JPH07198014A - Roller support construction for troidal type continuously variable transmission - Google Patents

Abstract

PURPOSE:To avoid a problem in the case of supporting a power roller by a roller support member through a pivot shaft and movably support the power roller in the axial direction of input, output discs in a simple constitution, in a troidal type continuously variable transmission disposed between the input, output discs with the power roller in its pressure contact state. CONSTITUTION:A recessed groove 91 arranged opposite to a rotary shaft direction of input, output discs is formed on a trunnion in such a constitution as having an input disc and an output disc oppositely disposed coaxialy and a power roller 35 disposed between both discs in pressure contact state and a trunnion 38 for rotatably supporting this power roller 35. The lace member 94 of a thrust bearing 93 for supporting the power roller 35 is slidably supported in the recessed groove 91 along the rotary shaft direction of both discs.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a toroidal type continuously variable transmission, and more particularly to a support structure for a power roller arranged in a pressure contact state between input and output disks.

[0002]

2. Description of the Related Art A toroidal type continuously variable transmission mounted on an automobile or the like includes an input disk to which an engine output is input, an output disk coaxially arranged with the input disk, the output disk and the input disk. A power roller disposed so as to be capable of tilting in a state of being pressed against both the disks and the disk, transmitting torque input to the input disk to the output disk through the power roller, and tilting the power roller. By doing so, the contact position with respect to both disks is changed to adjust the speed ratio steplessly. In this type of toroidal type continuously variable transmission, the torque reaction force or the disk side acts. In order to support the power roller against the pressing force reaction force, the power roller is rotatably supported by using a supporting member called a trunnion. Going on.

By the way, in this type of toroidal transmission, misalignment between the input and output disks and the power roller is absorbed, and the pressing force applied to the power roller from the input disk side and the output disk side are applied. In order to balance the pressing force with the power roller,
It may be allowed to move freely along the direction of the axis of rotation of the output disc.

In the past, for example, the actual exploitation 64-27563.
As shown in the publication, the power roller is rotatably supported by a pivot shaft which is rotatably supported by a trunnion.

[0005]

However, in the case where the power roller is supported by the trunnion via the pivot as in the prior art described in the above publication, the following inconvenience may occur. It has been found.

That is, since the pivot shaft has an eccentric shape, it is difficult to process and the cost of parts is high.
Moreover, since it is necessary to form a hole in the trunnion for supporting the pivot shaft, the trunnion also becomes large and heavy in order to secure the strength.

Particularly, since the center of rotation of the power roller on the pivot shaft is arranged in a state of being offset from the center of attachment to the trunnion, the center of rotation of the power roller oscillates with respect to the trunnion, and Not only in the axial direction, but also in the direction orthogonal to the plane including the rotary shafts of both disks and the rotary shaft of the power roller. Then, due to the displacement in the orthogonal direction, the tilt angle of the power roller varies, and the gear ratio cannot be controlled accurately.

The present invention avoids the above problems in the case where the power roller is supported by the roller support member via the pivot shaft in the toroidal type continuously variable transmission in which the power roller is arranged in a pressure contact state between the input and output disks. Then, the object is to support the power roller with a simple structure so as to be movable in the axial direction of the output disk.

[0009]

That is, the invention according to claim 1 of the present application (hereinafter referred to as the first invention) is such that an input disk and an output disk that are coaxially opposed to each other and a pressure contact state between both disks are provided. In a toroidal type continuously variable transmission having a power roller arranged and a roller supporting member for rotatably supporting the power roller, a concave groove oriented in the rotational axis direction of the input and output disks is formed in the roller supporting member. A race bearing member, which is formed and supports the power roller, is slidably supported in a concave groove of the roller support member along the rotational axis directions of the both disks.

The invention according to claim 2 of the present application (hereinafter referred to as the second invention) is, in addition to the configuration of the first invention, a race member of a thrust bearing for rotatably supporting the power roller, and a roller support. A linear bearing that facilitates movement of the race member along the groove is provided between the groove and the bottom surface of the groove of the member.

[0011]

According to the above structure, the following effects can be obtained.

That is, in both the first and second aspects of the invention, a groove is formed in the roller supporting member and oriented in the direction of the rotation axis of the output disk, and the power roller is rotatably supported in the groove. Since the thrust bearing race member is slidably supported along the rotation axis direction of the both disks, it is assumed that the thrust bearing race member slides along the groove to absorb misalignment. However, the rotation center of the power roller does not move in the direction orthogonal to the plane including the rotation center and the rotation center of the output disk. This allows
There is no variation in the tilt angle of the power roller,
Good control ratio of the gear ratio can be obtained.

Moreover, since the race bearing member of the thrust bearing need only be supported in the concave groove formed in the roller supporting member, the pivot shaft which is difficult to process is not required, the supporting structure of the power roller is simplified, and the manufacturing cost is reduced. Will be reduced. Further, since it is not necessary to provide a hole for supporting the pivot shaft in the roller supporting member, the roller supporting member is lightweight and compact.

In particular, according to the second aspect of the invention, the movement of the race member along the concave groove is promoted between the race bearing thrust member supporting the power roller and the bottom surface of the concave groove of the roller supporting member. Since the linear bearing is installed, the race member moves smoothly against the thrust force acting through the power roller.

[0015]

Embodiments of the present invention will be described below with reference to the drawings.

As shown in FIG. 1, a power train 1 of an automobile according to this embodiment has an engine 2, a torque converter 3 connected to an engine output shaft 2a, and a deceleration to which the output of the torque converter 3 is transmitted. A device 10 and a toroidal type continuously variable transmission 30 into which the output of the engine 2 is input by bypassing the torque converter 3, and the output of the speed reducer 10 or the toroidal type continuously variable transmission 30 or both of them. Is transmitted to the drive wheels (not shown) via the output shaft 4.

The torque converter 3 has a pump 3b integrated with a casing 3a connected to the engine output shaft 2a.
And a turbine 3c that is arranged to face the pump 3b and is driven by the pump 3b via hydraulic oil, and a stator 3d that is interposed between the turbine 3c and the pump 3b to perform a torque increasing action. A front end side of a hollow shaft 3f, which is externally fitted to a turbine shaft 3e that rotates integrally with the turbine 3c, is connected to the stator 3d via a one-way clutch 3g. An intermediate portion of the hollow shaft 3f extending rearward is fixed to the transmission casing 60, and a rear end portion of the hollow shaft 3f and a rear end portion of the turbine shaft 3e are connected to the speed reducer 10.
Are linked to. Further, an oil pump 5 is provided at the shaft end of a pump shaft 3h which is fitted onto the hollow shaft 3f and has one end connected to the casing 3a. The oil pump 5 is provided via the casing 3a. It is driven by the engine 2.

The speed reducer 10 has a first planetary gear mechanism 11 and a second planetary gear mechanism 12 which are coaxially arranged in series with the turbine shaft 3e, and the first planetary gear mechanism 12 is arranged on the engine 2 side. The planetary gear mechanism 11 is for backward movement, and the second planetary gear mechanism 12 is for forward movement.

The first planetary gear mechanism 11 is of a single pinion type and has a sun gear 13 coupled to the turbine shaft 3e. A carrier 15 for rotatably supporting a pinion 14 meshing with the sun gear 13 is provided. The output shaft 4 is connected to the hollow shaft 3f fixed to the transmission casing 60, and the ring gear 16 meshing with the pinion 14 is arranged on the same axis as the turbine shaft 3e via the reverse clutch 17. Are linked to.

The second planetary gear mechanism 12 is of a double pinion type, and the inner pinion 18 has the first pinion.
It is integrated with the pinion 14 of the planetary gear mechanism 11, and the sun gear 13 of the first planetary gear mechanism 11 is shared by the sun gear of the second planetary gear mechanism 12. The carrier 20 that fixedly supports the inner pinion 18 and the outer pinion 19 is integrated with the carrier 15 of the first planetary gear mechanism 11 and is fixed to the transmission casing 60 via the hollow shaft 3f. Further, the ring gear 21 constituting the second planetary gear mechanism 12 is connected to the output shaft 4 via a forward clutch 22 and a one-way clutch 23. Therefore, when the reverse clutch 17 is engaged, the output of the turbine shaft 3e is transmitted to the output shaft 4 via the first planetary gear mechanism 11 and rotationally drives the output shaft 4 in the backward direction. When the forward clutch 22 is engaged, the output of the turbine shaft 3e is transmitted to the output shaft 4 via the second planetary gear mechanism 12, and the output shaft 4
Is driven to rotate in the forward direction.

On the other hand, the toroidal type continuously variable transmission 30
Has a pair of first and second continuously variable transmission mechanisms 31 and 32 arranged in series on the output shaft 4. These continuously variable transmission mechanisms 31 and 32 have the same structure, and are arranged such that an input disk 33 rotatably provided on the output shaft 4 with respect to the shaft 4 and the input disks 33 face each other. An output disk 34 that rotates integrally with the output shaft 4, and both disks 33, 34 between the output disk 34 and the input disk 33.
And a pair of power rollers 35, 35 which are arranged so as to be rotatable and tiltable in a state of being in contact with each other.

Then, the first and second continuously variable transmission mechanisms 3
Between the input discs 33, 33 of 1, 32, an intermediate disc 36 which is rotatable relative to the input discs 33 is arranged, and the intermediate disc 36 and the respective input discs 33, 33 are arranged. A plurality of loading cams 37 ... 37 are respectively interposed therebetween.
The loading cams 37 ... 37 are configured such that the larger the input torque input from the engine 2 to each input disk 33, the greater the pressing force of each cam 37 against each input disk 33.

Further, an input shaft 51 for inputting the output of the engine 2 to each input disk 33 via the intermediate disk 36 is arranged parallel to the output shaft 4. A first gear 52 is provided on the front end side of the input shaft 51,
The first gear 52 is meshed with an intermediate gear 53, and the intermediate gear 53 is meshed with an output gear 55 connected to the pump shaft 3h via a power distribution clutch 54. Further, on the rear end side of the input shaft 51,
Input gear 5 provided integrally with the intermediate disc 36
A second gear 57 that meshes with the gear 6 is integrally provided.
Therefore, when the power distribution clutch 54 is engaged, the output of the engine 2 bypassing the torque converter 3 is output via the input gear 56 to the first and second continuously variable transmission mechanisms 31 of the toroidal continuously variable transmission 30. , 32 are input to the input disks 33, 33, and the rotation of the input disks 33, 33 is changed at a predetermined speed ratio (reduction ratio) according to the tilt angle of the power rollers 35, ... It is adapted to be transmitted to the output disks 34, 34.

Next, the structure of the first and second continuously variable transmission mechanisms 31 and 32 constituting the toroidal type continuously variable transmission 30 will be described more specifically. Since the first continuously variable transmission mechanism 31 and the second continuously variable transmission mechanism 32 have the same configuration as described above, the first continuously variable transmission mechanism 31 will be described as a representative.

That is, as shown in FIG. 2, the first continuously variable transmission mechanism 31 is provided with first and second trunnions 38 and 39 arranged in the vertical direction, and these trunnions 38 and 39 are provided. The power rollers 35, 35 are rotatably supported.

On the other hand, on the support member 62 attached to the transmission casing 60 via the link post 61, the upper ends of the first and second trunnions 38 and 39 rotate via spherical bearings 63 and 63, respectively. A lower end portion of each of the first and second trunnions 38 and 39 is spherically supported by a support member 66 that is freely supported and is attached to a partition wall 64 that is integral with the transmission casing 60 via a link post 65. It is rotatably supported via bearings 67, 67. The trunnion shafts 40a and 40b extending in the direction orthogonal to the output shaft 4 are integrally attached to the first and second trunnions 38 and 39, respectively. These trunnion shafts 40a, 40
The front end sides of b respectively penetrate the partition wall 64 and project into the lower space covered with the oil pan 68.

Further, the partition wall 64 is provided with first and second hydraulic cylinders 71 and 72 for sliding the first and second trunnions 38 and 39 up and down. The hydraulic cylinders 71 and 72 are divided into upper hydraulic chambers 71a and 72a and lower hydraulic chambers 71b and 72b by partition walls 64a and 64a formed on the partition wall 64, respectively. Of these, the upper and lower hydraulic chambers 71 in the first hydraulic cylinder 71 on the first trunnion 38 side
Annular transmission pistons 73a and 73b, which are loosely fitted to the trunnion shaft 40a, are inserted in the a and 71b, respectively. A thrust bearing 41 is provided between the speed change piston 73a inserted in the upper hydraulic chamber 71a and the lower end of the first trunnion 38. In addition, the speed change piston 7 inserted in the lower hydraulic chamber 71b.
On the lower surface of 3b, a thrust bearing 42 is provided adjacent to the lower end of the trunnion shaft 40a. At the lower end portion of the trunnion shaft 40a, a recess cam 81 constituting a shift control mechanism 80 is spline-fitted adjacent to the thrust bearing 42, and the trunnion is in contact with the lower surface of the boss. The circlip 43 mounted on the shaft 40a supports the recess cam 81 and the speed change piston 73b.

On the other hand, the upper and lower hydraulic chambers 72a, 72b in the second hydraulic cylinder 72 on the second trunnion 39 side.
Also in this case, annular speed change pistons 74a and 74b, which are respectively loosely fitted to the trunnion shaft 40b, are inserted. Even in this case, the upper hydraulic chamber 72
A thrust bearing 41 is interposed between the speed change piston 74a inserted in a and the lower end of the second trunnion 39. Further, the thrust bearing 4 externally mounted on the lower end portion of the trunnion shaft 40b is provided on the lower surface of the speed change piston 74b inserted in the lower hydraulic chamber 72b.
2 are arranged adjacent to each other, and by mounting the circlip 43 on the trunnion shaft 40b in a state of being in contact with the lower end of the collar 44 arranged adjacent to the thrust bearing 42, the collar 44 and the shift piston 74b are It is supported.

Therefore, for example, the first hydraulic cylinder 71
In the lower hydraulic chamber 71b.
If it is made relatively higher than the operating pressure of
The first trunnion 38 is pushed up by the speed change piston 73a inserted in and is slid upward.
On the other hand, if the operating pressure of the upper hydraulic chamber 71a is made relatively lower than the operating pressure of the lower hydraulic chamber 71b, the trunnion shaft 40a is pushed down by the speed change piston 73b inserted in the lower hydraulic chamber 71b. Therefore, the first trunnion 38 slides downward accordingly.

Here, in the first cylinder 71, as shown in FIG. 3, for example, the center C1 of the lower hydraulic chamber 71b accommodating the speed change piston 73b is set to the center O1 of the trunnion shaft 40a of the first trunnion 38. , Partition wall 64
It is provided so as to be located closer to the center of the transmission casing 60. Also, regarding the first cylinder 72, for example, the lower hydraulic chamber 7 that houses the speed change piston 74b is used.
The center C2 of 2b is provided so as to be located in the center of the partition wall 64 or the transmission casing 60 with respect to the center O2 of the trunnion shaft 40b of the second trunnion 39. As a result, the piston diameter can be increased without increasing the size of the transmission casing 60, which reduces the required hydraulic pressure and improves the transmission efficiency.

Next, the shift control mechanism 80 that changes the gear ratio by controlling the supply and discharge of the operating pressure to and from the hydraulic chambers of the first and second hydraulic cylinders 71 and 72 in the first continuously variable transmission mechanism 31. The configuration of will be described.

That is, the first and second hydraulic cylinders 7 are provided on the lower surface of the partition wall 64 via the intermediate body 69.
The valve body 83 of the shift control valve 82 that switches the supply and discharge of the operating pressure to and from 1, 72 is fixed. A sleeve 84 is inserted in the valve body 83 so as to be movable in the axial direction, and a spool 85 is inserted in the sleeve 84 so as to be movable in the axial direction.

A stepping motor 86 is fixed to the side wall of the oil pan 68, and a conversion mechanism 87 for converting rotational movement into reciprocating movement is interlocked with a rotary shaft 86a of the stepping motor 86. The base end side of the sleeve 84 is connected to the conversion mechanism 87. Therefore, when the stepping motor 86 is rotationally driven, the sleeve 84 moves back and forth in the axial direction.

On the other hand, in front of the shift control valve 82, an L-shaped link 88 which is swingably supported is arranged. One end side of the L-shaped link 88 is disposed in contact with a cam surface of a precess cam 81 fixed to the lower end portion of the trunnion shaft 40a of the first trunnion 38, and the other end side of the link 88 is connected to the spool 85. Is engaged with the front end side of the. A spring 89 is arranged on the base end side of the spool 85, and the urging force of the spring 89 urges the tip end side of the spool 85 to contact the L-shaped link 88.

In that case, in the steady state, the pressing force acting on the tip end side of the spool 85 of the shift control valve 82 via the L-shaped link 88 and the pressing force by the spring 89 acting on the base end side of the spool 85. The force is balanced, and the supply and discharge of the operating pressure to and from the first and second hydraulic cylinders 71 and 72 are maintained in a stopped state.

If the stepping motor 86 is driven to move the sleeve 84 to the right (direction a) in the drawing in FIG. 2, the line pressure from a hydraulic source (not shown) causes the first hydraulic cylinder to move. While being guided to the lower hydraulic chamber 71b in 71 and the upper hydraulic chamber 72a in the second hydraulic cylinder 72, the operating pressure of the upper hydraulic chamber 71a in the first hydraulic cylinder 71 and the lower hydraulic chamber 72b in the second hydraulic cylinder 72 is exhausted. Will be done. As a result, the first trunnion 38 in the first continuously variable transmission mechanism 31 slides downward, and the second trunnion 39 slides upward.
The power roller 35 attached to the trunnion 38, 39,
When the contact position of 35 changes, a tilting force is generated. In that case, for example, if the input disk 33 is rotating in the b direction in FIG. 2, the power roller 35 on the first trunnion 38 side is rotated in the c direction in FIG. 1, and the second trunnion 39 side is rotated. Power roller 35 is d
Will rotate in the direction. As a result, the gear ratio (reduction ratio) between the input and output disks 33, 34 in the first continuously variable transmission mechanism 31 increases.

Then, the three-dimensional displacement of the first trunnion 38 due to the tilting of the power roller 35 is converted into the vertical displacement of the trunnion shaft 40a, and the displacement is transferred to the L-shaped link 89 via the precess cam 81. When transmitted, the link 89 is moved in the clockwise direction (e in FIG. 2).
Direction). Therefore, when the spool 85 moves to the right (direction f) against the urging force of the spring 89, and when the spool 85 moves by the movement amount of the sleeve 84, the movement stops and the shift control ends. . As a result, a predetermined target gear ratio corresponding to the operation amount of the stepping motor 85 is realized.

Here, the support structure of the power roller which constitutes the characteristic part of the present invention will be described. The first and second
Since the trunnions 38 and 39 have the same configuration, the first trunnion 38 will be described as a representative.

That is, for example, in the first trunnion 38, as shown in FIGS. 2 and 4, the input and output disks 33,
A concave groove 91 having a U-shaped cross section, which is oriented in the direction of the rotation axis of 34, is formed along the output shaft 4. On the other hand, the power roller 3
The race member 94, which constitutes the thrust bearing 93 by the back surface of the No. 5 and the plurality of balls 92, 92, is formed in the concave groove 91.
Is fitted so as to be rollable along the rotation axis direction.
Here, in the first trunnion 38, a concave hole 95 is formed at the bottom surface of the concave groove 91, and the concave hole 9 is formed.
A linear bearing 96 for promoting the movement of the race member 94 or the power roller 35 along the concave groove 91 is fitted to the shaft 5.

In this embodiment, a shaft portion 94a is provided at the center of the race member 94, and the power roller 35 is rotatably supported by the shaft portion 94a via a radial bearing 97. .

According to this structure, the power roller 3
The axial forces of the trunnion shafts 40a, 40b in the first and second trunnions 38, 39 acting on the shafts 5, 35 are the radial bearings 97, 97, the shaft portions 94a, 94.
It is received by the side surfaces of the concave grooves 91, 91 formed in the first and second trunnions 38, 39 via the a and the race members 94, 94. In addition, the power rollers 35, 3
The force in the direction orthogonal to the axial direction of the trunnion shafts 90a and 90b acting on the ball 5 is
And the race members 94, 94 through the concave grooves 91, 9
It will be received on the bottom of 1.

Since the recessed grooves 91, 91 are inserted and oriented along the rotational axis direction of the output disks 33, 34, for example, as shown by the arrow in FIG. It is possible to freely slide along the concave groove 91 together with the member 94, absorbs misalignment due to deformation and assembly of the input and output disks 33 and 34, and to the power roller 35. Thus, the pressing force applied from the input disk side and the pressing force applied from the output disk side are always equalized.

Moreover, the thrust bearings 93, 9 are
3 race members 94, 94 to the first and second trunnions 3
Since it is only necessary to support the concave grooves 91, 91 formed in the 8, 39, the pivot shaft which is difficult to process becomes unnecessary,
The support structure of the power rollers 35, 35 is simplified, and the manufacturing cost is reduced. Further, holes for supporting the pivot shaft are formed in the first and second trunnions 38,
Since it is not necessary to provide the trunnion 38, 39, the trunnions 38, 39 are made lightweight and compact.

In particular, as in the embodiment, the power roller 35
The race member 9 between the race members 94, 94 of the thrust bearings 93, 93 for supporting the race bearings and the bottom surfaces of the concave grooves 91, 91 of the first and second trunnions 38, 39.
Since the linear bearings 96, 96 that promote the movement of the grooves 4, 94 along the concave grooves 91, 91 are provided, the race members 94, 94 resist the thrust force acting through the power rollers 35, 35. Then it will move smoothly.

Next, another embodiment of the toroidal type continuously variable transmission will be described.

That is, as shown in FIG. 6, also in this embodiment, the pair of first and second continuously variable transmission mechanisms 131 and 132 constituting the toroidal continuously variable transmission 130 are provided on the output shaft 104. An input disk 133 that is rotatably provided with respect to 104, and an output disk 134 that is arranged so as to face each input disk 133 and rotates integrally with the output shaft 104.
And a pair of power rollers 13 disposed between the output disk 134 and the input disk 133 so as to be rotatable and tiltable while being in contact with the disks 133 and 134, respectively.
5,135. The input disks 133, 1 in the first and second continuously variable transmission mechanisms 131, 132
An intermediate disk 136, which is rotatable relative to the input disks 133 and 133, is arranged between 33, and the intermediate disk 136 and each of the input disks 133, 133.
133 and a plurality of loading cams 137 ... 137
Will be installed.

In this case, in this embodiment, the intermediate disk 136 is composed of the first and second members 136a and 136b on both sides which are separably arranged, and the intermediate member 136c interposed therebetween. At the same time, as shown by the hatching in the figure, the intermediate member 136c is made of a material having a linear expansion coefficient larger than that of the other constituent elements. In this case, as shown by the solid line in FIG. 7, the pressing force of the loading cams 137 ... 137 at room temperature is larger than the required pressing force at room temperature shown by the broken line, and the high temperature shown by the chain double-dashed line due to thermal expansion. The pressing force is set so that the required pressing force can be obtained when (for example, 140 ° C.).

According to this, there is an advantage that the pressing force at normal temperature is reduced and the transmission efficiency is improved.

[0049]

As described above, according to the present invention, in a toroidal type continuously variable transmission in which a power roller is arranged between the input and output discs in a pressure contact state, the toroidal type continuously variable transmission is inserted into the roller support member and is arranged in the rotational axis direction of the output disc. A misalignment is formed because the race bearing of the thrust bearing that rotatably supports the power roller is slidably supported along the rotational axis direction of the both disks by forming a recessed groove that is oriented. Even if the race member of the thrust bearing slides along the groove to absorb the above,
The rotation center of the power roller does not move in the direction orthogonal to the plane including the rotation center and the rotation center of the output disk. As a result, there is no variation in the tilt angle of the power roller, and good gear ratio control accuracy can be obtained.

Moreover, since the race bearing member of the thrust bearing need only be supported in the concave groove formed in the roller supporting member, the pivot shaft which is difficult to process is not required, the supporting structure of the power roller is simplified, and the manufacturing cost is reduced. Will be reduced. Further, since it is not necessary to provide a hole for supporting the pivot shaft in the roller supporting member, the roller supporting member is lightweight and compact.

In particular, according to the second aspect of the invention, the movement of the race member along the concave groove is promoted between the race bearing thrust member supporting the power roller and the bottom surface of the concave groove of the roller supporting member. Since the linear bearing is installed, the race member moves smoothly against the thrust force acting through the power roller.

[Brief description of drawings]

FIG. 1 is a skeleton diagram of an automobile power train according to the present invention.

FIG. 2 is a cross-sectional view of a first continuously variable transmission mechanism constituting the toroidal type continuously variable transmission taken along the line AA in FIG.

FIG. 3 is a sectional view of a main part of the partition wall taken along line BB in FIG.

FIG. 4 is an enlarged sectional view of an essential part showing the structure of the first trunnion and its surroundings.

5 is a view taken along the line CC in FIG.

FIG. 6 is a schematic view showing another embodiment of the toroidal transmission.

FIG. 7 is a characteristic diagram of pressing force with respect to input torque in the embodiment.

[Explanation of symbols]

 33 input disc 34 output disc 35 power roller 38 first trunnion 39 second trunnion 91 concave groove 93 thrust bearing 94 race member 96 linear bearing

Claims (2)

[Claims] 1. A toroidal type having an input disk and an output disk coaxially opposed to each other, a power roller arranged in a pressure contact state between both disks, and a roller support member rotatably supporting the power roller. A roller supporting structure for a continuously variable transmission, wherein the roller supporting member is formed with a concave groove oriented in the rotational axis direction of the input / output disk, and a race bearing thrust member supporting the power roller is provided. A roller support structure for a toroidal type continuously variable transmission, wherein the roller support member is supported in a groove of the roller support member so as to be slidable along the rotational axis directions of the both disks. 2. A linear bearing for promoting the movement of the race member along the concave groove between the race bearing thrust member rotatably supporting the power roller and the bottom surface of the concave groove of the roller supporting member. The roller support structure of the toroidal type continuously variable transmission according to claim 1, wherein the roller support structure is provided.