JPH11223175A - Plunger type hydraulic unit - Google Patents

Abstract

(57) [Problem] To maintain a fitted state between a slipper and a plunger main body even when a cylinder rotates at high speed with a low oil pressure in a cylinder hole. SOLUTION: A plunger 200 slidably inserted into a cylinder hole is provided with a slipper 205 formed in a bowl shape.
And a plunger body 20 slidably fitted to the slipper
1 and a connecting member 210 for connecting the slipper 205 and the plunger body 201. First axial force F pressing plunger body by hydraulic pressure in cylinder bore
1, a second axial force F2 for separating the plunger body by the hydraulic pressure of the sliding contact surface with the slipper, and a third axial force F3 for separating the plunger body by centrifugal force applied to the slipper when the cylinder rotates, F1> (F2 + F3) ) Is provided to control the oil pressure in the cylinder hole so as to satisfy the relationship expressed by (1).

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to a plunger type hydraulic unit comprising a swash plate plunger type hydraulic pump, a motor and the like.

[0002]

2. Description of the Related Art In a swash plate plunger type hydraulic pump or motor, for example, a plurality of cylinder holes are formed in a rotatable cylinder in an annular arrangement surrounding a rotation shaft thereof, and each of the cylinder holes is slidable in the axial direction. The plunger is fitted with a cylindrical plunger, and a slipper attached to the end of the plunger is slid on the swash plate. In the case of a hydraulic pump, the slipper slides on the swash plate by rotating the cylinder, and the plunger reciprocates in the cylinder hole to discharge and suck oil. Supplies and discharges oil into and from a cylinder hole, thereby moving a slipper on a swash plate and rotating a cylinder.

[0003] Such a swash plate plunger type hydraulic pump,
In the motor, the slipper rotates with the plunger about the rotation axis of the cylinder while slidingly contacting the swash plate. Therefore, the slipper needs to be connected to the end of the plunger so as to swing freely. Regarding the structure of such a connecting part,
Conventionally, various proposals have been made.
The structures disclosed in JP-A-8-189377, JP-B-60-3350, and U.S. Pat. No. 4,454,802 are known.

In these, a slipper formed in a bowl shape having a convex spherical outer peripheral surface and a concave spherical inner peripheral surface, and having a ring-shaped slipper surface at an axial end, and an axial one end. A plunger main body formed into a hollow cylindrical shape having a spherical concave surface slidably fitted to the convex spherical outer peripheral surface and having an inside opening at the other end in the axial direction; A connecting member extending in the axial direction from the portion and disposed through one end and the slipper, and connecting the slipper and the plunger body in a state where the convex spherical outer peripheral surface and the spherical concave surface are slidably fitted. A disclosed plunger assembly is disclosed.

[0005]

In the case where the slipper and the plunger body are connected via such a connecting member, a slight gap is provided between the slipper and the plunger body in order to ensure the slidability between the slipper and the plunger body. Clearance is provided.
However, when such a clearance is provided, when the cylinder rotates at a high speed, the centrifugal force applied to the slipper acts in a direction to separate the slipper from the plunger main body, and the convex spherical outer peripheral surface of the slipper and the spherical concave surface of the plunger main body. It has been found that a gap may be formed when the fitting portions are separated from each other, and a phenomenon that oil leaks through the gap may occur. Normally, the oil pressure in the cylinder hole acts to press the plunger body toward the slipper, so such a problem of oil leakage does not occur when the oil pressure in the cylinder hole is high. It was also found that such a problem occurs when the hydraulic pressure in the inside is low and the cylinder rotates at high speed.

For example, a swash plate plunger type hydraulic pump and a motor constitute a hydraulic continuously variable transmission mechanism, and a mechanical power transmission mechanism is provided in parallel with the hydraulic continuously variable transmission mechanism to constitute a continuously variable transmission. 2. Description of the Related Art There is known a continuously variable transmission provided with a lock-up mechanism for transmitting power only by a transmission mechanism. In such a continuously variable transmission, when the lock-up mechanism is operated, power transmission by the hydraulic continuously variable transmission mechanism is not performed, so that the hydraulic pressure between the hydraulic pump and the motor decreases, and the hydraulic pump and the motor cylinder A high-speed rotation state occurs. In this case, the slipper separates from the plunger body due to the centrifugal force applied to the slipper, and the problem of oil leakage as described above tends to occur.

Incidentally, Japanese Utility Model Laid-Open No. 58-189377,
In the structure disclosed in Japanese Utility Model Publication No. Sho 60-3350, a spring is provided to pull the slipper toward the plunger main body via the connecting member, so that the slipper is hard to separate from the plunger main body. However, while the centrifugal force acting on the slipper changes according to the rotation speed of the cylinder, the spring force is a constant value. For this reason,
To prevent oil leakage as described above, it is necessary to set a spring force so that the slipper does not separate when the maximum centrifugal force acts, and the spring force becomes too large, and the slipper and the plunger body However, there is a problem that the slidability of the sliding member is deteriorated and the sliding portion is worn.

The present invention has been made in view of such a problem.
Provided is a plunger-type hydraulic unit in which the slipper and the plunger main body are maintained in a fitted state even when the cylinder rotates at a high speed with a low hydraulic pressure in the cylinder hole, and the problem of oil leakage does not occur. The purpose is to:

[0009]

In order to achieve the above object, according to the present invention, there is provided a cylinder which rotates about a rotation axis, and a cylinder formed in an annular arrangement surrounding the rotation axis of the cylinder and extending in the axial direction. A plunger-type hydraulic unit includes a plunger slidably inserted into the hole and a swash plate with which the tip of the plunger abuts slidably. The plunger is formed into a bowl shape having a convex spherical outer peripheral surface and a concave spherical inner peripheral surface, and has a ring-shaped slipper surface slidingly in contact with a swash plate at an axial end portion; A plunger body having a spherical concave surface formed slidably fitted with a convex spherical outer peripheral surface at one end, and a head slidingly in contact with the concave spherical inner peripheral surface at one axial end,
A leg integrally connected to the head and extending to the other end in the axial direction penetrates the slipper, protrudes into the plunger body, and includes a connecting member fixedly connected to the plunger body. Then, the first axial force F1 for pressing the plunger body toward the slipper by the oil pressure in the cylinder hole and the second force acting to separate the plunger body from the slipper by the oil pressure acting in the bowl-shaped internal space of the slipper. An axial force F2 and a third axial force F3 acting to separate the plunger body from the slipper by centrifugal force applied to the slipper when the cylinder rotates in the plunger hydraulic unit are expressed by the following equation (1). Hydraulic pressure control means for controlling the hydraulic pressure in the cylinder hole so as to satisfy the following.

[0010]

F1> (F2 + F3) (1)

As long as the relation of the expression (1) is satisfied, even if a centrifugal force acts on the slipper due to the rotation of the cylinder, the slipper and the plunger main body are always kept in contact by the oil pressure in the cylinder hole. Thus, there is no possibility that a gap is formed between the two and oil leaks. In addition, by the hydraulic control by the hydraulic control means, the force for pressing the slipper against the plunger body can always be set to be constant or optimal, and the slidability between the two can be ensured, and the wear between the two can be effectively reduced. It is possible to reliably prevent the two from separating from each other.

In the above construction, the leg of the connecting member is connected to the plunger body so as to be movable in the axial direction.
Further, a plunger may be provided in the plunger body by providing an urging member for applying an axial force in a direction of pressing the head against the concave spherical inner peripheral surface to the connecting member. In this case,
A first axial force F1 for pressing the plunger body toward the slipper by the hydraulic pressure in the cylinder hole, and a second axial force F2 for separating the plunger body from the slipper by the hydraulic pressure acting in the bowl-shaped internal space of the slipper. When,
The third function acts to separate the plunger body from the slipper by centrifugal force applied to the slipper when the cylinder rotates.
A fourth force acting to press the slipper against the plunger body by the axial force F3 and the urging member via the connecting member.
Hydraulic control means for controlling the hydraulic pressure in the cylinder hole is used so that the axial force F4 satisfies the relationship represented by the following equation (2).

[0013]

(F1 + F4)> (F2 + F3) (2)

Since the fourth axial force F4 by the urging member (for example, a spring) always acts to press the slipper against the plunger body, even if the oil pressure in the cylinder hole is zero, both have a certain pressing force. In this case, there is no gap between the two. However, when a centrifugal force is applied to the slipper due to rotation of the cylinder, the hydraulic pressure in the cylinder hole is set by the hydraulic control means such that the centrifugal force opposes the force (F3) in the direction of separating the slipper from the plunger body. You. For this reason, the force of the urging member may be a small value, and the urging force of the urging member almost eliminates the possibility that the slipperiness between the slipper and the plunger main body is impaired, or that wear is generated between the two. Can be.

The present invention is particularly arranged such that at least one of the hydraulic stepless transmission is composed of a variable displacement swash plate plunger type hydraulic pump and a hydraulic motor, and the hydraulic stepless transmission is arranged in parallel. This is suitable for use in a plunger type hydraulic unit including a mechanical power transmission mechanism for performing mechanical power transmission and a lock-up mechanism for performing power transmission only by the mechanical power transmission mechanism. In this case, the hydraulic control means controls the hydraulic pressure such that the hydraulic pressure in the cylinder hole of the hydraulic pump or the hydraulic motor satisfies the above formula (1) or (2) when the lock-up mechanism is operated.

When the lock-up mechanism is operated, there is no power transmission via the hydraulic stepless transmission mechanism, so the hydraulic pressure in the closed circuit connecting the hydraulic pump and the motor decreases. Since the control means adjusts the hydraulic pressure in the closed circuit so as to satisfy the above equations (1) and (2), even if a centrifugal force acts on the slipper, a gap is generated between the slipper and the plunger body. Is reliably prevented.

[0017]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Preferred embodiments of a plunger type hydraulic unit according to the present invention will be described below. FIG. 1 shows a configuration of a continuously variable transmission T which constitutes a plunger type hydraulic unit according to the present invention. As can be seen from this figure, in this example, the transmission T is used as an FF drive system or an RR drive system. The continuously variable transmission T is a so-called hydromechanical continuously variable transmission, and is configured by combining a mechanical transmission unit 1 and a hydrostatic continuously variable transmission unit 2. The engine E that drives the continuously variable transmission T is disposed on the opposite side of the continuously variable transmission unit 2 with the mechanical transmission unit 1 interposed therebetween.

The mechanical transmission unit 1 includes a power split device 3,
The power transmission device 4 and the final reduction gear 5 are disposed in the first casing 1c. The power split device 3 includes the engine E
Transmission shaft 9 connected to the output shaft 7 through a torque damper 8, and a carrier 11 directly connected to the input shaft 9.
And a pump input shaft 10 facing the carrier 11 and extending coaxially with the input shaft 9. Further, a plurality of pinion shafts 12 are integrally provided on the carrier 11 at positions revolving around the input shaft 9.
2, a pair of large and small binion gears 1 integrally connected to each other.
3, 14 are rotatably arranged. Pump input shaft 1
At 0, a small-diameter sun gear 15 meshing with the large-diameter pinion 13 is connected and arranged, and a large-diameter sun gear 16 meshing with the small-diameter pinion 14 is rotatably arranged on the pump input shaft 10.

An intermediate drive gear 18 integrally connected to the large-diameter sun gear 16 is rotatably disposed on the pump input shaft 10, and an intermediate driven gear 19 meshing with the intermediate drive gear 18 is connected to the motor output shaft 17. It is arranged. The drive and driven gears 18, 1
9 constitutes a power transmission device 4. A final drive gear 20 is also connected to the motor output shaft 17, and a final driven gear 21 having a differential mechanism 22 built therein meshes with the final drive gear 20, thereby constituting the final reduction gear 5. Left and right wheel drive shafts 23 are provided from the differential mechanism 22.
L and 23R extend, and a driving force is transmitted to left and right wheels (not shown) via these drive shafts 23L and 23R to drive the vehicle.

The continuously variable transmission unit 2 comprises a variable swash plate plunger type hydraulic pump 24, a variable swash plate plunger type hydraulic motor 25, and a control panel 27 having a hydraulic closed circuit 26 for interconnecting them. . The control panel 27 is joined to the side of the mechanical transmission unit 1 and rotatably supports the pump input shaft 10 and the motor output shaft 17. For this reason, the control panel 27 is arranged between the mechanical transmission unit 1 and the hydraulic pump 24 and the hydraulic motor 25.

The hydraulic pump 24 has a pump cylinder 28 coaxially connected to the pump input shaft 10 and rotatably slidably contacting a valve plate surface 27a of a control panel 27.
A plurality of pump plungers 30 slidably fitted and inserted into a plurality of cylinder holes 29 formed in an annular arrangement surrounding the rotation axis of the pump cylinder 28; Shoe 3 attached
1 is a slidably movable pump swash plate 32 that slidably contacts
It is comprised including. That is, the hydraulic pump 24 is a variable displacement type swash plate plunger pump.

The pump swash plate 32 is orthogonal to the pump input shaft 10 (orthogonal to the plane of FIG. 1).
, And an upright position orthogonal to the pump input shaft 10 (in this case, a pump swash plate angle α = 0) as shown by a solid line in the figure, and a right and left swing inclination as shown by a two-dot chain line. The predetermined left and right maximum tilt positions (α = αR (MAX),
αF (MAX)), whereby the reciprocating stroke of the pump plunger 30 when the pump cylinder 28 is rotated changes. Even if the pump input shaft 10 is rotationally driven by the engine E in the upright position, the reciprocating stroke is zero and the pump discharge oil amount is zero, and the reciprocating stroke increases as the swing angle increases toward the maximum tilt position. As a result, the pump discharge amount increases.
Note that the discharge direction is reversed depending on whether the vehicle is tilted to the left or right, and the forward / rearward direction of the vehicle is determined by the tilt direction, as described later.

The hydraulic motor 25 is connected to the motor output shaft 17 coaxially and is provided with a motor cylinder 34 that is rotatably slidably mounted on the valve plate surface 27 a of the control panel 27.
A plurality of motor plungers 200 slidably fitted and inserted into a plurality of cylinder holes 35 formed in an annular array surrounding the rotation axis of the motor cylinder 34; And a variable slidable motor swash plate 38 with which a shoe 37 attached to the slidably abuts. That is, the hydraulic motor 25 is a variable displacement type swash plate plunger motor.

The motor swash plate 38 is orthogonal to the motor output shaft 17 (orthogonal to the plane of FIG. 1).
, And an upright position (motor swash plate angle β = 0) orthogonal to the motor output shaft 17 as shown by a two-dot chain line in the figure, and a rocking inclination to the right as shown by a solid line. It is possible to swing between a predetermined maximum tilt position (β = β (MAX)).

The hydraulic pump 24 and the hydraulic motor 25
Is coupled to the first casing in which the control panel 27 and the mechanical transmission unit 1 are accommodated.

[0026]

[Operation of Transmission] The operation of the continuously variable transmission having the above configuration will be described. When the engine E is driven, the engine output is transmitted from the output shaft 7 to the transmission via the torque damper 8, and the transmission input shaft 9 and the carrier 11 are driven to rotate at the same speed as the engine output shaft 7. When the carrier 11 is rotationally driven, the engine power is changed to the large and small diameter pinions 13,
The transmission is divided into small and large sun gears 15 and 16 via the transmission 14.

The division of the power is performed by the hydraulic pump 24.
The relationship between the swash plate angles α and β and the total speed ratio e of the transmission is shown in FIG. 2 and will be described with reference to FIG. The total speed ratio e
Is the ratio between the input and output rotational speeds of the transmission T, and is determined by equation (3). The vertical axis in FIG. 2 represents the pump and motor swash plate angles, with the plus side indicating rightward swing and the minus side indicating leftward swing. The horizontal axis represents the total speed ratio e, with the plus side indicating the speed ratio in the forward direction and the minus side indicating the speed ratio in the reverse direction. In the figure, the solid line indicates the pump swash plate angle, and the broken line indicates the motor swash plate angle.

[0028]

## EQU3 ## Total speed ratio e = (No) / (Ni) (3) where Ni: rotation speed of transmission input shaft 9 No: rotation speed of final driven gear 21

Here, the pump swash plate 32 is in the upright position (α =
0) and the motor swash plate 38 is at the maximum swing position (β = β (M
AX)), the pump cylinder 28 is freely rotatable and the discharge is zero, and the motor cylinder 34 is hydraulically locked and fixedly held because there is no oil supplied from the hydraulic pump 24. For this reason, the large-diameter sun gear 16
When the intermediate drive gear 18 is stationary, the small-diameter sun gear 15 (and the pump input shaft 10 and the pump cylinder 28 connected thereto) freely rotates in accordance with the rotation of the carrier 11, and the engine output is idly consumed and the left and right wheels are consumed. It is not transmitted to the drive shafts 23L and 23R. This state is the state shown by the vertical line a in FIG. 2, the total speed ratio e = 0, and the transmission T is in an infinite speed ratio.

However, this state is a state that is set when the shift lever (shift lever operated by the driver in the driver's seat) position is in the driving range such as the D and R ranges. When the shift lever is in the P or N range position, the motor swash plate angle β is set to 0, the motor cylinder 34 is also made free to rotate, and control for creating a neutral state is performed.

When the pump swash plate 32 is swung to the right in this state, hydraulic oil is started to be discharged from the hydraulic pump 24 in response to the swing, and the discharged hydraulic oil is fed to the hydraulic motor 25.
And the motor output shaft 17 (and the motor cylinder 34) of the hydraulic motor 25 is driven. When the rotational driving force at this time is transmitted from the motor output shaft 17 to the wheels via the left and right wheel drive shafts 23L and 23R, the wheels are driven in the forward direction. The rotation speed of the motor output shaft 17 increases as the pump swash plate angle α increases, and when this becomes the maximum swash plate angle αF (MAX), the state shown by the vertical line b in FIG. 2 is reached. Therefore, the overall speed ratio e increases from zero (vertical line a) to e1 (vertical line b). However, when the rotation of the motor output shaft 17 increases in this way, mechanical power transmission (via the power transmission device) via the small-diameter pinion 14, the large-diameter sun gear 16, the intermediate drive gear 18, and the intermediate driven gear 19 is performed. Simultaneously, the rotation of the pump input shaft 10 is correspondingly reduced.

When the pump swash plate angle reaches the maximum swash plate angle αF (MAX) (when the state of the vertical line b is reached), the motor swash plate angle β
Is swung so as to gradually decrease from the maximum angle. As a result, the rotation speed of the motor output shaft 17 further increases from the state of the vertical line b, and becomes maximum when the motor swash plate angle β becomes zero (upright position) (the state of the vertical line c, The total speed ratio becomes e2).

However, as described above, as the rotation speed of the motor output shaft 17 increases, the mechanical power transmission performed via the power transmission device 4 also increases, and the pump input shaft 10 The rotation of the power split device 3 and the power transmission device 4 is reduced so that the rotation of the pump input shaft 10 (and the pump cylinder 28) becomes zero when the motor swash plate angle β becomes zero (upright position). The gear ratio is set. When the motor swash plate angle β becomes zero (upright position), the motor cylinder 34 is in a freely rotatable state, and the pump cylinder 28 is in a hydraulic lock state and is held stationary. Therefore, in this state (the state of the vertical line c), only mechanical power transmission is performed by the power transmission device 4. In this state, the lock-up brake is actuated, as will be described later, to mechanically limit the operation of the continuously variable transmission unit 2 including the hydraulic pump and the motor, and to reliably transmit the power only by the power transmission device 4. .

On the other hand, when the pump swash plate 32 is swung to the left from the state of the vertical line a, the hydraulic oil is discharged from the hydraulic pump 24 in the hydraulic closed circuit 26 in the opposite direction. Therefore, the supply of the hydraulic oil drives the motor output shaft 17 (and the motor cylinder 34) of the hydraulic motor 25 in the opposite direction (reverse direction). The rotation speed of the motor output shaft 17 increases as the pump swash plate angle α increases,
When this becomes the maximum swash plate angle αR (MAX), the state shown by the vertical line d in FIG. 2 is obtained. Therefore, the total speed ratio e is zero (vertical line a)
To e3 (negative value).

[0035]

[Hydraulic Closed Circuit for Power Transmission] The hydraulic closed circuit 26 and the control hydraulic circuit thereof in the hydraulic type continuously variable transmission unit 2 will be described with reference to FIG. In this figure, the hydraulic pump 24 and the hydraulic motor 25 are symbolized and represented, and a first oil passage 26a connecting one port 24b of the hydraulic pump 24 and one port 25a of the hydraulic motor 25,
The other port 24a of the hydraulic pump 24 and the hydraulic motor 25
And a second oil passage 26b connecting the other port 25b to the other port 25b.

As described above, the pump swash plate 32 of the hydraulic pump 24 can swing right and left around the upright position (neutral position). The sucked hydraulic oil is discharged from the port 24b and supplied to the port 25a of the hydraulic motor 25 to drive the hydraulic motor 25 to rotate in the forward direction. After the driving, the hydraulic oil is discharged from the port 25b and supplied to the port 24a,
Circulated in the closed circuit 26. At this time, the hydraulic motor 25
If the wheels are driven by the rotational drive of the
The oil pressure in the oil passage 26a becomes high corresponding to the driving force, and the oil pressure in the second oil passage 26b becomes low. On the other hand, in the case where the rotation of the wheels is decelerated by the engine braking action, such as when coasting, the second
The oil pressure in the oil passage 26b becomes high corresponding to the engine braking force, and the oil pressure in the first oil passage 26a becomes low.

When the pump swash plate 32 is swung to the left (reverse direction), a flow of hydraulic oil is generated which is completely opposite to that described above, and the hydraulic motor is driven to rotate in the reverse direction. At this time, the hydraulic pressures in the first and second oil passages 26a and 26b also have the opposite relation to the above.

As described above, the driving force is transmitted between the hydraulic pump and the motor. The hydraulic oil circulated in the hydraulic closed circuit 26 generates heat according to the power transmission, the oil temperature increases, and Some of the oil accumulates and part of the oil leaks into the tank through gaps in the plunger. Therefore, part of the hydraulic oil in the closed hydraulic circuit is replaced to cool, replenish, and clean (flushing) the hydraulic oil. It is supposed to do. for that reason,
The hydraulic oil in the oil tank 41 is supplied to the suction filter 4
A charge pump 43 for supplying the first line 100 via the second line 100 is provided. The charge pump 43 is driven directly by the engine E, and has a discharge amount proportional to the engine speed.

From the charge pump 43 to the first line 100
Is adjusted by the regulator valve 60 to a predetermined line pressure PL. First line 10
0 is branched as shown in FIG.
A modulator valve 65 composed of a pressure reducing valve is connected to 0a, and regulates the oil pressure of the second line 101 connected to the output side of the modulator valve 65 to a predetermined modulation pressure Pm. The configuration of the modulator valve 65 is shown in FIG. 5, and is configured by disposing a spool 66 in a state of being biased leftward by a spring 67 in a housing. First branch line 100a connected to port 65a
The operating oil pressure of the control line 10
The pressure is reduced to a hydraulic pressure (constant hydraulic pressure) that balances with the hydraulic pressure of 1 to generate a modulated pressure Pm in the control line 101. It should be noted that the crosses in the figure mean that the connection to the drain is made.

The second line 101 is also branched into a plurality of lines.
The first linear solenoid valve 51 is connected to the branch second line 101a. The first linear solenoid valve 51 adjusts the modulation pressure Pm based on the control current (I),
As shown in FIG. 6, the control pressure PCL is proportional to the control current (I).
Is acted on the regulator valve 60 via the control line 110.

The structure of the regulator valve 60 is shown in FIG.
A spring 62 for urging the spool 61 to the left is provided in the housing. The housing is provided with a plurality of ports 60a to 60e as shown, the ports 60a and 60b are connected to the first line 100, the port 60c is connected to the charge line 130,
The port 60d is connected to the discharge line 131, and the port 60d
e is connected to the control line 110.

For this reason, the spool 61 has the internal communication hole 61.
line pressure P from the first line 100 to the left end through
L, the biasing force of the spring 62 and the control pressure P
Receive CL. Since the control pressure PCL can be regulated by the first linear solenoid valve 51 as described above, the line pressure PL is shown in FIG. Is controllable. In the regulator valve 60, the line pressure PL is regulated in this manner, but at this time, the excess oil first moves to the right by the spool 61 and the boat 60c
Side communicates with the port 60a side and flows to the charge line 130, and when there is excess oil, the port 60d side communicates with the port 60b side and is discharged to the discharge line 131.

As shown in FIG. 3, the charge line 130 includes charge supply lines 105a and 106a connected to the first and second oil passages 26a and 26b,
4a and 44b are connected. For this reason, the hydraulic oil flowing to the charge line 130
0 to one of the first and second oil passages 26a, 26b through one of the check valves 44a, 44b to the lower-pressure oil passage, thereby supplying hydraulic oil into the hydraulic closed circuit 26. Will be

The hydraulic oil discharged to the discharge line 131 is cooled by the oil cooler 151 and then returned to the tank 41 through the lubrication unit 152.

As shown in FIG. 3, charge discharge lines 105b and 106b are respectively connected to the first and second oil passages 26a and 26b constituting the hydraulic closed circuit 26.
The shuttle valve 70 is connected to these discharge lines 105b and 106b. The configuration of the shuttle valve 70 is shown in FIG. 8, and includes a spool 71 slidably disposed in the housing left and right, and a pair of springs 72 and 73 for urging the spool 71 from left and right. .

The hydraulic pressures of the two discharge lines 105b and 106b act on the right end and the left end of the spool 71, respectively, and when one of the first and second oil passages 26a and 26b becomes high pressure and the other becomes low pressure. , Spool 71
Is moved by being pressed by the pressing force on the high pressure side. As a result, the side of the ports 70a and 70b that receives the low-pressure hydraulic oil communicates with the port 70c, and the hydraulic oil in the low-pressure oil passage is discharged to the discharge line 132. As a result, an amount of hydraulic oil corresponding to the supply of hydraulic oil in the hydraulic closed circuit 26 is discharged, and cooling, flushing, and the like of the hydraulic oil are performed. However, the low pressure relief valve 74 is connected to the discharge line 132.
The oil pressure in the low-pressure side oil passage is set by the relief valve 74. The discharge line 132
The hydraulic oil discharged to the tank is also cooled by the oil cooler 151 and then returned to the tank 41 through the lubrication unit 152.

The hydraulic closed circuit 26 further includes high-pressure relief valves 75F, 75R for setting the maximum hydraulic pressure in the first and second oil passages 26a, 26b.
5c and 106c are connected and arranged. Since these high-pressure relief valves have the same structure, one of the valves 75F will be described as an example with reference to FIG.

This valve 75F has two independent spools 76 and 77 in the housing, and a first spool 76
Is urged rightward by the first spring 79a,
When moved to the right by this urging force, port 75a and port 7
5b, and when the port is moved to the left against this urging force, both ports 75a and 75b are communicated. The port 75a communicates with the first oil passage 26a via the relief line 105c,
Port 75b communicates with branch charge line 130a. The first spool 76 is provided with an orifice 76a, and the hydraulic pressure of the first oil passage 26a acting on the port 75a acts on the left and right sides of the first spool 76 in a steady state, and reduces the urging force of the first spring 79a. As a result, the first spool 76 moves rightward.

On the other hand, the second spool 77 is urged leftward by the second spring 79b. Second spring 7
9b urges the closing valve member 78 rightward at the right end, whereby the closing valve member 78
The communication passage 79c communicating with the insertion space a and having the orifice is closed. The port 75c of the second spool 77 facing the left end is connected to the second linear solenoid valve 52 via the control line 107. The second linear solenoid valve 52 regulates the line pressure PL from the branch line 100b and supplies a control oil pressure PCH corresponding to the control current to the control line 107.
The hydraulic pressure acting on the left end of the spool 77 can be controlled based on the control current of the second linear solenoid valve 52.

In this high pressure relief valve 75F,
The oil pressure in the first oil passage 26a acts on the closing valve member 78 from the right, and the urging force of the second spring 79b acts from the left. By the way, the second spring 79b has a left end portion in contact with the second spool 77 and the control line 107.
From the control oil pressure PCH from the above. As can be seen from this, if the control oil pressure PCH is controlled by the second linear solenoid valve 52, the force pressing the closing valve member 78 rightward can be controlled.

The closing valve member 78 closes the communication passage 79c by receiving such a rightward pressing force.
When the oil pressure in the first oil passage 26a acting on the first oil passage 9c becomes higher than this pressing force, the closing valve member 78 is moved to the left, and the communication passage 79c is moved.
To the drain. As a result, an oil flow is generated through the orifice 76a of the first spool 76, and a hydraulic pressure difference is generated between the left and right of the first spool 76, the first spool 76 is moved leftward, and the ports 75a and 75b communicate with each other. As a result,
The hydraulic oil in the first oil passage 26a is supplied to the branch charge line 130.
and is sent to the second oil passage 26b via the check valve 44b.

That is, when the oil pressure in the first oil passage 26a becomes equal to or higher than a predetermined pressure, the closing valve member 78 is opened, the first spool 76 moves leftward, and the hydraulic oil in the first oil passage 26a is turned to the low pressure side. The oil is discharged to the second oil passage 26b and keeps the oil pressure in the first oil passage 26a at a predetermined pressure or less. When the hydraulic pressure in the second oil passage 26b is high, the other high-pressure relief valve 75R
As a result, the oil is discharged to the first oil passage 26a on the low pressure side, and the inside of the second oil passage 26b is suppressed from becoming a predetermined pressure or more. As described above, the high pressure relief valves 75F and 75R prevent the oil pressure in the first and second oil passages 26a and 26b from becoming more than a predetermined pressure. And the third linear solenoid valve 5
It can be variably set by the current control of 2, 53. In addition,
In this example, as shown in FIG. 10, the high-pressure relief pressures (PHF, PHR) can be variably set in proportion to the control current (I).

The hydraulic pump 24 and the hydraulic motor 2 of this embodiment
A valve plate having a shape as shown in FIG. 11 is formed on the valve plate surface 27a of the control panel 27 of No. 5. The valve plates of the hydraulic pump and the hydraulic motor are different in size and the like, but have the same basic shape and the same role.
This will be described with reference to FIG.

In the hydraulic pump 24, the end surface of the pump cylinder 28 rotates in the direction of arrow A (clockwise) in the figure by driving the engine while slidingly contacting the valve plate 150. Here, when the pump swash plate 32 is tilted forward, the pump plunger 30
Is reciprocated in the cylinder bore. At this time, the pump plunger 30 is moved to the top dead center (T.
DC), is located at the bottom dead center (BDC) at the upper end, and in the process of moving from top dead center to bottom dead center in the left half, sucks oil, and from bottom dead center to top dead center in the right half During the moving stroke, oil is discharged.

For this reason, a semicircular first port 151 connected to the second oil passage 26b is formed in the left half of the valve plate 150, and a second semicircular second port connected to the first oil passage 26a is formed in the right half. A port 152 is formed. Here, on the inlet side in the rotation direction of the first port 151 and the inlet side in the rotation direction of the second port 152, each cylinder hole 29 starts to communicate with each port 151, 152 according to the rotation of the pump cylinder 28. Main notches 151a and 152a for suppressing a sudden change in hydraulic pressure are formed to extend toward the inlet side. Furthermore, these main notches 151a, 15
Subnotches 153 and 154 are formed independently as shown in FIG.

The sub notches 153 and 154 are separated from the main notches 151a and 152a and the ports 151 and 152 and are independent of each other. However, sub notch 15
3, 154 communicate with the suction and second ports 151, 152 via short-circuit oil passages 155a, 155b and 156a, 156b, respectively, as indicated by broken lines in the figure. However, the variable notch valve 8
0A and 80B are provided to communicate with the short-circuit oil passage.
Control the interruption.

The variable notch valves 80A and 80B have the same configuration and are configured as shown in FIG. 12 (in this figure, the variable notch valve 80A is shown as an example). The valve 80A is provided with a valve spool 81 slidably disposed in the housing left and right.
82 for urging the valve spool to the right and the valve spool 8
1 and a support spool 83 which is slidably fitted to the left end portion of the support spool 83. The housing is provided with a first port 80a communicating with the short-circuit oil passage 155a, a second port 80b communicating with the short-circuit oil passage 155b, and a third port 80c communicating with the control line 102. Valve spool 81
Is biased by the spring 82 and moves rightward as shown in the figure, the right end of the spool 81 blocks the first and second ports 80a and 80b, and when the valve spool 81 moves leftward, both ports 80a and 80a 80b communicates. The hydraulic pressure acting on the first port 80a is applied to the space 8 between the first port 80a and the support spool 83 through a small hole 81a in the spool 81.
4, the valve spool 81 does not receive a thrust force due to the hydraulic pressure.

The control line 102 is, as shown in FIG.
It is connected to the branch second line 101b via the orifice 45a and to the open / close control solenoid valve 45. The opening / closing control solenoid valve 45 is connected to the control line 102.
Is a valve that can be opened to the drain, and the inside of the control line 102 is reduced in pressure by opening the valve to the drain.
On the other hand, when the control line 102 is disconnected from the drain by the opening / closing control solenoid valve 45, the control line 10
2, a modulation pressure Pm from the second branch line 101b is generated.

When the control line 102 is maintained at a low pressure by the open / close control solenoid valve 45, the valve spool 81 is urged by the spring 82 to move rightward within the variable notch valve 80A to move short-circuit oil passages 155a, 155.
5b is cut off. Therefore, at this time, the sub notch 153 does not operate. On the other hand, when the modulation pressure Pm is generated in the control line 102 by the opening / closing control solenoid valve 45, the valve spool 81 is moved leftward by this oil pressure, the short-circuit oil passages 155a and 155b communicate, and the sub notch 153 becomes usable.

Here, the variable notch valve 80A
Has been described, the same control is simultaneously performed on the variable notch valve 80B. The hydraulic motor 25 is also provided with a similar sub-notch. The opening and closing control solenoid valve 46 variably sets the oil pressure in the control line 103 to control the operation of the variable notch valves 80C and 80D, thereby controlling the use of the sub-notch. Done.

[0061]

[Swash Plate Angle Control System] Next, a swing angle control system of the pump swash plate 32 and the motor swash plate 38 will be described with reference to FIG. This control is performed using the line pressure PL from the first branch line 100d and the modulation pressure Pm from the second branch line 101c. Note that the circled symbols a and b in FIGS. 3 and 13 are connected to each other.

In order to swing the pump swash plate 32,
A pair of servo cylinders 92a, 92b are provided as shown in the figure, and servo control lines 121, 12
2, the pump control valve 84 is connected. The valve 84 comprises a four-way valve, and distributes and supplies the line pressure PL from the first branch line 100d to the servo control lines 121 and 122 according to the position of the spool 85. The spool 85 is urged rightward by a spring 86, receives hydraulic pressure acting on a left end port 84a, and its position is set by a balance between the two.

That is, if the hydraulic pressure acting on the left end port 84a is controlled, the position of the spool 85 can be controlled, whereby the operation of the servo cylinders 92a and 92b can be controlled to control the angle of the pump swash plate 32. It is. Therefore, the pump control oil pressure PCP from the fourth linear solenoid valve 54 is supplied to the left end port 84a via the control line 111. The fourth linear solenoid valve 54 is a valve that regulates the modulation pressure Pm from the branch second line 101c based on the control current, creates a pump control oil pressure PCP proportional to the control current, and supplies this to the control line 111. . Therefore, the operation control of the pump control valve 84 is performed by controlling the energizing current of the fourth linear solenoid valve 54, and the angle of the pump swash plate 32 can be controlled.

The same applies to the angle control of the motor swash plate 38.
The motor control valve 87 is connected via servo control lines 123 and 124 to servo cylinder holes 95a and 95b in which a pair of servo cylinders 96a and 96b are slidably disposed.
Is connected. Like the pump control valve 84, the valve 87 can control the hydraulic pressure applied to the left end port 87a to control the position of the spool 88, thereby controlling the operation of the servo cylinders 96a and 96b and controlling the motor tilt. Angle control of the plate 38 is possible. Left end port 87
The motor control oil pressure PCM is supplied to a from the fifth linear solenoid valve 55 via the control line 112. Therefore, the operation control of the motor control valve 87 is performed by controlling the energizing current of the fifth linear solenoid valve 55, and the angle control of the motor swash plate 38 is possible.

The hydraulic pump 24 of this embodiment is further provided with a lock-up brake 93 for holding the pump cylinder 28 stationary.
Is provided. As described above, in the state where the pump swash plate angle α = αF (MAX) and the motor swash plate angle β = 0 on the forward side (state of the vertical line c in FIG. 2), theoretically (transmission loss When there is no rotation), the rotation of the pump cylinder 28 becomes zero and only mechanical power transmission is performed by the power transmission device 4. However, since there is actually a loss due to oil leakage or the like, the pump cylinder 28 slightly rotates, and the power transmission by the power transmission device 4 is reduced accordingly. Therefore, in such a case, the pump cylinder 28 is forcibly held stationary by the lock-up brake 93, and only the mechanical power transmission is performed to improve the transmission efficiency.

The lock-up brake 93 is composed of a wet multi-plate type brake, and operates the brake by receiving the pressing force of the piston 104. The lock-up oil pressure PLB for applying a pressing force to the piston 104 is generated by the sixth linear solenoid valve 56 and is supplied through the lock-up line 113. The sixth linear solenoid valve 56 is capable of producing a lock-up hydraulic pressure PLB proportional to the control current, whereby the lock-up brake 93 can be arbitrarily controlled from partial engagement to full engagement. .

When the lock-up brake 93 is completely engaged, power is transmitted only by the power transmission device 4 and no driving force acts on the hydraulic stepless transmission unit 2. descend. At this time,
The oil pressure in the oil pressure closed circuit 26 becomes the oil pressure set by the low pressure relief valve 74 connected via the shuttle valve 70. That is, when the lock-up brake 93 is engaged, the oil pressure in the oil pressure closed circuit 26 can be controlled by the low pressure relief valve 74. In the present invention, as will be described later, the hydraulic pressure in the hydraulic closed circuit 26 is controlled to maintain the fitting state between the slipper and the plunger body in the pump and the motor plunger. For this reason, the low-pressure relief valve 74 is constituted by an electromagnetic control valve, and its relief pressure can be variably set.

[0068]

[Plunger Configuration] In the continuously variable transmission T configured as described above, the hydraulic pump 24 constituting the continuously variable transmission unit 2
The configuration of the plunger assembly used for the hydraulic motor 25 will be described below. However, only the dimensions of the pump plunger 30 and the motor plunger 200 are different, and the configuration is the same. Therefore, the structure of the motor plunger 200 will be described in detail using the motor plunger 200 as an example.

This motor plunger 200 is shown in FIG. 14. The motor plunger 200 has a hollow cylindrical plunger body 201 with a bottom and a bowl-shaped slipper 205.
And a connecting member 2 for connecting these two members 201 and 205.
And 10. The plunger body 201 has a bottomed hollow cylindrical shape having an opening 201a at the end opposite to the swash plate (left side in the figure), and has a spherical concave surface 201b at the bottom, that is, at the right end in the figure. Further, a through hole 201c penetrating in the axial direction is formed at the bottom.

A slipper 205 is fitted on the right end of the plunger main body 201.
It is shown in FIG. The slipper 205 is a plunger body 2
01 outer peripheral surface 20 fitted with spherical concave surface 201b
5a and a concave spherical inner peripheral surface 2 concentric with the outer peripheral surface 205a.
05b, and has a ring-shaped slipper surface 205c at the right end in the axial direction. In addition, a through hole 207 that passes through the slipper 205 in the axial direction is formed to have a tapered shape that spreads to the left as shown in the figure.

The connecting member 210 includes a rod member 211 and a head member 215. The rod member 211 has a flange-shaped lid 211a at the left end of the rod-shaped main body. The lid 211a is attached so as to cover the opening 201a of the plunger main body 201, and is welded to the plunger main body 201. The main body of the rod member 211 is inserted into the cylindrical hollow space 202 of the plunger main body 201 and the through hole 2 is formed.
The slipper 205 is inserted into the through-hole 207 of the slipper 205 with the convex spherical outer surface 205a fitted to the spherical concave surface 201b, and protrudes into the concave spherical inner peripheral surface 205b of the slipper 205. The head member 215 has a spherical convex surface 215a that fits in sliding contact with the concave spherical inner peripheral surface 205b, and has a through hole that penetrates in the axial direction. The through-hole has an opening that expands in a tapered shape on the right side.

After the main body of the rod member 211 is inserted into the through hole 207 of the slipper 205 as described above, it is further inserted into the through hole of the head member 215. And the tip 211
b is caulked and fixed so that it can be spread out into the tapered opening of the through hole. Thereby, the rod member 21
1 and the head member 215, and as a result, the plunger body 201 and the slipper 205 are connected to the connecting member 21 in a state where the convex spherical outer peripheral surface 205a is fitted to the spherical concave surface 201b.
Linked by 0. The rod member 211 has a communication hole 212 penetrating therethrough in the axial direction, and the oil in the motor cylinder hole 35 is introduced as lubricating oil to the slipper 205 through the communication hole 212.

The motor plunger 2 thus configured
The balance of the forces acting on 00 will be described with reference to FIG. The motor plunger 200 receives the hydraulic pressure in the motor cylinder hole 35, and receives a first axial force F1 that presses the plunger body 201 toward the slipper 205 (presses rightward in the axial direction in the drawing). This force F1 is represented by F1 = Pc × Ac, where Ac is the cross-sectional area of the plunger body 201 at right angles to the axis, and Pc is the hydraulic pressure in the cylinder hole 35.

The hydraulic oil in the motor cylinder hole 35 is introduced to the slipper through the communication hole 212 of the rod member 211.
This oil pressure also acts between the convex spherical outer peripheral surface 205a of the slipper 205 and the spherical concave surface 201b of the plunger main body 201.
1 acts as a second axial force F2 acting to separate them.

Normally, the hydraulic pressure in the motor cylinder hole 35 is high, and the convex spherical outer surface 205 a
Since the fitting area of the plunger body 201 and the spherical concave surface 201b of the plunger body 201 is smaller than the pressure receiving area of the plunger body 201 in the cylinder hole 35, and the hydraulic pressure acting on this fitting portion is lower than the hydraulic pressure in the cylinder hole 35, F1 > F2
Is established, the plunger body 201 is pressed toward the slipper 205, and the motor plunger 200 is pressed against the sliding surface 38a of the motor swash plate 38. The difference between the two axial forces F1 and F2 increases as the oil pressure in the cylinder hole 35 increases, and the difference decreases as the oil pressure decreases.

Here, as described above, when the lock-up brake 93 is engaged, the hydraulic pressure in the motor cylinder hole 35 is low, so that the first and second axial forces F
The difference between 1 and F2 becomes smaller. In this state, the motor cylinder 34 is rotated at a high speed.
The centrifugal force Fc corresponding to this rotation acts on the rotation. Since the plunger body 201 is slid into the cylinder hole 35, the centrifugal force is received by the motor cylinder 34.

However, since the centrifugal force Fc acting on the slipper 205 is swingably attached to the tip (right end) of the plunger body 201, the convex spherical outer peripheral surface 205a of the slipper 205 and the spherical concave surface 201b of the plunger body 201 And acts on the outermost end point P in the fitting portion.
This centrifugal force Fc is determined by the rotation radius R of the plunger 200 with respect to the rotation center of the motor cylinder 34 and the slipper 205.
And the rotational angular velocity ω of the cylinder 34, and Fc = M · R · ω ² .

This centrifugal force Fc is a force acting in the outer peripheral direction at the position G of the center of gravity of the slipper 205, and at the point P, is a resultant force in the direction connecting the position G of the center of gravity and the point P,
The third axial force F3 acting to separate the plunger body 201 from the slipper 205 is generated by the axial component force F3.
(= Fc / tan θ). For this reason, in this state, the first axial force F1 acts as a force for pressing the plunger body 201 against the slipper 205, and the second axial force F1
2 and the third axial force F3 act as a force for separating the plunger body 201 from the slipper 205.

In this case, when the lock-up brake 93 is engaged, the oil pressure in the motor cylinder hole 35 is low, and the difference between the first and second axial forces F1 and F2 is small. Third axial force F due to centrifugal force
3 is added, the plunger body 201 is moved to the slipper 205.
Force (= F2 + F3) becomes greater than the first axial force F1, and the plunger body 201 is pulled out of the slipper 205.
The oil in the cylinder hole 35 may leak from the gap.

From the above, in the present apparatus,
When the lock-up brake 93 is in operation, the relief pressure setting of the low-pressure relief valve 74 is increased to control the hydraulic pressure in the cylinder hole 35 to be increased. That is, the low-pressure relief valve 74 constitutes the hydraulic control means in the claims. By increasing the oil pressure in the cylinder hole 35 in this manner, F1
> (F2 + F3) is always maintained, and the plunger main body 201 is reliably prevented from separating from the slipper 205. Specifically, since the centrifugal force Fc is obtained from the rotation speed of the motor as can be seen from the equation,
The relief pressure is controlled according to the rotation speed of the motor.
At the time of lock-up operation in which a small state of F1 is assumed in advance, the relief pressure setting is increased. For this reason, FIG.
As shown in (1), the control oil pressure from the low pressure control device 74b acts on the low pressure relief valve 74 to perform such relief pressure control.

Therefore, the relief pressure set by the low-pressure relief valve 74 is a pressure that increases according to the rotation speed of the motor cylinder 34. Further, the first axial force F1
Is too large than the resultant force of the second and third axial forces F2 and F3, the plunger body 201 and the slipper 20
5, the surface pressure of the sliding contact portion with the portion 5 becomes too high, and the slidability is reduced, and a problem of wear of this portion occurs. Therefore, the first
The relief pressure of the low-pressure relief valve 74 is controlled in accordance with the rotation speed of the cylinder 34 so that the difference between the axial force F1 and the resultant force (F2 + F3) of the second and third axial forces F2 and F3 becomes a predetermined value. I do.

Instead of motor plunger 200 having the above structure, motor plunger 30 having a structure as shown in FIG.
0 can also be used. This motor plunger 300
Is a bottomed hollow cylindrical plunger body 301, a bowl-shaped slipper 305, a connecting member 310 connecting these two members 301, 305, and an axially movably inserted and arranged inside the hollow space of the plunger body 301. Provided support member 32
And a compression spring 325 disposed between the plunger body 301 and the support member 320. The plunger main body 301 has a hollow cylindrical shape with a bottom opening at the end opposite to the swash plate, and has a spherical concave surface 301b formed at the bottom, that is, at the right end in the drawing, and further has an axial direction at the bottom. A penetrating through-hole 301c is formed.

A slipper 305 is fitted on the right end of the plunger main body 301, and this part is enlarged in FIG.
It is shown in FIG. The slipper 305 is the plunger body 3
01 convex spherical outer surface 30 fitted with the spherical concave surface 301b
5a and a concave spherical inner peripheral surface 3 concentric with the outer peripheral surface 305a.
05b, and has a ring-shaped slipper surface 305c at the right end in the axial direction. Further, a through hole 307 that passes through the slipper 305 in the axial direction is formed to have a tapered shape that spreads to the left as illustrated.

The connecting member 310 is integrally formed of a rod member 311 and a head member 315. A head member 315 is formed at the right end of the rod member 311, and the rod member 315 is in contact with the through-hole 307 of the slipper 305 and the plunger in a state where the head member 315 is in sliding contact with the concave spherical inner peripheral surface 305 b of the slipper 305. It extends through the through hole 301c of the main body 301 and extends into the hollow space of the plunger main body 301. Further, a through hole 321 formed in the support member 320 is provided.
The left end 313 is pushed out and is joined to the support member 320. Thus, the plunger body 301 and the slipper 305 are connected by the connection member 310 and the support member 320 in a state where the convex spherical outer peripheral surface 305a is fitted to the spherical concave surface 301b. A communication hole 312a and a communication groove 312b are formed in the rod member 311. The oil in the motor cylinder hole 35 is introduced as lubricating oil to the slipper 305 through the communication hole 312a and the communication groove 312b. The compression spring 325 is disposed in the hollow space of the plunger main body 301 before the support member 320 is inserted, and the fourth axial force F for urging the support member 320 to the left with respect to the plunger main body 301.
Add 4

The motor plunger 3 configured as described above
The balance of the forces acting on 00 will be described with reference to FIG. The motor plunger 300 receives a hydraulic pressure in the motor cylinder hole 35 and receives a first axial force F1 that presses the plunger body 301 toward the slipper 305 (presses rightward in the axial direction in the drawing).

The convex spherical outer peripheral surface 305a of the slipper 305
This hydraulic pressure also acts between the plunger body 301 and the spherical concave surface 301b of the plunger body 301, and this acts as a second axial force F2 acting to separate the plunger body 301 from the slipper 305.

The centrifugal force Fc acting on the slipper 305 when the motor cylinder 34 is rotated is, as in the case described above, the convex spherical outer peripheral surface 30 of the slipper 305.
It acts on the outermost end point P in the fitting portion between 5a and the spherical concave surface 301b of the plunger body 301. This centrifugal force Fc
Is the force acting in the outer peripheral direction at the position G of the center of gravity of the slipper 305, and at the point P, a resultant force in the direction connecting the position G of the center of gravity and the point P. The third axial force F3 (= Fc / tan θ) acting so as to separate 301 is obtained.

Further, the urging force F4 of the compression spring 325 is
Acts as a force to attract the slip 305 to the plunger body 301 via the connecting member 310. This force is referred to as a fourth axial force F4, which acts on the plunger in the right direction in the drawing.

In this case, the first and fourth axial forces F
1 and F4 are the forces pressing the plunger body 301 against the slipper 305, and the second and third axial forces F2 and F3
Is the force for separating the plunger body 301 from the slipper 305. Therefore, (F1 + F4) <(F2 + F3)
Then, the plunger body 301 may be separated from the slipper 305, and the oil in the cylinder hole 35 may leak from this gap.

For this reason, also in this device, when the lock-up brake 93 is operating, the relief pressure of the low-pressure relief valve 74 is set to be high, and the control to increase the oil pressure in the cylinder hole 35 is performed. F1 + F4)
> (F2 + F3) is always maintained, and the plunger body 301 is reliably prevented from separating from the slipper 305. In this case, since the force F3 corresponding to the centrifugal force can be dealt with by setting the relief pressure of the low-pressure relief valve 74, the urging force F4 of the compression spring 325 may be a minimum necessary value. The surface pressure of the sliding contact portion between the plunger body 301 and the slipper 305 does not become too high, and the slidability of the two can be ensured and the wear of this portion can be prevented. Note that the maximum value of the first axial force F1 is a hydraulic pressure that can be generated mechanically, and is a hydraulic pressure that is naturally determined from the swash plate angle, the rotation speed, the endurance pressure, and the like.

In the case of the motor plunger 300 shown in FIG. 16, since the support member 320 is inserted and disposed in the hollow space of the plunger body 301, the volume that becomes hollow, that is, the hydraulic oil in the cylinder hole 35 is reduced. The inflow volume is small. The oil flowing into the hollow space does not affect the amount of oil sucked and discharged due to the reciprocating motion of the plunger, and is affected by the compressibility due to the oil pressure in the cylinder hole 35, which causes a reduction in volumetric efficiency. Therefore, by reducing the hollow volume in this way, it is possible to prevent a decrease in volume efficiency.

[0092]

As described above, according to the present invention,
A first axial force F1 for pressing the plunger body toward the slipper by the hydraulic pressure in the cylinder hole, and a second axial force F2 for separating the plunger body from the slipper by the hydraulic pressure acting in the bowl-shaped internal space of the slipper. When,
And a third axial force F3 acting to separate the plunger body from the slipper by centrifugal force applied to the slipper when the cylinder rotates in the plunger-type hydraulic unit;
Formula F1> (F2 + F3) (1) Since the hydraulic pressure in the cylinder hole is controlled by the hydraulic pressure control means so as to satisfy the relationship expressed by the following expression, the centrifugal force acts on the slipper by rotation of the cylinder. Even in this case, the slipper and the plunger body are always kept in contact with each other due to the oil pressure in the cylinder hole, and it is possible to reliably prevent the oil from leaking due to the formation of a gap between the two. In particular, by the hydraulic control by the hydraulic control means, the force for pressing the slipper against the plunger body can always be set to be constant or optimal, and the slidability between the two can be ensured and the wear between the two can be effectively reduced. Can be reliably prevented from separating from each other.

In the above configuration, the leg of the connecting member is connected to the plunger body so as to be movable in the axial direction.
Furthermore, a plunger may be provided in the plunger main body by providing an urging member that applies an axial force in a direction of pressing the head against the concave spherical inner peripheral surface to the connecting member. In this case,
The above-mentioned first to third axial forces F1 to F3 and the fourth axial force F4 acting to press the slipper against the plunger body via the connecting member by the biasing member are expressed by the following equation (F1).
+ F4)> (F2 + F3) (2) The oil pressure in the cylinder hole is controlled by the oil pressure control means so as to satisfy the relationship expressed by the following expression. Therefore, in this case, since the fourth axial force F4 by the urging member (for example, a spring) always acts to press the slipper against the plunger body, even if the oil pressure in the cylinder hole is zero, both of them are pressed to some extent. The fitting is performed with a force. Even in such a case, no gap is generated between the two.

Further, in this case, when the cylinder rotates and a centrifugal force is applied to the slipper, the centrifugal force causes the slipper to separate from the plunger body (F).
Since the hydraulic pressure in the cylinder hole facing 3) can be set by the hydraulic control means, the force of the urging member may be a small value, and the sliding force between the slipper and the plunger body is generated by the urging force of this urging member. It is possible to almost eliminate the possibility that the mobility may be impaired or wear may occur between the two.

In particular, in the present invention, at least one of the hydraulic stepless transmission is composed of a variable displacement type swash plate plunger type hydraulic pump and a hydraulic motor, and is disposed in parallel with the hydraulic stepless transmission. Suitable for use in a plunger type hydraulic unit including a mechanical power transmission mechanism for performing mechanical power transmission and a lock-up mechanism for performing power transmission only by the mechanical power transmission mechanism. The means controls the hydraulic pressure so that the hydraulic pressure in the cylinder hole of the hydraulic pump or the hydraulic motor satisfies the above formula (1) or (2) when the lock-up mechanism is operated. When the lock-up mechanism is activated, there is no power transmission via the hydraulic continuously variable transmission mechanism, so the hydraulic pressure in the closed circuit connecting the hydraulic pump and the motor decreases, but the hydraulic control means Since the oil pressure in the closed circuit is adjusted so as to satisfy the above equations (1) and (2), even when a centrifugal force acts on the slipper, it is ensured that a gap is generated between the slipper and the plunger body. Is prevented.

[Brief description of the drawings]

FIG. 1 is a schematic view showing the configuration of a vehicular hydraulic continuously variable transmission constituting a plunger type hydraulic unit according to the present invention.

FIG. 2 is a graph showing a relationship between a swash plate angle of a pump and a motor and an overall speed ratio in the hydraulic continuously variable transmission.

FIG. 3 is a hydraulic circuit diagram showing a configuration of a hydraulic closed circuit and a control hydraulic circuit system of the hydraulic continuously variable transmission.

FIG. 4 is a schematic sectional view showing a configuration of a regulator valve used in the control hydraulic circuit system.

FIG. 5 is a schematic sectional view showing a configuration of a modulator valve used in the control hydraulic circuit system.

FIG. 6 is a graph showing a relationship between a control current (I) and a control pressure (Pc) in a first linear solenoid valve.

FIG. 7 is a graph showing a relationship between a control pressure (P) and a line pressure (PL) in a regulator valve.

FIG. 8 is a schematic sectional view showing a configuration of a shuttle valve used in the control hydraulic circuit system.

FIG. 9 is a schematic sectional view showing a configuration of a high-pressure relief valve used in the control hydraulic circuit system.

FIG. 10 is a graph showing a relationship between a high pressure relief pressure (PH) by the high pressure relief valve and a control current (I) of the second and third linear solenoid valves.

FIG. 11 is a schematic diagram showing a valve plate surface shape of a valve plate of a hydraulic pump and a configuration in which a variable notch valve is provided.

FIG. 12 is a schematic sectional view showing a configuration of a variable notch valve.

FIG. 13 is a hydraulic circuit diagram showing a swing angle control system of a swash plate of a pump and a motor.

FIG. 14 is a cross-sectional view showing a configuration of a motor plunger constituting the hydraulic pump.

FIG. 15 is an enlarged sectional view showing a slipper portion of the motor plunger.

FIG. 16 is a cross-sectional view showing a different example of a motor plunger constituting the hydraulic pump.

FIG. 17 is an enlarged sectional view showing a slipper portion of the motor plunger.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Mechanical transmission unit 2 Continuously variable transmission unit 3 Power split device 4 Power transmission device 5 Final reduction gear 24 Hydraulic pump 25 Hydraulic motor 26 Hydraulic closed circuit 32 Pump swash plate 38 Motor swash plate 60 Regulator valve 65 Modulator valve 70 Shuttle valve 75 High pressure Relief valve 80 Variable notch valve 84 Pump control valve 87 Motor control valve 93 Lock-up brake 200, 300 Motor plunger 201, 201 Plunger main body 205, 305 Slipper 210, 310 Connecting member

Claims (3)

[Claims] A cylinder that rotates about a rotation axis;
It comprises a plunger slidably inserted into a cylinder hole formed in the annular arrangement surrounding the rotation axis of the cylinder and extending in the axial direction, and a swash plate with which the tip of the plunger abuts slidably. In the plunger type hydraulic unit, the plunger is formed in a bowl shape having a convex spherical outer peripheral surface and a concave spherical inner peripheral surface, and has a slipper surface which is in sliding contact with the swash plate at an axial end. A plunger body having at one axial end a spherical concave surface slidably fitted to the convex spherical outer peripheral surface; and a head slidingly contacting the concave spherical inner peripheral surface at one axial end. And a connecting member integrally connected to the head and extending to the other end in the axial direction penetrates the slipper and is connected to the plunger body. Hydraulically forward A first axial force F1 for pressing the plunger body toward the slipper, a second axial force F2 acting to separate the plunger body from the slipper by a hydraulic pressure acting in the bowl-shaped internal space of the slipper, In the plunger-type hydraulic unit, a third axial force F3 acting to separate the plunger main body from the slipper by a centrifugal force applied to the slipper when the cylinder rotates is expressed as follows: F1> (F2 + F3) (1) A plunger-type hydraulic unit comprising: hydraulic control means for controlling a hydraulic pressure in the cylinder hole so as to satisfy a relationship represented by: 2. A hydraulic continuously variable transmission mechanism comprising at least one of a variable displacement swash plate plunger hydraulic pump and a hydraulic motor, at least one of which is provided in parallel with the hydraulic continuously variable transmission mechanism. A mechanical power transmission mechanism disposed to perform mechanical power transmission; and a lockup mechanism configured to perform power transmission only by the mechanical power transmission mechanism, wherein the hydraulic control unit includes the lockup mechanism. 2. The plunger-type hydraulic according to claim 1, wherein the hydraulic control is performed such that the hydraulic pressure in a cylinder hole of the hydraulic pump or the hydraulic motor satisfies the expression (1) when the hydraulic pump is operating. 3. unit. 3. A slipper, wherein the plunger is formed in a bowl shape having a convex spherical outer peripheral surface and a concave spherical inner peripheral surface, and has a slipper surface slidingly contacting the swash plate at an axial end. A plunger body having a spherical concave surface formed at one end in the axial direction to be slidably fitted to the convex spherical outer peripheral surface; and a head slidingly contacting the concave spherical inner peripheral surface at one axial end. A connecting member integrally connected to the head and extending to the other end in the axial direction penetrates the slipper and is connected to the plunger main body. A first axial force F1 for pressing the plunger body toward the slipper by hydraulic pressure in the cylinder hole; and a biasing member for applying a force pressing the inner surface of the concave spherical surface to the connecting member. Inside the bowl of the slipper A second axial force F2 acting to separate the plunger main body from the slipper by a hydraulic pressure acting in the internal space; and a centrifugal force applied to the slipper when the cylinder rotates in the plunger type hydraulic unit. A third axial force F3 acting to separate the plunger body, and a fourth force acting to press the slipper against the plunger body by the urging member via the connecting member.
It is characterized by having the urging member in which the fourth axial force F4 is set so that the axial force F4 satisfies the relationship expressed by the following expression: (F1 + F4)> (F2 + F3) (2) The plunger-type hydraulic unit according to claim 1 or 2, wherein