JPH08145138A - Troidal type continuously variable transmission - Google Patents

Abstract

PURPOSE: To provide a troidal type continuously variable transmission with which the durability of a power roller, etc., is enhanced and the power loss is reduced. CONSTITUTION: A troidal type continuously variable transmission is composed of an input disc 33 and output disc 34 installed opposingly on the output shaft 4, a power roller 35 installed between the two discs in the pressure contacted condition with possibility of tilting, and a trunnion which supports the power roller rotatably. The rotational center C of the power roller 35 is located offset toward the input disc 33 with respect to the center of tilting in the initial set condition with a gearing ratio of one and with no load applied.

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 is known as a transmission mounted on an automobile or the like. This toroidal type continuously variable transmission includes an input disc to which an engine output is input, an output disc coaxially arranged to face the input disc, as disclosed in Japanese Patent Laid-Open No. 193454/1989. A single or a plurality of toroidal transmission units having a power roller tiltably arranged in a state of being pressed against both the output disks and the input disk are provided, and by tilting the power rollers, The contact position of the power roller with respect to the disk is changed to change the speed ratio steplessly.In this type of toroidal type continuously variable transmission, for example, a rotational force is converted to a thrust force by a loading cam. By converting and transmitting to the input disc side, the contact between the input and output discs and the power roller Had a major pressing force between the faces with a high surface pressure state, thereby transmitting power by traction oil that transition to a glassy state.

[0003]

By the way, in this type of toroidal continuously variable transmission, the power roller is referred to as a trunnion in order to support it against a torque reaction force and a pressing force reaction force acting from the disk side. In order to rotatably support the power roller using the support member and to tilt the power roller, the trunnion is inserted into the rotation axis of the power roller, and the axis perpendicular to the plane including the rotation axis of the output disk ( Hereinafter, it will be rotatably supported as a trunnion center). In that case, conventionally, the initial set state (no load, reduction ratio =
Since the relationship between the two is set so that the center of rotation of the power roller in 1) (hereinafter referred to as the power roller center) passes through the trunnion center, the following inconvenience occurs especially in the deceleration range where the gear ratio is large. It has been found.

That is, for example, as shown in FIG. 11, the input disk 1 constituting the toroidal transmission unit 101.
If the pressing force of Fa is applied to 02 via the loading cam 103, the power roller 104
With respect to, the pressing force Fb acts via the contact ellipse Si with the input disk 102, and the pressing force Fc acts via the contact ellipse So with the output disk 105. Then, the resultant force Fs of these pressing forces Fb and Fc acts on the trunnion 106 that supports the power roller 104. In that case, when the transmission torque is large,
Input / output disks 102, 105 to power roller 10
Since the pressing forces Fb and Fc acting on 4 also become large,
The output disc 105 is actually deformed as shown by the chain double-dashed line and supports the power roller 104.
06 will also be deformed in the direction of the resultant force Fs. Therefore, the power roller center C is deviated from the state of the one-dot chain line passing through the trunnion center O to the output disc 105 side as shown by the two-dot chain line, and the input disc 102 and the output disc 105 and the power roller 104.
The contact ellipses Si and So between and move from the ideal gear shift state indicated by the one-dot chain line to the state indicated by the two-dot chain line.

That is, as shown in FIG. 12, for example, the input side contact point Pi between the input disk 102 and the power roller 104 is geometrically viewed as the center of curvature Ci of the input disk 102 and the curvature of the power roller 4. It exists on the line connecting the center Cp.

In this case, when the output disc 105 is deformed as shown by the chain double-dashed line, the input disc 102 is moved in the axial direction as shown by the chain double-dashed line by the pressing force Fa acting from the loading cam 103. Along with this, the center of curvature Ci of the input disk 102 also moves in the axial direction, as indicated by the symbol (a). On the other hand, in order to equalize the pressing force between input and output, the power roller 104 is
As indicated by reference numeral (a), since it is movably supported in the direction perpendicular to the power roller center C,
The center of curvature Cp of the power roller 4 after the deformation moves in a direction in which the above moving direction (a) and the deformation direction of the trunnion 106 indicated by the symbol (c) are combined.

Therefore, the contact point Pi on the input side between the input disk 102 and the power roller 104 is defined by the center of curvature Ci of the input disk 102 after deformation and the center of curvature Cp of the power roller 104 as shown by the chain double-dashed line. They will move to the connecting line.

For this reason, the edge portion 104a of the power roller 104 rides on the input disk 102, for example, to lower the durability, and the power roller center C deviates from the trunnion center O as described above. The tilting moment acts and causes power loss.

This is supported by the geometrical analysis results shown in FIG.

That is, a predetermined reduction ratio (for example, 2.4)
When the maximum load is applied in this state, the power roller center C is closer to the output disk side by the distance x from the power roller center Co in the ideal gear shift state shown by the broken line.
It is assumed that only parallel translation is performed. In this case, the input side contact point Pi between the input disc and the power roller moves clockwise with respect to the input side contact point Pi ′ in the ideal speed change state, and the output between the output disc and the power roller is increased. The side contact point Po moves counterclockwise with respect to the output side contact point Po ′ in the ideal gear shift state. In this case, since the input disc is moved a large amount, the input side contact point Pi is largely moved, and the edge boundary line y of the power roller is attached to the input side contact ellipse Si centered on the contact point Pi. Will enter.

If the vertical component force of the pressing force Fb on the input disk side along the power roller center direction is Fi and the vertical component force of the pressing force Fc on the output disk side along the power roller center direction is Fo, Since the displacement amount of the power roller center C is x, the tilting moment M acting on the power roller is expressed by the following relational expression (1).

M = (Fi + Fo) · x (1) In that case, the perpendicular drawn from the input side contact point Pi to the power roller center C is Hi, and the perpendicular drawn from the output side contact point Po to the power roller center C is If Ho, the foot of the perpendicular line Hi comes closer to the trunnion center side set to the origin 0 than the foot of the perpendicular line Ho.
If b and Fc are substantially equal to each other, the vertical component force Fi of the pressing force Fb on the input disk side becomes smaller than the vertical component force Fo of the pressing force Fc on the output disk side. Therefore, as indicated by the symbol (d), a counterclockwise tilting moment M centered on the origin 0 (the trunnion center O) is generated.

The amount of movement and the tilting moment of these contact points increase as the amount of deformation of the output disk or trunnion increases.

According to the present invention, there are provided an input disk and an output disk which are coaxially opposed to each other, a power roller tiltably arranged between the disks in a pressure contact state, and a trunnion which rotatably supports the power roller. The toroidal type continuously variable transmission has the above-mentioned problems caused by the deformation of the output disc and the trunnion, and prevents the edge portion of the power roller from riding on the disc to improve the durability of the power roller and the like. And let
An object of the present invention is to reduce power loss caused by a large tilting moment acting on the power roller.

[0015]

That is, the invention according to the present application is directed to an input disk and an output disk that are coaxially opposed to each other, a power roller that is tiltably arranged between the disks in a pressure contact state, and a power roller. In a toroidal type continuously variable transmission having a trunnion that rotatably supports a roller, the rotation center of the power roller is set at a gear ratio of 1 and no load is applied to the input disk side with respect to the tilt center thereof. It is characterized in that it is arranged in an offset state.

[0016]

According to the above construction, the following operation can be obtained.

Even if the center of rotation of the power roller moves to the side of the output disc due to deformation of the output disc or trunnion in the high load deceleration range, the center of rotation of the power roller with respect to the tilt center in the initial set state. Since it is offset to the input disc side, the actual deviation amount of the rotation center of the power roller with respect to the tilting center is reduced, and thus, for example, the contact point between the input disc and the power roller is moved to the edge side of the power roller. Is suppressed, and the edge of the power roller is prevented from riding on the input disc.
The durability of the power roller will be improved. Further, since the deviation amount of the rotation center of the power roller with respect to the tilt center is reduced, the tilting moment acting on the power roller is reduced, and the power loss is also reduced.

[0018]

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

As shown in FIGS. 1 and 2, 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 an output of the torque converter 3 transmitted thereto. Reducer 10
And the toroidal type continuously variable transmission 30 to which the rotation of the engine 2 is input, and the output of the speed reducer 10 or the toroidal type continuously variable transmission 30 or both of them is output via the output shaft 4. It is adapted to be transmitted to a drive wheel (not shown).

The torque converter 3 has a pump 3b integrated with a casing 3a connected to the engine output shaft 2a.
And a turbine 3c arranged to face the pump 3b,
A turbine shaft 3e, which has a stator 3d interposed between the turbine 3c and the pump 3b, rotates integrally with the turbine 3c, and a one-way clutch 3f externally fitted to the turbine shaft 3e and at one end. The first hollow shaft 3g integrated with the transmission casing 60 by connecting the stator 3d via the speed reducer 10
It is connected to. Further, the first hollow shaft 3g
An oil pump 5 is provided at the shaft end of the second hollow shaft 3h, which is externally fitted to the casing 3a and has one end connected to the casing 3a.
The engine 2 is driven via the.

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.
The first planetary gear mechanism 11 arranged on the side is for backward movement,
Further, the second planetary gear mechanism 12 is for forward movement. The first planetary gear mechanism 11 is of a single pinion type and is connected to the turbine shaft 3e.
3, the carrier 15 that rotatably supports the pinion 14 that meshes with the sun gear 13 is a transmission casing 6
The ring gear 1 is connected to the first hollow shaft 3g fixed to 0 and meshes with the pinion 14.
6 through the reverse clutch 17 through the turbine shaft 3
It is connected to the output shaft 4 arranged on the same axis as e.

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. Further, 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 first hollow shaft 3g. There is. Further, the ring gear 21 that constitutes the second planetary gear mechanism 12 is
It 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 rotationally driven in the forward direction.

On the other hand, the toroidal type continuously variable transmission 30
Has a pair of first and second toroidal transmission units 31, 32 arranged in series on the output shaft 4. These toroidal transmission units 31 and 32 have the same structure, and an input disc 33 is provided on the output shaft 4 so as to be rotatable with respect to the shaft 4, and an output is provided so as to face each input disc 33. Output disk 3 that rotates together with shaft 4
4 and a pair of power rollers 35, 35 disposed between the output disk 34 and the input disk 33 so as to be rotatable and tiltable while being in contact with both disks 33, 34, respectively.
Have and.

An intermediate disk 36, which is rotatable relative to the input disks 33, is arranged between the input disks 33, 33 of the first and second toroidal transmission units 31, 32. A plurality of loading cams 37 ... 37 are interposed between the intermediate disc 36 and the input discs 33, 33, respectively. The loading cams 37 ... 37 are configured such that the larger the input torque input 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 at an end portion of the input shaft 51 located on the torque converter 3 side, the first gear 52 is meshed with an intermediate gear 53, and the intermediate gear 53 is
The second hollow shaft 3 via the power distribution clutch 54.
It is meshed with the output gear 55 connected to h. Also,
At the other end of the input shaft 51, the intermediate disk 3
A second gear 57 that meshes with the input gear 56 that is integrally provided with the gear 6 is integrally provided. Therefore, when the power distribution clutch 54 is engaged, the output of the engine 2 is transmitted via the input gear 56 to the toroidal type continuously variable transmission 3
1st and 2nd toroidal transmission units 31 and 3 at 0
2 is input to each input disk 33, 33, and the rotation of each input disk 33, 33 is changed at a predetermined speed change ratio (reduction ratio) according to the tilt angle of each power roller 35. It is adapted to be transmitted to 34, 34.

Next, the structure of the first and second toroidal transmission units 31 and 32 constituting the toroidal type continuously variable transmission 30 will be described more specifically. In addition, the first toroidal transmission unit 31 and the second toroidal transmission unit 32.
Since has the same configuration as described above, the first toroidal transmission unit 31 will be described as a representative.

That is, the first toroidal transmission unit 3
As shown in FIG. 3, a first and a second trunnions 38, 39 arranged in the up-down direction are provided on the power source 1, and the power roller 35 is attached to these trunnions 38, 39 via eccentric shafts 40, 40. , 35 are rotatably supported. The support member 62 attached to the transmission casing 60 via the link post 61 includes the first and second support members 62.
The upper ends of the trunnions 38 and 39 are spherical bearings 6 respectively.
The first and second trunnions 38 are attached to a support member 66 that is rotatably supported via the shafts 3, 63 and is attached to a partition wall 64 integral with the transmission casing 60 via a link post 65. , 39 are rotatably supported by spherical bearings 67, 67, respectively. The first and second trunnions 38, 39 have the trunnion shaft 4 extended in the direction orthogonal to the output shaft 4.
1a and 41b are integrally attached to each other.
The tip ends of these trunnion shafts 41a and 41b respectively penetrate the partition wall 64 and pass through the oil pan 68.
It projects into the lower space covered with.

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 speed change pistons 73a and 73b, which are loosely fitted to the trunnion shaft 41a, are inserted in the a and 71b, respectively. A thrust bearing 42 is interposed 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 43 is provided adjacent to the lower end of the trunnion shaft 41a. Then, at the lower end portion of the trunnion shaft 41a, a recess cam 81 constituting the shift control mechanism 80 is spline-fitted adjacent to the thrust bearing 43 and is in contact with the lower surface of the boss portion. The circlip 44 mounted on the trunnion shaft 41a supports the recess cam 81 and the shift piston 73b.

On the other hand, the upper and lower hydraulic chambers 72a and 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 41b, are inserted. Even in this case, the upper hydraulic chamber 72
A thrust bearing 42 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 41b is provided on the lower surface of the speed change piston 74b inserted in the lower hydraulic chamber 72b.
3 are arranged adjacent to each other, and by mounting the circlip 44 on the trunnion shaft 41b in a state of being in contact with the lower end of the collar 45 arranged adjacent to the thrust bearing 43, the collar 45 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 41a is pushed down by the speed change piston 73b inserted in the lower hydraulic chamber 71b. Therefore, the first trunnion 38 slides downward accordingly.

Next, the shift control mechanism 80 for changing the gear ratio by controlling the supply and discharge of operating pressure to and from the hydraulic chambers of the first and second hydraulic cylinders 71 and 72 in the first toroidal transmission unit 31. The configuration 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 arranged in contact with a cam surface of a recess cam 81 fixed to the lower end portion of the trunnion shaft 41a 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.

According to this structure, 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 base end side of the spool 85 are exerted. The pressing force of the spring 89 balances with the pressure of the first and second hydraulic cylinders 7.
It is configured such that the supply and discharge of the operating pressure to and from the motors 1 and 72 are stopped.

If the stepping motor 86 is driven to move the sleeve 84 to the right (direction a) in the drawing of FIG. 3, 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 toroidal transmission unit 31
The first trunnion 38 in the above slides downward, and the second trunnion 39 slides upward,
When the contact positions of the power rollers 35, 35 attached to the first and second trunnions 38, 39 change, a tilting force is generated. In this case, for example, if the input disk 33 is rotating in the counterclockwise b direction in FIG. 3, the power roller 35 on the first trunnion 38 side is rotated in the c direction in FIG. The power roller 35 on the trunnion 39 side rotates in the d direction. As a result, the gear ratio (reduction ratio) between the input and output disks 33, 34 in the first toroidal transmission unit 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 41a, and the displacement is transferred to the L-shaped link 88 via the precess cam 81. When transmitted, the link 88 is moved in the clockwise direction (e
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 86 is realized.

In this case, if the power roller center is biased toward the output disk 34 side with respect to the trunnion center, a tilting moment acts on the power roller in the direction of increasing the gear ratio, and the tilting of the power roller is caused. The gear shift control mechanism 80 is actuated to the speed-up side to cancel the moment, and power is consumed unnecessarily.

Therefore, in this embodiment, the first,
As shown in the enlarged view of FIG. 4, for example, the power roller 35 forming the second toroidal transmission units 31 and 32 is set such that the power roller center C with respect to the trunnion center O in the initial set state with a gear ratio of 1 and no load. Are arranged on the input disc side with a predetermined offset amount z.

In this embodiment, as shown in FIG. 2, the intermediate disk 36 is composed of the first and second members 36a and 36b which are separably arranged, and both of these members 36a and 36a. A dish plate 48 is provided between 36b.

Next, for example, the offset amount z is set to-
The operation of the embodiment will be described by taking the case of setting 0.5 mm as an example. Here, the code defines the output disk side as (+) and the input disk side as (-) with respect to the trunnion center O.

That is, in the initial set state in which the gear ratio is 1 and no load is applied, as shown in FIG. 5, the input side and output side contact points Pi and Po do not move. When the power roller is tilted to the deceleration side (reduction ratio = 2.4) from this state, the contact points Pi and Po move in the direction opposite to the position where the edge exists, as shown in FIG.

When the maximum load is applied while the power roller is tilted in this way, the power roller center C is displaced toward the output disk as shown in FIG. 7 due to the deformation of the output disk, the trunnion, etc. Accordingly, the input-side and output-side contact points Pi, Po move to the edge side. In that case, the absolute amount of movement of the power roller center C and the input-side and output-side contact points Pi, Po with respect to the amount of deformation of the output disk or trunnion is the same as the conventional one, but the power roller center C is previously Since it is offset to the input disk side, the deviation amount x with respect to the trunnion center O becomes smaller than that in the state where there is no offset. As a result, the movement of the input-side and output-side contact points Pi and Po to the edge side is suppressed, and for example, the edge boundary line y of the power roller is attached to the contact ellipse Si centered on the input-side contact point Pi. Does not enter.

Further, since the deviation amount x of the power roller center C with respect to the trunnion center O is smaller than that in the state where there is no offset, the tilting moment M acting on the power roller is also reduced.

As a comparative example, the offset amount z is, for example, +
The case of setting to 0.5 mm will be described.

That is, with respect to the initial set state,
When the power roller is tilted to the deceleration side (reduction ratio = 2.4), as shown in FIG. 8, the input side and output side contact points Pi,
Po moves toward the edge.

When a maximum load is applied in this state, the power roller center C is displaced in the output disc direction as shown in FIG. 9, and accordingly the input side and output side contact points Pi, Po are further moved to the edge side. As a result, for example, the edge boundary line y of the power roller enters the contact ellipse Si centered on the input side contact point Pi, and the tilting moment M of the power roller also increases.

FIG. 10 is a characteristic diagram showing the relationship between the offset amount and the movement amount of the input disk in the unloaded state.

As is clear from this characteristic diagram, the larger the offset amount z is, the larger the moving amount of the input disc in the initial set state is. Therefore, it is necessary to increase the stroke amount absorbed by the dish plate 48. Further, as shown in Table 1 below, when the offset amount z is set to -1.0 mm, the tilting moment at the maximum load is greatly reduced, but in the reverse direction at the partial load in the normal range. The tilting moment of acts greatly. Therefore, the offset amount z is -0.1
About -0.5 mm is desirable.

[0050]

[Table 1]

[0051]

As described above, according to the present invention, even when the rotation center of the power roller moves to the output disk side due to the deformation of the output disc or the trunnion in the high load deceleration region, the rotation center of the power roller is However, in the initial set state, since it is offset to the input disk side with respect to the tilt center, the actual deviation amount of the rotation center of the power roller with respect to the tilt center is reduced, and thus, for example, the input disk and the power roller are The movement of the contact point between the edges to the edge side of the power roller is suppressed,
Since the edge portion of the power roller is prevented from riding on the input disc, the durability of the power roller and the like is improved. Further, since the deviation amount of the rotation center of the power roller with respect to the tilt center is reduced, the tilting moment acting on the power roller is reduced, and the power loss is also reduced.

[Brief description of drawings]

FIG. 1 is a skeleton diagram showing a power train of an automobile according to an embodiment.

FIG. 2 is a sectional view of a toroidal type continuously variable transmission.

FIG. 3 is a cross-sectional view of the first toroidal transmission unit taken along the line AA in FIG.

FIG. 4 is an enlarged sectional view of a main part of the toroidal transmission unit.

FIG. 5 is a schematic diagram showing a geometric analysis result in an initial set state when the offset amount is set to −0.5 mm.

FIG. 6 is a schematic diagram showing a geometric analysis result in a no-load state when the power roller is tilted to the deceleration side.

FIG. 7 is a schematic diagram showing a geometric analysis result in a maximum load state when the power roller is tilted to the deceleration side.

FIG. 8 is a schematic diagram showing a geometric analysis result in an unloaded state when the power roller is tilted to the deceleration side when the offset amount is set to +0.5 mm.

FIG. 9 is a schematic diagram showing a geometric analysis result in a maximum load state when the power roller is also tilted to the deceleration side.

FIG. 10 is a characteristic diagram showing a relationship between an offset amount and a movement amount of an input disk in a no-load state.

FIG. 11 is a schematic view of a toroidal transmission unit showing a conventional problem.

FIG. 12 is a schematic view showing a mechanism of movement of a contact point.

FIG. 13 is a schematic diagram showing a geometrical analysis result in a maximum load state when the conventional power roller is tilted to the deceleration side.

[Explanation of symbols]

 30 Toroidal type continuously variable transmission 33 Input disk 34 Output disk 35 Power roller 38,39 Trunnion O Trunnion center C Power roller center

Claims (1)

[Claims] 1. A toroidal device having an input disk and an output disk that are coaxially opposed to each other, a power roller that is tiltably arranged in a pressure contact state between both disks, and a trunnion that rotatably supports the power roller. In the continuously variable transmission, the center of rotation of the power roller is arranged offset from the tilt center toward the input disk in the initial set state with a gear ratio of 1 and no load. A toroidal type continuously variable transmission characterized in that