JP2006077814A - Spindle rotation support device for machine tool - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To realize a structure capable of preventing skidding damage from occuring on a rolling bearing 6b at a base end side supporting a base end of a spindle and preventing excessive lubricating oil at low costs. <P>SOLUTION: A cylindrical roller bearing is used as the rolling bearing 6b at the base end side. A retainer 19 comprising the rolling bearing 6b at the base end side has pockets 18a and 18b at a plurality of circumferential locations. The number of pockets 18a and 18b is set to be an even number which is twice the number of cylindrical rollers 16 and 16, and the pockets 18b and 18b not internally retaining the cylindrical rollers 16 and 16 are arranged at every other location. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention is incorporated in various machine tools such as a machining center, and rotates to support a spindle that rotates at high speed with a tool or a work piece supported and fixed at the tip thereof, with respect to a fixed housing. The present invention relates to an improvement of a support device.

  The spindle of the machine tool rotates at a high speed during use, and is affected by the heat generated at the contact portion (working portion) between the tool and the workpiece, so that its entire length expands and contracts. That is, based on the centrifugal force accompanying high-speed rotation, the main shaft is elastically deformed in a direction in which its entire length is reduced instead of increasing its diameter. Further, due to the influence of the heat, the entire length of the main shaft becomes longer due to thermal expansion or becomes shorter due to thermal contraction. On the other hand, it is important for precision machining to prevent the position of the tip of the main shaft from being displaced regardless of changes in the rotational speed and temperature of the main shaft.

  In particular, in the case of a so-called motor built-in type machine tool in which the rotor of the electric motor is fixed to the outer peripheral surface of the intermediate portion of the main shaft and this main shaft itself functions as the rotating shaft of the electric motor, The rotation speed of this is fast, and the change speed (acceleration, deceleration) is remarkable. Therefore, in the case of the main shaft of such a motor built-in type machine tool, it is necessary to use a rolling bearing for supporting the main shaft that can withstand high-speed operation. In addition, since the change in the overall length of the main shaft accompanying the change in the operating condition of the machine tool and the heat generation of the rotor and the stator constituting the electric motor becomes significant, the structure needs to smoothly absorb this change. There is.

  For this reason, conventionally, the tip end portion of the main shaft is supported on the housing by the tip side rolling bearing unit to prevent axial displacement of the main shaft and to support the radial load applied to the front end portion of the main shaft. A radial load applied to the base end portion of the main shaft is supported by a base end side rolling bearing on the base end side of the main shaft, but the axial displacement of the main shaft is supported in an allowable manner. 16 to 17 show two examples of a conventional structure of a spindle support device for a spindle of a machine tool having such a function.

  First, in the case of the first example shown in FIG. 16, the tip end portion (left end portion in FIG. 16) of the main shaft 1 is combined with a pair of angular type (or deep groove type) ball bearings 2 and 2 on the back side. It is supported with respect to the fixed housing 4 by the tip side rolling bearing unit 3 formed as described above. The inner rings constituting each of the ball bearings 2 and 2 are externally fitted and fixed to the outer peripheral surface of the tip end portion of the main shaft 1 by an interference fit. Similarly, the outer rings are fixed and fitted to the inner peripheral surface of the housing 4 by a clearance fit. The axial displacement is prevented by the step portion 22 provided on the inner peripheral surface of the housing 4 and the presser lid 23 fitted and fixed to the housing 4. Accordingly, the tip side rolling bearing unit 3 supports the tip end portion of the main shaft 1 so as to prevent axial displacement and to support a radial load applied to the tip end portion of the main shaft 1 so as to be supported. On the other hand, the base end side of the main shaft 1 (the right end side in FIG. 16) is connected to the base end side rolling bearing unit 6 formed by combining a pair of angular type (or deep groove type) ball bearings 5 and 5 on the back side. It supports with respect to the fixed housing 4a. The housing 4a may be integral with or separate from the housing 4. The inner rings constituting each of the ball bearings 5 and 5 are externally fitted and fixed to the outer peripheral surface of the base end portion of the main shaft 1 by an interference fit. Similarly, the outer rings are internally fitted to the inner peripheral surface of the housing 4a by a gap fit. Yes. Therefore, the base end side rolling bearing unit 6 supports the radial load applied to the base end portion of the main shaft 1 but supports the axial displacement of the main shaft 1 in an allowable manner. Further, a rotor 9 is fixed between the distal end side rolling bearing unit 3 and the proximal end side rolling bearing unit 6 at an intermediate portion of the main shaft 1, and an outer peripheral surface of the rotor 9 and a fixed portion such as a housing are fixed. The electric motor 11 is configured to face the inner peripheral surface of the supported stator 10. During operation of the machine tool, the stator 10 is energized to rotate the main shaft 1 at a high speed.

  In the case of the second example shown in FIG. 17, the tip end portion (left end portion in FIG. 17) of the main shaft 1 is combined with a pair of angular type (or deep groove type) ball bearings 2 and 2 in parallel. It supports with respect to the fixed housing 4 by the front end side rolling bearing unit 3a comprised. The inner rings constituting each of the ball bearings 2 and 2 are externally fitted and fixed to the outer peripheral surface of the tip end portion of the main shaft 1 by an interference fit. Similarly, the outer rings are fixed and fitted to the inner peripheral surface of the housing 4 by a clearance fit. The axial displacement is prevented by the step portion 22 provided on the inner peripheral surface of the housing 4 and the presser lid 23 fitted and fixed to the housing 4. Therefore, the distal end side rolling bearing unit 3a prevents the axial displacement of the distal end portion of the main shaft 1 toward the proximal end side (the right side in FIG. 4) and supports the radial load applied to the distal end portion of the main shaft 1 so as to be supported. is doing. On the other hand, the base end side of the main shaft 1 (the right end side in FIG. 4) is a base end side rolling bearing unit 6a formed by combining a pair of angular type (or deep groove type) ball bearings 5 and 5 in parallel. It supports with respect to the fixed housing 4a. The direction of the contact angle of each ball bearing 5, 5 on the proximal end side of the main shaft 1 is opposite to the direction of the contact angle of each ball bearing 2, 2 on the distal end side of the main shaft 1. Further, the inner rings constituting each of the ball bearings 5 and 5 are externally fitted and fixed to the outer peripheral surface of the base end portion of the main shaft 1 by an interference fit. The sleeve 7 is internally fitted and fixed. Further, a spring 8 is provided between the sleeve 7 and the housing 4a so that the sleeve 7 is given elasticity in the direction in which the preload of the ball bearings 5 and 5 is applied. Therefore, the base end side rolling bearing unit 6a supports the base end portion of the main shaft 1 with a radial load applied to the base end portion, but supports the axial displacement of the main shaft 1 in an allowable manner. The attachment of the electric motor 11 to the intermediate part of the main shaft 1 is the same as in the case of the first example shown in FIG.

  As described above, in the case of the conventional high-speed rotation support device for machine tool main shafts, the ball bearings 5 and 5 are used for the base end side rolling bearing units 6 and 6a for supporting the base end side of the main shaft 1. is doing. This is because, in the case of a rotation support device for a spindle of a motor built-in type machine tool, the load applied to the base end side rolling bearing units 6 and 6a is small, and skidding is performed when a general roller bearing is used. This is because damage is likely to occur. That is, the load (cutting force) generated in the processing of the workpiece, which occurs at the contact portion between the tool or workpiece supported on the tip of the spindle 1 and the workpiece or tool, is almost on the tip side. The rolling bearing units 3 and 3a are supported and hardly transmitted to the base end side rolling bearing units 6 and 6a. And most of the load transmitted to the base end side rolling bearing units 6 and 6a is a case of a horizontal machine tool in which the main shaft 1 is arranged in the horizontal direction, and only a part of its own weight of the main shaft 1 is used. is there. Further, in the case of a vertical machine tool in which the main shaft 1 is arranged in the vertical direction, the base end side rolling bearing units 6 and 6a do not apply even the own weight of the main shaft 1 and only an unbalanced load. Very light load.

  In this way, when a general roller bearing is used as the base end side rolling bearing unit 6 or 6a to which only a light load is applied during high-speed rotation, the rolling surface of the roller that does not rotate when pressed against the outer ring raceway by centrifugal force is obtained. The above-mentioned skid damage that wears by rubbing against the inner ring raceway occurs. Such skid damage occurs not only in roller bearings but also in ball bearings. In the case of ball bearings, the preload is applied to prevent skid damage and to support the spindle 1 that rotates at high speed. Can be done. Therefore, conventionally, ball bearings are used as the base end side rolling bearing units 6 and 6a. Conversely, in general, in the case of roller bearings, if preload is applied (negative internal clearance) to prevent skid damage, damage such as seizure occurs with high-speed rotation. For this reason, conventionally, a roller bearing is rarely used as a base end side rolling bearing for supporting the base end portion of the main shaft 1 of a high-speed machine tool.

  However, in the case of the first example shown in FIG. 16 among the conventional structures shown in FIGS. 16 to 17 described above, the axial displacement of the outer rings of the ball bearings 5 and 5 accompanying the expansion and contraction of the main shaft 1 is not necessarily limited. It may not be performed smoothly. That is, the entire length of the main shaft 1 expands and contracts with changes in rotational speed and temperature, and the outer ring of each of the ball bearings 5 and 5 slides in the axial direction with respect to the inner peripheral surface of the housing 4a. Since the outer peripheral surface of these outer rings and the inner peripheral surface of the housing 4a are not necessarily smooth surfaces and the lubrication state may not be good, the axial displacement of the outer ring may not be performed smoothly. In particular, when the rotational speed of the main shaft 1 changes abruptly, there is a high possibility that such an axial displacement becomes defective. If the axial displacement is not smoothly performed, the outer peripheral surface edge of each outer ring meshes with the inner peripheral surface of the housing 4a, and a stick is generated. The preload becomes excessive, and there is a possibility that the ball bearings 5 and 5 may be damaged such as seizure.

  On the other hand, in the case of the second example shown in FIG. 17, in addition to the problem that the axial displacement as described above cannot always be performed smoothly, resonance occurs at the base end side rolling bearing unit 6a. In addition, there is a problem that the structure is complicated and the cost is increased. That is, the sleeve 7 that is pressed in the axial direction by a certain elasticity by the spring 8 resonates at a specific rotational speed based on the minute vibration of the main shaft 1 that rotates at high speed, and axial vibration is likely to occur. Further, since the sleeve 7 and the spring 8 are provided, the structure is complicated and the cost is increased. A ball spline may be provided between the outer peripheral surface of the sleeve 7 and the inner peripheral surface of the housing 4a in order to smoothly perform the axial displacement of the sleeve 7 accompanying the expansion and contraction of the main shaft 1. The structure becomes more complex and the cost is higher.

  In view of such circumstances, as disclosed in Patent Document 1, in recent years, as a base end side rolling bearing, an inner ring raceway provided on the outer peripheral surface of the inner ring, and an outer ring raceway provided on the inner peripheral surface of the outer ring, A rotary support device for a spindle of a machine tool is incorporated that incorporates a non-circular cylindrical roller bearing in which at least one of the tracks is assembled to at least the outer peripheral surface of the main shaft and the inner peripheral surface of the housing. . According to such a rotation support device for a spindle of a machine tool disclosed in Patent Document 1, the distance between both raceways of a cylindrical roller bearing that is a base end side rolling bearing can be narrowed only in a part in the circumferential direction. The rolling surface of the cylindrical roller existing in the portion can be elastically pressed against these both tracks simultaneously (with a preload applied). For this reason, a plurality of cylindrical rollers constituting the cylindrical roller bearing does not cause a so-called revolving slip that revolves without rotating, thereby preventing the occurrence of the skid damage.

  In addition, by using a cylindrical roller bearing that incorporates ceramic cylindrical rollers as the base-end rolling bearing that constitutes the rotation support device for the spindle of the machine tool, damage such as seizure occurs even when preload is applied. It is also considered that the main shaft can be supported at a higher speed.

  As described above, in the case of the structure disclosed in Patent Document 1, at least one of the outer ring raceway and the inner ring raceway is used as the base end side rolling bearing, at least the outer peripheral surface of the main shaft and the inner periphery of the housing. By using a non-circular cylindrical roller bearing when assembled to the surface, it is possible to prevent the occurrence of skid damage. However, when implementing the structure disclosed in Patent Document 1, it is necessary to make at least one of the outer ring raceway and the inner ring raceway non-circular in the assembled state. Manufacturing may be cumbersome and cost may increase. In addition, when using a cylindrical roller bearing that incorporates a ceramic cylindrical roller in the rotation support device for the spindle of a machine tool, the number of cylindrical rollers is lower because the ceramic cylindrical roller is less productive than the ceramic ball. Is the same as that generally used in the past, the manufacturing cost will be significantly increased. For example, in the case of a cylindrical roller bearing incorporating a ceramic cylindrical roller, it is manufactured several times the manufacturing cost of this ball bearing compared to a ball bearing using ceramic balls having the same diameter as this cylindrical roller. costly.

  Also, in the rotation support device for the main spindle of the machine tool, grease lubrication is used for the base end side rolling bearing in order to keep the heat generation sufficiently small, or the base end side rolling of the machine tool that rotates the main shaft at a higher speed. In the case of a bearing, oil-air lubrication or oil mist lubrication is used. However, when oil-air lubrication or oil mist lubrication is used, if the lubricating oil supplied to the inside of the base end side rolling bearing is not smoothly discharged to the outside, the lubricating oil becomes excessive. For this reason, the internal friction resistance of the lubricating oil is increased, and the temperature of the rolling bearing rises with the increase of the heat of stirring due to the stirring resistance. It occurs in. In particular, when a cylindrical roller bearing is used as the base end side rolling bearing, the seizure is more likely to occur.

  On the other hand, when a general cylindrical roller bearing is used for the rotation support device for the spindle of the machine tool, the moment generated by the dynamic unbalance of each cylindrical roller, and each of these cylinders accompanying the revolving motion of each of these cylindrical rollers. When the posture of each of these cylindrical rollers changes due to the moment generated with the fluctuation of the skew angle of the rollers, the rotation and revolution of these cylindrical rollers become unstable. As a result, the friction between the end surfaces of these cylindrical rollers and the flanges formed on the races increases, which causes abnormal wear and seizure at the contact portions between these end surfaces and the flanges. .

Under such circumstances, as disclosed in Patent Document 2, the gyro moment S of each cylindrical roller is determined by the moment U generated by the dynamic unbalance of each cylindrical roller and the skew angle of each cylindrical roller. Each cylinder has a sliding contact point between the end face of each cylindrical roller and the inner side surface of the flange, based on each moment S, U, and C. The product Q · V of the force Q acting in the direction in which the end face of the roller and the inner surface of each flange are pressed against each other and the sliding speed V at this sliding contact is 588 N · m / s (= 60 kgf · m / s) ) The following cylindrical roller bearings are conceivable, and some of them are actually used in a rotation support device for a spindle of a machine tool that rotates at high speed. According to such a rotation support device for a spindle of a machine tool using the cylindrical roller bearing disclosed in Patent Document 2, it is possible to prevent abnormal wear and seizure from occurring on the axial end surface of the cylindrical roller and the inner surface of the flange. I can plan.
Further, in Patent Document 3, a cylindrical roller bearing is used for a rotation support device for a spindle of a machine tool, and a cage for holding each cylindrical roller is used as an outer ring guide, and a pair of annular portions constituting the cage is provided. An invention is disclosed in which the size of a predetermined coefficient (cage movement guide clearance coefficient) AI defined by the radial length, the axial length of each annular portion, and the PCD of a plurality of cylindrical rollers is regulated. . The invention disclosed in Patent Document 3 as described above is a problem caused by an increase in roller load due to a negative internal clearance of the bearing due to an increase in temperature difference between the inner and outer rings during high-speed rotation of the spindle of the machine tool. The purpose is to eliminate. According to the invention disclosed in Patent Document 3, by restricting the structure and dimensions of the cage as described above, it is possible to improve the lubricity of the cylindrical roller bearing and to suppress the deformation and breakage of the cage. To support the main shaft rotating at high speed. However, in the case of the invention disclosed in Patent Document 3 as described above, many cylindrical rollers are incorporated in the same manner as in the case of cylindrical roller bearings that have been generally used in the past. Repeated stress is applied. For this reason, there is still room for improvement with respect to extending the life when the cylindrical roller bearing is operated with a preload applied. That is, when the cylindrical roller bearing is operated with a preload applied so that the radial gap is negative, the life tends to be shortened as the absolute value of the negative radial gap increases.
As prior art documents related to the present invention, there are Patent Documents 4 and 5 in addition to Patent Documents 1 to 3.

JP 2001-193744 A Japanese Patent No. 3289711 Japanese Patent Laid-Open No. 2004-3577 Utility Model Registration No. 2587897 Japanese Patent Laid-Open No. 11-6526

  The rotation support device for a spindle of a machine tool according to the present invention has a structure that uses a cylindrical roller bearing as a base end side rolling bearing in view of the above-described circumstances, and can rotate at high speed. Low-cost, high-reliability and durability structure that can prevent skidding damage and can prevent excessive lubrication, and can extend the service life even when used with preload applied. Invented to realize the above.

The rotation support device for a spindle of a machine tool according to the present invention is similar to the rotation support device for a spindle of a conventional machine tool described above, and is fixed to a spindle that rotates around its axis, and an outer peripheral surface of a tip portion of the spindle. A distal-end side rolling bearing unit provided between the inner peripheral surface of the housing and preventing the axial displacement of the main shaft and supporting a radial load applied to the distal end portion of the main shaft; and a proximal-end outer peripheral surface of the main shaft; A base-side rolling bearing is provided between the inner peripheral surface of the fixed housing and supports a radial load applied to the base end portion of the main shaft but allows axial displacement of the main shaft.
In particular, in the rotation support device for a main spindle of a machine tool according to the present invention, the base end side rolling bearing is a cylindrical roller bearing. And this base end side rolling bearing is rotatably provided between the inner peripheral surface of the outer ring and the outer peripheral surface of the inner ring, and a cage having a plurality of pockets in the circumferential direction, A plurality of cylindrical rollers rotatably provided between an outer ring raceway provided on the inner peripheral surface of the outer ring and an inner ring raceway provided on the outer peripheral surface of the inner ring while being held in a pocket. Prepare. The number of pockets is an even number that is twice the number of these cylindrical rollers, and the cylindrical rollers are rotatably held only in some of these pockets, and the remaining pockets have members. The remaining pockets are in a state where the inner peripheral surface and the outer peripheral surface of the retainer are communicated with each other, leaving an empty state without providing any.

According to the rotation support device for a spindle of a machine tool of the present invention configured as described above, the following operations and effects (1) to (5) can be achieved.
(1) It is possible to prevent the occurrence of skidding damage to the base end side rolling bearing that rotates at high speed under light load, and to improve the oil drainage of the lubricating oil supplied to the inside, to suppress heat generation and to cause seizure. Can be difficult.
That is, the number of cylindrical rollers of the cylindrical roller bearing that is the base end side rolling bearing is set to 1/2 of the number of pockets, and with the setting of empty pockets, the inertia mass decreases, the rolling resistance decreases, etc. As a result, the rotational resistance decreases, and the load applied to one cylindrical roller increases. As a result, the so-called revolving slip, in which each cylindrical roller slides around the inner ring while sliding without sufficiently rolling, is less likely to occur, and skidding damage is less likely to occur. In addition, as the empty pockets are set, the lubricating oil supplied to the inside circulates through these empty pockets. Moreover, the space | interval of adjacent cylindrical rollers can be expanded regarding the circumferential direction of a base end side rolling bearing. Therefore, the lubricating oil can be easily discharged to the outside, and the state of excessive lubricating oil can be eliminated. In addition, since the number of cylindrical rollers is reduced, the amount of heat generated inside the proximal end side rolling bearing is reduced. As a result, the allowable seizure limit with respect to the negative radial gap increases, and seizure can be hardly caused even if the proximal-side rolling bearing is operated with some negative radial gap.
(2) Compared to the case where cylindrical rollers are held in all pockets, the number of repeated stresses is reduced by the smaller number of cylindrical rollers, and the service life even when operating with a negative radial gap and preload applied. Will grow.
(3) The axial displacement of the base end side of the main shaft accompanying the expansion and contraction of the main shaft can be smoothly performed, so that sticking can be hardly generated.
That is, the rolling surfaces of the plurality of cylindrical rollers, the outer ring raceway, and the inner ring raceway that slide with each other during the axial displacement are finished to be smooth and sufficiently lubricated. Therefore, the resistance to the axial displacement is extremely small, and the axial displacement can be smoothly performed to prevent the stick from being generated.
(4) Since the base end side rolling bearing is provided, a sleeve, a spring, and a ball spline are not required, so that the structure can be simplified and the cost can be reduced.
(5) Since no preload is applied to the base end side rolling bearing by a spring, minute vibrations of the main shaft and the base end side rolling bearing do not resonate regardless of fluctuations in the rotational speed of the main shaft. In addition, harmful vibrations such as chatter noise and noise can be prevented.
As a result, according to the rotation support device for a spindle of a machine tool of the present invention, even when the rotation speed of the spindle is increased in order to increase the work efficiency, the reliability and durability of the rotation support portion of the spindle are reduced with an inexpensive structure. Enough to secure.

When implementing this invention, Preferably, as described in Claim 2, each cylindrical roller which comprises the cylindrical roller bearing which is a base end side rolling bearing shall be made from ceramics.
According to this preferable configuration, it is possible to make it less likely to cause skid damage in the base end side rolling bearing. Moreover, since the number of cylindrical rollers is ½ of the number of pockets constituting the cage, an increase in cost can be kept low despite the use of expensive ceramic cylindrical rollers.

More preferably, as described in claim 3, the diameter of the rolling surface of each cylindrical roller constituting the cylindrical roller bearing which is the base end side rolling bearing is D, and the axial length of each cylindrical roller is defined as D. When L is L, these cylindrical rollers are short rollers that satisfy 0.5 ≦ L / D <1.
According to this more preferable configuration, it is possible to prevent abnormal wear and seizure from occurring on the axial end surface of the cylindrical roller and the inner surface of the flange formed on the race. Further, the manufacturing work of the base end side rolling bearing can be easily performed, and a sufficient load capacity can be secured.

  1 to 3 show an embodiment of the present invention. The tip end portion (left end portion in FIG. 1) of the main shaft 1 is fixed to a fixed housing 4 by a tip side rolling bearing unit 3 formed by combining a pair of angular type (or deep groove type) ball bearings 2 and 2 on the back side. I support it. The inner rings constituting each of the ball bearings 2 and 2 are externally fitted and fixed to the outer peripheral surface of the tip end portion of the main shaft 1 by an interference fit. Similarly, the outer rings are fixed and fitted to the inner peripheral surface of the housing 4 by a clearance fit. The axial displacement is prevented by the step portion 22 provided on the inner peripheral surface of the housing 4 and the presser lid 23 fitted and fixed to the housing 4. Accordingly, the tip side rolling bearing unit 3 supports the tip end portion of the main shaft 1 so as to prevent axial displacement and to support a radial load applied to the tip end portion of the main shaft 1 so as to be supported.

  On the other hand, the base end side (the right end side in FIG. 1) of the main shaft 1 is supported with respect to the fixed housing 4a by a base end side rolling bearing 6b as detailed in FIGS. The housing 4a may be integral with or separate from the housing 4. In the case of the rotation support device for a main spindle of a machine tool according to the present invention, the base-side rolling bearing 6b includes an inner ring 13 having a cylindrical inner ring raceway 12 on its outer peripheral surface and a cylindrical outer ring on its inner peripheral surface. A cylindrical roller bearing including an outer ring 15 having a raceway 14 and a plurality of cylindrical rollers 16 and 16 provided between the inner ring raceway 12 and the outer ring raceway 14 so as to roll freely. In the case of the present embodiment, these cylindrical rollers 16 and 16 are made of ceramics. Among the components of such a base side rolling bearing 6b, flanges 17 and 17 are formed at positions where the inner ring raceway 12 is sandwiched from both sides in the axial direction at both ends of the outer peripheral surface of the inner ring 13. The cylindrical rollers 16 and 16 are disposed between the flanges 17 and 17 to prevent the cylindrical rollers 16 and 16 from being displaced in the axial direction with respect to the inner ring 13. On the other hand, no flange is formed on the inner peripheral surface of the outer ring 15, and the cylindrical rollers 16, 16 can be displaced in the axial direction with respect to the outer ring 15. Therefore, the axial displacement of the inner ring 13 and the outer ring 15 is free.

  Further, the base-side rolling bearing 6b includes a cage 19 (shown in FIGS. 2 and 3 and omitted in FIG. 1) between the inner peripheral surface of the outer ring 15 and the outer peripheral surface of the inner ring 13. It is provided so that it can rotate freely. The retainer 19 has pockets 18a and 18b at a plurality of locations (12 locations in the illustrated example) at equally spaced positions in the circumferential direction. The cylindrical rollers 16 and 16 are held only in a part of the pockets 18a and 18b of the cage 19.

  That is, in the case of the present invention, the cylindrical rollers 16, 16 are provided by a half of the number of the pockets 18a, 18b (six in the illustrated example). In other words, the retainer 19 is provided with an even number of pockets 18a and 18b (12 in the illustrated example) that is twice the number of the cylindrical rollers 16 and 16. The cylindrical rollers 16 and 16 are not held in all the pockets 18a and 18b, but the cylindrical rollers 16 and 16 are held only in every other pocket 18a and 18a. That is, the pockets 18 a and 18 a holding the cylindrical rollers 16 and 16 and the pockets 18 b and 18 b not holding the cylindrical rollers 16 and 16 are alternately arranged in the circumferential direction of the cage 19.

  In the case of the present embodiment, the diameter (outer diameter) of the rolling surface of each of the cylindrical rollers 16, 16 is D, and the axial length of each of the cylindrical rollers 16, 16 is L (D , L refer to FIG. 7), and the cylindrical rollers 16 and 16 are short rollers that satisfy 0.5 ≦ L / D <1. For example, the dimensions of the cylindrical rollers 16 and 16 are regulated so as to satisfy L / D≈0.56.

  Of the base end side rolling bearing 6b as described above, the inner ring 13 is externally fixed to the outer peripheral surface of the base end portion of the main shaft 1 by an interference fit, and the outer ring 15 is also tightened to the inner peripheral surface of the housing 4a. The inner fitting is fixed by fitting. Therefore, the base-side rolling bearing 6b supports the base end portion of the main shaft 1 with a radial load applied to the base end portion, but the shaft of the main shaft 1 is based on the axial displacement of the inner ring 13 and the outer ring 15. Supports directional displacements freely.

  Further, in the illustrated example, a rotor 9 is fixed between the distal end side rolling bearing unit 3 and the proximal end side rolling bearing 6b at an intermediate portion of the main shaft 1, and an outer peripheral surface of the rotor 9 and a housing, etc. The electric motor 11 is configured to face the inner peripheral surface of the stator 10 supported by the fixed portion. During operation of the machine tool, the stator 10 is energized to rotate the main shaft 1 at a high speed.

According to the rotation support device for a spindle of a machine tool of the present invention configured as described above, the following operations and effects (1) to (5) can be achieved.
(1) The base-end side rolling bearing 6b that rotates at a high speed in a light load state is less likely to cause skid damage, and the oil drainage of the lubricating oil supplied to the inside can be increased, heat generation can be reduced, and seizure can be prevented. Can be difficult to generate.
That is, the number of cylindrical rollers 16 and 16 of the cylindrical roller bearing which is the base end side rolling bearing 6b is ½ of the number of pockets 18a and 18b, and the empty pockets 18b and 18b are set. Due to a decrease in inertial mass, a decrease in rolling resistance, and the like, the rotational resistance decreases, and the load applied to one cylindrical roller 16 increases. As a result, it is possible to prevent the so-called revolving slip, in which the cylindrical rollers 16 and 16 rotate around the inner ring 12 while sliding without sufficiently rolling, and to prevent skid damage. Further, as the empty pockets 18b and 18b are set, the lubricating oil supplied to the inside flows through these empty pockets 18b and 18b. Moreover, the space | interval of adjacent cylindrical rollers 16 and 16 can be expanded regarding the circumferential direction of the base end side rolling bearing 6b. Therefore, the lubricating oil can be easily discharged to the outside, and the state of excessive lubricating oil can be eliminated. Further, since the number of cylindrical rollers 16, 16 is reduced, the amount of heat generated inside the proximal end side rolling bearing 6b is reduced. As a result, the allowable seizure limit with respect to the negative radial gap increases, and seizure can hardly occur even if the proximal-side rolling bearing 6b is operated with some negative radial gap.

(2) Compared to the case where the cylindrical rollers 16 and 16 are held in all the pockets 18a and 18b, the number of repeated stresses is reduced by the smaller number of the cylindrical rollers 16 and 16, and preload is applied with a negative radial gap. Life is extended even when operated under load.
(3) The axial displacement of the base end side of the main shaft 1 accompanying the expansion and contraction of the main shaft 1 can be smoothly performed, so that sticks can be hardly generated.
That is, the rolling surfaces of the plurality of cylindrical rollers 16, 16 that slide on each other during the axial displacement, the outer ring raceway 14, and the inner ring raceway 12 are finished smoothly and are sufficiently lubricated. Therefore, the resistance to the axial displacement is extremely small, and the axial displacement can be smoothly performed to prevent the stick from being generated.
(4) Since the base end side rolling bearing 6b is provided, the sleeve 7, the spring 8, and the ball spline are not required as in the second example of the conventional structure shown in FIG. Can be simplified to reduce the cost.
(5) Since no preload is applied to the base end side rolling bearing 6b by a spring, the minute vibration of the main shaft 1 and the base end side rolling bearing 6b resonate regardless of the rotational speed fluctuation of the main shaft 1. It is possible to prevent generation of harmful vibration such as chatter noise and noise.

  As a result, according to the rotation support device for a main spindle of a machine tool of the present invention, even when the rotation speed of the main spindle 1 is increased in order to increase the work efficiency, the reliability of the rotation support portion of the main spindle 1 is low. And sufficient durability can be secured.

  In the case of the present embodiment, the cylindrical rollers 16 and 16 constituting the cylindrical roller bearing which is the base end side rolling bearing 6b are made of ceramics. For this reason, it is possible to make it difficult to cause skid damage in the base end side rolling bearing 6b. Moreover, since the number of the cylindrical rollers 16 and 16 is ½ of the number of the plurality of pockets 18a and 18b constituting the retainer 19, the use of the expensive ceramic cylindrical rollers 16 and 16 is involved. Therefore, the cost increase can be kept low.

  Furthermore, in the case of the present embodiment, the diameter of the rolling surface of each cylindrical roller 16, 16 constituting the cylindrical roller bearing which is the base end side rolling bearing 6b is D, and the axial direction of each cylindrical roller 16, 16 is D. When the length is L, the cylindrical rollers 16 and 16 are short rollers that satisfy 0.5 ≦ L / D <1. For this reason, it is possible to prevent abnormal wear and seizure from occurring between the axial end surfaces of the cylindrical rollers 16 and 16 and the inner surfaces of the flanges 17 and 17 formed on the outer peripheral surfaces of both ends of the inner ring raceway 12. Next, this reason will be described.

  When a general cylindrical roller bearing is used, it is inevitable that a so-called skew occurs in which each cylindrical roller rotates while the central axis of each cylindrical roller is not parallel to the central axis of the inner ring and outer ring. . When such a skew occurs, some measures are taken because the outer peripheral edge of each cylindrical roller and the inner peripheral surface of the inner ring or the inner peripheral surface of the outer ring are in sliding contact with each other. Unless otherwise, under severe conditions such as poor lubrication, significant wear (abnormal wear) may occur at both ends of the cylindrical rollers and the ridges.

In particular, the product of the pitch circle diameter d _m (mm) and the rotational speed _{^{n (min -1), d m}} · n is used in such conditions over 1.5 million, in the case of high-speed cylindrical roller bearings Since problems such as abnormal wear and seizure are likely to occur, this is an obstacle to increasing the speed of cylindrical roller bearings. In particular, in machine tools, there is a demand to increase the rotational speed of the spindle supported by cylindrical roller bearings in order to achieve higher performance. It is desired to reduce the amount of lubricant supplied. For these reasons, the above problems are likely to occur in the field of machine tools.

  Further, the cylindrical roller bearing for supporting a large radial load is not used in a state where the load is uniformly applied over the entire circumference. In other words, the cylindrical roller bearing in use is a load zone in which a part in the circumferential direction supports a radial load, and a non-load zone that does not receive a radial load on the opposite side in the diameter direction. Accordingly, during the operation of the cylindrical roller bearing, each cylindrical roller alternately passes through the load zone and the non-load zone along with the revolution movement. Since each of these cylindrical rollers is strongly held between the inner ring raceway and the outer ring raceway while being in the load zone, the posture (skew angle) hardly changes and remains in a stable state. Revolves while rotating. On the other hand, each of the cylindrical rollers located in the non-load zone is relatively free to change its posture without being restricted by the inner ring raceway and the outer ring raceway.

  The posture of each cylindrical roller while it is located in the non-load zone is the moment generated by the dynamic unbalance of each cylindrical roller, and the skew angle of each cylindrical roller accompanying the revolution motion of each cylindrical roller. If it changes due to the moment generated along with the fluctuation, the rotation and revolution of these cylindrical rollers become unstable. As a result, the friction between the end faces of these cylindrical rollers 16 and 16 and the heel increases, and the abnormal wear and seizure are likely to occur.

On the other hand, in the case of the present embodiment, as described above, the diameter of the rolling surface of each of the cylindrical rollers 16 and 16 constituting the cylindrical roller bearing that is the base end side rolling bearing 6b is D, When the axial length of the cylindrical rollers 16 and 16 is L, these cylindrical rollers 16 and 16 are short rollers that satisfy 0.5 ≦ L / D <1. Thus, it is known that when the dimensions of the cylindrical rollers 16 and 16 are regulated, the following relations [1] and [2] are satisfied.
[1] The gyro moment S generated in each cylindrical roller 16, 16 is accompanied by the variation in the moment U generated by the dynamic unbalance of each cylindrical roller 16, 16 and the skew angle of each cylindrical roller 16, 16. Or more than the sum of the generated moment C (S ≧ U + C).
[2] The cylindrical rollers 16 and 16 are skewed, and a pair of flanges 17 and 17 (one pair on both sides of the outer ring raceway) provided on the end surfaces of the cylindrical rollers 16 and 16 and on both sides of the inner ring raceway 12. When the flanges are provided, the end surfaces of the cylindrical rollers 16 and 16 and the flanges 17 are slidably contacted with the inner surfaces of the respective flanges) based on the moments S, U and C. , 17, the product Q · V of the force Q acting in the direction in which the inner surfaces of the 17 and 17 are pressed against each other and the sliding velocity V at the sliding contact is 588 N · m / s (= 60 kgf · m / s) or less. .

Of the moments S, U, and C, the gyro moment S is the center point of the cylindrical rollers 16 and 16 with the inertia moment about the rotation axis X of the cylindrical rollers 16 and 16 being I _X (FIG. 5). The inertia moment about the Z axis perpendicular to the rotation axis X of each cylindrical roller 16, 16 is I _Z (FIG. 5), the revolution angular velocity of each cylindrical roller 16, 16 is ω _C, and each cylindrical roller The rotation angular velocity of 16, 16 is ω _B , the skew angle of each cylindrical roller 16, 16 is positive (> 0) is ψ (see FIG. 10), and each of these cylindrical rollers 16, 16 rotates once. The variation of the skew angle during this time is assumed to be Δψ, and the center point of each cylindrical roller 16, 16 generated by the dynamic imbalance of each cylindrical roller 16, 16 is set to the rotation axis of each cylindrical roller 16, 16. Maume around the axis orthogonal to The door in the case of the I _U,
S = I _X · ω _C · ω _B · sinψ- (I _X -I _Z ) · ω _C ² · sinψ · cosψ --- (1)
It is represented by
Further, the maximum value U of the moment about the Z-axis generated by the dynamic unbalance of the cylindrical rollers 16, 16 is
U = I _U · ω _B ^2- (2)
It is represented by
Further, when the cylindrical rollers 16 are rotated, the end surfaces of the cylindrical rollers 16 and 16 are in sliding contact with the inner surfaces of the flanges 17 and 17 due to the skew of the cylindrical rollers 16 and 16. The maximum value C of the moment generated along with the fluctuation of the skew angle of each of these cylindrical rollers 16, 16 is
C = I _Z · Δψ · ω _B ^2- (3)
It is represented by

  When the relationship [1] is satisfied, the gyro moment S keeps the postures of the cylindrical rollers 16 and 16 stable. Further, when the relationship [2] is satisfied, the gyro moment S can be prevented from becoming excessively large, and the sliding contact between the end faces of the cylindrical rollers 16 and 16 and the inner faces of the flanges 17 and 17 can be prevented. By suppressing the friction, it is difficult to cause abnormal wear and seizure.

  These points will be described in more detail below. The following description will be made when the flanges 17 and 17 are formed on the outer peripheral surface of the inner ring 13. The same applies to the case where a collar is formed on the inner peripheral surface of the outer ring 15. First, the gyro moment S is set to be equal to or greater than the sum of the maximum value U of the moment and the maximum value C of the moment (S ≧ U + C), so that the postures of the cylindrical rollers 16 and 16 are stabilized. . In other words, in this embodiment, the gyro moment S generated when the cylindrical rollers 16 and 16 are skewed is positively used as a control factor for stabilizing the posture of the cylindrical rollers 16 and 16 located in the non-load zone. It is what you use.

In the case of the base end side rolling bearing 6b that constitutes the rotation support device for the main spindle of the machine tool of the present embodiment, the distance between the pair of flanges 17 and 17 formed at both ends of the outer peripheral surface of the inner ring 13 is the cylindrical roller 16, It is slightly larger than the length dimension over 16 axial directions. Therefore, there is a slight gap between the axial end surfaces of the cylindrical rollers 16 and 16 and the inner surfaces of the flanges 17 and 17, and the cylindrical rollers 16 and 16 are based on the gaps. In a skewed state, it revolves while rotating. As described above, the cylindrical rollers 16 and 16 revolve while rotating in a skewed state. As a result, the cylindrical rollers 16 and 16 precess as shown in FIG. 16, a gyro moment S having a magnitude as shown in the equation (1) acts.
In FIG. 4, H represents the angular momentum (vector) when the cylindrical rollers 16 and 16 are precessing. Ω _C represents the revolution angular velocity of each of the cylindrical rollers 16 and 16 as described above.

On the other hand, in the conventional specification of a cylindrical roller bearing that has been conventionally known, when this cylindrical roller bearing is used in a state where the outer ring is stationary and the inner ring is rotated, ω _B / (ω _C · cos ψ)> 1. Therefore, if an external force (a force other than the gyro moment S) is not taken into account, when each cylindrical roller 16 is skewed, the gyro moment S is in the same direction as each cylindrical roller 16 is skewed, that is, This works in a direction to promote skew and increase the skew angle ψ. FIG. 6 shows a range in which the gyro moment S acts in a direction to increase the skew angle ψ when each cylindrical roller 16 is skewed {1-I _Z / I _X ≦ ω _B / (ω _C · cos ψ) } And the range of I _Z / I _X ≧ 0 is indicated by hatching. In FIG. 6, the horizontal axis is the ratio I _Z / I _X of the inertia moments I _X and I _Z of the cylindrical rollers 16 and 16, and the vertical axis is ω _B / (ω _C · cos ψ).

On the other hand, a moment U _V due to dynamic imbalance (unbalance) generated around the Z axis acts on each cylindrical roller 16 with a magnitude represented by the following equation (4).
U _V = U · cos (ω _B t) --- (4)
In the equation (4), U which is the maximum value of the moment U _V is expressed by the equation (2).
The moment I _U is expressed by the following equation (5).
_{I U = (M 6/2} ) · e · L --- (5)
In this equation (5), M ₆ is the mass of each cylindrical roller 16, and e is the center of gravity of each cylindrical roller 16, as shown in FIG. Is the radial eccentricity of each cylindrical roller 16 in a state where it is assumed that the cylindrical roller 16 is present at both ends of each cylindrical roller 16, and L represents the axial length of each cylindrical roller 16, respectively. Yes.

Furthermore, each cylindrical roller 16 revolves while rotating while keeping the state where the outer peripheral edge portions of both axial end surfaces are pressed against the inner side surfaces of the flanges 17 and 17 by skewing. As a result, each cylindrical roller 16 has a moment C _V generated as the skew angle ψ changes based on the shape errors of the cylindrical rollers 16 and the flanges 17 and 17, and the following ( 6) Acts in the size as shown in the equation.
C _V = C · cos (ω _B t + θ) −−− (6)
In the equation (4), θ represents the phase difference of the moment C _V with respect to the moment U _V. In the equation (4), C, which is the maximum value of the moment C _V , is expressed by the equation (3).

Now, it is assumed that the outer peripheral edge portions of both end faces of the cylindrical rollers 16 are rotating in a state of being in contact with the inner side surfaces of the flanges 17 and 17, and the cylindrical rollers 16 are rotated from the flanges 17 and 17. If the received skew moment is T, the following equation regarding the skew angle ψ of each cylindrical roller 16 is obtained. However, the skew moment acting on each cylindrical roller 16 from both the inner ring and outer ring raceways 12 and 14 is ignored as being small.
I _Z · (d ² ψ / dt ² ) = S + U _V −T −−− (7)
Also, the left side of this equation (7) is
I _Z · (d ² ψ / dt ² ) = I _Z · d ² / dt ² {Δψ · cos (ω _B t + θ)}
= I _Z · {-Δψ · ω _B ² · cos (ω _B t + θ)}
= -C _V
Therefore, the above equation (7) can be rewritten as the following equation (8).
T = S + U _V + C _V −−− (8)
Now, considering that θ = 0, that is, the case where U _V and C _V are in phase with each other, the above equation (8) is
T = S + (U + C) .cos (ω _B t) −−− (9)
It becomes. That is, the skew moment T received by the cylindrical rollers 16 from the flanges 17 and 17 periodically varies while the cylindrical rollers 16 rotate once. This is shown in FIG.

Therefore, the skew moment T received by the cylindrical rollers 16 from the flanges 17 and 17 is maximized when the cylindrical rollers 16 are most skewed between the flanges 17 and 17. In the case where the roller 16 is pushed back from each of the flanges 17 and 17 in a direction in which the skew angle ψ is decreased, and each of the cylindrical rollers 16 is most skewed between the flanges 17 and 17, This is a case where the cylindrical rollers 16 are most pressed against the flanges 17 and 17 by the skew moment U _V due to the unbalance of the cylindrical rollers 16. In addition, the skew moment T received by the cylindrical rollers 16 from the flanges 17 and 17 is minimized when the skew angle ψ of the cylindrical rollers 16 is minimized and the skew of the cylindrical rollers 16 is reduced. When the angle ψ is increased, the moment C _V acts on the cylindrical rollers 16 in a direction in which the cylindrical rollers 16 move away from the flanges 17 and 17, and the skew of the cylindrical rollers 16 is increased. This is a case where the skew moment U _V acts on each cylindrical roller 16 in a direction in which each of the cylindrical rollers 16 moves away from the flanges 17 and 17 in a state where the angle ψ is the smallest.

  The case where the relationship S ≧ U + C is established between the moments S, U, and C as described above will be described. Since T ≧ 0 is always satisfied from the above equation (9), the cylindrical roller 16 is always pressed against the flanges 17 and 17 while maintaining a skew angle in a fixed direction, that is, the cylindrical roller 16 is always pressed against the flanges 17 and 17. It will be revolving around while being guided by, and will not leave 鍔 17,17. Thus, since each cylindrical roller 16 revolves while rotating in a stable state by being guided by the flanges 17 and 17, the cylindrical roller bearing has a non-load zone as described above. The cylindrical roller bearing can be operated in a stable state even when used in the above.

  That is, while passing through the non-load zone, each cylindrical roller 16 revolves while rotating with the axial end faces pressed against the inner surfaces of the flanges 17 and 17, and remains in the same posture (skew). Without changing the direction of the angle ψ, the robot enters the load zone again. For this reason, the behavior of each cylindrical roller 16 does not become unstable while each cylindrical roller 16 revolves once.

In addition to the moments U _V and C _V described above, misalignment between the inner ring 13 and the outer ring 15 may be included as factors that may be a force that changes the skew angle ψ of each cylindrical roller 16 during operation of the cylindrical roller bearing. Alternatively, a skew moment induced by a shape error (inclination) of the inner ring raceway 12 or the outer ring raceway 14 is also conceivable. However, skew moment caused by these, in the case where d _m · n is used in a high speed, such as more than 1,500,000, the above moment S, U, much smaller than C. Therefore, even if the skew moment based on the misalignment or the shape error is not taken into consideration, there is no particular problem.

Next, the end surfaces of the cylindrical rollers 16 and the inner surfaces of the flanges 17 and 17 are pressed against each other at the sliding contact points between the axial end surfaces of the cylindrical rollers 16 and the inner surfaces of the flanges 17 and 17. The product Q · V of the directional force Q and the sliding speed V at the sliding contact is set to 588 N · m / s (= 60 kgf · m / s) or less, and the friction at the sliding contact is suppressed and abnormal wear occurs. The point of preventing the occurrence of burn-in will be described.
As described above, when the gyro moment S is set to be equal to or greater than the sum of the maximum values U and C of the respective moments described above, the movement of each cylindrical roller 16 is stabilized, and the abnormal wear and seizure are observed from the surface. Can be prevented. However, if the gyro moment S is too large, the force Q that presses both axial end surfaces of the cylindrical rollers 16 against the inner surfaces of the flanges 17 and 17 becomes too large.

As a result of the excessive force Q, the Q · V value at the sliding contact expressed by the following equation (10) increases, and the abnormal wear and seizure are likely to occur.
Q · V = (M / B) · V --- (10)
In the equation (10), Q indicates that each cylindrical roller 16 is skewed as shown in FIGS. 9 to 10, and the outer peripheral edge portion of each axial end surface of each cylindrical roller 16 and each of the flanges 17 and 17. When the inner surface of the cylindrical roller 16 comes into contact with each other, the force (N) generated in the contact point G in the direction of pressing each other, V is the outer peripheral edge of each cylindrical roller 16 in the axial direction at the contact point G. The sliding speed (m / s) between the portion and the inner surface of each of the flanges 17 and 17 is shown. M is a skew moment (Nm) acting in the direction of decreasing the skew angle ψ of each cylindrical roller 16, and is the sum of the moments S, U, C described above (M = S + U + C). In addition, as shown in FIG. 11, B is the length (m) of the portion where the end surface of each cylindrical roller 16 and the inner surface of the flange 17 abut when each cylindrical roller 16 is skewed. Represents each.

  Under such a premise, the present inventor conducted an experiment to know the influence of the Q · V value on the wear of the axial end surface of each cylindrical roller 16 and the inner surface of each of the flanges 17 and 17. It was. As a result, if the Q · V value is 588 N · m / s (= 60 kgf · m / s) or less (Q · V ≦ 588 N · m / s), both axial end surfaces of the cylindrical rollers 16 and the above It was found that abnormal wear and seizure did not occur on the inner surface of each of the flanges 17 and 17. Therefore, when S ≧ U + C and the upper limit of the Q · V value generated by the sum of these moments (S + U + C) is 588 N · m / s because of the relationship between the moments S, U and C, It was confirmed that the above-mentioned abnormal wear and seizure could hardly occur.

  In FIG. 13, a range satisfying these two requirements is indicated by a diagonal lattice. The horizontal axis in FIG. 13 represents the portion caused by the gyro moment S in the Q · V value, and the vertical axis represents the sum of the moments U and C (U + C) of the Q · V value. ) Represents the parts generated by each.

  In FIG. 13, in a region where (U + C)> S indicated by hatching in a portion that does not satisfy the relationship [1] and [2] described above, as indicated by hatching in FIG. The end face of each cylindrical roller 16 is eccentrically worn. That is, in the region where (U + C)> S, the cylindrical roller 16 moves in an unstable manner with the operation of the cylindrical roller bearing, and therefore, the axial end surfaces of the cylindrical rollers 16 wear eccentrically. On the other hand, in the region where S ≧ U + C but Q · V> 588 N · m / s shown in the satin pattern in FIG. 13, the movement of each cylindrical roller 16 accompanying the operation of the cylindrical roller bearing is Because of stabilization, the axial end surfaces of these cylindrical rollers 16 are worn concentrically as shown by the oblique lines in FIG. However, even if it wears concentrically, as long as Q · V> 588 N · m / s, the amount of wear increases and seizure may occur.

  In the case of the present embodiment, the region shown by the oblique lattice in FIG. 13, that is, S ≧ U + C and Q · V ≦ 588 N · m / s can be obtained. The movement of the cylindrical roller 16 can be stabilized, and the risk of seizure can be reduced by suppressing wear. Then, by using a so-called short roller whose axial dimension L is smaller than the diameter (outer diameter) D of the rolling surface (L / D <1), two such relationships (S ≧ U + C, Q・ It was found that both V ≦ 588 N · m / s) can be satisfied.

  Moreover, each cylindrical roller 16 can be easily processed by satisfying 0.5 ≦ L / D. Further, by satisfying this relationship, a machining clearance (escape groove) is formed in a continuous portion between the inner ring raceway 12 of the inner ring 13 and the inner side surface of the flanges 17 and 17, which are the race rings formed with the flanges 17 and 17. Easy to do. Furthermore, the load capacity of the base end side rolling bearing 6b can be sufficiently secured. On the other hand, unlike the case of the present embodiment, when each cylindrical roller 16 satisfies 0.5> L / D, the manufacturing operation becomes extremely troublesome, and it is extremely difficult to secure a sufficient load capacity. become.

  Next, among the effects of the present embodiment, in order to confirm the effect that the seizure between the axial end surface of each cylindrical roller 16 and the inner surface of the flanges 17 and 17 can be hardly generated, An experiment conducted to confirm the effect when the number of rollers 16, 16 is the same as the number of pockets 18a, 18b of the cage 19 will be described. In the experiment, as shown in FIG. 14A, the diameter (outer diameter) D of the rolling surface is equal to the axial dimension L (see FIG. 7 for D and L) (L / D = 1). A ratio between a cylindrical roller bearing 21a (a comparative example) using a so-called equal-length equal-diameter roller as the roller 16a and the axial dimension L and the diameter (outer diameter) D of the rolling surface as shown in FIG. Is 0.5 or more and less than 1 (0.5 ≦ L / D <1), and a cylindrical roller bearing 21b (reference example) using a so-called short roller as the cylindrical roller 16 was used. The cylindrical roller bearings 21a and 21b used have a nominal number of N1014 and have an outer diameter of 110 mm, an inner diameter of 70 mm, and a width of 20 mm. The number of cylindrical rollers 16 and 16a is 18 in both cases. In such cylindrical roller bearings 21a and 21b, the inner ring 13 was rotated under a small amount of lubrication (oil-air lubrication), and the number of revolutions until seizure was measured. The result is shown in FIG.

  In FIG. 15 showing the experimental results, the ◇ marks indicate the carbonitriding when the isometric isometric roller made of carbonitriding steel as described in Patent Document 5 is used (in the case of the comparative example). The experimental results are shown when steel short rollers are used (in the case of the reference example). As is clear from this experiment, when short rollers are used, the outer ring temperature rise is small and seizure resistance is improved. From such experimental results, in the above-mentioned reference example, the number of cylindrical rollers 16 in which the number of pockets 18a and 18b of the retainer 19 is ½ of the number of pockets 18a and 18b belongs to the range of this embodiment. It was confirmed that the seizure could be made difficult to occur by suppressing the friction at the sliding contact between the end face of each of the flanges and the inner surface of each of the flanges 17 and 17.

  The inner surfaces of the pair of flanges provided on the cylindrical roller bearing that face each other are inclined in a direction in which the distance between the inner surfaces increases toward the outside in the diametrical direction. If the outer peripheral edge of the end face of the cylindrical roller abuts against each of the flanges, if the outer peripheral edge of the end face comes into contact with the inclined inner surface, a strong oil film is formed on the sliding contact portion due to the oil wedge effect. It becomes possible to form. For this reason, the above-mentioned abnormal wear and seizure can be prevented more effectively from occurring even under high-speed and minute lubrication.

1 is a schematic cross-sectional view showing an embodiment of the present invention. The expanded AA sectional view of Drawing 1 which omits and shows a part. BB sectional drawing of FIG. The schematic diagram for demonstrating the gyro moment added to a cylindrical roller at the time of operation | movement of a cylindrical roller bearing. Schematic diagram for explaining the moment of inertia of a cylindrical roller. The graph which shows the range where a gyro moment acts on the direction which tries to enlarge a skew angle when a cylindrical roller skews. The schematic diagram for demonstrating the dynamic imbalance of a cylindrical roller. The graph which shows the relationship between the gyro moment added to a cylindrical roller, and the moment which arises with the fluctuation | variation of the skew angle of this cylindrical roller. The fragmentary sectional view which shows the state which the cylindrical roller skewed. The figure seen from the upper part of FIG. The figure seen from the side of FIG. The end elevations which show two examples of the abrasion state of the end surface of a cylindrical roller. The graph for showing the range of a present Example. The fragmentary sectional view which shows two examples of the cylindrical roller bearing used for the experiment conducted in order to confirm the effect of a present Example. The graph which shows the result of the experiment conducted in order to confirm the effect of a present Example. FIG. 6 is a schematic cross-sectional view showing a first example of a conventional structure. The schematic sectional drawing which shows the 2nd example.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Main shaft 2 Ball bearing 3, 3a Front end side rolling bearing unit 4, 4a Housing 5 Ball bearing 6, 6a Base end side rolling bearing unit 6b Base end side rolling bearing 7 Sleeve 8 Spring 9 Rotor 10 Stator 11 Electric motor 12 Inner ring track 13 Inner ring 14 Outer ring raceway 15 Outer ring 16, 16a Cylindrical roller 17 鍔 18a, 18b Clearance 19 Cage 21a, 21b Cylindrical roller bearing 22 Step 23 Presser lid

Claims (3)

  Provided between the main shaft that rotates about its axis, and the outer peripheral surface of the front end of the main shaft and the inner peripheral surface of the fixed housing, prevent axial displacement of the main shaft and add to the front end of the main shaft It is provided between the tip side rolling bearing unit that supports the radial load and the base end side outer peripheral surface of the main shaft and the inner peripheral surface of the fixed housing, and supports the radial load applied to the base end portion of the main shaft. In a rotation support device for a main shaft of a machine tool provided with a base end side rolling bearing that allows axial displacement of the main shaft, the base end side rolling bearing is a cylindrical roller bearing, and the base end side rolling bearing is A cage that is rotatably provided between the inner circumferential surface of the outer ring and the outer circumferential surface of the inner ring, and has a plurality of pockets extending in the circumferential direction, and the cage is held in the pocket of the cage. Outer ring raceway provided on the inner peripheral surface of the outer ring and above A plurality of cylindrical rollers provided between the inner ring raceway and the inner ring raceway provided on the outer peripheral surface of the inner ring, the number of the pockets being an even number that is twice the number of the cylindrical rollers, The cylindrical rollers are movably held only in some of the pockets, and the remaining pockets are left empty without providing any members in the remaining pockets. A rotating support device for a spindle of a machine tool, characterized in that the outer peripheral surface is in communication with the outer peripheral surface.   The rotation support device for a spindle of a machine tool according to claim 1, wherein each cylindrical roller constituting a cylindrical roller bearing which is a base end side rolling bearing is made of ceramics.   When the diameter of the rolling surface of each cylindrical roller constituting the cylindrical roller bearing which is the base side rolling bearing is D and the axial length of each cylindrical roller is L, each cylindrical roller is 0.5 The rotation support device for a spindle of a machine tool according to claim 1 or 2, wherein the roller is a short roller that satisfies ≤L / D <1.

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