The present invention relates to a ball bearing, and in particular to a ball bearing to which oil is supplied through an outer ring thereof.
In recent years, with the aim of improving cutting efficiency of a machine tool spindle, the demand for speed up has been increased. Recently, in order to improve production efficiency, the spindle also needs to respond to 5-axis processing machines capable of processing a workpiece having a complicated shape without using a plurality of machine tools and without changing the stage. In the 5-axis processing machine, the spindle and a table is rotated, so that an axial length of the spindle is required to be shortened due to requirements of saving space by shortening a turning radius, or saving electric power by reducing inertia during turning or reducing weight.
Examples of methods for lubricating rolling bearings, which are often adopted for the machine tool spindle, include grease lubrication, oil air lubrication, oil mist lubrication, or the like. In general, the oil air lubrication is adopted in the field of high speed rotation (dmn is 800,000 or more). As conventional oil air lubrication, there has been know that an oil supply nozzle head 101 arrange on a lateral side of a bearing 100 shown in
In this way, it is necessary to separately provide an oil supply part such as the nozzle head 101 and the number of parts of the spindle is increased, which leads to an increase in the cost of the whole spindle and in labor and time of management. Since the nozzle head 101 is used, a shape of the outer ring spacer and a structure of a housing become complicated, and time and labor for designing and processing the spindle is increased. Further, since the nozzle head 101 is provided on an axial side surface side of the bearing, a spacer with a certain length is required, and an axial length of the spindle increases. As a result, a size of the machine tool increases, the spindle weight becomes heavier as the axial length increases, and a critical speed (the critical speed is a rotation speed calculated from a natural vibration frequency of the spindle, and if the spindle is rotated in a range of the critical speed, the vibration becomes large) of the spindle decreases. The supply of oil particles from the oil supply nozzle is obstructed by an air curtain (the air curtain is a wall of high speed air flow in a circumferential direction generated by friction between air and an outer diameter surface of an inner ring rotating at high speed) generated along with the high speed rotation, and as a result, it is difficult to reliably supply the lubricating oil to the interior of the bearing. Such conventional oil air lubrication is superior to grease lubrication in lubricity under high speed rotation, but the responsiveness is becoming more important as speed up has progressed.
As another oil air lubrication method, there has been know that, as shown in
Patent Document 1: JP-A-2013-79711
In a bearing to which oil is supplied through an outer ring thereof, since lubricating oil is directly supplied to the vicinity of a contact portion between rolling elements and inner and outer rings, functional problems may occur in the bearing according to supply positions. For example, in a case where a position of an raceway side opening of an axial hole of the outer ring overlaps a contact ellipse of a contact portion between an outer ring groove and a ball, a contact surface pressure between the outer ring groove and the ball increases in the vicinity of an edge of the opening and damage such as premature seizure is likely to occur in the bearing. Further, even if the position of the raceway side opening of the axial hole provided in the outer ring does not overlap the contact ellipse between the outer ring groove and the ball, when the lubricating oil is supplied to the vicinity of the contact ellipse at one time, damage due to a large amount of heat generated in the bearing may occur.
The present invention has been made in view of the above-described problems, and an object thereof is to provide a ball bearing that can provide better lubricity and rotation performance by appropriately setting an axial position of a radial hole provided in an outer ring of a bearing to which oil is supplied through the outer ring thereof, according to an intended use of the bearing.
The above object of the present invention is achieved by the following configurations.
(1) A ball bearing lubricated by lubricating oil, includes: an inner ring including an inner ring raceway groove on an outer circumferential surface thereof; an outer ring including an outer ring raceway groove on an inner circumferential surface thereof; and a plurality of balls rollably arranged between the inner ring raceway groove and the outer ring raceway groove,
in which the outer ring includes at least one radial hole that extends radially through the outer ring from an outer circumferential surface thereof to the inner circumferential surface thereof, and
in which when a minor angle formed relative to a rotation axis of the ball by a straight line that connects a center of the ball and an intersection point between a centerline of the radial hole and a surface of the ball is defined as λ, an axial position of the centerline of the radial hole is set to satisfy a relationship of 0°<λ≤60°.
(2) In the ball bearing described in (1), a counter bore is provided on the inner circumferential surface of the outer ring, the counter bore includes an inclined portion that gradually decreases in diameter from an axial end surface toward the outer ring raceway groove, and a straight portion that connects the inclined portion and the outer ring raceway groove and is parallel to a centerline of a rotation axis of the ball bearing, and
an inner diameter side opening of the radial hole is entirely positioned in the inclined portion.
(3) In the ball bearing described in (1), a counter bore is provided on the inner circumferential surface of the outer ring, the counter bore includes an inclined portion that gradually decreases in diameter from an axial end surface toward the outer ring raceway groove, and a straight portion that connects the inclined portion and the outer ring raceway groove and is parallel to a centerline of a rotation axis of the ball bearing, and
an inner diameter side opening of the radial hole is positioned to straddle the inclined portion and the straight portion.
(4) In the ball bearing described in (1), a counter bore is provided on the inner circumferential surface of the outer ring, the counter bore includes an inclined portion that gradually decreases in diameter from an axial end surface toward the outer ring raceway groove, and a straight portion that connects the inclined portion and the outer ring raceway groove and is parallel to a centerline of a rotation axis of the ball bearing, and
an inner diameter side opening of the radial hole is entirely positioned in the straight portion.
(5) In the ball bearing described in (1), a counter bore is provided on the inner circumferential surface of the outer ring, the counter bore includes an inclined portion that gradually decreases in diameter from an axial end surface toward the outer ring raceway groove, and a straight portion that connects the inclined portion and the outer ring raceway groove and is parallel to a centerline of a rotation axis of the ball bearing, and
an inner diameter side opening of the radial hole is positioned to straddle the straight portion and the outer ring raceway groove.
(6) In the ball bearing described in (1), a counter bore is provided on the inner circumferential surface of the outer ring, the counter bore includes an inclined portion that gradually decreases in diameter from an axial end surface toward the outer ring raceway groove, and a straight portion that connects the inclined portion and the outer ring raceway groove and is parallel to a centerline of a rotation axis of the ball bearing, and
an inner diameter side opening of the radial hole is entirely positioned in the outer ring raceway groove.
(7) In the ball bearing described in (1), a counter bore including an inclined portion that gradually decreases in diameter from an axial end surface to the outer ring raceway groove is provided on the inner circumferential surface of the outer ring, and
an inner diameter side opening of the radial hole is entirely positioned in the counter bore.
(8) In the ball bearing described in (1), a counter bore including an inclined portion that gradually decreases in diameter from an axial end surface to the outer ring raceway groove is provided on the inner circumferential surface of the outer ring, and
an inner diameter side opening of the radial hole is positioned to straddle the counter bore and the outer ring raceway groove.
(9) In the ball bearing described in (1), a counter bore including an inclined portion that gradually decreases in diameter from an axial end surface to the outer ring raceway groove is provided on the inner circumferential surface of the outer ring, and
an inner diameter side opening of the radial hole is entirely positioned in the outer ring raceway groove.
(10) In the ball bearing described in any one of (1) to (9), the axial position of the centerline of the radial hole is set to satisfy a relationship of 30°≤λ≤60°.
(11) In the ball bearing described in any one of (1) to (10), a recessed groove in communication with the radial hole is formed on the outer circumferential surface of the outer ring along a circumferential direction thereof.
(12) In the ball bearing described in (11), annular grooves are formed on the outer circumferential surface of the outer ring along the circumferential direction on both axial sides sandwiching the recessed groove, and an annular seal member is arranged in each of the annular grooves.
(13) In the ball bearing described in any one of (1) to (12), a diameter of the radial hole is 0.5 mm to 1.5 mm.
(14) The ball bearing described in any one of (1) to (13) is a bearing used for a machine tool spindle.
According to the ball bearing of the present invention, an outer ring thereof includes at least one radial hole that extends radially through the outer ring from an outer circumferential surface thereof to an inner circumferential surface thereof; and if a minor angle formed by a rotation axis of a ball and a straight line that connects a center of the ball and an intersection point between a centerline of the radial hole and a surface of the ball is defined as λ, an axial position of the centerline of the radial hole is set to satisfy a relationship of 0°<λ≤60°, so that more stable lubrication performance and rotation performance can be obtained in high speed applications.
Hereinafter, a ball bearing according to each embodiment of the present invention will be described in detail with reference to the drawings.
As shown in
For applications in which high speed machine tool spindles are used, the contact angle α is set at 15° to 30°. In order to suppress an adverse effect of expansion of a ring on a rotation side due to a centrifugal force during rotation of the bearing and decrease of internal radial clearance due to a temperature difference between the inner and outer rings, the contact angle α is preferably set to satisfy a relationship of 18°<α≤25°.
The ball bearing 10 is a bearing to which oil is supplied through an outer ring thereof, and the outer ring 12 is provided with a radial hole 15 that extends radially through the outer ring 12 from an outer circumferential surface thereof to the inner circumferential surface thereof. A recessed groove 16 in communication with the radial hole 15 is formed on the outer circumferential surface of the outer ring 12 along a circumferential direction thereof. Therefore, in the angular contact ball bearing 10, oil air lubrication is performed such that oil particles and lubrication air supplied from an oil supply path of a housing (not shown) are directly supplied to the ball 13 via the recessed groove 16 and the radial hole 15 of the outer ring 12.
Instead of being provided in the outer ring 12, the circumferential recessed groove may be formed at a position of an opening of the oil supply path that is in communication with the radial hole 15 on an inner peripheral surface of the housing.
In the embodiment, an axial position of a centerline X of the radial hole 15 is set as follows.
Here, in
That is, in the embodiment, an axial position of the centerline X of the radial hole 15 is set to satisfy a relationship of 0°<λ≤60° (in
That is, in respect of the axial position of the raceway groove side opening of the centerline X of the radial hole 15, when λ exceeds 0°, that is, the axial position of the centerline X of the radial hole 15 is set to closer to a bottom side of an outer ring groove than the point N in
When the centerline X of the radial hole 15 moves toward the side opposite to the bottom of the groove across the point N, a value of θ in
Meanwhile, when the centerline X of the radial hole 15 reaches a range of λ>60° in
In order to suppress the occurrence of the problem, it is desirable to supply a small amount of lubricating oil little by little to the vicinity of the contact portion and the contact portion between the ball 13 and the inner and outer rings 11, 12, the radial hole 15 is provided such that the centerline X of the radial hole 15 is in the range of 0°<λ≤60°, preferably 30°≤λ≤60° in
As compared with the case of 0°<λ≤60°, when the centerline X of the radial hole 15 reaches the range of λ>60° in
In particular, since the bearing (contact angle α=15° to 30°) used in the high speed machine tool spindle has a small contact angle, and a contact portion between the ball 13 and the outer ring 12 is close to the bottom of the outer ring groove, so that when the centerline X of the oil hole is in the range of λ>60° while setting the contact angle to an upper limit of 30°, the radial hole 15 overlaps the contact ellipse between the outer ring 12 and the ball 13, which may cause a large contact surface pressure between the vicinity of an edge of the radial hole 15 and the ball 13. Therefore, in high speed applications in which the ball bearing 10 of the embodiment is mainly used, the radial hole 15 is provided such that the centerline X of the radial hole 15 is in a range of 0°<λ≤60°, more preferably in a range of 30°≤λ≤60°, and thus more stable rotation performance can be obtained.
In the embodiment, since the contact angle α is set to 15° to 30°, the radial position of the raceway groove side opening of the centerline X of the radial hole 15 is formed within the range closer to a side (counter bore side) opposite to the contact angle than the bottom of the outer ring groove.
The reason why the contact angle α is set to 30° or less is as follows.
In the application of the high speed machine tool spindle where oil air lubrication or the like is adopted as in this embodiment, when the contact angle α is set to exceed 30°, the deviation between the rotation axis of the inner and outer rings 11 and 12 and the rotation axis of the ball 13 increases, slip such as spin slip and gyro slip of the contact portions between the ball 13 and the inner ring raceway groove 11a and between the ball 13 and the outer ring raceway groove 12a becomes conspicuous, and the centrifugal force generated by the revolution of the ball 13 also becomes large.
Therefore, an internal load of the bearing (corresponding to a pre-load during operation of the bearing) increases, and the contact pressure of the contact portion also increases. As a result, the contact pressure exceeds a limit PV value that is an index serving as a main factor of a seizure limit at high speed rotation of the bearing.
Therefore, in this application, the above problem can be avoided by setting the contact angle α to 30° or less.
Here, as shown in Table 1, analysis was carried out for bearings of multiple specifications with different contact angles under operation conditions of the spindle required for high speed machine tool spindles. The results are shown in Table 2.
The operation condition of the spindle is that a dmn value (the dmn value is a product of a pitch circle diameter (mm) of a rolling element and a rotation speed (min−1) of the bearing) is 1.3 million or more (standard use condition of a recent high speed spindle), and the centrifugal force generated by rotation is very high. Moreover, as described above, as the contact angle increases, the internal slip of the bearing (spin slip, gyro slip, etc.) increases.
Therefore, in order to lower the internal pre-load during operation at high speed rotation and to lower the PV value of the rolling contact portion, the contact angle is set to be large, and therefore it is necessary to reduce the pre-load (a stationary state means a state where the rotation speed is 0) after the bearing is assembled to the spindle (refer to Table 1).
That is, as a result of specification examination, as shown in Table 2, when the contact angle is 30° or less, even if the pre-load is added after being assembled to the spindle, the function of the spindle can be satisfied, but when the contact angle exceeds 30°, the pre-load cannot be added, and it is inevitable that an internal clearance is larger than 0 (so-called backlash state) after being assembled to the spindle. Therefore, rigidity of the spindle during operation cannot be secured, vibration is likely to occur, and rotation accuracy also deteriorates in a case of the machine tool spindle, thus resulting in poor processing accuracy.
In a state where there is no pre-load in the bearing, since there is backlash in the bearing, revolution slip (skidding that is a phenomenon in which no driving force is transmitted from a rotation ring to a rolling element, and an extremely large slip occurs at the contact portion) at the rolling contact portion is likely to occur during rapid acceleration operation or rapid deceleration operation of the spindle, which may lead to wear and seizure caused by abnormal temperature rise due to this phenomenon.
As can be seen from the estimation results, the contact angle of the bearing should be 30° or less, and under this condition, the effect of 0°<λ≤60° can be effectively demonstrated.
In
That is, since the opening of the oil hole is outside the outer ring raceway groove 12a, there is no possibility that the surface of the ball is scratched or possibility that scratches on the surface of the ball cause deterioration of acoustic performance or premature damage of the bearing.
In
In
As shown in
In the embodiment, the diameter of the radial hole 15 is set to 0.5 mm to 1.5 mm in consideration of the lubricating oil supply performance. In the embodiment, the radial hole 15 has a uniform diameter throughout the radial direction.
Therefore, according to the ball bearing 10 of the embodiment, if a minor angle formed relative to the rotation axis s of the ball 13 by a straight line that connects the center O of the ball 13 and the intersection point between the centerline X of the radial hole 15 and the surface of the ball 13 is defined as λ, the axial position of the centerline X of the radial hole 15 is set to satisfy the relationship of 0°<λ≤60°, so that more stable lubrication performance and rotation performance can be obtained in the ball bearings used for high speed applications.
Further, in the embodiment, as the modification shown in
Next, a ball bearing according to a second embodiment of the present invention will be described in detail with reference to
An angular contact ball bearing 10a of this embodiment is different from that of the first embodiment in the shape of the inner circumferential surface of the outer ring 12 on the counter bore side. That is, in the embodiment, the counter bore 12b includes an inclined portion (conical surface) 12b1 that gradually decreases in diameter from an axial end surface toward the raceway groove 12a side and a straight portion (cylindrical surface) 12b2 that connects the inclined portion 12b1 and the raceway groove 12a and is parallel to a centerline L (see
In order to prevent the bearings from being disassembled after assembling, the angular contact ball bearing 10a is provided with a portion referred to as a catching margin Δr shown in
Accordingly, in a case of this embodiment in which the straight portion 12b2 for defining the catching margin Δr is provided, the counter bore 12b includes only the inclined portion, and compared with a case where an edge is provided between the counter bore 12b and the outer ring raceway groove 12a, it is possible to facilitate dimension control of the catching margin Δr (since dimensional accuracy is easy to obtain during grinding, and the straight portion is provided, it is easy to measure a dimension D of the catching margin) and to suppress occurrence of ball scratches during assembling of the bearing.
In the embodiment, if a minor angle formed relative to the rotation axis s of the ball 13 by a straight line that connects the center O of the ball 13 and an intersection point between the centerline X of the radial hole 15 and the surface of the ball 13 is defined as λ, the axial position of the centerline X of the radial hole 15 may also be set to satisfy the relationship of 0°<λ≤60°, and preferably a relationship of 30°≤λ≤60°. Thus, in the ball bearings used for high speed applications, more stable lubricity and rotation performance can be obtained.
Specifically, in the embodiment, the axial position of the centerline X of the radial hole 15 may be designed at any one of the positions shown in
In
Since the radial hole 15 is provided in such a position, it is unnecessary to consider the burrs or the like of the opening of the oil hole when machining the raceway groove, and the processability of the outer ring raceway groove 12a is good. Since the opening of the oil hole does not overlap the contact ellipse between the ball 13 and the outer ring raceway groove 12a when the bearing is assembled to the spindle or during use, it is possible to suppress occurrence of ball scratches and abnormal increase in contact surface pressure.
In
In
Specifically, in the bearings of
In
In
The present invention is not limited to the above-described embodiments and may be appropriately modified, improved, or the like.
For example, the radial hole may penetrate the outer ring from the outer circumferential surface thereof to the inner circumferential surface thereof in the radial direction, and may be inclined in the circumferential direction in addition to the one formed along the radial direction (parallel to a radial cross-sectional plane) of the embodiment.
In the above embodiments, although the outer ring 12 is provided with one radial hole, it is not limited thereto and the outer ring 12 may be provided with a plurality of radial holes.
For lubricating oil supply to the radial hole of the outer ring, oil mist lubrication may be adopted in addition to the oil air lubrication. In some cases, oil jet lubrication can also be adopted. However, in a case of a grease supply method in which grease is fed from the radial hole 15 of the outer ring 12 by using a lubricant supply device on a peripheral portion of the bearing and the outside of the spindle, when the radial hole 15 is opened in the outer ring raceway groove 12a, the intersection portion between the counter bore 12b and the outer ring raceway groove 12a, or in the vicinity of the counter bore 12b on the outer ring raceway groove side, the grease that is a semisolid containing a thickener is supplied into the outer ring raceway groove 12a.
In this case, since the grease is caught in the outer ring raceway groove 12a, problems such as an increase in torque and abnormal heat generation occur due to stirring resistance. In particular, these problems are likely to occur in high speed rotation as in this embodiment. Accordingly, oil lubrication method for supplying a lubricant in which a thickener is not contained is desirable in the present invention.
Further, the ball bearing of the present invention is not limited to one applied to a machine tool spindle device, and can also be applied as a ball bearing of a general industrial machine or a high speed rotation device such as a motor.
The present application is based on a Japanese Patent Application No. 2016-150501 filed on Jul. 29, 2016, contents of which are incorporated herein as reference.
Number | Date | Country | Kind |
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2016-150501 | Jul 2016 | JP | national |
Filing Document | Filing Date | Country | Kind |
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PCT/JP2017/027105 | 7/26/2017 | WO | 00 |