 5 and a housing separate from the rotary shaft 5, so that the rotary shaft 5 extends not only in the axial direction but also in the radial direction. At this time, in order to prevent the internal clearance of the bearing from becoming negative, the materials of the collars 10 and 11 are appropriately set to provide a predetermined gap with the outer rings 41 and 41, and the outer rings 41 and 41 are fixed to the rotary shaft 5. I try to move it relative to. Therefore, although the internal clearance does not become negative due to the thermal expansion of the rotating shaft 5, the balls 4 are moved by the axial movement of the outer rings 41, 41.
There is a case where the wheels 2 and 42 are disengaged from the raceway grooves 41a and 41a of the outer rings 41 and 41, and so-called shoulder riding occurs. If this shoulder riding occurs, there is a problem that the lubricating film covering the surfaces of the balls 42, 42 peels off at the ends of the raceway grooves 41a, 41a, and as a result, the bearing life may be shortened.

Further, the rotating anode X-ray tube has a capacity of about 10,000.
0r. p. m. Of the inner and outer raceway raceways 43, 41a, 43, 41a due to the high speed rotational movement of the
The rolling of the balls 42, 42 between them cannot completely follow the inner and outer rings, and the balls 42, 42 and the inner ring raceway grooves 43, 43
Also, slippage occurs in both of the balls 42, 42 and the outer ring raceway grooves 41a, 41a. In a ball bearing in which the surfaces of the balls 42, 42 are provided with a solid lubricating film, the above-mentioned slippage leads to early wear of the lubricating film, resulting in a shorter bearing life.

The present invention has been made in view of the above problems, and at the time of high temperature and high speed, the slippage of the balls is reduced and the thermal influence on the balls is reduced, so that the rotating torque is small and the life is long. The purpose is to provide a full ball bearing.

[0010]

Generally, when the slips between the ball and the inner ring and between the ball and the outer ring are compared, the slippage between the ball and the outer ring is larger due to the higher peripheral speed. Therefore, in order to solve the above problems, the structure according to claim 1 of the present invention is a full ball bearing in which a plurality of balls are arranged between the race grooves of the inner and outer races, and the radius of curvature of the inner race groove is The full ball bearing is characterized in that it is set to have a radius of curvature larger than that of the outer raceway groove.

According to a second aspect of the present invention, there is provided a full ball bearing according to the first aspect, wherein the surface of the ball is coated with a film made of a solid lubricant.

[0012]

According to the above construction, the contact area between the inner ring raceway groove and the balls is reduced. Then, thermal expansion occurs, the rotating shaft, balls,
When the outer ring extends in the axial direction and the radial direction, the outer ring raceway groove is axially displaced relative to the inner ring raceway groove. At this time, the inner ring raceway groove is set to have a larger radius of curvature than the outer ring raceway groove, so that the balls can make rolling contact with the inner ring raceway groove along the cross-sectional shape of the inner ring raceway groove without riding on the shoulder. . The rate of heat conduction from the inner ring is also reduced.
Furthermore, since the balls are rolled and guided by the outer ring, slippage between the outer ring and the balls is reduced. Therefore, when a solid lubricating film is applied to the balls, the service life of the lubricating film is extended, which in turn extends the service life of the bearing.

[0013]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT A preferred embodiment of the present invention will now be described in detail with reference to the accompanying drawings. In the following explanation,
The same members as those of the conventional technique are designated by the same reference numerals and the description thereof is omitted. FIG. 1 is an enlarged view of a main part of a full ball bearing according to an embodiment of the present invention, and FIG. 2 is a schematic front view showing a main part of a rotary anode X-ray tube adopting the full ball bearing.

First, referring to FIG. 2, in the rotary anode X-ray tube according to this embodiment, a pair of full ball bearings 401, 40.
A rotary shaft 5 is rotatably supported by the housing 3 by means of 2. The outer rings 41, 41 are fitted to the housing 3 by a clearance fit, and are arranged so as to be axially movable with respect to the housing 3. Both ball bearings 401, 4
Reference numeral 02 denotes the same outer ring raceway grooves 41a and 41a as in the conventional example,
A large number of balls 42, 42 between the inner ring raceway grooves 51, 51 formed directly on the rotary shaft 5 are not restricted by the cage, that is, they approach (or separate) from each other.
It is an integral type deep groove type in which the interposition is allowed. As the balls 42 that can be used in this embodiment, those made of high speed steel (for example, Japanese Industrial Standard: SKH4) can be adopted, and those made of a ceramic material mainly containing silicon nitride are preferable. Is. The solid lubricant applied to the surface of the balls 42 is preferably Ag-IP (ion plating) or Pb-IP. The surface hardness of the rotating shaft 5 and the outer ring 41 is set to HRC62-66. The surface hardness of the ball 42 is HRC63-
68.

The linear expansion coefficient of the rotary shaft 5, the outer ring 41 and the balls 42 is 11.0 × 10 ^-6 . In the total ball bearings 401 and 402 of the present embodiment, the radius of curvature Pi of the inner ring raceway grooves 51 and 51 is set to Pi = (61% ± 2%) × Bd when the diameter of the balls 42 and 42 is Bd. ing.

On the other hand, the radius of curvature Po of the outer ring raceway grooves 41a, 41a is Po = (56% ± 2%) × Bd in consideration of wear of the solid lubricating film, as compared with a general bearing. It is set slightly larger. Also in this case, the radius of curvature Pi of the inner ring raceway grooves 51, 51 is set larger than the radius of curvature Po of the outer ring raceway grooves 41a, 41a. This reduces the contact area between the inner ring raceway groove 51 and the balls 42.

According to the above construction, when thermal expansion occurs and the rotary shaft 5, balls 42, 42, and outer rings 41, 41 extend in the axial and radial directions, the outer ring 41 with respect to the inner ring raceway groove 51.
Is relatively displaced in the axial direction so as to reduce the contact angle with the ball 42 arranged between the track groove 41a. At this time, the inner ring raceway groove 51 is set to have a larger radius of curvature than the outer ring raceway groove 41a (Pi> Po).
Can be in rolling contact with the outer ring raceway groove 41a and the inner ring raceway groove 51 along the cross sectional shape of the inner ring raceway groove 51 while being guided by the outer ring raceway groove 41a without riding on the shoulder. Therefore, compared with the conventional structure, slippage between the ball 42 and the raceway grooves 41a, 51 is reduced, wear of the lubricating film on the ball surface is reduced, and the life of the bearing is improved.

Further, the rate of heat conduction from the rotary shaft 5 is reduced, the temperature rise of the bearing is reduced, and the loss of the lubricating film is reduced. Further, since the contact area is small, the surface pressure is reduced, the rotational torque τ is reduced, and the wear of the lubricating film is also reduced.
As described above, according to the full ball bearing of the present embodiment, the sliding motion of the ball 42 is reduced, and in addition, the inner ring raceway groove 51 and the ball 4 are
Since the contact area with 2 becomes small, and further, the balls 42 make rolling contact with the inner ring raceway groove 51 along the cross-sectional shape of the inner ring raceway groove 51 without riding on the shoulder, thereby preventing the shoulder run-up. In addition, it is possible to prevent problems such as an increase in the rotation torque τ and peeling off of the lubricating coating, and thus it is possible to prevent the life of the full ball bearing from being shortened. In particular, since the rotation torque τ can be reduced, the spin of the balls 42 provided with the solid lubricating film can be suppressed. As a result, the lubricating film can be prevented from being damaged due to the friction with the raceway groove, and the solid lubricating life can be extended. It has the advantage of being long.

Moreover, according to the present embodiment, the rate of heat conduction from the rotary shaft 5 is reduced and the temperature rise of the bearing is reduced, so that it is less susceptible to thermal influence, and in this respect as well, the life is extended. Can contribute to. This structure is used for at least the first total ball bearing 401 closer to the target plate 2.
Has been applied to. This is because the heat effect of the rotary shaft 5 is greater in the first full ball bearing 401. However, in the present embodiment, also in the second total ball bearing 402, the radius of curvature Pi of the inner ring raceway groove 51 is set to be larger than the radius of curvature Po of the outer ring raceway groove 42a.

Next, in this embodiment, both full ball bearings 40
1 and 402, when the balls 42 are gathered in one direction along the raceway grooves 41a and 51, the circumferential clearance CC between one end and the other end of the row formed by the balls 42 has the following relational expression ( 1),
It is set to satisfy (2): CC = dm × sin {180 ° −Z × sin ⁻¹ (Bd / dm)} ... (1) 0.8 × Bd ≦ CC ≦ 2.0 × Bd (2) dm: Pitch circle diameter of balls 42 (see FIG. 3) Bd: Diameter of balls 42 (see FIG. 3) Z: Number of balls 42 Since the above configuration is adopted, the total ball bearing of this embodiment 40
In 1 and 402, the ball 4 is
It is possible to suppress the influence of the advance / delay of 2 on the rotation torque. For this reason, it is possible to prevent the balls 42 from interfering with each other, and it is possible to prevent abnormal noise and seizure from occurring. Also, the rotational torque τ of the bearing itself is reduced.

Next, in this embodiment, the load capacity of the second total ball bearing 402 is set smaller than that of the first total ball bearing 401. As described above, both full ball bearings 401,
When the target plate 2 is cantilevered via the rotary shaft 5 via the rotary shaft 5, the rotary shaft 5 is tilted and pulled toward the target plate 2, and as a result, the second full ball bearing 402 is preloaded. , The surface pressure acting on each individual ball is low.

However, when the thermal expansion is ignored, the first bearing device 401 close to the target plate 2 receives almost no radial load, so that only a specific ball receives the radial load, and the overhung condition occurs. become,
The surface pressure becomes high. As a result, both ball bearings 401, 402
If the specifications are completely the same, the life of the first full ball bearing 401 is shortened and the life of the product is controlled.

Therefore, in this embodiment, the second full ball bearing 4 is used.
By setting the load capacity of 02 to be smaller than that of the first total ball bearing 401, the surface pressure is distributed, and the rotational torque τ of the second total ball bearing 402 is reduced, so that the overall rotational torque is reduced. We are trying to reduce τ. As a first aspect thereof, it is preferable that the number of balls 42 of the second total ball bearing 402 be smaller than the number of first total ball bearings 401.

As another aspect, it is preferable to make the diameter of the balls 42 of the second total ball bearing 402 smaller than the diameter of the first total ball bearing 401. In any of the above-described modes, the load capacity of the first total ball bearing 401 does not change, and the rotational torque τ of the second total ball bearing 402 decreases, so that the rotational torque τ can be reduced as a whole. . This is because in the second full ball bearing 402, the so-called preload acts due to the axial load originating from the cantilever support, and as a result, even if the load capacity is set as described above and the load capacity is reduced, Although the life is not affected and the rotation torque τ can be reduced, the first full ball bearing 4
In 01, since the load capacity does not change, the life does not change. Therefore, the rotating torque τ can be reduced as a whole, and the life can be extended.

Table 1 shows the results of calculating specific numerical values in the above-mentioned embodiment. As the calculation conditions, the load of the center of gravity G shown in FIG. 2 is 3.5 KG, the axial distance D0 from the center of gravity to the first total ball bearing 401 is 63 cm, and the axial distance between both total ball bearings 401 and 402 is 50.8 cm for D1
When set to.

[0026]

[Table 1]

[Measurement Results] Rotational Torque τ According to the Present Embodiment
In order to investigate the life and the life, the rotation torque τ and the life ratio were measured under the following conditions. (1) Sample To measure the change in the rotation torque τ, A. Angular type full ball bearing (Model No. 608 or equivalent. Lubricant: coated with solid lubricant (AgIP treatment))
, The ball diameter Bd and the number of balls Z are made constant, and the clearance C
B. Only C is changed (test product ~), B. Using the above-mentioned angular type full ball bearing, the number of balls Z is set to be the same as the test product, the ball diameter Bd is set to be smaller than the test product, and only the clearance CC is changed (test product to), and C ． Using the above-mentioned angular type full ball bearing, the ball diameter Bd is set to be the same as the test product to, the number of balls Z is set to be smaller than the test product, and only the clearance CC is changed (test product). And Among the above-mentioned test products, the products equivalent to the embodiments of the present invention are.

Further, in order to measure the life ratio, the test products A to C, ... Are used as samples. (2) Test device The test device shown in FIG. 4 was adopted. In the figure, a rotary shaft 451 which is supported by a sample full ball bearing 400 is supported by a first air bearing 420 in a neutral state, and an outer ring of the full ball bearing 400 is fixed to a torque detection housing 430. did. This torque detection housing 4
The 30 is supported by the second air bearing 440 in a levitated state. Reference numeral 450 denotes an air load device arranged at the upper end of the rotating shaft 451 for applying a predetermined axial load (preload) to the full ball bearing 400. This axial load is adjusted arbitrarily. The rotating shaft 451 is rotationally driven by the air turbine 460 so that the speed can be adjusted.

The rotational torque acting on the torque detecting housing 430 is detected by a torque detector 470. The detection signal is input to the XY recorder 490 via the amplifier 480. The rotation speed of the rotary shaft 410 can be detected by the non-contact rotation speed detector 500 and the tachometer 510, and the detection signal is input to the XY recorder 490. X-
The Y recorder 490 outputs the detection result.

(3) Test Method (3-1) Measurement of Rotational Torque Using the test apparatus shown in FIG.
A preload (load in the axial direction) of f was applied, and it was rotated at a rotation speed of 4000 rpm at atmospheric pressure and room temperature.

(3-2) Life measurement Using the test apparatus shown in FIG.
Apply a preload of 15 kg / f (axial load),
It was rotated at a rotation speed of 3000 rpm at room temperature at 10 ⁻⁵ Pa. (4) Measurement results FIG. 5 is a plot diagram showing the measurement results of the above-described test products, the white ones show the numerical values of the rotational torque τ,
The blackened ones are plots of life ratios where the life of the test product is 1.

As shown in FIG. 5, it has been found that the rotational torque τ decreases as the clearance CC increases.
However, if the clearance CC exceeds 2.0,
It is expected that the life ratio will be shortened, and it has been found that the range of the above relational expression (1) is the most preferable to satisfy the requirements of both. Next, the ball diameter Bd set to be smaller than the test product (corresponding to column B: test product,
), The rotational torque τ was lower than that of the standard size ball diameter Bd.

Further, in the case where the number of balls Z is set to be smaller than that of the test product (corresponding to column C: test product), the rotational torque τ is lower than that corresponding to the above column B, and It was found that the life ratio did not decrease so much. In the above embodiment, the so-called integral type deep groove type full ball bearing has been described, but a deep groove type full ball bearing having an ordinary inner ring may be used. Further, it is not limited to the oblique contact ball bearing, and may be a deep groove ball bearing in which ball insertion grooves are formed in either or both of the inner and outer rings.

[0034]

As explained above, according to the full ball bearing of the present invention, the contact area between the ball and the inner ring raceway groove is reduced, and further, the ball is not ridden on the shoulder and the outer ring raceway groove is formed. By rolling contact with the inner and outer raceway grooves along the cross-sectional shape of the inner raceway groove while being guided by rolling, it is possible to prevent problems such as an increase in rotational torque due to slippage and shoulder riding, and peeling of the lubricating coating. So,
The remarkable effect that the life of the full ball bearing can be prevented from being shortened is achieved. In particular, since the rotation torque can be reduced, the spin of the ball provided with the solid lubricating film can be suppressed. As a result, damage to the lubricating film due to friction with the raceway groove can be suppressed, and the solid lubricating life can be extended. There is an advantage.

Moreover, according to the present invention, the rate of heat conduction from the rotating shaft is reduced and the temperature rise of the bearing is reduced, so that it is less susceptible to thermal influences, and this also contributes to extending the life. can do.

[Brief description of drawings]

FIG. 1 is an enlarged view of a main part of a full ball bearing according to an embodiment of the present invention.

FIG. 2 is a schematic front view showing a main part of a rotary anode X-ray tube that employs the full ball bearing.

FIG. 3 is a schematic side view of an essential part showing a part of the rotating anode X-ray tube in a cutaway manner.

FIG. 4 is a schematic diagram of a test apparatus used for measuring rotational torque and life.

FIG. 5 is a plot diagram showing a result of the above measurement.

FIG. 6 is a schematic front view showing a general rotating anode X-ray tube.

FIG. 7 is a schematic front view showing an enlarged main part of the rotary anode X-ray tube in FIG.

[Explanation of symbols] 5 rotating shaft 41a outer ring raceway groove 42 balls 51 inner ring raceway groove

Claims (2)

[Claims] 1. A full ball bearing having a plurality of balls arranged between raceway grooves of inner and outer races, wherein a radius of curvature of the inner raceway groove is set larger than a radius of curvature of the outer raceway groove. And full ball bearings. 2. The full ball bearing according to claim 1, wherein the surface of the ball is coated with a solid lubricant.