JP2014066184A - Variable capacity pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variable capacity pump capable of improving energy efficiency.SOLUTION: The variable capacity pump comprises: a control valve 4 for variably controlling an emission flow rate Q of a pump (pump part 2) by moving a valve body (spool 41) by a difference ΔPv of pressures Pv1, Pv2 applied to a first pressure reception part (pressure reception face of the spool 41 in an x-axial negative direction) and a second pressure reception part (pressure reception face of the spool 41 in an x-axial positive direction); a first flow passage (upstream side oil passage 31) connecting an upstream side of a throttle part (orifice 320) in the emission flow passage connected to an emission part (emission port 28) of the pump with the first pressure reception part side of the valve body of the control valve 4; and a second flow passage (downstream side oil passage 33) connecting a downstream side of the throttle part (orifice 320) with the second pressure reception part side of the valve body of the control valve 4. Pressure losses ΔP1, ΔP2 of the first flow passage or the second flow passage are adjusted so that characteristics of the emission flow rate Q under predetermined conditions (pressure, temperature and the like) are approximate to characteristics required by an object apparatus (CVT) to which the working fluid is supplied by the pump.

Description

  The present invention relates to a variable displacement pump.

  Conventionally, variable displacement pumps are known. For example, the pump described in Patent Document 1 includes a control valve that variably controls the discharge flow rate of the pump by the movement of the valve body, and the pressure on the upstream side and the pressure on the downstream side of the throttle portion in the discharge flow path from the pump. The discharge flow rate of the pump is controlled by moving the valve body by this pressure difference.

JP 2001-304139 A

  However, in the pump described in Patent Document 1, if the control flow rate is set in accordance with the maximum flow rate required under a predetermined condition, the discharge flow rate is wasted and the energy efficiency may be deteriorated. An object of the present invention is to provide a variable displacement pump capable of improving energy efficiency.

  In order to achieve the above object, the pump of the present invention adjusts the pressure loss in the flow path connected to the control valve so that the characteristic of the discharge flow rate with respect to a predetermined condition approximates the required characteristic.

  Therefore, waste of the discharge flow rate can be reduced and energy efficiency can be improved.

It is sectional drawing of the pump of Example 1. FIG. It is a characteristic view showing the relationship between the rotation speed of the pump of Example 1, and discharge flow volume. It is a characteristic view which shows the relationship between the pressure of the pump of Example 1, temperature, and the difference of pressure loss. It is a figure which shows the characteristic of the discharge flow volume with respect to the pressure and temperature of the pump of Example 1. FIG. It is a graph which shows the experimental result of the relational characteristic of the discharge pressure and discharge flow rate of the pump of Example 1.

[Example 1]
[Constitution]
The variable displacement pump of the first embodiment (hereinafter referred to as pump 1) is a variable displacement pump that can vary the discharge capacity (the amount of fluid discharged per revolution; hereinafter referred to as pump capacity), and is a hydraulic device for automobiles. It is used as a hydraulic pressure supply source for supplying hydraulic oil. Specifically, the pump 1 is used as a hydraulic pressure supply source for a continuously variable transmission, particularly a belt-type continuously variable transmission (CVT). The pump 1 is driven by a crankshaft of an internal combustion engine (engine), and sucks and discharges working fluid. Hydraulic fluid, specifically ATF (automatic transmission fluid) is used as the working fluid. Various valves controlled by the CVT control unit are provided in the control valve of the CVT to which the pump 1 is applied. The hydraulic oil discharged from the pump 1 is supplied to each part (primary pulley, secondary pulley, forward clutch, reverse brake, torque converter, lubrication / cooling system, etc.) of the CVT via the control valve. In the CVT control unit, the line pressure is appropriately controlled according to the driving conditions such as the accelerator opening, the engine speed, and the vehicle speed.

  The pump 1 has a pump unit 2 that sucks and discharges hydraulic oil and a control unit 3 that controls the pump capacity as an integral unit. FIG. 1 is a diagram showing a cross section of a part of the pump 1, showing a cross section of the pump part 2 excluding the housing member 40 taken along a plane perpendicular to the rotation axis O, and the control part 3 as the axis of the control valve 4. The partial cross section cut by the plane which passes through is shown. For convenience of explanation, the x-axis is provided in the direction in which the axis of the control valve 4 extends, and the side where the valve body (spool 41) is separated from the solenoid 5 is defined as the x-axis positive direction.

  The pump unit 2 includes, as main components, a drive shaft 20 driven by a crankshaft, a rotor 21 driven to rotate by the drive shaft 20, a vane 22 accommodated in the rotor 21 so as to protrude and retract, and a rotor 21. An annular cam ring 23 disposed so as to surround, an annular adapter ring 24 disposed so as to surround the cam ring 23, and disposed on the axial side surface of the cam ring 23 and the rotor 21, together with the cam ring 23, the rotor 21, and the vane 22. And a plate member (or housing member 40) that forms a plurality of pump chambers r. A cam ring 23 is swingably accommodated on the inner periphery of the adapter ring 24. A spring 27 is installed in a compressed state on the x-axis positive direction side of the adapter ring 24, and the cam ring 23 is constantly urged toward the x-axis negative direction side with respect to the adapter ring 24. A pin 25 is provided between the adapter ring 24 and the cam ring 23 to lock them. The cam ring 23 is supported with respect to the adapter ring 24 by a rolling surface 240 on which a pin 25 is provided, and is swingably installed with the rolling surface 240 as a fulcrum. A seal member 26 is installed on the inner periphery of the adapter ring 24 on the substantially opposite side of the pin 25 across the rotation axis O. When the cam ring 23 swings, the inner rolling surface 240 of the adapter ring 24 contacts the outer peripheral surface of the cam ring 23, and the seal member 26 slides on the outer peripheral surface of the cam ring 23. The rotor 21 is installed on the inner peripheral side of the cam ring 23. In the rotor 21, a plurality of grooves (slits 210) are formed radially. The vane 22 is a substantially rectangular plate member (blade), and is accommodated in each slit 210 so as to appear and retract.

  A space between the rotor 21 and the cam ring 23 and the plate member (or housing member 40) sandwiching them is partitioned into a plurality (11) of pump chambers (volume chambers) r by a plurality of vanes 22. The rotor 21 rotates in the clockwise direction of FIG. In a state where the center axis Oc of the cam ring 23 is decentered with respect to the rotation axis O (to the negative x-axis direction), the outer surface of the rotor 21 and the cam ring 23 are moved from the positive x-axis direction toward the negative x-axis direction. The distance in the rotor radial direction from the inner peripheral surface (the radial dimension of the pump chamber r) increases. In response to this change in distance, the vanes 22 appear and disappear from the slits 210 to separate the pump chambers r, and the pump chamber r on the x-axis negative direction side has a pump chamber r on the x-axis positive direction side. Rather than the volume. Due to the difference in the volume of the pump chamber r, the volume of the pump chamber r decreases as the rotor 21 rotates (the pump chamber r moves toward the positive x-axis direction) on the upper side of FIG. 1, the volume of the pump chamber r increases as the rotor 21 rotates (the pump chamber r moves toward the negative x-axis direction). Thus, the pump chamber r periodically expands and contracts while rotating around the rotation axis O in the clockwise direction. In the suction region where the pump chamber r expands in the rotation direction, hydraulic oil is sucked into the pump chamber r from a suction port 29 formed in the plate member (or the housing member 40). In the discharge region where the pump chamber r is reduced in the rotation direction, the sucked hydraulic oil is discharged from the pump chamber r to the discharge port 28 formed in the plate member (or the housing member 40).

  The control unit 3 of the pump 1 includes control chambers R1, R2, a hydraulic circuit, and an electromagnetic valve (control valve 4). The space between the inner peripheral surface of the adapter ring 24 and the outer peripheral surface of the cam ring 23 is sealed on both sides in the axial direction by plate members (or housing members 40), while the rolling surface 240 (and the outer periphery of the cam ring 23). The two control chambers R <b> 1 and R <b> 2 are liquid-tightly separated by a seal member 26 and a contact portion with the surface). On the outer peripheral side of the cam ring 23, the first control chamber R1 is separated on the x-axis negative direction side in which the eccentric amount δ of the cam ring 23 increases, and the x-axis positive direction side in which the eccentric amount δ decreases. The second control room R2 is separated from each other. The hydraulic circuit has an oil passage 30 or the like as a hydraulic oil passage (pipe) connecting the respective parts in the housing member 40. On the oil passage 32 connecting the discharge port 28 and the discharge oil passage 34 of the pump portion 2, a metering orifice (hereinafter simply referred to as an orifice) 320 as a throttle portion is provided. The oil passage 30 from the discharge port 28 branches to an upstream oil passage 31 and an oil passage 32 on the upstream side (discharge port 28 side) of the orifice 320, and the upstream oil passage 31 is connected to an upstream port 310 described later. Is done. The oil passage 32 branches downstream of the orifice 320 into a downstream oil passage 33 and a discharge oil passage 34. The downstream oil passage 33 is connected to a downstream port 330 described later, and the discharge oil passage 34 is connected to the CVT. Is done.

  The electromagnetic valve is an electromagnetic hydraulic control valve for variably controlling the capacity of the pump 1 and integrally includes a control valve 4 and a solenoid 5 as a valve device. The housing member 40 is formed with a substantially cylindrical valve accommodating hole 400 extending in the x-axis direction. The valve accommodating hole 400 accommodates the valve body and the like of the control valve 4. An upstream port 310 to which the discharge pressure on the upstream side of the orifice 320 is supplied is formed at the end on the x-axis negative direction side of the valve housing hole 400. In the valve accommodating hole 400, the first control oil passage 38 opens on the x-axis positive direction side of the upstream port 310. The first control oil passage 38 communicates with the first control chamber R1 of the pump unit 2 through a through hole that penetrates the adapter ring 24 in the radial direction. On the other hand, a downstream port 330 to which discharge pressure downstream of the orifice 320 (that is, pressure discharged to the CVT 100) is supplied is formed at the end of the valve housing hole 400 on the x-axis positive direction side. A second control oil passage 39 opens on the x-axis negative direction side of the downstream port 330 in the valve accommodation hole 400. The second control oil passage 39 communicates with the second control chamber R2 of the pump unit 2 through a through hole that penetrates the adapter ring 24 in the radial direction.

  The control valve 4 operates (displaces) the valve body to control the inflow / outflow of the working oil into the first control chamber R1 and the second control chamber R2. The control valve 4 is a spool valve, and a spool 41 as a valve body accommodated in a valve accommodation hole 400 as a cylinder so as to be displaceable (stroke) in the x-axis direction, and an x of the spool 41 in the valve accommodation hole 400. It is an elastic member installed in a compressed state on the axial positive direction side, and has a spring 43 as a return spring that constantly biases the spool 41 toward the negative x-axis direction side (solenoid 5 side). The x-axis positive direction end of the spring 43 is held at the bottom (retainer 42) of the valve housing hole 400 on the x-axis positive direction side. The spool 41 includes a first land portion 410 and a second land portion 411 for port shut-off (or for changing the port opening degree). The first land portion 410 is provided on the negative side of the spool 41 in the x-axis negative direction, and the second land portion 411 is provided on the end portion of the spool 41 in the positive direction of the x-axis. In order to facilitate the operation of the spool 41, a circumferential groove 44 is provided on the outer periphery of the first land portion 410, and a circumferential groove 45 is provided on the outer periphery of the second land portion 411.

  The diameter of the valve housing hole 400 (inner peripheral surface) is slightly larger than the diameters of the first land portion 410 and the second land portion 411. Inside the valve housing hole 400, a first pressure chamber 35 is defined on the x axis negative direction side of the first land portion 410, and a second pressure chamber 36 is defined on the x axis positive direction side of the second land portion 411. In addition, a drain chamber 37 is defined between the first land portion 410 and the second land portion 411. Regardless of the displacement of the spool 41, the upstream port 310 is always open in the first pressure chamber 35, and the downstream port 330 is always open in the second pressure chamber 36. The drain chamber 37 is always in communication with a drain oil passage (not shown) and is kept at a low pressure (open to atmospheric pressure). The hydraulic oil leaked from the first pressure chamber 35 to the drain chamber 37 through the gap between the valve housing hole 400 and the first land portion 410, and the valve housing hole 400 and the second land from the second pressure chamber 36. The hydraulic oil leaking into the drain chamber 37 through the gap with the part 411 is discharged from the drain oil passage.

  Displacement of the spool 41 in the x-axis direction changes the area (opening area of the oil passage) in which the opening portions (hydraulic oil supply / discharge holes) of the oil passages in the valve housing holes 400 are blocked by the land portions 410 and 411. As a result, the communication state or blocking state of each oil passage is switched. Each opening is arranged as follows. That is, in the state where the spool 41 is maximally displaced in the x-axis negative direction side, the opening (first control chamber R1) of the first control oil passage 38 is communicated with the first pressure chamber 35 by the first land portion 410. While being shut off, it communicates with the drain chamber 37. In the same state, the opening (second control chamber R2) of the second control oil passage 39 is communicated with the second pressure chamber 36 while the communication with the drain chamber 37 is blocked by the second land portion 411. As the spool 41 is displaced in the positive direction of the x-axis, the opening of the first control oil passage 38 (first control chamber R1) is blocked from communicating with the drain chamber 37, while the displacement amount exceeds a predetermined value. The area that is in communication with the first pressure chamber 35 and is blocked by the first land portion 410 decreases as the amount of displacement in the positive x-axis direction further increases (the opening area to the first pressure chamber 35 increases). To do. Further, the area of the opening (second control chamber R2) of the second control oil passage 39 that is blocked by the second land portion 411 increases as the spool 41 is displaced in the positive x-axis direction (second pressure chamber). When the amount of displacement exceeds a predetermined value, the communication with the second pressure chamber 36 is blocked. While the spool 41 is displaced to the maximum in the positive x-axis direction, the communication between the opening portion (first control chamber R1) of the first control oil passage 38 and the drain chamber 37 is blocked by the first land portion 410. The first pressure chamber 35 communicates with the first pressure chamber 35. In the same state, the opening (second control chamber R <b> 2) of the second control oil passage 39 is disconnected from the second pressure chamber 36 by the second land portion 411, and communicates with the drain chamber 37.

  By supplying hydraulic oil from the upstream oil passage 31, the pressure in the first pressure chamber 35 becomes a value corresponding to the oil pressure P1 upstream of the orifice 320. By supplying the hydraulic oil from the downstream oil passage 33, the pressure in the second pressure chamber 36 becomes a value corresponding to the oil pressure P 2 on the downstream side of the orifice 320. Due to the pressure drop when passing through the orifice 320, P2 becomes lower pressure than P1. The difference ΔPM (= P1−P2) between P1 and P2, that is, the difference between the hydraulic pressures upstream and downstream of the orifice 320 (hereinafter referred to as the differential pressure across the orifice) is the flow rate of hydraulic oil passing through the orifice 320, that is, the pump discharge flow rate Q. Increases in response to On the other hand, since the hydraulic oil leaks from the first pressure chamber 35 to the drain chamber 37, a flow of hydraulic oil is generated in the upstream oil passage 31, and a pressure loss ΔP1 occurs in the upstream oil passage 31. The pressure Pv1 in the one pressure chamber 35 is a value obtained by subtracting the pressure loss ΔP1 from the hydraulic pressure P1 when it is assumed that there is no pressure loss ΔP1 (Pv1 = P1−ΔP1). Similarly, when hydraulic oil leaks from the second pressure chamber 36 to the drain chamber 37, a flow of hydraulic oil occurs in the downstream oil passage 33, and a pressure loss ΔP2 occurs in the downstream oil passage 33. The pressure Pv2 in the second pressure chamber 36 is a value obtained by subtracting the pressure loss ΔP2 from the hydraulic pressure P2 when there is no pressure loss ΔP2 (Pv2 = P2-ΔP2). In other words, the difference ΔPv of the pressure Pv1 in the first pressure chamber 35 with respect to the actual pressure Pv2 in the second pressure chamber 36 is the pressure loss of the upstream oil passage 31 with respect to the downstream oil passage 33 from the differential pressure ΔPM before and after the orifice. (ΔP1−ΔP2) is subtracted (ΔPv = Pv1−Pv2 = ΔPM− (ΔP1−ΔP2)). In the present embodiment, the pressure loss ΔP1 in the upstream oil passage 31 is larger than the pressure loss ΔP2 in the downstream oil passage 33 by adjusting the length and diameter of the upstream oil passage 31 or the downstream oil passage 33. Set as follows. In other words, the pressure loss difference (ΔP1−ΔP2) is a positive value. The orifice front-rear differential pressure ΔPM is set larger than the pressure loss difference (ΔP1−ΔP2).

  As shown in FIG. 1, the solenoid 5 is installed in an installation portion 404 provided in an opening of the housing member 40 on the x-axis negative direction side of the valve housing hole 400. A seal member 404b is installed between the inner peripheral surface 404a of the installation portion 404 and the outer peripheral surface of the solenoid 5 (the x-axis positive direction protruding portion 5b), and the liquid tightness of the valve housing hole 400 is ensured. A plunger (rod) 5 a extends from the protruding portion 5 b of the solenoid 5 into the valve housing hole 400. The solenoid 5 is energized based on a command from the CVT control unit constituting the solenoid control means, thereby driving the plunger 5a in the positive direction of the x-axis with a thrust Fsol corresponding to the energization amount (command current). Specifically, by changing the pulse width of the drive voltage using the PWM method, a desired effective current is passed through the coil of the solenoid 5 to continuously change the driving force of the plunger 5a. The CVT control unit outputs a drive signal (drive voltage) to the coil at a predetermined cycle, and controls the current value supplied to the coil by making the energization time to the coil per cycle of the drive signal variable. The x-axis positive end of the plunger 5a is in contact with the x-axis negative end 412 of the spool 41, and the spool 41 is moved away from the load direction of the spring 43 by the electromagnetic force Fsol of the solenoid 5 (on the x-axis positive direction side). ) Is provided so that the displacement of the spool 41 can be controlled. As the current value energized to the solenoid 5 increases, the force Fsol that the solenoid 5 drives the spool 41 in the positive x-axis direction increases. The command current to the solenoid 5 is set to a small current value when the discharge flow rate required for the pump 1 is large, and is set to a large current value when the discharge flow rate required for the pump 1 is small according to the CVT request. Is done.

[Action]
Next, the operation of the pump 1 of Example 1 will be described. First, the capacity variable action of the pump 1 will be described with reference to FIGS. 1 and 2. FIG. 2 is a characteristic diagram showing the relationship between the rotational speed of the pump 1 of Example 1 and the discharge flow rate. The urging force Fspr by the spring 43 acts on the spool 41 of the control valve 4 in the x-axis negative direction, and the urging force Fv by the pressure difference ΔPv acting on both sides of the spool 41 in the x-axis direction acts in the x-axis positive direction. To do. Specifically, the urging force Fspr by the spring 43 is a stroke of the spool 41 from the state in which the spool 41 is displaced maximum in the negative x-axis direction (the amount of contraction of the spring 43 is minimum) to the positive x-axis direction. It depends on how much the spring 43 is contracted. The urging force Fv1 by the hydraulic pressure acting on the negative side of the spool 41 in the x-axis negative direction (pressing the spool 41 in the positive direction of the x-axis) is the pressure in the first pressure chamber 35 (the hydraulic pressure upstream of the orifice 320) Pv1. It is determined by the product of the pressure receiving area A1 on the negative side of the x-axis of the spool 41 corresponding to the diameter of the first land portion 410 (Fv1 = A1 × Pv1). The urging force Fv2 by the hydraulic pressure acting on the positive side of the spool 41 in the x-axis positive direction (pressing the spool 41 in the negative direction of the x-axis) is the pressure in the second pressure chamber 36 (the hydraulic pressure downstream of the orifice 320) Pv2. It is determined by the product of the spool 41 and the pressure receiving area A2 on the x axis positive direction side according to the diameter of the second land portion 411 (Fv2 = A2 × Pv2). Since A1 = A2 and Pv1 ≧ Pv2, Fv1 ≧ Fv2. The biasing force Fv due to the differential pressure ΔPv (= Pv1−Pv2) is the difference between Fv1 and Fv2 (Fv = Fv1−Fv2), and biases the spool 41 in the positive x-axis direction (the spool 41 in the positive x-axis direction). Act). For example, if the magnitude of the biasing force Fv in the positive x-axis direction due to the differential pressure ΔPv exceeds the biasing force Fspr in the negative x-axis direction by the spring 43 (Fv> Fspr), in other words, the biasing force Fv1 in the positive x-axis direction is x When the sum of the biasing forces Fspr and Fv2 in the negative axis direction is exceeded (Fv1> Fspr + Fv2), the spool 41 strokes in the positive x-axis direction. In accordance with this stroke, the spring 43 is compressed and Fspr is increased. The spool 41 stops when it strokes to a position where Fv (= Fv1−Fv2) and Fspr are balanced (Fv1 = Fspr + Fv2).

  The stop position of the spool 41 is on the x-axis negative direction side, the first land portion 410 blocks the upstream oil passage 31 and the first control oil passage 38, and the second land portion 411 contacts the downstream oil passage 33 and the first oil passage 33. When communicating with the 2 control oil passage 39, the hydraulic oil is not supplied to the first control chamber R1, but the hydraulic oil is supplied to the second control chamber R2. The stop position of the spool 41 is on the x-axis positive direction side, the first land portion 410 communicates with the upstream oil passage 31 and the first control oil passage 38, and the second land portion 411 communicates with the downstream oil passage 33 and the first oil passage 33. When the control oil passage 39 is cut off, hydraulic oil is not supplied to the second control chamber R2, but hydraulic oil is supplied to the first control chamber R1. The force that urges the cam ring 8 in the x-axis positive direction by the pressure in the first control chamber R1 and the force that urges the cam ring 8 in the x-axis negative direction by the pressure in the second control chamber R2 and the spring 27 When the sum of the forces urging in the negative axis direction is exceeded, the cam ring 8 swings in the positive x-axis direction, the eccentric amount δ decreases, and the pump capacity decreases. The force that urges the cam ring 8 in the x-axis positive direction by the pressure in the first control chamber R1 and the force that urges the cam ring 8 in the x-axis negative direction by the pressure in the second control chamber R2 and the spring 27 When the sum of the forces urging in the negative axis direction is below, the cam ring 8 swings in the negative x-axis direction, increasing the eccentricity δ and increasing the pump capacity. As the discharge flow rate Q of the pump 7 increases, the differential pressure ΔPv (biasing force Fv) increases and the spool 41 strokes in the positive direction of the x-axis. Therefore, the first control chamber is not the second control chamber R2. The hydraulic oil is supplied to R1. Therefore, the amount of eccentricity δ decreases, and the pump capacity decreases. As the discharge flow rate Q of the pump 7 decreases, the differential pressure ΔPv (biasing force Fv) decreases and the spool 41 strokes in the negative direction of the x-axis. Therefore, the second control chamber is not the first control chamber R1. The hydraulic oil is supplied to R2. Therefore, the amount of eccentricity δ increases and the pump capacity increases. In this way, the operation of the control valve 4 is controlled so that the pump capacity changes according to the discharge flow rate Q of the pump 7.

  As described above, the operation of the control valve 4 (displacement of the spool 41) is mechanically (difference in hydraulic pressure acting on both sides of the spool 41) in response to feedback of the discharge flow rate (control flow rate) Q of the pump unit 2 to be controlled. And controlled electronically by a thrust (electromagnetic force) acting on the spool 41 from the solenoid 5 in the direction opposite to the load direction of the spring 43 (the positive direction of the x-axis) based on a command from the CVT control unit. It is provided as possible. First, the operation of the control unit 3 when the solenoid 5 is in an inoperative state will be described. For example, in the initial stage of pump operation, when the pump rotational speed is relatively small and the pump discharge flow rate Q is relatively small, the differential pressure ΔPv (biasing force Fv) is not so large. Therefore, the cam ring 8 is in a predetermined eccentric state, and the flow rate Q increases with an inclination corresponding to the pump capacity as the rotational speed increases (see the fixed capacity region (a) in FIG. 2). The operation of the control valve 4 in the initial stage of the pump operation can be stabilized by installing the spring 43 and constantly energizing the spool 41 in the x-axis negative direction by the set load (urging force Fspr). If the rotational speed (flow rate Q) increases to some extent, the differential pressure ΔPv (biasing force Fv) increases, the eccentric amount δ decreases and the pump capacity decreases, so even if the rotational speed increases, the increase in the flow rate Q is suppressed. Will come to be. Specifically, a constant flow rate is maintained regardless of the rotational speed. When the flow rate Q falls below the predetermined flow rate, the differential pressure ΔPv (biasing force Fv) decreases, so the cam ring 8 is decentered again, and the flow rate Q is increased as appropriate (discharge flow rate control region (b) in FIG. )reference). Here, when the solenoid 5 is in an inoperative state, the force in the x-axis positive direction opposite to the biasing force Fspr in the x-axis negative direction by the spring 43 is only the urging force Fv by the differential pressure ΔPv. Therefore, unless the differential pressure ΔPv (flow rate Q) is relatively large, it is not possible to secure a force that causes the spool 41 to stroke in the x-axis positive direction (decreases the pump capacity). For this reason, as shown in FIG. 2 (a graph when the solenoid is off), after a relatively high flow rate Q is achieved, a constant flow rate is maintained.

  Next, the operation of the control unit 3 when the solenoid 5 is in the operating state will be described. When the solenoid 5 is energized, an electromagnetic force Fsol that urges the spool 41 in the positive x-axis direction is generated, so that the same action as that obtained by changing the urging force Fspr (set load) by the spring 43 to a small value can be obtained. For example, when the sum of the urging force Fv and the electromagnetic force Fsol due to the differential pressure ΔPv acting in the x-axis positive direction exceeds the urging force Fspr due to the spring 43 acting in the x-axis negative direction (Fv + Fsol> Fspr), the spool 41 Stroke in the x-axis positive direction. The spool 41 stops when it strokes to a position where Fv + Fsol and Fspr are balanced (Fv = Fspr−Fsol). Therefore, the spool 41 can be stroked in the positive direction of the x-axis from the stage where the flow rate Q is relatively small and the differential pressure ΔPv (biasing force Fv) is relatively small. Therefore, in the process of increasing the flow rate Q, the control valve 4 is switched at a timing earlier than when the solenoid 5 is in the non-operating state, and hydraulic oil is supplied to the first control chamber R1 instead of the second control chamber R2, and the eccentric amount δ Becomes smaller. For this reason, as shown in FIG. 2 (a graph when the solenoid is on), after a relatively low flow rate Q is achieved, a constant flow rate is maintained. In this way, the urging force generated by the solenoid 5 can control the above-described constant flow rate (hereinafter referred to as control flow rate) or the rotational speed at which the flow rate is constant, that is, the flow rate characteristic. Energy efficiency can be improved by controlling the flow rate characteristics in this way. In addition, although the structure where the urging | biasing direction of the spring 43 and the urging | biasing direction of the solenoid 5 are the same is also considered, there exists a possibility that a minimum control flow rate may be set high and energy efficiency may deteriorate. In this embodiment, since the biasing direction of the spring 43 and the biasing direction of the solenoid 5 are opposite to each other, the minimum control flow rate can be set low, and the energy efficiency can be improved.

  The hydraulic fluid discharged from the pump 1 is used in the CVT. In the CVT control unit, the line pressure is appropriately controlled according to the traveling situation such as the accelerator opening. Therefore, for example, when a high discharge flow rate is required according to the traveling situation, the current (electromagnetic force) energized to the solenoid 5 is reduced or zero, and when a low discharge flow rate is required, the solenoid 5 Increase the current (electromagnetic force) energized. Since it is such a structure, fail-safe property can be improved. That is, since the pump 1 is used for hydraulic pressure supply of an automatic transmission (CVT), it is desirable to ensure a high value (for example, maximum discharge flow rate) as the discharge flow rate when the solenoid 5 or the like fails. This is because the shortage of the discharge flow rate may cause slippage due to belt slip in the automatic transmission (CVT) or insufficient fastening force of the clutch or the like, which may affect the durability. Therefore, as described above, high reliability can be ensured by providing the discharge flow rate to be high when the solenoid is not operated. Specifically, a compression spring is used as the spring 43, and the solenoid 5 is installed on the side opposite to the side on which the set load of the spring 43 acts with the spool 41 interposed therebetween. By using the compression spring, it is possible to reduce the size and the configuration.

  Next, the operation of setting the pressure loss ΔP1 in the upstream oil passage 31 to be larger than the pressure loss ΔP2 in the downstream oil passage 33 will be described with reference to FIGS. FIG. 3 is a characteristic diagram showing the relationship between the pressure loss difference (ΔP1−ΔP2) and the pressure and temperature. FIG. 4 is a characteristic diagram showing the relationship between the control flow rate Q and the pressure and temperature. As shown in FIG. 3, the pressure loss difference (ΔP1−ΔP2) increases as the discharge pressure of the pump 1 (pressure in the oil passages 30 to 34) increases. In other words, compared with the pressure loss ΔP2 in the downstream oil passage 33, the pressure loss ΔP1 in the upstream oil passage 31 increases. Similarly, the pressure loss difference (ΔP1−ΔP2) increases as the temperature of the hydraulic oil increases. When the pressure loss difference (ΔP1−ΔP2) increases, the pressure difference ΔPv (= ΔPM− (ΔP1−ΔP2)) of the first pressure chamber 35 with respect to the second pressure chamber 36 decreases, and the spool 41 is moved in the x-axis by the differential pressure ΔPv. The biasing force Fv biased in the positive direction also decreases. For this reason, even with the same ΔPM (flow rate Q or number of revolutions), the spool 41 strokes more in the negative direction of the x-axis, and more hydraulic oil is supplied not to the first control chamber R1 but to the second control chamber R2. As a result, the amount of eccentricity δ increases. Therefore, as shown by the solid line in FIG. 4, the flow rate Q increases as the discharge pressure of the pump 1 or the temperature of the hydraulic oil increases. Here, the characteristic indicated by the solid line in FIG. 4 is also the characteristic (required characteristic) of the flow rate Q that is actually required by the CVT, which is the target device to which the pump 1 supplies hydraulic oil, according to the pressure and temperature. That is, in the CVT pump 1 of the present embodiment, the required flow rate characteristic is that the flow rate increases to the right with respect to the pump speed (engine speed), pressure, or temperature (the flow rate is increased according to these increases). The characteristic is such that the flow rate increases at high temperature and high pressure. Thus, in the pump 1, the pressure losses ΔP1 and ΔP2 in the upstream oil passage 31 or the downstream oil passage 33 are set so that the characteristics of the flow rate Q with respect to conditions such as pressure and temperature approximate the characteristics required by the CVT. It is adjusted. FIG. 5 is a graph showing the results of actual measurement of the relational characteristics between the control flow rate Q and the pressure in a prototype to which the configuration of this example is applied. As shown in FIG. 5, by adjusting the length and diameter of the upstream oil passage 31 or the downstream oil passage 33 to adjust the pressure loss ΔP1, ΔP2, specifically, the pressure loss difference (ΔP1-ΔP2), Unlike the conventional product to which the configuration of the example is not applied, the flow rate Q increases as the discharge pressure of the pump 1 increases.

  In other words, conventionally, variable displacement pumps that supply hydraulic oil to hydraulic equipment supply the maximum flow rate required by the hydraulic equipment under various conditions, such as pump speed, hydraulic oil temperature, and pressure. The pump characteristics were set so that However, since the hydraulic equipment discharges a large amount of flow even under conditions that do not actually require the flow, energy loss (power loss) for driving the pump increases and energy efficiency is improved. I can't. For example, in a case where the pump is driven by an internal combustion engine, the fuel efficiency improvement effect due to the variable displacement pump is reduced. For example, when the solid line in FIG. 4 indicates the required characteristics of a hydraulic device (CVT in this embodiment), the maximum flow rate Qmax (required under the worst conditions of pressure and temperature) is always realized regardless of the pressure and temperature. Thus, when the pump characteristics are set as shown by the broken line in FIG. 4, the difference between the set control flow rate Qmax and the flow rate Q actually required in accordance with the pressure or the like (flow rate in the region indicated by the oblique lines in FIG. 4) Is wasted and energy loss occurs. In addition, setting the control flow rate on the large flow rate side leads to increased noise due to cavitation, wear due to erosion, and the like. On the other hand, in the pump 1 of the present embodiment, the characteristics of the pump 1 are adjusted so that the flow rate Q changes according to conditions such as pressure and temperature. Specifically, the characteristics of the flow rate Q with respect to the above conditions are brought close to the required characteristics (solid line in FIG. 4) of the hydraulic equipment (CVT in this embodiment) to which the pump 1 supplies the working hydraulic pressure. This makes it possible to discharge the minimum flow rate required by hydraulic equipment (CVT) regardless of changes in the above conditions (increase or decrease in pressure or temperature), improving energy efficiency while achieving the required performance. can do. That is, by adjusting the characteristics of the pump 1 so as to be close to the required characteristics of the CVT, the loss region of the flow rate Q indicated by the diagonal lines in FIG. 4 can be reduced, thereby saving energy. When the pump 1 is driven by an internal combustion engine as in this embodiment, the fuel efficiency effect can be maximized.

  Note that the required characteristics of the hydraulic equipment (CVT in this embodiment) are not limited to those shown in FIG. 4, but various characteristics (for example, not linear but curved characteristics, and flow characteristics depending on other conditions such as the number of revolutions). Of course, the above-described effect can be obtained by adjusting the characteristics of the pump 1 to be close to the required characteristics. For example, the pressure loss ΔP1, ΔP2 in the upstream oil passage 31 or the downstream oil passage 33 may be adjusted so as to approach the characteristic of the control flow rate Q required by the CVT according to the pump rotation speed (engine rotation speed). Good.

  In the case of an electronically controlled variable displacement pump that has a solenoid and is capable of adjusting the control flow rate as in this embodiment, the flow rate that is required only when hydraulic equipment (CVT) actually requires the flow rate. It is also conceivable to tune the characteristics of the pump by controlling the solenoid so as to supply the pressure. However, in this case, the control becomes complicated and the capacity of the control unit increases. In the configuration in which the adjustment is performed with the current value, it is necessary to flow an extra current value by that amount, which ultimately causes deterioration of energy efficiency. On the other hand, in the pump 1 of the present embodiment, the characteristics of the pump 1 are adjusted by adjusting the length and diameter of the upstream oil passage 31 or the downstream oil passage 33, that is, by the mechanical configuration. . Therefore, the control configuration can be simplified and the capacity of the control unit can be reduced.

  Also, the hydraulic pressure supplied to the first pressure chamber 35 or the second pressure chamber 36 is not adjusted by the length or diameter of the upstream oil passage 31 or the downstream oil passage 33 but by other mechanical means. It is also conceivable to change according to conditions (pressure, temperature, etc.). However, in this case, the number of parts may increase and the pump 1 may be complicated. On the other hand, in the pump 1 of the present embodiment, the characteristics of the pump 1 are adjusted by adjusting the length and diameter of the upstream oil passage 31 or the downstream oil passage 33, that is, by tuning the existing configuration. It was decided to adjust. Therefore, an increase in the number of parts can be suppressed and the configuration of the pump 1 can be simplified. Note that the length or diameter of only one of the upstream oil passage 31 or the downstream oil passage 33 may be adjusted relative to the conventional one, or the length or diameter of both of the oil passages 31 and 33 may be adjusted. Good. Further, only one of the length and the diameter may be adjusted, or both the length and the diameter may be adjusted.

  In addition, by tuning a mechanical configuration that has conventionally existed, the hydraulic pressure supplied to the first pressure chamber 35 or the second pressure chamber 36 is changed according to various conditions (pressure, temperature, etc.) For example, the size of the clearance (valve clearance) between the spool 41 (the first land portion 410 or the second land portion 411) of the control valve 4 and the valve accommodating hole 400 is controlled, and the first pressure chamber 35 or the second pressure chamber 35 is controlled. It is also conceivable to adjust the amount of hydraulic fluid that leaks from the pressure chamber 36 to the drain chamber 37. However, in this case, since the amount of hydraulic oil leakage is increased or decreased, the efficiency of the pump 1 may be deteriorated. On the other hand, the pump 1 of this embodiment does not adjust the amount of hydraulic oil leakage (while keeping the amount of leakage as necessary as usual), but adjusts the pressure loss ΔP1, ΔP2 resulting from this leakage. Therefore, deterioration of pump efficiency can be suppressed. Further, in the case of a configuration in which the pressure loss is not adjusted, a variation in the control flow rate that does not depend on the orifice differential pressure ΔPM occurs according to the difference in the pressure loss. In order to reduce this variation, a means or process for controlling or mechanically adjusting is required. On the other hand, in the pump 1 of the present embodiment, variation in the control flow rate Q due to the pressure loss difference (ΔP1−ΔP2) can be suppressed in advance by intentionally adjusting the pressure losses ΔP1 and ΔP2.

  In addition, even if it is the structure which is not equipped with a solenoid, the said effect can be obtained. In this embodiment, by providing the solenoid 5, the energy efficiency can be improved by controlling the flow rate by the solenoid 5 as described above. In particular, a hydraulic device (CVT) that supplies hydraulic oil from the pump 1 of this embodiment is an automatic transmission mounted on a vehicle of an automobile. That is, the pump 1 supplies hydraulic oil used in an automatic transmission as a working fluid. Therefore, for example, when the flow rate of hydraulic oil required when the automatic transmission is operated is small on average, energy efficiency can be improved by setting the minimum control flow rate low.

  The pump 1 may supply hydraulic fluid to a hydraulic device other than the automatic transmission, for example, a power steering device that applies a steering assist force to the steering mechanism of the vehicle. The said effect can be acquired by adjusting. Here, since the valve clearance is generally set larger in the automatic transmission than in the power steering device (to prevent contamination), the upstream oil passage 31 and the downstream oil passage 33 are more likely to flow. Large pressure loss difference due to flow. Therefore, when the pump 1 is used in an automatic transmission (CVT) as in this embodiment, there is an advantage that the adjustment of the characteristics of the pump 1 using the pressure losses ΔP1 and ΔP2 is relatively easy.

[effect]
Hereinafter, the effects of the variable displacement pump 1 grasped from the first embodiment will be listed.
(1) A first pressure receiving portion (pressure receiving surface on the x-axis negative direction side of the spool 41) and a second pressure receiving portion (pressure receiving surface on the x-axis positive direction side of the spool 41) on which the pressure of the working fluid (hydraulic oil) acts. A control valve that includes a valve body (spool 41) that moves according to the difference ΔPv between the pressures Pv1 and Pv2 that act on these pressure receiving parts, and that variably controls the discharge flow rate Q of the pump (pump part 2) by the movement of the valve body 4 and a first flow path connecting the upstream side of the throttle section (orifice 320) in the discharge flow path connected to the discharge section (discharge port 28) of the pump and the first pressure receiving section side of the valve body of the control valve 4 ( A second flow path (downstream oil path 33) connecting the upstream oil path 31) and the downstream side of the throttle part (orifice 320) in the discharge flow path and the second pressure receiving part side of the valve body of the control valve 4. The characteristics of the discharge flow rate Q with respect to a predetermined condition (pressure, temperature, etc.) The pressure losses ΔP1 and ΔP2 in the first flow path or the second flow path were adjusted so as to approximate the characteristics required by CVT).
Therefore, with a simple configuration, energy efficiency can be improved while achieving the required performance.
(2) The pressure loss ΔP1, ΔP2 was adjusted by adjusting the length or diameter of the first flow path (upstream oil path 31) or the second flow path (downstream oil path 33).
Therefore, an increase in the number of parts can be suppressed and the configuration of the pump 1 can be simplified.
(3) The predetermined condition is the discharge pressure of the pump or the temperature of the working fluid.
Therefore, the flow rate characteristic with respect to the discharge pressure or temperature can be brought close to the required characteristic.
(4) The pump 1 supplies hydraulic oil for an automatic transmission (CVT) mounted on the vehicle.
Therefore, the flow characteristics of the pump 1 are brought close to those required by the automatic transmission (CVT), and the energy efficiency of the unit composed of the automatic transmission and the pump 1 is improved while achieving the required performance of the automatic transmission. can do. Further, it is relatively easy to adjust the characteristics of the pump 1 using the pressure losses ΔP1 and ΔP2.
(5) A rotor 21 that is driven to rotate by the drive shaft 20, a cam ring 8 that surrounds the rotor 21 and is swingably disposed, and a plurality of pump chambers r are formed so as to protrude and retract on the outer periphery of the rotor 21. And a control valve 4 for controlling the swing amount (eccentric amount δ) of the cam ring 8 by the hydraulic oil discharged from the pump chamber r. The control valve 4 controls the swing of the cam ring 8. A spool 41 is provided as a valve body for switching the oil passages (first and second control oil passages 38, 39) to the control chambers R1, R2.
Therefore, the present invention can be applied to the above type of vane pump.
(6) An elastic member (spring 43) for applying a set load to the spool 41 is provided.
Therefore, the operation of the control valve 4 in the initial stage of the pump operation can be stabilized and the flow rate Q can be controlled stably.
(7) A solenoid 5 for applying a desired urging force to the spool 41 is provided.
Therefore, the energy efficiency can be further improved by controlling the flow rate Q by the solenoid 5.

[Other Examples]
Although the variable displacement pump of the present invention has been described based on the embodiments, the specific configuration of the present invention is not limited to the embodiments, and there are design changes and the like without departing from the scope of the invention. Is included in the present invention. For example, the type of the variable displacement pump 1 is not limited to the vane pump. Further, the type of the control valve 4 is not limited to the spool valve.

DESCRIPTION OF SYMBOLS 1 Pump 2 Pump part 3 Control part 4 Control valve 41 Spool (valve body)
43 Spring (elastic member)
5 Solenoid 8 Cam ring 20 Drive shaft 21 Rotor 22 Vane 28 Discharge port (discharge part)
31 Upstream oil passage (first passage)
33 Downstream oil passage (second passage)
320 Orifice (throttle part)
r Pump room R1 1st control room R2 2nd control room

Claims (6)

It has a first pressure receiving part and a second pressure receiving part where the pressure of the working fluid acts, and has a valve body that moves due to the difference in pressure acting on these pressure receiving parts, and the discharge flow rate of the pump can be varied by the movement of the valve body A control valve to control,
A first flow path connecting the upstream side of the throttle in the discharge flow path connected to the discharge section of the pump and the first pressure receiving section side of the valve body of the control valve;
A second flow path connecting the downstream side of the throttle part in the discharge flow path and the second pressure receiving part side of the valve body of the control valve,
The pressure loss in the first flow path or the second flow path is adjusted so that the characteristic of the discharge flow rate with respect to the predetermined condition approximates the characteristic required by the target device to which the pump supplies the working fluid. Variable displacement pump.   The variable displacement pump according to claim 1, wherein the pressure loss is adjusted by adjusting a length or a diameter of the first flow path or the second flow path.   3. The variable displacement pump according to claim 1, wherein the predetermined condition is a pump discharge pressure or a working fluid temperature.   4. The variable displacement pump according to claim 1, wherein hydraulic oil for an automatic transmission mounted on a vehicle is supplied. A rotor driven to rotate by a drive shaft;
A cam ring that surrounds the rotor and is swingably disposed;
A plurality of vanes which are installed on the outer periphery of the rotor so as to be able to project and sink to form a plurality of pump chambers;
The control valve for controlling the swing amount of the cam ring by the hydraulic oil discharged from the pump chamber,
The control valve is
A spool as the valve body that switches an oil passage to a control chamber that controls swinging of the cam ring;
An elastic member for applying a set load to the spool;
A variable displacement pump according to any one of claims 1 to 4, further comprising a solenoid that applies a desired urging force to the spool. It has a first pressure receiving part and a second pressure receiving part where the pressure of the working fluid acts, and has a valve body that moves due to the difference in pressure acting on these pressure receiving parts, and the discharge flow rate of the pump can be varied by the movement of the valve body A control valve to control,
A first flow path connecting the upstream side of the throttle in the discharge flow path connected to the discharge section of the pump and the first pressure receiving section side of the valve body of the control valve;
A method for adjusting a flow rate characteristic of a variable displacement pump comprising: a second flow path that connects a downstream side of the throttle section in the discharge flow path and a second pressure receiving section side of the valve body of the control valve. There,
The pressure loss in the first flow path or the second flow path is adjusted so that the characteristic of the discharge flow rate with respect to the predetermined condition approximates the characteristic required by the target device to which the pump supplies the working fluid. To adjust the flow characteristics of the variable displacement pump.

Images