JP2008248705A - Horsepower control regulator, horsepower control device, and piston pump - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a horsepower regulator achieving equivalent horsepower characteristics of a swash plate type piston pump, on both of a high pressure side and a low pressure side of other pump. <P>SOLUTION: The horsepower regulator in a swash plate type piston pump 1 driven by engine output together with other pump, comprises: a valve housing 42; a control spool 35 disposed inside the valve housing, and moving with respect to the valve housing to adjust fluid pressure lead by a tilting actuator; a hole 72 in which discharge pressure of the swash plate type piston pump is lead to move the control spool in the control spool; a sleeve 81 disposed between the valve housing 42 and the control spool; and a port 64 to which signal pressure Pi is supplied according to the fluid pressure P3 from the other pump, and which applies the force by the signal pressure to the sleeve to move the sleeve. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a swash plate type piston pump, and its horsepower control regulator and horsepower control device.

2. Description of the Related Art Conventionally, a swash plate type piston pump including a horsepower control regulator that controls discharge pressure and discharge flow rate with equal horsepower characteristics so that horsepower (that is, output) becomes substantially constant is known (see Patent Document 1). The swash plate type piston pump is used in a hydraulic machine such as a mini excavator, and the swash plate type piston pump is driven by an output from an engine of the hydraulic machine.
JP 2002-202063 A

  However, in a system that combines a swash plate piston pump with another pump (gear pump, etc.) driven by the engine output, the pressure of the other pump is on the high pressure side (or low pressure side), and the equal horsepower of the swash plate piston pump On the other hand, if the characteristic is established, the equi-horsepower characteristic of the swash plate type piston pump will not be established on the low pressure side (or high pressure side) of the other pump, and the engine output cannot be used effectively. Occurs. This is because the horsepower characteristics (relationship between hydraulic pressure and discharge flow rate) of the swash plate type piston pump move horizontally in accordance with the pressures of other pumps.

  The present invention has been made in view of the above problems, and an object of the present invention is to provide a horsepower control regulator that establishes equal horsepower characteristics of a swash plate type piston pump on both the high-pressure side and the low-pressure side of another pump. And

  The present invention relates to a swash plate type piston pump that is driven together with another pump by engine output, and includes a cylinder block having a cylinder that houses a piston, and a swash plate that reciprocates the piston as the cylinder block rotates. A horsepower control regulator used in a swash plate type piston pump having a tilting actuator that tilts the swash plate by fluid pressure, and is slidably disposed inside the valve housing and inside the valve housing, A control spool for adjusting a fluid pressure guided to the tilting actuator according to a movement position with respect to the valve housing, and a hole for guiding a discharge pressure of the swash plate type piston pump in the control spool, and depending on the discharge pressure To apply force to the control spool and A hole for moving the control spool, a through hole for passing the discharge pressure of the swash plate type piston pump to the hole, a sleeve disposed between the valve housing and the control spool, and a fluid from the other pump A port that is supplied with a signal pressure in accordance with the pressure, applies a force generated by the signal pressure to the sleeve substantially parallel to a force in the spool movement direction applied by the discharge pressure, and moves the sleeve relative to the valve housing. The fluid pressure guided to the tilting actuator is corrected according to the relative position between the sleeve and the control spool.

  According to the invention, the equi-horsepower characteristics of the swash plate piston pump can be realized not only on the high-pressure side but also on the low-pressure side of other pumps (gear pumps, etc.), and the equi-horsepower characteristics of the swash plate-type piston pump over a wide range of hydraulic pressure from other pumps realizable.

  Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

  FIG. 1 shows an overall view of a swash plate type piston pump according to the first embodiment. The configuration of the swash plate type piston pump other than the regulator is the same as that described in Japanese Patent Laid-Open No. 2002-202063, and will not be described in detail here.

  As shown in FIG. 1, in the swash plate type piston pump 1, a cylinder block 3 and a swash plate 4 are accommodated in an internal space formed by a pump housing 2 and a cover 10. The cylinder block 3 is rotationally driven via a shaft 5. The rotation of the shaft 5 is transmitted from an engine (not shown) as a power source to one end thereof. In addition, this engine rotationally drives a gear pump and the like in addition to the piston pump 1.

In the cylinder block 3, a plurality of cylinders are arranged in parallel with the rotation axis O and arranged side by side at a constant interval on substantially the same circumference around the rotation axis O.
When the cylinder block 3 rotates, the piston of the cylinder reciprocates between the swash plate 4 and sucks and discharges hydraulic oil (fluid).

  In order to make the discharge amount of the piston pump 1 variable, the swash plate 4 is tiltably supported by the pump housing 2. As a tilt actuator that changes the tilt angle θ of the swash plate 4, a small-diameter piston (not shown) that drives the tilt angle θ in a direction that increases, and a large drive that drives the swash plate 4 in a direction that decreases the tilt angle θ. A diameter piston 23 is provided. In the large-diameter piston 23, the tilt angle θ of the swash plate 4 decreases as the hydraulic pressure Pc (fluid pressure) guided from the regulator 30 to the pressure chamber 28 decreases, that is, the discharge amount Q (discharge flow rate) decreases. Driven by. The large-diameter piston 23 is supported so as to be slidable substantially parallel to the rotation axis O through a guide sleeve fixed to the pump housing 2. In the pump housing 2, there is a feedback pin 32 that follows the swash plate 4 and is displaced substantially parallel to the rotation axis O of the cylinder block 3.

  As shown in FIG. 2, the regulator 30 constituting the horsepower control device adjusts the hydraulic pressure Pc guided to the large-diameter piston 23 in conjunction with the feedback pin 32. The feedback pin 32 and the regulator 30 are disposed on an axis L substantially parallel to the rotation axis O. The feedback pin 32 is slidably supported in the hole 2A of the pump housing 2, and its end surface abuts on the ball of the swash plate 4 or the cylindrical pin 31, and follows the tilt of the swash plate 4 in the direction of the axis L. Displace.

  The regulator 30 controls the horsepower (ie, output) of the piston pump 1 by adjusting the hydraulic pressure Pc guided to the large-diameter piston 23 according to the discharge pressure P of the piston pump 1. The regulator 30 controls the outer and inner control springs (springs) 33 and 34 (that is, elastic bodies) that press the feedback pin 32 against the swash plate 4 side, and the hydraulic pressure Pc guided to the pressure chamber 28 of the large-diameter piston 23. A spool 35, a stepped shaft portion 36, and a sleeve 81 are provided, and these are arranged on the shaft L in series. The control spool 35, the stepped shaft portion 36, the valve housing 42, and the sleeve 81 are arranged coaxially, and their axes are the axis L. The control spool 35 is slidably mounted on the cylindrical sleeve 81. A hole in the direction of the axis L is provided inside the control spool 35, and a stepped shaft portion 36 is disposed in this hole, and the control spool 35 can also slide with respect to the stepped shaft portion 36. The sleeve 81 is disposed between the control spool 35 and the valve housing 42 and is slidable with respect to the control spool 35 and the valve housing 42. The spring 82 is disposed in contact with the rear end 81 b of the sleeve 81 in the direction of the axis L, and is interposed between the valve housing 42 and the sleeve 81. The spring 82 expands and contracts in the direction of the axis L according to the signal pressure Pi applied to the sleeve 81 from the sleeve pressing port 64. The signal pressure Pi is a hydraulic pressure supplied by a signal pressure supply system (signal pressure supply means) (not shown) according to the gear pump hydraulic pressure P3 (the hydraulic pressure from the gear pump or the discharge pressure of the gear pump), and the increase in the gear pump hydraulic pressure P3. It increases according to. The signal pressure Pi may increase in proportion to the gear pump hydraulic pressure P3, and can be simply obtained by reducing or increasing the gear pump hydraulic pressure P3 by a known means. Most simply, the signal pressure Pi may be equal to the gear pump hydraulic pressure P3, and the gear pump hydraulic pressure P3 may be guided to the sleeve pressing port 64 and used as the signal pressure Pi.

  The coiled outer and inner control springs 33, 34 are interposed between a retainer 37 that engages the spherical end of the feedback pin 32 and a flange 38 formed separately from the control spool 35. . An inner control spring 34 (second spring) having a small winding diameter is disposed inside an outer control spring 33 (first spring) having a large winding diameter. The natural length (free length) of the outer control spring 33 is longer than the natural length of the inner control spring 34. In a state where the tilt angle θ of the swash plate 4 is maximized, the outer control spring 33 is interposed between the retainer 37 and the flange portion 38 in a compressed state, while the inner control spring 34 has a retainer at one end. 37 or the flange portion 38 is interposed in a floating state. That is, as long as the outer control spring 33 is long, only the outer control spring 33 contracts, and when the outer control spring 33 contracts beyond the natural length of the inner control spring 34, the outer and inner control springs 33 and 34 Both are compressed. As a result, as the swash plate 4 tilts beyond a predetermined angle, both ends of the inner control spring 34 come into contact with the retainer 37 and the collar 38 and are compressed, and are applied to the swash plate 4 via the feedback pin 32. The elastic force from the inner and outer control springs 33 and 34 is increased stepwise.

  Four ports (a first port 51, a second port 52, a third port 53, and a fourth port 54) are formed on the outer periphery of the valve housing 42. The ports 51, 52, 53, 54 can communicate with the port of the sleeve 81 through the first, second, third, and fourth communication holes 51a, 52a, 53a, 54a of the valve housing 42, respectively. .

  The first port 51 is supplied with a gear pump hydraulic pressure P3 from a gear pump driven by the engine. The second port 52 communicates with the discharge port and is supplied with the discharge pressure P of the piston pump 1. The discharge pressure P is also supplied to the hydraulic chamber of the small diameter piston. The third port 53 communicates with the pressure chamber 28 of the large diameter piston 23 and supplies the hydraulic pressure Pc thereto. A signal pressure Pi is applied to the fourth port 54 according to the gear pump hydraulic pressure P3 by a signal pressure supply system (not shown).

  Four ports (a fifth port 91, a sixth port 92, a seventh port 93, and a sleeve pressing port 64) are formed on the outer periphery of the sleeve 81. The fifth port 91, the sixth port 92, the seventh port 93, and the sleeve pressing port 64 are the first communication hole 51a, the second communication hole 52a, the third communication hole 53a, and the fourth communication hole. 54a communicates with each other. A signal pressure Pi is supplied to the sleeve pressing port 64 via the fourth port 54 and the communication hole 54a. The sleeve 81 is provided with a first through hole 91a, a second through hole 92a, and a third through hole 93a that penetrate the sleeve 81. The first through hole 91 a, the second through hole 92 a, and the third through hole 93 a communicate with the fifth port 91, the sixth port 92, and the seventh port 93.

  In the sleeve pressing port 64, the sleeve 81 receives force by the signal pressure Pi according to the difference between the pressure receiving area on the front end 81a side and the pressure receiving area on the rear end 81b side. In the present embodiment, in the sleeve pressing port 64, since the pressure receiving area on the front end 81a side with respect to the signal pressure Pi is smaller than the pressure receiving area on the rear end 81b side, the direction of the force is the direction toward the rear end 81b (the sleeve 81 is connected to the valve). The direction of drawing into the housing 42). Accordingly, the force applied by the signal pressure Pi is substantially parallel to and opposite to the force in the moving direction of the spool applied by the discharge pressure P. For simplicity, in the sleeve pressing port 64, the pressure receiving area on the tip 81a side of the sleeve 81 with respect to the signal pressure Pi may be zero.

  A first oil groove 61, a second oil groove 62, and a third oil groove 63 are formed on the outer periphery of the control spool 35. The control spool 35 is provided with a first through hole 71, a second through hole 72, and a third through hole 73 that penetrate the control spool 35. The first through hole 71, the second through hole 72, and the third through hole 73 communicate with the first oil groove 61, the second oil groove 62, and the third oil groove 63. The third through hole 73 is connected to the inner hole 74 of the control spool 35 in the axis L direction. Tank pressure Pt is supplied to the inner hole 74. Further, on the bottom 35 b side of the control spool 35, an oil chamber 56 to which the tank pressure Pt is supplied via the inner hole 45 of the valve housing 42 is provided between the valve housing 42 and the control spool 35.

  In the second through-hole 72, the pressure receiving area of the control spool 35 with respect to the discharge pressure P of the piston pump 1 is equal to the area of the front end portion 36a of the stepped shaft portion 36 from the bottom 35b side of the control spool 35 to the front end 35a side. growing. This is because the hole of the control spool 35 corresponding to the stepped shaft portion 36 reaches the second through hole 72. As a result, the control spool 35 receives a force toward the front end 35 a in the direction of the axis L (force in a direction of pushing the control spool 35 from the valve housing 42) due to the discharge pressure P.

  The diameter of the stepped shaft portion 36 is smaller at the tip than at the root portion, and between the first oil groove 61 and the second oil groove 62 and between the first oil groove 61 and the bottom portion 35b. The inner diameter of the control spool 35 decreases. Therefore, in the first through hole 71, the pressure receiving area of the control spool 35 with respect to the gear pump hydraulic pressure P3 is larger on the tip 35a side than on the bottom 35b side of the control spool 35. For this reason, the control spool 35 receives a force toward the tip 35a in the direction of the axis L (force to push the control spool 35 from the valve housing 42) by the gear pump hydraulic pressure P3.

  The control spool 35 is switched between the first and second open positions and the closed position (steady position) as shown in the figure by sliding in the direction of the axis L. The hydraulic pressure (fluid pressure) Pc guided to the tilt actuator is corrected in accordance with the relative position between the control spool 35 and the sleeve 81. In the first open position, the control spool 35 communicates the second through hole 92a and the third through hole 93a with the second oil groove 62, and supplies the discharge pressure P to the pressure chamber 28 of the large diameter piston 23. Lead. In the second open position, the control spool 35 communicates the third through-hole 93a and the third through-hole 73 by the third oil groove 63 to the pressure chamber 28 of the large-diameter piston 23 in the pump housing 2. The tank pressure Pt is derived. On the other hand, in the closed position, the communication between the second through hole 92a and the third through hole 93a is blocked by the land 68 of the control spool 35 (located between the second and third oil grooves 62, 63). In the closed position, the resultant force applied to the control spool 35 by the hydraulic pressure (that is, the discharge pressure P and the gear pump hydraulic pressure P3) and the elastic force of the control springs 33 and 34 are balanced.

  When the signal pressure Pi supplied to the sleeve pressing port 64 increases in accordance with the increase in the gear pump hydraulic pressure P3, the sleeve 81 is pulled toward the rear end 81b, that is, into the valve housing 42 while contracting the spring (elastic body) 82. Move in the direction (left side of FIG. 2). At this time, if the inner control spring 34 is in a floating state, the difference dL between the free length of the inner control spring 34 and the length of the outer control spring 33 increases. Conversely, when the signal pressure Pi supplied to the sleeve pressing port 64 decreases in accordance with the decrease in the gear pump hydraulic pressure P3, the sleeve 81 is pushed out toward the front end 81a, that is, from the valve housing 42 while extending the spring 82 ( Move to the right side of FIG. At this time, if the inner control spring 34 is in a floating state, the difference dL between the free length of the inner control spring 34 and the length of the outer control spring 33 decreases.

  Thus, the sleeve 81 takes a position corresponding to the signal pressure Pi in the direction of the axis L. Further, since the signal pressure Pi increases (decreases) in accordance with the increase (decrease) of the gear pump hydraulic pressure P3, the sleeve 81 takes a position in the axis L direction according to the gear pump hydraulic pressure P3.

  While the control spool 35 performs the control operation of the hydraulic pressure Pc guided to the pressure chamber 28 of the large-diameter piston 23, the positional relationship between the sleeve 81 and the closed position (steady position) of the control spool 35 is maintained substantially constant. Therefore, the closed position (steady position) of the control spool 35 moves in the same direction as the sleeve 81 as the sleeve 81 moves according to the gear pump hydraulic pressure P3 (movement according to the signal pressure Pi). That is, when the gear pump hydraulic pressure P3 increases, the signal pressure Pi increases, and the closed position (steady position) of the control spool 35 is directed toward the bottom 35b, that is, the direction in which the control spool 35 is pulled into the valve housing 42 (see FIG. 2). Move to the left). Conversely, when the gear pump hydraulic pressure P3 decreases, the signal pressure Pi decreases, and the closed position (steady position) of the control spool 35 is directed toward the tip 35a, that is, the direction in which the control spool 35 is pushed out from the valve housing 42 (FIG. 2). Move to the right).

  When the discharge pressure P of the piston pump 1 rises and the resultant force applied to the control spool 35 increases beyond the elastic force of the control springs 33 and 34, the control spool 35 compresses the control springs 33 and 34 while rotating the shaft. Sliding in the direction and switching from the closed position to the first open position. Then, the discharge pressure P is guided to the pressure chamber 28 of the large-diameter piston 23, and the large-diameter piston 23 tilts the swash plate 4 in the direction of decreasing the tilt angle θ, and the discharge amount Q of the piston pump 1 decreases. To do. When the large-diameter piston 23 tilts the swash plate 4, conversely, the feedback pin 32 moves to push the control springs 33, 34, that is, to push the control spool 35 back from the first open position to the closed position. To do. When in the closed position, the introduction of the discharge pressure P to the large-diameter piston 23 is stopped, and the tilt of the swash plate 4 is stopped. When the control spool 35 is pushed back to the closed position, the resultant force applied to the control spool 35 balances with the elastic force of the control springs 33 and 34 while the control springs 33 and 34 are contracted from before the discharge pressure P increases. Yes.

  On the contrary, when the discharge force P of the piston pump 1 decreases and the resultant force applied to the control spool 35 becomes lower than the elastic force of the control springs 33 and 34, the control spool 35 extends the control springs 33 and 34. It slides in the axial direction and switches from the closed position to the second open position. Then, the tank pressure Pt is guided to the pressure chamber 28 of the large-diameter piston 23, the large-diameter piston 23 moves in the retracting direction, and the swash plate 4 is tilted in the direction in which the tilt angle θ is increased. The discharge amount Q increases. When the large-diameter piston 23 tilts the swash plate 4, conversely, the feedback pin 32 moves away from the control springs 33 and 34, that is, in a direction to push the control spool 35 back from the second open position to the closed position. To do. When in the closed position, the introduction of the tank pressure Pt to the large-diameter piston 23 is stopped, and the tilt of the swash plate 4 is stopped. When the control spool 35 is returned to the closed position, the resultant force applied to the control spool 35 is balanced with the elastic force of the control springs 33 and 34 in a state where the control springs 33 and 34 are extended from before the discharge pressure P is reduced. ing.

  Thus, the pump discharge amount Q (discharge flow rate) decreases as the discharge pressure P of the piston pump 1 increases, and the pump discharge amount Q increases as the discharge pressure P decreases. Further, the elastic force generated from the control springs 33 and 34 increases stepwise as the discharge pressure P increases. For this reason, it is possible to realize an equal horsepower characteristic (a characteristic in which the product of the discharge pressure P and the discharge amount Q is substantially constant) in which the relationship between the discharge pressure P and the discharge amount Q of the piston pump 1 is approximately inversely proportional.

  Here, the influence of the gear pump hydraulic pressure P3 on the horsepower characteristics (that is, the relationship between the discharge pressure P and the discharge amount Q) will be described. In the prior art of FIG. 3, since there is no sleeve pressing port 64 and sleeve 81, when the gear pump hydraulic pressure P3 decreases, the horsepower characteristic of the piston pump 1 moves horizontally to the high pressure side along the P axis of FIG. When the hydraulic pressure P3 increases, the horsepower characteristic of the piston pump 1 moves horizontally along the P axis to the low pressure side. This is because the gear pump hydraulic pressure P3 and the discharge pressure P of the piston pump 1 apply the same direction force to the control spool 35, so that the same tilt angle θ (that is, the same discharge amount Q) is obtained with the low gear pump hydraulic pressure P3. This is because a higher discharge pressure P is required than in the case of the high gear pump hydraulic pressure P3 in order to achieve this. As a result, when the gear pump hydraulic pressure P3 is low, the discharge pressure P of the piston pump 1 increases and the horsepower can be increased. However, since the horsepower characteristic of the piston pump 1 moves horizontally, if the high-pressure gear pump hydraulic pressure P3 is set so that the equal horsepower characteristic of the piston pump 1 is established, the low-pressure gear pump hydraulic pressure P3 has the equal horsepower characteristic of the piston pump 1. There is a problem that cannot be realized. Conversely, if the low-pressure gear pump hydraulic pressure P3 is set so that the equal horsepower characteristics of the piston pump 1 are established, the high-pressure gear pump hydraulic pressure P3 cannot realize the equal horsepower characteristics of the piston pump 1.

  However, in the present invention shown in FIG. 4, since the sleeve pressing port 64 and the sleeve 81 are present, the horsepower characteristic moves diagonally when the gear pump hydraulic pressure P3 decreases or increases. This is because, in order to realize the same tilt angle θ (that is, the same discharge amount Q) at the low gear pump hydraulic pressure P3, a higher piston pump discharge pressure P is required than in the case of the high gear pump hydraulic pressure P3. In addition, when the gear pump hydraulic pressure P3 decreases, the closed position (steady position) of the control spool 35 moves in a direction to be pushed out from the valve housing 42, and the control spool 35 moves the collar portion 38 in a direction to contract the outer control spring 33. It is for pushing and moving. Bending point in the horsepower characteristic (point A in FIG. 4) where the length of the outer control spring 33 contracts beyond the natural length of the inner control spring 34 and both the outer and inner control springs 33 and 34 begin to compress. Therefore, since the outer control spring 33 is contracted at the lower gear pump hydraulic pressure P3 than at the higher gear pump hydraulic pressure P3, the change dQ (inclination) from the maximum discharge amount (maximum tilt angle) to the bending point. The change in angle (dθ) is decreasing (dQ1> dQ2 in FIG. 4). For this reason, the bending point (point A in FIG. 4) moves diagonally on the PQ plane, and the horsepower characteristic also moves diagonally. Thereby, the equi-horsepower characteristic of the swash plate type piston pump 1 can be realized not only on the high-pressure side but also on the low-pressure side of the gear pump oil pressure P3, and the equi-horsepower characteristic of the swash plate type piston pump 1 can be realized over a wide range of the gear pump oil pressure P3.

  Conversely, when the gear pump hydraulic pressure P3 increases, the closed position (steady position) of the control spool 35 moves in a direction in which the control spool 35 is pulled into the valve housing 42, and the control spool 35 extends in the direction in which the outer control spring 33 is extended. The collar part 38 is moved. Considering the bending point in the horsepower characteristics (point A in FIG. 4), since the outer control spring 33 extends at the higher gear pump hydraulic pressure P3 than at the lower gear pump hydraulic pressure P3, the maximum discharge amount (maximum tilt angle). The change in discharge amount dQ (change in tilt angle dθ) from to the bending point increases. For this reason, the bending point (point A in FIG. 4) moves diagonally on the PQ plane, and the horsepower characteristic also moves diagonally. Thereby, the equi-horsepower characteristic of the swash plate type piston pump 1 can be realized not only on the low pressure side but also on the high pressure side of the gear pump oil pressure P3, and the equihorse force characteristic of the swash plate type piston pump 1 can be realized over a wide range of the gear pump oil pressure P3.

  The present invention is not limited to the above-described embodiment, and it is obvious that various modifications can be made within the scope of the technical idea.

  INDUSTRIAL APPLICABILITY The present invention can be used for a horsepower control regulator or a horsepower control device of a piston pump in a hydraulic machine having a swash plate type piston pump that can be driven by engine output and other pumps (gear pumps, etc.).

It is sectional drawing of the whole swash plate type piston pump. It is sectional drawing of the horsepower control regulator which concerns on this invention. It is a figure which shows the horsepower control characteristic (relationship between discharge pressure and discharge flow rate) of the swash plate type piston pump which concerns on a prior art. It is a figure which shows the horsepower control characteristic (relationship of discharge pressure and discharge flow rate) of the swash plate type piston pump which concerns on this invention.

Explanation of symbols

33 Outer control spring 34 Inner control spring 35 Control spool 36 Stepped shaft 42 Valve housing 51, 52, 53, 54 First, second, third, fourth port 51a, 52a, 53a, 54a First, second Second, third, and fourth communication holes 61, 62, 63 First, second, and third oil grooves 64 Sleeve pressing ports 71, 72, 73 First, second, and third through holes 81 Sleeve 82 Spring 91, 92, 93 Fifth, sixth, seventh ports 91a, 92a, 93a First, second, third through holes

Claims (8)

A swash plate type piston pump driven together with other pumps by engine output, a cylinder block having a cylinder for housing the piston, a swash plate for reciprocating the piston as the cylinder block rotates, and a fluid pressure A horsepower control regulator used in a swash plate type piston pump having a tilt actuator for tilting the swash plate by
A valve housing;
A control spool that is slidably disposed inside the valve housing and adjusts a fluid pressure guided to the tilting actuator according to a moving position with respect to the valve housing;
A hole through which the discharge pressure of the swash plate type piston pump is guided in the control spool, and applies a force to the control spool in accordance with the discharge pressure to move the control spool with respect to the valve housing;
A sleeve having a through hole for passing the discharge pressure of the swash plate type piston pump to the hole, and a sleeve disposed between the valve housing and the control spool;
A signal pressure is supplied in accordance with a fluid pressure from the other pump, and a force due to the signal pressure is applied to the sleeve substantially parallel to a force in a spool moving direction applied by the discharge pressure, and the sleeve is added to the valve. A port to be moved relative to the housing;
With
A horsepower control regulator characterized in that a fluid pressure guided to the tilting actuator is corrected in accordance with a relative position between the sleeve and the control spool.   The horsepower control regulator according to claim 1, wherein the fluid pressure is supplied to the port so that the signal pressure becomes a fluid pressure from the other pump.   The horsepower control regulator according to claim 1, wherein the force applied to the sleeve by the signal pressure is opposite to the force applied to the control spool by the discharge pressure.   The horsepower control regulator according to claim 3, wherein the steady position of the control spool moves in the same direction as the sleeve as the sleeve moves. Furthermore, the first and second springs that can apply a force to the control spool opposite to the force applied to the control spool by the discharge pressure while the pressing force from the swash plate acts,
The free length of the first spring is longer than the free length of the second spring,
5. The difference between the free length of the second spring and the length of the first spring is increased by the movement of the steady position of the control spool, thereby extending the first spring. Horsepower control regulator. Furthermore, a second hole through which fluid pressure from the other pump is guided in the control spool, and a force is applied to the control spool by the fluid pressure, and the valve housing is applied to the valve housing according to the force. A second hole for moving the control spool;
The horsepower control regulator according to any one of claims 1 to 5, wherein the sleeve has a through hole that allows fluid pressure from the other pump to pass through the second hole.   A swash plate type piston pump having the horsepower control regulator according to any one of claims 1 to 6. A swash plate type piston pump driven together with other pumps by engine output, a cylinder block having a cylinder for housing the piston, a swash plate for reciprocating the piston as the cylinder block rotates, and a fluid pressure A horsepower control regulator used in a swash plate type piston pump having a tilt actuator for tilting the swash plate by
A valve housing;
A control spool that is slidably disposed inside the valve housing and adjusts a fluid pressure guided to the tilting actuator according to a moving position with respect to the valve housing;
A hole through which the discharge pressure of the swash plate type piston pump is guided in the control spool, and applies a force to the control spool in accordance with the discharge pressure to move the control spool with respect to the valve housing;
A sleeve having a through hole for passing the discharge pressure of the swash plate type piston pump to the hole, and a sleeve disposed between the valve housing and the control spool;
Means for supplying a signal pressure in accordance with a fluid pressure from the other pump;
A signal pressure is supplied in accordance with a fluid pressure from the other pump, and a force due to the signal pressure is applied to the sleeve substantially parallel to a force in a spool moving direction applied by the discharge pressure, and the sleeve is added to the valve. A port to be moved relative to the housing;
With
A horsepower control device characterized in that a fluid pressure guided to the tilting actuator is corrected in accordance with a relative position between the sleeve and the control spool.