JP3360396B2 - Ball bearing device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION A ball bearing device according to the present invention is used for rotatably supporting a rotary shaft of a screw compressor.

[0002]

2. Description of the Related Art A ball bearing device as shown in FIG. 1 is conventionally used to support a high-speed rotating shaft such as a rotary shaft of a screw compressor. This ball bearing device is provided between an outer peripheral surface of a rotating shaft 2 to which a rotor 1 constituting a screw compressor is fixed and an inner peripheral surface of a housing 3. Incidentally, 4 in FIG. 1 is a roller bearing for supporting a load F _r in the radial direction.

[0003] ball bearing apparatus is the subject of the present invention is intended for supporting an axial load F _a over the axial direction of the rotary shaft 2 (lateral direction in FIG. 1), the first respectively are angular type , And the second ball bearings 5 and 6 are combined. The contact angle α of these first and second ball bearings 5 and 6
The directions of (see FIG. 2 to be described later) are opposite to each other, and a front combination (hereinafter referred to as “D
F ". ). For this reason, when the rotating shaft 2 is to be displaced to the left in FIG. 1, the first ball bearing 5 on the left in FIG. The right second ball bearing 6 supports the axial load to prevent the rotating shaft 2 and the rotor 1 from being displaced with respect to the housing 3.

[0004] Incidentally, as a combination type in which an axial load in any direction can be supported by combining a pair of angular type ball bearings, the first and second types are opposite to the structure shown in FIG. There is a so-called back combination, in which the back surfaces of the ball bearings are combined so as to face each other. However, when such a back-to-back combination structure is employed, the interval between the action points on the inner ring side of the first and second ball bearings increases, and the bending rigidity against the inclination of the rotating shaft increases. . Therefore, when the back combination is used to support a member having a relatively large inclination, such as the rotating shaft 2 of the screw compressor, an excessive surface pressure is likely to act on the action point portion. Excessive surface pressure is not preferable because it causes abnormal heat generation and shortens the fatigue life of the ball rolling surface and the raceway surface.

The ball bearing device incorporated in the screw compressor includes a rolling surface of balls 11 and 12 constituting first and second ball bearings 5 and 6, an inner ring 10 and an inner ring raceway 15 on the outer peripheral surface. 15, outer ring 13, outer ring raceway 1 on inner peripheral surface of 13
4 and 14 are operated with a plus gap (a gap that actually exists). The reason for this is also to prevent heat generation and reduction in fatigue life. Screw compressors ball bearing apparatus to which the present invention is incorporated, ultrafast (e.g. d _m n is from 700,000 to 2,000,000 mm ·
rpm). For this reason, when a preload is applied to each of the balls 11 and 12 and the first and second ball bearings 5 and 6 having a so-called negative gap are used, these first and second ball bearings 5 and 6 are used based on the preload. The internal load of the ball bearings 5 and 6 increases,
Again, it causes abnormal heat generation and shortens the fatigue life.

Therefore, the ball bearing device incorporated in the screw compressor uses the angular type first and second ball bearings 5 and 6 each having a plus clearance in the DF. Further, the rotary shaft 2 which is supported by the ball bearing apparatus, in use of the screw compressor, axial load F _a substantially constant direction with the rotation of the rotor 1 (in Fig. 1 from right to left) Is added. This is because the direction of the pressure applied to the rotor 1 is determined. still,
In FIG. 1, reference numerals 7 and 8 denote spacers and reference numeral 9 denotes a presser.

By the way, as described above, the ball bearing device in which the first and second ball bearings 5 and 6 each of which is of the angular type is combined with the DF, when the rotating shaft 2 rotates at a high speed, A sufficient bearing life cannot always be obtained. Regarding the cause of the shortening of the bearing life with high speed rotation,
This will be described with reference to FIG.

FIG. 2 is a view for explaining the force applied to the first and second ball bearings 5 and 6 based on the centrifugal force generated by high-speed rotation. When the rotating shaft 2 (FIG. 1) is rotated at a high speed, the balls 11, 1 constituting the first and second ball bearings 5, 6 are formed.
2, a centrifugal force F _c1 is applied. The balls 11 and 12 constituting the first and second ball bearings 5 and 6 have the centrifugal force F
Based on _c1 , they are pressed against the outer ring raceways 14, 14 with a force of Q _in1 from the direction of the contact angle α. And these first,
The balls 11, 12 constituting the second ball bearings 5, 6 are _subjected to the axial component component _{Fain 1} of the force Q _in1 to each outer raceway 1.
4, 14 are pressed in the axial direction.

Then, an internal axial load corresponding to F _ain1 + F _ain1 is generated in the first and second ball bearings 5 and 6 which are combined in series with each other due to this component force. The contact surface pressure with the balls 11 and 12 is increased, and the fatigue life of the first and second ball bearings 5 and 6 is reduced. The calculation of the fatigue life of a ball bearing in consideration of the centrifugal force of the ball during high-speed rotation is described in “The Life o ASME” published in Trans. ASME issued in July 1952.
f High-Speed Ball Bearings ”.

In order to prevent the life from being shortened due to the above-mentioned causes, the balls 11, 12 constituting the first and second ball bearings 5, 6 are formed.
Is made of lightweight ceramic material, or ball 11,
Attempts have also been made to reduce the centrifugal force _Fc1 by using a small diameter as 12. Further, attempts have been made to reduce the axial component F _ain1 based on the centrifugal force F _c1 by reducing the contact angles α of the first and second ball bearings 5 and 6. .

[0011] However, the balls 11 and 12, when made by a large ceramic material modulus, in the first ball bearing 5 side in particular supporting the external loads (e.g. the axial load F _a), the outer ring raceway The contact surface pressure between 14 and ball 11 is
It increases more than when a ball made of bearing steel is used. As a result, the fatigue life of the first ball bearing 5 decreases, and the fatigue life of the entire ball bearing device formed by combining the first and second ball bearings 5 and 6 decreases. Also, if small diameter balls 11 and 12 are used or the contact angle α is reduced, the axial load capacity (basic dynamic load rating) of the first and second ball bearings 5 and 6 is reduced, and the tires are also fatigued. The life will be shortened.

In order to reduce the internal axial load based on such centrifugal force and extend the fatigue life of the first and second ball bearings 5, 6, JP-A-5-248431 and JP-A-5-248431.
Japanese Patent Publication No. 280482 discloses a ball bearing device as shown in FIGS. In the case of the ball bearing apparatus is provided in axial load F _a the bears load applied from the outside to the rotary shaft 2 (left side in FIGS. 3-4) at the time of use, the contact angle of the first ball bearing 5a alpha than _1, the axial load F anti-load side without supporting the _a small contact angle alpha ₂ of the second ball bearing 6a provided (right side in FIGS. 3-4) at the time of use (alpha
₂ <α ₁ ).

As described above, the contact angle α of the first ball bearing 5a
₁ By larger than the contact angle alpha ₂ of the second ball bearing 6a, can be sufficiently ensured load capacity of the first ball bearing 5a. In this case, the contact angle α ₁ is the same as the contact angle α of the first and second ball bearings 5 and 6 incorporated in the conventionally generally known ball bearing device (α = α). _{If 1} ), the axial component force acting on the first ball bearing 5a based on the centrifugal force is _Fain1 . Further, the axial component force applied to the second ball bearing 6a is _Fain2 . In this case, the contact angle α _{2 of} the second ball bearing 6a
Is smaller than the contact angle α1 of the _first ball bearing 5a.
For this reason, the axial component force _Fain2 applied to the second ball bearing 6a is smaller than the axial _component force _Fain1 acting on the second ball bearing 6 of the conventional device ( _Fain2 < _Fain2).
ain1 ). As a result, the first and second ball bearings 5a, 6a
Occurs to the ball bearing apparatus comprising a combination of an internal axial load based on the centrifugal _force, F ain1 + F ain2 next, the internal axial load _F ain1 + F ain1 occurring conventional apparatus
Smaller than.

For this reason, the first and second ball bearings 5a, 6
The contact surface pressure between the balls 11, 12 constituting the a and the outer raceways 14, 14 decreases. And the first and second ball bearings 5a,
The reduction in the fatigue life of the ball bearing 6a can be suppressed, and the reduction in the life of the ball bearing device formed by combining the first and second ball bearings 5a, 6a can be suppressed. Load capacity of the axial direction of the second ball bearing 6a having a small contact angle alpha ₂ is reduced, the second ball bearing 6a is for supporting the light load applied to rare, only have the function as a kind of back-up bearing Since there is no load capacity reduction, there is no particular problem.

[0015]

[SUMMARY OF THE INVENTION However, when the ball bearing apparatus described in JP-like, such as described above, the contact angle alpha ₂ to the load side of the second ball bearing 6a of counter-load-side ball bearing 5a to smaller than the contact angle alpha _1, these first and second ball bearings 5a, contact angle alpha ₁ of 6a, it is necessary to strictly manage the alpha _2. Therefore, these first and second ball bearings 5
Not only do the manufacturing costs of the devices a and 6a increase, but also the following problems occur.

That is, a ball bearing device as shown in FIGS.
When mounting between the outer peripheral surface of the rotary shaft 2 of the screw compressor and the inner peripheral surface of the housing 3, the mounting direction must be correct. When the mounting direction reversed (when mounted in the direction of supporting the axial load F _a by the second ball bearing 6a), the durability is remarkably deteriorated. However, the contact angle of the assembled ball bearing cannot be easily confirmed from the outside visually. For this reason, if care is not taken at the time of assembly work, there is a possibility that the compressor manufacturer will assemble the ball bearing device shipped from the bearing manufacturer in the opposite direction.

Of course, if a mark indicating the mounting direction is provided by a bearing maker, such erroneous mounting can be prevented. However, the provision of the mark complicates the manufacturing operation of the ball bearing device and increases the manufacturing cost. Cause. The ball bearing device of the present invention was invented in view of such circumstances. still,
As a combination ball bearing that changes the number of balls between the load side and the non-load side, "PUMPAC The M" issued by MRC Bearing Services
The one described in RC Bearing System is also conventionally known. However, in the case of the structure described in this, instead of making the counter-load-side ball smaller in diameter than the load-side ball, the number of the counter-load-side ball is made larger than the load-side ball. I have. This structure is intended to release the preload of the ball bearing, and is fundamentally different from the present invention in any of the object, configuration, and effects.

[0018]

A ball bearing device according to the present invention comprises:
Like the above-described conventional ball bearing device, first and second balls provided between the outer peripheral surface of the shaft and the inner peripheral surface of the housing are each of an angular type and have different contact angle directions. Equipped with bearings. An axial load applied from outside to the shaft and the housing in a substantially constant direction during use is supported by the first ball bearing.

In particular, the ball bearing device of the present invention is characterized in that the number m of the balls constituting the second ball bearing is smaller than the number n of the balls constituting the first ball bearing.

[0020]

The ball bearing device of the present invention configured as described above has the following features.
Since the number n of balls constituting the first ball bearing that supports an axial load applied from the outside during use is sufficiently large and the load capacity of the first ball bearing is large, a sufficiently large axial load can be supported. In addition, the number m of balls constituting the second ball bearing that does not support this axial load during use is small,
Since the sum of the internal axial loads based on the centrifugal force applied to these balls is the internal axial load generated in the second ball bearing, the increase in the internal axial load is suppressed. For this reason, it is possible to suppress a decrease in fatigue life of the entire ball bearing device formed by combining the pair of angular type first and second ball bearings.

Furthermore, if the first and second ball bearings are arranged in DF and have a positive clearance, no preload is applied even during operation of the ball bearing device, and heat generation and fatigue life based on the preload are reduced. There is no drop. Also, the allowable limit for the inclination of the shaft is high.

[0022]

FIG. 5 shows a first embodiment of the present invention. A non-contact type seal such as a labyrinth seal is provided between the outer peripheral surface of the rotary shaft 2 and the inner peripheral surface of the housing 3 in order from the side of the rotor 1 housed in the housing 3. Is a seal device 19 such as a contact-type seal such as a mechanical seal, a roller bearing 4 for supporting a radial load, and a first ball bearing 5a and a second ball bearing 6a constituting the ball bearing device of the present invention. And are arranged in series with each other.

In the case of the ball bearing device of the present invention, the number m of the balls 12 constituting the second ball bearing 6a on the non-load side is replaced by the number m of the balls 11 constituting the first ball bearing 5a on the load side. n (m <n). That is, in the case of the ball bearing device of the present invention, the number m of the balls 12 constituting the second ball bearing 6a is several m.
And the internal axial load applied from each ball 12 to the outer ring 13 constituting the second ball bearing 6a is reduced. By reducing the number m of the balls 12, the second ball bearing 6
Although the load capacity of "a" is small, the load applied to the second ball bearing 6a is small, so that practically sufficient durability can be ensured even when the number of balls 12 is reduced to several meters.

The ball 1 constituting the second ball bearing 6a as described above
2 is the number n of the balls 11 constituting the first ball bearing 5a.
When the number is smaller than the above, the small number m is increased to the large number n of 7
It is preferable to regulate to 0 to 80% (m = (0.7 to 0.8) n). When more than 80% of the balls 12 are incorporated, the effect of reducing the internal axial load cannot always be obtained sufficiently. On the other hand, if it is less than 70%, the interval between the balls 12 adjacent in the circumferential direction becomes too wide, and there is a possibility that smooth rotation is hindered.

The first ball bearing 5 constituting the ball bearing device
The relationship between the calculated life L _5a of the second ball bearing 6a and the calculated life L _{6a of} the second ball bearing 6a is set to 3L _5a ≦ L _6a.
Calculated life L _{T of a} ball bearing device composed of a and 6a
It is preferable from the viewpoint of securing The reason is as follows. The calculated life of a ball bearing device composed of a combination of multiple ball bearings is not the same as the calculated life of the shortest ball bearing (short life bearing), but is shorter than this.
Being affected by the calculated life of a ball bearing with a long calculated life (long life bearing), and the longer the calculated life of a long life bearing,
Be calculated life L _T of the ball bearing device approaches the calculated life of the short life the bearings are known in the art.

On the other hand, the other hand it is difficult to increase the calculated life L _5a of the first ball bearing 5a for receiving the axial load, in use, the second ball bearing 6 is not subjected to such axial load
It is easy to extend the calculated life _L6a of a. Therefore,
If the ball bearing apparatus of the present invention, the calculated life L _5a of the first ball bearing 5a considered calculated life of short life bearing, calculated life L _6a a long life bearing of calculated life of the second ball bearing 6a Can be considered. Therefore, these first and second ball bearings 5a,
6a the calculated life L _5a, the ratio of L _6a and (L _6a / L _5a), the calculated life of the calculated life L _T and the first ball bearing 5a of both ball bearings 5a, ball bearing apparatus comprising a combination of 6a FIG. 6 shows the relationship between L _5a and the ratio (L _T / L _5a ).
FIG. 6 is based on a conventionally known formula for calculating the life of a combination bearing. As is apparent from FIG. 6, if the calculated life L _6a of the second ball bearing 6a is set to be three times or more the calculated life L _5a of the first ball bearing 5a, the calculated life L _T of the ball bearing device is reduced to the first life. It can be of more than 80% of the calculated life L _5a of the ball bearing 5a.

Of course, if the calculated life L _6a of the second ball bearing 6a is made much longer than the calculated life L _5a of the first ball bearing 5a, the calculated life L _T of the above-mentioned ball bearing device will be increased to the first. Ball bearing 5
It can be closer to the calculated life _L5a of a. However,
Even if the calculated life L _6a of the second ball bearing 6a is extended significantly more than three times the calculated life L _5a of the first ball bearing 5a, compared to the cost required to extend the calculated life L _6a , prolonging effect of the calculated life L _T as obtained thereby ball bearing apparatus is small. Therefore, considering the overall cost, it is appropriate that the calculated life L _6a of the second ball bearing 6a exceeds three times the calculated life L _5a of the first ball bearing 5a. still,
Each of these calculated lifetimes is described in The Life of High-Speed Ba
ll Bearings.

On the other hand, the spacer 7 is provided between the inner ring 20 of the roller bearing 4 and the first ball bearing 5a, and the outer ring 21 of the roller bearing 4 and the outer ring 13 of the first ball bearing 5a. A nozzle ring 22 is sandwiched between the nozzle rings. The portions of both sides of the nozzle ring 22 that are closer to the inner periphery are formed as conical concave inclined surfaces that are inclined toward each other toward the inner peripheral edge. The nozzle holes 23 are opened at one or more positions in the circumferential direction of both inclined surfaces. Each of the nozzle holes 23 is formed in a direction perpendicular to the inclined surface. The provision of the inclined surface is for facilitating the operation of forming the nozzle holes 23 in the inclined direction. Accordingly, each of the nozzle holes 23 is open obliquely inward in the diameter direction. Lubricating oil can be freely fed into these nozzle holes 23 via an oil supply passage 24 formed in the housing 3 and the nozzle ring 22.

On the other hand, the first and second ball bearings 5a, 6a
The retainers 16, 16 are incorporated in the
12 is rotatably held. Each of these cages 16,
Reference numeral 16 denotes a so-called machined retainer, which is formed by forming pockets 18 for holding the balls 11 and 12 in the cylindrical main portions 17 and 17 in a freely rolling manner. Such retainers 16, 16 are respectively provided with first and second ball bearings 5a,
The inner ring 6a is mounted inside the outer rings 13, 13 by a so-called outer ring guide. That is, the outer peripheral surfaces of the retainers 16, 16 in the axial direction (left-right direction in FIG. 5) and the outer races 13, 13
Are opposed to each other via a minute gap.

The reason why the guide state of each of the retainers 16 and 16 is set as the outer ring guide is as follows. That is,
General state of use of the ball bearing apparatus, i.e., d _m n is 7
If the pressure is not more than 100,000 and the first and second ball bearings 5a, 6a are preloaded and vibration is less generated during operation, it is not necessary to particularly restrict the guiding conditions of the retainers 16, 16. Absent. On the other hand, in the case of the subject to, the ball bearing device incorporated in the screw compressor of the present invention, d _m n is 70
It is operated in a super-high speed range exceeding 10,000 with a positive axial clearance. Moreover, the vibration increases due to the fluctuation of the axial load peculiar to the screw compressor. In such a state of use, in order to suppress the wear of the guide surfaces of the retainers 16, 16, it is necessary to provide an outer ring guide capable of suppressing the inclination of the retainers 16, 16.

On the other hand, when the cages 16, 16 are rolling element guides or inner ring guides, the rigidity is lower than that of the inner ring 10 and the outer ring 13 due to centrifugal force during operation. ) As the diameters of the retainers 16 and 16 are elastically increased, the interval between the guide surfaces is increased, and sufficient guidance cannot be performed. When vibrations during operation are applied to the cages 16 in this state, the surfaces of the cages 16 and 16 come into contact with each other with a small area, which causes inconvenience such as abnormal wear. When the outer ring guide is used, the retainers 16, 1
Even if the diameter of 6 becomes wider, the guide surface (cage 16,
The distance between the outer peripheral surface of the outer ring 16 and the inner peripheral surface of the outer rings 13, 13 does not increase. Since lubricating oil exists between the guide surfaces, the guide surfaces do not directly rub against each other. Therefore, by using the cages 16 and 16 as outer ring guides, abnormal wear of the guide surfaces can be effectively prevented. In addition, the retainers 16, 1
If a copper alloy cage is used as 6, this cage 1
By increasing the strength of the retainers 6, 16, the retainers 16, 16 can be effectively prevented from being damaged. Incidentally, as long as d _m n in use is not so high, therefore the centrifugal force applied to the cage 16, 16 is limited, as in the second embodiment shown in FIG. 7, each retainer 16, 16 May be used as the outer ring one shoulder guide.

Further, in the case of the first embodiment shown in FIG. 5, the inner peripheral surfaces of both ends of each of the retainers 16, 16 in the axial direction (left-right direction in FIG. 5) have inner diameters closer to the edges. The conical concave inclined surface becomes larger. Accordingly, the inner rings 10, 10 constituting the first and second ball bearings 5a, 6a
The width dimension of the space existing between the outer peripheral surface and the inner peripheral surface of each of the retainers 16 and 16 in the diametrical direction (vertical direction in FIG. 5) increases toward the opening end. As a result, the lubricating oil ejected from the nozzle holes 23, 23
The efficiency of feeding into the second ball bearings 5a, 6a is improved.
Therefore, even when the ball bearing device constituted by these ball bearings 5a and 6a is operated at an ultra-high speed, these first and second ball bearings are used.
A sufficient amount of lubricating oil can be taken into the second ball bearings 5a and 6a.

The reason why the direction in which the nozzle holes 23 are opened and the shapes of the inner peripheral surfaces of the retainers 16 and 16 are as described above is as follows. That is, the ball bearing device of the present invention
The life can be greatly extended as compared with the above-described conventional device. Therefore, in the case where the required life is greatly exceeded when the product is manufactured in the same size as the conventional one, the first and second ball bearings 5 are required.
The outer diameter can be reduced without changing the inner diameter by reducing the dimension series of a and 6a. As a result, the size and cost of the ball bearing device can be reduced, but the internal space of a ball bearing having a small size series is narrow. For this reason, it is difficult to perform jet lubrication (a lubricating method of breaking a wall of air generated inside a bearing based on high-speed rotation and feeding lubricating oil into the bearing) which is generally performed in a screw compressor. Therefore, the jet lubrication is enabled by adopting the opening direction and the inner peripheral surface shape as described above.

The angle of inclination of the nozzle hole 23 with respect to the rotating shaft 2 is optimally about 13 degrees, but can be set in the range of 10 to 20 degrees. Further, the nozzle ring 2 having the nozzle hole 23 and having a function as an outer ring spacer is also provided.
2 has a hardness of HRc 56 to 66 (more preferably HRc 60
-62). This is because the two sides of the nozzle ring 22 which come into contact with each other and the outer races 21 and 1 of the roller bearing 4 and the first ball bearing 5a due to the specific vibration generated during the operation of the screw compressor.
This is for preventing the occurrence of fretting wear on the end face 3.

Although the illustrated embodiment shows an example in which a pair of angular ball bearings are combined with a DF, _a plurality of first ball bearings for supporting an axial load Fa during operation are provided. parallel a plurality of first ball bearing these combined with, and one second ball bearing that does not support the axial load F _a may be combined with DF. Furthermore, if it is an application other than the screw compressor, it can be used in a back-to-back combination or with a slight preload applied depending on the use conditions.

[0036]

Since the present invention is constructed and operated as described above, the internal axial load is reduced while the load capacity against the external axial load is sufficiently secured, and the durability of the ball bearing device is improved. I can do it. In addition, it is possible to increase the tolerance for the inclination of the rotating shaft and to surely prevent abnormal wear due to slippage. For this reason, more excellent durability can be reliably obtained. Also, since the number of balls can be easily confirmed visually from the outside, there is no case where the assembling direction is mistaken even if no mark is provided. Therefore, it is possible to save time and effort for providing a mark, and to reduce the cost of the ball bearing device.

[Brief description of the drawings]

FIG. 1 is a sectional view showing a first example of a conventional device.

FIG. 2 is a sectional view of an essential part of the same.

FIG. 3 is a sectional view showing the second example.

FIG. 4 is a sectional view of an essential part of the same.

FIG. 5 is a sectional view of a main part showing the first embodiment of the present invention.

FIG. 6 shows the ratio between the calculated life of the first ball bearing and the calculated life of the second ball bearing, and the ratio between the calculated life of the ball bearing device formed by combining them and the calculated life of the first ball bearing. FIG.

FIG. 7 is a sectional view of a main part showing a second embodiment of the present invention.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Rotor 2 Rotating shaft 3 Housing 4 Roller bearing 5, 5a First ball bearing 6, 6a Second ball bearing 7, 8 Spacer 9 Retainer 10 Inner ring 11, 12 Ball 13 Outer ring 14 Outer ring raceway 15 Inner ring raceway 16 Holding Container 17 Main part 18 Pocket 19 Sealing device 20 Inner ring 21 Outer ring 22 Nozzle ring 23 Nozzle hole 24 Oil supply passage

──────────────────────────────────────────────────続 き Continued on front page (58) Field surveyed (Int. Cl. ⁷ , DB name) F16C 19/00-19/56 F16C 33/30-33/66

Claims (1)

(57) [Claims] A first and a second ball bearings each provided between an outer peripheral surface of a shaft and an inner peripheral surface of a housing, each of which is of an angular type and whose contact angle directions are different from each other. In a ball bearing device in which an axial load applied from the outside in a substantially constant direction between the shaft and the housing is supported by the first ball bearing, the number m of balls constituting the second ball bearing is equal to the first m. A ball bearing device, wherein the number is smaller than the number n of balls constituting the ball bearing.