JPH07259950A - Hydraulic continuously variable transmission - Google Patents

Abstract

PURPOSE:To improve the service life of a thrust bearing by reducing the reaction force of a plunger for a motor to be applied to a thrust bearing to support a rear cover, and reduce the cost by dispensing with a thrust bearing exclusive for a supporting shaft. CONSTITUTION:In a hydraulic continuously variable transmission, a ring 49 is interposed between an outward flange-shaped supporting part 25b formed on a supporting shaft 25 to support a hydraulic distributing ring 29 and a rear cover 23, the supporting part 25b is fitted to a fitting surface 49b of the ring 49 in a relatively movable manner in the axial direction, the rear end surface of the ring 49 is locked directly or through a buffer 55 to a stationary case member on the rear side, and the ring 49 is supported by the inner circumferential hole of the rear cover 23 in a rotatable and relatively movable manner in the axial direction through a bearing 48. Apart of the reaction force of the plunger for the motor is received by a stationary case member through the ring 49 to reduce the thrust force on the bearing.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a fixed displacement hydraulic pump and a hydraulic closed circuit which is connected to a discharge portion and a suction portion of the hydraulic pump via a supply side oil passage and a return side oil passage, respectively. The present invention relates to a hydraulic continuously variable transmission including a variable displacement type swash plate hydraulic motor, and particularly to a support structure such as a distribution ring support shaft provided therein.

[0002]

2. Description of the Related Art FIG. 20 shows a structure of a distribution ring supporting shaft of a conventional hydraulic continuously variable transmission (Japanese Patent Application No. 5-130385). The rear cover 23 is integrally connected to the front motor cylinder block via the valve body 3 so as to rotate integrally with the valve body 3 and the motor cylinder block, and an oil chamber 24 is formed in the rear cover 23. ing. The hydraulic distribution ring support shaft 25 arranged in the oil chamber 24 is arranged substantially coaxially with the input shaft 1, and is rotatably supported by the rear cover 23 via a radial bearing 327, and is arranged behind the rear cover 23. A thrust bearing 326 is arranged between the side step portion and the side bearing portion, thereby restricting movement in the axial rearward direction while ensuring rotation.

On the eccentric head 25a at the front end of the support shaft 25,
A bottomed tubular hydraulic distribution ring 29 is fitted, whereby the oil chamber 24 is moved to the outer oil chamber 24a on the outer side of the distribution ring 29.
And an inner oil chamber 24b on the inner side. The inner oil chamber 24b communicates with a supply side oil passage of a hydraulic closed circuit formed between a hydraulic pump and a hydraulic motor, and the outer oil chamber 24a communicates with a return oil passage.

[0004]

In FIG. 20, the axial load applied to the distribution ring support shaft 25 is different between when the engine drive travels and when the engine brake travels.
When the engine is driven, the support shaft 25 is pushed rearward in the axial direction by the high hydraulic pressure of the inner oil chamber 24b, and this load is
From the thrust bearing 326 and the outer oil chamber 24a to the valve body 350
They are balanced by the force of being guided inward and pushing the support shaft 25 back. In this engine drive state, the inner oil chamber 24b has a higher pressure than the outer oil chamber 24a, and since the pressure receiving area on the inner oil chamber 24b side is large,
The thrust bearing 326 is always pushed rearward by the support shaft 25.

However, when the engine is braked, the inner oil chamber 24b becomes lower in pressure than the outer oil chamber 24a, so that the support shaft 25 is pulled forward and presses the distribution ring 29. Therefore, the thrust bearing 326 is placed in a state where there is a gap in the axial direction.

When the engine braking state is changed to the engine drive state described above, the inner and outer oil chambers 2 are
Since the pressure levels of 4b and 24a are suddenly reversed, the support shaft 25 is suddenly pushed rearward and strongly collides with the thrust bearing 326.

The reaction force to the rear of the motor plunger is mainly handled by the thrust bearing 328 via the rear cover 23. Generally, in order to improve the transmission efficiency of the hydraulic transmission, the pressure acting on the motor plunger is
The thrust bearing 328 is set as high as possible, so that the reaction force is also large, and therefore the load applied to the thrust bearing 328 is extremely large.
No. 8 requires a bearing with high pressure resistance and an expensive price.

As a countermeasure against this, the support shaft 25 is replaced by the one shown in FIG.
The structure for abutting and supporting the end surface of the rear cover 23 by the thrust bearing 326 as described above is abolished, and the rear end portion of the support shaft 25 is directly brought into contact with the stationary case member (transmission case or the like), so that the thrust bearing 328 The load acting on can be reduced. This is disclosed, for example, in JP-A-53-8846.
It is used in the inventions described in No. 0 and the like.

However, even if the thrust force applied to the support shaft 25 is shared by the stationary case member in this way, the load applied to the thrust bearing 328 for supporting the rear cover during running of the engine drive is reduced, but the engine brake is used. At this time, since the support shaft 25 is pulled in to the front side (distribution ring side), the rearward thrust force due to the reaction force of the motor plunger applied to the thrust bearing 328 is not reduced.

Further, each time the engine braking and the engine driving are repeated, the support shaft 25 collides with the stationary case member, causing damage and causing abnormal noise.

[0011]

[Object of the Invention]

(1) The thrust force due to the reaction force of the motor plunger applied to the bearing for supporting the rear cover is reduced to reduce the life of the bearing or to reduce the cost of parts during engine driving and engine braking, or their repeated states. Aim to reduce. (2) Thrust bearings dedicated to the distribution ring support shaft will be abolished to reduce costs, improve durability, and prevent abnormal noise. (3) Also, the inclination of the distribution ring is prevented to improve the sealing performance between the distribution ring and the end surface of the valve body.

[0012]

[Means for Solving the Problems]

[Invention of Claim 1] In FIG. 5, a fixed displacement hydraulic pump P is connected to a discharge portion and a suction portion of the hydraulic pump via a supply-side oil passage B1 and a return-side oil passage B2, respectively. A variable displacement swash plate hydraulic motor M that constitutes a hydraulic closed circuit is provided. In FIG. 1, the rear cover 23
Is the axial direction of the motor plunger
The valve body 3 is provided on the rear side of the motor cylinder block 4.
Through a bearing 7 and are rotatably supported by a stationary case member (rear support wall 22a) via a bearing 7. The hydraulic distribution ring 29 is rotatably slidably arranged on the rear surface of the valve body, and the inside of the rear cover is a part of the supply side oil passage B1 and the outer side is a part of the return side oil passage B2. It is partitioned into a square oil chamber 24a. The distribution ring support shaft 25 is supported by the front end eccentric head 25a so as to be movable in the axial direction, presses the distribution ring 29 forward, and projects rearward from the inside of the rear cover to be connected to the shift mechanism 10. In such a hydraulic continuously variable transmission,
In FIG. 12, the front surface of the support shaft 25 is the outer oil chamber 24.
forming an outward flange-shaped support portion 25b facing a,
An annular ring 49 is interposed between the bearing portion 25b and the inner peripheral support hole of the rear cover 23. On the inner peripheral surface of the ring 49, a fitting surface 49b is formed in the front half portion via a step portion 49a, and the outer peripheral surface of the support portion 25b is fitted to the fitting surface 49b so as to be relatively movable in the axial direction. The rear end surface of the ring 49 is locked to the stationary case member (46) on the rear side either directly or via a cushioning material 55. The outer peripheral surface of the annular ring 49 is supported in the inner peripheral support hole of the rear cover 23 via a bearing 48 so as to be rotatable and relatively movable in the axial direction.

[Invention of Claim 2] In addition to the invention of Claim 1, in FIG. 12, a relative axial direction between the support portion 25b and the ring 49 is provided between the rear surface of the support portion 25b and the ring 49. A damper oil chamber 87 whose volume changes by the movement is formed.

[Invention of Claim 3] In addition to the invention of Claim 1, a protrusion 25c for pressing the approximate center of pressure in the distribution ring 29 forward is provided in the approximate center of the front surface of the head of the support shaft 25. Is forming.

[0015]

When the engine is driven, the inner oil chamber 24b becomes higher in pressure than the outer oil chamber 24a as shown in FIG.
Is pushed rearward with respect to the annular ring 49 and the rear cover 23, but this movement is performed gently by the damper action of the damper oil chamber 87, and the rear surface of the bearing portion 25b engages with the annular step portion 49a. . The rearward reaction force of the motor plunger applied to the rear cover 23 is distributed and received by the rear end edge of the support shaft 25, the rear end edge of the circular ring 49 (buffer material 55), and the bearing 7. , The load due to the reaction force applied to the bearing 7 that rotates relatively is reduced.

During engine braking, as shown in FIG.
The outer oil chamber 24a has a higher pressure than the inner oil chamber 24b, and the distribution ring support shaft 25 is pulled forward with respect to the rear cover 23 and the like. At this time, the annular ring 49 receives the pressure of the high-pressure oil in the outer oil chamber 24a on its front annular pressure receiving surface and is pressed against the stationary case member 46. That is, the valve body 3 is pushed forward through the hydraulic pressure in the outer oil chamber 24a, thereby reducing the thrust force due to the reaction force of the motor plunger applied to the bearing 7.

At this time, the front end protrusion 25c of the support shaft 25
Presses the approximately center of pressure of the distribution ring 29 forward to uniformly press the front sealing surface of the distribution ring 29 against the end surface of the valve body 3 to prevent the distribution ring 29 from rising and maintain the sealing performance. To do.

[0018]

As described above, according to the invention described in claim 1 of the present application, (1) the bearing for supporting the rear cover 23 in any state of engine driving and engine braking or repeated states thereof. It is possible to reduce the thrust force due to the reaction force of the motor plunger related to 7, and it is possible to extend the life of the thrust bearing 7 or reduce the cost of parts.

(2) Since the support shaft 25 is supported by the stationary case member via the ring 49 facing the outer oil chamber 24a, a thrust bearing dedicated to the distribution ring support shaft (bearing 328 in FIG. 20). ) Can be abolished, and cost reduction, durability improvement and prevention of abnormal noise can be achieved.

(3) According to the invention described in claim 2, in addition to the above effects, since the damper oil chamber 87 is formed between the bearing portion 25b and the ring 49, when the engine brake is changed to the engine drive. In, the collision (impact) between the circular ring 49 and the support shaft 25 due to the rearward movement of the support shaft 25
Can be alleviated, and the effect of further improving durability and preventing abnormal noise can be obtained.

(4) According to the invention of claim 3, in addition to the above effects, when the distribution ring 29 is pressed against the end surface of the valve body 3 during engine braking, the center of the hydraulic repulsive force of the distribution ring 29 is concentrated. By pressing it forward, the inclination of the distribution ring 29 can be prevented and the sealing performance between the distribution ring and the end surface of the valve body can be improved.

[0022]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS FIG. 1 is an overall vertical cross-sectional view of a hydraulic continuously variable transmission to which the present invention is applied, in which a motor plunger projection direction (left side in the drawing) is defined as an axial forward direction. There is.

First, the overall layout will be described. A variable displacement type swash plate type hydraulic motor (axial type plunger motor) M and a fixed displacement type swash plate type hydraulic pump (axial type plunger pump) P are arranged in this order from the front side on the outer periphery of the speed change input shaft 1 on the drive source side. And are arranged in series in the axial direction.

More specifically, the motor swash plate 19, the motor cylinder block 4, the pump swash plate 15 and the pump cylinder block 2 are arranged in this order from the front side, and the pump cylinder block 2 is radially outward. Is provided with an intermediate drum 5 which covers the same at intervals, and between the intermediate drum 5 and the pump cylinder block 2, both 5, 2
A multi-plate hydraulic direct coupling clutch C1 for directly coupling the above is arranged.

A valve body 3 having a hydraulic circuit is arranged at the rear side of the intermediate drum 5 and the pump cylinder block 2, and a rear cover 23 is arranged at the rear side of the valve body 3, and the valve body 3 and the rear cover 23 are , Is integrally fastened together with the intermediate drum 5 to the motor cylinder block 4 (FIG. 4), rotates integrally with the motor cylinder block 4, and the axial thrust force is also directly transmitted to each other.
The rear cover 23 will be described in detail later, but the rear support wall portion 22a of the transmission case 22 is mounted via the rolling ball bearing 7.
Is rotatably supported on the rear cover 2 and is immovable rearward.
An output gear 53 is integrally formed at the rear end of the wheel 3, and the output gear 53 meshes with a wheel-side gear 54. In the rear cover 23, an oil chamber 24 in the rear cover is provided with an oil chamber 24 in the outer side.
a hydraulic distribution ring 29 for switching between forward and backward movement, which is divided into a and an inner oil chamber 24b, a distribution ring support shaft 25 for eccentrically supporting the hydraulic distribution ring 29, an inner pipe 31 and the like installed inside the support shaft 25. There is.

A rear case 46 is fixed to the rear surface of the rear support wall portion 22a of the transmission case 22, and a cover case 47 is fixed to the rear side thereof. The mechanism 10 and the clutch valve 72 for short-circuiting the inner and outer oil chambers are provided.

On the radially outer side of the motor M and the pump P, a shift control hydraulic actuator, that is, a control ram, which controls the tilt angle of the motor swash plate 19 to perform automatic shift control and speed ratio lock control. A1 is arranged.

The outline of the hydraulic pump P will be described. The inner circumference of the pump cylinder block 2 is spline-fitted to the input shaft 1 and rotates integrally with the input shaft 1. The pump cylinder block 2 has an odd number (for example, 5) of cylindrical holes 41 formed at equal intervals in the circumferential direction, and each cylindrical hole 41 is formed parallel to the input shaft 1. It opens toward the front. A bottomed cylindrical pump plunger 14 is fitted in each cylindrical hole 41 so as to be slidable in the axial direction, and the front end spherical surface portion of each pump plunger 14 is in contact with the pump swash plate 15. . The pump swash plate 15 serves as a guide for the reciprocating motion of the pump plunger 14, and is installed on the rear end slope of the motor cylinder block 4 via a thrust bearing 16. A coil spring 17 for biasing the pump plungers 14 forward is provided in each pump plunger 14.
Is contracted between the rear end surface of the cylindrical hole and
The pump swash plate 15 is pressed to prevent the pump swash plate 15 from falling off, and the pump cylinder block 2 is pressed to the rear valve body 3 to improve the sealing performance of the sliding seal surface 2b at low rotation speed. Further, the self-suction capability is imparted to the hydraulic pump P by urging the pump plunger 14 forward.

The outline of the hydraulic motor M will be described. The inner peripheral side of the motor cylinder block 4 is rotatably fitted to the input shaft 1 via an annular oil passage 50, and is supported by the transmission case 22 via bearings 44 and the like. The motor cylinder block 4 has a plurality of (for example, 9) cylindrical holes 43 formed at equal intervals in the circumferential direction. The cylindrical holes 43 are formed in parallel with the input shaft 1. It opens toward the front. A cylindrical motor plunger 18 having a bottom is fitted into each cylindrical hole 43 so as to be slidable in the axial direction and projectable forward.
The front-end spherical surface of is in contact with the motor swash plate 19. The motor swash plate 19 serves as a guide for the reciprocating motion of the motor plunger 18, and is supported by the swash plate holder 21 on the front side via the rolling elements 20 and the like. A coil spring 37 for biasing the motor plungers 18 forward is contracted between the motor plungers 18 and the rear end surface of the cylindrical hole, thereby imparting a self-suction force to the hydraulic motor M.

An outline of the HST circuit (closed circuit) formed between the hydraulic motor M and the hydraulic pump P will be described. Regarding the hydraulic pump P, the valve body 3 is provided with oil passages 148 and 149 which connect the pump cylindrical hole 41 and the inner oil chamber 24b or the outer oil chamber 24a, and the cylindrical hole 41 in the discharge stroke is The hydraulic oil is discharged to the inner oil chamber 24b through the oil passage 148 in communication with the inner oil chamber 24b, the cylindrical hole 41 in the suction stroke communicates with the outer oil passage 24a, and through the oil passage 149. The hydraulic oil is sucked from the outer oil chamber 24a.

Regarding the hydraulic motor M, the valve body 3,
In the intermediate drum 5 and the motor cylinder block 4, each motor cylindrical hole 43 and the inner oil chamber 24b or the outer oil chamber 2 is provided.
4a is formed with oil passages 150 and 151, and the motor cylindrical hole 43 in the expansion stroke is formed in the inner oil chamber 24b.
Hydraulic oil is sent under pressure from the inner oil chamber 24b into the motor cylindrical hole 43, while the motor oil cylindrical hole 4 in the discharge stroke is discharged.
3 communicates with the outer oil chamber 24a and discharges the working oil from the motor cylindrical hole 43 to the outer oil chamber 24a.

A part of FIG. 5 shows a simplified diagram of the HST circuit (closed circuit) formed between the pump P and the motor M. The hydraulic pump P is connected to the engine 11 via the input shaft 1. The hydraulic oil supply side oil passage B1 which connects the discharge portion of the hydraulic pump P and the inlet portion of the hydraulic motor M is the oil passage 14
8, the inner oil chamber 24b, the oil passage 150, and the like, and have a high pressure when the engine is driven. Return-side oil passage B2 that connects the outlet of the hydraulic motor M and the suction of the hydraulic pump P
Is composed of the oil passage 149, the outer oil chamber 24a, the oil passage 151, etc., and is normally at a low pressure when the engine is driven. Further, during engine braking, the hydraulic motor M drives the hydraulic pump P, and the return side oil passage B
2 becomes high pressure and the supply side oil passage B1 becomes low pressure.

The clutch valve 72 is a valve that eventually short-circuits the two oil passages B1 and B2 in a closed circuit. A centrifugal vehicle speed governor valve 76 is provided in the return-side oil passage B2 via an oil passage 75.
The inlet port of the vehicle speed governor valve 76 is connected to the outlet port of the vehicle speed governor valve 76 via the oil passage 77.
It communicates with 8.

The respective oil passages B1 and B2 of the HST circuit communicate with the oil passage 94 via check valves 95b and 95a, respectively, which oil passage 94 is connected to the secondary pressure oil passage 102, the primary pressure regulating valve 121 and It is connected to the charging pump 100 via the primary pressure oil passage 101.

The structure of the centrifugal vehicle speed governor valve 76 will be described. In FIG. 1, a valve case 80 of the centrifugal vehicle speed governor valve 76 is fixed to a radial fitting hole of the intermediate drum 5, and a spool valve body 81 is fitted in the valve case 80 so as to be movable in the radial direction. And the valve spool valve element 81 has a spring 8
The valve is closed by being urged inward in the radial direction by 2. When the vehicle speed reaches a constant high speed value, the spool valve element 81 moves outward in the radial direction against the spring 82 to open the valve.

Oil passage 75 on the inlet side of the vehicle speed governor valve 76
Is formed in the intermediate drum 5 and the valve body 3 and communicates with the outer oil chamber 24a. A signal oil passage 83 formed in the motor cylinder block 4 communicates with the oil passage 77 on the outlet side of the governor valve 76, and the oil passage 83 is connected to the transmission case 22 via the annular oil passage 50. The oil passage 90 communicates with the oil passage 90 in the front wall, and as shown in FIG. 5, the oil passage 90 communicates with the hydraulic oil chamber of the stopper cylinder 115 of the governor lever Y.

FIG. 3 is a sectional view taken along the line III-III of FIG. 2, in which the oil passage 77 connected to the outlet of the vehicle speed governor valve 76 is bent in a U shape and then extends rearward as shown in FIG. The hydraulic oil chamber 78 of the direct coupling clutch C1.

A shift mechanism 10 for switching between forward and reverse travel will be described with reference to FIG. A gear portion 32 is integrally formed at a rear end portion of the support shaft 25. The gear portion 32 meshes with a fan-shaped shift gear 60, and the shift gear 60 has an upper end portion rotatably supported by a switching shaft 56. In addition, it has a long hole 60a that is long in the vertical direction in the middle part. The eccentric cylindrical protrusion 58a of the intermediate gear 58 is engaged with the elongated hole 60a.
The intermediate gear 58 is rotatably supported by a support pin 59 fixed to the rear case 46, and meshes with the drive gear 57. The drive gear 57 is fixed to the switching shaft 56 and rotates integrally with the switching shaft 56. The switching shaft 56 is rotatably supported by the rear case 46 and the cover case 47, extends to the outside of the case, and is interlocked with an appropriate operating member such as an operating lever.

That is, by rotating the switching shaft 56 from the outside, the intermediate gear 58 is rotated via the drive gear 57, and the eccentric cylindrical protrusion 58a causes the shift gear 60 to move through the elongated hole 60a. The support shaft 25 and the hydraulic pressure distribution ring 29 are rotated through the gear portion 32 by swinging around.

FIG. 18 is an enlarged vertical rear view of the shift mechanism 10. The drive gear 57 has a gear portion 57a meshing with the intermediate gear 58 and a circumferential surface 57 centered on the switching axis.
b, the circumferential surface 57b is provided with a forward positioning groove 61a, a neutral positioning groove 61b, and a backward positioning groove 61c for clarifying the forward, neutral and reverse shift positions. , Positioning grooves 61a, 61b, 61
A steel ball 62 for positioning engages with c in a disengageable manner. The steel ball 62 is fitted in a hole 63 formed in the cover case 47 and is urged toward the drive shaft center by a compression coil spring 64, and the elastic force of the compression coil spring 64 causes, for example, forward positioning. It meshes with the groove 61a.

The respective positioning grooves 61a, 61b, 61c are
It has a substantially U-shape, a V-shape, or a sloped shape, and when a rotational torque above a certain level is applied to the drive gear 57,
Due to the cam action of the inclined surface of the groove 61a or the like, the steel ball 62 can be pushed away and rotated against the elastic force of the spring 64.

On both sides of the shift gear 60 in the rotating direction, there are provided a forward position regulating stopper 85 and a fixed wall type backward position regulating stopper 86 capable of adjusting the protrusion amount.

In addition to the positioning mechanism for the steel balls 62 and the stoppers 85 and 86, the shift mechanism 10 is provided with a neutral lock actuator 129 for completely locking the drive gear 57 at the neutral position. It is configured as follows. A rod supporting hole 68 and a cylinder 71 communicating with the rod supporting hole 68 are formed in the cover case 47,
In the cylinder 71, the piston portion 66 of the cylindrical rod 66 is provided.
a is slidably fitted, and the rod 66 is fitted in the support hole 68 in a liquid-tight state and penetrates through the support hole 68 to project toward the center of the drive gear. A plug 73 is screwed to the cylinder 71, and a spring 67 compressed between the plug 73 and the piston portion 66a urges the rod 66 toward the center of the drive gear. The lock groove 70 formed on the surface 57b is engaged. Lock groove 70
Is formed in a U shape, while the tip end surface of the rod 66 is formed in a flat shape.
When is engaged, it does not come off due to the rotational torque of the drive gear 57.

The inside of the cylinder 71 is partitioned by a piston portion 66a into an oil chamber 69 and a spring chamber 74, and the spring chamber 74 communicates with the inside of the case 47 through a relief passage 65. As shown in FIG. 5, the oil chamber 69 is connected to the inlet port of the neutral lock actuation valve (solenoid valve) 126 via the oil passage 106.
It is connected to the secondary pressure oil passage 102. When hydraulic oil is pumped to the oil chamber 69, the rod 66 moves backward against the spring 67, and the neutral lock state is released.

The solenoid of the neutral lock operating valve 126 is
As shown in FIG. 11, the brake actuation detection switch 226 is connected, and when the brake is actuated (when the brake is on), the neutral lock actuation valve 126 is closed, and as shown in FIG.
02, the hydraulic oil is pressure-fed to the oil chamber 69 through the oil passage 106. That is, the rod 66 is retracted when the brake is operated.

Referring to FIG. 12, the support structure and the internal structure of the rear cover 23 will be specifically described. Rear cover 23
A bearing mounting surface 23b is formed on the outer periphery of the intermediate portion in the axial direction of the bearing via a step portion 23a, the inner race of the rolling ball bearing 7 is fitted to the mounting surface 23b, and the front end surface is in the step portion 23a. Abutting. The outer race of the rolling ball bearing 7 is fitted on the inner peripheral surface of the boss 6 formed on the rear support wall 22a of the transmission case 22, and the rear end surface is in contact with the rear inward flange 6a of the boss 6.

The hydraulic distribution ring support shaft 25 has a hollow hole penetrating in the front-rear direction and is arranged substantially coaxially with the input shaft 1. A cylindrical head portion 25a that is eccentric from the support shaft center is integrally formed at the front end portion of the support shaft, and a clutch oil hole 51 for the main clutch that radially penetrates in a cross shape is formed at a rear side portion of the head portion 25a. A fitting support portion 25b which is concentric with the support shaft and extends outward is formed at the rear side portion thereof, and further extends rearward to pass through the rear case 46, and the rear end portion is
The rotation sleeve 34 for the clutch valve 72 is rotatably and axially movably supported by an O-ring on the large diameter portion of the inner peripheral surface of the rotation sleeve 34.

The inner pipe 31 extends rearward of the rear end edge of the support shaft 25, is fitted and supported by the latter half small diameter portion of the inner peripheral surface of the rotating sleeve 34, and further extends rearward to cover the cover case. It is supported by a boss portion 47a of 47 and is regulated by a positioning and fixing bolt 38 so as not to rotate in the axial direction. An annular oil passage 33 is formed between the outer peripheral surface of the inner pipe 31 and the inner peripheral surface of the support shaft 25, and the front end portion of the annular oil passage 33 is provided with an oblique oil passage 84 to the outer oil chamber 24 a. And a rear end portion thereof communicates with the annular oil chamber 42 between the rear end portion and the rotating sleeve 34. The outer peripheral surface of the rotating sleeve 34 is provided with a rolling ball bearing 40 and a bearing supporting hole 47c of the cover case 47.
It is rotatably supported and is immovable in the axial direction.

The support shaft 25 is rotated by the shift mechanism 10 from the neutral position in FIG. 18 to the F side or the R side about the support shaft center by a certain angle, whereby the distribution ring 29 is moved to the forward position about the support shaft center. (Figure 17) or reverse position (Figure 19)
12 to each port P0 (FIG. 18) of the oil passages 150 and 151 opening on the sliding contact surface of the valve body 3 in FIG. 12, whereby the inner oil chamber 24b and the outer oil chamber 24b are opened. The communication relationship of the oil chamber 24a is switched, and the HST circuit is set to the forward or reverse state.

The charging hydraulic pressure supply path will be described.
In FIG. 12, supply oil passages 94a and 94b for HST hydraulic oil replenishment are formed in the cover case 47 which is a stationary member, and the secondary pressure oil is provided via the check valves 95a and 95b and the oil passage 94, respectively. It communicates with the road 102. The upper end of the one oil passage 94b communicates with the inner oil chamber 24b through the hole at the rear end of the inner pipe 31 and the inside of the inner pipe 31. The upper end of the other oil passage 94a communicates with the rear oil chamber 45 of the rotating sleeve 34, and the oil chamber 45 has an axial oil groove 35 formed in the inner peripheral surface of the rotating sleeve 34. Through the oil chamber 42
Is in communication with. That is, the oil passage 9 of the cover case 47
4a is an oil chamber 45, an oil groove 35, an oil chamber 42, an annular oil passage 33.
And is always communicated with the outer oil chamber 24a through the oblique oil passage 84.

Thus, the inner oil chamber 24b and the outer oil chamber 2 are
4a communicates with the oil passages 94b, 94a in the cover case 47, which is a stationary case member, through the passage in the inner pipe 31 of the support shaft 25 that rotates only during the shift operation, and the check valve 9
By communicating with the charging pump 100 (FIG. 5) via 5a, 95a, etc., a complicated sealing structure such as a device provided between a rotating member and a stationary member is not required, and charging is performed with a simple sealing structure. Hydraulic pressure can be supplied to the HST circuit.

The specific structure of the clutch valve 72 will be described.
A clutch oil hole 36 is formed in a fitting portion of the inner pipe 31 with the rotating sleeve 34, and when the oil groove 35 is aligned with the oil hole 36 by the rotation of the rotating sleeve 34, an inner side is formed. Inside the pipe 31, through the oil hole 36, the oil groove 35, the oil chamber 42, the annular oil passage 33, and the diagonal oil passage 84, the inner and outer oil passages 2
4b and 24a are connected. That is, the clutch valve 72
Open to short circuit the HST circuit. On the other hand, by rotating the rotating sleeve 34, the position of the oil groove 35 is shifted in the circumferential direction to close the oil hole 36.
The connection between b and 24a is cut off. The oil groove 35 is provided so as to connect the oil chambers 42 and 45 at both ends of the rotating sleeve 34 to each other, and the charging hydraulic pressure from the check valve 95a is applied to the outer oil chamber 2.
Lead to 4a.

The rotating sleeve 34 has its lever portion 34a.
Is connected to a hydraulic actuator (not shown) that operates with governor hydraulic pressure that changes depending on the rotation speed of the engine 11, and opens when idling and closes when the engine speed increases. Further, when the direct coupling clutch is operated, the oil groove 35 and the inside of the inner pipe 31 are opened to communicate with each other.

The detailed support structure of the support shaft 25 will be described.
As described above, the fitting support portion 25b formed on the support shaft 25 in the outward flange shape has the front surface facing the outer oil chamber 24a,
An annular ring 49 is arranged on the outer periphery of the bearing portion 25b, and a fitting surface 49b is formed on the inner peripheral surface of the annular ring 49 via a forward facing step portion 49a, and the fitting surface 49b is sealed. The support portion 25b is fitted and supported via a ring so as to be rotatable and relatively movable in the axial direction. The outer peripheral surface of the annular ring 49 is rotatably supported in the rear inner peripheral support hole 39 of the rear cover 23 via a roller rolling bearing 48, and the outer peripheral surface front end 49c is directed inward in the rear cover 23 via a seal ring. It is fitted in the fitting portion 23d in a liquid-tight state so as to be relatively movable in the axial direction. The rear end surface of the ring 49 is in contact with the front end surface of the rear case 46 via a disk-shaped cushioning material 55 such as resin.

Between the step portion 49a of the annular ring 49 and the rear surface of the support portion 25b, there is formed a damper oil chamber 87 whose volume is changed by the axial movement of the support shaft 25, and the damper oil chamber 87 is supported. A narrow annular oil passage 88 between the outer peripheral surface of the shaft 25 and the rear inner peripheral surface of the ring 49 communicates with the oil sump in the cover case 47.

Outer diameter of the support portion 25b (fitting portion 4 of the ring 49)
The bore diameter 9b) D1 is determined to such an extent that the pressing force of the distribution ring 29 required during engine braking can be maintained. That is, if the outer diameter D1 of the support portion 25b is increased, the support shaft 2
The force of pushing 5 forward is small, and conversely, the force of pushing 5 is large. On the other hand, the outer diameter D2 of the seal portion of the front end portion 49c of the outer peripheral surface of the annular ring 49 is equal to the outer diameter D2 in both the engine braking state and the engine driving state.
The total thrust force around the support shaft 25 is set as large as possible within the limit that it does not exceed the axial reaction force of the motor plunger 18 and does not push back the rear cover 23 and the like.

The reaction force of the motor plunger 18 works backward regardless of whether the engine is being driven or the engine is being braked. This reaction force is applied to the rear end edge of the support shaft 25 and the rear end surface of the ring 49. The bearing 7 is being received by the bearing 7, but since the bearing 7 is relatively rotating, the rear case 4 which is a stationary case member via the cushioning member 55 so that reaction force is not concentrated on the bearing 7.
The load on the bearing 6 is reduced, and the backward reaction force of the motor plunger on the bearing 7 is reduced.

Increasing the seal diameter (outer diameter) D2 of the annular ring 49 increases the pressure receiving area of the annular ring 49 facing the outer oil chamber 24a, so that the working oil in the outer oil chamber 24a is passed through. As a result, the force for pushing the entire body (valve body 3) to the left increases, and the load on the bearing 7 against the reaction force of the motor plunger 18 decreases accordingly. Particularly, during engine braking, it is effective when the outer oil chamber 24a has a high pressure and the support shaft 25 is pulled in.

On the front end face of the support shaft head 25a, the head 2 is
A protrusion 25c that slightly protrudes forward is formed at the center of the 5a so as to push substantially the central portion of the hydraulic distribution ring 29 during engine braking.

When the engine is driven, the inner oil chamber 24b becomes higher in pressure than the outer oil chamber 24a as shown in FIG. 13, and the support shaft 25 is pushed rearward with respect to the ring 49 and the rear cover 23. Due to the damper action of the chamber 87, the chamber gradually moves rearward so that the rear surface of the support portion 25b has an annular step portion 49.
engage a. The rearward reaction force of the motor plunger applied to the rear cover 23 is caused by the rear end edge of the support shaft 25 and the annular ring 4.
The rear end edge 9 and the bearing 7 are distributed and received, thereby reducing the load applied to the thrust bearing 7 that relatively rotates.

During engine braking, as shown in FIG.
The outer oil chamber 24a has a higher pressure than the inner oil chamber 24b, and the distribution ring support shaft 25 is pulled forward with respect to the rear cover 23 and the like. At this time, the annular ring 49 receives the pressure of the high-pressure oil in the outer oil chamber 24a at its front annular pressure receiving surface and is pressed against the stationary case member 46, whereby the reaction force of the motor plunger applied to the bearing 7 is generated. Reduces thrust power.

Further, when the distribution ring 29 is pressed against the end surface of the valve body 3 during engine braking, the center of the hydraulic repulsion force is concentratedly held to prevent the distribution ring 29 from tilting and to prevent the distribution ring 29 and the body from colliding. Maintains the sealing performance between the end faces.

A concrete structure of the control ram A1 as a hydraulic actuator for automatic shift control will be described with reference to FIGS. In FIG. 14, control ram A
1 is a three-stage piston type, and has a small-diameter main piston 161, which is a front stage, an intermediate-diameter first auxiliary piston 162, which is a middle stage, and a large-diameter second auxiliary piston 163, which is a rear stage. The cylinders 164 are provided in series. The multi-stage cylinder 164 has five stages of tubular portions 171,1.
72, 173, 174, 175 are integrally formed, and the diameter is sequentially increased from the first cylindrical portion 171 at the front end to the rear side.

The main piston 161 is formed in the shape of a bottomed cylinder whose rear end is open, is connected to the upper end of the swash plate holder 21 via the connecting link 12, and the main piston 1161 moves forward. Along with this, the swash plate holder 21 stands up.

The main piston 161 has an outer circumference at the front end fitted to the inner circumference of the first tubular portion 171 via a seal so as to be movable in the axial direction, and an enlarged diameter fitting at the rear via an annular step 161a. Is formed, is fitted to the inner peripheral surface of the second tubular portion 172 so as to be movable in the axial direction, and constitutes an annular first oil chamber 181 between the second tubular portion 172 and the main piston 161. .

The first auxiliary piston 162 is formed in a bottomed cylinder shape having an opening at the front end, and has an annular step 162 at the rear end.
a is formed. The front end is the main piston 16
The rear end edge 1 is formed to have substantially the same shape and can be brought into contact with it. The rear enlarged diameter seal portion is fitted to the inner peripheral surface of the third tubular portion 173 so as to be movable in the axial direction. Main piston 161
The second oil chamber 182 is formed between the first auxiliary piston 162 and the first auxiliary piston 162. A cylindrical centering guide protrusion 162b that protrudes rearward is formed at the center of the rear end of the first auxiliary piston 162, and is fitted into the hole of the front wall of the second auxiliary piston 163 via a seal.

The second auxiliary piston 163 is formed in a bottomed cylindrical shape having a rear end opened, and a step portion 163a is formed in the front portion. The front outer peripheral seal portion is the fourth tubular portion 17
4 is fitted to the inner peripheral surface of 4 so as to be movable in the axial direction, and the front edge is the first
The auxiliary piston 162 is formed in the same shape as the rear edge of the auxiliary piston 162, and can come into contact with it. The rear end edge of the second auxiliary piston 163 can come into contact with the bottom wall surface (rear case 46) of the cylinder 164. A third oil chamber 183 is formed between the first auxiliary piston 162 and the second auxiliary piston 163. Further, the fourth oil chamber 1 is provided on the rear side of the second auxiliary piston 163.
84 are formed.

The oil chambers 181, 182, 183 and 184 have first, second, third and fourth ports 191, 19 respectively.
2, 193 and 194 are formed. Each port 19
1, 192, 193, and 194 are respectively connected to the engine governor valve 112 (FIG. 5) via various valve mechanisms described later, and the engine rotation governor hydraulic pressure is selectively supplied according to various running conditions.

The difference between the pressure receiving areas of the fourth oil chamber 184 and the first oil chamber 181 is that the governor hydraulic pressure is supplied from the second port 192 to the second oil chamber 182 in FIG. 14 and the main piston 161 moves forward. Main piston 161 when doing
Is equal to the pressure receiving area of.

Further, the difference between the forward thrust of the third oil chamber 183 and the rearward thrust of the first oil chamber 181 is determined by the second port 19
The pressure receiving area is set so as to be substantially equal to the forward thrust from the governor hydraulic pressure from 2.

According to the control ram A1, the speed ratio can be locked at three kinds of speed ratio positions, and automatic shift control can be performed according to the wheel side load (motor plunger side load). Hereinafter, each speed ratio lock position and the state during normal operation will be described.

[At the first lock position (maximum reduction ratio L1)-
FIG. 14] As shown by the solid line arrow in FIG. 14, when the governor hydraulic pressure is pumped only from the first port 191 to the first oil chamber 181, the other ports 192, 193, 194 are opened. Due to the hydraulic pressure in the chamber 181, all pistons 161,
162 and 163 are simultaneously pushed rearward to reach the maximum contracted state. That is, the swash plate holder 21 is locked at the maximum tilt position (maximum reduction ratio L1).

[During Normal Running] As shown by the broken line in FIG. 14, the engine governor hydraulic pressure is pumped only from the second port 192 to the second oil chamber 182, and the other ports 191, 193.
When 194 is opened, the main piston 161 attempts to move in the front-rear direction in proportion to the governor hydraulic pressure entering the second oil chamber 182, and the tilted position of the motor swash plate 19 is automatically controlled. Both auxiliary pistons 162 and 163 are biased rearward by the hydraulic oil in the second oil chamber 182 and locked at the rear positions.

The main piston 161 has the maximum stroke S
1 is moved forward to move the first stopper surface 201 of the cylinder 164.
When the step portion 161a of the main piston 161 comes into contact with,
The motor swash plate 19 is in a substantially upright state.

When the load on the wheel side (motor side) increases from this state, the pressure in the HST supply side oil passage B1 rises,
The motor plunger 18 tilts the motor swash plate 19 to increase the reduction ratio.

[Second Lock Position (Intermediate Speed Ratio L2 Close to Maximum Reduction Ratio) -FIG. 15] In FIG. 15, the first and fourth ports 191, 194 to the first and the first ports are indicated by solid line arrows. 4 Governor hydraulic pressure is sent to the oil chambers 181 and 184,
If the second and third ports 192 and 193 are opened,
Since the second auxiliary piston 163 has a large pressure receiving area,
It moves forward by the hydraulic pressure entering the fourth oil chamber 184 from the fourth port 194, and the step portion 163a of the second auxiliary piston 163 comes into contact with the third stopper surface 203 of the cylinder 164 and is locked.

On the other hand, the main piston 161 is pushed rearward by the governor hydraulic pressure entering the first oil chamber 181 from the first port 191, contacts the front end edge of the first auxiliary piston 162, and passes through the first auxiliary piston 162. And is locked by the second auxiliary piston 163 and locked at the speed ratio L2 in FIG.

The force for pushing the motor swash plate 19 to the upright side is
This is the difference between the forward thrust of the hydraulic oil in the fourth oil chamber 184 and the backward thrust of the first oil chamber 181. The fourth oil chamber 184
As described above, the difference in the pressure receiving area between the first oil chamber 181 and the first oil chamber 181 is that the pressure received by the main piston 161 when the governor hydraulic pressure is supplied from the second port 192 to the second oil chamber 182 and the main piston 161 moves forward. Since the area is equal to the area, the governor hydraulic pressure is supplied only from the first port 191 to the first oil chamber 181 from the state of FIG. 14 to the fourth port 194.
When the governor hydraulic pressure is added to the 3), the thrust characteristics are the same as in normal operation (when pumping only to the second port 192).
The two pistons 161, 162, 163 move forward, raise the swash plate 19 for the motor, and reach the state shown in FIG.
That is, the automatic shifting is performed with the same characteristics as in the normal operation, and finally the vehicle is locked in the state of the speed ratio L2 in FIG.

In the state of FIG. 15, when the vehicle load increases, for example, when the vehicle is going uphill, kickdown is required. However, since the hydraulic pressure of the motor plunger 18 of the HST circuit is high, the motor plunger 18 is inevitably high. When the pressure to the front of the motor increases and the force that tilts the motor swash plate 19 increases, the second auxiliary piston 163 is pushed back together with the main piston 161 and the first auxiliary piston 162, Downshifted. Even if the load increases, if the locked state is kept completely at the second lock position, the driving force will be overwhelmed by the resistance and the engine will stall. The correction is made by the movement of the piston 163. That is,
It is normally locked, but when the load increases, it can be naturally released to shift down.

[Third lock position (speed ratio L3 close to the top side) -FIG. 16] As shown by solid line arrows in FIG. 16, from the first and third ports 191, 193 to the first and third oil chambers. 18
Governor hydraulic pressure is sent to 1,183, and the remaining port 19
When the valves 2 and 194 are opened, from the locked state at the maximum tilt position in FIG. 14, the third oil chamber 18 is passed through the third port 193.
When the governor hydraulic pressure is applied to No. 3 to shift to the state of FIG. 16, the difference between the thrust to the front of the third oil chamber 183 and the thrust to the rear of the first oil chamber 181 is the normal traveling as described above. Since the pressure receiving area is set so as to be substantially equal to the forward thrust by the governor hydraulic pressure from the second port 192 at the time, the characteristics from the position of the maximum reduction ratio L1 to the state of FIG. The swash plate 19 for the motor can be changed by. When the load on the wheel side suddenly increases, the main piston 161 is moved rearward, the gear ratio is increased, and downshifting is performed, as in the case of normal operation.

In the hydraulic continuously variable transmission of the embodiment, control of automatic speed change, control of shift lock, control for preventing extra pressure from occurring in an extra portion when the direct coupling clutch is engaged, and the like are performed. The hydraulic circuit of the control system is shown in Figs.
It is shown in 0. The charging hydraulic pressure is used not only for driving the control system and for operating the direct coupling clutch, but also for replenishing the leakage of the HST circuit as described above.

In FIG. 5, as the solenoid valve for control,
The following five are mainly arranged, all of which are normally-closed solenoid valves. Drive mode switching pilot valve 120 Control ram switching pilot valve 105 First shift lock pilot valve 108 Second shift lock pilot valve 110 Neutral lock operating valve 126

5 to 10, the shift mechanism 10 is shown.
The drive gear 57 and the fan-shaped shift gear 60 are depicted as one member. Further, among the oil passages, the oil passage for pilot is shown by a broken line, and the oil passage under pressure is drawn thick.

In FIG. 11, the solenoids of the solenoid valves are connected to the following switches. The solenoid of the traveling mode selector pilot valve 120 is connected in series with a traveling mode selector switch 220 and a direct coupling clutch operation detection switch 221, and the traveling mode selector switch 22 is connected.
The positive terminal of 0 is connected to the power supply 204, and is also switchably connected to the economy mode contact Ec, the power mode contact Pa, and the speed ratio lock mode contact Lo.

Control ram switching pilot valve 105
A forward detection switch 205, a clutch valve actuation detection switch 206, and an accelerator pedal depression operation detection switch 207 are connected in parallel to the solenoid of FIG.
o is connected.

Each solenoid of the first shift lock pilot valve 108 and the second shift lock pilot valve 110 is
It is connected to each contact portion of the shift lock position changeover switch 208, and both shift lock pilot valves 108, 11 are connected.
A reverse detection switch 209 is connected between the solenoids 0, and the reverse detection switch 209 can be switched between forward and neutral detection contacts F and N and a reverse detection contact R.
The positive terminal of the shift lock position changeover switch 208 is connected to the lock contact Lo of the traveling mode changeover switch 220.

As described above, the brake operation detecting switch 226 is connected to the solenoid of the neutral lock operating valve 126.

In FIG. 5, the charging pump 100
A primary pressure regulating valve 121 and an engine governor valve 112 are connected to the primary pressure oil passage 101 connected to the discharge side of. 2 connected to the outlet port of the primary pressure regulating valve 121
A secondary pressure regulating valve 122 is connected to the secondary pressure oil passage 102, and the oil chamber of the neutral lock actuator 129 is connected to the secondary pressure oil passage 102 via the orifice 91 and the oil passage 106 as described above, and the throttle via the oil passage 107. The oil chamber of the quantity detecting cylinder 92 is connected. The moving rod of the throttle lever push-back cylinder 92 contacts the manual throttle lever 93, and when the rod is projected, the throttle lever 93 is pushed back to the idling state.

The secondary pressure oil passage 102 is connected to the oil passage 9 as described above.
4, through check valves 95b, 95a and oil passages 94b, 94a, respectively connected to the discharge side oil passage B1 and the return side oil passage B2 of the HST circuit. The clutch valve 72, the vehicle speed governor valve 76, the direct coupling clutch C1 and the like provided in the HST circuit are connected as described above.

The outlet port of the engine governor valve 112 has an oil passage 96 directly reaching the inlet port of the governor pressure switching valve 113, and an oil passage 98 reaching another inlet port of the governor pressure switching valve 113 via the orifice 97. , An oil passage 99 leading to the inlet port of the control ram switching valve 104 and the hydraulic oil chamber of the clutch valve 72. Further, the oil passage 98 is branched into an oil passage reaching the hydraulic oil chamber of the control ram switching valve 104 and an oil passage 118 reaching the inlet port of the control ram switching pilot valve 105.

The moving rod of the engine governor valve 112 is in contact with one end of the governor lever Y. As the engine speed increases, the mechanical governor 117
The other end side of the governor lever Y is pushed to rotate in the direction of the arrow T1 by a predetermined angle, and the moving rod of the engine governor valve 112 is pushed in to open the valve, thereby gradually increasing the engine governor pressure.

At one end of the governor lever Y, the pressure increasing assist cylinder 1 is provided on the side opposite to the engine governor valve 112.
14 is arranged, and the mechanical governor 11 is provided at the other end.
A stopper cylinder 115 is arranged on the side opposite to 7. The oil chamber of the boosting assist cylinder 114 is connected to the outlet port of the governor pressure switching valve 113. The oil chamber of the stopper cylinder 115 is connected to the outlet port of the vehicle speed governor valve 76 via the signal oil passage 90. When the vehicle speed rises and the vehicle speed governor valve 76 opens, the rod projects and the governor lever Y moves. By suppressing the rotation in the T1 direction, the engine governor pressure is reduced.

The outlet-side first port 104a of the control ram switching valve 104 is connected to the first port 191 of the control ram A1 via the first oil passage 131, and the outlet-side second port 104b is controlled via the oil passage 132. Ram A
1 is connected to the second port 192. The first oil passage 1
31 is branched in the middle and connected to the inlet ports of the first and second shift lock switching valves 109 and 111, respectively, and further branched to the first and second via orifices 133 and 134 and oil passages 135 and 136. The second shift lock switching valves 109 and 111 communicate with the respective hydraulic oil chambers. The inlet ports of the first and second shift lock pilot valves 108 and 110 are connected to the oil passages 135 and 136, respectively. First and second shift lock pilot valves 108, 11
The 0 outlet port is connected to the tank.

The outlet port of the first shift lock switching valve 109 is connected to the fourth port 194 of the control ram A1 via the oil passage 138, and the outlet port of the second shift lock switching valve 111 is connected via the oil passage 139. Control ram A
1 is connected to the third port 193.

The primary pressure regulating valve 121 is provided with a pressure reducing actuator 123.
And the working oil chamber of the governor pressure switching valve 113 are connected to the secondary pressure oil passage 102 via the oil passage 140 and the orifice 141. The oil passage 140 includes the oil passage 14
Driving mode switching pilot valve that can connect 0 to tank 1
20 are connected.

A series of hydraulic control from the idle state of the engine to starting, acceleration, low speed and high speed running will be described.

FIG. 5 shows the state of the engine at idle rotation before starting. [Conditions] Engine rotation: Idle shift position: Neutral drive mode: Economy speed ratio lock: Release (normal speed change mode)

[State of Solenoid Valve] Travel mode switching pilot valve 120 ON (economy mode) Control ram switching pilot valve 105 ON (neutral shift position) First shift lock pilot valve 108 OFF Second shift lock pilot valve 110 OFF Neutral lock Actuating valve 126 on (neutral lock)

The pressure oil from the charging pump 100 is
First, it enters the primary pressure oil passage 101 through the filter 103. The inside of the oil passage 101 is regulated to a primary pressure (1.3 MPa) by the primary pressure regulating valve 121, reaches the engine governor valve 112, and is temporarily shut off by the engine governor valve 112 in the closed state. .

Since the brake is not activated,
The neutral lock operating valve 126 is opened by moving the spool in the X2 direction, draining the oil passages 106 and 107,
The pressure is reduced. The pressure oil in the secondary pressure oil passage 102 is supplied to the oil passages 106 and 107 through the orifice 91, but is drained through the neutral lock operating valve 126 as described above, and the oil passages 106 and 2 are also drained. Since the orifice 91 is arranged between the secondary pressure oil passages 102, the pressure in the oil passages 106 and 107 is substantially zero. Therefore,
The oil chamber 69 of the neutral lock actuator 129 is decompressed to near 0, and the rod 66 of the actuator 129 projects by the force of the spring 67 as shown in FIG.
The lock mechanism 70 is engaged to lock the shift mechanism 10 in the neutral state. Further, the cylinder 92 in FIG. 5 is also in a non-operating state and is retracted.

The pressure oil in the secondary pressure oil passage 102 is supplied into the closed circuit of the HST through the oil passage 94.

Since the engine governor valve 112 is still closed, the governor hydraulic pressure is not supplied from the valve 112, so that there is no supply of pilot hydraulic oil from the oil passage 99 to the clutch valve 72, and the clutch valve 72 is is open. As a result, the oil passages B1 and B2 of the HST circuit (oil chamber 24
A and 24b) are short-circuited, and the motor M is not operating. Further, since the vehicle speed is 0, the centrifugal vehicle speed governor valve 76 is closed, so that the direct coupling clutch C1 is not supplied with the pressure oil and the signal oil passage 90 is not supplied with the signal pressure oil.

FIG. 6 shows a state in which the vehicle is accelerating to the middle and low speed ranges after starting. Engine: Medium, high vehicle speed: Medium, low speed (direct coupling clutch off) Shift position: Forward drive mode: Economy speed ratio lock: Release (normal speed change mode)

[State of Solenoid Valve] Travel mode switching pilot valve 120 ON (economy mode) Control ram switching pilot valve 105 OFF First shift lock pilot valve 108 OFF Second shift lock pilot valve 110 OFF Neutral lock operating valve 126 OFF ( (Only when the brake is activated)

When the brake is depressed from the state of FIG. 5,
The neutral lock actuating valve 126 moves in the X1 direction and is closed, and the pressure in the oil passage 106 rises. As a result, the pressure in the oil chamber 69 of the actuator 129 rises, and the spring 6
The rod 66 is retracted against 7 and the drive gear 5 of FIG.
Remove the rod 66 from the lock groove 70 of 7. This allows
The neutral lock state of the shift mechanism 10 is released, and the shift mechanism 10 can be shifted forward or backward.

When the neutral lock actuating valve 126 of FIG. 5 is closed by stepping on the brake, the oil passage 107 rises as the oil passage 106 rises, so that the rod of the cylinder 92 projects and the throttle lever 93 moves. Hold the throttle closed to keep it idle. The above-mentioned throttle lever 93 is a manual type and is normally stopped and held at a predetermined idle position at the time of forward or reverse shift. However, if the accelerator position is at a middle opening or a high opening, input rotation is performed. Since the speed is too high and a large start shock is involved, the throttle lever 93 is automatically returned to the idle position by stepping on the brake.

After shifting the shift mechanism to the forward or reverse position while maintaining the brake depression state, the rod 66 of the actuator 129 is disengaged from the lock groove 70, and the circumferential surface 57b of the drive gear 57 is moved as shown in FIG. Since I was riding on, the lock remains unlocked even when the brake is released.

By increasing the opening degree of the throttle lever 93 and increasing the engine speed, the protrusion amount of the rod 117a of the mechanical governor 117 shown in FIG. 6 is increased, the governor lever Y is rotated in the T1 direction, and the engine governor valve is rotated. 11
By pressing the second rod 112a, the hydraulic oil proportional to the square of the engine rotation speed is generated.
It is discharged from 2.

This hydraulic oil (governor hydraulic pressure) enters the assist cylinder 114 through the orifice 97, the oil passage 98 and the governor pressure switching valve 113, and the governor lever Y.
Amplifies the rotational power of the T1 direction. Therefore, the output is increased more than pressing with the mechanical governor 117. For example, normally, when a governor hydraulic pressure of 0.6 MPa is output at 3000 rpm, assisting by the assist cylinder 114 raises the pressure to approximately 1.2 MPa. That is, the proportional constant (gain) with respect to the rotation is increased.

Since the traveling mode is economy, the governor hydraulic pressure is increased and supplied to each port 192 of the control ram A1 even when the engine speed is low.
It will push the piston 161 (Fig. 14) of
The swash plate 19 for the motor easily stands upright, and the reduction ratio quickly changes to the top side. That is, by matching the low engine rotation speed and changing the gear ratio to the top side, the fuel consumption is reduced although the driving force is reduced.

The governor hydraulic pressure is also supplied to the hydraulic oil chamber of the clutch valve 72 through the oil passage 99, and the clutch valve 72 is closed when the engine reaches a predetermined rotation speed. As a result, pump pressure rises in the supply-side oil passage B1, the HST circuit functions, and the motor M rotates.

In the initial process of changing from FIG. 5 to FIG. 6, the governor hydraulic pressure is low and the control ram switching valve 1
When 04 is still pushed by the spring 146 as shown in FIG. 5, the governor hydraulic pressure from the oil passage 99 is the first oil passage 131.
Then, the pressure is fed to the first port 191 of the control ram A1. As a result, the motor swash plate 19 is locked at the maximum tilt position (maximum reduction ratio). That is, when the vehicle starts, the maximum reduction ratio is achieved.

However, immediately after the above state, the rotation of the engine rises and the governor hydraulic pressure rises.
Immediately, the control ram switching valve 104 changes as shown in FIG. Then, pressure oil is supplied from the oil passage 99 to the second port 192 of the control ram A1 through the oil passage 132, while the pressure oil to 191 is drained to the first port.

Since the brake is of course released when the vehicle starts moving, the cylinder 92 is in the released state and the throttle lever 93 can be freely operated.

FIG. 7 shows that the vehicle speed increases and the vehicle speed governor valve 7 increases.
6 shows the state where 6 is opened and the direct coupling clutch C1 is engaged. [Conditions] Engine speed: Medium, high speed Vehicle speed: High speed (direct clutch on) Shift position: Forward drive mode: Economy mode Speed ratio lock: Release

[State of solenoid valve] Travel mode switching pilot valve 120 OFF (economy mode) Control ram switching pilot valve 105 ON First shift lock pilot valve 108 OFF Second shift lock pilot valve 110 OFF Neutral lock operating valve 126 ON

As the vehicle speed increases, the centrifugal type vehicle speed governor valve 76 opens (communicates), the hydraulic oil 78 of the direct coupling clutch C1 is supplied with the pressure oil of the HST circuit, the direct coupling clutch C1 is connected, and the signal is transmitted. The working oil (vehicle speed governor hydraulic pressure) is supplied to the working oil chamber of the stopper cylinder 115 via the working oil passage 90, and the rod of the cylinder 115 is projected to press the governor lever Y in the T2 direction. And close the governor hydraulic pressure to zero. That is, the operation of the governor hydraulic pressure at the time of direct connection is stopped. Therefore, the governor hydraulic pressure is not supplied from the outlet port of the engine governor valve 112, and the governor hydraulic pressure is not supplied to the hydraulic oil chamber of the clutch valve 72 via the oil passage 99. Therefore, the clutch valve 72 is opened and both oil passages of the HST circuit are opened. A short circuit occurs between them. As a result, in the direct connection operation state, even if the motor swash plate 19 is slightly inclined, the rise of the hydraulic pressure is released through the clutch valve 72.
It prevents the increase of load and prevents the loss of energy.

When the centrifugal vehicle speed governor valve 76 is opened as described above and pressure oil is supplied to the signal oil passage 90, the direct coupling clutch operation detection switch (pressure switch) 221 shown in FIG. As shown in FIG. 7, the driving mode switching pilot valve 120 is turned off. As a result, the oil passage 1
The pressure rises in 40, and the rod of the actuator 123 is projected to lower the primary set pressure of the primary pressure regulating valve 121 from, for example, 1.3 MPa to 0.8 MPa. Therefore, the burden on the charging pump 100 is reduced,
Energy saving is achieved. At the same time, pressure oil is also supplied to the hydraulic oil chamber of the governor pressure switching valve 113, and the valve 1
13 is switched to a straight section (upper section in the figure), and the assist cylinder 114 is in a non-pressurized state (a state in which the governor lever is not pressurized).

In the state of FIG. 7, the engine governor valve 11
Since governor hydraulic pressure is not supplied from No. 2, any port 191, 192, 19 is connected to the control ram A1.
No pressure oil is supplied to 3,194. Therefore, the shift thrust of the control ram A1 disappears, and the motor swash plate 19 is simply pressed by the motor plunger 18.
Maintained in a substantially vertical position.

FIG. 8 shows a state in which the speed ratio of the control ram A1 is locked at the maximum reduction ratio L1 and running work is performed at a constant vehicle speed. [Conditions] Engine rotation: Medium, high speed Vehicle speed: Medium, low speed (direct coupling clutch off) Shift position: Forward drive mode: Speed ratio lock mode Speed ratio lock: L1 (maximum reduction ratio)

[State of solenoid valve] Travel mode switching pilot valve 120 OFF (speed ratio lock mode) Control ram switching pilot valve 105 ON First shift lock pilot valve 108 ON (L
1 designation) 2nd shift lock pilot valve 110 ON (L
1 designation) Neutral locking valve 126 ON

As a precondition, since the vehicle speed is medium or low, the direct coupling clutch C1 is off, the governor valve 112 is open, and the governor valve 112 supplies the governor hydraulic pressure. The control ram changeover pilot valve 105 is a running mode changeover switch 220 (FIG. 11).
Is always on due to the speed ratio lock mode position, the pressure oil in the oil passages 116 and 118 is drained, and the pressure does not rise in the oil passage portion on the downstream side of the orifice 97. Therefore, the pilot pressure does not act on the hydraulic oil chamber of the control ram switching valve 104, and the governor hydraulic pressure discharged from the governor valve 12 passes from the oil passage 99 to the control ram switching valve 104 and from the oil passage 131 to the control ram. A1 first port 191, first shift lock switching valve 109 and second shift lock switching valve 1
Pumped to 11 inlet ports.

However, since the first and second shift lock pilot valves 108 and 110 for the pilots of the first and second shift lock change valves 109 and 111 are open, both shift lock change valves 109 and 111 are opened. Is closed and, eventually, the oil passage 1 is connected to the control ram A1.
Pressure oil is supplied from 31 only to the first port 191. As a result, the control ram is locked at the maximum reduction ratio L1 as shown in FIG.

As described above, the lock operation is effective only when the direct coupling clutch C1 is turned off at the time of traveling in the middle and low speeds. On the other hand, the locking operation is performed at the time of high speed traveling.
When is connected, the lock does not work no matter how much the lock operation is performed in order to prevent a sudden decrease in vehicle speed. That is, as shown in FIG. 7, when the direct coupling clutch is engaged, the first and second shift lock pilot valves 108, 11
Even if 0 is opened and locked at the maximum reduction ratio L1, the governor valve 112 does not supply the governor hydraulic pressure, so the state as shown in FIG. 8 does not occur and the pressure is not applied to the first port 191 of the control ram A1. No oil supply and no lock. That is, even if the shift lock operation is performed during high-speed traveling, the lock is not performed, and abrupt shift down can be avoided.

When the vehicle speed decreases and the direct coupling clutch C1 is released, the governor valve 112 opens and the governor hydraulic pressure is supplied. Therefore, when the shift lock operation is performed as shown in FIG. 1 port 191
The pressure oil is supplied to and locked at the maximum reduction ratio 1.

FIG. 9 shows a state in which the speed ratio of the control ram A1 is locked by the second reduction ratio L2. [Conditions] Engine rotation: Medium, high speed Vehicle speed: Medium, low speed (direct coupling clutch off) Shift position: Forward drive mode: Speed ratio lock mode Speed ratio lock: L2 (intermediate reduction ratio)

[State of solenoid valve] Travel mode switching pilot valve 120 OFF (speed ratio lock mode) Control ram switching pilot valve 105 ON First shift lock pilot valve 108 OFF (L
2 designation) 2nd shift lock pilot valve 110 ON (L
2 designation) Neutral lock valve 126 on

The state is the same as the state of FIG. 8 except that the first shift lock pilot valve 108 is closed. The governor hydraulic pressure discharged from the governor valve 112 passes from the oil passage 99 to the control ram switching valve 104, and from the oil passage 131 to the first port 191 of the control ram A1 and the inlet port of the first shift lock switching valve 109. It is supplied to the inlet port of the second shift lock switching valve 111.

Since the second shift lock pilot valve 110 for the pilot of the second shift lock switching valve 111 is open, the second shift lock switching valve 111 is in the closed state. Oil passage 1
At the same time as the pressure oil is supplied from 31 to the first port 191, the pressure oil is supplied to the fourth port 194 via the first shift lock switching valve 109 in the ON state, and the second reduction ratio is set as shown in FIG. Locked to L2.

FIG. 10 shows a state in which the speed ratio is locked by the third speed reduction ratio L3. [Conditions] Engine rotation: Medium, high speed Vehicle speed: Medium, low speed (direct coupling clutch off) Shift position: Forward drive mode: Speed ratio lock mode Speed ratio lock: L3

[State of solenoid valve] Travel mode switching pilot valve 120 OFF (speed ratio lock mode) Control ram switching pilot valve 105 ON First shift lock pilot valve 108 ON (L
3 designation) 2nd shift lock pilot valve 110 off (L
3 designation) Neutral lock actuating valve 126 ON

The state is the same as the state of FIG. 8 except that the second shift lock pilot valve 110 is closed. The governor hydraulic pressure discharged from the governor valve 112 passes from the oil passage 99 to the control ram switching valve 104, and from the oil passage 131 to the first port 191 of the control ram A1 and the inlet port of the first shift lock switching valve 109. It is supplied to the inlet port of the second shift lock switching valve 111.

Since the first shift lock pilot valve 108 for the pilot of the first shift lock switching valve 109 is opened, the first shift lock switching valve 109 is in the closed state, so that the control ram A1 is Oil passage 1
At the same time as the pressure oil is supplied from 31 to the first port 191, the pressure oil is supplied to the third port 193 via the second shift lock switching valve 111 in the ON state, and the third reduction ratio is set as shown in FIG. Locked to L3.

[Brief description of drawings]

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

2 is an enlarged cross-sectional view taken along the line II-II of FIG.

FIG. 3 is a sectional view taken along line III-III in FIG.

FIG. 4 is a partial vertical cross-sectional view of the same transmission as in FIG. 1, but with another section.

FIG. 5 is a hydraulic circuit diagram showing the entire control mechanism of the hydraulic continuously variable transmission according to the present invention and showing a state during idle rotation.

FIG. 6 is the same hydraulic circuit diagram as FIG. 5, showing the state at the start of normal operation.

7 is a hydraulic circuit diagram similar to FIG. 5, showing a normal operation state when a direct coupling clutch is connected.

FIG. 8 is a hydraulic circuit diagram similar to FIG. 5, showing a state in which the speed ratio is locked at the maximum reduction ratio.

FIG. 9 is the same hydraulic circuit diagram as FIG. 5, showing a state in which the speed ratio is locked at the second reduction ratio.

10 is a hydraulic circuit diagram similar to FIG. 5, showing a state in which the speed ratio is locked at a third reduction ratio.

11 is a wiring diagram of various solenoid valves used in the control mechanism of FIG.

FIG. 12 is an enlarged vertical cross-sectional view showing the state of the support shaft portion of FIG. 1 during engine brake.

FIG. 13 is an enlarged vertical cross-sectional view showing the state of the support shaft portion of FIG. 1 when the engine is driven.

FIG. 14 is a longitudinal sectional view showing a hydraulic actuator (control ram) for gear shift control, showing a state in which the speed ratio is locked at the maximum reduction ratio.

FIG. 15 is a hydraulic actuator for shift control, comprising:
FIG. 6 is a hydraulic circuit diagram showing a state in which a speed ratio is locked at a second reduction ratio.

FIG. 16 is a hydraulic actuator for shift control,
FIG. 6 is a hydraulic circuit diagram showing a state in which a speed ratio is locked at a third reduction ratio.

FIG. 17 is an enlarged view of the rear surface of the shift mechanism, showing the forward shift state in a vertical section.

FIG. 18 is an enlarged view of the rear surface of the shift mechanism, showing the neutral shift state.

FIG. 19 is an enlarged longitudinal rear view showing a shift mechanism in a reverse shift state.

FIG. 20 is a vertical sectional view of a conventional example.

[Explanation of symbols]

 1 Input Shaft 3 Valve Body 4 Motor Cylinder Block 7 Bearing 18 Motor Plunger 23 Rear Cover 24 Oil Chamber 24a Outer Oil Chamber 24b Inner Oil Chamber 25 Distribution Ring Support Shaft 25a Eccentric Head 25b Bearing 25c Projector 29 Hydraulic Distribution Ring 46 Rear case (stationary case member) 48 Bearing 49 Circular ring 49a Step portion 49b Fitting surface 87 Damper oil chamber P Hydraulic pump M Hydraulic motor B1 Supply side oil passage B2 Return side oil passage

Claims (3)

[Claims] 1. A fixed displacement hydraulic pump (P), and a supply-side oil passage (B1) for each of a discharge portion and a suction portion of the hydraulic pump (P).
And a return side oil passage (B2) to form a closed hydraulic circuit, and a variable displacement type swash plate hydraulic motor (M), and a motor cylinder block ( 4) On the rear side, a rear cover (23) is integrally connected in the rotational direction and the axial direction via the valve body (3) and is rotatably supported by the stationary case member via the bearing (7). , Which is rotatably slidably arranged on the rear surface of the valve body and becomes a part of the supply side oil passage (B1) inside the rear cover and becomes an inner oil chamber (24b) and a return side oil passage (B2). Outer oil chamber (24
a) and a hydraulic distribution ring (29) which is partitioned into a) and a front end eccentric head (25a) which supports the hydraulic distribution ring (29) so that it can be pressed forward and axially movable, and protrudes from the inside of the rear cover.
In a hydraulic continuously variable transmission equipped with a support shaft (25) connected to (0), the support shaft (25) has an outward flange-shaped bearing (the front surface faces the outer oil chamber (24a)). 25b) forming the bearing
A ring (49) between the (25b) and the inner peripheral support hole of the rear cover (23).
And a fitting surface (49b) is formed on the inner peripheral surface of the ring (49) in the first half portion via the step (49a), and the fitting surface (49b) has a support portion (25b). The outer peripheral surface is fitted so as to be relatively movable in the axial direction, and the ring
(49) The rear end surface is locked to the stationary case member on the rear side directly or via the cushioning material (55), and the outer peripheral surface of the ring (49) is rotated to the inner peripheral support hole of the rear cover (23). A hydraulic continuously variable transmission characterized by being supported by a bearing (48) so as to be freely movable relative to each other in the axial direction. 2. The hydraulic continuously variable transmission according to claim 1, wherein the bearing (25b) is provided between the rear surface of the bearing (25b) and the ring (49).
A hydraulic continuously variable transmission characterized in that a damper oil chamber (87) having a variable volume is formed by relative axial movement between the ring and the ring (49). 3. The hydraulic continuously variable transmission according to claim 1, wherein a substantially central portion of the front surface of the head of the support shaft (25) presses the substantial pressure center of the distribution ring (29) forward. ) Is formed, the hydraulic continuously variable transmission.