JP2004150563A - Supporting structure of spindle and machine tool - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a supporting structure of a spindle which can be used at a high rotational speed, and a machine tool. <P>SOLUTION: The spindle 40 of a machining center is so constituted that the front end part 40a side, to which a tool is mounted, is rigidly supported by bearings 42, 43, the terminal end part 40b side is supported by a bearing 44 having a damper part 61, and as a result, especially, vibrations on the front end part 40a side, to which the tool is mounted, are effectively suppressed. The damper part 61 is so constituted that oil is sealed between a pair of O-rings provided between a holder 41 and a bearing part. <P>COPYRIGHT: (C)2004,JPO

Description

[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a machine tool, and more particularly, to a support structure for a spindle on which a tool is mounted.
[0002]
[Prior art]
2. Description of the Related Art In various machine tools such as a machining center, a tool for processing a workpiece to be processed is mounted on a spindle (spindle), and the spindle is rotated to rotate the tool and hold the spindle side and the workpiece. By moving the side, predetermined processing is performed.
As shown in FIG. 12, in a machine tool, a spindle 1 on which a tool is mounted is usually supported by three bearings 2A, 2B, and 2C. The tool (not shown) is mounted on the tip 1a of the spindle 1 via a chuck mechanism (not shown). A rotor 3 is integrally provided between the bearings 2B and 2C of the main shaft 1, and the main shaft 1 is rotated around its axis by magnetism generated between the rotor 3 and a stator 4 provided on the outer periphery thereof. It has become.
Here, as the bearings 2A and 2B, those having a large restraining force in both the radial direction and the thrust direction are used, while the bearing 2C on the end portion 1b side of the main shaft 1 has a restraining force in the thrust direction with respect to the radial direction. Use small bearings. Thus, a structure is provided in which the displacement in the thrust direction (axial direction) is released on the end portion 1b side of the main shaft 1.
[0003]
Conventionally, many machining centers or the like perform machining by rotating the main spindle 1 at about 20,000 rpm.
In recent years, in order to improve productivity, it is desired that the rotation speed of the main shaft 1 be set to a high speed rotation of 30,000 rpm or more.
[0004]
What is very important in rotating the spindle 1 at high speed is the balance adjustment of the spindle 1. Conventionally, although not limited to the main spindle 1 of a machine tool, an object rotating at a high speed is adjusted in balance by some method. In the case of the spindle 1, usually, the unbalance is measured when the spindle 1 is rotated at a predetermined rotation speed (for example, 500 to 600 rpm), and a hole is drilled or ground in a part of the spindle 1 according to the result. This adjusts the rotational balance.
Since the rotational balance as described above is measured at a rotational speed very low with respect to the rotational speed of the spindle 1, the so-called static balance of the spindle 1 is adjusted.
On the other hand, it is conceivable to actually incorporate the main spindle 1 into a machine tool and balance the rotation of the main spindle 1 in this state. As such a method, there is a method in which a screw hole is provided at an end of a main spindle 1 incorporated in a machine tool, and a rotation balance is adjusted in a state where a weight whose weight is adjusted is inserted into the screw hole. (For example, refer to Patent Document 1).
[0005]
[Patent Document 1]
JP-A-6-126588
[0006]
[Problems to be solved by the invention]
However, if the spindle 1 is to be rotated at a higher speed than before, even if the above-described balance adjustment is performed, there is a limit to the high-speed rotation.
That is, the main shaft 1 has a natural frequency determined by the spring constant and the mass in a system in which the main shaft 1 itself has a spring constant and the rotor 3 integrally attached to the main shaft 1 has a mass. Then, as shown in FIG. 13, resonance occurs at a specific rotation speed according to the natural frequency (the rotation speed at this time is referred to as “critical speed”), and the vibration of the main shaft 1 increases. If the vibration of the main spindle 1 increases, the deflection of the tip of the tool mounted on the main spindle 1 increases, which directly leads to a reduction in machining accuracy.
For this reason, conventionally, the main shaft 1 must be rotated with the number of rotations lower than the critical speed as a use area, and there is a limit to the high-speed rotation of the main shaft 1.
[0007]
In addition, when an attempt is made to increase the rotation speed, the following problems are likely to become apparent.
First, with the configuration shown in FIG. 12A, as shown in FIG. 12B, the main shaft 1 has a small vibration in the middle portion (portion (A) in FIG. 12B). The mode is large in vibration at the tip 1a and the end 1b. For this reason, there is a disadvantage that sensitivity to unbalance is high (when an element causing the unbalance acts, large vibration is likely to occur).
In addition, a tool is mounted on the tip 1a of the main spindle 1. However, not only the main spindle 1 but also the tool has an element that causes an imbalance. It can be lost. For this reason, when the unbalance occurs at the tip 1a of the main shaft 1 having high sensitivity to the unbalance, the possibility that the vibration increases is high.
Furthermore, when a workpiece to be machined is cut by a tool during machining, a cutting force acts on the main shaft 1 from the tool. Then, in the main shaft 1 having high sensitivity to unbalance, the vibration of the main shaft 1 is easily amplified by the cutting force, and as a result, machining accuracy is reduced.
[0008]
In addition, it is necessary to maintain not only the static balance of the main spindle 1 alone but also the dynamic balance at the time of rotation while the main spindle 1 is supported by the bearings 2A, 2B, and 2C. Therefore, as a matter of course, it is necessary that the alignment is performed with high accuracy while the main shaft 1 is held by the bearings 2A, 2B, and 2C. If the alignment is not accurate, an external force acts on the main shaft 1 from the bearings 2A, 2B, 2C during high-speed rotation, which also affects the vibration of the main shaft 1.
For this reason, it is required to increase the precision of each component of the bearings 2A, 2B, and 2C, and to attach the bearings 2A, 2B, and 2C with very high precision. However, the bearings 2A, 2B, and 2C support three points. In the main shaft 1, it is difficult to accurately align the bearings 2A, 2B, 2C. This is also one of the factors that limit the number of rotations of the spindle 1.
[0009]
By the way, as shown in FIG. 14, the resonance points corresponding to the natural frequency of the main shaft 1 include primary, secondary, tertiary,... Therefore, as shown in FIG. 15, the vibration mode that changes according to the actual rotational speed is affected by all of these multiple orders, and the mode has a very complicated form.
Even when the balance of the main shaft 1 is adjusted, it is necessary to perform adjustment corresponding to these plural-order vibrations. In particular, in the case of the main shaft 1 which is long in the axial direction, it is necessary to adjust the balance on a plurality of surfaces (surfaces perpendicular to the axis). However, aside from the static balance taken by the main spindle 1 alone, in the case of the dynamic balance taking the main spindle 1 set in the bearings 2A, 2B, 2C, it is practically difficult to correct the balance in the middle part of the main spindle 1. However, as described in Patent Document 1, the balance must be corrected only at the tip 1a and the end 1b of the spindle 1, and the balance cannot be adjusted as desired.
As a result, the influence of the higher-order mode remains, and it is difficult to obtain a predetermined accuracy.
[0010]
The present invention has been made based on such a technical problem, and an object of the present invention is to provide a spindle support structure and a machine tool that can be used at a higher rotation speed.
[0011]
[Means for Solving the Problems]
With this object in mind, the spindle support structure of the present invention rotatably supports a spindle to which a tool is attached at one end via three or more bearings with a holder, and rotates the spindle around its axis by a motor. Of the three or more bearings, the bearing located at the other end of the main shaft includes a vibration damping unit that dampens radial vibration of the main shaft.
It is preferable that the vibration damping unit has a vibration damping ability with respect to a vibration in a specific frequency region according to the natural frequency of the main shaft.
As the vibration damping unit, for example, there is a configuration in which a pair of seal members is interposed between a bearing and a holder facing each other at an interval, and damper oil is sealed between the pair of seal members.
In addition, a configuration may be provided in which a displacement absorbing portion that extends obliquely between the bearing and the holder facing each other at an interval and absorbs displacement in a direction in which the bearing and the holder approach and separate from each other is provided. Such a displacement absorbing portion may be provided with two portions that extend obliquely, and may have a substantially rectangular cross section as a whole.
Further, the vibration damping portion has a contact surface that is supported by the holder and abuts on the outer peripheral surface of the bearing, and is deformed in a direction substantially perpendicular to the contact surface, thereby displacing the bearing and the holder in a direction in which the bearing approaches and separates from each other. May be provided with a diaphragm portion for absorbing the pressure.
Here, a slit or the like can be formed in the displacement absorbing portion or the diaphragm portion to adjust the rigidity.
[0012]
The present invention also provides a work holding unit for holding a work to be machined, a main shaft for holding a tool for working the work at one end and rotating the tool around an axis, and controlling relative movement of the work holding unit and the main shaft. And a controller that performs the same. In this machine tool, the main shaft is rotatably supported by three or more bearings, and among the three or more bearings, the bearing located on the other end side of the main shaft includes a vibration damping portion that attenuates radial vibration of the main shaft. It is characterized by having.
Further, the machine tool may be characterized in that the spindle is used at a rotation speed exceeding a critical speed determined based on the natural frequency of the spindle. More specifically, the rotation speed is 30,000 rpm or more.
[0013]
BEST MODE FOR CARRYING OUT THE INVENTION
Hereinafter, the present invention will be described in detail based on embodiments shown in the accompanying drawings.
FIG. 1 is a diagram for explaining a configuration of a machining center according to the present embodiment.
As shown in FIG. 1, the machining center 10 includes a workbench 20, a processing head 30, and a controller (not shown) for controlling operations of the workbench 20 and the processing head 30.
The workbench 20 includes a work table (work holding unit) 22 for holding a work to be processed on a base block 21. The work table 22 is configured to move along the linear guide 23 by driving a feed motor (not shown). Further, the work table 22 is provided with a rotation mechanism 24 that rotates about a vertical axis, whereby the machining center 10 can be a four-axis processing machine.
[0014]
The processing head 30 includes a base block 31, a column block 32, and a spindle block 33.
On the base block 31, a pair of two linear guides 34 extending in a horizontal direction (hereinafter referred to as a Y direction) orthogonal to the moving direction of the work table 22 is provided. The column block 32 extends vertically upward with respect to the work table 22, and is movably provided along a linear guide 34 by a feed motor (not shown). The column block 32 is provided with a pair of two linear guides 35 extending in a vertical direction (this is defined as a Z direction), and the spindle block 33 moves along the linear guides 35 by driving a feed motor (not shown). It is provided as possible.
The spindle block 33 includes a main shaft 40 on which a tool for performing processing such as cutting on a work held on the work table 22 is mounted. The main shaft 40 is driven to rotate about an axis.
[0015]
The movement of the work bench 22 of the workbench 20 along the X direction, the movement of the column block 32 of the processing head 30 along the Y direction, and the movement of the spindle block 33 along the Z direction as described above are input in advance. The operation is performed by controlling the operation of a feed motor (not shown) by a controller (not shown) based on a numerical control program. As a result, the work held on the work table 22 and the tool mounted on the spindle 40 relatively move in three directions of X, Y, and Z. Then, while rotating the tool mounted on the spindle 40 about the axis of the spindle 40, the workpiece and the tool are relatively moved in three directions based on a machining program input in advance, so that the workpiece and the tool are held on the work table 22. A predetermined process is performed on the work.
[0016]
The spindle 40 provided in the spindle block 33 has a support structure as shown in FIG. 2 in detail.
As shown in FIG. 2, the main shaft 40 is supported by a holder 41 fixed to a spindle block 33 via bearings 42, 43, 44 so as to be rotatable around its axis. In a portion between the bearing 43 located at the intermediate portion of the main shaft 40 and the bearing 44 located at the end portion 40b of the main shaft 40, a rotor 45 constituting a motor is integrally provided on the outer peripheral surface of the main shaft 40. I have. A stator 46 that constitutes a motor is provided at a position on the inner peripheral surface of the holder 41 that faces the rotor 45. The main shaft 40 is rotationally driven at a predetermined rotation speed around the axis by the motor constituted by the rotor 45 and the stator 46.
[0017]
Now, of the three bearings 42, 43, and 44 that rotatably hold the main shaft 40, the bearing 42 located at the tip end portion 40a of the main shaft 40 and the bearing 43 located at the intermediate portion have the same radial shape as in the related art. Those having a large restraining force in both the direction and the thrust direction are used. For example, as shown in FIG. 3, an inner race 50a fitted on the outer peripheral surface of the main shaft 40, an outer race 50b fitted on the inner peripheral surface of the holder 41, and a roller 50c rotating between the inner race 50a and the outer race 50b. This is a so-called general bearing 50.
[0018]
On the other hand, as shown in FIG. 4, the bearing 44 located at the end portion 40 b of the main shaft 40 has a bearing portion 60 and a damper portion (vibration damping portion) 61.
The bearing part 60 includes an inner race 60a fitted on the outer peripheral surface of the main shaft 40, an outer race 60b located on the outer peripheral side thereof, and a roller 60c rotating between the inner race 60a and the outer race 60b.
[0019]
On the other hand, the damper part 61 has the following configuration.
A short cylindrical ring plate 62 having a predetermined width in the axial direction of the main shaft 40 is fitted on the outer peripheral side of the outer race 60b.
The outer diameter of the ring plate 62 is set to be smaller than the inner diameter of the portion of the holder 41 facing the ring plate 62 by a predetermined dimension, so that the clearance between the ring plate 62 and the holder 41 has a predetermined dimension. C is formed.
On the outer peripheral surface of the ring plate 62, a pair of two grooves 62a, 62a continuous in the circumferential direction are formed. On the other hand, on the inner peripheral surface of the holder 41, two pairs of grooves 41a, 41a are formed at positions corresponding to the grooves 62a, 62a. O-rings (seal members) 63 made of, for example, a rubber material are provided in the grooves 62 a of the ring plate 62 and the grooves 41 a of the holder 41. Thus, the O-rings 63, 63 are arranged at predetermined intervals in the axial direction of the main shaft 40. A space surrounded by the outer peripheral surface of the ring plate 62, the inner peripheral surface of the holder 41, and the two O-rings 63, 63 is formed, and an oil (damper oil) 64 is sealed in this space. ing. The enclosed oil 64 functions as the damper portion 61.
[0020]
As shown in FIGS. 5A and 5B, in the bearing 44 having the damper portion 61, the O-rings 63 made of, for example, a rubber material function as springs having an elastic coefficient k. The filled oil 64 functions as a dashpot (damping member) having a damping coefficient c. Since the O-rings 63, 63 and the oil 64 are interposed over the entire circumference between the ring plate 62 and the holder 41, they have a suspension effect in the entire circumferential direction of the main shaft 40. become. Then, in each direction, a system as shown in FIG. 5B is provided.
[0021]
As shown in FIG. 6A, as described above, of the bearings 42, 43, and 44 that support the main shaft 40, only the bearing 44 on the terminal end 40b side has the damper portion 61. I have. As a result, the main shaft 40, which is supported at three points, is "rigidly" supported at two points (portions of the bearings 42 and 43) on the tip portion 40a side, whereas it is "softly" supported on the end portion 40b side. It is the configuration that was done.
In such a configuration, the damper portion 61 of the bearing 44 exerts its suspension effect when the vibration of the main shaft 40 increases due to the O-rings 63 and 63 having the elastic coefficient k and the oil 64 having the damping coefficient c, In particular, vibration at the resonance point generated when the main shaft 40 rotates is suppressed.
[0022]
Here, generally, as shown in FIG. 6B, the Q factor (Qf) corresponding to the sharpness of the resonance is as follows, where a represents the amplitude at the resonance point and k represents the elastic coefficient of the main shaft 40.
a = k · Qf
It is. Also,
Qf = 1 / 2ζ
(Ζ is the damping ratio),
ζ = c / [2 · (mk) ^1/2 ]
(M is the equivalent mass of the rotor 45).
Therefore, as the damping coefficient c increases, the Q factor · Qf decreases, and thereby the amplitude a at the resonance point decreases.
That is, by appropriately setting the damping coefficient c of the oil 64 in the damper section 61, the vibration at the resonance point can be effectively suppressed.
[0023]
Now, with the support structure of the main shaft 40 as described above, the resonance can be effectively suppressed by the damper portion 61. Therefore, the main shaft 40 is rotated in a region exceeding the critical speed determined according to the natural frequency of the main shaft 40. It becomes possible.
In a region where the vibration is not large, the main shaft 40 is in a state where it is supported at two points of the bearings 42 and 43 on the tip portion 40a side. The support effect of the bearing 44 is increased by the suspension effect exerted by the damper portion 61, so that the bearing is supported at three points of the bearings 42, 43 and 44. In other words, statically supports two points, and dynamically supports three points. That is, it can be said that the main shaft 40 has particularly high dynamic rigidity.
In such a support structure, as shown in FIG. 6C, the vibration mode of the main shaft 40 has a node near the distal end portion 40a, small vibration at the distal end portion 40a and the intermediate portion, and increases at the end portion 40b side. The form is shown. Therefore, the vibration at the tip end portion 40a side where the tool is mounted is smaller than that in the related art (see FIG. 12B), and it is possible to suppress a reduction in machining accuracy due to a run-out at the tool tip end. Also, by having the vibration mode as shown in FIG. 6C, the sensitivity to unbalance is reduced, and various tools are attached to and detached from the tip portion 40a of the main shaft 40, and a cutting force acting during machining is input. Also, it is difficult for vibration to increase.
[0024]
Also, as described above, the main shaft 40 is substantially firmly supported at two points of the bearings 42 and 43 on the side of the distal end portion 40a except for the critical speed region where resonance occurs, and the bearing 44 on the side of the end portion 40b is rigidly supported. , And is more flexibly supported. Therefore, compared with the conventional three-point support configuration, there is no need to severely align the main shaft 40 and the bearings 42, 43, 44, and the alignment can be performed easily. Moreover, since the end portion 40b side is flexibly supported by the bearing 44, the influence of the thermal expansion of the main shaft 40 and the like is much smaller than in the conventional case.
[0025]
Further, by supporting the end portion 40b side of the main shaft 40 with the bearing 44 having the damper portion 61, as shown in FIG. , A mode having fewer inflection points than that shown in FIG. Therefore, even when the dynamic balance is obtained particularly in a state where the main shaft 40 is actually incorporated in the spindle block 33, the balance can be easily corrected even with a small correction surface such as the front end portion 40a and the end portion 40b of the main shaft 40.
[0026]
As described above, the main shaft 40 of the machining center 10 is rigidly supported by the bearings 42 and 43 on the tip end 40a side on which the tool is mounted, and supported by the bearing 44 having the damper portion 61 on the end end 40b side. In particular, it is possible to effectively suppress vibration particularly on the tip end portion 40a side where the tool is mounted, and it is possible to achieve high-precision cutting performance. Moreover, the damping effect exerted by the oil 64 sealed in the damper portion 61 can suppress the resonance amplitude of the main shaft 40, so that the ultra-high speed rotation range of 30000 to 50,000 rpm, which exceeds the dangerous speed conventionally difficult, is achieved. It can be used in Therefore, the machining center 10 can have a higher productivity than ever before.
[0027]
By the way, in the above-mentioned embodiment, the oil 64 sealed between the bearing part 60 and the holder 41 is used as the damper part 61. However, if the damping effect can be effectively exerted, other oils may be used. The structure can be appropriately adopted.
The following shows another example of such a structure.
[Other examples]
The damper portion (vibration damping portion) 70 shown in FIG. 8 is provided between the outer race 60 b of the bearing portion 60 and the holder 41 of the bearing 44 supporting the outer peripheral portion of the main shaft 40. The damper portions 70 are provided on both sides of the bearing portion 60 in the width direction, and are formed of an inner peripheral ring 71 and an outer peripheral ring 72, respectively. The inner peripheral ring 71 and the outer peripheral ring 72 are integrally joined at mating surfaces 71a, 72a, and have a substantially "U" shape when viewed in cross section.
The inner peripheral ring 71 has a base end portion 71b fixed to a side surface (or an outer peripheral surface) of the outer race 60b, and a slot-shaped or tapered intermediate portion (which extends obliquely toward the inner peripheral surface of the holder 41 when viewed in cross section). A displacement absorbing portion) 71c, and a mating surface 71a is formed at the tip of the intermediate portion 71c. The outer peripheral ring 72 has a base end 72b fixed to the inner peripheral surface of the holder 41, and has a file-shaped or tapered intermediate portion (displacement absorbing portion) 72c that extends diagonally toward the bearing portion 60 when viewed in cross section. A mating surface 72a is formed at the tip of the intermediate portion 72c.
The inner peripheral ring 71 and the outer peripheral ring 72 of the damper portion 70 are configured such that when the main shaft 40 and the holder 41 are relatively moved in the radial direction, the intermediate portions 71c and 72c in the shape of a chip pot are elastically deformed. A repulsive force is generated, thereby damping the vibration of the main shaft 40. For this reason, for the inner peripheral ring 71 and the outer peripheral ring 72, the desired vibration damping effect can be obtained according to the material (elastic coefficient), the length and thickness of the intermediate portions 71c and 72c, and the natural frequency of the main shaft 40 and the like. Is appropriately set and selected in advance.
[0028]
[Another example]
The damper portion (vibration damping portion) 80 shown in FIG. 9A is provided between the outer race 60 b of the bearing portion 60 and the holder 41 of the bearing 44 supporting the outer peripheral portion of the main shaft 40. The damper portion 80 extends on the outer peripheral surface of the outer race 60b to an intermediate portion (diaphragm portion) 80a having a contact surface that is in contact with the outer race 60b via a collar 81, and extends on both sides toward the main shaft 40 (inner peripheral side). And a protruding flange portion 80b. In this damper portion 80, flange portions 80b on both sides are fixed to a pair of support portions 83 projecting from the holder 41 to the inner peripheral side, and O is provided between the support portion 83 and the flange portion 80b. A ring 84 is provided.
The intermediate portion 80a has a width sufficiently larger than the widths of the outer race 60b and the collar 81, and when the main shaft 40 and the holder 41 attempt to move relative to each other in the radial direction, the intermediate portion 80a is moved like a diaphragm. By resiliently deforming in a direction substantially perpendicular to the surface, a repulsive force is generated, thereby damping the vibration of the main shaft 40. For this reason, the material (elastic coefficient) of the damper portion 80 and the width and thickness of the intermediate portion 80a are appropriately set in advance so that the desired vibration damping effect can be obtained according to the natural frequency of the main shaft 40 and the like. Selected.
Furthermore, the space formed between the intermediate portion 80a and the holder 41 can be filled with oil 85 that functions as a damper. The oil 85 is injected from an inlet (not shown). At the time of the injection, air existing in a space formed between the intermediate portion 80a and the holder 41 is discharged from a discharge port (not shown). After the injection, the injection port and the discharge port are sealed, so that oil 85 Is enclosed.
The oil 85 may be circulated for lubrication of the main shaft 40 and the bearing 60 as well as the space formed between the intermediate portion 80a and the holder 41.
As shown in FIG. 9B, the intermediate portion 80a has a groove extending in a direction connecting the flange portions 80b on both sides and not penetrating the intermediate portion 80a in order to facilitate the elastic deformation. 82 can also be formed.
[0029]
The damper part 80 shown in FIG. 10 has the same configuration as that of FIG. 9A, and is provided between the outer race 60 b of the bearing part 60 and the holder 41 of the bearing 44 supporting the outer peripheral part of the main shaft 40. ing. The damper portion 80 is formed on the outer peripheral surface of the outer race 60b by an intermediate portion 80a having a contact surface in contact with the collar 81 via a collar 81, and flange portions 80b extending toward the main shaft 40 on both sides thereof. ing.
The damper portion 80 is supported by a pair of O-rings 86 mounted in a groove formed in the holder 41. That is, the main shaft 40 is positioned by these O-rings 86.
Further, the space surrounded by the intermediate portion 80a, the holder 41, and the O-ring 86 can be filled with oil that functions as a damper. This oil is injected from an injection port (not shown). At the time of this injection, air existing in a space formed between the intermediate portion 80a and the holder 41 is discharged from a discharge port (not shown), and after the injection, the injection port and the discharge port are sealed, so that oil is injected into this space. Encapsulate.
Further, the oil may be circulated not only for the space formed between the intermediate portion 80a and the holder 41 but also for lubrication of the main shaft 40 and the bearing portion 60.
The intermediate portion 80a has a width that is sufficiently larger than the width of the outer race 60b and the collar 81. When the main shaft 40 and the holder 41 attempt to move relative to each other in the radial direction, the intermediate portion 80a becomes like a diaphragm. As a result, a repulsive force is generated by being elastically deformed in a direction substantially perpendicular to the surface, whereby the vibration of the main shaft 40 is attenuated. At this time, the vibration damping effect can be obtained by the oil filled in the space surrounded by the intermediate portion 80a, the holder 41, and the O-ring 86. Also, the vibration damping effect of the O-ring 86 can be obtained by appropriately setting its material.
[0030]
The damper part 80 shown in FIG. 11A is provided between the outer race 60 b of the bearing part 60 and the holder 41 of the bearing 44 supporting the outer peripheral part of the main shaft 40. The damper portion 80 is formed on the outer peripheral surface of the outer race 60b by an intermediate portion 80a having a contact surface in contact with the collar 81 via a collar 81, and flange portions 80b extending toward the main shaft 40 on both sides thereof. ing.
As shown in FIG. 11B, a slit 87 extending in a direction connecting the flange portions 80b on both sides and penetrating the intermediate portion 80a is formed in the intermediate portion 80a.
Further, the space between the intermediate portion 80a and the holder 41 can be filled with oil that functions as a damper.
An O-ring 88 located outside the slit 87 is provided between the outer peripheral surface of the collar 81 and the inner peripheral surface of the intermediate portion 80a to prevent oil from leaking.
When the main shaft 40 and the holder 41 attempt to move relative to each other in the radial direction, the intermediate portion 80a generates a repulsive force by elastically deforming the intermediate portion 80a in a direction substantially orthogonal to its surface like a diaphragm. Thus, the vibration of the main shaft 40 is attenuated. At this time, the vibration damping effect can be obtained also by the oil filled in the space surrounded by the intermediate portion 80a and the holder 41.
[0031]
In the above-described embodiment, a three-axis machining center is used as an example. However, a similar structure can be applied to a four-axis, five-axis machining center or a spindle of another machine tool. .
Further, as long as the support structure of the main shaft 40 is the same, the configuration of the machining center and other parts of various machine tools may be any.
In addition, the configuration described in the above embodiment can be selected or changed to another configuration as appropriate without departing from the gist of the present invention.
[0032]
【The invention's effect】
As described above, according to the present invention, it is possible to effectively suppress the vibration of the main spindle particularly at the tip end side where a tool is mounted, and to realize a machine tool having high-precision machining performance. it can. In addition, the resonance of the main shaft can be suppressed by the damping effect of the damper part, so it can be used in the ultra-high speed rotation region exceeding the dangerous speed, which was difficult previously, and the machine tool can be more productive. Can be included.
[Brief description of the drawings]
FIG. 1 is a perspective view showing a schematic configuration of a machine tool according to the present embodiment.
FIG. 2 is a sectional view showing a support structure of a main shaft.
FIG. 3 is a view showing a bearing at a tip end side of a main shaft.
FIG. 4 is a view showing a bearing with a damper provided on the end side of the main shaft.
FIG. 5 is a model diagram of a bearing having a damper portion.
FIGS. 6A and 6B are diagrams showing a state of generation of vibration of a main shaft supported by a bearing having a damper portion, wherein FIG. 6A is a diagram schematically showing a main shaft support structure, and FIG. FIG. 3C is a diagram showing the relationship between the number and the amplitude, and FIG.
FIG. 7 is a diagram illustrating first-, second-, and third-order vibration modes.
FIG. 8 is a view showing another example of a main shaft support structure.
FIG. 9 is a view showing still another example of the support structure of the main shaft.
FIG. 10 is a view showing still another example of the support structure of the main shaft.
FIG. 11 is a view showing still another example of the support structure for the main shaft.
12A is a diagram illustrating a conventional spindle support structure, and FIG. 12B is a diagram illustrating a primary vibration mode of the spindle.
FIG. 13 is a diagram showing the relationship between the rotation speed and amplitude of a spindle in a conventional spindle support structure.
FIG. 14 is a diagram illustrating a relationship between a rotation speed including a plurality of vibrations and an amplitude in a conventional spindle support structure.
FIG. 15 is a diagram showing primary, secondary, and tertiary vibration modes in a conventional spindle support structure.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 10 ... Machining center, 22 ... Work stand (work holding part), 30 ... Processing head, 33 ... Spindle block, 40 ... Spindle, 40a ... Tip part, 40b ... End part, 41 ... Holder, 42, 43, 44 ... Bearing, 45 ... rotor, 50 ... bearing, 60 ... bearing part, 61, 70, 80 ... damper part (vibration damping part), 63 ... O-ring (seal member), 64 ... oil (damper oil), 71 ... inner ring , 72 ... outer peripheral ring, 71c, 72c ... middle part (displacement absorbing part), 80a ... middle part (diaphragm part)

Claims (8)

A spindle on which a tool is mounted at one end;
Three or more bearings for supporting the main shaft,
A holder that rotatably supports the main shaft via the bearing,
A motor for rotating the main shaft about its axis,
With
The main shaft support structure, wherein, among the three or more bearings, a bearing located on the other end side of the main shaft includes a vibration damping portion that dampens radial vibration of the main shaft. 2. The main shaft support structure according to claim 1, wherein the vibration damping portion has a vibration damping ability with respect to a vibration in a specific frequency region according to a natural frequency of the main shaft. 3. The vibration damping portion, a pair of seal members interposed between the bearing and the holder facing each other at an interval,
A damper oil sealed between the pair of seal members,
The main shaft support structure according to claim 1 or 2, further comprising: The vibration damping unit may include a displacement absorbing unit that extends obliquely between the bearing and the holder facing each other at an interval and absorbs displacement in a direction in which the bearing and the holder approach and separate from each other. The main shaft support structure according to claim 1 or 2, wherein The vibration damping portion is supported by the holder, has a contact surface that comes into contact with the outer peripheral surface of the bearing, and deforms in a direction substantially perpendicular to the contact surface so that the bearing and the holder approach and separate from each other. The main shaft support structure according to claim 1 or 2, further comprising a diaphragm portion that absorbs displacement in the direction. A work holding unit for holding a work to be machined,
A spindle for holding the tool for processing the work at one end, and rotating the tool about its axis;
A controller for controlling the relative movement of the work holding unit and the spindle,
The main shaft is rotatably supported by three or more bearings, and among the three or more bearings, a bearing located on the other end side of the main shaft includes a vibration damping unit that attenuates radial vibration of the main shaft. A machine tool characterized in that: The machine tool according to claim 6, wherein the spindle is used at a rotation speed exceeding a critical speed determined based on a natural frequency of the spindle. The machine tool according to claim 7, wherein the rotation speed is equal to or greater than 30,000 rpm.

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