JP2003172343A - Multi-point contact ball bearing for supporting ball screw - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a multi-point contact ball bearing suitable for a bearing supporting a ball screw shaft of a ball screw. <P>SOLUTION: If setting the diameter of a ball 10c as D, the curvature of an outer ring side track groove 10e as Re, the curvature of an inner ring side track groove 10f as Ri, and a contact angle between the ball 10c and track grooves 10a, 10b as α, the contact angle α is made to satisfy an expression of 12.5°<α≤32.5°, and the curvature rates of the track grooves 10e, 10f with regard to the diameter of the ball 10c is made to satisfy an expression of Re/D =Ri/D=0.54 to 0.61. Annular seal elements 12 are provided on both end portions of the outer ring 10a or the inner ring 10b, and an annular space formed between the outer ring 10a and the inner ring 10b is sealed by the annular seal element 12. <P>COPYRIGHT: (C)2003,JPO

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an electronic component mounting apparatus for mounting or mounting electronic components such as semiconductor chips and liquid crystals on a substrate in an IT-related product manufacturing process or assembly process, an electric discharge machine, and a laser. The present invention relates to a ball screw supporting multi-point contact ball bearing that supports a ball screw shaft of a precision positioning ball screw used in a non-cutting machine tool such as a processing machine or a light cutting machine such as a tapping center.

[0002]

2. Description of the Related Art Conventionally, a ball screw shaft of a precision positioning ball screw used in an electronic component mounting apparatus, a machine tool or the like is often supported by a bearing device as shown in FIG. This bearing device is a double row combination bearing 1
A bearing housing 2 for sealing the periphery of the two-row combination bearing 1 and oil seals 3 provided at both ends of the bearing housing 2 in the axial direction.
21 is composed of two angular ball bearings 4, as shown in FIG. These angular ball bearings 4 are composed of an outer ring 4a, an inner ring 4b, balls 4c, etc., and the inner ring 4b is
As shown in FIG. 20, the pressing ring 6 and the bearing nut 7 are pressed and fixed to the end surface of the large-diameter shaft portion 5a formed at the end portion of the ball screw shaft 5. The bearing housing 2 is composed of a cylindrical housing body 2a and an annular outer ring retainer 2c which is bolted to the flange portion 2b of the housing body 2a.

By the way, the angular ball bearing has the following features as compared with the deep groove ball bearing. The number of balls per row is large, and the load carrying capacity and rigidity are large. The shape of the cage is strong and high-speed rotation is possible. A single row can only load an axial load in one direction, but a combination of two rows of angular ball bearings makes it possible to load an axial load in both directions.

By adjusting the gap g (see FIG. 21) formed between the end faces of the two angular ball bearings, it is easy to eliminate the internal clearance of the bearings and generate an internal load in advance (so-called fixed position preload). Is. By applying a preload, it is possible to improve bearing rigidity and obtain high rotational accuracy. Since the angular contact ball bearings have the above characteristics, they are used as bearings that support spindles in machine tools such as lathes, milling machines, and grinders, as well as machines that require highly accurate positioning (for example, machine tools). It is also widely used for precision ball screws for machines, precision ball screws for high-speed electronic component mounting machines, ball screws for transfer robots, etc.). However, when an angular ball bearing is used as a bearing for supporting a ball screw, a load load in both axial directions acts on the ball screw shaft, so a single row angular ball bearing alone can support the load acting on the ball screw shaft. Can not do it. Therefore, in the ball screw used as a means for moving the bed of the cutting type or grinding type machine tool, the displacement due to the reaction force of the processing load is minimized,
In addition, in order to obtain the required processing accuracy, a combination of thrust angular contact ball bearings with extremely high axial rigidity and a contact angle of about 60 ° in 2 to 4 rows is used as a ball screw support bearing. In many cases. In this case, the load applied in the radial direction is received by the linear guide and dovetail groove used in combination with the ball screw.

On the other hand, electrolytic machining type machine tools such as electric discharge machines and laser machines, tapping centers for machining screw holes,
Alternatively, in an electronic component mounting apparatus or the like, since the load applied during actual processing is relatively small, a bearing for supporting the ball screw shaft has, for example, an inner diameter of about 6 mm to 40 mm and a contact angle of about 15 ° to 30 °. In many cases, standard type angular contact ball bearings (dimension series 70xx, 72xx, etc.) are combined in two rows and a comparatively small preload of several tens N to several hundreds N is applied.

[0006]

As described above, when the angular ball bearing is used as a ball screw supporting bearing, a fixed position preload must be applied, and therefore, a two-row combination is a condition. For this reason, the space for the shaft on which the bearing is mounted and the space for the housing are doubled as compared with the use in a single row, which leads to an increase in the space of the entire machine.

In the assembly line and mounting line for IT parts and the like, a large number of component mounting apparatuses and transfer robots are used according to the process, so the above-mentioned increase in space poses a problem in line construction. Further, when the double-row combination angular contact ball bearing is used in several places in one machine, there is a concern that interference between the ball screw supporting bearings may become a problem. In addition to this, recently, it has been required to increase the mounting speed of components to shorten the tact time, and there is also a very high demand for reducing the inertial force at the time of rapid acceleration / deceleration by reducing the weight of the movable part.

Further, when a conventional ball screw supporting bearing (double-row combination angular contact ball bearing) is used in an electrolytic machining type machine tool such as an electric discharge machine or a laser machining machine, fine particles generated during electrolytic machining are generated. There is a risk that sludge, electrolytic solution, etc. may enter the inside of the bearing through the annular gap formed between the outer ring and the inner ring. Further, in an electronic component mounting apparatus that mounts an electronic component such as a semiconductor chip on a substrate, a semiconductor chip left in the line may inadvertently enter the inside of the bearing.

The present invention has been made in view of the above problems, and an object thereof is to replace the double row angular contact ball bearing currently used as a ball screw supporting bearing. In addition, it is an object of the present invention to provide a multi-point contact ball bearing for supporting a ball screw, which is capable of accurate positioning and preloading, and which can improve dustproofness and waterproofness.

[0010]

In order to achieve the above-mentioned object, the invention of claim 1 is a multipoint having a large number of balls which contact the raceway grooves formed in the outer ring and the inner ring at three or more points. The contact ball bearing is characterized in that an annular seal body is provided at both ends of the outer ring or the inner ring, and an annular gap formed between the outer ring and the inner ring is sealed by the annular seal body.

According to a second aspect of the present invention, in the ball screw supporting multipoint contact ball bearing according to the first aspect, the diameter of the ball is D, the curvature of the raceway groove is R, and the contact angle between the ball and the raceway groove. Where α is α, the contact angle α is 12.5 ° <α ≦ 1
The curvature ratio R / D of the raceway groove to the diameter of the ball is set to 7.5 °, and R / D = 0.54 to 0.555.

According to a third aspect of the present invention, in the ball screw supporting multipoint contact ball bearing according to the first aspect, the diameter of the ball is D, the curvature of the raceway groove is R, and the contact angle between the ball and the raceway groove. Where α is α, the contact angle α is 17.5 ° <α ≦ 2
It is characterized in that the curvature ratio R / D of the raceway groove to the diameter of the ball is 2.555 <R / D ≦ 0.57.

According to a fourth aspect of the present invention, in the ball screw supporting multi-point contact ball bearing according to the first aspect, the diameter of the ball is D, the curvature of the raceway groove is R, and the contact angle between the ball and the raceway groove. Where α is α, the contact angle α is 22.5 ° <α ≦ 2
The curvature ratio R / D of the raceway groove with respect to the diameter of the ball is set to 7.5 ° and 0.57 <R / D ≦ 0.59.

According to a fifth aspect of the present invention, in the ball screw supporting multipoint contact ball bearing according to the first aspect, the diameter of the ball is D, the curvature of the raceway groove is R, and the contact angle between the ball and the raceway groove. Where α is α, the contact angle α is 27.5 ° <α ≦ 3
The curvature ratio R / D of the raceway groove with respect to the diameter of the balls is set to 2.5 ° and 0.59 <R / D ≦ 0.61.

[0015]

The multipoint contact ball bearing for supporting a ball screw according to the present invention, as shown in FIG. 16 or 17, has an outer ring 10a.
Also, the raceway grooves 10e and 10f formed in the inner ring 10b are multi-point contact ball bearings in which the balls 10c contact at three or four points, and either the outer ring 10a or the inner ring 10b is divided into two in the radial direction. . Further, the ball screw supporting multi-point contact ball bearing according to the present invention is provided with an annular seal body at both ends of the outer ring or the inner ring, and the annular seal body seals the annular gap formed between the outer ring and the inner ring. It has a different structure.

Here, by providing annular seal bodies at both ends of the outer ring or the inner ring, prevention of fine sludge from entering the inside of the bearing during electrolytic machining such as electric discharge machining or laser machining is carried out. Intrusion prevention In the case of electronic component mounting equipment, prevention of intrusion of fine semiconductor chips into the bearing. Prevention of axial detachment of a separate inner ring (when a ball screw support unit is used, especially handling is improved). The effect of improving the assemblability when assembling the bearing is obtained.

Generally, in general applications, the bearing is mainly used for inner ring rotation, and a stationary radial load is often applied (so-called inner ring rotation load). In this case, tighten the inner ring to prevent fretting that tends to occur on the inner ring.
Make the outer ring a gap fit. For this reason, when the bearing is assembled into the shaft or the housing, a method of first press-fitting or shrink-fitting the inner ring onto the shaft and then inserting the outer ring into the housing is often adopted. However, ball screw support bearings are used only for axial loads and not radial loads, so
A method is adopted in which the outer ring and the inner ring are set to a clearance fit (h or H fit eye) and the axial direction is securely fastened with a bearing nut or the like. In addition to such fitting, the shaft may be a long ball screw, and the bearing is prefabricated as a ball screw support unit inserted in the housing, and that unit is used as a clearance fit screw shaft. Often, a built-in method of inserting is adopted. In such a case, the multi-point contact ball bearing having a separate inner ring is extremely poor in assembling. Further, if the inner ring is accidentally separated / dropped, foreign matter may enter the bearing or dent marks may be generated. However, the multi-point contact ball bearing according to the present invention is used as a ball screw bearing. As a result, the above problems can be prevented.

As shown in FIG. 18 or FIG. 19, between the adjacent end faces of the inner ring divided into two in the radial direction,
An appropriate preload clearance (Δa) is geometrically formed by three-point or four-point contact (bearing contact angle α), and like the two-row combination angular contact ball bearing, the end of the inner ring or outer ring can be Appropriate preload can be applied by pressing and fixing the parts in the axial direction.

[0019]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 is a sectional view showing a structure of a ball screw supporting bearing according to a first embodiment of the present invention,
As shown in the figure, the ball screw supporting bearing according to the first embodiment includes a four-point contact ball bearing 10 that supports the ball screw shaft 5, and a bearing housing 2 that seals the periphery of the ball bearing 10. An oil seal 3 is provided on both ends of the bearing housing 2 in the axial direction.

FIG. 2 is an axial sectional view of the four-point contact ball bearing 10. As shown in the figure, a four-point contact ball bearing 10
Is composed of an outer ring 10a, an inner ring 10b, balls 10c, a cage 10d, etc., and a raceway groove 10e is formed on the inner peripheral surface of the outer ring 10a along the circumferential direction of the outer ring 10a. As shown in (a) of FIG. 17, the axial section of the raceway groove 10e is formed into two arcs having a radius of curvature of Re, that is, a Gothic arch shape, whereby the ball 10c is formed in the raceway groove 10e. It comes in contact with two points.

On the other hand, the raceway groove 10 is formed on the outer peripheral surface of the inner ring 10b.
f is formed along the circumferential direction of the inner ring 10b. As shown in FIG. 17B, the axial section of the raceway groove 10f is formed into two arcs having a radius of curvature of Ri, that is, a Gothic arch shape, whereby the ball 1 is formed.
0c contacts the raceway groove 10f at two points.

The inner ring 10b of the four-point contact ball bearing 10 is shown in FIG.
As shown in FIG. 5, the ball screw shaft 5 is pressed and fixed to the end surface of the large-diameter shaft portion 5a formed at the end portion thereof by the pressing ring 6 and the bearing nut 7. Further, the inner ring 10b has its axial center portion divided into two in the radial direction, and between the adjacent end faces of the divided inner ring, there is a geometrical four-point contact (bearing contact angle α) with an appropriate preload. Gap 11 (Fig. 1
8) is formed.

The four-point contact ball bearing 10 is provided with a pair of annular seal bodies 12 (see FIG. 2) that seal the annular gap formed between the outer ring 10a and the inner ring 10b. These annular seal bodies 12 are formed in an annular shape, and the outer peripheral edge portion of each annular seal body 12 is detachably engaged with a seal mounting groove 13 formed on the inner peripheral surface of the outer ring 10a. The portion 12a is formed. Further, the annular seal body 12 is formed of an elastic material such as resin, and each annular seal body 12
As shown in FIG. 3, a Rabins seal portion 12b which is elastically contacted with the seal groove 14 formed on the outer peripheral surface of the inner ring 10b is formed on the inner peripheral edge portion of the. With the above-described configuration, when the inner ring is divided into two parts, it is possible to improve the waterproof property and the dustproof property and prevent the inner ring from being disassembled.

FIG. 4 is a perspective view showing a part of the ball 10c and the holder 10d. As shown in FIG.
d has a crown shape and is formed of a synthetic resin material such as polyamide, polyacetal, polyphenylene sulfide, or the like. In addition, if it becomes necessary to increase the load capacity, a total ball design without a cage is possible. The bearing housing 2 is a cylindrical housing body 2a.
And an annular outer ring retainer 2c bolted to the flange portion 2b of the housing body 2a.

In such a structure, the diameter of the ball 10c is D, the curvatures of the raceway grooves 10e and 10f are Re and Ri, and the ball 1 is
When the contact angle between 0e and the raceway grooves 10e and 10f is α,
The four-point contact ball bearing 10 according to the present embodiment has a contact angle α of 1
Within the range of 2.5 ° <α ≦ 32.5 °, and the curvature ratio Re / D, R of the raceway grooves 10e, 10f to the diameter of the ball
i / D is in the range of Re / D = Ri / D = 0.54 to 0.61.

By the way, when the multi-point contact ball bearing is used as a ball screw support bearing, the points required for the multi-point contact ball bearing are that the dynamic torque is suppressed within a certain range, and the smooth rotation start-up and rotation start-up are achieved. It is a point that can get downhill. In order to satisfy such a requirement, the present inventors conducted computer analysis of various combination conditions, and found Examples A to D having specifications as shown in Table 1 from the analysis results.

[0027]

[Table 1]

In Table 1, Example A has a contact angle α of 1
2.5 ° <α ≦ 17.5 ° and the diameter D of the ball 10c
The curvature ratios fi and fe of the orbital grooves 10e and 10f (hereinafter referred to as "groove curvature ratio") with respect to 0.54≤fi≤0.55
No. 5, 0.54 ≦ fe ≦ 0.555, and the contact angle α is 17.5 ° <α ≦ 22.
5 °, and the groove curvature ratios fi and fe are 0.555 <fi
It is a 4-point contact ball bearing with ≦ 0.57, 0.555 <fe ≦ 0.57. Further, in Example C, the contact angle α was 22.5.
° <α ≦ 27.5 °, and the groove curvature ratios fi and fe are 0.57 <fi ≦ 0.59 and 0.57 <fe ≦ 0.59.
And the contact angle α is 2 in Example D.
7.5 ° <α ≦ 32.5 °, and the groove curvature ratios fi and f
e is 0.59 <fi ≦ 0.61, 0.59 <fe ≦ 0.
It is a four-point contact ball bearing 61.

Examples A to D shown in Table 1 and a conventional example (double row angular contact ball bearing (7010A, bearing inner diameter: 50
Table 2 shows the ball diameter, the number of balls, the ball PCD, and the selected values for the preload clearance in mm).

[0030]

[Table 2]

The contact angle α of the conventional example E (condition in which the dynamic torque is the largest in the manufacturing range in the conventional bearing) and Examples A to D shown in Table 2 is α _E = 30 °, α _A = 17. 5 °,
α _B = 22.5 °, α _C = 27.5 °, α _D = 32.5
And the groove curvature ratios fe and fi are fe _E = fi _E = 0.
52, fe _A = fi _A = 0.54, fe _B = fi _B =
0.556, fe _C = fi _C = 0.571, fe _D = f
Table 3 shows the dynamic torque when i _D = 0.591.

[0032]

[Table 3]

From the results shown in Table 3, the contact angle α is 17.5 ° ≦
With α ≦ 32.5 °, the groove curvature ratios fe and fi are set to fe = fi
= 0.54 to 0.591, the dynamic torque Tr becomes Tr
= 0.28 to 0.282 N · m, which is the same as that of the conventional example E.
It can be seen that the value is the same as or lower than Luk. Next
And the contact angle α of Examples A to D is_A= 15 °, α_B= 2
0 °, α_C= 25 °, α_D= 30 °, groove curvature ratio fe
And fi to fe_A= Fi_A= 0.55, fe _B= Fi_B
= 0.56, fe_C= Fi_C= 0.58, fe_D= Fi
_DTable 4 shows the dynamic torque when = 0.60.
You

[0034]

[Table 4]

From the results shown in Table 4, the contact angle α is 15 ° ≦ α ≦
The groove curvature ratios fe and fi are set to 30 ° and fe = fi = 0.5.
When it is set to 5 to 0.60, the dynamic torque Tr becomes Tr = 0.24.
Up to 0.255 N · m, the dynamic speed of the conventional example E shown in Table 2
It can be seen that the value is lower than Luk. Next, conventional example E
(The conventional bearing has the smallest dynamic torque within the manufacturing range.
Condition) and the contact angle α of Examples A to D is α_E= 29 °, α
_A= 12.6 °, α_B= 17.6 °, α_C= 22.6
°, α_D= 27.6 °, and the groove curvature ratios fe and fi are f
e _E= Fi_E= 0.53, fe_A= Fi_A= 0.55
5, fe_B= Fi_B= 0.57, fe_C= Fi_C= 0.
59, fe_D= Fi_DMotion when = 0.61
The torque is shown in Table 5.

[0036]

[Table 5]

From the results shown in Table 5, the contact angle α is 12.6 ° ≦
With α ≦ 27.6 °, the groove curvature ratios fe and fi are set to fe = fi
= 0.555 to 0.61, the dynamic torque Tr becomes Tr
= 0.215 to 0.222 N · m, which is larger than the dynamic torque of Conventional Example E shown in Table 2. That is, by adopting Examples A to D shown in Table 1, it becomes possible to suppress the dynamic torque within the manufacturing performance range of the conventional example.

Based on the results of Tables 3 to 5, the relationship between the dynamic torque Tr, the contact angle α, and the groove curvature ratios fi and fe is three-dimensionally analyzed. The results are shown in FIG. 5, and the dynamic torque Tr and the groove curvature are shown in FIG. Ratio f
The result of analyzing the relationship between i and fe is shown in FIG. As is apparent from the analysis results of FIGS. 5 and 6, it is necessary to combine the contact angle α and the groove curvature ratios fi and fe with appropriate values in order to suppress the dynamic torque. Further, the contact angle α is set to 12.6 ° ≦ α ≦ 27.6 °, and the groove curvature ratio fe,
When fi is set to fe = fi = 0.555 to 0.61, the dynamic torque becomes smaller. However, if the dynamic torque is too small, it cannot be stopped quickly due to the deceleration inertia when the rotation of the ball screw is stopped. There is a possibility of overshooting the indicated position. Further, the axial rigidity of the bearing becomes small, and again, overshooting is likely to occur. Further, since the axial load capacity also becomes small, there arises a problem that the rolling fatigue life becomes short. From this, in order to suppress the dynamic torque within the range of variation of the conventional example and obtain smooth rotation rising and falling characteristics,
It is understood that it is necessary to combine the contact angle α and the groove curvature ratios fe and fi within the range shown in Table 1.

Next, the contact angle α of Examples A to D is α_A= 1
5 °, α_B= 20 °, α_C= 25 °, α_D= 30 °
The groove curvature ratios fe and fi to fe_A= Fi_A= 0.5
5, fe _B= Fi_B= 0.56, fe_C= Fi_C= 0.
58, fe_D= Fi_D= 0.60
Table 6 shows the PV value of the bearing.

[0040]

[Table 6]

From the results of Table 6, the contact angle α was set to 15 ° ≦ α ≦
The groove curvature ratios fe and fi are set to 30 ° and fe = fi = 0.5.
By setting it to 5 to 0.60, the PV value PVi between the inner ring and the ball becomes PVi = 47 to 80 MPa · m / s, and the PV value PVe between the outer ring and the ball becomes PVe = 32 to 58 MP.
It is found that the value is a · m / s, which is higher than that of the conventional example.

FIG. 7 is a diagram showing the relationship between the PV value of the inner ring, the groove curvature ratio, and the contact angle based on the results of Table 6, and FIG.
It is a figure which shows the relationship between an inner ring PV value and a groove curvature ratio based on the result of. 9 is a diagram showing the relationship between the outer ring PV value, the groove curvature ratio and the contact angle based on the results of Table 6, and FIG. 10 is the same as the outer ring PV value and the groove curvature ratio based on the results of Table 6. It is a figure which shows a relationship.

As can be seen from FIGS. 7 to 10, the magnitude of the PV value affects not only the dynamic torque and the bearing temperature rise but also the wear resistance of the contact portion. Recently, high-speed rotation of a ball screw is required for an electronic component mounting apparatus for mounting an electronic component such as a semiconductor chip on a substrate and an electrolytic processing machine such as an electric discharge machine and a laser machine. In order to satisfy the wear resistance characteristics under unstable sliding conditions during sudden acceleration and sudden deceleration, it is necessary to make the PV value smaller. Conventional Example E and Examples A to D
The PV values in Table 1 are summarized in Table 6. When the dynamic torque value is constant, the smaller the contact angle and the smaller the groove curvature ratio, the smaller the PV value (Example A). All specifications are below the limit PV value judged from past bearing performance, so there are very few problems of wear, but in response to the weight of the ball screw rotating part and the G generated during sudden acceleration and deceleration, Depending on the required axial rigidity required value and the bearing temperature rise characteristic, any of the devised specifications (Example A
~ D) may be selected.

Next, the contact angle α of Examples A to D is set to α_A= 1
5 °, α_B= 20 °, α_C= 25 °, α_D= 30 °
The groove curvature ratios fe and fi to fe_A= Fi_A= 0.5
5, fe _B= Fi_B= 0.56, fe_C= Fi_C= 0.
58, fe_D= Fi_D= 0.60
Axial displacement of bearing ad (μm / 1000N: 1000N
Axial displacement when a axial load is applied)
Table 7 shows the results of the analytical calculation.

[0045]

[Table 7]

From the results of Table 7, the contact angle α is set to 15 ° ≦ α ≦
The groove curvature ratios fe and fi are set to 30 ° and fe = fi = 0.5.
It can be seen that the axial displacement ad of each bearing is ad = 4 to 14 by setting the value to 5 to 0.60, and in any case, the axial displacement ad is equal to or larger than the axial displacement of the conventional example. The axial rigidity of the ball screw affects not only the bearing for supporting the ball screw but also the rigidity between the ball screw shaft and the nut and the rigidity of the ball screw shaft itself. Also, the required rigidity value changes depending on the type of machine used and the usage conditions.

From the results shown in Table 7, the larger the contact angle and the larger the groove curvature ratio (Example D), the greater the axial rigidity. In consideration of the above-mentioned PV value requirement, any of the invention specifications (Examples A to D) may be selected according to the required axial rigidity requirement. FIG. 11 shows three types of preload clearances (Δα = −5 μm, −10 μm, −15 μm) for dynamic torque when an axial load is externally applied to a preloaded multi-point contact ball bearing (corresponding to Example B). It is a figure which shows the result analyzed by the conditions of. It can be seen that in any of the preload gaps, the dynamic torque tends to increase temporarily with an increase in the axial load, then decrease and increase again. This is because the axial elastic deformation of the bearing was caused by the axial load, and the contact point between the inner and outer rings and the ball changed from four-point contact at the time of initial external load (0N) to three-point or two-point contact. . In this way, by shifting from the initial four-point contact to the two-point contact as the axial load increases, the dynamic torque can be reduced as a result even if the axial load increases.

Further, as can be seen from FIG. 11, if the preload gap is set to be small, the maximum value of the dynamic torque during the transitional period when the contact shifts from four points to two points can be reduced. This means that if you replace it with the actual operating conditions, you can select the appropriate preload clearance.
This means that the increase in dynamic torque due to the reciprocating inertia load during sudden acceleration and sudden deceleration can be reduced by decreasing the number of contact points. As a result of various analyses, in order to minimize the axial deviation of the ball screw and obtain smooth rotation characteristics, as a result of considering actual assumed load conditions, the preload clearance should be 0-
It was found that the dynamic torque reduction effect is brought about by setting the thickness to about 10 μm.

From the above, the ball 10e and the raceway groove 10
The contact angle α with e and 10f is 12.5 ° <α ≦ 32.5 °
And the raceway grooves 10e, 10 with respect to the diameter of the ball 10e.
The curvature ratios fe and fi of f are fe and fi = 0.54 to 0.6
By setting the value to 1, the dynamic torque can be suppressed to a low value and smooth rotation rising and falling characteristics can be obtained. Therefore, the four-point contact ball bearings of Examples A to D are suitable as a ball screw supporting bearing. It can be used. In particular, Examples A and D have a slightly larger deterioration in function in terms of PV value and axial rigidity as compared with conventional products, and Example B or Example C is well balanced, It can be used more suitably as a bearing for supporting a ball screw.

Further, unlike the above-mentioned conventional example, it is not necessary to use a combination of single-row axial ball bearings in two rows, so that the ball screw of the ball screw can be installed in about half the mounting space as compared with the conventional one. The shaft can be supported. Furthermore, in the above-described embodiment, the outer ring 10a or the inner ring 10b.
Since the annular seal bodies 12 are provided at both ends of the bearing and the annular gap formed between the outer ring 10a and the inner ring 10b is sealed by the annular seal bodies 12, foreign matter, liquid, etc. are prevented from entering the bearing. it can.

Next, the test bearings A to
E (bearing inner diameter: 12 mm, bearing outer diameter: 28 mm, ball pitch circle diameter: 20 mm) was used to determine the dynamic torque characteristics of the ball bearing. Rotation speed: 10 to 5000 min ^-1 (5000 min)
^-1 is equivalent to 100,000 as the bearing dmn value), lubrication: grease (base oil viscosity; 15 × 10 ⁻⁶ m ² / s) was tested under the conditions shown in FIG. 12, and the temperature rise characteristics of the ball bearing. FIG. 13 shows the result of testing under the same conditions as described above.

[0052]

[Table 8]

From the test results shown in FIGS. 12 and 13, the rotation speed of Examples A to D in Table 8 was 5000 min ^−1.
The dynamic torque and bearing temperature rise (bearing temperature-room temperature) of the conventional angular contact ball bearing (conventional example E) when (the maximum rotation speed range assumed when used as a ball screw multi-point contact ball bearing) It can be seen that they show almost the same values as those of.

Further, from the test results shown in FIG. 12, in Examples A to D of Table 8, the change of the dynamic torque is smaller than that of the conventional example E when the rotation speed shifts from the low speed region to the high speed region. I understand. Therefore, in Examples A to D of Table 8, inertia fluctuations due to dynamic torque fluctuations during sudden acceleration / deceleration can be suppressed, and as a result, overshoot or the like when the ball screw suddenly stops can be more effectively prevented. be able to.

From the above, the multi-point contact ball bearing for supporting a ball screw according to the present invention can achieve space saving while maintaining dynamic torque characteristics and temperature rise characteristics equivalent to or higher than those of angular contact ball bearings. You can The present invention is not limited to the above embodiment. For example, in the above-described first embodiment, the contact type annular seal body is used as the annular seal body that seals the annular gap formed between the outer ring and the inner ring, but in the second embodiment shown in FIG. As described above, the annular gap formed between the outer race 10a and the inner race 10b may be sealed by the non-contact type annular seal body 15. Further, in the above-described first embodiment, an annular seal body made of an elastic material such as resin is used as a means for sealing the annular gap formed between the outer race and the inner race, but the third embodiment shown in FIG. As in the embodiment, the annular seal body 16 made of metal may be used. Further, in each of the above-described embodiments, the 4-point contact ball bearing is used as the ball screw supporting bearing, but a 3-point contact ball bearing may be used instead of the 4-point contact ball bearing (see FIGS. 16 and 19). ).

[0056]

As described above, according to the present invention,
As a ball screw support bearing that requires precise positioning, it is possible to replace the same function as a double row combination angular contact ball bearing with a single row multipoint contact ball bearing. As a result, the installation space required is about half that of the conventional product, the design space can be saved by reducing the machine space, the cost can be reduced by simplifying the structure, the operating part can be made lighter, and the tact time improves production efficiency. You can also contribute to up. Also, by providing annular seals at both ends of the outer ring or inner ring, and by sealing the annular gap formed between the outer ring and the inner ring by these annular seals, it is possible to prevent foreign matter from entering and to prevent dust and water. Can be improved.
Furthermore, the inner ring can be prevented from falling off, and the handling and assembly can be improved.

[Brief description of drawings]

FIG. 1 is a sectional view showing a structure of a ball screw supporting bearing according to a first embodiment of the present invention.

FIG. 2 is an axial sectional view of the four-point contact ball bearing shown in FIG.

3 is a sectional view showing a part of the annular seal body shown in FIG.

FIG. 4 is a perspective view showing a part of a cage.

FIG. 5 is a diagram showing a three-dimensional relationship between the dynamic torque, the contact angle, and the groove curvature ratio based on the results of Tables 3 to 5.

FIG. 6 is a diagram showing the relationship between dynamic torque and groove curvature ratio based on the results of Tables 3-5.

FIG. 7 is a diagram showing three-dimensionally the relationship between the inner ring PV value, the contact angle, and the groove curvature ratio based on the results of Table 6.

FIG. 8 is a diagram showing a relationship between an inner ring PV value and a groove curvature ratio based on the results of Table 6.

FIG. 9 is a diagram showing a three-dimensional relationship between the outer ring PV value, the contact angle, and the groove curvature ratio based on the results of Table 6.

FIG. 10 is a diagram showing the relationship between the outer ring PV value and the groove curvature ratio based on the results of Table 6.

FIG. 11 is a diagram showing a result of analyzing a dynamic torque when an axial load is applied from the outside.

FIG. 12 is a diagram showing the relationship between the dynamic torque and the rotational speed of a ball bearing.

FIG. 13 is a diagram showing a relationship between a bearing temperature rise (bearing temperature-room temperature) and a rotation speed of a ball bearing.

FIG. 14 is an axial cross-sectional view of a ball screw supporting bearing according to a second embodiment of the present invention.

FIG. 15 is an axial cross-sectional view of a ball screw supporting bearing according to a third embodiment of the present invention.

16A and 16B are views showing the structure of a three-point contact ball bearing, wherein FIG. 16A is an axial sectional view of an outer ring, and FIG. 16B is an axial sectional view of an inner ring.

FIG. 17 is a view showing the structure of a four-point contact ball bearing, (a) is an axial sectional view of the outer ring, and (b) is an axial sectional view of the inner ring.

FIG. 18 is an axial sectional view of a four-point contact ball bearing.

FIG. 19 is an axial sectional view of a three-point contact ball bearing.

FIG. 20 is a sectional view showing a conventional ball screw supporting bearing.

21 is an axial cross-sectional view of the double-row combination ball bearing shown in FIG.

[Explanation of symbols]

2 bearing housing 3 oil seal 5 ball screw shaft 6 Presser ring 7 Bearing nut 10 4-point contact ball bearing 10a outer ring 10b inner ring 10c ball 10d cage 10e Track groove 10f Track groove 12 Annular seal body 13 Seal mounting groove 14 Seal groove 15 Annular seal body 16 Annular seal body

   ─────────────────────────────────────────────────── ─── Continued front page    F term (reference) 3J012 AB04 BB03 CB01 EB14 FB10                       HB01 HB02                 3J016 AA02 AA03 BB03 CA02                 3J062 AA22 AA28 AB22 AC07 BA14                       CD34                 3J101 AA04 AA32 AA42 AA54 AA62                       BA64 FA01 FA41 GA31 GA55

Claims (5)

[Claims] 1. A raceway groove formed in an outer ring and an inner ring is provided with 3
A multi-point contact ball bearing having a large number of balls contacting at points or more, wherein annular seal bodies are provided at both ends of the outer ring or the inner ring, and the annular seal body forms an annular ring between the outer ring and the inner ring. A multi-point contact ball bearing for supporting a ball screw, characterized by having a sealed gap. 2. The contact angle α is 12.5 ° <α ≦ 17.5, where D is the diameter of the ball, R is the curvature of the raceway groove, and α is the contact angle between the ball and the raceway groove. And the curvature ratio R / D of the raceway groove to the diameter of the ball is R / D =
0.54 to 0.555 is set, The claim 1 characterized by the above-mentioned.
A multipoint contact ball bearing for supporting a ball screw as described. 3. The contact angle α is 17.5 ° <α ≦ 22.5, where D is the diameter of the ball, R is the curvature of the raceway groove, and α is the contact angle between the ball and the raceway groove. And the curvature ratio R / D of the raceway groove to the diameter of the ball is 0.555.
The multi-point contact ball bearing for supporting a ball screw according to claim 1, wherein <R / D ≦ 0.57. 4. The contact angle α is 22.5 ° <α ≦ 27.5, where D is the diameter of the ball, R is the curvature of the raceway groove, and α is the contact angle between the ball and the raceway groove. And the curvature ratio R / D of the raceway groove to the diameter of the ball is 0.57 <
The multipoint contact ball bearing for supporting a ball screw according to claim 1, wherein R / D ≦ 0.59. 5. The contact angle α is 27.5 ° <α ≦ 32.5, where D is the diameter of the ball, R is the curvature of the raceway groove, and α is the contact angle between the ball and the raceway groove. And the curvature ratio R / D of the raceway groove to the diameter of the ball is 0.59 <
The multipoint contact ball bearing for supporting a ball screw according to claim 1, wherein R / D ≦ 0.61.