JP2003139073A - Variable displacement plunger pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variable displacement plunger pump which saves on energy, and reduces a cost by reducing a useless flow rate in a high rotation area. SOLUTION: This variable displacement plunger pump has a piston of a small diameter control spool 455 and a large diameter control spool 456 for energizing a cam ring 403 in the eccentric direction, a reducing operation side communicating passage 461 for introducing pump delivery pressure to the small diameter control spool 455, and an increasing operation side communicating passage 462 for introducing pump internal pressure to the large diameter control spool 456, and sets a pump delivery quantity characteristic to a rotor rotating speed by having a nonvariable area where a pump delivery quantity is in proportion to the rotor rotating speed, a first variable area for restraining the pump delivery quantity from increasing in proportion to the rotor rotating speed in a higher rotation area than this nonvariable area, and a second variable area for reducing more than the pump delivery quantity in the first variable area to an increase in the rotor rotating speed in a further higher rotation area than the first variable area.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a variable displacement plunger pump, and more particularly to a pump suitable for being applied to a driving force distribution device for distributing the driving force of an engine to left and right driving wheels.

[0002]

2. Description of the Related Art Conventionally, there has been disclosed a variable displacement plunger pump in which the discharge flow rate is made substantially constant when the pump exceeds a predetermined rotational speed by eccentricizing a cam ring.
It is known from Japanese Patent Publication No. 200240. This conventional variable displacement plunger pump has a piston as an eccentric mechanism that urges the cam ring in the eccentric direction, and guides the pressure before and after the orifice arranged in the discharge flow path to both end faces of the piston, and the liquid on both end faces of the piston is guided. It was designed to operate eccentrically based on the pressure difference.

[0003]

However, in the above-mentioned prior art, the flow rate is substantially the same as that in the variable start region even when the pump rotation speed is high and the pressure is high. If was not required, the pump energy was lost. In particular, since the amount of heat generated is large in the high pressure region, when a large amount of fluid is discharged at high pressure, it becomes necessary to cool the fluid, and the provision of a cooling means causes an increase in cost. Further, in the plunger pump, it is possible to provide an orifice for restricting the flow rate at the time of high pressure and thereby suppress the flow rate, but in that case, the number of parts increases.

The present invention has been made by paying attention to the above-mentioned conventional problems, and it is possible to reduce a wasteful flow rate in a high rotation range to save energy and reduce costs. It is an object of the present invention to provide a pump, and further to facilitate setting of flow rate characteristics in this variable displacement plunger pump.

[0005]

To achieve the above object, a variable displacement plunger pump according to the present invention has a rotor that is rotationally driven, a cam ring that is eccentrically operated with respect to the rotor, and a radial shape formed on the rotor. , A cylinder chamber connected to the suction passage and the discharge passage, and mounted in the cylinder chamber so as to move forward and backward, and moves along the inner peripheral surface of the cam ring as the rotor rotates to make a reciprocating stroke to move fluid to the cylinder chamber. A plunger for inhaling and exhaling
A piston that biases the cam ring in an eccentric direction, an orifice inserted in the discharge passage, and a pressure that is upstream of the orifice in a direction that reduces the eccentric amount of the cam ring with respect to one end of the piston. A variable displacement plunger pump comprising: an operating communication passage; and an increasing operating communication passage that applies pressure downstream of the orifice to the other end of the piston in a direction to increase the eccentricity of the cam ring. The pump discharge amount characteristic with respect to the rotor rotation speed formed by eccentrically operating the cam ring by the piston has a non-variable region in which the pump discharge amount is proportional to the rotor rotation speed, and a pump in a rotation region higher than this non-variable region. In the first variable region in which the discharge amount is suppressed from increasing in proportion to the rotor rotation speed, and in the higher rotation region than the first variable region, the rotor rotation speed is increased. It was technique, characterized in that the setting and a second variable region to reduce than the pump discharge amount in the first variable region with increasing number.

The invention described in claim 2 is the same as claim 1
In the variable displacement plunger pump described in (1), the pressure receiving area on the other end side of the piston that receives the pressure from the increasing operation side communication passage is smaller than the pressure receiving area on one end side that receives the pressure from the decreasing operation side communication passage. At the same time as increasing the rotational speed of the second variable region, the force acting in the direction of increasing the eccentricity of the cam ring based on the area difference is applied to the variable load acting in the direction of decreasing the eccentricity of the cam ring generated by each plunger. The technique is characterized in that the second variable region is obtained by setting it smaller than the above.

According to a third aspect of the present invention, in the variable displacement plunger pump according to the second aspect, the second variable region is set such that the pressure receiving area on one end side of the piston is A and the pressure introduced to this one end side. Is a force F1 = A (P + ΔP) acting on the piston when (P + ΔP), a pressure receiving area on the other end side of the piston is B, and a pressure introduced to the other end is P, which acts on the piston. The relationship between the force F2 = BP and the fluctuating load F acting on the cam ring is F> F2-
The technique is characterized in that the pressure receiving area difference B-A is set to be F1.

[0008]

In the variable displacement plunger pump of the present invention, the pump discharge amount increases in proportion to the rotor rotational speed in the non-variable region where the rotor rotational speed is low, but the rotor rotational speed is In the first variable region, the piston operates in a direction to reduce the eccentric amount of the cam ring,
It is possible to prevent the pump discharge amount from increasing in proportion to the rotor rotation speed and improve the operation efficiency. This point is similar to the conventional technique. On the other hand, in the present invention, the rotor rotation speed is higher than the first variable range in the second rotation range.
When the variable region is reached, the piston operates in a direction to further reduce the eccentric amount of the cam ring, and reduces the pump discharge amount in the first variable region as the rotor speed increases. By setting the second variable region, it is possible to reduce a wasteful flow rate at high rotation speed, and prevent heat generation and save energy.

According to the second aspect of the present invention, the setting of the second variable region is a balance between the area difference between the one end side and the other end side formed on the piston and the variable weight acting on the cam ring by the operation of the plunger. According to the invention of claim 3, wherein the second variable region can be easily set, the setting of the second variable region acts on the piston in a direction to reduce the eccentric amount of the cam ring. Force F
1, the pressure receiving area difference B-A is set such that the relationship between the force F2 acting on the piston in the direction of increasing the eccentricity of the cam ring and the variable load F acting on the cam ring is F> F2-F1. Since this is done, it becomes easy to set the pressure receiving area, that is, the second variable region.

[0010]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the drawings. The variable displacement plunger pump of the present embodiment is applied to the driving force distribution device PD that distributes the output of the engine to the left and right driving wheels. FIG. 1 is an overall view showing a driving force distribution device PD, in which an engine output (not shown) is transmitted from a transmission (not shown) and a propeller shaft via the driving force distribution device PD, and further to the left and right from the wheel shafts 16 and 17. It is configured to be transmitted to the wheel.

In the figure, 1 is a housing. The housing 1 includes a stepped cylindrical input housing portion 2 located on the input side and an input housing portion 2 located on the output side.
And an output housing portion 3 having a stepped tubular shape that is integrally provided with the output housing portion 3 and that extends leftward and rightward. Further, the output housing portion 3 includes a tubular body portion 3A located at the center, and stepped tubular cover portions 3B and 3C provided on both left and right sides of the body portion 3A.

Further, the output housing portion 3 has a body portion 3A.
Is located inside the cover, and the stepped cylindrical portion 3D is provided between the cover portion 3C and the cover portion 3C.
Is provided, and between the cover portion 3C and the stepped cylinder portion 3D,
A hydraulic motor 18 described later is rotatably arranged. The stepped cylindrical portion 3D extends in the axial direction from the cover portion 3C side toward the cover portion 3B side, and the tip end side thereof is formed so as to avoid interference with the input gear 4A described later.

In the drawing, reference numeral 4 denotes a propulsion shaft, which is rotatably provided inside the input housing portion 2 of the housing 1 and is connected to the propeller shaft. Further, the body 3 of the output housing 3 is attached to the propulsion shaft 4.
An input gear 4A is provided on the tip side extending inside A, and the input gear 4A is meshed with a ring gear 7 described later.

In the figure, reference numeral 5 is a differential mechanism, which is located on the cover portion 3B side and is rotatably provided inside the output housing portion 3. The differential mechanism 5 is composed of a planetary gear mechanism G, and the ring gear 7 meshed with the input gear 4A is fixed by bolts or the like on the outer peripheral side of a differential case 6 forming the outer shell thereof. Then, the differential case 6 is rotationally driven by the propulsion shaft 4 via the input gear 4A and the ring gear 7.
At this time, the driving force is distributed and transmitted to the left and right wheel shafts 16 and 17 by the sun gear 13 and the carrier 8 described later.

Here, the differential mechanism 5 and the ring gear 6A of the planetary gear mechanism G formed on the inner circumference of the differential case 6 over the entire circumference thereof and the planetary gear arranged in the differential case 6 so as to be rotatable relative to each other. A carrier 8 of the gear mechanism G and a plurality of first planetary gears 1 rotatably supported by the carrier 8 and meshed with a ring gear 6A.
0, and a second planetary gear 12 meshed with the first planetary gear 10 and rotatably supported by the carrier 8.
And a sun gear 13 meshed with these second planetary gears 12 and rotatable relative to the carrier 8. The sun gear 13 is spline-coupled so as to rotate integrally with the wheel shaft 16. Therefore, the rotation of the differential case 6 depends on the rotation of the first and second planetary gears 1.
Rotation in the same direction is transmitted to the sun gear 13 via 0 and 12.

When the vehicle is running straight, the rotational force of the differential case 6 is evenly distributed to the sun gear 13 and the carrier 8 of the planetary gear mechanism G, and the left and right wheel shafts 1
6 and 17 rotate at the same rotation speed. Further, when the vehicle is turning, one of the wheel shafts 16 and 17 rotates faster than the other due to a reaction force received by a wheel (not shown) from the road surface or the like on the differential case 6 side with respect to the sun gear 13 and the carrier 8. Rotational forces are transmitted with different rotation ratios. As a result, the wheel on the inside of the turn has a relatively low rotation speed, and the wheel on the outside of the turn has a relatively high rotation speed, thereby exhibiting a differential function of enhancing the cornering performance of the vehicle.

Next, the relationship between the differential mechanism 5 and the wheel shafts 16 and 17 will be described. The sun gear 13 is connected to the wheel shaft 16 via a spline connecting portion, and
Rotates integrally with the sun gear 13. Further, the wheel shaft 16 has a spline shaft portion 16A inserted into the output housing portion 3, and the spline shaft portion 16A is connected to the cylinder block 25 in a non-rotating state. On the other hand, the carrier 8 is connected via a sleeve 14 to a motor case 19 formed so as to enclose the wheel shaft 16 and the cylinder block 25.

The hydraulic motor 18 is juxtaposed side by side with the differential mechanism 5 in the left-right direction inside the output housing portion 3. The hydraulic motor 18 is, for example, a radial piston hydraulic motor, and includes a motor case 19, a cylinder block 25, a piston 27, and a passage block 32.
And a cylinder 2 formed in a cylinder block 25
By supplying and discharging pressure oil from the hydraulic pump 40 to the piston 6 and applying a driving force to the piston 27, a relative rotational force is applied between the motor case 19 and the cylinder block 25.
A relative rotational force is applied to the left and right wheel shafts 16 and 17.

A directional control valve 42 is provided between the hydraulic motor 18, the hydraulic pump 40 and the tank T, and oil is discharged from the hydraulic pump 40 through connection circuits 41A and 41B connected to the hydraulic motor 18. By selectively connecting or disconnecting the discharge passage 417 and the drain passage 415 connected to the tank T, it is possible to execute the supply / discharge of the pressure oil to / from the hydraulic motor 18 and the stop of the supply / discharge. .

The hydraulic pump 40 has a suction passage 416.
The oil in the tank T is sucked from the tank and discharged into the discharge passage 417. Further, the discharge pressure of the hydraulic pump 40 is variably controlled according to the number of rotations, whereby the hydraulic motor 1
The driving force of 8 can be changed. Further, when the discharge pressure exceeds the set pressure, the relief valve 43 allows it to escape to the tank T.

Next, the structure of the hydraulic pump 40 will be described. FIG. 2 is a sectional view of the hydraulic pump 40, in which 401 is a rotor, 403 is a cam ring, and 404 is a casing.

A plurality of plunger holes 402 are radially formed on the outer periphery of the rotor 401, a drive shaft (not shown) is provided at one end in the axial direction, and a boss portion (not shown) is provided at the other end. The drive shaft and the boss portion are the casing 40.
4 is rotatably supported. In addition, a plunger 408 having a bottomed cylindrical shape is housed in the plunger hole 402 so as to be retractable in a state of being urged in the outer diameter direction by a spring 409. A supply / discharge passage 410 communicates with the bottom of the plunger hole 402.

Each of the supply / discharge passages 410 can communicate with the suction passage 416 and the discharge passage 417. That is, the casing 404 has a suction groove formed in the upper half of the arc overlapping the rotation trajectory of the supply / discharge passage 410 and connected to the suction passage 416, and a lower half of the same arc shape. A discharge groove to which the discharge passage 417 is connected is formed, and the supply / discharge passage 4 is formed according to the rotation of the rotor 401.
10 is connected to the suction passage 416 or the discharge passage 417.

The cam ring 403 is the outer ring 4
18 and the inner ring 419 are rotatably engaged with each other via a bearing 420, and the inner peripheral surface of the inner ring 419 is brought into contact with the top surface of each plunger 418.
That is, the inner peripheral surface of the inner ring 419 is a cam surface that guides the retracting movement of each plunger 408, and each plunger 408 and the plunger hole 402 form a pump chamber between the rotor 401 and the cam ring 403. . In addition, the outer ring 418 has one end of the casing 40.
4 is rotatably supported by a pin 421, and an input protrusion 422 protruding to the other end is engaged with a discharge amount control device 423 described later. The cam ring 403 changes the eccentric amount between the center of the inner ring 419 and the center of the rotor 401 by rotating about the pin 421, and thereby changes the stroke of each plunger 408, that is, the discharge amount of each pump chamber. Is what you can do. That is, FIG. 2 shows a state in which the cam ring 403 is most eccentric, whereas FIG. 3 shows a state in which the cam ring 403 swings from the position shown in FIG. There is. In the state where the amount of eccentricity shown in FIG. 3 is reduced, the stroke amount of each plunger 408 is reduced as compared with the state shown in FIG. 2, and the discharge amount is reduced accordingly.

The discharge amount control device 423 performs the operation of changing the eccentric amount of the cam ring 403 as described above. The structure of the discharge amount control device 423 will be described below. The discharge amount control device 423 includes a small diameter cylinder 451 and a large diameter cylinder 452 formed on both sides of the input protrusion 422 in the casing 404,
Plug 45 for closing the open ends of the cylinders 451 and 452
3, 454, a small diameter control spool 455 and a large diameter control piston 455, which are slidably accommodated in the respective cylinders 451 and 452, and whose front end surfaces are in contact with the input projections 422. Control spool 4
56 and springs 457, 458 for biasing the spools 455, 456 in the direction of pressing the input protrusion 422.
And the main orifice 42 inserted in the discharge passage 417
4, a reduction operation side communication passage 461 that connects the upstream side of the main orifice 424 and the small diameter cylinder 451, and the downstream side of the main orifice 424 and the large diameter cylinder 452.
And the communication passages 462 for increasing communication between
61 and 462, and damper orifices 463 and 464 for absorbing pulsation provided in the middle. The springs 457 and 458 are set such that the biasing force of the spring 458 that biases the large-diameter control spool 456 is set larger, and the cam ring 403 is biased to the maximum eccentricity state shown in FIG. 2 in the initial state. . On the other hand, the small-diameter control spool 455 is operated toward the direction (rightward in the drawing) in which the eccentric amount of the cam ring 403 is reduced by the hydraulic pressure (internal pressure of the pump) obtained from the decreasing-operation side communication passage 461, and the large-diameter control spool 455 is operated. 456 is configured to operate in a direction (leftward in the drawing) in which the amount of eccentricity of the cam ring 403 is increased by the hydraulic pressure obtained from the increasing operation side communication passage 462.

Next, the operation of the driving force distribution device PD will be described. When the engine is driven, this rotation output is
The propeller shaft propagates to the differential case 6 of the differential mechanism 5 via the propulsion shaft 4, and the rotation of the differential case 6 is transmitted to the sun gear 13 via the planetary gears 10 and 12 as rotation in the same direction. Further, each planetary gear 1 is attached to the carrier 8 of the differential mechanism 5.
Revolutions of 0 and 12 are transmitted, whereby the carrier 8 also rotates in the same direction as the differential case 6.

When the vehicle is running straight, the rotational force on the side of the differential case 6 is applied to each planetary gear 1.
It is evenly distributed to the sun gear 13 and the carrier 8 through 0 and 12, and the sun gear 13 and the carrier 8 rotate the left and right wheel shafts 16 and 17 at the same rotational speed, so that the vehicle passes through the left and right wheels. It is driven to run.

Further, during steering operation (turning) of the vehicle, one of the wheel shafts 16 and 17 rotates faster than the other due to a reaction force received by the left and right wheels from the road surface.
The rotational forces on the side of the differential case 6 are transmitted to the sun gear 13 and the carrier 8 at different rotational ratios, so that the wheels on the inside of the turn have a relatively low rotational speed with respect to the outside of the turn, thereby improving the cornering performance of the vehicle. Can be increased.

By the way, since the distribution of the driving force to the left and right wheel shafts 16 and 17 by the differential mechanism 5 depends on the reaction force received from the road surface by the left and right wheels, for example, when a slip occurs or the like. Loses its drive distribution function,
It may not always be possible to improve the running stability of the vehicle.

Therefore, in such a case, the control unit (not shown) outputs a control signal to the hydraulic motor 18 by outputting a control signal to the directional control valve 42 based on the detection signals of various sensors for detecting the operating state of the vehicle. Supply and discharge. As a result, the hydraulic motor 18 causes the supply / discharge passage 33A of the passage block 32,
When pressure oil is supplied to and discharged from each cylinder 26 in the cylinder block 25 via 33B and each oil passage 28, the piston 27 reciprocates in each cylinder 26, and the hydraulic motor 1
The motor case 19 of 8 and the cylinder block 25 rotate relative to each other.

As a result, the left and right wheel shafts 16 and 17 are
Rotational forces due to relative rotation between the motor case 19 and the cylinder block 25 are applied in opposite directions, and the driving force applied from the propulsion shaft 4 to the differential mechanism 5 is applied to the left and right wheel shafts 16 according to the rotational force from the hydraulic motor 18. , 17 are actively distributed. Then, by the distribution of the driving force at this time, for example, yaw control for generating a yaw moment in an arbitrary direction in the vehicle is executed, or LSD control for limiting the left-right differential of the rear wheel RW is executed. be able to. Note that the yaw control and the LSD control themselves are not a feature of the present invention, and thus detailed description thereof will be omitted.

Next, the operation of the hydraulic pump 40 of this embodiment will be described. In the present embodiment, when the rotation speed of the rotor 401 of the hydraulic pump 40 increases and the amount of oil passing through the main orifice 424 increases, the discharge passage 4
The discharge pressure, which is the pressure of 17, increases. This discharge pressure does not become higher than the set value by the relief valve 43, and since the main orifice 424 is provided, when the discharge amount in the hydraulic pump 40 increases,
The pump internal pressure upstream of the main orifice 424 rises.

Here, in the discharge amount control device 423, when the pressure receiving area of the small diameter control spool 455 is A and the pressure supplied from the reducing operation side communication passage 461 is (P + ΔP), the small diameter control spool 455 is operated. Acting force FS
Is FS = A (P + ΔP). On the other hand, the pressure receiving area of the large-diameter control spool 456 is A + ΔA, and the increasing operation side communication passage 4
When the pressure supplied from 62 is P, the force FB acting on the large diameter control spool 456 is FB = (A + ΔA) ·
P.

Therefore, when the rotor rotational speed is low and the discharge amount is small to some extent, the hydraulic pressure difference ΔP before and after the main orifice 424 is small. Therefore, FB> FS
Therefore, the cam ring 403 is maintained in a state where the amount of eccentricity is large. Therefore, as shown in the rotational speed-discharge amount characteristic diagram of FIG. 4, a region where the rotor rotational speed N is low (this is the non-variable region H1
In this case, the discharge amount Q is proportional to the increase of the rotor speed.
Rises.

When the rotation speed becomes higher than the non-variable region H1, the hydraulic pressure difference ΔP of the main orifice 424 becomes large. Therefore, FS> FB, and the small diameter control spool 455 pushes the input protrusion 422 in a direction to reduce the eccentric amount of the cam ring 403. Therefore, the discharge amount Q of the hydraulic pump 40 is suppressed.

In the conventional variable displacement pump, in the region where the rotor speed N is higher than the non-variable region H1, the small diameter control spool 455 adjusts the eccentric amount of the cam ring 403 regardless of the discharge pressure P being low or high. The discharge amount is reduced and the discharge amount is set to be substantially constant.

On the other hand, in the present embodiment, as shown in FIG. 3, in the region where the rotor speed N is higher than the non-variable region, the discharge amount Q is suppressed as the rotor speed N increases. The first variable region H2 and the second variable region H3 in which the discharge amount Q is reduced as the rotor pressure N increases as the discharge pressure P becomes higher are set.

The setting of the second variable region H3 is performed by setting the force FS generated in each of the control spools 455 and 456 described above.
FB and the fluctuating load F that acts on the rotor 401 from the plunger 408.

The variable load F will be described below.
As shown in FIGS. 5 and 6, each plunger 408 generates a reaction force that pushes the inner circumference of the cam ring 403 outward in the compression stroke, and the component force F of this reaction force centered on the pin 421 is generated.
1, F2, F3, F4 are generated, and these component forces F1, F
The resultant force of 2, F3 and F4 is referred to as a variation weight F, and this variation weight F acts in a direction in which the cam ring 403 is swung, that is, eccentric. Figure 5 here
Indicates that the fluctuating load F is the minimum, and in this case, (-) F1 * L1 + F2 * L2> (+) F3 * L3 + F
4 × L4, and the fluctuating load F acts in the direction (-direction) to reduce the eccentricity amount. By the way, the left side of the above inequality is M1,
When the right side is M2, the variation weighting F is F = (M1-M
2) / L. On the other hand, FIG. 6 shows a state where the fluctuating load F is maximum, and (−) F1 × L1 + F2 × L2 <(+) F3 × L3 + F
4 × L4, and the fluctuating load F acts in the direction (+ direction) of increasing the eccentricity. In this way, the variable load F is
It changes depending on the rotation angle of 01 and increases as the discharge pressure increases.

FIG. 7 shows changes in the fluctuating load F corresponding to the rotor rotation angle. As shown in the figure, the fluctuating load F is
It changes depending on the rotation angle of the rotor 401. With respect to this fluctuating load F, in the present embodiment, the above-mentioned second variable region H
In the rotor rotation range of 3, the force that acts in the direction in which the cam ring 403 is eccentric based on the area difference ΔA in the discharge amount control device 423 is a value smaller than the maximum value Fmax of the fluctuating load F, specifically, the maximum value. Is set to a value within the range of 80 to 90%. That is, in the second variable region, the force FS = A (P + Δ) acting on the small diameter control spool 455.
P) and the force FB acting on the large diameter control spool 456 =
(A + ΔA) · P and the variation weight F are Fmax> FB
It is set to −FS, and this can be set by the area difference ΔA between the small diameter control spool 455 and the large diameter control spool 456.

Therefore, in the second variable region H3 in which the rotor speed N is increased, the variable load F is increased, the eccentric amount of the cam ring 403 is decreased, and the discharge amount Q is decreased as shown in FIG. The discharge amount Q decreases more as the discharge pressure P increases.

In the above-described embodiment, for example, in a situation where the wheel speed of the inner wheel is higher than that of the outer wheel and there is a risk of excessive understeer characteristics, etc.
The control shifts to LSD control for reducing the speed difference between the left and right wheels to execute control for softening the vehicle behavior. At this time, in the hydraulic pump 40, the rotor 4 is operated at high speed.
When the rotational speed of 01 becomes high and the discharge pressure becomes high, the discharge amount Q is narrowed down based on the characteristic shown in FIG.
Therefore, it is possible to significantly reduce the flow rate of high-pressure oil in the high rotation range as compared with the related art, to suppress the amount of heat generation and to suppress the energy consumption.

Although the embodiments have been described with reference to the drawings, the present invention is not limited to the embodiments. For example, in the embodiment, the piston divided into the small-diameter control spool 455 and the large-diameter control spool 456 is shown. However, as in the technique described in the prior art publication, one piston is provided. You may make it form two pressure receiving surfaces. In the embodiment, the desired discharge amount characteristic is set by the balance between the pressure receiving area difference ΔA between the small diameter control spool 455 and the large diameter control spool 456 and the variable load. Including the case where only one piston is used, the hydraulic pressure acting on the piston is set to the same pressure receiving area.
It is also possible to adjust and set the operating forces FS and FB by adjusting the hydraulic pressure acting on the piston by providing an orifice (including a variable orifice) in the middle of the hydraulic pressure transmission path. Further, in the embodiment, the example in which the variable displacement plunger pump of the present invention is applied to the driving force distribution device PD has been shown, but it can be applied to other devices such as a power steering device.

[Brief description of drawings]

FIG. 1 is a sectional view showing a driving force distribution device to which a variable displacement plunger pump according to an embodiment is applied.

FIG. 2 is a cross-sectional view showing a hydraulic pump as a variable displacement plunger pump according to an embodiment.

FIG. 3 is a sectional view showing a hydraulic pump as a variable displacement plunger pump according to the embodiment.

FIG. 4 is a rotation speed-discharge amount characteristic diagram of the embodiment.

FIG. 5 is an explanatory diagram showing a generation state of a fluctuating load corresponding to a rotor rotation angle.

FIG. 6 is an explanatory diagram showing a generation state of a fluctuating load corresponding to a rotor rotation angle.

FIG. 7 is a variable load characteristic diagram corresponding to a rotor rotation angle.

[Explanation of symbols]

40 hydraulic pump 401 rotor 402 Plunger hole 403 cam ring 404 casing 408 Plunger 409 spring 410 supply / discharge passage 415 drain passage 416 Inhalation passage 417 Discharge passage 418 outer ring 419 Inner Ring 420 bearing 421 pin 422 Input protrusion 423 Discharge rate control device 424 Main orifice 451 small diameter cylinder 452 large diameter cylinder 453,454 plug 455 Small diameter control spool 456 large diameter control spool 457,458 springs 461 Reduced working side communication passage 462 Increase working side communication passage 463,464 Damper orifice 456 large diameter control spool G planetary gear mechanism PD driving force distribution device T tank

Continued front page    F term (reference) 3H040 AA03 BB01 CC09 CC18 CC19                       CC22 DD03 DD21 DD33 DD37                 3H044 AA02 BB05 CC19 CC26 DD10                       DD11 DD24 DD33

Claims (3)

[Claims] 1. A rotor that is rotationally driven, a cam ring that eccentrically operates with respect to the rotor, a cylinder chamber that is radially formed in the rotor and that is connected to an intake passage and a discharge passage, and is movable back and forth in the cylinder chamber. A plunger that is mounted and moves along the inner peripheral surface of the cam ring with the rotation of the rotor to reciprocate to draw and discharge fluid into and from the cylinder chamber; and a piston that biases the cam ring in an eccentric direction. An orifice inserted in the discharge passage, a reducing actuating communication passage for applying a pressure upstream of the orifice to the one end side of the piston in a direction to reduce the eccentric amount of the cam ring, and a downstream side of the orifice. An increasing working side communication passage for applying a pressure in the direction of increasing the eccentricity of the cam ring to the other end side of the piston. A variable displacement plunger pump, wherein the piston discharge amount characteristic with respect to the rotor rotation speed formed by the piston eccentrically operating a cam ring has a non-variable region in which the pump discharge amount is proportional to the rotor rotation speed, and Also in the high rotation range, the first variable range in which the pump discharge amount is suppressed from increasing in proportion to the rotor rotation speed,
A variable displacement plunger pump, characterized in that it has a second variable region for reducing the pump discharge amount in the first variable region with respect to an increase in the rotor rotational speed in a higher rotational region than in the variable region. 2. The variable displacement plunger pump according to claim 1, wherein the pressure receiving area on the other end side of the piston that receives the pressure from the increasing operation side communication passage, and the pressure from the decreasing operation side communication passage are received. At the rotational speed of the second variable region, the force acting in the direction of increasing the eccentricity of the cam ring based on the area difference is applied to the eccentricity of the cam ring generated by each plunger. The variable displacement plunger pump is characterized in that the second variable region is obtained by setting it to be smaller than the fluctuating load acting in the direction of decreasing. 3. The variable displacement plunger pump according to claim 2, wherein the second variable region is set such that the pressure receiving area on one end side of the piston is A and the pressure guided to this one end side is (P + ΔP). Force acting on the piston of F1 = A (P + ΔP)
And the pressure receiving area on the other end side of the piston is B and the pressure introduced to the other end side is P, the relationship between the force F2 = BP acting on the piston and the fluctuating load F acting on the cam ring is A variable displacement plunger pump characterized in that a pressure receiving area difference BA is set so that F> F2-F1.