JP2008082360A - Toroidal type continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To carry out securement of a gear ratio width and efficiency improvement (fuel consumption improvement) while preventing enlargement. <P>SOLUTION: An output side disc 10 is composed of a single member, and an output gear 9b is directly formed on an outer circumference part of the output side disc. Small diameter both ends of such an output side disc 10 are rotatably supported by pure thrust ball bearings 23, 23 with respect to struts 20a, 20a fixed to an inner face of an actuator body 21. At least three radial needle bearings 29a, 29b, 29c are provided between an inner circumference face of the output side disc 10 and an outer circumference face of an input rotary shaft 1. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to an improvement of a toroidal continuously variable transmission that is used, for example, as an automatic transmission for an automobile or as a transmission for adjusting the operating speed of various industrial machines such as a pump, while preventing an increase in size. Therefore, it is intended to ensure the transmission gear ratio range and increase the efficiency (improvement of fuel consumption).

  For example, as described in Patent Documents 1 to 6, the use of a toroidal type continuously variable transmission as an automatic transmission for automobiles has been studied and implemented in part. FIG. 3 shows a so-called double cavity type of such a toroidal type continuously variable transmission. This toroidal-type continuously variable transmission has a pair of input sides, each of which is an outer disk described in the claims, at two positions in the axial direction of the input rotary shaft 1 which is the rotary shaft described in the claims. The disks 2a and 2b are supported. These input side disks 2a, 2b are each a toroidal curved surface (concave arc-shaped concave surface) with respect to the input rotation shaft 1, and are on the input side corresponding to one axial side surface recited in the claims. In a state where the side surfaces 3 and 3 are opposed to each other, concentric and synchronized rotation is supported freely.

  Further, the intermediate portion of the input rotary shaft 1 is inserted through a through hole 6 provided in a partition wall portion 5 installed in a casing 4 housing a toroidal type continuously variable transmission. A cylindrical output cylinder 7 is rotatably supported by a pair of rolling bearings 8, 8 on the inner diameter side of the through hole 6, and an output gear 9 is fixed to the outer peripheral surface of the intermediate portion of the output cylinder 7. Has been established. Further, a pair of output side disks 10a and 10b corresponding to the inner disks described in the claims are respectively connected to the protruding portions of both ends of the output cylinder 7 from the outer side surfaces of the partition wall 5 by splines. By the engagement, the rotation synchronized with the output cylinder 7 is freely supported. In this state, the output side inner surfaces 11, 11 of the output side disks 10a, 10b, each of which is a toroidal curved surface (concave arc-shaped concave surface) and corresponding to both side surfaces in the axial direction recited in the claims. Is opposed to the input side inner surfaces 3 and 3.

  In addition, a plurality (generally two or three) of power rollers 12 are provided in a portion (cavity) between the input side and output side inner side surfaces 3 and 11 around the input rotation shaft 1. , 12 are arranged. Each of these power rollers 12 and 12 has a spherical convex surface on the peripheral surfaces 13 and 13 that come into contact (rolling contact) with the inner surfaces 3 and 11 on both the input and output sides, and is described in the claims. The trunnions 14 and 14 corresponding to the supporting members are supported on the inner surfaces of the trunnions 14 and 14 so as to be freely rotatable and slightly oscillating. Each of the trunnions 14 and 14 is provided with a swinging displacement about a pivot that is twisted with respect to the input rotary shaft 1.

  When changing the gear ratio between the input side disks 2a and 2b and the output side disks 10a and 10b, the trunnions 14 and 14 are displaced in the axial direction of the pivot (front and back direction in FIG. 3). As a result, rolling contact portions between the peripheral surfaces 13 and 13 of the power rollers 12 and 12 and the input and output inner surfaces 3 and 11 of the input and output disks 2a, 2b, 10a and 10b ( The direction of the tangential force acting on the traction section changes (side slip occurs in the rolling contact section). As the direction of the force changes, the trunnions 14 and 14 swing (tilt) about the pivot, and the peripheral surfaces 13 and 13 of the power rollers 12 and 12 and the input side and output The contact positions of the side disks 2a, 2b, 10a and 10b with the input side and output side inner surfaces 3 and 5 change.

  For example, as shown in FIG. 3, the peripheral surfaces 13, 13 of each of the power rollers 12, 12 are connected to the radially inward portions of the input side inner surfaces 3, 3 of the input side disks 2a, 2b and the output side disk. If it is brought into rolling contact with the radially outer portions of the output side inner surfaces 11, 11 of 10a, 10b, the gear ratio between the two disks 2a, 2b, 10a, 10b becomes the deceleration side. On the other hand, conversely to FIG. 3, the peripheral surfaces 13 and 13 of the power rollers 12 and 12, and the radially outward portions of the input side inner surfaces 3 and 3 of the input side disks 2 a and 2 b, If the rolling contact is made with the radially inner portions of the output side inner surfaces 11, 11 of the output side disks 10a, 10b, the gear ratio between the two disks 2a, 2b, 10a, 10b will be increased. Become.

  During operation of the toroidal-type continuously variable transmission as described above, one input side disk 2a is connected via a loading cam type pressing device 16 by a drive shaft 15 connected to a power source such as an engine. Rotating drive. As a result, the pair of input-side disks 2a and 2b supported at both ends of the input rotation shaft 1 rotate synchronously while being pressed in a direction approaching each other. Then, this rotation is transmitted to both the output side disks 10a and 10b via the power rollers 12 and 12, and is taken out from the output gear 9.

  In the case of the structure shown in FIG. 3 as described above, similarly to the structure described in Patent Document 1, the output gear 9 and another gear 17 that meshes with the output gear 9 by a pair of rolling bearings 8, 8. The output side disks 10a and 10b are positioned in the axial direction (supporting axial displacement) while supporting the load applied to the meshing portion. That is, a thrust load (axial load) Fa and a radial load Fr are applied to the meshing portions of the gears 9 and 17 based on the meshing of the helical gears constituting the gears 9 and 17. Further, based on the thrust load Fa, the output gear 9 is caused to have a moment M (in the direction of shifting the center axis of the output gear 9 and the center axis of the input rotary shaft 1) in the direction in which the output gear 9 is inclined (turned down). M = P · Fa) is applied when the moment load M and the pitch circle diameter of the output gear 9 are 2P. The rolling bearings 8 and 8 position the output disks 10a and 10b while supporting the loads Fa and Fr and the moment M, respectively. However, in the case of such a structure, since the output gear 9 and the partition wall 5 and further the rolling bearings 8 and 8 are provided between the output side disks 10a and 10b, the toroidal type The axial dimension of the entire stage transmission increases.

  On the other hand, in Patent Document 2, as shown in FIG. 4, the output side disk 10 is constituted by a single member having a shape such that a pair of output side disks 10a and 10b (FIG. 3) are coupled and fixed. In addition, a structure in which an output gear 9a is provided on the outer peripheral surface of the intermediate portion of the output side disk 10 is described. In the case of such a structure, compared to the structure shown in FIG. 3, the output gear 9, the partition wall 5, and the rolling bearings 8 and 8 (FIG. 3) are not provided. The axial dimension of the entire machine can be reduced. However, in the case of the structure shown in FIG. 4, the output side disk 10 is not positioned in the axial direction (axial displacement is not prevented). Therefore, the pressing force applied to the rolling contact portion between the inner side surfaces 3 and 11 of the input side and output side disks 2a, 2b and 10 and the peripheral surfaces 13 and 13 of the power rollers 12 and 12 is different between the cavities. May be different from each other.

  That is, based on the pressing force generated by the loading cam type pressing device 16, the pressing force applied to the output side disk 10 is equal from the power rollers 12, 12 existing in the respective cavities (in a direction to cancel each other). Join. However, in addition to such a pressing force, a thrust load Fa generated at the meshing portion between the output gear 9a and another gear 17a meshing with the output gear 9a is applied to the output side disk 10. Then, based on the thrust load Fa, the output side disk 10 tends to be displaced in a direction in which the thrust load Fa is applied. For this reason, there is a possibility that the pressing force of each rolling contact portion of the cavity on the side to which the thrust load Fa is applied becomes larger than the pressing force on the side to which the thrust load Fa is not applied. Note that Patent Document 3 describes a structure in which an output gear and another gear meshing with the output gear are each constituted by a helical gear. If such a structure is adopted, it is considered that the thrust load Fa based on the meshing can be canceled (cancelled) at the meshing portion between the spur gears. However, it is considered that the above-mentioned spur gears are troublesome to process and may increase the cost, and the thrust load based on the meshing cannot be completely reduced to zero. That is, there is a possibility that a thrust load Fa is applied to the meshing portion based on a phase shift between teeth that are inclined in opposite directions and that constitute the helical gear.

  On the other hand, in Patent Document 4, as shown in FIG. 5, an integrated output side disk 10 is arranged around the intermediate part of the input side rotating shaft 1, and a pair of radial needle bearings 18, 18 and 18 A structure in which thrust needle bearings 19 and 19 are rotatably supported is described. That is, both ends on the small diameter side of the output side disk 10 are supported by the thrust needle bearings 19 and 19 with respect to a pair of support columns 20 and 20 fixed to the inner surface of the casing 4a, and the input rotary shaft 1 A structure is described in which the radial needle bearings 18 and 18 corresponding to the roller bearings described in the claims are provided between the outer peripheral surface and the inner peripheral surface of the output side disk 10. According to such a structure described in Patent Document 4, the axial dimension can be reduced, and the axial positioning of the output side disk 10 can be achieved by the thrust needle bearings 19 and 19. And since these thrust needle bearings 19 and 19 can support the thrust load Fa generated at the meshing portion of the output gear 9a, it is possible to prevent the pressing force applied to each rolling contact portion of any cavity from increasing. it is conceivable that.

  However, in the case of the structure described in Patent Document 4 as described above, the following inconvenience may occur. That is, the thrust load Fa and the radial load Fr (see FIG. 3) are applied to the meshing portion of the output gear 9a as described above. Then, on the basis of the thrust load Fa, the output side disk 10 provided with the output gear 9a is inclined (tilted) in the direction of the output side disk 10 (the central axis of the output side disk 10 and the input rotating shaft 1). Moment M) in the direction of shifting from the center axis. Such a moment M constitutes the thrust needle bearings 19, 19 in the thrust needle bearings 19, 19 provided between the both ends of the small diameter side of the output side disk 10 and the struts 20, 20. Applied as a force in the direction of skewing each needle. As a result, each of the thrust needle bearings 19, 19 becomes difficult to rotate smoothly based on the skew (obstructing the rotation of the output side disk 10), which may cause a decrease in transmission efficiency.

  Further, Patent Document 4 also describes a structure in which the radial needle bearings 18 and 18 are omitted, and tapered roller bearings are used instead of the thrust needle bearings 19 and 19. However, in the case of such a structure, even if the above-described skew can be prevented, the tapered roller bearings are provided at both ends on the small diameter side of the output side disk 10, and therefore the diameters at both ends on the small diameter side are It cannot be avoided that it grows. When the diameters at both ends on the small diameter side increase in this way, the width in the radial direction of the output side inner surfaces 11, 11 that are the traction surfaces of the output side disk 10 decreases, and the transmission ratio width of the toroidal continuously variable transmission May be smaller. Although it is conceivable to increase the diameter of the output side disk 10 and to secure a width in the radial direction of the output side inner surfaces (traction surfaces) 11 and 11, the radial dimension of the toroidal type continuously variable transmission is It is unfavorable because it increases and becomes larger as a whole.

  Patent Document 5 describes a structure in which an output cylinder supporting a pair of output side disks is supported by a pair of radial ball bearings on a support ring of a support column provided in a casing. However, in the case of this structure, since the output side disk is not an integral type, the axial dimension is increased, and the load applied to the output side disk is supported only by the pair of radial ball bearings. There is a possibility that the radial ball bearing becomes large (it is necessary to make the allowable load large). When each radial ball bearing is increased in size as described above, the radial ball bearings (support rings of the supporting columns supporting the bearings) and the power rollers are likely to interfere with each other, resulting in a reduction in the gear ratio range. There is a possibility that the same inconvenience as in the case of the structure described in Patent Document 4 (a structure in which tapered roller bearings are employed) may occur.

Japanese Patent Laid-Open No. 1-229158 JP 11-63139 A JP 2004-197793 A Japanese Patent Laid-Open No. 2001-116097 JP 2003-207005 A Japanese Patent Laid-Open No. 11-141537

  The toroidal-type continuously variable transmission of the present invention takes into account the above-mentioned circumstances, increasing the size, i.e., suppressing the increase in the axial and radial dimensions of the inner disk such as the output-side disk, while reducing the gear ratio width. It was invented to realize a structure capable of ensuring and improving efficiency (improvement of fuel consumption).

The toroidal-type continuously variable transmission of the present invention includes, for example, a casing, a rotating shaft, a pair of outer disks, an inner disk, and a plurality of support members, as in the conventional structure shown in FIGS. (Trunnion) and a plurality of power rollers.
Of these, the casing houses the toroidal-type continuously variable transmission, and the rotating shaft is rotatably supported in the casing.
Each of the outer disks is supported on both ends of the rotating shaft so as to freely rotate in synchronization with the rotating shaft in a state where the axial side surfaces of the outer disks are opposed to each other. Yes.
Further, the inner disk rotates relative to the rotating shaft in a state in which both axial side surfaces having a circular arc cross section are opposed to one axial side surface of each outer disk around the intermediate portion of the rotating shaft. Is supported freely.

Each of the supporting members is twisted with respect to the rotating shaft in a plurality of positions (in the respective cavities) between the axial side surfaces of the inner disk and the axial side surfaces of the outer disks in the axial direction. Oscillating displacement about the pivot at the position is freely provided.
The power rollers are rotatably supported by the support members, and the circumferential surfaces of the spherical convex surfaces are in contact with both axial side surfaces of the inner disk and one axial side surface of each outer disk. (Rolling contact).
And the small diameter side edge part of the said inner side disk is rotatably supported with the rolling bearing with respect to the member which fixed to the inner surface of the said casing. At the same time, the inner disk is formed of a single member and is integrated, and a power transmission gear is provided concentrically with the inner disk on the outer peripheral edge of the inner disk.
This gear is constructed by, for example, a separate member from the inner disk as in the conventional structure shown in FIGS. 4 and 5 described above, and is coupled and fixed to the inner disk. It can also be formed directly on the periphery.

In particular, in the toroidal type continuously variable transmission of the present invention, the rolling bearing is a thrust ball bearing (more preferably a pure thrust ball bearing), and the outer peripheral surface of the rotating shaft and the outer peripheral surface of the rotating shaft. At least three roller bearings (for example, radial needle bearings) are provided between the inner peripheral surface of the inner disk and the inner disk. The number of these roller bearings is set according to the radial load to be supported and the magnitude of the moment. Specifically, for example, as the radial load is larger, the number of roller bearings located on the center side of the inner disk is increased.
In this case, more preferably, as described in claim 2, the crowning amount of each roller (for example, a needle) of the roller bearing located on the small-diameter side end portion side of the inner disk is also set to the same. Increase the crowning amount of each roller of the roller bearing located at the center of the inner disk. For example, the crowning amount of each roller of a pair of roller bearings positioned at both axial ends of the inner disk is made larger than the crowning amount of each roller of the remaining roller bearings.
Further preferably, as described in claim 3, among the roller bearings, a radial clearance of the roller bearing located on the small diameter side end side of the inner disk (from the diameter of the outer ring raceway to the diameter of the inner ring raceway). The value obtained by subtracting twice the diameter) is made smaller than the radial clearance of the roller bearing located on the center side of the inner disk. For example, the radial gaps of the pair of roller bearings located at both axial ends of the inner disk are made smaller than the radial gaps of the remaining roller bearings.

  According to the toroidal type continuously variable transmission of the present invention configured as described above, among the loads applied to the inner disk based on the meshing of the gears, the thrust load is similarly applied to the radial load and the moment by the pair of thrust ball bearings. Each can be supported by at least three roller bearings. Of these roller bearings, the radial load can be supported by all the roller bearings, and the moment can be supported by the roller bearings located on the small-diameter end side of the inner disk. In this case, even if the inner disk tends to fall (tilt) based on the moment, the thrust ball bearings do not cause a skew unlike the conventional structure described above. For this reason, the ball bearing does not hinder the rotation of the inner disk (the inner disk can be smoothly rotated). Each thrust ball bearing receives only the above thrust load (radial load and moment are supported by each roller bearing), and each of these thrust ball bearings is made large (with a large load capacity). There is no need. For this reason, it can prevent that the diameter of the small diameter side edge part of the said inner side disk becomes large, and can ensure sufficient gear ratio width | variety also with a small diameter inner side disk. For this reason, it is possible to secure a gear ratio range and increase efficiency (improve fuel efficiency) while suppressing an increase in the size of the toroidal continuously variable transmission.

Further, when the configuration described in claim 2 is adopted, even if the inner disk tends to fall (tilt) due to the moment, each roller of the roller bearing positioned on the small diameter end side of the inner disk. The surface pressure can be made uniform, and the occurrence of galling and excessive surface pressure (edge load) can be prevented. For this reason, it is possible to improve the durability of each of these rollers.
Further, when the configuration described in claim 3 is adopted, the moment of the load applied to the inner disk is positively caused by the roller bearing having a small radial clearance located on the small diameter end portion side of the inner disk. Can be supported. As a result, the load applied to other bearings (other roller bearings, ball bearings) can be reduced, and the bearing loss can be reduced and the durability can be improved.

  1 and 2 show an example of an embodiment of the present invention. The feature of this example is that by devising the support structure of the output side disk 10 which is the inner side disk described in the claims, the transmission ratio width is ensured and the efficiency is improved (the fuel efficiency is improved) while suppressing the increase in size. ). Since the structure and operation of the other parts are substantially the same as those of the conventional structure shown in FIGS. 3 to 5 described above, the overlapping description will be omitted or simplified, and the following description will focus on the characteristic parts of this example.

  In the case of this example, a state where a pair of support columns 20a and 20a are supported in a fixed part such as a casing via an actuator body 21 in each cavity respectively present on both axial sides of the output side disk 10 Is provided. Then, both axial ends of the output side disk 10 are provided on the support rings 22 and 22 provided in the intermediate portions of the respective struts 20a and 20a, and pure thrust ball bearings 23 which are thrust ball bearings described in claims. , 23 are rotatably supported. Each of these pure thrust ball bearings 23, 23 is rotated by a cage 26 between a plurality of balls 25, 25 between the respective raceways provided on the mutually opposing side surfaces of the pair of raceways 24 a, 24 b. The contact angle is set to 90 °.

  Short cylindrical ridges 27 and 27 are formed over the entire circumference of the outer ring surfaces (side surfaces opposite to each other) of the race rings 24a and 24b. The pure thrust ball bearings 23, 23 are fitted into the support rings 22, 22 and the small-diameter ends of the output side disk 10 without rattling. Positioning in the radial direction is intended. Further, shim plates 36, 36 are sandwiched between the outer surfaces of one of the race rings 24a, 24a and the support rings 22, 22, thereby positioning the pure thrust ball bearings 23, 23 in the axial direction. ing. In this state, a desired preload is applied to each of these pure thrust ball bearings 23 and 23. Accordingly, the output side disk 10 is rotatably supported in a state where positioning in the radial direction and the axial direction is achieved between the support columns 20a, 20a provided in pairs in each cavity.

  Further, between the inner peripheral surface of the through hole 28 provided at the center of the output side disk 10 supported as described above and the outer peripheral surface of the input rotary shaft 1 corresponding to the rotary shaft described in the claims. Further, at least three (three in this example) radial needle bearings 29a, 29b, and 29c, each corresponding to the roller bearing described in the claims, are provided. Each of these radial needle bearings 29a, 29b, 29c includes an outer ring raceway 30, an inner ring raceway 31, needles 32, 32 each corresponding to a roller described in the claims, and a retainer 33. Of these, each outer ring raceway 30 is constituted by the inner peripheral surface of the through hole 28. The inner ring raceway 31 is constituted by the bottom surfaces of recesses 34, 34 provided in the outer peripheral surface of the input rotary shaft 1 so as to be recessed from the outer peripheral surface. Such radial needle bearings 29a, 29b, and 29c are provided at both axial end portions (small-diameter side end portions) of the output side disk 10 and axially central portions, respectively.

In the case of this example, among the radial needle bearings 29a, 29b, and 29c, the crowning of the needles 32 and 32 of the pair of radial needle bearings 29a and 29b located at both axial ends of the output side disk 10 is provided. The amounts δ _{a and} δ _b (FIG. 2) are made larger than the crowning amounts δ _c of the needles 32 and 32 of the radial needle bearing 29c (the rest) which are also located in the central portion in the axial direction (δ _{a ,} Δ _b > δ _c ). Of the radial needle bearings 29a, 29b, 29c, a radial gap between a pair of radial needle bearings 29a, 29b located at both axial ends of the output side disk 10 (from the diameter of the outer ring raceway 30, the inner ring raceway The value obtained by subtracting the diameter of 31 and twice the diameter of the central portion in the axial direction of the needle 32) is made smaller than the radial clearance of the radial needle bearing 29c which is also located in the central portion in the axial direction. . The radial rotary bearings 29a, 29b, and 29c support the input rotary shaft 1 with respect to the output-side disk 10 so as to be freely rotatable and axially displaced. Such an input rotating shaft 1 has one input (left side in FIG. 1) according to the pressing force generated by the hydraulic pressing device 16a (according to the expansion / contraction of the hydraulic chamber 35 constituting the pressing device 16a). Along with the other input side disk 2b (to the right in FIG. 1), the side disk 2a is displaced in the axial direction.

  Further, an output gear 9b for power transmission corresponding to the gear described in the claims is provided concentrically with the output side disk 10 at the outer peripheral edge of the output side disk 10. In the case of this example, the output side disk 10 is constituted by a single member. That is, the output side inner surfaces 11 and 11 that are concave surfaces having an arc cross section forming the output side disk 10 are formed on both side surfaces in the axial direction of a single member, and the output side disk 10 is integrally formed. Forming. In the case of this example, the output gear 9b is directly formed on the outer peripheral edge of the output side disk 10, and the output side disk 10 and the output gear 9b are integrally formed. When such an output-side disk 10 is manufactured, necessary processing such as cutting, grinding, heat treatment or the like is performed on a single material having a shape such that a pair of disks are coupled and fixed as shown in FIG. The output side disk 10 is assumed. Incidentally, since such a method for manufacturing the output side disk 10 is described in, for example, Patent Document 6 and the like, detailed description thereof is omitted.

  According to the toroidal type continuously variable transmission of this example configured as described above, the thrust load among the loads applied to the output side disk 10 based on the meshing of the output gear 9b is converted into a pair of pure thrust ball bearings 23, 23. Similarly, the radial load and moment are supported by the three radial needle bearings 29a, 29b, and 29c, respectively. Of these radial needle bearings 29a, 29b, and 29c, a pair of radial needles positioned on both ends of the output-side disk 10 on the small-diameter side are subjected to the radial load by all the radial needle bearings 29a, 29b, and 29c. The moments are supported by the needle bearings 29a and 29b, respectively. In this case, even if the output-side disk 10 tends to fall (tilt) based on the moment, the balls constituting the pure thrust ball bearings 23, 23 have a skew as in the conventional structure described above. It never happens.

Therefore, these pure thrust ball bearings 23 and 23 do not hinder the rotation of the output side disk 10 (the output side disk 10 can be smoothly rotated). The pure thrust ball bearings 23, 23 receive only the thrust load (the radial loads and moments are supported by the radial needle bearings 29a, 29b, 29c). It is not necessary to make the size large (with a large allowable load). Further, it is not necessary to superimpose the main body portions of the pure thrust ball bearings 23 and 23 on the output side disk 10 and the input rotary shaft 1 in the radial direction. For this reason, it is possible to prevent the diameters of both ends on the small diameter side of the output side disk 10 from becoming large, and a sufficient transmission ratio width can be secured even with the small diameter output side disk 10. For this reason, it is possible to secure a gear ratio range and increase efficiency (improve fuel efficiency) while suppressing an increase in the size of the toroidal continuously variable transmission.
Although it is possible to adopt a thrust angular contact ball bearing instead of the pure thrust ball bearings 23, 23, in this case, the thrust angular contact ball bearing may receive a radial load. It is necessary to make the permissible load large. Further, in the case of the same diameter, since the allowable amount with respect to the thrust load is larger in the pure thrust ball bearing than in the angular type, it is preferable to adopt the pure thrust ball bearings 23 and 23 also from this aspect.

In the case of this example, as described above, the crowning amounts δ _a , δ _b , δ _c (FIG. 2) of the needles 32, 32 of the radial needle bearings 29a, 29b, 29c are used as the output side disk. Are regulated (δ _a , δ _b > δ _c ) in accordance with the positional relationship in the axial direction. For this reason, even if the output side disk 10 tends to fall (tilt) based on the moment, the surface pressure of the needles 32, 32 of the radial needle bearings 29a, 29b, 29c can be made uniform. It is possible to prevent galling and excessive surface pressure (edge load). Therefore, it is possible to improve the durability of the needles 32, 32 and the radial needle bearings 29a, 29b, 29c.

  In the case of this example, the radial clearances of the radial needle bearings 29a, 29b, and 29c are also restricted according to the positional relationship in the axial direction of the output side disk 10. That is, since the radial gap between the radial needle bearings 29a and 29b located on the small diameter end portion side of the output side disk 10 is reduced, the moment M applied to the output side disk 10 by the radial needle bearings 29a and 29b. Can be actively supported. Accordingly, the load applied to the radial needle bearing 29c and the pure thrust ball bearings 23, 23, which are other bearings, can be reduced, and the loss of these bearings 29c, 23 can be reduced and the durability can be improved.

Sectional drawing which cuts and shows one example of embodiment of this invention in a 90 degree different plane from FIG. The partial expanded sectional view seen from the same direction as Drawing 1 which takes out and shows the needle and cage which constitute a radial needle bearing. Sectional drawing which shows the 1st example of a conventional structure. Sectional drawing which shows the 2nd example. Sectional drawing which shows the 3rd example.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Input rotating shaft 2a, 2b Input side disk 3 Input side inner surface 4, 4a Casing 5 Bulkhead part 6 Through-hole 7 Output cylinder 8 Rolling bearing 9, 9a, 9b Output gear 10, 10a, 10b Output side disk 11 Output side inside Side surface 12 Power roller 13 Circumferential surface 14 Trunnion 15 Drive shaft 16, 16a Pressing device 17, 17a Another gear 18 Radial needle bearing 19 Thrust needle bearing 20, 20a Strut 21 Actuator body 22 Support ring 23 Pure thrust ball bearing 24a, 24b Track Wheel 25 Ball 26 Cage 27 Projection part 28 Through-hole 29a, 29b, 29c Radial needle bearing 30 Outer ring raceway 31 Inner ring raceway 32 Needle 33 Cage 34 Recess 35 Recess 35 Hydraulic chamber 36 Shim plate

Claims (3)

  A rotating shaft supported rotatably in the casing, and a rotating shaft supported on both ends of the rotating shaft in a state where the axial side surfaces of the rotating shaft are opposed to each other. A pair of outer disks supported so as to be able to freely rotate in a synchronized manner, and a state where both axial side surfaces having an arcuate cross section are opposed to one axial side surface of each outer disk around the intermediate portion of the rotating shaft And a plurality of inner disks that are supported to freely rotate relative to the rotating shaft, and a plurality of the inner disks in the axial direction between the axial side surfaces of the inner disks and the axial side surfaces of the outer disks. A support member that is freely provided with a swinging displacement centered on a pivot that is twisted with respect to the shaft, and a circumferential convex surface that is rotatably supported by each of the support members, and has a spherical convex surface. A power roller that is in contact with both axial sides and one axial side of each outer disk, and is rotatable by a rolling bearing with respect to a member in which the small-diameter end of the inner disk is fixed to the inner surface of the casing. A toroidal-type stepless in which the inner disk is formed of a single member and a gear for power transmission is provided on the outer peripheral edge of the inner disk concentrically with the inner disk. In the transmission, the rolling bearing is a thrust ball bearing, and at least three bearings are provided between the outer peripheral surface of the rotary shaft and the inner peripheral surface of the inner disk facing the outer peripheral surface of the rotary shaft. Toroidal-type continuously variable transmission, characterized by the provision of a roller bearing.   Among each roller bearing, the crowning amount of each roller of the roller bearing located on the small-diameter end side of the inner disk is made larger than the crowning amount of each roller of the roller bearing located on the center side of this inner disk. A toroidal continuously variable transmission according to claim 1.   Of each roller bearing, the radial clearance of the roller bearing located on the small-diameter side end side of the inner disk is made smaller than the radial clearance of the roller bearing located on the center side of the inner disk. 2. A toroidal continuously variable transmission according to any one of the above.

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