JPH06294462A - Toroidal type continuously variable transmission - Google Patents

Abstract

PURPOSE:To provide a toroidal type continuously variable transmission provided with a clutch which can control the oil pressure of a start clutch to a reasonable value without requiring a complicate device and difficult adjustment. CONSTITUTION:This transmission is provided with power rollers 45a-45d for transmitting the torque of input discs 43f, 43r to output discs 44f, 44r, hydraulic mechanisms 79a-79d, 80a-80d for moving power rollers 45a-45d in a direction nearly orthogonally acrossing the rotating shaft of the output discs 44f, 44r in order to slantly roll the power rollers 45a-45d and a start clutch 150 connected to the output discs 44f, 44r side. In a toroidal type continuously variable transmission C by which the hydraulic mechanisms 79a-79d, 80a-80d can generate a drive power for moving the power rollers by using the pressure difference PH-PL between two hydraulic chambers, the control pressure of a control valve 160 for controlling the clutch pressure of the start clutch 150 is adjusted by using the pressure difference PH-PL between the two hydraulic chambers.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a toroidal type in which a clutch is provided on the output side and the clutch pressure of this clutch is controlled by utilizing the pressure generated by a hydraulic mechanism for driving a power roller. The present invention relates to a continuously variable transmission.

[0002]

2. Description of the Related Art In recent years, compared to a belt type continuously variable transmission,
A toroidal type continuously variable transmission has been attracting attention because it has high durability and can transmit a large torque.
In a toroidal type continuously variable transmission, a power roller arranged between an input side cone disc and an output side cone disc is swung to change the gear ratio between the input disc and the output disc. Has been done.
Such a toroidal type continuously variable transmission is often combined with a torque converter utilizing hydraulic pressure, but it is conceivable to use a starting clutch instead of the torque converter in order to increase transmission efficiency.

The starting clutch is disclosed in Japanese Patent Laid-Open No. 62-25155.
It corresponds to the forward or reverse clutch of Japanese Patent Publication No. 9 and is adapted to switch between a state in which the rotational output from the engine is connected to the wheels and a state in which the rotational output is disconnected without passing through a fluid coupling. In such a starting clutch,
In order to connect the clutch smoothly and start the vehicle smoothly, the full connection must be made via the half-clutch state. For that purpose, the clutch hydraulic pressure, which is the pressing pressure of the starting clutch, needs to have a time waveform as shown in FIG. That is, the clutch hydraulic pressure is changed from the state where the hydraulic pressure is zero to the half-clutch state through the shelf pressure,
It must change to rise to the pressure required for full connection. Further, the shelf pressure in the half-clutch state needs to be set higher as the input torque is larger, as shown in FIG.

[0004]

In order to generate such a shelf pressure, conventionally, a clutch pressure control valve is used as shown in FIG. 13, and the control pressure is controlled by an electronically controlled hydraulic valve such as a linear solenoid valve. It was designed to be controlled. However, such a method has problems that the number of parts is increased, the cost is increased, and the control system is complicated. In addition, there is a problem in that the tuning becomes complicated because it is necessary to tune the rising timing of the engine torque and the rising timing of the clutch pressure when starting.

Therefore, the present invention has been made in view of the above-mentioned problems, and an object of the present invention is to set the hydraulic pressure of the starting clutch to an appropriate value without requiring a complicated device or complicated adjustment. It is an object of the present invention to provide a toroidal type continuously variable transmission including a clutch that can be controlled.

[0006]

In order to solve the above-mentioned problems and to achieve the object, a toroidal type continuously variable transmission according to the present invention is a cone-shaped input disk to which power is input, and the input disk. A cone-shaped output disc disposed coaxially and outputting power to the outside, and a cone face of the input disc and a cone of the output disc, which are disposed between the cone faces of the input disc and the output disc that face each other. A power roller for transmitting the rotational force of the input disc to the output disc by rotating while contacting both surfaces, and for changing the gear ratio of the input disc and the output disc by tilting the power roller. A hydraulic mechanism for moving the power roller in a direction substantially orthogonal to the rotation axis of the output disc, and a starting clutch connected to the output disc side. In the toroidal continuously variable transmission, wherein the hydraulic mechanism is configured to generate a driving force for moving the power roller by using a pressure difference between two hydraulic chambers provided in the hydraulic mechanism. The control pressure of the control valve for controlling the clutch pressure is adjusted using the differential pressure between the two hydraulic chambers.

[0007]

Since the toroidal type continuously variable transmission according to the present invention is constructed as described above, it operates as follows.
That is, as a reaction force when the rotational force of the input disc is transmitted to the output disc, a force for moving the power roller in the tangential direction of these discs acts on the power roller. This force is proportional to the transmitted torque. Therefore, a differential pressure is generated in the two hydraulic chambers of the hydraulic mechanism that moves and drives the power roller to generate a force that balances this reaction force, and this differential pressure is proportional to the transmission torque. Therefore, if the control pressure of the control valve that controls the clutch pressure using this differential pressure is adjusted, the hydraulic mechanism for driving the power roller that is originally necessary can be provided without using a complicated linear solenoid valve as in the past. Utilizing this, it is possible to obtain the control pressure of the clutch that is proportional to the transmission torque.

[0008]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT A preferred embodiment of the present invention will be described in detail below with reference to the accompanying drawings. FIG. 1 is a sectional view showing the structure of a toroidal type continuously variable transmission according to an embodiment. Figure 1
In the toroidal type continuously variable transmission C,
The first toroidal transmission mechanism 7 is arranged on the front side (left side in the drawing) so as to surround the transmission output shaft 5 extending in the front-rear direction of the vehicle in which the vehicle is mounted, and the second toroidal transmission mechanism is arranged on the rear side.
And a toroidal transmission mechanism 8. And the first
And the engine rotational force input from the input cam 48 disposed in the middle of the second toroidal type speed change mechanisms 7 and 8 is decelerated or accelerated through the first and second toroidal type speed change mechanisms 7 and 8. , Is output from the transmission output shaft 5. By dividing the toroidal type transmission mechanism into two in this way, the transmission can be downsized while taking a large torque capacity, as compared with the case where one toroidal type transmission mechanism receives all the transmission torque. be able to.

Here, the first and second toroidal type speed change mechanisms 7 and 8 are arranged symmetrically in the front and rear direction.
Since the configurations and functions of the two are basically the same, the corresponding members are denoted by the same reference numerals, and in principle, each member of the first toroidal type speed change mechanism 7 is provided with a subscript f and the second toroidal type. Each member of the speed change mechanism 8 is attached with a subscript r. However,
The members such as rollers, which are provided in twos in each toroidal transmission mechanism 7 and 8, cannot be distinguished only by the subscripts f and r. Therefore, the respective members disposed on the left and right of the first toroidal transmission mechanism 7 can be distinguished. To subscripts a and b, and subscripts c and d to the respective members arranged on the left and right of the second toroidal transmission mechanism 8. Therefore, in the following, the description made for one member also applies in principle to other members having the same number but different subscripts.

The first toroidal transmission mechanism 7 has a first input disk 43f loosely fitted around the transmission output shaft 5.
And the first output disc 44 fixed to the transmission output shaft 5.
f and the first and second power rollers 45 for transmitting the torque of the first input disk 43f to the first output disk 44f.
a and 45b are provided. Then, the first input disc 43f is engaged with the input cam 48 having the driven gear 42 formed on the outer periphery through the first cam roller 49f, and the larger the input torque to the first input disc 43f is, The 1-input disk 43f is strongly pressed against the first and second power rollers 45a and 45b. This pressing force is applied to the first input disk 43f and the first output disk 44f and the first and second power rollers 45.
A frictional force is generated between the first and second rollers 45a and 45b, and the rotational force of the first input disk 43f causes the first and second rollers 45a and 45b to rotate.
It is transmitted to the first output disk 44f via b.

First and second power rollers 45a and 45b
Are each rotatable about the axis Y, and their peripheral surfaces are brought into contact with the cone-shaped curved surface of the first input disk 43f and the cone-shaped curved surface of the first output disk 44f. . Therefore, the rotation torque of the first input disk 43f is transmitted to the first output disk 44f via the first and second power rollers 45a and 45b. Here first
The gear ratio (torque ratio) in the transmission of the rotational force from the input disc 43f to the first output disc 44f is equal to the radius R1 of the first input disc 43f at the position in contact with the first and second power rollers 45a and 45b. It is determined by the ratio R2 / R1 of the radius R2 of the first output disk 44f.

Then, the first and second rollers 45a, 45
The contact position between b and both discs 43f and 44f is determined by the tilt angle of the first and second rollers 45a and 45b in the plane parallel to the paper surface of the drawing, and is controlled by a hydraulic mechanism described later. By changing the tilt angle, the gear ratio can be arbitrarily set within a predetermined range. next,
The configuration of the toroidal transmission C will be described more specifically.

As shown in FIG. 1, in the first toroidal transmission mechanism 7, the first output disk 244f is spline-fitted to the transmission output shaft 5. Further, the first output disc 44f is rotatably supported by the transmission case 22 via the first bearing 47f while being positioned by the ring-shaped positioning member 46 fitted to the transmission output shaft 5. . The second output disk 44r of the second toroidal transmission mechanism 8 has a radially enlarged portion 5g integrally formed on the outer periphery of the transmission output shaft 5 and the transmission case 22.
It is positioned by a second bearing 47r that is provided between the second bearing 47r and rotatably supports the transmission output shaft 5.

Between the first output disc 44f and the second output disc 44r, there are first and second input discs 43.
The f and 43r are arranged close to each other so that their rear surfaces are opposed to each other, and the first and second cam rollers 49f and 49r are respectively provided between the f and 43r and both input disks 43f and 43r. Here, the first and second cam rollers 49f and 49r
Is the input cam 48 and the first and second input disks 4
When the 3f and 43r rotate relative to each other, it has a function of generating a pressing force for pressing the first and second input disks 43f and 43r toward the first and second output disks 44f and 44r, respectively. And the second input disk 43f,
The greater the input torque to 43r, the greater the pressing force of the first and second cam rollers 49f and 49r against the first and second input disks 43f and 43r.

First and second input disks 43f, 43r
Between the first and second input disks 43f, 43r, with the output shaft 5 of the transmission loosely fitted, and both ends in contact with the back surfaces of the first and second input disks 43f, 43r, respectively.
And the engaging member 50 that is spline-fitted is arranged. Then, the engaging member 50 and the second input disk 43
A disc spring 51 is provided between the disc and the r, and the disc spring 51 allows the first input disc 43f and the second input disc 43r to be formed.
And are preloaded in the directions away from each other. The disc spring 51 abuts against the back surface of the second input disk 43r and urges the second input disk 43r toward the second output disk 44r.
The first input disk 43f is urged toward the first output disk 44f, so that the first input disk 43f and the first output disk 4
4f, and between the second input disk 43r and the second output disk 44r, a predetermined preload is applied.

A start clutch 150 is connected to the transmission output shaft 5. Next, a hydraulic mechanism for tilting the first to fourth power rollers 45a to 45d will be described with reference to FIGS. The first toroidal transmission mechanism 7 includes the first and second power rollers 45a,
First and second trunnions 59a and 59b that rotatably support 45b are provided. The first and second trunnions 59a and 59b rotatably support the first and second power rollers 45a and 45b via the first and second eccentric shafts 60a and 60b, respectively. Further, the first and second trunnions 59a and 59b respectively have first and second shaft members 6 that extend downward (in a direction orthogonal to the transmission output shaft 5).
1a and 61b are attached.

First and second power rollers 45a, 45b
An upper connection member 62 is attached to the transmission case 22 slightly above. On the other hand, the first and second
Below the power rollers 45a and 45b, a lower connecting member 63 is attached to the partition wall portion 53 fixed to the transmission case 22. Then, the upper connecting member 6
Due to the first and second shaft holes 65a and 65b formed in 2, the upper ends of the first and second trunnions 59a and 59b are respectively connected to the first and second upper spherical bushes 64a and 64b.
It is rotatably supported via. On the other hand, due to the first and second shaft holes 67a and 67b formed in the lower connecting member 63, the lower end portions of the first and second trunnions 59a and 59b respectively have the first and second lower spherical bushes 66a and 66b.
It is rotatably supported via b.

Further, the first and second shaft members 61a, 61b
The lower part of the lower housing 56 attached to the lower surface of the upper housing 55 penetrates the opening 55g of the upper housing 55 attached to the lower surface of the partition wall 53.
The recess 56g of the first support bearing 5
It is rotatably supported via 4a and 54b. In the partition wall portion 53, the first and second trunnions 59a, 59a,
First and second hydraulic cylinders 76a and 76b are provided to operate 59b, and these first and second hydraulic cylinders 76a and 76b are respectively formed by a partition wall portion 53g forming a part of the partition wall tool 53. It is divided into top and bottom. The first and second upper pistons 77a and 77b are fitted in the upper and lower portions of the first and second hydraulic cylinders 76a and 76b, respectively, and the first and second lower pistons 78a and 78b are inserted in the lower portion.
78b is inserted. Therefore, the first and second upper pistons 77a and 77b and the partition 53g define the first and second upper hydraulic chambers 79a and 79b, respectively, while the first and second lower pistons 78a and 78b and the partition are separated from each other. Part 5
3g and the first and second lower hydraulic chambers 80a and 80a, respectively.
0b is defined.

Here, the first and second upper hydraulic chambers 79a,
When hydraulic pressure is applied to 79b, the first and second upper pistons 77a and 77b displace the first and second trunnions 59a and 59b upward, while the first and second trunnions 59a and 59b are displaced.
When the hydraulic pressure is applied to the second lower hydraulic chambers 80a and 80b, the first and second trunnions 59a and 59b are displaced downward by the first and second lower pistons 78a and 78b. There is. When the first and second trunnions 59a and 59b are vertically displaced in this way, the first and second trunnions 59a and 59b are accordingly displaced according to the displacement amount.
The power rollers 45a and 45b are tilted so that the gear ratio of the first toroidal transmission mechanism 7 is changed. Also,
Along with this, the first and second trunnions 59a and 59b rotate about their axes.

The first to fourth upper hydraulic chambers 79a to 79 are provided.
The gear ratio control device that controls the gear ratio by controlling the supply of hydraulic pressure to the d and the first to fourth lower hydraulic chambers 80a to 80d, and the specific gear shift operation by the gear ratio control device will be described later. explain in detail. Further, in order to back up the synchronization of the rotations of the first and second trunnions 59a and 59b when the hydraulic mechanism fails, the two trunnions 59a and 59b are backed up.
a, 59b (59c, 59d), the interlocking wire 57,
58 is wrapped around.

The upper connecting member 62 has a first upper positioning hole 68f formed at an intermediate portion between the first axial hole 65a and the second axial hole 65b, and an intermediate portion between the third axial hole 65c and the fourth axial hole 65d. A second upper positioning hole 68r is formed. And
The first support portion 69f integrally formed with the transmission case 22 is inserted into the first upper positioning hole 68f. In addition,
The first roller lubrication member 70f is attached to the first support portion 69f using the first attachment member 71. Further, the upper spherical bearing 75r attached to the second support portion 69r is inserted into the second upper positioning hole 69r. The second roller lubrication member 70r is provided on the second support portion 69r.
Are attached using the second attachment member 71r.
Thus, the upper coupling member 62 is fixed or positioned with respect to the transmission case 22 by the first support portion 69f and the upper spherical bearing 75r.

In the lower connecting member 63, a first lower positioning hole 72f is formed in an intermediate portion between the first shaft hole 67a and the second shaft hole 67b, and an intermediate portion between the third shaft hole 67c and the fourth shaft hole 67d. A second lower positioning hole 72r is formed in the portion. And
In the first and second lower positioning holes 72f and 72r, respectively,
The first and second mounting bolts 74 are provided on the upper surface of the partition wall 53.
The first and second lower spherical bearings 73f and 73r fixed by using f and 74r are inserted. In this way, the lower connecting member 63 includes the first and second lower spherical bearings 73.
The f and 73r fix or position the partition wall portion 53 (on the transmission case side).

Hereinafter, the first to fourth upper hydraulic chambers 79a to 79 will be described.
A gear ratio control device that controls the speed ratio by controlling the supply of hydraulic pressure to the d and the first to fourth lower hydraulic chambers 80a to 80d will be described. In this gear ratio control device, the first to fourth upper hydraulic chambers 79a to 79d and the first to fourth lower hydraulic chambers 80a to 80d will be described later from the gear ratio control valve V according to the operating state. The hydraulic pressure is supplied through the hydraulic circuit. This gear ratio control valve V is
As will be described in detail later, a sleeve 83 is fitted in the valve body 82, and the spool 8 is inserted in the sleeve 83.
4 is a so-called three-layer valve fitted with a spring 85, a rotating member 86, a pin member 87 and the like as an operating mechanism, and is driven by a stepping motor 88 that operates according to a signal from a control unit (not shown). The hydraulic pressure (line pressure) that is controlled and received in the source pressure receiving port P1 according to the operating state is passed through the shift-up control port P2 or the shift-down control port P3 to the predetermined hydraulic chambers 79a to 79d, 81a. Hydraulic pressure is supplied to ~ 80d. The gear ratio control valve V is provided with feedback means 90.

As shown in FIGS. 3 to 5, in the gear ratio control valve V, a portion of the lower housing 56 is part of the valve body 8.
2, and a sleeve 83 which can reciprocate in the valve body axial direction in the valve body 82.
Further, the spool 84 is fitted in the sleeve 83 so as to be reciprocally movable in the axial direction of the valve body. Here, the sleeve 83 has a main port 83i which is always in communication with the source pressure receiving port P1 and a first port 83j which is always in communication with the upshift control port P2.
And a second port that always communicates with the downshift control port P3.
A port 83k is formed. Also, the spool 84
Includes an annular groove 84i that is always in communication with the main port 83i, and a first groove that is adjacent to the left and right of the groove 84i.
And second lands 84j and 84k are formed. Here, the gear shifting operation (when upshift or downshift is not performed, the first and second land portions 84j, 84)
k is adapted to close the inner openings of the first and second ports 83j and 83k, respectively (FIG. 5 shows this state).

The stepping motor 8 is attached to the side wall of the oil pan 36 attached to the lower end of the transmission case 22.
8, a rotary member 86 is fixed and connected to a rotary shaft 88i of the stepping motor 88. A male screw portion 86i is formed near the tip portion of the rotating member 86, and a female screw collar 92 is screwed onto the male screw portion 86i. A pin member 87 is fixed to the female threaded collar 92, and both ends of the pin member 87 are
Locked by the pair of upper and lower groove portions 82i formed in the valve body 82, the female threaded collar 92 moves in the valve body axial direction. Further, the pin member 87 causes the collar 92 with the female screw and the sleeve 83 to interlock in the axial direction of the valve body.

Inside the sleeve 83, between the female threaded collar 92 and the spool 84, a spring 85 is provided which applies a biasing force to the two in a direction to separate them from each other in the axial direction of the valve body. That is, the spring 85 constantly urges the female threaded collar 92 in the stepping motor direction (leftward in FIG. 5). In the following description, this direction (leftward in FIG. 5) is simply referred to as “left” for convenience, unless otherwise specified.
The opposite direction (rightward in FIG. 5) will be simply referred to as “right”. Therefore, the sleeve 83 interlocking with the female screw collar 92 is always biased to the left. In addition,
Of course, the spring 85 always urges the spool 84 to the right.

When the rotary member 86 rotates with the rotation of the stepping motor 88, the female screw collar 92, the rotation of which is restricted by the pin member 87, is moved in the axial direction of the valve body, and along with this, the sleeve. 83 moves in the axial direction of the valve body, and the main port 83i
To communicate with the first port 83j or the second port 83k via the groove 84i, and the hydraulic oil (hydraulic pressure) in the source pressure receiving port P1 is output to the shift-up control port P2 or the shift-down control port P3. It has become like. Note that the hydraulic oil here is hydraulic oil accompanied by a predetermined hydraulic pressure, and the same applies below.

Specifically, for example, at the time of upshifting, the stepping motor 88 rotates forward at a rotation angle corresponding to the pulse, and the sleeve 83 moves rightward accordingly, and at this time, the main port 83i moves the groove 84i. The hydraulic oil of the source pressure receiving port P1 is communicated with the first port 83j via the up port control port P2. In this case, the hydraulic oil at the downshift control port P3 is released to the drain passage 99.

On the other hand, at the time of downshifting, the stepping motor 88 reversely rotates at a rotation angle corresponding to the pulse, and the sleeve 83 moves leftward accordingly, and the main port 8
3i communicates with the second port 83k through the groove 84i, and the working oil (hydraulic pressure) of the source pressure receiving port P1 is output from the downshift control port P3. In this case, the hydraulic oil of the shift-up control port P2 is released to the drain passage 99.

At the time of downshifting, since the sleeve 83 is constantly biased to the left by the spring 85, the torque load of the stepping motor 88 is reduced by this biasing force, and the drive limit speed becomes high (step out). Is less likely to occur) and the sleeve 83 quickly moves to the left, so that the shift responsiveness is improved. At the time of shift-up, the urging force of the spring 85 increases the torque load of the stepping motor 88, so that the drive limit speed of the stepping motor 88 is lowered and the gear shift response is slightly lowered. ,
Since no responsiveness is required, no trouble occurs.

The right end portion of the spool 84 and the first shaft member 61a
A feedback means 90 is provided between the lower end of the and. The feedback means 90 is provided with a recess cam 100 that is fixed to the first shaft member 61a and rotates integrally with the first shaft member 61a. The recess cam 100 has an inclined surface 100i. In addition, the rotary shaft 1 provided at a predetermined position of the upper housing 55
A first arm 102i and a second arm 102j are attached to 01. Here, the tip portion of the first arm 102i engages with the inclined surface 100i of the recess cam 100,
The tip of the second arm 102j is engaged with the slit 84m formed at the right end of the spool 84.

The basic function of the feedback means 90 is the same as that of a commonly used feedback means, so a detailed description thereof will be omitted, but in general, the following process will be used to process each of the rollers 45a to 45d. The tilt angle (gear ratio) is maintained at the target tilt angle (target gear ratio). That is, at the time of gear shifting, when the stepping motor 88 rotates by an angle corresponding to the target tilt angle (target gear ratio), the sleeve 83 moves in the valve body axial direction by an amount corresponding to this rotation angle, and a predetermined hydraulic pressure is applied. Chamber 79
Hydraulic pressure is supplied to a to 79d and 80a to 80d, and the rollers 45a to 45d tilt to the target tilt angles. On the other hand, when the rollers 45a to 45d are tilted in this way, correspondingly, the trunnions 59a to 59d and the shaft members 61a to 61d are correspondingly rotated.
61d rotates, and accordingly, the recess cam 100 rotates. At this time, the rotating shaft 101 is rotated by the precess cam 100 through the first arm 102i, and the spool 84 is moved by the rotating shaft 101 through the second arm 102j in the same direction as the moving direction of the sleeve 83. When the tilt angles of the rollers 45a to 45d reach the target tilt angle, the movement amount of the spool 84 becomes equal to the movement amount of the sleeve 83, and the main port 83i and the first port 83j or Communication with the second port 83k is cut off, and the hydraulic chambers 79a to 79d, 8
The supply of hydraulic pressure to 0a to 80d is stopped, the change of the tilt angle is stopped, and the tilt angle is maintained at the target tilt angle.

However, when such a feedback operation is performed, the tip of the second arm 102j is moved to the spool 8
When the second arm 102j is displaced in the direction away from No. 4, that is, to the right, the spring 84 causes the spool 84 to follow the displacement of the second arm 102j. Hereinafter, hydraulic oil (hydraulic pressure) is supplied from the hydraulic pressure supply source to the gear ratio control valve V, and the hydraulic pressure chambers 79a to 79d are supplied from the gear ratio control valve V.
A hydraulic circuit that supplies hydraulic oil to 80a to 80d will be described.

In this hydraulic circuit, as shown in FIG. 6, after the hydraulic oil in the oil pan 36 is discharged from the oil pump 17, the line pressure controller 122 adjusts the hydraulic oil to a predetermined pressure (original pressure). After that, it is supplied to the source pressure receiving port P1 of the gear ratio control valve V via the source pressure supply passage 123. The hydraulic oil output from the shift-up control port P2 flows into the common shift-up oil passage 130, and thereafter, the first to fourth branch shift-up oil passages 130a.
~ 130d through the first lower hydraulic chamber 80a,
The second upper hydraulic chamber 79b, the third lower hydraulic chamber 80c, and the fourth upper hydraulic chamber 79d are supplied. On the other hand, the hydraulic oil output from the downshift control port P3 flows into the common downshift oil passage 131, and thereafter, through the first to fourth branched downshift oil passages 131a to 131d, the first upper side, respectively. Hydraulic chamber 79a, second lower hydraulic chamber 80b, third upper hydraulic chamber 79c, and fourth lower hydraulic chamber 8
0d and so on.

In such a gear ratio control device, at the time of downshifting, the sleeve 83 moves to the right in FIG. 6, the hydraulic oil of the source pressure receiving port P1 is output from the downshift control port P3, and this hydraulic oil is It is supplied to the first upper hydraulic chamber 79a, the second lower hydraulic chamber 80b, the third upper hydraulic chamber 79c, and the fourth lower hydraulic chamber 80d. It should be noted that the gear ratio control valve V in FIG. 6 is shown to have a left-right positional relationship opposite to that of the gear ratio control valve V in FIGS. 3 and 5.

At this time, the first and third trunnions 59
a and 59c are displaced upward, the second and fourth trunnions 59b and 59d are displaced downward, and the trunnion 5
By the displacement of 9a to 59d, the first to fourth rollers 45a
~ 45d changes to the deceleration side. At this time each trunnion 5
9a to 59d, the shaft members 61a to 61d rotate, but with the rotation of the first shaft member 61a, the recess cam 100 causes the first arm 102i in contact with the inclined surface 100i and the rotating shaft 101. Then, the spool 84 is moved to the right through the second arm 102j until the communication between the source pressure receiving port P1 and the downshift control port P3 is blocked. In this way, the tilt angle is maintained at the target tilt angle.

Here, when the sleeve 83 moves rightward, the spring 83 biases the sleeve 83 rightward, so that the torque load of the stepping motor 88 is reduced and the shift responsiveness is improved. As described above. On the other hand, when shifting up, the sleeve 83 moves leftward in FIG. 6, and the source pressure receiving port P
1 hydraulic oil is output from the shift-up control port P2, and this hydraulic oil is used as a first lower hydraulic chamber 80a, a second upper hydraulic chamber 79b, a third lower hydraulic chamber 80c, and a fourth upper hydraulic chamber. It is supplied to the chamber 79d.

At this time, the first and third trunnions 59
a and 59c are displaced downward, the second and fourth trunnions 59b and 59d are displaced upward, and the trunnion 5
By the displacement of 9a to 59d, the first to fourth rollers 45a
~ 45d changes to the speed increasing side (overdrive side).
At this time, each trunnion 59a to 59d (each shaft member 61a
~ 61d) is rotated, and the first cam 102a in contact with the inclined surface 100i of the precess cam 100, the rotating shaft 101, and the second arm 102j are operated in accordance with the rotation of the first shaft member 61a. However, in this case, unlike the case of the above-mentioned shift down, the tip end portion of the second arm 102j is displaced in the direction away from the spool 84 (leftward in FIG. 6). Therefore, the spool 84 is moved leftward in FIG. 6 by the urging force of the spring 85, and the source pressure receiving port P1 and the shift-up control port P
It can be moved to the left until the communication of 2 is cut off. In this way, the tilt angle is maintained at the target tilt angle. Next, the configuration of the shelf pressure control hydraulic system for controlling the shelf pressure of the hydraulic clutch 150, which is a characteristic part of this embodiment, will be described.

As already described in the section of the prior art, in the starting clutch, in order to smoothly connect the clutch and start the vehicle smoothly, the complete connection must be made via the half-clutch state. I won't. For that purpose, the clutch hydraulic pressure, which is the pressing pressure of the starting clutch, needs to have a time waveform as shown in FIG. That is,
The clutch hydraulic pressure must change so that the hydraulic pressure rises from zero to half-clutch state to the pressure required for complete engagement. Further, the shelf pressure in the half-clutch state needs to be set higher as the input torque is larger, as shown in FIG.

FIG. 7 shows a control hydraulic system for generating such a shelf pressure. 7, in order to explain the principle of clutch shelf pressure control, the first to fourth power rollers 45a to 45d and the first to fourth upper hydraulic chambers 79a.
To 79d and the first to fourth lower hydraulic chambers 80a to 80d are schematically illustrated, and in the following description, for convenience, a hydraulic chamber that generates a force to move the power rollers 45a to 45d upward, That is, the first to fourth upper hydraulic chambers 79
The hydraulic pressure generated in a to 79d is represented by PH, and the power roller 4
5a to 45d are hydraulic chambers that generate a force to move the lower side, that is, first to fourth lower hydraulic chambers 80a to 80
Let PL denote the hydraulic pressure generated at d. The input disks 43f and 43r, the power rollers 45a to 45d, and the output disks 44f and 44r are supposed to rotate in the directions indicated by arrows A, B, and C in the figure.

In the toroidal type continuously variable transmission C schematically represented as above, the first to fourth power rollers 45 are provided.
As a reaction force of the transmission torque, a force F proportional to the transmission torque acts on the a to 45d in the direction indicated by the white arrow. Therefore, the hydraulic oil pressure PH output from one of the upshift control port P2 and the downshift control port P3 of the three-layer valve V (gear ratio control valve V), the upshift control port P2, and the downshift control port P2. The pressure PL of the hydraulic oil output from the other one of the control ports P3 for use is balanced with the force F acting on the power rollers 45a to 45d to generate a force for holding the power rollers 45a to 45d in a predetermined position. Therefore, the differential pressure .delta.P = (PH-PL) is always generated. Since the force F is proportional to the transmission torque, the differential pressure δP is also proportional to the transmission torque. By inputting this differential pressure to the clutch pressure control valve, the rack pressure of the clutch is controlled to a value proportional to the transmission torque. It can be done.

Next, the structure of the control hydraulic system shown in FIG. 7 will be described. In FIG. 7, the hydraulic oil in the oil pan 36 is discharged from the oil pump 152 and then input to the pressure reducing valve (reducing valve) 154 to be reduced to a predetermined reducing pressure. The hydraulic oil reduced to the reducing pressure is input to the linear solenoid valve 156, and the linear solenoid valve 156 generates a pilot pressure for driving and controlling the pressure regulating valve (regulator valve) 158 by changing the duty ratio of the solenoid. To do. On the other hand, the hydraulic oil output from the oil pump 152 is directly input to the pressure regulating valve 158, and this hydraulic oil is regulated by the pilot pressure from the linear solenoid valve 156, and is converted into line pressure from the pressure regulating valve 158. Is output. The hydraulic oil regulated to a predetermined line pressure is input to the three-layer valve V (gear ratio control valve V), and by the control as described above, the three-layer valve V for the shift-up control port P2 or the shift-down valve. Pressures PH and PL are output from the control port P3. These hydraulic pressure PH
, PL differential pressure δP = (PH −PL) is the power roller 4
It works as a pressure to move 5a to 45d up and down.

On the other hand, the hydraulic oil having the pressures PH and PL is
It is input to the clutch pressure control valve 160. A line pressure is input to the clutch pressure control valve 160, and the line pressure is adjusted by the clutch pressure control valve 160 to a pressure proportional to the differential pressure δP = (PH-PL) between the pressures PH and PL. , Is input to the starting clutch 150. As a result, the shelf pressure of the starting clutch 150 becomes equal to the differential pressure δ.
Pressure proportional to P, that is, toroidal type continuously variable transmission C
The pressure is controlled to be proportional to the transmission torque of.

Next, FIG. 8 and FIG. 9 are side sectional views showing a concrete structure of the clutch pressure control valve 160. Figure 8,
In FIG. 9, the clutch pressure control valve 160 is configured by fitting a spool 164 in a valve body 162 that constitutes the main body of the clutch pressure control valve 160 so as to be capable of reciprocating in the axial direction of the valve body 162. . The spool 164 is provided with a weak compression spring 16
6 always urges to the right in the figure. FIG. 8 shows the spool 164 moved to the right end of the valve body 162, and FIG. 9 shows the spool 164 moved to the left end of the valve body 162. Here, in the valve body 162, a line pressure receiving port P4 for receiving the line pressure (original pressure), a first working pressure receiving port P5 for receiving the hydraulic oil of pressure PH from the three-layer valve V, A second working pressure receiving port P6 for receiving working oil of pressure PL from the three-layer valve V, and a clutch pressure control port P7 for guiding the line pressure to the starting clutch 150,
From the clutch pressure control port P7, the feedback pressure (F /
A feedback port P8 for circulating (B pressure),
The toroidal type continuously variable transmission C is provided with a drain port P9 for releasing the clutch pressure when no torque is applied.

As shown in FIG. 8, the spool 164 has a first line closing port P4 for closing the line pressure receiving port P4 when the spool 164 moves to the right end of the valve body 162 in the figure.
The land portion 164a and the spool 164 as shown in FIG.
And a second land portion 164b that closes the drain port P9 when the valve body 162 moves to the left end in the figure. An annular groove 164c is formed between the first land portion 164a and the second land portion 164b to introduce the hydraulic oil entering from the line pressure receiving port P4 to the clutch pressure control port P7.

The first working pressure receiving port P5, the second working pressure receiving port P6 and the feedback port P8 are
Orifices 167, 168 and 170 are attached respectively to cut the high frequency component of the hydraulic pressure of the hydraulic oil, and DC
Only the components are introduced into the valve body 162. Next, the operation of the clutch pressure control valve configured as described above will be described. First, when the driver steps on the accelerator pedal to start the vehicle, the engine torque, that is, the input torque of the toroidal type continuously variable transmission C increases. Then, the control oil pressures PH and PL for supporting and holding the power rollers 45a to 45d change. In the state shown in FIG. 7, a force F is exerted in the direction of lowering the power rollers 45a to 45d downward in the figure. Therefore, in order to support this force F, the control hydraulic pressure PH needs to be higher than PL. is there. When the control oil pressure PH becomes higher than the oil pressure PL, in the clutch pressure control valve 160, the spool 164 is pushed to the oil pressure PH from the state where the spool 164 is located at the right end of the valve body 162 as shown in FIG. And the line pressure receiving port P4 is gradually opened. Since the receiving areas of the spools 164 that receive the hydraulic pressures PH and PL are the same, the opening amount of the line pressure receiving port P4 is determined by the differential pressure δP = (PH-PL) between the control hydraulic pressures PH and PL. The hydraulic fluid flowing into the groove 164c from the line pressure receiving port P4 is adjusted to a pressure proportional to the differential pressure .delta.P and introduced into the starting clutch 150 from the clutch pressure control port P7. At this time, since the clutch pressure is returned from the feedback port P8 through the orifice 168 into the valve body 162, the hydraulic pressure from which the high frequency component is removed by the orifice 168 becomes the back pressure of the spool 164, and the operation of the spool 164 becomes stable. The opening amount of the line pressure receiving port P4 is accurately controlled. The differential pressure δP is, as described above, the toroidal type continuously variable transmission C.
As a result, the pressure of the hydraulic oil introduced into the starting clutch 150 becomes a value proportional to the input torque, and the shelf pressure of the starting clutch is set to a value proportional to the magnitude of the input torque. Becomes

As shown in FIG. 10, even if the input torque is the same, the value of the differential pressure δP changes according to the change of the gear ratio, and the low side (deceleration side) where high torque is required.
Becomes higher, and becomes lower on the overdrive side (acceleration side) where torque is not required so much. Therefore, by configuring as in this embodiment, not only the clutch pressure according to the magnitude of the input torque can be obtained, but also the clutch pressure according to the gear ratio can be obtained. Especially in this embodiment,
Starting clutch 15 on the output side of the toroidal-type continuously variable transmission C
Since 0 is provided, the clutch rack pressure of the starting clutch 150 at the subsequent stage is controlled in accordance with the gear ratio changed by the transmission C, so that the clutch corresponding to both the input torque and the gear ratio is consistent. Shelf pressure will be obtained.

It should be noted that the present invention can be applied to modifications and variations of the above embodiments without departing from the spirit of the invention. For example, in the above-described embodiment, an example in which the invention is applied to a starting clutch without a fluid coupling is described, but the present invention is not limited to this, and is disclosed in Japanese Patent Laid-Open No. 62-251559 with a fluid coupling. It can also be applied to the control of the forward or reverse clutch of the type described above.

[0049]

As described above, according to the toroidal type continuously variable transmission of the present invention, the following effects can be obtained. That is, as a reaction force when the rotational force of the input disc is transmitted to the output disc, a force for moving the power roller in the tangential direction of these discs acts on the power roller. This force is proportional to the transmitted torque. Therefore, a differential pressure is generated in the two hydraulic chambers of the hydraulic mechanism that moves and drives the power roller to generate a force that balances this reaction force, and this differential pressure is proportional to the transmission torque. Therefore, if the control pressure of the control valve that controls the clutch pressure using this differential pressure is adjusted, the hydraulic mechanism for driving the power roller that is originally necessary can be provided without using a complicated linear solenoid valve as in the past. Utilizing this, it is possible to obtain the control pressure of the clutch that is proportional to the transmission torque.

[Brief description of drawings]

FIG. 1 is a cross-sectional view showing a configuration of a toroidal type continuously variable transmission according to an embodiment.

FIG. 2 is a side sectional view of the toroidal type continuously variable transmission shown in FIG.

3 is a cross-sectional view taken along the line AA of the toroidal continuously variable transmission shown in FIG.

FIG. 4 is a cross-sectional view taken along the line BB of the toroidal type continuously variable transmission shown in FIG.

FIG. 5 is a side sectional view of a three-layer structure gear ratio control valve including a sleeve.

FIG. 6 is a diagram showing a hydraulic circuit of a hydraulic mechanism of the toroidal-type continuously variable transmission according to one embodiment.

FIG. 7 is a diagram showing a control hydraulic system for generating a shelf pressure of a starting clutch.

FIG. 8 is a side sectional view showing a specific structure of a clutch pressure control valve.

FIG. 9 is a side sectional view showing a specific configuration of a clutch pressure control valve.

FIG. 10 is a diagram showing a relationship between a differential pressure of a hydraulic mechanism driving a power roller and an input torque.

FIG. 11 is a diagram showing a relationship between clutch pressure and time when the vehicle is started.

FIG. 12 is a diagram showing the relationship between the shelf pressure of the clutch and the input torque.

FIG. 13 is a view showing a conventional shelf pressure control mechanism for a clutch.

[Explanation of symbols]

 C Toroidal continuously variable transmission V Gear ratio control valve P1 Source pressure receiving port P2 Shift up control port P3 Shift down control port 17 Oil pump 82 Valve body 83 Sleeve 84 Spool 85 Spring 88 Stepping motor 90 Feedback means 150 Starting clutch 162 Valve body 164 Spool 166 Compression spring 167, 168, 170 Orifice

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Terauchi Politics, 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Mazda Corporation

Claims (1)

[Claims] 1. A cone-shaped input disk to which power is input, a cone-shaped output disk arranged coaxially with the input disk and outputting power to the outside, and the input disk and the output disk facing each other. A power roller that is disposed between the cone surfaces and that rotates while contacting both the cone surface of the input disk and the cone surface of the output disk to transmit the rotational force of the input disk to the output disk. Connected to the output disk side is a hydraulic mechanism for moving the power roller in a direction substantially orthogonal to the rotation axis of the output disk in order to tilt the power roller to change the gear ratio between the input disk and the output disk. And a moving clutch that moves the power roller by using a differential pressure between two hydraulic chambers provided in the hydraulic mechanism. In a toroidal-type continuously variable transmission that is configured to generate a driving force that causes the control pressure of a control valve that controls the clutch pressure of the starting clutch to be adjusted by using a differential pressure between the two hydraulic chambers. A characteristic toroidal type continuously variable transmission.