JP2024140919A - Variable capacity hydraulic rotary machine - Google Patents

Abstract

To stably match a neutral position of a regulator (for example, spool) with an arbitrary tilt angle of a tilt member.SOLUTION: A swash plate type hydraulic pump 1 comprises: a swash plate 11; a tilt actuator 16 that changes a tilt angle of the swash plate 11 by a servo piston 20 that moves due to a pressure difference between a pressure in a first hydraulic chamber 17C and a pressure in a second hydraulic chamber 17D formed in a tilt cylinder 17; and a regulator 23 that adjusts the pressure supplied to the first hydraulic chamber 17C or the second hydraulic chamber 17D according to a command pressure, and the swash plate type hydraulic pump has a dead zone where a servo piston 20 maintains a constant position relative to the tilt cylinder 17 until the command pressure supplied to the regulator 23 reaches a predetermined pressure. The servo piston 20 maintains a constant position within the tilt cylinder 17 while the command pressure supplied to the regulator 23 is in the dead zone, so that a neutral position of a spool 28 and a tilt angle of the swash plate 11 can be stably matched.SELECTED DRAWING: Figure 4

Description

This disclosure relates to a variable displacement hydraulic rotating machine that is mounted as a hydraulic pump or hydraulic motor on construction machinery such as hydraulic excavators and wheel loaders.

Construction machinery such as hydraulic excavators are equipped with variable displacement hydraulic rotating machines such as hydraulic pumps and hydraulic motors. The variable displacement hydraulic rotating machine has a capacity variable section including a tilting member such as a swash plate, and the capacity is variably controlled by driving the tilting member to tilt by a tilting actuator. The tilting actuator includes a tilting cylinder and a servo piston slidably inserted into the tilting cylinder. The servo piston is connected to the tilting member via a connecting section, and moves within the tilting cylinder in response to the servo pressure supplied to the tilting cylinder via a regulator, thereby tilting the tilting member and changing the tilting angle.

When such a variable displacement hydraulic rotating machine is used in a hydraulic closed circuit such as an HST (Hydro Static Transmission), the tilting member is tilted in both directions around a neutral position where the tilting angle is zero. When the tilting member is held in the neutral position, the variable displacement hydraulic rotating machine does not discharge or supply/drain pressurized oil. On the other hand, when the tilting member is tilted in one direction or the other direction from the neutral position, the variable displacement hydraulic rotating machine switches the discharge direction or supply/drain direction of the pressurized oil appropriately depending on the tilting angle (tilting direction) of the tilting member (see Patent Document 1).

Japanese Patent Application Publication No. 8-261158

In a variable displacement hydraulic rotary machine according to the prior art, the regulator is provided with a spool that displaces in response to a command pressure in order to change the angle of the tilting member. When the command pressure is less than a predetermined pressure, the spool does not displace and remains in a neutral position. The regulator is adjusted so that the tilting angle of the tilting member when the spool is in the neutral position is the desired angle, but it is difficult to match the desired tilting angle because it is affected by manufacturing errors of the parts that make up the tilting actuator, regulator, etc. Furthermore, the operation (tilt amount) becomes unstable when the tilting angle of the tilting member is tilted in one direction or the other around the neutral position of the spool, resulting in a problem of reduced accuracy when controlling the tilting angle of the tilting member.

The object of the present invention is to provide a variable displacement hydraulic rotary machine that can stably match the neutral position of the regulator (e.g., a spool) to any tilt angle of the tilt member.

The present invention is a variable displacement hydraulic rotating machine comprising a tilting member that changes the discharge capacity according to the tilting angle, a tilting cylinder and a servo piston that forms a first hydraulic chamber and a second hydraulic chamber in the tilting cylinder, a tilting actuator that changes the tilting angle of the tilting member by the servo piston that moves in the tilting cylinder according to the pressure difference between the pressure in the first hydraulic chamber and the pressure in the second hydraulic chamber, and a regulator that adjusts the pressure supplied to the first hydraulic chamber or the second hydraulic chamber according to a command pressure to move the servo piston, characterized in that the servo piston has a dead zone in which it maintains a constant position relative to the tilting cylinder until the command pressure supplied to the regulator reaches a predetermined pressure.

According to the present invention, while the command pressure supplied to the regulator is between zero and a predetermined pressure, the servo piston is in a dead zone and the tilt angle of the tilt member is held constant, so the neutral position of the regulator (e.g., a spool) and the tilt angle of the tilt member can be stably matched.

1 is a cross-sectional view showing a swash plate type hydraulic pump to which an embodiment of the present invention is applied. 2 is a cross-sectional view of the tilt actuator as viewed from the direction of arrows II-II in FIG. 1. FIG. 2 is a cross-sectional view showing the tilt actuator together with a regulator, a feedback link, etc. FIG. 4 is an enlarged cross-sectional view of a regulator in FIG. 3 . 5 is a cross-sectional view showing main parts such as an adjuster, a spool, and a spring mechanism in FIG. 4. FIG. 2 is a hydraulic circuit diagram including a regulator and a tilt actuator. FIG. 4 is a characteristic diagram showing the relationship between a command pressure and a spool position. FIG. 4 is a characteristic diagram showing the relationship between the position of a spool and the position of a servo piston. FIG. 4 is a characteristic diagram showing the relationship between a command pressure and a position of a servo piston.

The following is a detailed explanation of an embodiment of the variable displacement hydraulic rotary machine of the present invention, using an example in which it is applied to a variable displacement swash plate type hydraulic pump, with reference to the attached drawings.

The swash plate type hydraulic pump 1 shown in FIG. 1 is used in a hydraulic closed circuit such as an HST (Hydro Static Transmission), and is a bi-directional tilt type hydraulic pump that tilts a swash plate 11 (described later) in one direction and the other direction around a neutral position where the tilt angle is zero. The swash plate type hydraulic pump 1 has a casing 2 that forms an outer shell. The casing 2 is configured to include a stepped cylindrical casing body 3 with a bottom 3A at one end and a lid 4 that closes the other end of the casing body 3. The casing body 3 is provided with an actuator mounting portion 3B that protrudes outward from between the bottom 3A and the lid 4, and the actuator mounting portion 3B is formed with an opening 3C that opens to the outside. The actuator mounting portion 3B is provided with a tilt actuator 16 (described later), and a feedback link 47 (described later) is inserted into the opening 3C. The lid 4 of the casing 2 is formed with supply and exhaust passages 15A and 15B (described later).

A rotating shaft 5 is rotatably provided inside the casing 2. One axial side of the rotating shaft 5 is rotatably supported via a bearing on the bottom 3A of the casing body 3. The other axial side of the rotating shaft 5 is rotatably supported via a bearing on the cover 4. One axial end of the rotating shaft 5 protrudes from the bottom 3A of the casing body 3 to the outside. One end (protruding end) of the rotating shaft 5 is connected to a prime mover (not shown) such as an engine, and is rotated by the prime mover.

The cylinder block 6 is splined to the outer periphery of the rotating shaft 5 inside the casing 2 and rotates integrally with the rotating shaft 5. The cylinder block 6 is formed with a plurality of cylinders 7 that are spaced apart circumferentially and extend in the axial direction. A piston 8 is slidably inserted into each of the cylinders 7. The pistons 8 reciprocate within each cylinder 7 as the cylinder block 6 rotates.

These multiple pistons 8 reciprocate between a bottom dead center position where they protrude (extend) axially from the cylinder 7, and a top dead center position where they contract into the cylinder 7. During one rotation of the cylinder block 6, the multiple pistons 8 repeat an intake stroke in which they draw in hydraulic oil while sliding from top dead center to bottom dead center within the cylinder 7, and a discharge stroke in which they pressurize and discharge the sucked in hydraulic oil while sliding from bottom dead center to top dead center.

A shoe 9 is swingably attached to each of the protruding ends of the multiple pistons 8. The shoe 9 is pressed against the smooth surface 11A of the swash plate 11 by the pressing force (hydraulic force) from each piston 8. The shoe 9 rotates together with the rotating shaft 5, cylinder block 6, and pistons 8, and slides on the smooth surface 11A of the swash plate 11 in a ring-shaped trajectory.

A swash plate support member 10 is fixed to the bottom 3A of the casing body 3. The swash plate support member 10 is disposed on the back side of the swash plate 11 and has a concavely curved tilt sliding surface 10A that supports the swash plate 11 so that it can tilt. The tilt sliding surfaces 10A form a pair on either side of the rotating shaft 5, and each of them slidably abuts against the back side of the swash plate 11.

The swash plate 11, which serves as a tilting member, is provided within the casing 2. The swash plate 11 is attached to the bottom 3A side of the casing body 3 via a swash plate support member 10. The back side of the swash plate 11 (bottom 3A side) abuts against the tilting sliding surface 10A of the swash plate support member 10. The front side of the swash plate 11 (cylinder block 6 side) is a smooth surface 11A, and the shoe 9 slides on this smooth surface 11A. A shaft insertion hole 11B is provided in the center of the swash plate 11, and the rotating shaft 5 is rotatably inserted into the shaft insertion hole 11B.

The swash plate 11 is tiltably supported by the swash plate support member 10 and is tilted by the tilt actuator 16 described later. The swash plate hydraulic pump 1 is a double tilt type hydraulic pump, and can be tilted in one direction and the other direction around a neutral position where the tilt angle of the swash plate 11 is almost zero. Such a double tilt type swash plate hydraulic pump 1 can be connected to a hydraulic motor (not shown) by, for example, a hydraulic closed circuit, and the rotation speed and rotation direction of the hydraulic motor can be appropriately switched according to the tilt angle of the swash plate 11. When the tilt angle of the swash plate 11 is held at zero, the rotating shaft 5 of the swash plate hydraulic pump 1 rotates without load, and the discharge amount of pressure oil becomes zero. On the other hand, when the tilt actuator 16 tilts the swash plate 11 in one direction or the other direction around the neutral position, pressure oil according to the tilt direction and tilt angle is discharged from the swash plate hydraulic pump 1. For example, if a hydraulic motor (not shown) for traveling is connected to the swash plate type hydraulic pump 1, the hydraulic motor for traveling can be rotated in the forward or reverse direction.

The tilt lever 12 is integrally formed with the swash plate 11. The tilt lever 12 extends from the swash plate 11 toward the servo piston 20 described below. A pin 12A protrudes from the tip of the tilt lever 12, and the pin 12A is rotatably inserted into a through hole 13A formed in the center of the connecting member 13. The connecting member 13 slidably fits into a connecting groove 20C provided in the servo piston 20, and moves in a direction crossing the axial center of the servo piston 20 within the connecting groove 20C as the servo piston 20 moves in the axial direction. As a result, the axial displacement of the servo piston 20 is transmitted to the swash plate 11 via the tilt lever 12, and the swash plate 11 is tilted in accordance with the axial displacement of the servo piston 20.

The valve plate 14 is attached to the cover 4 of the casing 2. The valve plate 14 constitutes a switching valve plate that slides against the end face of the cylinder block 6. The valve plate 14 is provided with a pair of supply and discharge ports 14A and 14B. The supply and discharge ports 14A and 14B are formed as arc-shaped (eyebrow-shaped) long holes centered on the axial center of the rotating shaft 5, and are arranged to surround the periphery of the rotating shaft 5. The supply and discharge ports 14A and 14B are switched to suction ports or discharge ports depending on the tilt direction of the swash plate 11. For example, when the supply and discharge port 14A becomes a low-pressure side suction port, the supply and discharge port 14B becomes a high-pressure side discharge port. On the other hand, when the tilt direction of the swash plate 11 is reversed, the supply and discharge port 14A becomes a high-pressure side discharge port, and the supply and discharge port 14B becomes a low-pressure side suction port.

A pair of supply and discharge passages 15A, 15B are formed in the cover 4 of the casing 2. These supply and discharge passages 15A, 15B supply and discharge hydraulic oil (pressurized oil) to and from the cylinder 7 via the supply and discharge ports 14A, 14B of the valve plate 14. That is, when the rotating shaft 5 rotates in the casing 2, the piston 8 reciprocates in each cylinder 7 in conjunction with the rotation of the cylinder block 6, so that hydraulic oil is sucked into the cylinder 7 from one of the supply and discharge passages 15A, 15B, and pressurized hydraulic oil (pressurized oil) is discharged to the other of the supply and discharge passages 15A, 15B.

The tilt actuator 16 is provided on the actuator mounting portion 3B of the casing body 3. The tilt actuator 16 includes a tilt cylinder 17 and a servo piston 20. The tilt cylinder 17 is formed on the actuator mounting portion 3B of the casing body 3. The servo piston 20 is slidably inserted into the tilt cylinder 17, and tilts the swash plate 11, thereby driving the swash plate 11 to tilt in both directions, with the neutral position in between.

The tilt cylinder 17 has a pair of cylinder holes 17A, 17B (see FIG. 2). The pair of cylinder holes 17A, 17B are arranged concentrically with the same hole diameter and extend in a direction perpendicular to the axial center of the rotating shaft 5. The cylinder holes 17A, 17B are closed by side covers 18, 19 attached to the casing body 3 using bolts or the like. The side cover 18 that closes the cylinder hole 17A has a cylindrical protrusion 18A that protrudes into the tilt cylinder 17 and is formed concentrically with the cylinder hole 17A. The side cover 19 that closes the cylinder hole 17B has a cylindrical protrusion 19A that protrudes into the tilt cylinder 17 and is formed concentrically with the cylinder hole 17B.

The servo piston 20 is inserted into the tilt cylinder 17 so as to be slidable in the axial direction, and forms a first hydraulic chamber 17C and a second hydraulic chamber 17D within the tilt cylinder 17. As shown in Figures 2 and 3, the servo piston 20 has a solid cylindrical shape in the middle in the axial direction (length direction), and a cylindrical one-side cylindrical portion 20A is formed on one axial side, and a cylindrical other-side cylindrical portion 20B is formed on the other axial side. The one-side cylindrical portion 20A and the other-side cylindrical portion 20B have the same outer diameter dimensions.

The one-side cylindrical portion 20A is slidably inserted into the cylinder bore 17A of the tilt cylinder 17, and forms a first hydraulic pressure chamber 17C between it and the side cover 18. The other-side cylindrical portion 20B is slidably inserted into the cylinder bore 17B of the tilt cylinder 17, and forms a second hydraulic pressure chamber 17D between it and the side cover 19. The servo piston 20 moves axially within the tilt cylinder 17 in response to the pressure difference between the pressure in the first hydraulic pressure chamber 17C and the pressure in the second hydraulic pressure chamber 17D.

A connecting groove 20C and a link mounting groove 20D are provided in the axial middle of the servo piston 20 so as to face each other in the radial direction. A pin 12A of the tilt lever 12 is connected to the connecting groove 20C of the servo piston 20 via a connecting member 13. A lever member 49 of a feedback link 47 (described later) is engaged with the link mounting groove 20D of the servo piston 20.

The first hydraulic chamber 17C of the tilt cylinder 17 is provided with a first servo spring 21 that presses the servo piston 20 toward the second hydraulic chamber 17D. The first servo spring 21 includes a guide shaft 21A, a spring guide 21B, and a compression spring 21C. The guide shaft 21A is, for example, a cylindrical body, and is fixed in a state where it protrudes from the side cover 18 into the first hydraulic chamber 17C by screwing a bolt 21D inserted into the inner periphery side into the protrusion 18A of the side cover 18. The spring guide 21B is formed in a stepped cylindrical shape having a large diameter cylindrical portion 21B1 and a small diameter cylindrical portion 21B2. The outer diameter dimension of the large diameter cylindrical portion 21B1 is set smaller than the inner diameter of the one-side cylindrical portion 20A of the servo piston 20, and the outer diameter dimension of the small diameter cylindrical portion 21B2 is set equal to the outer diameter of the protrusion 18A of the side cover 18. A shaft insertion hole 21B3 is formed in the center of the small diameter cylindrical portion 21B2, penetrating in the axial direction. The spring guide 21B is supported movably on the guide shaft 21A by inserting the guide shaft 21A through the shaft insertion hole 21B3, and is prevented from coming off in the axial direction by the head of the bolt 21D.

Compression spring 21C is provided between side cover 18 and spring guide 21B. Specifically, compression spring 21C is disposed on the outer periphery of small diameter cylindrical portion 21B2 of spring guide 21B and protrusion 18A of side cover 18, and presses (biases) spring guide 21B in a direction away from side cover 18. Here, compression spring 21C is compressed between side cover 18 and spring guide 21B in a state where it is reduced from its free length, and presses spring guide 21B with a predetermined set load (initial load).

The second hydraulic chamber 17D of the tilt cylinder 17 is provided with a second servo spring 22 that presses the servo piston 20 toward the first hydraulic chamber 17C. The second servo spring 22, like the first servo spring 21, is composed of a guide shaft 22A, a spring guide 22B, and a compression spring 22C. The guide shaft 22A is fixed to the side cover 19 using a bolt 22D and protrudes into the second hydraulic chamber 17D. The spring guide 22B has a large diameter cylindrical portion 22B1, a small diameter cylindrical portion 22B2, and a shaft insertion hole 22B3, and is supported movably on the guide shaft 22A by inserting the shaft insertion hole 22B3. The compression spring 22C is provided between the side cover 19 and the spring guide 22B, and presses the spring guide 22B in a direction away from the side cover 19. Here, the compression spring 22C is compressed between the side cover 19 and the spring guide 22B in a state where it is shorter than its free length, and presses the spring guide 22B with a predetermined set load (initial load).

When the servo piston 20 is held in the neutral position, the tilt angle of the swash plate 11 is zero (neutral position). With the servo piston 20 held in the neutral position, a predetermined clearance (not shown) is formed between the servo piston 20 and the spring guide 21B in the first hydraulic pressure chamber 17C of the tilt cylinder 17, and a predetermined clearance (not shown) is formed between the servo piston 20 and the spring guide 22B in the second hydraulic pressure chamber 17D. Therefore, when the pressure in the first hydraulic pressure chamber 17C becomes greater than the pressure in the second hydraulic pressure chamber 17D by supplying servo pressure to the first hydraulic pressure chamber 17C, the servo piston 20 moves toward the second hydraulic pressure chamber 17D by the amount of the clearance formed in the second hydraulic pressure chamber 17D. Then, when the load acting on the servo piston 20 due to the pressure difference between the pressure in the first hydraulic chamber 17C and the pressure in the second hydraulic chamber 17D becomes greater than the set load of the second servo spring 22 (compression spring 22C), the servo piston 20 moves further toward the second hydraulic chamber 17D while compressing the compression spring 22C.

On the other hand, when servo pressure is supplied to the second hydraulic chamber 17D of the tilt cylinder 17, and the pressure in the second hydraulic chamber 17D becomes greater than the pressure in the first hydraulic chamber 17C, the servo piston 20 moves toward the first hydraulic chamber 17C by the amount of the clearance formed in the first hydraulic chamber 17C. Then, when the load acting on the servo piston 20 due to the pressure difference between the pressure in the second hydraulic chamber 17D and the pressure in the first hydraulic chamber 17C becomes greater than the set load of the first servo spring 21 (compression spring 21C), the servo piston 20 moves further toward the first hydraulic chamber 17C while compressing the compression spring 21C.

In this way, when the pressure difference between the pressure in the first hydraulic chamber 17C and the pressure in the second hydraulic chamber 17D is equal to or less than the set load of the first servo spring 21 (compression spring 21C) and the second servo spring 22 (compression spring 22C), the servo piston 20 maintains a fixed position without moving within the tilt cylinder 17. That is, in this embodiment, the second dead zone of the servo piston 20 is set by the set load of the first servo spring 21 (compression spring 21C) and the second servo spring 22 (compression spring 22C). Note that, if the pressure difference between the pressure in the first hydraulic chamber 17C and the pressure in the second hydraulic chamber 17D no longer occurs due to some reason, the servo piston 20 is configured to maintain the neutral position by the first servo spring 21 and the second servo spring 22.

The regulator 23 is provided in the casing body 3 of the swash plate type hydraulic pump 1 adjacent to the actuator mounting portion 3B, and adjusts (controls) the servo pressure supplied to the tilt actuator 16 according to the command pressure (pilot pressure) supplied to the pilot oil chamber 32 or the pilot oil chamber 38 described below. The regulator 23 has a regulator casing 24 attached to the actuator mounting portion 3B, and the regulator casing 24 is arranged so as to cover the opening 3C formed in the actuator mounting portion 3B of the casing body 3. In other words, the inside of the regulator casing 24 is connected to the inside of the casing body 3 via the opening 3C.

As shown in FIG. 4, the regulator 23 is configured as a hydraulic servo valve including a regulator casing 24, a sleeve 27, a spool 28, a pilot cylinder 29, an adjuster 33, a spring mechanism 39, and a feedback link 47.

The regulator casing 24 is formed in a stepped cylindrical shape extending in the axial direction of the spool 28. Inside the regulator casing 24, a sleeve sliding hole 24A located in the middle of the axial direction (length direction), an adjuster mounting hole 24B located on one side in the axial direction, and a cylinder mounting hole 24C located on the other side in the axial direction are formed on the same axis. The adjuster mounting hole 24B opens at one end 24D of the regulator casing 24 in the axial direction, and an adjuster 33 is attached to it. The cylinder mounting hole 24C opens at the other end 24E of the regulator casing 24 in the axial direction, and a pilot cylinder 29 is attached to it. The diameter of the cylinder mounting hole 24C is set to be larger than the sleeve sliding hole 24A and smaller than the adjuster mounting hole 24B. In addition, a link insertion hole 24F for inserting a feedback link 47 is formed between the sleeve sliding hole 24A and the adjuster mounting hole 24B in the regulator casing 24.

One side cover 25 is attached to one axial end 24D of regulator casing 24 using bolts or the like. Bolt holes (female thread holes) 25A are formed in one side cover 25, and bolt holes 25A are arranged concentrically with adjuster mounting hole 24B. Another side cover 26 is attached to the other axial end 24E of regulator casing 24 using bolts or the like. Bolt holes 26A are formed in other side cover 26, and bolt holes 26A are arranged concentrically with cylinder mounting hole 24C.

The regulator casing 24 is formed with a pump port 24G and a pair of actuator ports 24H, 24J. The pump port 24G and the actuator ports 24H, 24J are arranged in line in the axial direction of the sleeve sliding hole 24A, and each opens into the sleeve sliding hole 24A. The pump port 24G is arranged in the middle between the actuator ports 24H and 24J, and is connected to a pump line 53 described later. The actuator port 24H is connected to the first hydraulic chamber 17C of the tilt cylinder 17 via a supply and exhaust line 57B described later, and the actuator port 24J is connected to the second hydraulic chamber 17D of the tilt cylinder 17 via a supply and exhaust line 57A described later. In addition, a pair of pilot ports 24K, 24L are formed in the regulator casing 24. The pilot port 24K opens into the adjuster mounting hole 24B, and is connected to a pilot line 55B described later. The pilot port 24L opens into the cylinder mounting hole 24C and is connected to the pilot line 55A, which will be described later.

The sleeve 27 is inserted into the sleeve sliding hole 24A of the regulator casing 24 and is movable in the axial direction. The sleeve 27 is formed in a cylindrical shape, and the inner periphery of the sleeve 27 forms a spool sliding hole 27A. The sleeve 27 has three sets of oil holes 27B, 27C, and 27D that penetrate in the radial direction, and these three sets of oil holes 27B, 27C, and 27D each extend radially from the periphery of the spool sliding hole 27A. In addition, a link engagement portion 27E is provided on one axial end side (adjuster 33 side) of the sleeve 27, and the engagement pin 48 of the feedback link 47 engages with this link engagement portion 27E.

Of the three sets of oil holes 27B, 27C, and 27D, the oil hole 27B located in the middle (between oil holes 27C and 27D) communicates with pump port 24G of regulator casing 24 and is blocked from actuator ports 24H and 24J when spool 28 is in neutral position. Oil hole 27C located on one axial side of sleeve 27 communicates with actuator port 24H of regulator casing 24 and is blocked from pump port 24G and actuator port 24J. Oil hole 27D located on the other axial side of sleeve 27 communicates with actuator port 24J of regulator casing 24 and is blocked from pump port 24G and actuator port 24H.

The spool 28 is movably disposed within the regulator casing 24 via a sleeve 27. The spool 28 moves axially relative to the sleeve 27 to connect or disconnect the pump port 24G of the regulator casing 24 to or from the actuator port 24H or 24J. This adjusts the servo pressure supplied to the first hydraulic chamber 17C or the second hydraulic chamber 17D of the tilt actuator 16, and controls the amount of movement of the servo piston 20. Here, in this embodiment, the amount of movement of the spool 28 and the amount of movement of the servo piston 20 are set to have a one-to-one relationship.

The spool 28 is formed as a stepped cylinder extending in the axial direction as a whole, and has a large diameter shaft portion 28A located on one axial side and a small diameter shaft portion 28B located on the other axial side. The small diameter shaft portion 28B of the spool 28 is slidably inserted into the spool sliding hole 27A of the sleeve 27. In addition, a disk-shaped flange portion 28C having an outer diameter dimension larger than that of the large diameter shaft portion 28A is provided between the large diameter shaft portion 28A and the small diameter shaft portion 28B (see FIG. 5). The flange portion 28C of the spool 28 constitutes a spool side engagement member with which the other side spring bearing 44 described below engages.

The large diameter shaft portion 28A of the spool 28 protrudes from the sleeve 27 and is disposed within the adjuster 33. The piston 35 described below abuts the end portion 28D of the spool 28 on the side of the large diameter shaft portion 28A, and the piston 31 described below abuts the end portion 28E of the small diameter shaft portion 28B. In addition, the other side spring retainer 44 abuts against the flange portion 28C of the spool 28.

Lands 28F and 28G are formed on the outer circumferential surface of the small diameter shaft portion 28B of the spool 28. The land 28F corresponds to the oil hole 27C of the sleeve 27, and the land 28G is located at a position corresponding to the oil hole 27D of the sleeve 27. The width (axial length) of the lands 28F and 28G is formed to be equal to the hole diameter of the oil holes 27C and 27D. Therefore, when the spool 28 is in the neutral position, the oil holes 27C and 27D of the sleeve 27 are blocked from the spool sliding hole 27A by the lands 28F and 28G of the spool 28. On the other hand, when the spool 28 moves axially from the neutral position within the spool sliding hole 27A, the oil holes 27C and 27D of the sleeve 27 communicate with the spool sliding hole 27A.

The pilot cylinder 29 is fixed to the cylinder mounting hole 24C of the regulator casing 24 so that its axial position can be adjusted, and constitutes a part of the regulator casing 24. The pilot cylinder 29 is formed in a bottomed cylinder shape with one axial end 29A as an open end and the other axial end 29B as a closed end. A piston insertion hole 29C is formed in the center of the pilot cylinder 29, and the piston insertion hole 29C opens at one end 29A of the pilot cylinder 29, and the other end 29B of the pilot cylinder 29 forms a bottom 29D. A port hole 29E is formed at the other end 29B of the pilot cylinder 29, penetrating radially through the piston insertion hole 29C. The port hole 29E is connected to the pilot port 24L of the regulator casing 24.

The cylinder abutment member 30 is attached to the other side cover 26 of the regulator casing 24. The cylinder abutment member 30 protrudes into the cylinder mounting hole 24C of the regulator casing 24, and this protruding end abuts the other end 29B of the pilot cylinder 29. The cylinder abutment member 30 is formed using a hexagonal socket bolt or the like, and the outer circumferential surface of the cylinder abutment member 30 is provided with a male thread 30A that screws into the bolt hole 26A of the other side cover 26. The cylinder abutment member 30 abuts against the other end 29B of the pilot cylinder 29 in the cylinder mounting hole 24C by screwing the male thread 30A into the bolt hole 26A of the other side cover 26. Therefore, the position of the port hole 29E of the pilot cylinder 29 can be adjusted relative to the pilot port 24L of the regulator casing 24 according to the amount of screwing of the cylinder abutment member 30. The pilot cylinder 29 can then be positioned by screwing a lock nut 30B onto the cylinder abutment member 30.

The piston 31 is slidably inserted into the piston insertion hole 29C of the pilot cylinder 29. One end 31A of the piston 31 protrudes from one end 29A of the pilot cylinder 29 toward the end 28E of the spool 28.

The pilot oil chamber 32 is located between the other end 31B of the piston 31 and the bottom 29D of the piston insertion hole 29C, and is formed in the piston insertion hole 29C of the pilot cylinder 29. The pilot oil chamber 32 is connected to the pilot port 24L of the regulator casing 24 via the port hole 29E of the pilot cylinder 29. Therefore, when the command pressure (pilot pressure) introduced into the pilot port 24L is supplied to the pilot oil chamber 32 through the port hole 29E of the pilot cylinder 29, the piston 31 presses the spool 28 in a direction away from the pilot cylinder 29.

The adjuster 33 is fixed to the adjuster mounting hole 24B of the regulator casing 24 so that its axial position can be adjusted, and constitutes a part of the regulator casing 24. The adjuster 33 has a small diameter cylindrical portion 33A and a large diameter cylindrical portion 33B, and is formed in a stepped cylindrical shape extending in the axial direction. A stopper fitting hole 33C into which a stopper 36 described later is inserted is formed in the center of the small diameter cylindrical portion 33A, and the stopper fitting hole 33C extends to the large diameter cylindrical portion 33B. A spool accommodation hole 33E having a hole diameter larger than the stopper fitting hole 33C and a spring accommodation hole 33F constituting a part of the spool accommodation hole 33E are formed on the end portion 33D side of the large diameter cylindrical portion 33B. The spring accommodation hole 33F has a hole diameter larger than the spool accommodation hole 33E. The large diameter shaft 28A of the spool 28 is accommodated in the spool accommodation hole 33E, and the large diameter shaft 28A of the spool 28 and the spring mechanism 39 described below are accommodated in the spring accommodation hole 33F. The boundary between the spool accommodation hole 33E and the spring accommodation hole 33F is an annular step 33G, which constitutes an adjuster side engagement member with which the one-side spring receiver 42 described below engages.

A piston insertion hole 33H is formed in the large diameter cylindrical portion 33B adjacent to the spool accommodation hole 33E. One end of the piston insertion hole 33H opens into the stopper fitting hole 33C, and the other end of the piston insertion hole 33H opens into the spool accommodation hole 33E. The piston insertion hole 33H has a smaller hole diameter than the stopper fitting hole 33C, and a piston 35 described below is inserted into it. A port hole 33J is formed in the large diameter cylindrical portion 33B of the adjuster 33, penetrating radially through the stopper fitting hole 33C, and the port hole 33J is connected to the pilot port 24K of the regulator casing 24.

A male screw 33K is provided on the outer peripheral surface of the small diameter cylindrical portion 33A of the adjuster 33. The male screw 33K is screwed into the bolt hole 25A of the one side cover 25. Therefore, by adjusting the amount of screwing of the male screw 33K into the one side cover 25 (bolt hole 25A), the position of the adjuster 33 can be adjusted relative to the regulator casing 24. This allows the zero position to be adjusted. The adjuster 33 can be positioned by abutting the lock nut 34 screwed onto the male screw 33K against the one side cover 25. In addition, a female screw 33L is formed on the inner peripheral surface of the stopper fitting hole 33C provided in the small diameter cylindrical portion 33A of the adjuster 33.

The piston 35 is slidably inserted into the piston insertion hole 33H of the adjuster 33. One end 35A of the piston 35 protrudes into the stopper fitting hole 33C, and the other end 35B of the piston 35 protrudes into the spool housing hole 33E.

The stopper 36 is movably fitted into the stopper fitting hole 33C of the adjuster 33. One end 36A of the stopper 36 is provided with a tool engagement portion 36B that engages with a tool such as a screwdriver, and a male thread 36C is formed on the outer circumferential surface of the one end 36A side of the stopper 36. The stopper 36 is fitted into the stopper fitting hole 33C of the adjuster 33 by screwing the male thread 36C into the female thread 33L of the adjuster 33, and the other end 36D of the stopper 36 abuts against one end 35A of the piston 35.

Therefore, by adjusting the amount of threading of the male thread 36C into the female thread 33L, the length of the piston 35 that protrudes into the spool housing hole 33E of the adjuster 33 can be adjusted. Then, by abutting the lock nut 37 screwed onto the male thread 36C of the stopper 36 against the end of the small diameter cylindrical portion 33A of the adjuster 33, the stopper 36 can be positioned relative to the adjuster 33.

The pilot oil chamber 38 is located between the other end 36D of the stopper 36 and the piston 35, and is formed in the stopper fitting hole 33C of the adjuster 33. The pilot oil chamber 38 is connected to the pilot port 24K of the regulator casing 24 via the port hole 33J of the adjuster 33. Therefore, when the command pressure (pilot pressure) introduced into the pilot port 24K is supplied to the pilot oil chamber 38 through the port hole 33J of the adjuster 33, the piston 35 presses the spool 28 in a direction away from the stopper 36.

Next, we will explain the spring mechanism 39 used in this embodiment.

The spring mechanism 39 is provided between the large diameter shaft portion 28A of the spool 28, which is disposed in the spring accommodation hole 33F of the adjuster 33, and the adjuster 33. As shown in FIG. 5, the spring mechanism 39 is composed of the flange portion 28C of the spool 28 and the shaft retaining ring 40 as the spool side engaging member, the annular step portion 33G of the adjuster 33 and the hole retaining ring 41 as the adjuster side engaging member, the one side spring retainer 42, the other side spring retainer 44, and the compression spring 46. The spring mechanism 39 applies a spring force to the spool 28 in the direction opposite to the pressing force acting from the piston 31 or the piston 35.

The shaft retaining ring 40 is attached to the outer peripheral surface of the end 28D side of the large diameter shaft portion 28A constituting the spool 28. The shaft retaining ring 40 constitutes a spool side engaging member with which the one side spring bearing 42 engages. The outer diameter dimension of the shaft retaining ring 40 is set smaller than the hole diameter of the spool accommodation hole 33E provided in the adjuster 33. The hole retaining ring 41 is attached to the inner peripheral surface of the end 33D side of the large diameter cylindrical portion 33B constituting the adjuster 33. The hole retaining ring 41 constitutes an adjuster side engaging member with which the other side spring bearing 44 engages. The inner diameter dimension of the hole retaining ring 41 is set larger than the outer diameter dimension of the flange portion 28C provided on the spool 28.

The one-side spring bearing 42 is inserted into the outer periphery of the large-diameter shaft portion 28A close to the shaft retaining ring 40 and is movable in the axial direction. The one-side spring bearing 42 has an annular flange portion 42A against which one end of the compression spring 46 abuts. The outer diameter of the flange portion 42A is set to be smaller than the hole diameter of the spring accommodation hole 33F provided in the adjuster 33 and larger than the hole diameter of the spool accommodation hole 33E. Therefore, the one-side spring bearing 42 is movable relative to the large-diameter shaft portion 28A of the spool 28 only within the spring accommodation hole 33F by engaging the flange portion 42A with the annular step portion 33G of the adjuster 33. In addition, an annular one-side spacer 43 that fits onto the outer periphery of the large-diameter shaft portion 28A is disposed between the shaft retaining ring 40 and the one-side spring bearing 42.

The other-side spring bearing 44 is inserted into the outer periphery of the large-diameter shaft portion 28A close to the hole retaining ring 41 and is movable in the axial direction. The other-side spring bearing 44 has an annular flange portion 44A against which the other end of the compression spring 46 abuts. The inner diameter of the other-side spring bearing 44 is set to be larger than the outer diameter of the large-diameter shaft portion 28A and smaller than the outer diameter of the flange portion 28C. The outer diameter of the flange portion 44A is set to be smaller than the hole diameter of the spring accommodating hole 33F. In addition, an annular other-side spacer 45 that fits into the inner periphery of the spring accommodating hole 33F is disposed between the hole retaining ring 41 and the other-side spring bearing 44. The inner diameter of the other-side spacer 45 is set to be larger than the outer diameter of the flange portion 28C. Here, as shown in FIG. 5, when the spool 28 is held in the neutral position, for example, a predetermined axial clearance ΔX is formed between the flange 28C of the spool 28 and the other side spring bearing 44.

The compression spring 46 is disposed on the outer periphery of the large diameter shaft portion 28A of the spool 28, and is provided between the flange portion 42A of the one-side spring retainer 42 and the flange portion 44A of the other-side spring retainer 44. The compression spring 46 is disposed between the one-side spring retainer 42 and the other-side spacer 45 in a state where it is reduced from its free length, and is compressed between the shaft retaining ring 40 and the hole retaining ring 41 with a predetermined set load (spring load). The compression spring 46 applies a spring force to the spool 28 in the opposite direction to the pressing force acting from the piston 31 or piston 35.

Therefore, when a command pressure is supplied to the pilot oil chamber 32 and the spool 28 is pressed against the piston 31, the spool 28 first moves toward the adjuster 33 by the amount of clearance ΔX. Then, when the load acting on the spool 28 from the piston 31 becomes greater than the set load of the spring mechanism 39 (compression spring 46), the spool 28 moves further toward the adjuster 33 while compressing the compression spring 46. On the other hand, when a command pressure is supplied to the pilot oil chamber 38 and the spool 28 is pressed against the piston 35, the spool 28 first moves toward the pilot cylinder 29 by the amount of clearance ΔX. Then, when the load acting on the spool 28 from the piston 35 becomes greater than the set load of the spring mechanism 39 (compression spring 46), the spool 28 moves further toward the pilot cylinder 29 while compressing the compression spring 46.

In this embodiment, the amount of movement of the spool 28 and the amount of movement of the servo piston 20 are set to have a one-to-one relationship. Therefore, when the load acting on the spool 28 from the piston 31 or piston 35 is equal to or less than the set load of the spring mechanism 39 (compression spring 46), the spool 28 maintains a constant position, and the servo piston 20 also maintains a constant position without moving within the tilt cylinder 17. That is, in this embodiment, the dead zone of the servo piston 20 is set by the set load of the spring mechanism 39 (compression spring 46).

The feedback link 47 is provided between the sleeve 27 of the regulator 23 and the servo piston 20. When the swash plate 11 of the swash plate type hydraulic pump 1 performs a tilting operation, the feedback link 47 moves the sleeve 27 axially relative to the spool 28 so as to feedback control the sleeve 27 of the regulator 23 in response to the tilting operation.

One end of the feedback link 47 in the longitudinal direction is engaged with the link engagement portion 27E of the sleeve 27 via an engagement pin 48. A springy lever member 49 is integrally attached to the other end of the feedback link 47 in the longitudinal direction, and this lever member 49 engages with the link attachment groove 20D of the servo piston 20. A fulcrum pin 50 is inserted through the middle of the longitudinal direction of the feedback link 47, and the fulcrum pin 50 is attached, for example, near an opening 3C provided in the actuator attachment portion 3B of the casing body 3. As a result, the feedback link 47 connects the sleeve 27 and the servo piston 20 through the opening 3C, and rotates (swings) around the fulcrum pin 50, thereby transmitting the axial displacement of the servo piston 20 to the sleeve 27.

Next, the hydraulic circuit including the tilt actuator 16 and regulator 23 will be described with reference to Figures 3 and 6.

The pilot pump 51 discharges the oil stored in the tank 52 into the pump line 53 as servo pressure. The pump line 53 is connected to the pump port 24G of the regulator 23 (regulator casing 24), and the servo pressure is supplied from the pump port 24G into the spool sliding hole 27A of the sleeve 27.

Two branch lines 53A and 53B are branched and connected to the pump line 53. The branch lines 53A and 53B are connected to pilot lines 55A and 55B via tilt control valves 54A and 54B. The pilot line 55A is connected to the pilot oil chamber 32 in the pilot cylinder 29 via the pilot port 24L of the regulator 23. The pilot line 55B is connected to the pilot oil chamber 38 in the adjuster 33 via the pilot port 24K of the regulator 23.

The tilt operation valves 54A and 54B are, for example, electromagnetic proportional valves. The tilt operation valves 54A and 54B are normally held in the return position (a), and are switched to the supply position (b) by applying a current corresponding to the target tilt angle and tilt amount of the swash plate 11 from a controller (not shown) in response to the operator's operation. In the return position (a), the tilt operation valve 54A connects the pilot line 55A to the tank 52 via the tank line 56A, and in the supply position (b), the branch line 53A connects to the pilot oil chamber 32 via the pilot line 55A. In the return position (a), the tilt operation valve 54B connects the pilot line 55B to the tank 52 via the tank line 56B, and in the supply position (b), the branch line 53B connects to the pilot oil chamber 38 via the pilot line 55B. Here, the tilt control valves 54A and 54B are switched in proportion to the current value, so the command pressure from the branch lines 53A and 53B is supplied to the pilot lines 55A and 55B in a state in which the pressure is variably controlled (adjusted) according to the amount of operation of the tilt control valves 54A and 54B.

Servo pressure supply and exhaust pipes 57A and 57B are provided between the regulator 23 (regulator casing 24) and the tilt actuator 16. Supply and exhaust pipe 57A connects between the actuator port 24J of the regulator 23 and the second hydraulic chamber 17D of the tilt cylinder 17. Supply and exhaust pipe 57B connects between the actuator port 24H of the regulator 23 and the first hydraulic chamber 17C of the tilt cylinder 17.

The servo piston 20 of the tilt actuator 16 moves axially within the tilt cylinder 17 in response to the pressure difference between the pressure in the first hydraulic chamber 17C and the pressure in the second hydraulic chamber 17D. The axial movement of the servo piston 20 is transmitted to the swash plate 11 of the swash plate type hydraulic pump 1 via the tilt lever 12, and the swash plate 11 performs a tilting operation in response to the amount of movement of the servo piston 20. Meanwhile, the axial movement of the servo piston 20 is transmitted to the sleeve 27 of the regulator 23 via the feedback link 47, and the sleeve 27 is feedback controlled to follow the tilting operation of the swash plate 11.

The swash plate type hydraulic pump 1 according to this embodiment has the configuration described above. Next, we will explain the operation of the regulator 23, tilt actuator 16, etc. when tilting the swash plate 11 in one direction and the other direction from the neutral position where the tilt angle is zero.

When the swash plate 11 of the swash plate type hydraulic pump 1 is driven to tilt in one direction from the neutral position, for example, the tilt operation valve 54A is held in the return position (a) and the tilt operation valve 54B is switched from the return position (a) to the supply position (b).

As a result, the command pressure flowing from pump line 53 to branch line 53B is supplied from tilt control valve 54B through pilot line 55B to pilot oil chamber 38 of regulator 23 (adjuster 33). At this time, because tilt control valve 54A is held in the return position (a), the inside of pilot line 55A is kept at tank pressure, and pilot oil chamber 32 of regulator 23 (pilot cylinder 29) is also kept at tank pressure. Therefore, piston 35 receives the command pressure in pilot oil chamber 38 and applies a pressing force to spool 28 in the direction of arrow A in FIG. 3.

At this time, as shown in FIG. 5, a spring mechanism 39 is provided between the large diameter shaft portion 28A and the adjuster 33, and a compression spring 46 is compressed with a predetermined set load between the shaft retaining ring 40 attached to the large diameter shaft portion 28A and the hole retaining ring 41 attached to the adjuster 33. In addition, a predetermined axial clearance ΔX is formed between the flange portion 28C of the spool 28 and the other side spring retainer 44.

As a result, as shown in the characteristic line 58 in FIG. 7, when the command pressure Pp is in the range of pressure Pp0 or less (Pp≦Pp0) corresponding to the set load of the spring mechanism 39 (compression spring 46), the position X of the spool 28 is displaced from the neutral position (zero) when the swash plate 11 is in the neutral position to a position X1 corresponding to the amount of movement of the clearance ΔX formed between the flange 28C and the other side spring bearing 44 by the spool 28, and this position X1 is maintained. Then, when the command pressure Pp exceeds the pressure Pp0 corresponding to the set load of the compression spring 46, the spool 28 moves in the direction of the arrow A in FIG. 3 against the spring force of the compression spring 46 (by compressing the compression spring 46), and the position X of the spool 28 increases from the position X1 as the command pressure Pp increases (the amount of movement of the spool 28 increases).

When the spool 28 moves in the direction of arrow A along the sleeve 27, the oil hole 27B of the sleeve 27 communicates with the oil hole 27D between the lands 28F and 28G of the spool 28. This connects the pump port 24G of the regulator casing 24 to the actuator port 24J. Therefore, the servo pressure supplied to the pump port 24G from the pilot pump 51 via the pump line 53 is supplied to the second hydraulic chamber 17D of the tilt cylinder 17 via the actuator port 24J and the supply/exhaust line 57A.

When servo pressure from the pilot pump 51 is supplied to the second hydraulic chamber 17D of the tilt cylinder 17, the servo piston 20 moves in the direction of arrow C in FIG. 3. As a result, the oil in the first hydraulic chamber 17C of the tilt actuator 16 is discharged through the supply and exhaust pipe 57B and the actuator port 24H of the regulator casing 24 to the sleeve sliding hole 24A, and is returned to the tank 52 through a drain passage (not shown).

The swash plate type hydraulic pump 1 according to this embodiment is set so that the amount of movement of the spool 28 and the amount of movement of the servo piston 20 have a one-to-one relationship, as shown by characteristic line 59 in FIG. 8. Therefore, when the position X of the spool 28 is displaced from the neutral position (zero) when the swash plate 11 is in the neutral position to position X1, the position S of the servo piston 20 is displaced to position S1 corresponding to the amount of movement of the spool 28, relative to the neutral position (zero) when the swash plate 11 is in the neutral position.

Therefore, as shown in the characteristic line 60 in FIG. 9, when the command pressure Pp is in the range of pressure Pp0 or less (Pp≦Pp0) corresponding to the set load of the spring mechanism 39 (compression spring 46), the position S of the servo piston 20 first moves from the neutral position (zero) to position S1 and maintains this position S1. The amount of movement of the servo piston 20 at this time corresponds to the amount of movement of the spool 28 through the clearance ΔX formed between the flange 28C and the other side spring bearing 44 (position X1 of the spool 28). Then, when the command pressure Pp exceeds the pressure Pp0 corresponding to the set load of the spring mechanism 39 (compression spring 46), the position S of the servo piston 20 increases from position S1 as the command pressure Pp increases (the amount of movement of the servo piston 20 increases), similar to the spool 28.

Here, a predetermined clearance (not shown) is formed between the servo piston 20 and the spring guide 21B in the first hydraulic chamber 17C of the tilt cylinder 17. Therefore, as the command pressure Pp increases, the servo piston 20 is displaced from position S1 to position S2, which is moved by the clearance, without compressing the compression spring 21C of the first servo spring 21. Then, when the command pressure Pp exceeds the pressure corresponding to the set load of the first servo spring 21 (compression spring 21C), the servo piston 20 moves in the direction of arrow C against the spring force of the compression spring 21C (compressing the compression spring 21C), and the position S of the servo piston 20 further increases from position S2 as the command pressure Pp increases (the movement amount of the servo piston 20 further increases).

As shown in FIG. 9, in this embodiment, the amount of movement of the servo piston 20 while it is displaced from position S1 to position S2, i.e., the amount of movement of the servo piston 20 within the second dead zone set by the set load of the first servo spring 21 (compression spring 21C), is greater than the amount of movement of the servo piston 20 while it is displaced from zero to position S1, i.e., the amount of movement of the servo piston 20 within the dead zone set by the set load of the spring mechanism 39 (compression spring 46).

While the servo piston 20 is held at position S1, the servo piston 20 does not tilt the swash plate 11, and the swash plate 11 maintains a neutral position where the tilt angle is nearly zero. When the tilt angle of the swash plate 11 is close to zero and the pressing force from the piston 31 acting on the spool 28 of the regulator 23 is within the dead zone set by the set load of the spring mechanism 39, the spool 28 maintains the neutral position without moving, and the supply of servo pressure to the second hydraulic chamber 17D of the tilt cylinder 17 is cut off. Therefore, when the tilt angle of the swash plate 11 is close to zero, for example, even if there is a manufacturing error in the parts that make up the regulator 23 or an error in the command pressure supplied to the pilot ports 24K and 24L, the spool 28 can stably maintain the neutral position. As a result, the neutral position of the spool 28 and the tilt angle of the swash plate 11 can be stably matched, and the tilt angle of the swash plate 11 can be stabilized when the tilt angle is close to zero.

In addition, when the command pressure Pp supplied to the pilot oil chamber 38 of the regulator 23 exceeds the pressure Pp0 corresponding to the set load of the spring mechanism 39 (compression spring 46), servo pressure is supplied to the second hydraulic chamber 17D of the tilt cylinder 17, and the servo piston 20 moves from position S1. Here, while the servo piston 20 is displaced from position S1 to position S2, the servo piston 20 moves the amount of the clearance formed between the servo piston 20 and the spring guide 21B in the first hydraulic chamber 17C of the tilt cylinder 17, so it is not subjected to the spring force of the first servo spring 21.

At this time, the servo piston 20 is subjected to the pressure difference ΔPs between the command pressure in the first hydraulic chamber 17C and the command pressure in the second hydraulic chamber 17D, and the external force F applied to the swash plate 11 due to the hydraulic force from the piston 8, inertial force, etc. Therefore, when the servo piston 20 is between position S1 and position S2 and the target tilt angle of the swash plate 11 is a relatively small tilt angle close to zero, the servo piston 20 stops at a position where the pressure difference ΔPs and the external force F are balanced (ΔPs = F), and the swash plate 11 changes to the target tilt angle.

Furthermore, when the command pressure Pp supplied to the pilot oil chamber 38 of the regulator 23 exceeds the pressure corresponding to the set load of the first servo spring 21 (compression spring 21C) of the tilt cylinder 17, the servo piston 20 moves from position S2 in the direction of arrow C while compressing the compression spring 21C. At this time, the servo piston 20 is subjected to the differential pressure ΔPs between the command pressure in the first hydraulic pressure chamber 17C and the command pressure in the second hydraulic pressure chamber 17D, the hydraulic force from the piston 8, the external force F applied to the swash plate 11 due to inertial force, etc., and the spring force Kx of the compression spring 21C (the product of the spring constant K and the deflection amount x of the compression spring 21C). Therefore, when the servo piston 20 is in a position beyond position S2 and the target tilt angle of the swash plate 11 is relatively large, the servo piston 20 stops at a position where the differential pressure ΔPs, the sum of the external force F and the spring force Kx of the compression spring 21C are balanced (ΔPs = F + Kx), and the swash plate 11 changes to the target tilt angle.

In this embodiment, the amount of movement of the servo piston 20 in the second dead zone set by the set load of the first servo spring 21 is set to be greater than the amount of movement of the servo piston 20 in the dead zone set by the set load of the spring mechanism 39. As a result, when the servo piston 20 is between position S1 and position S2, the servo piston 20 can change the tilt angle of the swash plate 11 based on the balance between the differential pressure ΔPs between the command pressure in the first hydraulic pressure chamber 17C and the command pressure in the second hydraulic pressure chamber 17D and the external force F applied to the swash plate 11, regardless of the spring force of the first servo spring 21 (compression spring 21C). Therefore, even if there is an error (variation) in the spring force Kx of the first servo spring 21 due to a manufacturing error of the compression spring 21C, in the range where the target tilt angle of the swash plate 11 is a relatively small tilt angle close to zero, it is not affected by the error in the spring force Kx of the first servo spring 21. As a result, the tilt angle of the swash plate 11 can be stably controlled within a relatively small range from zero (when the tilt angle is close to zero).

In this way, when the servo piston 20 moves in the direction of the arrow C, the movement of the servo piston 20 is transmitted to the sleeve 27 of the regulator 23 via the feedback link 47. That is, the regulator 23 is feedback-controlled so that the movement of the servo piston 20 is transmitted via the feedback link 47 to slide and displace the sleeve 27 in the same direction as the spool 28. As a result, the sleeve 27 is feedback-controlled to a position where the oil holes 27C, 27D are blocked (closed) from the spool sliding hole 27A by the lands 28F, 28G of the spool 28. In this way, by stopping the supply of servo pressure from the regulator 23 to the tilt actuator 16, the servo piston 20 can be held at any position, and the swash plate 11 of the swash plate type hydraulic pump 1 can be held in a state tilted in one direction from the neutral position.

On the other hand, when the swash plate 11 of the swash plate type hydraulic pump 1 is driven to tilt in the other direction from the neutral position, the tilt operation valve 54B is held in the return position (a) and the tilt operation valve 54A is switched from the return position (a) to the supply position (b). This causes the command pressure (pilot pressure) Pp to be supplied to the pilot oil chamber 32 of the regulator 23, and the pilot oil chamber 38 is maintained at tank pressure. Therefore, the piston 31 receives the command pressure Pp in the pilot oil chamber 32 and presses the end 28E on the small diameter shaft portion 28B side of the spool 28.

In this way, when a pressing force acts from the piston 31 on the spool 28 in the direction of arrow B in FIG. 3, a spring force acts from the compression spring 46 of the spring mechanism 39 on the spool 28 in the opposite direction (arrow A direction) to the pressing force of the piston 31. Therefore, when the pressing force of the piston 31, i.e., the command pressure Pp in the pilot oil chamber 32, is smaller than the set load of the compression spring 46, the spool 28 does not move. Then, when the command pressure Pp in the pilot oil chamber 32 becomes equal to or greater than the set load of the compression spring 46, the spool 28 moves in the direction of arrow B against the spring force of the compression spring 46.

This allows servo pressure to be supplied from the regulator 23 to the first hydraulic chamber 17C of the tilt cylinder 17, and the servo piston 20 can be moved in the direction of arrow D in FIG. 3. The movement of the servo piston 20 is transmitted to the sleeve 27 of the regulator 23 via the feedback link 47, and the sleeve 27 is feedback-controlled to a position where the oil holes 27C, 27D are blocked from the spool sliding hole 27A by the lands 28F, 28G of the spool 28. In this way, by stopping the supply of servo pressure from the regulator 23 to the tilt actuator 16, the servo piston 20 can be held at any position, and the swash plate 11 of the swash plate type hydraulic pump 1 can be held in a state tilted in the other direction from the neutral position.

Thus, in the embodiment, the variable displacement hydraulic rotary machine includes a swash plate 11 that changes the discharge capacity according to the tilt angle, a tilt cylinder 17 and a servo piston 20 that forms a first hydraulic pressure chamber 17C and a second hydraulic pressure chamber 17D in the tilt cylinder 17, a tilt actuator 16 that changes the tilt angle of the swash plate 11 by the servo piston 20 that moves in the tilt cylinder 17 due to the pressure difference between the pressure in the first hydraulic pressure chamber 17C and the pressure in the second hydraulic pressure chamber 17D, and a regulator 23 that adjusts the pressure supplied to the first hydraulic pressure chamber 17C or the second hydraulic pressure chamber 17D according to a command pressure to move the servo piston 20. The variable displacement hydraulic rotary machine is characterized by having a dead zone in which the servo piston 20 maintains a constant position relative to the tilt cylinder 17 regardless of the magnitude of the command pressure until the command pressure supplied to the regulator 23 reaches a predetermined pressure.

With this configuration, in the dead zone where the command pressure supplied to the regulator 23 ranges from zero to a predetermined pressure, the servo piston 20 maintains a constant position within the tilt cylinder 17. This allows the tilt angle of the swash plate 11 to be maintained constant, and the neutral position of the regulator 23 (e.g., spool 28) and the tilt angle of the swash plate 11 can be stably matched.

In the embodiment, the regulator 23 is provided with a spool 28 that is movably provided in the regulator casing 24 via a sleeve 27 and that adjusts the pressure supplied to the first hydraulic chamber 17C or the second hydraulic chamber 17D by moving in response to a command pressure, pistons 31, 35 that are provided in the regulator casing 24 and that press the spool 28 against the sleeve 27 by the command pressure supplied to the regulator casing 24, and a spring mechanism 39 that is provided between the regulator casing 24 and the spool 28 and that applies a spring force to the spool 28 in the opposite direction to the pressing force from the pistons 31, 35, and the dead zone is set by the set load of the spring mechanism 39. With this configuration, the tilt angle of the swash plate 11 can be kept constant while the command pressure supplied to the regulator 23 is within the dead zone set by the set load of the spring mechanism 39.

In the embodiment, the spring mechanism 39 includes an adjuster 33 having a spool accommodation hole 33E that accommodates one axial end side of the spool 28 and is fixed in the regulator casing 24 so as to be adjustable in position, a spool side engagement member (flange portion 28C and shaft retaining ring 40) provided on the spool 28 within the spool accommodation hole 33E, an adjuster side engagement member (annular step portion 33G and hole retaining ring 41) provided on the adjuster 33 within the spool accommodation hole 33E, a pair of spring receivers (one-side spring receiver 42 and other-side spring receiver 44) that are attached to the spool 28 within the spool accommodation hole 33E and are movable in the axial direction of the spool 28 between the spool side engagement member and the adjuster side engagement member, and a compression spring 46 compressed between the pair of spring receivers. According to this configuration, a single compression spring 46 provided on one axial end of the spool 28 can apply spring forces in opposite directions to the pressing force acting on the spool 28 from the piston 31 and the pressing force acting on the spool 28 from the piston 35. Therefore, the dead zone of the command pressure supplied to the regulator 23 to drive the piston 31 or the piston 35 can be set equally by the set load of the compression spring 46.

In the embodiment, the tilt actuator 16 has a first servo spring 21 that presses the servo piston 20 toward the second hydraulic pressure chamber 17D and a second servo spring 22 that presses the servo piston 20 toward the first hydraulic pressure chamber 17C, and has a second dead zone in which the servo piston 20 maintains a constant position relative to the tilt cylinder 17 by the set loads of the first servo spring 21 and the second servo spring 22 until the pressure difference between the pressure in the first hydraulic pressure chamber 17C and the pressure in the second hydraulic pressure chamber 17D reaches a predetermined pressure. According to this configuration, when the command pressure supplied to the regulator 23 exceeds the dead zone set by the set load of the spring mechanism 39, the spool 28 moves relative to the sleeve 27, and servo pressure is supplied to the first hydraulic pressure chamber 17C or the second hydraulic pressure chamber 17D. At this time, if the pressure difference between the pressure in the first hydraulic chamber 17C and the pressure in the second hydraulic chamber 17D is within the second dead zone, the servo piston 20 moves within the tilt cylinder 17 without receiving the spring force of the first servo spring 21 or the second servo spring 22, and the tilt angle of the swash plate 11 can be changed.

In the embodiment, the amount of movement of the servo piston 20 in the second dead zone is greater than the amount of movement of the servo piston 20 in the dead zone. According to this configuration, from the time when the servo piston 20 starts to move in the tilt cylinder 17 until it receives the spring force of the first servo spring 21 or the second servo spring 22, that is, when the tilt angle of the swash plate 11 is changed from zero to a relatively small tilt angle, the servo piston 20 can tilt the swash plate 11 regardless of the spring forces of the first servo spring 21 and the second servo spring 22. As a result, in a range where the tilt angle of the swash plate 11 is relatively small (when the tilt angle is close to zero), the tilt angle of the swash plate 11 can be stably controlled.

In the embodiment, a variable displacement swash plate hydraulic pump 1 is exemplified as a variable displacement hydraulic rotating machine equipped with a tilt actuator 16, a regulator 23, etc. However, the present invention is not limited to this, and can be applied to, for example, a variable displacement swash plate hydraulic motor, and can also be applied to a variable displacement bent axis hydraulic pump and a bent axis hydraulic motor.

1. Swash plate type hydraulic pump (variable displacement hydraulic rotary machine)
11 swash plate (tilting member)
16 Tilt actuator 17 Tilt cylinder 17C First hydraulic pressure chamber 17D Second hydraulic pressure chamber 20 Servo piston 21 First servo spring 22 Second servo spring 23 Regulator 24 Regulator casing 27 Sleeve 28 Spool 28C Flange (spool side engagement member)
31, 35 Piston 33 Adjuster 33E Spool receiving hole 33G Annular step portion (adjuster side engagement member)
39 Spring mechanism 40 Shaft retaining ring (spool side engagement member)
41 Hole retaining ring (adjuster side engagement member)
42 One side spring support (spring support)
44 Other side spring holder (spring holder)
46 Compression spring

Claims (5)

a tilting member that changes a discharge capacity according to a tilting angle;
a tilt actuator having a tilt cylinder and a servo piston which defines a first hydraulic pressure chamber and a second hydraulic pressure chamber within the tilt cylinder, the servo piston moving within the tilt cylinder depending on a pressure difference between the first hydraulic pressure chamber and the second hydraulic pressure chamber, and which changes the tilt angle of the tilt member by the servo piston;
a regulator that adjusts a pressure supplied to the first hydraulic chamber or the second hydraulic chamber in response to a command pressure to move the servo piston,
a dead zone in which the servo piston maintains a constant position relative to the tilt cylinder until the command pressure supplied to the regulator reaches a predetermined pressure. the regulator includes a spool that is movably provided within a regulator casing via a sleeve, the spool moving in response to the command pressure to adjust the pressure supplied to the first hydraulic pressure chamber or the second hydraulic pressure chamber;
a piston provided in the regulator casing, the piston pressing the spool against the sleeve by the command pressure supplied into the regulator casing;
a spring mechanism provided between the regulator casing and the spool for applying a spring force to the spool in a direction opposite to the pressing force from the piston,
2. The variable displacement hydraulic rotary machine according to claim 1, wherein the dead zone is set by a set load of the spring mechanism. the spring mechanism includes an adjuster having a spool receiving hole that receives one end side of the spool in the axial direction and is fixed in the regulator casing so as to be positionally adjustable;
a spool side engaging member provided on the spool within the spool receiving hole;
an adjuster side engagement member provided on the adjuster within the spool receiving hole;
a pair of spring bearings attached to the spool in the spool receiving hole and movable in the axial direction of the spool between the spool side engaging member and the adjuster side engaging member;
3. The variable displacement hydraulic rotary machine according to claim 2, further comprising a compression spring compressed between said pair of spring bearings. The tilt actuator includes a first servo spring that presses the servo piston toward the second hydraulic chamber;
a second servo spring that presses the servo piston toward the first hydraulic chamber,
4. A variable displacement hydraulic rotary machine according to claim 1, 2 or 3, characterized in that the servo piston has a second dead band region in which the servo piston is maintained at a constant position relative to the tilt cylinder by the set loads of the first servo spring and the second servo spring until a pressure difference between the pressure in the first hydraulic chamber and the pressure in the second hydraulic chamber reaches a predetermined pressure. The variable displacement hydraulic rotary machine according to claim 4, characterized in that the amount of movement of the servo piston within the second dead zone is greater than the amount of movement of the servo piston within the dead zone.

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