JP2709239B2 - Shift control device for hydraulic continuously variable transmission - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a swash plate type hydraulic pump and a swash plate type hydraulic motor which are connected via a hydraulic circuit and have a direct connection hydraulic clutch which operates according to the rotation speed of the hydraulic motor. The present invention also relates to a shift control device for a hydraulic continuously variable transmission that performs a continuously variable shift control by changing an inclination angle of a motor swash plate.

[0002]

2. Description of the Related Art Japanese Patent Publication No. 42-19728 discloses a hydraulic type continuously variable transmission having a hydraulic clutch for direct connection. When the inclination angle of a swash plate for a motor is substantially upright, that is, the reduction ratio is substantially When the ratio becomes 1: 1, the hydraulic valve for direct connection is operated by a switching valve interlocked with this to directly connect the hydraulic motor and the hydraulic pump.

[0003]

The above prior arts include:
Although a forward / reverse switching mechanism is not disclosed, if this conventional technique is applied to reverse rotation, when the hydraulic motor rotates backward, the rotation direction of the hydraulic motor and the rotation direction of the hydraulic pump (input shaft) are in different directions. Becomes In such a reverse rotation state, the motor swash plate is in an upright state,
When the direct-connection hydraulic clutch is operated, the rotation of the hydraulic motor is braked by a hydraulic pump (input shaft) that rotates in a different direction from the rotation of the hydraulic motor, resulting in a large loss of power transmission force.

[0004]

SUMMARY OF THE INVENTION It is an object of the present invention to improve the transmission efficiency during forward high-speed rotation by providing a direct-connection hydraulic clutch, to automatically perform speed change control in accordance with the engine speed and motor load, and to transmit power during reverse rotation. This is to prevent loss.

[0005]

The motor swash plate 19 is moved so that the motor swash plate 19 moves upright due to an increase in the engine rotation speed, while the motor swash plate 19 moves to the tilting side by the load of the hydraulic motor M. The hydraulic circuit includes a hydraulic distribution ring 29 that is rotatable to switch from a neutral position to a forward position and a reverse position.
The rotational position of the distribution ring is set so that the rotation direction of the hydraulic pump P and the rotation direction of the hydraulic motor M are the same direction during forward rotation of the motor and different directions during reverse rotation. The centrifugal governor valve G, which increases the pressure of the hydraulic oil in accordance with the increase in the centrifugal force due to the increase in the speed and supplies it to the oil chamber of the hydraulic clutch C2,
A hydraulic clutch for directly connecting the motor rotation speed even when the reverse rotation torque is larger than the forward rotation tilt moment and the reverse engine rotation speed is maximum, provided in the oil supply oil passage 2 The hydraulic distributing ring 2 is set so that a large tilting moment is generated to the extent that C2 does not operate.
The turning angle from the neutral position 9 to the reverse position is set large.

[0006]

The swash plate for the motor acts on the upright side by the governor hydraulic actuator whose working oil pressure rises substantially in proportion to the increase of the engine speed, for example, the square of the engine speed, and the pressing force of the motor plunger. , The tilt angle is automatically controlled in accordance with the engine speed and the motor load, whereby the rotation of the hydraulic motor is controlled continuously. When the rotation speed of the hydraulic motor increases during forward rotation, the governor valve opens due to centrifugal force and starts supplying hydraulic oil to the directly-connected hydraulic clutch. The hydraulic oil pressure supplied from the governor valve increases substantially in proportion to the square of the rotation speed of the hydraulic motor due to the pressure adjusting action of the governor valve. At high speeds, there is room for kickdown, and the clutch is completely clutched and becomes directly connected. On the other hand, during reverse rotation,
The tilting angle of the motor plunger with respect to the motor swash plate is larger than that in the forward state, and the motor rotation speed can be suppressed to a range where the direct-connection hydraulic clutch does not operate even at the maximum engine rotation. Since the motor swash plate is tilted, even if the operating hydraulic pressure of the hydraulic actuator increases due to the increase in engine speed,
The motor swash plate cannot rise upright beyond the direct coupling clutching point, and the motor swash plate is maintained at a large inclination angle (large reduction ratio). Therefore, the direct connection hydraulic clutch does not operate, and it is possible to prevent the hydraulic motor from being braked by rotation of the hydraulic pump (input shaft).

[0007]

FIG. 1 is an overall longitudinal sectional view (excluding a shift portion) of a hydraulic continuously variable transmission to which the present invention is applied. In order to facilitate understanding, FIG. Assume as described in.

First, the overall layout will be described. The front end of the transmission input shaft 1 is linked to an engine (not shown). The swash plate type hydraulic motor (axial plunger motor) M and the swash plate type hydraulic pump (axial) The plunger pumps) P are arranged in series so as to be aligned in the front-rear direction (axial direction). More specifically, the motor swash plate 1 is sequentially arranged from the front side.
9. The cylinder block 4 for the motor, the swash plate 15 for the pump, and the cylinder block 2 for the pump are arranged in series so as to be aligned in the front-rear direction (axial direction).
An intermediate drum 5 is disposed radially outward of the intermediate drum 5 at an interval, and a direct connection hydraulic clutch C2 is disposed between the intermediate drum 5 and the pump cylinder block 2. A valve body 3 having a hydraulic circuit is disposed behind the intermediate drum 5 and the cylinder block 2 for the pump.
A rear cover 23 is fixed to the rear side of the valve body 3. Inside the rear cover 23, a hydraulic distribution ring 29 for switching between forward and backward movement, which partitions the space 24 inside the rear cover 23 into an outer space 24a and an inner space 24b, a distribution ring support shaft 25 for supporting the same, and a main body. A spool 30 for clutch switching and the like are arranged. A hydraulic actuator A1 that controls the inclination angle of the motor swash plate 19 and performs automatic shift control is disposed above and in front of the motor swash plate 19.

An outline of the hydraulic pump P will be described. The pump cylinder block 2 has its inner peripheral side spline-fitted to the outer periphery of the rear end of the input shaft 1, and rotates integrally with the input shaft 1. An odd number, for example, five cylindrical holes 71 are formed in the pump cylinder block 2 at equal intervals in the circumferential direction. Each cylindrical hole 71 is formed in parallel with the input shaft 1, and Open to face. Each cylindrical hole 7
Each of the pump plungers 14 is fitted with a bottomed cylindrical pump plunger 14 so as to be axially slidable and protrudable forward, and a gentle spherical portion 14a is formed at the front end of each pump plunger 14. And is in contact with the front pump swash plate 15. The pump swash plate 15 serves as a guide for the reciprocating motion of the pump plunger 14, and includes a thrust bearing 16.
Rear end slope 4a of motor cylinder block 4 through
It is installed in. In each pump plunger 14, a coil spring 17 for urging the plunger forward is contracted between the pump plunger 14 and the rear end face of the cylindrical hole. In addition to preventing the pump cylinder block 2 from falling off, the pump cylinder block 2 is pressed against the rear valve body 3 to improve the sealing performance of the sliding seal surface 2b at low rotation, and further urges the pump plunger 14 forward. Thus, the self-suctioning ability is given to the hydraulic pump P.

The outline of the hydraulic motor M will be described. The motor cylinder block 4 is integrally fastened to the valve body 3 via the intermediate drum 5, and its inner peripheral side is rotatably fitted to the input shaft 1 via the annular oil passage 101, and The case 22 is supported via a bearing 72 and the like. In the motor cylinder block 4, an odd number of, for example, nine cylindrical holes 73 are arranged at equal intervals in the circumferential direction.
The cylindrical hole 73 is formed parallel to the input shaft 1 and opens forward as in the case of the hydraulic pump P. A bottomed cylindrical motor plunger 18 is fitted into each cylindrical hole 73 so as to be axially slidable and protrudable forward, and a loose spherical portion is formed at the front end of each motor plunger 18. 18a is formed and is in contact with the front motor swash plate 19. The motor swash plate 19 serves as a guide for reciprocating movement of the motor plunger 18, and is supported by a front swash plate holder 21 via a rolling element 20. Plunger 1 for each motor
A coil spring 37 for urging the coil spring 8 forward is contracted between the coil spring 37 and the rear end face of the cylindrical hole, thereby applying a self-sucking force to the hydraulic motor M.

The structure of the oil passage will be described. In FIG. 1, a front end of an annular oil passage 101 formed on the outer periphery of the input shaft 1 is connected to an HST charging feed pump or the like (not shown) via an annular oil chamber 102, and is connected to an annular oil chamber 102 from an oil pan or the like. The working oil is supplied to the annular oil passage 101 via the. At the rear end of the annular oil passage 101, a direct-connection hydraulic clutch cooling oil passage 103 extending through the input shaft 1 to the direct-connection hydraulic clutch C2, a motor cylinder block 4, an intermediate drum 5, and a valve body. 3 and the hydraulic oil chamber 105 of the hydraulic clutch C2 for direct connection.
And a charge that reaches the outer space 24a in the rear cover 23 through the motor cylinder block 4, the intermediate drum 5, and the valve body 3 as shown in FIG. Oil passage 108 communicates.

At the outlet of the charging oil passage 108 to the outer space 24a, there is provided a check valve 74 that opens to the outer space 24a.
Has a valve seat portion 75 provided in the valve body 3 and a steel ball 45 seated on the seat portion 75 is incorporated. The check valve 74 also tilts the valve seat 75 so that the centrifugal force generated in the steel ball 45 due to the rotation of the valve body 3 acts on the steel ball 45 rearward (in the valve opening direction). . The charging oil passage 108 is formed with a branch oil passage 108a reaching the inner space 24b, and the outlet of the branch oil passage 108a to the inner space 24b is opened in the direction of the inner space 24b. A check valve 77 is provided. The check valve 77 includes a valve seat 46, a compression coil spring 48, a steel ball 49 seated on the valve seat 46 by the elastic force of the compression coil spring 48, and the holding rod 4.
7 and the like, and are arranged at a substantially central portion of the valve body 3.

An outline of an oil passage in which the hydraulic oil supplied from the charging oil passage 108 to the space 24 through the hydraulic pump P shown in FIG. 1 reaches the hydraulic motor M and returns to the space 24 again will be described. . Regarding the hydraulic pump P, the valve body 3 communicates with the pumping cylindrical hole 71 in the suction stroke and the outer space 24a through the outer space 24a and into the pump cylindrical hole 71 as shown in the upper part of FIG. The pump-side suction oil passage 110 for sucking hydraulic oil into the pump cylinder hole 71 and the inner space portion 24b of the discharge stroke as shown in the lower part communicate with the pump cylinder hole 71 so as to be inward from the pump cylinder hole 71. A pump-side discharge oil passage 111 for discharging hydraulic oil is formed in the space 24b. Regarding the hydraulic motor M, the valve body 3,
The intermediate drum 5 and the motor cylinder block 4 have nine motor-side oil passages 11 communicating with the respective motor cylindrical holes 73.
The motor-side oil passage 113 communicates with the space 24 and the cylindrical hole 73 for each motor, but the rear end port of the motor-side oil passage 113 has the same circle centered on the rotation axis. The rotation of the valve body 3, which is arranged on the circumference and rotates integrally with the motor cylinder block 4, causes
As shown on the lower side, the cylindrical hole 73 for the motor in the expansion stroke is
The hydraulic oil is press-fitted into the motor cylindrical hole 73 from the inner space portion 24b in communication with the inner space portion 24b, while the motor cylindrical hole 73 in the discharge stroke is connected to the outer space portion 2 as shown above.
4a to the outer space 24 through the motor cylindrical hole 73.
The hydraulic oil is discharged to a.

The swash plate holder 21 shown in FIG. 1 is mounted on a transmission case 22 via a trunnion shaft 22 a perpendicular to the input shaft 1.
, And can be tilted about the trunnion axis as a rotation center. The tilt with respect to a plane perpendicular to the input axis is controlled by the tilting moment by the hydraulic actuator A1 for automatic transmission control and the motor plunger 18 described later. Change. Thereby, the stroke amount of the motor plunger 18 can be adjusted, and the motor capacity can be changed steplessly. The gear ratio of the hydraulic motor M to the hydraulic pump P is given by the following equation. Gear ratio = Pump rotation speed / Motor rotation speed = 1 + Motor capacity / Pump capacity As is clear from the above equation, if the motor capacity is changed from 0 to a certain value, the gear ratio can be changed from 1 to a certain required value. it can. The motor capacity is the motor plunger 1
When the motor swash plate 19 is in a posture perpendicular to the input shaft center, the stroke becomes zero, so that the motor capacity becomes zero and the gear ratio is one. As shown in FIG. 1, the speed ratio increases as the motor capacity increases as the vehicle inclines.

A detailed structure for controlling the inclination angle of the swash plate holder 21 will be described. Hydraulic actuator A1
A piston 35 having a spherical sliding portion is slidably fitted in a cylinder 36 formed integrally with the transmission case 22, and the piston 35 is fitted with a rear end spring receiving ring of the cylinder 36. Coil spring 81 contracted between
Urged forward. A rod is integrally formed with the piston 35, and a rod end 34 projecting rearward is connected to the piston 35.
4 is a motor swash plate holder 21 via a ball joint mechanism.
Is connected to the upper end. Oil chamber 82 in cylinder 36
Is the transmission case 2 through the throttle 39 for adjusting the flow rate.
2 communicates with the oil passage 120 in the wall. The oil passage 120
Is connected to a so-called governor hydraulic pressure source whose pressure changes substantially in proportion to the square of the rotation speed of the engine, and is supplied with hydraulic oil at the above pressure. Accordingly, the operating oil pressure in the oil chamber 82 of the cylinder 36 increases in response to the increase in the engine rotation speed (substantially proportional to the square of the engine rotation speed), and the coil spring 8
Then, the motor swash plate 19 is raised to a state perpendicular to the input shaft 1 against the input shaft 1.

On the other hand, the trunnion shaft 2 of the swash plate holder 21
The position 2a is eccentric to the opposite side (lower side in the figure) to the hydraulic actuator side with respect to the motor rotation axis, and
The motor swash plate 19 is set at a rearward position, so that a forward pressing load of the motor plunger 18 causes the motor swash plate 19 to tilt the motor swash plate 19 in the T1 direction to increase the reduction ratio. A tilting moment acts. The tilting moment increases as the motor load increases and the operating oil pressure in the motor plunger 18 increases, and the tilting moment of the motor swash plate 19 increases to increase the gear ratio (reduction ratio). I do.
That is, as described above, the governor oil pressure in the cylinder oil chamber 82 acting to raise the motor swash plate 19 in response to an increase in the engine rotational speed (substantially proportional to the square of the engine rotational speed) is used as the shift control device. A motor swash plate 19 eccentrically arranged so as to be inclined by a pressing load (reaction force) of the motor plunger 18 corresponding to an increase in the motor load, and a coil spring 81 for urging the piston 35 in a direction in which the inclination angle increases. And the gear ratio can be automatically controlled in accordance with the balance between them.

The structure inside the rear cover 23 will be described. As described above, the rear cover 23 is integrally fastened to the rear end face of the valve body 3 so as to rotate integrally with the valve body 3.
The space portion 24 is formed between the inner surfaces of the two. The hydraulic distribution ring support shaft 25 disposed in the space portion 24 is disposed substantially coaxially with the input shaft 1 and has a radial bearing 2.
The thrust bearing 26 is rotatably supported by the rear cover 23 via the rear cover 7 and the rear cover 23. The thrust bearing 26 is disposed between the rear cover 23 and the rear step portion. Restricts movement. Support shaft 25
A cylindrical portion 25a eccentric from the support axis is integrally formed at the front end of the cylinder, and the bottomed cylindrical hydraulic distribution ring 29 is fitted on the outer peripheral surface of the eccentric cylindrical portion 25a, whereby The space 24 in the rear cover 3 is divided into the distribution ring 29 as described above.
Outer space portion 24a on the outer side and inner space portion 2 on the inner side
4b. The front end face of the hydraulic distribution ring 29 is slid by being pushed to the rear end face of the valve body 3 by the hydraulic oil pressure in the inner space 24b, and the pressing strength automatically changes due to the pressure. It has a gap compensation function. Then, by the rotation of the support shaft 25, the hydraulic distribution ring 29 can be freely switched between a forward position F, a neutral position N, and a reverse position R as shown in FIG. I have.

In FIG. 4 showing a partial sectional view taken along the line IV-IV of FIG. 3, a line 22'a is parallel to the trunnion shaft 22a.
A vertical line (hereinafter referred to as a “separation line”) perpendicular to this line 22′a, which is a line passing through the motor axis,
A top dead center (most contraction point) U1 and a bottom dead center (most extension point) U2 of the motor plunger 18 are present. Assuming that the X1 direction is the forward rotation direction, the neutral position N of the distribution ring 29 is set at a position where the center of the distribution ring 29 passes on the above-mentioned forward / rearward separation line L and is most biased toward the top dead center U1. I have. The forward position F is set at a position rotated by an angle θf (approximately 90 °) from the neutral position N toward the motor forward rotation direction X1 around the support axis, while the reverse position R is shifted from the neutral position N. It is set at a position rotated by an angle θr larger than the angle θf around the support shaft center in the motor reverse rotation direction X2, and the difference between the two angles (θr−θf) = θa is, for example, approximately 5 °. Is set. That is, Kn is the distribution ring center at the neutral position, Kf is the distribution ring center at the forward position,
Assuming that Kr is the center of the distribution ring at the reverse position, in the illustrated embodiment, the distribution ring center Kf at the time of forward movement is substantially equal to the line 22 '.
a, the center of the distribution ring Kr when the vehicle is moving backward is displaced toward the bottom dead center U2 side from the line 22'a, and the line 22'a
Is approximately 5 °.

Of the pump-side oil passages 110 and 111, the pump-side suction oil passage 110 communicating with the outer space portion 24a has two
The rear end port is located at the upper outer end, and the distribution ring 29 is always in the outer space 24 regardless of the forward position F, the neutral position N, or the reverse position R.
a. Three pump-side discharge oil passages 111 communicating with the inner space 24b are formed, and their rear end ports are located in the lower part near the axis, and the distribution ring 29 is in the forward position F and the neutral position N. Alternatively, in any state of the reverse position R, it always communicates with the inner space 24b.

On the other hand, the rear ports of the nine motor-side oil passages 113 are arranged at an intermediate position in the radial direction at equal intervals on the same circumference. The range from the bottom dead center U2 to the top dead center U1 around the forward rotation direction X1 (the right range in the figure) is W1 and the top dead center U
The range (the left side range in the figure) from 1 to the bottom dead center U2 around the forward rotation direction X1 is W2. When the distribution ring 29 is in the neutral position N, the motor-side oil passage 113 substantially in the range closer to the bottom dead center U2 than the line 22'a communicates with the outer space 24a,
The motor-side oil passage 113 within the range of the top dead center U1 communicates with the inner space 24b, and the hydraulic motor M does not rotate. When the distribution ring 29 is at the forward position F, the motor-side oil passage 113 in the range W1 communicates with the outer space 24a, and the motor-side oil passage 113 in the range W2 communicates with the inner space 24b. That is, during forward rotation, the motor-side oil passage 113 in the range W1 communicates with the motor cylindrical hole 73 in the discharge stroke, and the motor-side oil passage 113 in the range W2 communicates with the motor cylindrical hole 73 in the expansion stroke. And rotates the hydraulic motor M in the forward rotation direction X1. When the distribution ring 29 is at the reverse position R, the motor-side oil passage 113 within the range W'1 shifted from the range W1 by an angle θa in the X2 direction.
Communicates with the inner space 24b, and a motor-side oil passage 113 in a range W'2, which is shifted from the range W2 by an angle θa in the X2 direction, communicates with the outer space 24a. That is, when the vehicle is traveling in reverse, the motor-side oil passage 113 in the range W′1 communicates with the motor cylindrical hole 73 in the expansion stroke, and the motor-side oil passage 113 in the range W′2 is used for the motor in the discharge stroke. It communicates with the cylindrical hole 73, and rotates the hydraulic motor M in the reverse rotation direction X2.
In this case, the virtual center of the pressing load of the motor plunger 18 is closer to the lower dead center U2 side (upper side) than in the forward rotation.
As a result, the plunger pressing load below the trunnion shaft 22a (top dead center U1 side) is reduced by θa / 180, and the plunger pressing load is higher than the trunnion shaft 22a (bottom dead center U2). Side), the plunger pressing load is increased by θa / 180, and the tilting moment is increased as compared with the tilting moment at the time of forward movement. In other words, the pressing load of the motor plunger 18 at the time of reverse movement is obtained by taking the turning angle θr at the reverse position larger than the turning angle θf at the forward position and displacing it toward the lower dead center U2 (upward). Is moved upward from the forward movement, and the tilting moment with respect to the motor swash plate 19 is increased as compared with the forward movement.
Therefore, the force for tilting the motor swash plate 19 is increased. In this embodiment, the hydraulic actuator A1 shown in FIG.
By adjusting the load of the spring 81, the motor plunger 1 can be rotated at the maximum speed even when the engine speed is maximized.
The displacement angle θa is set so that the force for tilting the motor swash plate 19 due to the tilting moment of 8 is larger than the force for raising the motor swash plate 19 to the vertical side by the hydraulic actuator A1. During reverse rotation,
At all times, the motor swash plate 19 is kept at the maximum tilt position (maximum speed ratio position), whereby the hydraulic motor M rotates in a low speed range, and the governor valve G that operates in association with the rotation of the hydraulic motor M opens. Therefore, the direct-connection hydraulic clutch C2 is not operated.

The front end face (sliding contact face) of the distribution ring 29 has a pair of wedge-shaped grooves 7 on the inner peripheral end side and one pair on the outer peripheral end side.
8a and 78b are formed. The formation positions of the grooves 78a and 78b are located near the top dead center U1 and the bottom dead center U2 when the distribution ring 29 is located at the forward position F or the reverse position R. Are formed at positions protruding from the inside and the outside, respectively.
As a result, when each motor-side oil passage 113 passes through the top dead center U1 and the bottom dead center U2, one of the inner and outer spaces is gradually closed by the distribution ring 29, and then gradually opened to the other space. It is supposed to.

The structure of the direct connection hydraulic clutch C2 will be described. In FIG. 1, the direct connection hydraulic clutch C2 is a wet multi-plate clutch,
And a plurality of separator plates 7 alternately arranged with the friction plates 6. The friction plates 6 axially slidably engage with splines 2 a formed on the outer periphery of the pump cylinder block 2. do it,
It rotates integrally with the cylinder block 2. At both ends in the axial direction of the hydraulic clutch C2 for direct connection, a pair of pressure plates 9 which can sandwich the plates 6 and 7 in the axial direction are arranged, and the pressure plates 9 are contracted between them. The two pressure plates 9 are urged in a direction away from each other in the axial direction, that is, in a clutch-off side by a coil spring (not shown).

A hydraulic piston 11 for operating the direct connection hydraulic clutch C2 is axially slidably fitted in the annular concave portion of the valve body 3, protrudes forward and projects between the pressure plates 9 between the two plates 6, 6. 7, the hydraulic motor M and the hydraulic pump P can be clutch-on (in a directly connected state).

Referring to FIG. 5, the structure of the centrifugal governor valve G of the direct connection hydraulic clutch C2 will be described. The centrifugal governor valve G is provided on the motor cylinder block 4 in the middle of the direct connection hydraulic clutch hydraulic oil supply oil passage 106 and includes a stepped moving spool 40 and a fixed sleeve 41. The fixed sleeve 41 is fixed to a hole 84 formed in the motor cylinder block 4. A stepped moving spool 40 is fitted in the sleeve 41 so as to be movable in the radial direction. The spool 40 is biased radially inward (closed position).
Reference numeral 121 denotes a drain hole which communicates with the outside when closed to prevent oil leakage from the
It comes back inside. The motor cylinder block 4 rotates, and the stepped moving spool 40 is rotated by centrifugal force.
Moves outward against the spring 42, and when the rotation speed reaches a predetermined value, the inlet-side passage 122 and the outlet-side passage 123 communicate with each other, and hydraulic oil flows into the annular step surface portion 40a of the spool 40. . Since the spool 40 has an area difference between the small diameter portion and the large diameter portion (the area of the step surface portion), the hydraulic oil in the sleeve 41 generates a force for pushing the spool 40 radially inward. Due to the balance between the inward pushing force due to the area difference and the outward force due to the centrifugal force, a hydraulic oil pressure substantially proportional to the square of the rotation speed is generated, and the direct-connection hydraulic clutch C2
To the oil chamber 105. The operating oil pressure to the direct-connection hydraulic clutch C2 increases as the rotation speed increases, whereby the operating oil is controlled so as to be in a half-clutch state at a medium speed and a complete clutching state at a high speed. .

Referring to FIG. 3, the structure of the main clutch C1 for interrupting the flow of the hydraulic oil between the hydraulic pump P and the hydraulic motor M will be described. The main clutch spool 30 includes the support shaft 2
5 is rotatably fitted to the support shaft 25 in the inner peripheral hole, and the thrust load balls 31 and 32
It is supported so that it can rotate smoothly even in the axial load. At the front end of the spool 30, a vertical oil passage 115 that opens to the inner space 24b and extends backward from the spool axis, and an oil passage 116 that extends radially outward from the rear end of the vertical oil passage 115 Is formed, while the support shaft 25 has
An oil passage 117 extending outward and opening to the outer space 24a is formed at the same axial position as the oil passage 116, and the rotation of the spool 30 causes the oil passages 116, 11 to rotate.
The main clutch C1 is turned on and off by shutting off and communicating with each other.

The operation of the entire hydraulic transmission and the operation related to the present invention will be described. When the pump cylinder block 2 is rotated by the rotation of the input shaft 1, high-pressure hydraulic oil flows from the cylindrical hole 71 that accommodates the pump plunger 14 in the discharge stroke to the inner space 24 b through the pump-side discharge oil passage 111. Is discharged, and the high-pressure hydraulic oil passes through a part of the motor-side oil passage 113 and is press-fitted into a cylindrical hole 73 that houses the motor plunger 18 in the expansion stroke. On the other hand, the hydraulic oil discharged from the cylindrical hole 73 that houses the motor plunger 18 in the discharge stroke to the outer space 24a through the remaining motor-side oil passage 113 passes through the pump-side suction oil passage 110. The liquid is returned to the pump cylinder hole 71 having the pump plunger 14 in the suction stroke. Meanwhile, the reaction torque given by the pump plunger 14 in the discharge stroke to the motor cylinder block 4 via the pump swash plate 15 and the reaction torque received by the motor plunger 18 in the expansion stroke from the motor swash plate 19 are obtained. The motor cylinder block 4 is rotated. The motor swash plate 19 is driven by the governor hydraulic actuator A1, whose working oil pressure rises substantially in proportion to the square of the engine rotation speed, on the upright side by the governor hydraulic actuator A1, and the tilting moment due to the pressing load of the motor plunger. The inclination angle is controlled in accordance with the motor load and the motor load.

When the hydraulic motor M is stopped, the governor valve G is closed because no centrifugal force is applied. That is, the oil chamber 105 of the direct-connection hydraulic clutch C2
The supply of hydraulic oil to is shut off. When the rotation speed of the hydraulic motor M increases during forward rotation, the governor valve G opens due to the centrifugal force, and starts supplying hydraulic fluid to the directly-connected hydraulic clutch C2. Due to the pressure adjusting action of the governor valve G, the hydraulic oil pressure supplied from the governor valve G increases substantially in proportion to the square of the rotation speed of the hydraulic motor. At high speeds, the clutch is completely clutched, leaving a room for kickdown. On the other hand, at the time of reverse rotation, the tilting moment of the motor plunger 18 toward the reduction ratio increasing side with respect to the motor swash plate 19 becomes larger than that at the time of the forward movement state. Even if the oil pressure increases, the motor swash plate 19 cannot be raised to the vertical side, and the motor swash plate 19 is maintained at the maximum inclination angle (maximum reduction ratio). Therefore, the direct connection hydraulic clutch C2 does not operate, and the hydraulic pump P
Can be prevented from damping the rotation of the hydraulic motor M.

[0028]

[Another embodiment] (1) Reverse rotation angle θr of distribution ring 29
Is such that even when the tilting moment during reverse movement is greater than the tilting moment during forward movement, and the engine speed during reverse movement is at the maximum, at least the motor rotation is suppressed to below the direct coupling clutch point. Therefore, for example, when the forward rotation angle θf is set to be smaller than 90 °,
θr is larger than the above θf, but may be smaller than 90 ° in some cases.

[0029]

As described above, according to the present invention, the acting force on the upright side by the governor hydraulic actuator A1 and the tilting moment due to the pressing force of the motor plunger are used in accordance with the engine speed and the motor load.
Continuously variable speed control is possible. During forward rotation, the direct connection hydraulic clutch C2
To automatically increase the power transmission efficiency during high-speed rotation. On the other hand, when the reverse rotation is performed, how much the engine speed is increased in a state where the hydraulic pump P and the hydraulic motor M are rotating in different directions. , The motor swash plate 19 can be maintained at a large inclination angle,
As a result, the operation of the direct connection hydraulic clutch C2 is prevented, and loss of power transmission during reverse rotation can be prevented.

[Brief description of the drawings]

FIG. 1 is an overall vertical sectional view of a hydraulic continuously variable transmission to which the invention of the present application is applied.

FIG. 2 is an enlarged vertical cross-sectional view of a charging oil passage circuit portion of the same hydraulic transmission as FIG.

FIG. 3 is an enlarged partial view of a portion III in FIG. 1;

FIG. 4 is a sectional view taken along line IV-IV of FIG. 3;

FIG. 5 is an enlarged longitudinal sectional view of the governor valve.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Input shaft 2 Pump cylinder block 4 Motor cylinder block 15 Pump swash plate 18 Motor plunger 19 Motor swash plate 29 Hydraulic distribution ring A1 Hydraulic actuator P Swash plate hydraulic pump M Swash plate hydraulic motor C2 Hydraulic clutch for direct connection G Governor valve

Claims (1)

(57) [Claims] 1. A swash plate type hydraulic pump P and a swash plate type hydraulic motor M are connected via a hydraulic circuit, and a direct connection hydraulic clutch C2 which operates according to the rotation speed of the hydraulic motor is provided. In a hydraulic continuously variable transmission that performs stepless shift control by changing the inclination angle of the swash plate 19, the motor swash plate 19 moves the motor swash plate 19 to the upright side due to an increase in the engine rotation speed. By the load of the motor M to move to the tilting side opposite to the above, the hydraulic circuit includes:
A hydraulic distributing ring 29 is provided which is rotatable from a neutral position to a forward position and a reverse position. The rotation direction of the hydraulic pump P and the rotation direction of the hydraulic motor M are the same when the motor rotates forward.
At the time of reverse rotation, the distribution ring rotation position is set so as to be in the different direction, and the pressure of the hydraulic oil is increased in accordance with the increase in the centrifugal force due to the increase in the rotation speed of the hydraulic motor, so that the oil of the direct connection hydraulic clutch C2 A centrifugal governor valve G to be supplied to the chamber is provided in the oil supply oil passage of the hydraulic clutch C2 for direct connection, and as a tilting moment at the time of backward movement, is larger than the tilting moment at the time of forward movement and at the time of reverse movement. The rotation angle from the neutral position to the reverse position of the hydraulic distribution ring 29 so that a large tilting moment is generated such that the motor rotation speed is limited to a range where the direct-connection hydraulic clutch C2 does not operate even at the maximum engine rotation speed. The transmission control device for a hydraulic continuously variable transmission, wherein the gear ratio is set large.