JP2019163699A - Pump unit - Google Patents

Abstract

To provide a pump unit which is improved in an energy-saving effect.SOLUTION: The pump unit is provided that has variable delivery flow rate by controlling an eccentric amount of a cam ring via a solenoid valve, the solenoid valve comprises: a solenoid part which is drive-controlled according to a change of an energization amount; and a valve part provided at a variable orifice part, the position of valve part being changed by the change of the energization amount of the solenoid part. The valve part has a first region whose flow passage area becomes minimum according to a change of a position of the variable orifice part, a second region whose flow passage area is changed according to a change of a position of the valve part, and a third region in which the position of the valve part is located between the first region and the second region, and an amount of the change of the flow passage area caused by the change of the position of the valve part is smaller than that of the second region.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a pump device used for, for example, a power steering device of an automobile.

  A variable displacement pump is used to control the hydraulic pressure of a power steering device of an automobile.

  The electronically controlled variable displacement pump variably controls the inherent discharge amount of the variable displacement pump by controlling the eccentric amount of the cam ring via a solenoid. In variably controlling the specific discharge amount of the variable displacement pump, based on the vehicle speed signal detected by the vehicle speed sensor provided on the wheel or the like and the steering angle signal detected by the steering angle sensor provided on the steering device. By controlling the solenoid to variably control the inherent discharge amount of the variable displacement pump, unnecessary drive torque of the variable displacement pump is reduced, and energy saving is achieved.

  As such a prior art, there exists a thing described in patent document 1, for example.

JP 2007-92761 A

  However, the technique described in Patent Document 1 has room for improvement in order to save energy because the control range of the orifice differential pressure for controlling the eccentric amount of the cam ring is narrow.

  The objective of this invention is providing the pump apparatus which solved the said subject and improved the energy-saving effect.

  According to the present invention, in one aspect thereof, a pump device that varies the discharge flow rate of the pump by controlling the eccentric amount of the cam ring via the solenoid valve, the solenoid valve responding to a change in the energization amount. A solenoid unit that is driven and controlled, and a valve unit that is provided in the variable orifice unit and changes its position in accordance with a change in the energization amount of the solenoid unit, the valve unit responding to a change in position in the variable orifice unit A first area in which the flow path area is minimized, a second area in which the flow path area changes according to a change in the position of the valve section, and a position of the valve section in the first area. The third region is located between the second regions, and the amount of change in the flow path area due to the change in the position of the valve unit is smaller than the second region. To have.

  ADVANTAGE OF THE INVENTION According to this invention, the pump apparatus which improved the energy saving effect can be provided.

1 is a system configuration diagram of a variable displacement pump according to an embodiment of the present invention. FIG. 2 is a configuration block diagram of an ECU shown in FIG. 1. It is a control block diagram of ECU shown in FIG. 1 showing 1st Example of this invention. It is a flowchart showing the control content of the solenoid valve which concerns on 1st Example of this invention. It is a figure which concerns on 1st Example of this invention, Comprising: It is a time chart based on the control flow shown in FIG. It is a figure which concerns on 1st Example of this invention, Comprising: It is the graph showing the relationship between a vehicle speed and steering angular acceleration. It is a figure which concerns on 1st Example of this invention, Comprising: It is the graph showing the relationship between a vehicle speed and correction | amendment addition flow volume. It is a graph showing the relationship between the vehicle speed at the time of non-steering and a pump discharge flow rate. 1 is a cross-sectional view of a solenoid according to a first embodiment of the present invention. FIG. 10 is a cross-sectional view of the solenoid taken along section line XX in FIG. 9. FIG. 10 is a cross-sectional view of the solenoid taken along section line XI-XI in FIG. 9. FIG. 10 is a cross-sectional view of the solenoid taken along section line XI-XI in FIG. 9. FIG. 10 is a cross-sectional view of the solenoid taken along section line XI-XI in FIG. 9. It is a figure which shows the relationship between the specific discharge amount of the pump which concerns on a comparative example, and the movement amount of a valve part. It is a figure which shows the relationship between the specific discharge amount of the pump which concerns on a present Example, and the movement amount of a valve part. It is a figure which shows the relationship between the vehicle state which concerns on a present Example, a foot brake, a side brake, a shift position, and a pump discharge amount. It is sectional drawing of the solenoid which concerns on 2nd Example based on this invention. It is sectional drawing of the solenoid which concerns on 2nd Example based on this invention. It is sectional drawing of the solenoid which concerns on 2nd Example based on this invention. It is sectional drawing of the solenoid which concerns on 3rd Example based on this invention.

  Embodiments of a variable displacement pump according to the present invention will be described below in detail with reference to the drawings. In the following embodiments, the variable displacement pump is applied to a hydraulic power steering device as a conventional vehicle steering device.

  1 to 8 show a first embodiment of the present invention. First, the power steering apparatus to which a variable displacement pump (hereinafter simply referred to as “pump”) according to the present invention is applied will be described. As shown in FIG. 1, the power steering device has one end side linked to the steering wheel 1 so as to be integrally rotatable, an input shaft 2 for steering input from the driver, and one end side of the rack and pinion mechanism 4 described later. The other end side of the input shaft 2 is connected to the other end side of the input shaft 2 via a torsion bar (not shown) so as to be relatively rotatable, and the torsion is based on the steering input from the input shaft 2. A pair of pressure chambers P1 and P2 (described later) that are interposed between the output shaft 3 that outputs the steering by the reaction force of the torsional deformation of the bar and the output shaft 3 and the steered wheels, and the interior of which is separated by the piston. A power cylinder 5 that assists the steering output by the output shaft 3 by supplying the hydraulic fluid by selecting it to the hydraulic fluid chamber), a reservoir tank 6 that stores the hydraulic fluid supplied to the power cylinder 5, A pump 10 that sucks up the hydraulic fluid stored in the reservoir tank 6 and pumps it to the pair of pressure chambers P1, P2 of the power cylinder 5, and the input shaft 2 and the output shaft 3 by the torsional deformation of the torsion bar. And a control valve 7 that controls the amount of hydraulic fluid supplied to the power cylinder 5 in accordance with the relative rotation amount (twist amount of the torsion bar) of both the shafts 2 and 3. It is mainly composed.

  The rack and pinion mechanism 4 has a predetermined range in the axial direction of a rack shaft 8 disposed so as to be substantially orthogonal to the pinion teeth (not shown) formed on the outer periphery of one end portion of the output shaft 3 and one end portion of the output shaft 3. The rack shaft 8 is formed in mesh with a rack tooth (not shown), and the rack shaft 8 moves in the left-right direction in FIG. 1 in accordance with the rotation direction of the output shaft 3. As the rack shaft 8 moves in the left-right direction, a knuckle (not shown) linked to both ends of the rack shaft 8 is pushed and pulled, thereby changing the direction of the steered wheels.

  The power cylinder 5 includes a cylinder shaft 5a formed in a substantially cylindrical shape, and a rack shaft 8 as a piston rod penetrating along the axial direction. The power cylinder 5 is a cylinder by a piston (not shown) fixed to the outer periphery of the rack shaft 8. A first pressure chamber P1 and a second pressure chamber P2 which are a pair of pressure chambers are separated in the tube 5a. A propulsive force for the rack shaft 8 is generated based on the hydraulic pressure acting on the first pressure chamber P1 and the second pressure chamber P2, thereby assisting the steering output. Specifically, the first pressure chamber P1 and the second pressure chamber P2 are connected to the reservoir tank 6 and the pump 10 through the control valve 7 through the first to fourth pipes 9a to 9d. The hydraulic fluid discharged from the pump 10 is selectively supplied to one of the first pressure chamber P1 and the second pressure chamber P2, and the hydraulic fluid in the other pressure chamber is returned to the reservoir tank 6. It has become.

  The pump 10 is a vane-type variable displacement pump, and has a pump housing 11 having a pump element accommodating space 11a that is a substantially cylindrical space therein, and is rotatably supported by the pump housing 11. A drive shaft 12 that is rotationally driven by a driving force from an external engine, and is accommodated in the pump element accommodating space 11a, and is driven to rotate counterclockwise in FIG. A pump element 13 that performs a so-called pump action of discharging the sucked hydraulic fluid, and an outer peripheral side of the pump element 13 in the pump element accommodating space 11a are provided so as to be eccentric (movable) with respect to the axis of the drive shaft 12, Based on this eccentric amount, a substantially annular shape that makes the discharge flow rate per rotation of the pump element 13 (hereinafter referred to as “specific discharge amount”) variable. The difference between the first fluid pressure chamber 21a and the second fluid pressure chamber 21b, which will be described later, based on the axial position of the spool valve 15a provided inside the pump ring 11 and slidably housed therein. A control valve 15 that controls the amount of eccentricity of the cam ring 14 by changing the pressure, and a first pressure chamber 15b described later based on a control current that is housed and fixed in the pump housing 11 and output from an ECU 40 (controller) described later. , And a solenoid valve 16 that is a solenoid for controlling the specific discharge amount by changing the differential pressure in the second pressure chamber 15c. The control valve 15 is connected to the pipe 6 b and the suction passage 30. The pump housing 11 of the present embodiment includes a front housing, a rear housing, a pressure plate, and an adapter ring.

  The pump element 13 is rotatably accommodated on the inner peripheral side of the cam ring 14, and the rotor 17 is rotationally driven by the drive shaft 12, and the outer periphery of the rotor 17 is radially notched along the radial direction. A plurality of pump chambers are provided in a space formed between the cam ring 14 and the rotor 17 by being provided so as to be movable in the plurality of slits, protruding outward when the rotor 17 rotates and slidingly contacting the inner peripheral surface of the cam ring 14. The vane 18 of the substantially rectangular plate shape which divides 20 is comprised. A plurality of pump chambers 20 are formed with suction ports 13a and discharge ports 13b. The suction port 13 a opens to a suction region in which the volume of the pump chamber 20 increases with the rotation of the drive shaft 12 among the plurality of pump chambers 20, and is connected to the suction passage 30. The suction passage 30 communicates with the control valve 15. The discharge port 13 b is open to a discharge region where the volume of the pump chamber 20 decreases with the rotation of the drive shaft 12 among the plurality of pump chambers 20, and is connected to the discharge passage 25.

  The cam ring 14 is supported by a swing fulcrum pin 22 through a support groove having a substantially semicircular cross section formed in a notch on the outer periphery of the cam ring 14, and the swing fulcrum pin 22 serves as a fulcrum in the left-right direction in FIG. It can swing. Then, the cam ring 14 swings in the left-right direction in the figure, whereby the volume of each pump chamber 20 increases or decreases, thereby changing the specific discharge amount. Further, a seal member 23 is disposed on the outer peripheral side of the cam ring 14 at a position substantially opposite to the swing fulcrum pin 22 in the radial direction, and the cam ring is configured by the seal member 23 and the swing fulcrum pin 22. A first fluid pressure chamber 21a on the left side and a second fluid pressure chamber 21b on the right side in FIG. Further, in addition to the pressure in the first fluid pressure chamber 21a and the second fluid pressure chamber 21b, the cam ring 14 is acted on by the spring force of the coil spring 24 disposed in the second fluid pressure chamber 21b. The spring force of 24 is always biased toward the first fluid pressure chamber 21a side, that is, the side where the eccentric amount of the cam ring 14 is maximized.

  The first fluid pressure chamber 21 a is a space provided on the radially outer side of the cam ring 14 in the radial direction of the rotational axis of the drive shaft 12, and the amount of eccentricity between the rotational axis of the drive shaft 12 and the center of the inner peripheral edge of the cam ring 14. It is provided in the part where the volume decreases as the value increases.

  The second fluid pressure chamber 21 b is a space provided on the radially outer side of the cam ring 14 in the radial direction of the rotation axis of the drive shaft 12, and the amount of eccentricity between the rotation axis of the drive shaft 12 and the center of the inner peripheral edge of the cam ring 14. It is provided in the part where the volume increases as the value increases.

Thus, the cam ring 14 is provided to be movable in the pump element accommodation space 11a based on the pressure difference between the first fluid pressure chamber 21a and the second fluid pressure chamber 21b.
The control valve 15 includes a spool valve 15a slidably accommodated in a valve hole 11b formed in the pump housing 11 so that the inside of the valve hole 11b and the first pressure chamber 15b on the left side in FIG. And the second pressure chamber 15c. The second pressure chamber 15c is provided with a coil spring 15d that urges a spool valve 15a that operates by a differential pressure between the first pressure chamber 15b and the second pressure chamber 15c.
The hydraulic pressure upstream of the solenoid valve 16 acts on the first pressure chamber 15b, and the hydraulic pressure downstream of the solenoid valve 16 acts on the second pressure chamber 15c. That is, a discharge passage 25 connected to the discharge-side pump chamber 20 via the discharge port 13b is branched into a first discharge passage 25a and a second discharge passage 25b, and the first discharge passage 25a is a fixed orifice portion. The first pressure chamber 15b is connected to the first pressure chamber 15b via a discharge pressure 48 so that the discharge pressure acts on the first pressure chamber 15b. On the other hand, the second discharge passage 25b opens to the outside on the downstream side of the solenoid valve 16 disposed in the middle and is connected to the second pressure chamber 15c, and the second pressure chamber 15c and the outside are decompressed by the solenoid valve 16. The hydraulic pressure is applied. With this configuration, when the spool valve 15 a is positioned on the left side in FIG. 1, a low pressure as a suction pressure is introduced into the first fluid pressure chamber 21 a, and the cam ring 14 is a spring of the coil spring 24. When the spool valve 15a is positioned on the right side in the figure while being pressed by force and held in the maximum eccentric state, a high pressure as a discharge pressure is introduced into the first fluid pressure chamber 21a. Thus, the cam ring 14 moves in a direction in which the amount of eccentricity decreases against the spring force of the coil spring 24.

  The solenoid valve 16 is connected to a vehicle-mounted ECU 40, and the steering angle calculated based on the steering angle, vehicle speed, engine speed, foot brake, side brake, shift position, and steering angle input to the ECU 40. Drive control is performed based on information such as acceleration. The solenoid valve 16 is provided with a variable metering orifice 28 including a fixed orifice portion 26 and a variable orifice portion 27. The variable orifice portion 27 is provided on the variable orifice portion 27 based on information from the ECU 40. The flow path area of the variable orifice portion 27 is changed by the provided valve portion 16b to variably control the differential pressure across the variable metering orifice 28, that is, the differential pressure between the pressure chambers 15b and 15c in the control valve 15. Thus, the specific discharge amount is controlled by controlling the axial position of the spool valve 15a of the control valve 15 to control the eccentric amount of the cam ring 14. The second discharge passage 25 b is branched into a second discharge passage 25 b 1 communicating with the variable orifice portion 27 and a second discharge passage 25 b 2 communicating with the fixed orifice portion 26. A fixed orifice portion 29 is provided in the second discharge passage 25b connected to the second pressure chamber 15c.

  The ECU 40 is supplied with electric power from an in-vehicle battery 31 via an ignition switch 32. The ECU 40 also detects a steering angle sensor 33 for detecting a steering angle by a driver and a vehicle speed. A vehicle speed sensor 34 that detects the engine speed, a foot brake sensor 36 that detects the operating condition of the foot brake, a side brake sensor 37 that detects the operating condition of the side brake, and a position of the shift position. A shift position sensor 38 and the like are connected, and information from each of the sensors 33, 34, 35, 36, 37, and 38 is input.

  More specifically, in the ECU 40, as shown in FIG. 2, a steering angle sensor provided on the input shaft 2 of the power steering apparatus is connected to an MPU (Micro Processor Unit) 50 that drives and controls the solenoid valve 16. The steering angle signal from 33, the vehicle speed signal from the vehicle speed sensor 34 provided in the brake control device provided for each wheel outside the figure, and the engine rotation from the engine rotation sensor 35 provided in the engine control apparatus outside the figure. The numerical signal is given through the CAN interface 41, and the MPU 50 outputs a PWM drive control signal for driving the solenoid valve 16 based on the outputs of the sensors 33, 34, 35, 36, 37, and 38. It has become. The power supply to the MPU 50 is performed from the battery 31 through the fuse 39, the ignition switch 32, the diode 42, and the regulator 43. Here, the regulator 43 normally reduces the battery voltage of about 12V to 5V that is the operating voltage of the MPU 50.

  Further, the PWM drive control signal from the MPU 50 is given to a field effect transistor (FET) 44 as a so-called switching means, and this FET 44 is supplied from the battery 31 through a fuse 39, an ignition switch 32, and a diode 42. The excitation current is supplied to the solenoid portion 16a of the solenoid valve 16 by switching the current to be switched based on the PWM drive control signal.

  Here, the solenoid portion 16a of the solenoid valve 16 has one end connected to the FET 44 and the other end connected to the ground via a current detection resistor 45, and responds to the current flowing through the solenoid portion 16a. The voltage generated across the resistor 45 is supplied to the MPU 50 through the amplifier (AMP) 46 as an actual supply current signal. A free wheel diode 47 provided in parallel with the solenoid unit 16a is connected to the solenoid unit 16a.

  As shown in FIG. 3, the MPU 50 includes a vehicle speed calculation unit 51 that calculates a vehicle speed V based on a vehicle speed signal from a vehicle speed sensor 34, and a steering angle that calculates a steering angle θ based on a steering angle signal from a steering angle sensor 33. A calculation unit 52, a steering angular velocity calculation unit 53 that calculates a steering angular velocity ω based on the steering angle θ calculated by the steering angle calculation unit 52, and a steering angular velocity ω calculated by the steering angular velocity calculation unit 53. A basic discharge flow rate Qω_CMD is calculated based on the steering angular acceleration calculation unit 54 that calculates the steering angular acceleration ωd, the steering angular velocity ω calculated by the steering angular velocity calculation unit 53, and the vehicle speed V calculated by the vehicle speed calculation unit 51. Calculated by the basic discharge flow rate calculation unit 55, the corrected addition flow rate calculation unit 56 that calculates the corrected addition flow rate Qωd_CMD based on the vehicle speed V calculated by the vehicle speed calculation unit 51, and the basic discharge flow rate calculation unit 55. A target command current ICMD that is a command current for supplying the target discharge flow rate QCMD obtained by adding the corrected addition flow rate Qωd_CMD calculated by the corrected addition flow rate calculation unit 56 to the basic discharge flow rate Qω_CMD to the solenoid valve 16 is calculated. PWM based on the PI control based on the difference between the target current calculation unit 57 that performs and the target command current ICMD calculated by the target current calculation unit 57 and the actual supply current Ireal flowing through the solenoid unit 16a detected by the solenoid current detection unit 58. A PI control unit 59 that calculates the duty, and a PWM signal output unit 60 that outputs a PWM drive control signal to the FET 44 based on the PWM duty calculated by the PI control unit 59 are provided.

  Based on the PWM duty calculated by the PI controller 59, the solenoid valve 16 is driven by the FET 44 via the solenoid driving means 61. The solenoid driving means 61 has a function of interrupting the output when the temperature of the solenoid driving means 61 exceeds a predetermined value, a function of limiting the energization amount when an overcurrent flows, and the like. Yes.

  The basic discharge flow rate calculation unit 55 obtains a basic discharge flow rate Qω_CMD from a predetermined map outside the figure based on the vehicle speed V and the steering angular velocity ω. In other words, the basic discharge flow rate calculation unit 55 includes a basic energization amount calculation circuit according to the present invention that is used for calculation of the basic energization amount serving as a base for drive control of the solenoid valve 16 based on the vehicle speed V and the steering angular velocity ω. Constitute. The relationship between the vehicle speed V, the steering angular velocity ω, and the basic discharge flow rate Qω_CMD is such that the basic discharge flow rate Qω_CMD decreases as the vehicle speed V increases, and the steering angular velocity is within the same vehicle speed. The characteristic is such that the basic discharge flow rate Qω_CMD increases as ω increases.

  The corrected added flow rate calculation unit 56 obtains a corrected added flow rate Qωd_CMD from a vehicle speed-corrected added flow rate map as shown in FIG. The relationship between the vehicle speed V and the corrected added flow rate Qωd_CMD is basically such that the corrected added flow rate Qωd_CMD decreases as the vehicle speed V increases, as shown in FIG. Among these, in the high speed running state where the vehicle speed V is equal to or higher than the first predetermined value V1, and in the vehicle stop state or low speed running state where the vehicle speed V is equal to or lower than the second predetermined value V2 lower than the first predetermined value V1, the corrected addition flow rate Qωd_CMD. Is constant.

  In this way, basically, by reducing the corrected addition flow rate Qωd_CMD as the vehicle speed V increases, an appropriate steering assist force according to the vehicle speed V that can stabilize the vehicle behavior against sudden turning is applied. It is a characteristic. Of these, in a high-speed running state where the vehicle speed V is equal to or higher than the first predetermined value V1, the correction addition flow rate Qωd_CMD is kept constant, thereby improving the steering stability and destabilizing the vehicle behavior during the high-speed running. Suppression can be achieved. In addition, when the vehicle speed V is equal to or lower than the second predetermined value V2, it is not necessary to consider instability of the vehicle behavior due to an increase in steering performance, so that the corrected addition flow rate Qωd_CMD is constant. In other words, by maximizing the corrected addition flow rate Qωd_CMD, it is possible to improve the steering responsiveness when the vehicle is stopped or when traveling at a low speed.

  The target current calculation unit 57 adds the basic discharge flow rate Qω_CMD calculated by the basic discharge flow rate calculation unit 55 and the corrected addition flow rate Qωd_CMD calculated by the correction addition flow rate calculation unit 56, and based on a predetermined map (not shown). A target command current ICMD to be supplied to the solenoid valve 16 is obtained. In other words, the target current calculation unit 57 adds the corrected additional flow rate Qωd_CMD that depends on the steering angular acceleration ωd to the basic discharge flow rate Qω_CMD that depends on the steering angular velocity ω, thereby obtaining the total energization amount supplied to the solenoid valve 16. The total energization amount calculation circuit according to the present invention is configured together with the corrected addition flow rate calculation unit 56 for calculation of the energization amount having an increase rate larger than the basic energization amount.

  Further, the MPU 50 has a steering angle sudden turning determination unit that determines whether or not the vehicle is in a sudden turning state based on the vehicle speed V calculated by the vehicle speed calculation unit 51 and the steering angular acceleration ωd calculated by the steering angular acceleration calculation unit 54. 62. The rudder angle sudden turn determination unit 62 is linked to the corrected addition flow rate calculation unit 56 and the target current calculation unit 57 via a predetermined signal switching unit 63, and the rudder angle sudden turn determination unit 62 determines that it is in a sudden turning state. If it is determined, the rapid turn determination flag Fc is set to “1”, and the corrected addition flow rate Qωd_CMD calculated by the correction addition flow rate calculation unit 56 is directly used as the target current calculation unit 57 via the signal switching unit 63. On the other hand, when the steering angle sudden turning determination unit 62 does not determine that the steering is suddenly steered, the sudden addition determination flag Fc is set to “0”, and the correction addition calculated by the correction addition flow rate calculation unit 56 is performed. The flow rate Qωd_CMD is switched to zero via the signal switching means 63 and output to the target current calculation unit 57.

Further, the MPU 50 includes a steering angle sensor failure determination unit 64 that determines an abnormality (failure) of the steering angle sensor 33 based on a steering angle signal from the steering angle sensor 33. And the determination result by this rudder angle sensor failure determination part 64 is output to the rudder angle sudden rotation determination part 62, and when the rudder angle sensor 33 has an abnormality in the rudder angle sensor failure determination part 64 If it is determined, the determination flag Fe is set to “1” to stop the correction control, while the steering angle sensor failure determination unit 64 determines that the steering angle sensor 33 is normal. “0” is set to the determination flag Fe, and the correction control is performed according to the determination of the steering angle sudden rotation determination unit 62. As described above, when the steering angular acceleration ωd shows an abnormal value, the correction control using the steering angular acceleration ωd is stopped, so that the pump is used for proper pump control and the safety of the power steering apparatus is ensured. be able to.
Hereinafter, specific control contents of the solenoid valve 16 based on the rudder angle sudden turning determination by the MPU 50 will be described with reference to FIG.

  As shown in the figure, first, initialization is performed (step S101), and after reading the actual supply current Ireal flowing through the solenoid unit 16a (step S102), the steering angle sensor is based on the steering angle signal from the steering angle sensor 33. A failure determination process 33 is performed (step S103). Although detailed illustration of the failure determination process is omitted, when the steering angle sensor 33 is determined to be abnormal, the correction control is stopped as described above, and the process proceeds to Step S111 described later. On the other hand, when it is determined that the steering angle sensor 33 is normal, the steering angle θ is read in the next step S104, the steering angular velocity ω is calculated from the read steering angle θ (step S105), and then this calculation is performed. A steering angular acceleration ωd is calculated based on the steering angular velocity ω (step S106). Thereafter, after the vehicle speed V is read (step S107), the following sudden turn determination process is performed.

  That is, in the next step S108, after calculating the steering angle sudden rotation determination threshold value ωdth from the vehicle speed-steering angular acceleration map as shown in FIG. 6 based on the vehicle speed V and the steering angular acceleration ωd, the calculated steering angle rapid rotation determination is performed. Based on the use threshold value ωdth, it is determined whether or not the condition of the absolute value of steering angular acceleration | ωd | ≧ the steering angle sudden turn determination threshold value ωdth is satisfied (step S109). Here, when it is determined that the condition is satisfied, a corrected addition flow rate Qωd_CMD corresponding to the vehicle speed V is calculated (step S110). On the other hand, when it is determined that the condition is not satisfied, the corrected addition flow rate Qωd_CMD is calculated. “0” is set (step S111).

  Then, after performing the sudden rotation determination process, a basic discharge flow rate Qω_CMD corresponding to the steering angular velocity ω is calculated in the next step S112, and a correction addition corresponding to the steering angle sudden rotation determination result is added to the calculated basic discharge flow rate Qω_CMD. The target discharge flow rate QCMD is calculated by adding the flow rate Qωd_CMD (step S113). Thereafter, an upper limit process is performed for limiting the upper limit value for the calculated target discharge flow rate QCMD (step S114), and after the process, the upper limit is continuously maintained for a predetermined time, and this is performed at a predetermined ratio. By performing the gradually decreasing peak hold and the gradually decreasing process, the command discharge flow rate Qout is calculated based on the values obtained by these processes (step S115). Thereafter, after calculating the target command current ICMD supplied to the solenoid valve 16 based on the calculated command discharge flow rate Qout (step S116), the calculated target command current ICMD and the actual supply current Ireal flowing through the solenoid unit 16a are calculated. The PWM duty is calculated from the difference using PI control (step S117), and the PWM drive control signal is output to the solenoid valve 16 based on the calculated PWM duty (step S118), and the program ends. It will be.

  Subsequently, the control content based on the above flow will be described with reference to the time chart shown in FIG. 5. When the absolute value | ωd | of the steering angular acceleration exceeds the threshold value ωdth in the steering angle sudden turn determination process (t1 in the figure). The corrected addition flow rate Qωd_CMD corresponding to the vehicle speed V is added to the basic discharge flow rate Qω_CMD. At this time, the basic discharge flow rate Qω_CMD calculated based on the steering angular velocity ω is still in a low state due to the control delay. Therefore, the corrected addition flow rate Qωd_CMD is output as the command discharge flow rate Qout. As a result, the actual discharge flow rate Qreal of the pump can be increased at an early stage as compared with the conventional case where the basic discharge flow rate Qω_CMD is the command discharge flow rate Qout (broken line in the column showing the actual discharge flow rate Qreal in the figure). Thereafter, when the basic discharge flow rate Qω_CMD increases following the increase in the steering angular velocity ω, the basic discharge flow rate Qω_CMD corresponding to the steering angular velocity ω is added to the command discharge flow rate Qout. When Qout reaches the target value of the basic discharge flow rate Qω_CMD, the peak hold process is started (t2 in the figure).

  Eventually, when the absolute value of the steering angular acceleration | ωd | falls below the threshold value ωdth (t3 in the figure), “0” is input to the corrected addition flow rate Qωd_CMD accordingly, but the command discharge flow rate Qout is Even if the target value of the basic discharge flow rate Qω_CMD exceeds the peak by the peak hold process, the target value of the basic discharge flow rate Qω_CMD is held (t4 in the figure). Thereafter, when a predetermined time elapses from the start of the peak hold process (t5 in the figure), the process is terminated and the gradual reduction process is started, so that the steering angular acceleration ωd that does not exceed the threshold value ωdth during that time. Even if the steering operation is performed, the command discharge flow rate Qout is gradually decreased at a predetermined rate regardless of the steering operation and returned to the initial value (t6 in the figure).

  As described above, in this embodiment, the discharge flow rate correction of the pump 10 is performed with the steering angular acceleration ωd that reflects the steering responsiveness requested by the driver. Therefore, the discharge flow rate correction is performed based on the steering angular velocity ω. Compared to the conventional case, the pump 10 can increase the specific discharge amount earlier and ensure the necessary discharge flow rate (see hatching in the column showing the actual discharge flow rate Qreal in FIG. 5). As a result, pump control according to the steering response required by the driver is possible.

  In other words, the state in which the steering angular acceleration ωd is large is a state in which the driver is steered suddenly and a state in which the discharge flow rate of the pump 10 needs to be increased. Based on the steering angular acceleration ωd as a reference, the target discharge flow rate QCMD (command discharge flow rate Qout) is determined based on the increase rate of the basic discharge flow rate Qω_CMD by determining whether or not the steering angular acceleration ωd exceeds the threshold value ωdth. The discharge flow rate of the pump 10 can be increased earlier than before by performing correction control so that the increase rate of the basic energization amount increases, that is, the increase rate of the total energization amount becomes larger than the increase rate of the basic energization amount. This makes it possible to ensure high responsiveness when supplying hydraulic fluid to the power steering device.

  Moreover, in the correction control, the target value of the target discharge flow rate QCMD and the target value of the basic discharge flow rate Qω_CMD when the steering angular acceleration ωd is equal to or greater than the threshold value ωdth are equal, that is, the target value of the total energization amount Since the basic energization amount is set to be equal to the target value, the increase in the total energization amount with respect to the basic energization amount is used only to increase the responsiveness of the solenoid valve 16. There is no possibility of changing the magnitude of the steering assist force with respect to the steering operation. For this reason, the uncomfortable feeling in the steering operation can be suppressed.

  In the case of the pump 10, in the engine idle state where the vehicle is stopped, in the non-steering state, the specific discharge amount is 5 L / min at a maximum as shown in FIG. As shown in FIG. 7, a corrected addition flow rate Qωd_CMD of 7 L / min at the maximum is added. As described above, when the engine is in the idling operation state, the solenoid valve 16 is controlled so that the specific discharge amount is smaller in the non-steering state than in the steering state. Pump load can be reduced while ensuring the occurrence.

  Further, since the pump 10 is configured not to directly drive the cam ring 14 by the solenoid valve 16 but to drive the spool valve 15a of the control valve 15, the weight of the solenoid valve 16 to be driven can be reduced. As a result, the drive response by the solenoid valve 16 can be improved. As a result, the steering response according to the power steering apparatus is further improved.

  In recent years, there has been a demand for further reduction in fuel consumption of automobiles for the purpose of improving environmental performance. In order to reduce fuel consumption, in addition to improving engine efficiency, it is effective to reduce the load on equipment that uses the engine as a drive source. In a power steering device, it is required to reduce a pump load that is a supply source of hydraulic fluid. For example, when the foot brake or the side brake is operating in the engine idle state where the vehicle is stopped, and the shift lever is at the position of neutral N or parking P, the possibility of steering operation is low. There is no problem even if the specific discharge amount of the pump is made smaller. In the conventional pump, since the control range of the orifice differential pressure is narrow, there is room for improvement in reducing the inherent discharge amount of the pump in the non-steering state.

  Hereinafter, the configuration of this embodiment for further reducing the pump load and realizing a smooth return to the steering state will be described with reference to the drawings.

  9 is a cross-sectional view of the solenoid according to the first embodiment of the present invention, FIG. 10 is a cross-sectional view of the solenoid along the cutting line XX of FIG. 9, and FIGS. 11 to 13 are cutting lines XI-XI of FIG. It is sectional drawing of a solenoid.

  As shown in FIGS. 9 to 13, the solenoid valve 16 includes a solenoid part 16 a, a plunger 71, and a stator 72 inside a housing member 70. The housing member 70 is provided with a connector portion 73. Inside the connector portion 73, a terminal 73a is provided and is electrically connected to the ECU 40.

  The solenoid unit 16a is an electromagnetic coil that generates a magnetic force when energized, and is supplied with a drive current via a terminal 73a. The solenoid portion 16a generates magnetic lines of force when energized to form a magnetic field, thereby causing the stator 72 to be magnetically attracted to the stator 72. The plunger 71 is a movable core formed of a soft magnetic metal material, and is supported on the inner peripheral side of the solenoid portion 16a so as to be movable (stroked) in the X-axis direction.

  The plunger 71 is disposed so as to be movable in the axial direction within a predetermined range in response to a thrust (magnetic attraction force in the positive direction of the X-axis) generated by the magnetic circuit formed by energizing the solenoid portion 16a. A rod 74, which is a rod-like member, is fixed to the shaft center of the plunger 71 by press fitting or the like. The rod 74 is installed so as to penetrate through a hole formed through the axial center of the plunger 71, and protrudes from both axial ends of the plunger 71.

  The positive end of the rod 74 in the X-axis direction is provided so as to be able to contact the valve portion 16b. The rod 74 transmits the thrust in the positive X-axis direction acting on the plunger 71 to the valve portion 16b according to the drive current supplied to the solenoid portion 16a, and biases the valve portion 16b.

  The variable orifice part 27 has a cylindrical part 75 formed in a hollow, and a valve part 16b and a spring 76 are disposed therein. The spring 76 has a non-linear characteristic. The valve portion 16 b moves in the cylindrical portion 75 of the variable orifice portion 27 and changes the flow path area according to the change in the position of the valve portion 16 b in the variable orifice portion 27. The spring 76 urges the valve portion 16b toward the negative X-axis direction (solenoid portion 16a side).

  When the solenoid portion 16a is energized, the plunger 71 overcomes the urging force of the spring 76 and moves the valve portion 16b in the positive direction of the X axis. The solenoid part 16a changes the magnetic attraction to the plunger 71 by controlling the drive current supplied to the solenoid part 16a, and changes the position of the valve part 16b.

  The valve portion 16b is formed with a plurality of variable orifice passage portions 77 (77a, 77b) which are through holes through which the hydraulic fluid passes. In this embodiment, the opening area of the variable orifice passage portion 77a is larger than the opening area of the variable orifice passage portion 77b.

  Next, the flow path configuration will be described. A joining portion 78 is connected to the cylindrical portion 75 of the variable orifice portion 27. As shown in FIG. 10, the diameter of the merging portion 78 is formed larger than the diameter of the cylindrical portion 75. In other words, the merging portion 78 is arranged so as to cover the outer periphery of the cylindrical portion 75.

  The second discharge passage 25b (25b2) and the fourth pipe 9d are connected to the junction 78 through the fixed orifice 26 (FIGS. 9 and 10). Furthermore, the merging portion 78 is connected to the second pressure chamber of the control valve 15 via the fixed orifice portion 29 (FIG. 11). In other words, the merging portion 78 is joined by the second discharge passages 25b1 and 25b2, and constitutes the second discharge passage 25b as the discharge passage 25.

  As shown in FIG. 9, the discharge passage 25 is branched into a first discharge passage 25a and a second discharge passage 25b1, and the second discharge passage 25b1 is connected to the spring 76 side of the cylindrical portion 75 in the variable orifice portion 27. ing. As shown in FIGS. 11 to 13, the first discharge passage 25 a is connected to the first pressure chamber 15 b of the control valve 15. A pipe 79 is connected to the control valve 15, and the pipe 79 is connected to the pipe 6 b and the suction passage 30.

  Next, operations of the solenoid valve 16 and the control valve 15 will be described with reference to FIGS.

  As shown in FIG. 11, when a drive current is supplied to the solenoid portion 16a (when energized), the valve portion 16b overcomes the urging force of the spring 76 and moves in the X axis positive direction, that is, the left side in FIG. To do. At this time, the valve portion 16b of the variable orifice portion 27 is joined by the variable orifice passage portion 77b, which is a smaller hole than the variable orifice passage portion 77a, among the plurality of variable orifice passage portions 77a and 77b formed in the valve portion 16b. It moves to a position that overlaps the portion 78. The hydraulic fluid that has passed through the second discharge passage 25b1 passes through the variable orifice passage portion 77b formed in the valve portion 16b of the variable orifice portion 27, is discharged to the junction portion 78, and is supplied through the fixed orifice portion 26. It merges with the hydraulic fluid in the second discharge passage 25b2. The hydraulic fluid supplied from the second discharge passage 25b1 and the second discharge passage 25b2 flows into the second pressure chamber 15c of the control valve 15 through the fixed orifice portion 29. The opening of the valve portion 16b in the variable orifice portion 27 is minimized, and the hydraulic fluid flowing into the second pressure chamber 15c of the control valve 15 is also minimized. In the present embodiment, a region where the flow area of the variable orifice portion 27 is minimum is referred to as a first region.

  When the variable orifice passage portion 77b formed in the valve portion 16b of the variable orifice portion 27 is in a position where it overlaps with the merging portion 78, the pressure in the first pressure chamber 15b is larger than that in the second pressure chamber 15c (the first pressure chamber 15b). 11), the pressure in the first pressure chamber 15b overcomes the urging force of the coil spring 15d, and the spool valve 15a is on the right side in FIG. Move to. High pressure is introduced into the first fluid pressure chamber 21a, and the cam ring 14 overcomes the spring force of the coil spring 24 and is held in the minimum eccentric state. In other words, the cam ring 14 is in a maximum eccentric state where the rotation axis of the drive shaft 12 and the center of the inner periphery of the cam ring 14 are the maximum with respect to the position where the rotation axis of the drive shaft 12 and the center of the inner periphery of the cam ring 14 coincide. Move to the opposite negative eccentricity. The negative eccentric state occurs when the valve portion 16b is open to the merging portion 78 only in the first region. When the cam ring 14 is in the minimum eccentric state, the inherent discharge amount of the pump decreases.

  Next, the case where the drive current is not supplied to the solenoid unit 16a will be described with reference to FIG. As shown in FIG. 12, when the drive current is not supplied to the solenoid portion 16a (when the energization amount is zero), the valve portion 16b is attached to the negative direction of the X axis, that is, toward the solenoid portion 16a by the biasing force of the spring 76. Be forced. At this time, the plurality of variable orifice passage portions 77 a and 77 b formed in the valve portion 16 b of the variable orifice portion 27 are in a position overlapping the merging portion 78. The hydraulic fluid that has passed through the second discharge passage 25 b 1 is discharged from the plurality of variable orifice passage portions 77 a and 77 b formed in the valve portion 16 b of the variable orifice portion 27 to the joining portion 78 and is supplied via the fixed orifice portion 26. And the working fluid in the second discharge passage 25b2. The hydraulic fluid supplied from the second discharge passage 25b1 and the second discharge passage 25b2 flows into the second pressure chamber 15c of the control valve 15 through the fixed orifice portion 29. The opening of the valve portion 16b in the variable orifice portion 27 is maximized, and the hydraulic fluid flowing into the second pressure chamber 15c of the control valve 15 is also maximized. A region where the flow path area of the variable orifice portion 27 is maximum (larger than the first region) is referred to as a second region.

  When the plurality of variable orifice passage portions 77a and 77b formed in the valve portion 16b of the variable orifice portion 27 are in a position overlapping the joining portion 78, a differential pressure between the first pressure chamber 15b and the second pressure chamber 15c is generated. Since the biasing force of the coil spring 15d is smaller than the pressure of the first pressure chamber 15b, the spool valve 15a is positioned on the left side in FIG. A low pressure, which is a suction pressure, is introduced into the first fluid pressure chamber 21a, and the cam ring 14 is pressed by the spring force of the coil spring 24 and held in the maximum eccentric state. When the cam ring 14 is in the maximum eccentric state, the inherent discharge amount of the pump increases.

  The solenoid valve 16 of the present embodiment is configured such that the second region opens to the merging portion 78 when the energization amount to the solenoid portion 16a is zero. According to the present embodiment, the solenoid valve 16 is a so-called normally open valve, so that a necessary flow rate can be ensured even when the solenoid valve 16 is poorly energized.

  Next, means for controlling the valve portion 16b to an intermediate position will be described with reference to FIG. As shown in FIG. 13, when a drive current is supplied to the solenoid portion 16a, the valve portion 16b overcomes the urging force of the spring 76 and moves to the X axis positive direction, that is, the left side in FIG. At this time, of the plurality of variable orifice passage portions 77 a and 77 b formed in the valve portion 16 b of the variable orifice portion 27, a part (for example, half) of the variable orifice passage portion 77 a and the variable orifice passage portion 77 b are joined to the joining portion 78. Overlapping position. The working fluid that has passed through the second discharge passage 25b1 is discharged from the variable orifice passage portion 77a formed in the valve portion 16b of the variable orifice portion 27 (for example, half) and the variable orifice passage portion 77b to the joining portion 78, It merges with the hydraulic fluid in the second discharge passage 25b2 supplied through the fixed orifice portion 26. The hydraulic fluid supplied from the second discharge passage 25b1 and the second discharge passage 25b2 flows into the second pressure chamber 15c of the control valve 15 through the fixed orifice portion 29. The opening of the valve portion 16b in the variable orifice portion 27 is intermediate, and the hydraulic fluid flowing into the second pressure chamber 15c of the control valve 15 is also intermediate. A region where the flow path area of the variable orifice portion 27 is between the first region and the second region and the flow channel area is smaller than the second region is referred to as a third region.

  When a part (for example, half) of the variable orifice passage portion 77a formed in the valve portion 16b of the variable orifice portion 27 and the variable orifice passage portion 77b overlap with the merging portion 78, the second pressure chamber 15c In comparison, since the pressure in the first pressure chamber 15b is large (the differential pressure between the first pressure chamber 15b and the second pressure chamber 15c is large), the pressure in the first pressure chamber 15b overcomes the urging force of the coil spring 15d, and the spool The valve 15a is located on the right side in FIG. 11, that is, in the negative X-axis direction. The hydraulic fluid having a pressure lower than that in the second region is introduced into the first fluid pressure chamber 21a, so that the cam ring 14 overcomes the spring force of the coil spring 24 and the eccentric state becomes smaller than that in the second region. Retained. When the eccentric state of the cam ring 14 is reduced, the inherent discharge amount of the pump is reduced. In this embodiment, the amount of overlap with the merging portion 78 in the variable orifice passage portion 77a is halved, but may be, for example, about 1/3 of the overlap amount. In this case, the variable orifice passage portion 77a in the third region has a smaller width in the direction perpendicular to the moving direction of the valve portion 16b than the variable orifice passage portion 77a in the second region. According to the present embodiment, since the width in the direction perpendicular to the moving direction of the valve portion 16b, that is, the opening width is small in the third region, the change in the opening area with respect to the change in the stroke of the valve portion 16b is reduced. While moving the valve portion 16b, it is possible to obtain a state in which the flow rate recovery response is good while the flow rate is small.

  Thus, in this embodiment, the position of the valve portion 16b in the variable orifice portion 27 is changed by changing the energization amount to the solenoid portion.

  The variable orifice passage portions 77a and 77b are provided in the first region, the second region, or the third region, and are joined together according to a change in the position of the valve portion 16b in the variable orifice portion 27. The opening amount of the variable orifice passage portions 77a and 77b with respect to 78 can be changed. The variable orifice passage portions 77a and 77b in the third region have the amount of change in the overlap amount between the variable orifice passage portions 77a and 77b and the joining portion 78 in accordance with the change in the position of the valve portion 16b in the variable orifice portion 27. It is smaller than the second region. According to the present embodiment, by changing the overlap amount of the variable orifice passage portion 77 with respect to the stroke change of the valve portion 16b, the responsiveness of returning the flow rate while keeping the flow rate small while moving the valve portion 16b. Can get a good condition.

  In addition, the variable orifice passage portion 77b is always open to the joining portion 78 in the first region, the second region, and the third region, and the variable orifice passage portion 77b functions as a fixed orifice passage portion. Have. In other words, the valve portion 16b of the present embodiment includes a fixed orifice passage portion and a variable orifice passage portion which are through holes through which the hydraulic fluid passes. The variable orifice passage portion 77a is provided in a region where the second region and the third region are combined. In this embodiment, by setting the first region as a so-called fixed orifice, a minimum flow rate can be ensured regardless of a change in the position of the valve portion.

  Further, the variable orifice passage portion 77 a includes a first opening in which all of the openings overlap with the merging portion 78, and a second opening in which a part of the opening overlaps with the merging portion 78. Yes. The variable orifice passage portion 77b functioning as a fixed orifice is made smaller than the area of the first opening or the second opening. According to the present embodiment, by sufficiently reducing the opening area of the variable orifice passage portion 77b functioning as a fixed orifice, the discharge amount in a state where the discharge amount is minimum can be sufficiently reduced.

  Next, the relationship between the specific discharge amount of the pump and the movement amount of the valve portion 16b will be described with reference to FIGS. 14A and 14B. FIG. 14A is a diagram showing the relationship between the specific discharge amount of the pump according to the comparative example and the movement amount of the valve portion, and FIG. 14B is a diagram showing the relationship between the specific discharge amount of the pump according to this embodiment and the movement amount of the valve portion. It is.

  In FIG. 14A, in the steering state in which the steering operation is performed when the vehicle speed is low, the solenoid unit is controlled so that the specific discharge amount of the pump is increased. When the vehicle speed increases and the vehicle travels normally, there is a low possibility that the steering operation will be performed. In the steering preparation state in the comparative example, the specific discharge amount of the pump is maintained in a predetermined state in preparation for a smooth transition to the steering state, so it is difficult to further reduce the specific discharge amount of the pump. It was.

  On the other hand, in this embodiment, as shown in FIG. 14B, the first region has a minimum discharge state in which the specific discharge amount of the pump is minimized. In this embodiment, since the pump has a minimum discharge state in which the specific discharge amount of the pump is minimized, it is possible to suppress the pump from operating more than necessary and to save energy. For example, when the configuration of this embodiment is applied to a power steering device for an automobile, fuel consumption can be reduced.

  In this embodiment, according to various conditions, the solenoid valve 16 is controlled to the first region, the second region, and the third region to control the specific discharge amount of the pump. In particular, when the steering assist for the steering operation is not necessary, the inherent discharge amount of the pump is minimized. Hereinafter, conditions for minimizing the specific discharge amount of the pump will be described with reference to FIG.

  FIG. 15 is a diagram illustrating the relationship between the vehicle state, the foot brake, the side brake, the shift position, and the pump discharge amount according to the present embodiment.

  In FIG. 15, there are patterns from A to F, for example. In the pattern A, the vehicle speed sensor 34 detects that the vehicle is stopped, the foot brake sensor 36 detects that the foot brake is ON, and the ECU 40 receives these signals. The vehicle is in a state of stopping at a signal. “-” In the figure does not matter. In such a state, since the possibility of performing the steering operation is low, the solenoid valve 16 is controlled so that the specific discharge amount of the pump is minimized.

  In the pattern B, the vehicle speed sensor 34 detects that the vehicle is stopped, the side brake sensor 37 detects that the side brake is ON, and the ECU 40 receives these signals. The engine is idling and the vehicle is parked. In such a state, since the possibility of performing the steering operation is low, the solenoid valve 16 is controlled so that the specific discharge amount of the pump is minimized.

  In the pattern C, the vehicle speed sensor 34 detects that the vehicle is stopped, the shift position sensor 38 detects that the shift position is in the neutral N or parking P position, and the ECU 40 receives these signals. Yes. The engine is idling and the vehicle is parked. In such a state, since the possibility of performing the steering operation is low, the solenoid valve 16 is controlled so that the specific discharge amount of the pump is minimized.

  In pattern D, the foot brake is in an OFF state. This is a state in which the vehicle starts running from the pattern A state where the vehicle is stopped waiting for a signal. In such a state, the driver may perform a steering operation to turn right or left at the intersection. Therefore, the solenoid valve 16 is controlled so that the specific discharge amount of the pump is increased.

  In pattern E, the side brake is in an OFF state. This is a state in which the vehicle starts running from the pattern B state where the vehicle is parked or stopped. In such a state, there is a possibility that the driver performs a steering operation and returns the vehicle from the parking / stopping position to the main line of the roadway, so the solenoid valve 16 is controlled so that the specific discharge amount of the pump is increased.

  In pattern F, the shift position is in the drive D state. This is a state where the vehicle starts running from the state of the pattern C where the vehicle is parked or stopped. In such a state, there is a possibility that the driver performs a steering operation and returns the vehicle from the parking / stopping position to the main line of the roadway, so the solenoid valve 16 is controlled so that the specific discharge amount of the pump is increased.

  In this embodiment, as shown in FIGS. 1 and 2, the vehicle speed sensor 34, the foot brake sensor 36, the side brake sensor 37, and the shift position sensor 38 are provided, and signals relating to the driving state of the vehicle detected by these sensors are sent to the ECU 40. And the solenoid valve 16 is controlled to control the specific discharge amount of the pump.

  According to the present embodiment, the first region of the valve portion 16b has a minimum discharge amount, the third region has a small discharge amount, and the flow rate recovery (increase) responsiveness is secured. It is possible to realize three states, that is, a state where a necessary flow rate is ensured in this area.

  In addition, according to the present embodiment, the power steering device can reduce the pump driving load and the fuel (electric power) consumption by minimizing the discharge amount of the pump when the steering assist is not necessary. . Accordingly, by selecting and using the first, second, or third region in accordance with the driving state or steering state of the vehicle, it is possible to ensure both steering response and energy saving.

  Further, in the present embodiment, the ECU 40 drives and controls the solenoid valve 16 so that only the first region opens to the junction portion 78 based on a signal related to the foot brake, side brake, shift position, or vehicle speed of the vehicle. I am doing so. Since the need for steering assistance is low while the vehicle is stopped, the discharge amount of the pump can be minimized. This determination while the vehicle is stopped can be determined based on the foot brake, side brake, shift position, and vehicle speed.

  Furthermore, the solenoid valve 16 may be controlled using a signal from the steering angle sensor 33. For example, when the steering angle sensor 33 detects that the steering shaft is at the stroke end, the ECU 40 receives the signal, and the solenoid valve 16 b opens the valve portion 16 b to the merging portion 78 only in the first region. The drive is controlled.

  When the steered shaft is in the vicinity of the stroke end, the need for steering assistance is low, and therefore the energy saving effect can be improved by minimizing the pump discharge amount.

  Further, the ECU 40 may drive and control the solenoid valve 16 so that the third region opens to the junction portion 78 when the vehicle speed is not less than a predetermined vehicle speed and the steering operation is not performed. When the steering operation is not performed, a large discharge flow rate is not required, and therefore it is preferable that the discharge flow rate be lower than the state in which the second region is used. On the other hand, since the vehicle speed is equal to or higher than the predetermined vehicle speed, the steering operation is performed, and the necessary discharge flow rate may increase suddenly. Therefore, the steering response can be ensured by increasing the discharge flow rate as compared with the state in which only the first region is used, and setting the position of the valve portion closer to the second region.

  Furthermore, the ECU 40 may drive-control the solenoid valve 16 according to a change in the load applied to the steered wheel shaft. As a means for detecting the load applied to the steered shaft, for example, a torque sensor that detects torsion of a torsion bar is used. When the load applied to the steering shaft of the steered wheel (for example, the front turning shaft) is large (for example, when the load is large), the steering load increases. Therefore, the steering load can be reduced by increasing the discharge flow rate. it can. On the other hand, when the load applied to the steering shaft is small, the energy saving effect can be improved by reducing the discharge flow rate. Thus, by controlling the solenoid in accordance with the change in the load applied to the steering shaft, it is possible to achieve both reduction of the steering load and improvement of the energy saving effect.

  Next, a second embodiment according to the present invention will be described with reference to FIGS. 16 to 18 are sectional views of a solenoid according to the second embodiment of the present invention.

  The second embodiment differs from the first embodiment in that a spool valve 16c is used as the valve portion of the solenoid valve 16. The same components as those in the first embodiment are denoted by the same reference numerals, and detailed description thereof is omitted.

  The spool valve 16c is formed with a concave portion 16c1 that is recessed from the outside in the radial direction of the spool valve toward the inside, and in the concave portion 16c1 from the solenoid portion 16a toward the spring 76 in the moving direction of the spool valve 16c. An inclined portion 16c2 is formed such that the width becomes narrow in a direction perpendicular to the right angle. The hydraulic fluid discharged from the variable orifice passage portion 77 is discharged from the concave portion 16 c 1 overlapping the merging portion 78 to the merging portion 78.

  As shown in FIG. 16, when a drive current is supplied to the solenoid unit 16a (when energized), the spool valve 16c overcomes the urging force of the spring 76 and moves in the positive direction of the X axis, that is, the left side of FIG. To do. At this time, the spool valve 16c of the variable orifice portion 27 moves to a position where the inclined portion 16c2 which is a part of the concave portion 16c1 formed in the spool valve 16c overlaps the merging portion 78. The hydraulic fluid that has passed through the second discharge passage 25b1 passes through the variable orifice passage portion 77 formed in the spool valve 16c of the variable orifice portion 27, and is discharged from the inclined portion 16c2 formed in the concave portion 16c1 to the joining portion 78. It merges with the hydraulic fluid in the second discharge passage 25b2 supplied through the fixed orifice portion 26. The hydraulic fluid that has passed through the variable orifice passage portion 77 is squeezed by the inclined portion 16c2 and discharged to the junction portion 78. In other words, the opening of the spool valve 16c in the variable orifice portion 27 is minimized. The hydraulic fluid supplied from the second discharge passage 25b1 and the second discharge passage 25b2 flows into the second pressure chamber 15c of the control valve 15 through the fixed orifice portion 29. Since the opening of the spool valve 16c in the variable orifice portion 27 is minimum, the hydraulic fluid flowing into the second pressure chamber 15c of the control valve 15 is also minimum. A region where the flow path area of the variable orifice portion 27 is minimized is referred to as a first region.

  When the inclined portion 16c2 of the concave portion 16c1 formed in the spool valve 16c of the variable orifice portion 27 is in a position overlapping the merging portion 78, the pressure of the first pressure chamber 15b is larger than that of the second pressure chamber 15c ( Since the pressure difference between the first pressure chamber 15b and the second pressure chamber 15c is large), the pressure of the first pressure chamber 15b overcomes the biasing force of the coil spring 15d, and the spool valve 15a of the control valve 15 is on the right side in FIG. That is, it is located in the X axis negative direction. High pressure is introduced into the first fluid pressure chamber 21a, and the cam ring 14 overcomes the spring force of the coil spring 24 and is held in the minimum eccentric state. In other words, the cam ring 14 is in a maximum eccentric state where the rotation axis of the drive shaft 12 and the center of the inner periphery of the cam ring 14 are the maximum with respect to the position where the rotation axis of the drive shaft 12 and the center of the inner periphery of the cam ring 14 coincide. Move to the opposite negative eccentricity. The negative eccentric state occurs when the inclined portion 16c2 of the recess 16c1 formed in the spool valve 16c is open to the merging portion 78 only in the first region. When the cam ring 14 is in the minimum eccentric state, the inherent discharge amount of the pump decreases.

  Next, the case where the drive current is not supplied to the solenoid unit 16a will be described with reference to FIG. As shown in FIG. 17, when the drive current is not supplied to the solenoid portion 16a (when the energization amount is zero), the spool valve 16c is attached to the negative direction of the X axis, that is, toward the solenoid portion 16a by the biasing force of the spring 76. Be forced. At this time, the concave portion 16c1 and the inclined portion 16c2 formed in the spool valve 16c of the variable orifice portion 27 are moved to a position where they overlap the merging portion 78. The hydraulic fluid that has passed through the second discharge passage 25b1 passes through the variable orifice passage portion 77, and is discharged from the concave portion 16c1 and the inclined portion 16c2 formed in the spool valve 16c of the variable orifice portion 27 to the merging portion 78 to be fixed to the fixed orifice portion. 26 joins the hydraulic fluid in the second discharge passage 25b2 supplied via the second discharge passage 25b2. The hydraulic fluid supplied from the second discharge passage 25b1 and the second discharge passage 25b2 flows into the second pressure chamber 15c of the control valve 15 through the fixed orifice portion 29. The opening of the valve portion 16b in the variable orifice portion 27 is maximized, and the hydraulic fluid flowing into the second pressure chamber 15c of the control valve 15 is also maximized. A region where the flow path area of the variable orifice portion 27 is maximum (larger than the first region) is referred to as a second region.

  When the concave portion 16c1 and the inclined portion 16c2 formed in the spool valve 16c of the variable orifice portion 27 are in a position overlapping the merging portion 78, the differential pressure between the first pressure chamber 15b and the second pressure chamber 15c is small. The biasing force of the coil spring 15d exceeds the pressure in the first pressure chamber 15b, and the spool valve 15a of the control valve 15 is located on the left side in FIG. A low pressure, which is a suction pressure, is introduced into the first fluid pressure chamber 21a, and the cam ring 14 is pressed by the spring force of the coil spring 24 and held in the maximum eccentric state. When the cam ring 14 is in the maximum eccentric state, the inherent discharge amount of the pump increases.

  The solenoid valve 16 of the present embodiment is configured such that the second region opens to the merging portion 78 when the energization amount to the solenoid portion 16a is zero. According to the present embodiment, the solenoid valve 16 is a so-called normally open valve, so that a necessary flow rate can be ensured even when the energization is defective.

  Next, means for controlling the valve portion 16b to an intermediate position will be described with reference to FIG. When a drive current is supplied to the solenoid part 16a, the spool valve 16c overcomes the biasing force of the spring 76 and moves in the X-axis positive direction, that is, the left side in FIG. At this time, a part of the recessed portion 16c1 formed in the spool valve 16c of the variable orifice portion 27 and the inclined portion 16c2 move to a position where the joining portion 78 overlaps. The hydraulic fluid that has passed through the second discharge passage 25b1 passes through the variable orifice passage portion 77 formed in the spool valve 16c of the variable orifice portion 27, and is discharged from a part of the recess portion 16c1 and the inclined portion 16c2 to the joining portion 78, It merges with the hydraulic fluid in the second discharge passage 25b2 supplied through the fixed orifice portion 26. The hydraulic fluid supplied from the second discharge passage 25b1 and the second discharge passage 25b2 flows into the second pressure chamber 15c of the control valve 15 through the fixed orifice portion 29. The opening of the spool valve 16c in the variable orifice portion 27 is in the middle, and the hydraulic fluid flowing into the second pressure chamber 15c of the control valve 15 is also in the middle. A region where the flow path area of the variable orifice portion 27 is between the first region and the second region and the flow channel area is smaller than the second region is referred to as a third region.

  When a part of the concave portion 16c1 formed in the spool valve 16c of the variable orifice portion 27 and the inclined portion 16c2 overlap with the merging portion 78, the pressure in the first pressure chamber 15b is higher than that in the second pressure chamber 15c. (The pressure difference between the first pressure chamber 15b and the second pressure chamber 15c is large), the pressure in the first pressure chamber 15b overcomes the biasing force of the coil spring 15d, and the spool valve 15a of the control valve 15 is shown in FIG. Move to the right side of, i.e., in the negative direction of the X-axis. A pressure lower than that in the second region is introduced into the first fluid pressure chamber 21a, and the cam ring 14 is held in a state where the eccentric state is smaller than that in the second region by overcoming the spring force of the coil spring 24. The When the eccentric state of the cam ring 14 is reduced, the inherent discharge amount of the pump is reduced.

  Next, a third embodiment according to the present invention will be described with reference to FIG. FIG. 19 is a cross-sectional view of the solenoid according to the third embodiment of the present invention, showing the state of the third region.

  The valve portion 16b is formed with a plurality of variable orifice passage portions 77 (77a, 77b) which are through holes through which the hydraulic fluid passes. In this embodiment, the opening area of the variable orifice passage portion 77a is larger than the opening area of the variable orifice passage portion 77b.

  When a drive current is supplied to the solenoid part 16a (when energized), the valve part 16b is moved so that the variable orifice passage part 77b of the valve part 16b overlaps the merging part 78. This state is the first region.

  Further, when the drive current is not supplied to the solenoid portion 16a (when no current is supplied), the valve portion 16b is moved so that the variable orifice passage portion 77a of the valve portion 16b overlaps the merging portion 78. This state is the second area. The operations of the control valve 15 and the pump in the first region and the second region are the same as those in the first and second embodiments.

  When controlling the valve portion 16b to an intermediate position, as shown in FIG. 19, when a drive current is supplied to the solenoid portion 16a, the valve portion 16b overcomes the urging force of the spring 76, so Move to the left side of 19. At this time, the valve portion 16b of the variable orifice portion 27 moves to a position where the variable orifice passage portion 77 (77a, 77b) does not overlap the junction portion 78. In other words, in the third region, the variable orifice passage portion 77 has a width of zero in a direction perpendicular to the moving direction of the valve portion 16b.

  In this way, according to the present embodiment, even if the position of the valve portion 16b changes, the opening area in the third region is zero, so that the opening area changes with respect to the stroke change of the valve portion 16b. Less (zero).

(Effect of each embodiment)
In recent years, there has been a demand for further reduction in fuel consumption of automobiles for the purpose of improving environmental performance. In order to reduce fuel consumption, in addition to improving engine efficiency, it is necessary to reduce the load on equipment that uses the engine as a drive source. As one way of reducing the load on equipment using an engine as a drive source, the present embodiment focuses on improving the pump device of the power steering device.

  In this embodiment, in the pump device, the pump housing 11 includes a pump element accommodating space 11a, a suction passage 30, a discharge passage 25, a suction port 13a, and a discharge port 13b. The suction passage 30 is connected to the suction port 13a. The discharge passage 25 is connected to the discharge port 13b. The pump housing 11, the drive shaft 12 rotatably provided in the pump housing 11, and the drive shaft 12 are provided with a plurality of slits. A rotor 17, a plurality of vanes 18 movably provided in each of the plurality of slits, and a cam ring 14, which are formed in an annular shape and provided in the pump element accommodation space 11 a. A plurality of pump chambers 20 are formed together with the rotor 17 and the plurality of vanes 18, and the first fluid pressure chamber 2 is formed in the pump element accommodating space 11a. a and a second fluid pressure chamber 21b are formed, and the suction port 13a is open to a suction region of the plurality of pump chambers 20 where the volume of the pump chamber 20 increases with the rotation of the drive shaft 12, The discharge port 13b opens to a discharge region in which the volume of the pump chamber 20 decreases with the rotation of the drive shaft 12 among the plurality of pump chambers 20, and the first fluid pressure chamber 21a is the rotation axis of the drive shaft 12. In the radial direction of the cam ring 14, and is provided in a portion where the volume decreases as the amount of eccentricity between the rotation axis of the drive shaft 12 and the center of the inner peripheral edge of the cam ring 14 increases. The second fluid pressure chamber 21b is a space provided on the radially outer side of the cam ring 14 in the radial direction of the rotational axis of the drive shaft 12, and includes the rotational axis of the drive shaft 12 and the center of the inner peripheral edge of the cam ring 14. Bias The cam ring 14 is provided in a portion where the volume increases as the amount increases, and the cam ring 14 is movable in the pump element accommodating space 11a based on the pressure difference between the first fluid pressure chamber 21a and the second fluid pressure chamber 21b. The cam ring 14 and the control valve 15 provided are provided with a spool valve 15a, a first pressure chamber 15b, and a second pressure chamber 15c. The first pressure chamber 15b is arranged in the moving direction of the spool valve 15a. The hydraulic fluid that is provided on one side of the spool valve 15a and that is discharged from the discharge port 13b is introduced, and the second pressure chamber 15c is located between the first pressure chamber 15b and the spool valve 15a in the moving direction of the spool valve 15a. A hydraulic fluid downstream is introduced in the direction of the flow of hydraulic fluid from the variable orifice portion 27 provided on the opposite side and provided in the discharge passage 25, and the spool valve 15 a is controlled based on the difference between the hydraulic pressure of the hydraulic fluid in the first pressure chamber 15b and the hydraulic pressure of the hydraulic fluid in the second pressure chamber 15c, and can control the hydraulic pressure of the hydraulic fluid in the first fluid pressure chamber 21a. The control valve 15 and the solenoid valve 16 include a solenoid part 16a and a valve part 16b. The solenoid part 16a is driven and controlled in accordance with a change in the energization amount, and the position of the valve part 16b in the variable orifice part 27 The valve portion 16b is provided in the variable orifice portion 27, and according to the change in the position in the variable orifice portion 27, the flow path area is minimized according to the position in the variable orifice portion 27. Region, the second region in which the flow path area changes according to the change in the position of the valve portion 16b, and the position of the valve portion 16b is located between the first region and the second region. Variation of the flow area due to the change in the position of the cage the valve portion 16b has a a valve portion 16b having a third area smaller than the second region.

  According to the present embodiment, the first region of the valve portion 16b has a minimum discharge amount, the third region has a small discharge amount, and the flow rate recovery (increase) responsiveness is secured. It is possible to realize three states, that is, a state where a necessary flow rate is ensured in this area.

  Further, in the present embodiment, in the above, the valve portion 16b includes a variable orifice passage portion 77 that is a through-hole through which the hydraulic fluid passes, and the variable orifice passage portion 77 includes the first region, the second region, Alternatively, provided in the third region, the opening amount of the variable orifice passage portion 77 with respect to the discharge passage 25 (merging portion 78) can be changed in accordance with the change in the position of the valve portion 16b in the variable orifice portion 27. The variable orifice passage portion 77 in the third region has an amount of change in the overlap amount between the variable orifice passage portion 77 and the discharge passage 25 (merging portion 78) accompanying a change in the position of the valve portion 16b in the variable orifice portion 27. It is smaller than the second region.

  According to the present embodiment, by changing the overlap amount of the variable orifice passage portion 77 with respect to the stroke change of the valve portion 16b, the responsiveness of returning the flow rate while keeping the flow rate small while moving the valve portion 16b. Can get a good condition.

  Further, in this embodiment, in the above, the variable orifice passage portion 77 in the third region has a smaller width in the direction perpendicular to the moving direction of the valve portion 16b than the variable orifice passage portion 77 in the second region. .

  According to the present embodiment, since the width in the direction perpendicular to the moving direction of the valve portion 16b, that is, the opening width is small in the third region, the change in the opening area with respect to the change in the stroke of the valve portion 16b is reduced. While moving the valve portion 16b, it is possible to obtain a state in which the flow rate recovery response is good while the flow rate is small.

  In the present embodiment, in the above description, the variable orifice passage portion 77 in the third region has a width in a direction perpendicular to the moving direction of the valve portion 16b set to zero.

  According to the present embodiment, even if the position of the valve portion 16b changes, the opening area in the third region is zero, so that the change in the opening area with respect to the change in the stroke of the valve portion 16b is reduced (zero). be able to.

  In this embodiment, the valve portion 16b includes a fixed orifice passage portion (variable orifice passage portion 77b) and a variable orifice passage portion 77a, which are through-holes through which the hydraulic fluid passes, and includes a fixed orifice passage portion (variable orifice passage portion). The passage portion 77b) is provided in the first region, and is always open to the discharge passage 25 (merging portion 78). The variable orifice passage portion 77a includes the second region and the third region. Are provided in the combined area.

  According to the present embodiment, by setting the first region as a so-called fixed orifice, it is possible to ensure the minimum flow rate regardless of the change in the position of the valve portion 16b.

  In this embodiment, the variable orifice passage portion 77a includes a first opening and a second opening, and the fixed orifice passage portion (variable orifice passage portion 77b) is the first opening portion. Or it made smaller than the opening area of the said 2nd opening part.

  According to the present embodiment, by sufficiently reducing the opening area of the fixed orifice passage portion (variable orifice passage portion 77b), the discharge amount in the state where the discharge amount is minimum can be sufficiently reduced.

  In this embodiment, the valve portion is the spool valve 16c in the above.

  According to this embodiment, it becomes easy to adjust the opening area with respect to the stroke of the solenoid valve 16.

  In the present embodiment, the solenoid valve 16 includes the spring 76, and the solenoid portion 16a biases the valve portion 16b against the biasing force of the spring 76, whereby the valve portion 16b in the variable orifice portion 27 is provided. The spring 76 has a non-linear characteristic.

  According to the present embodiment, since the amount of change in the stroke of the valve portion 16b with respect to the thrust change in the solenoid portion 16a can be made a non-linear characteristic, the thrust change in the solenoid portion 16a in the second region and the third region. A difference can be provided in the change of the opening area with respect to.

  Further, in the present embodiment, in the above description, the cam ring 14 has the maximum rotational axis of the drive shaft 12 and the center of the inner peripheral edge of the cam ring 14 relative to the position where the rotational axis of the drive shaft 12 and the center of the inner peripheral edge of the cam ring 14 coincide. When the valve portion 16b is open to the discharge passage 25 (merging portion 78) only in the first region, the negative eccentric state is obtained. It was made to become.

  According to the present embodiment, when the cam ring 14 is moved even in the first region, the change in the discharge amount accompanying the change in the eccentric amount of the cam ring 14 can be suppressed by using the minus eccentric state.

  In this embodiment, the solenoid valve 16 is configured such that the second region opens to the discharge passage 25 (merging portion 78) when the energization amount to the solenoid portion 16a is zero.

  According to the present embodiment, the solenoid valve 16 is a so-called normally open valve, so that a necessary flow rate can be ensured even when the energization is defective.

  In this embodiment, in the above, the pump device can supply hydraulic fluid to a power steering device mounted on a vehicle, and the power steering device has a pair of hydraulic fluid chambers separated by pistons. A steering force can be applied to the steered wheels by selecting and supplying the hydraulic fluid to the pair of hydraulic fluid chambers, and the solenoid valve 16 is driven and controlled by a controller (ECU 40). Can receive signals relating to the driving state of the vehicle.

  According to the present embodiment, the power steering apparatus can reduce the pump driving load and the fuel (electric power) consumption by minimizing the discharge amount when the steering assistance is not necessary. Accordingly, by selecting and using the first, second, or third region in accordance with the driving state or steering state of the vehicle, it is possible to ensure both steering response and energy saving.

  Further, in the present embodiment, in the above, the controller (ECU 40) is configured such that only the first region is in the discharge passage 25 (merging portion 78) based on a signal related to the foot brake, side brake, shift position, or vehicle speed of the vehicle. The solenoid valve 16 is driven and controlled to open.

  According to this embodiment, since the necessity of steering assistance is low while the vehicle is stopped, the discharge amount of the pump can be minimized. This determination while the vehicle is stopped can be determined based on the foot brake, side brake, shift position, and vehicle speed.

  Further, in the present embodiment, in the above, the controller (ECU 40) drives and controls the solenoid valve 16 so that only the first region opens in the discharge passage when the steered shaft is at the stroke end. did.

  According to the present embodiment, when the steered shaft is in the vicinity of the stroke end, the necessity for assisting steering is low, so that the energy saving effect can be improved by minimizing the pump discharge amount.

  Further, in the present embodiment, in the above, the controller (ECU 40) drives and controls the solenoid valve 16 so that the third region opens in the discharge passage when the steering speed is not less than a predetermined vehicle speed. I made it.

  In this embodiment, when the steering operation is not performed, a large discharge flow rate is not required, and therefore, the discharge flow rate is set to be smaller than the state in which the second region is used. On the other hand, since the vehicle speed is equal to or higher than the predetermined vehicle speed, the steering operation is performed, and the necessary discharge flow rate may increase suddenly. Thus, in this embodiment, the steering flow response is ensured by increasing the discharge flow rate as compared with the state in which only the first region is used, and by setting the position of the valve unit to be close to the second region. Can do.

  Further, in the present embodiment, in the above, the controller (ECU 40) drives and controls the solenoid valve 16 in accordance with a change in the load applied to the steered shaft of the steered wheels.

According to the present embodiment, when the load applied to the steering shaft (for example, the front turning shaft) of the steered wheel is large (for example, when the load is large), the steering load increases. Therefore, the steering is increased by increasing the discharge flow rate. The load can be reduced. On the other hand, when the load applied to the steering shaft is small, the energy saving effect can be improved by reducing the discharge flow rate. Thus, by controlling the solenoid in accordance with the change in the load applied to the steering shaft, it is possible to achieve both reduction of the steering load and improvement of the energy saving effect.
(Aspect of pump device based on embodiment)
As a pump apparatus based on the embodiments described above, for example, the following modes can be considered.

In the pump device, the pump housing includes a pump element accommodation space, a suction passage, a discharge passage, a suction port, and a discharge port, and the suction passage is connected to the suction port, and the discharge passage is Connected to the discharge port,
The pump housing, a drive shaft rotatably provided on the pump housing, a rotor provided on the drive shaft and having a plurality of slits, and a plurality of movably provided in each of the plurality of slits The vane and the cam ring are formed in an annular shape and are provided in the pump element accommodation space, and form a plurality of pump chambers together with the rotor and the plurality of vanes, and the pump element accommodation space. The first fluid pressure chamber and the second fluid pressure chamber are formed, and the suction port is opened to a suction region in which the volume of the pump chamber increases with rotation of the drive shaft among the plurality of pump chambers. The discharge port is open to a discharge region in which the volume of the pump chamber decreases with rotation of the drive shaft among the plurality of pump chambers, and the first fluid pressure chamber is The space provided on the radially outer side of the cam ring in the radial direction of the rotational axis of the drive shaft, and the volume increases as the amount of eccentricity between the rotational axis of the drive shaft and the center of the inner peripheral edge of the cam ring increases. The second fluid pressure chamber is a space provided on the radially outer side of the cam ring in the radial direction of the rotational axis of the drive shaft, and the second fluid pressure chamber is a space provided on the rotational axis of the drive shaft. The cam ring is provided in a portion where the volume increases as the amount of eccentricity with the center of the inner peripheral edge of the cam ring increases, and the cam ring is based on a pressure difference between the first fluid pressure chamber and the second fluid pressure chamber. The cam ring, a control valve, which is movably provided in the element accommodating space, includes a spool valve, a first pressure chamber, and a second pressure chamber, and the first pressure chamber is the spool. Valve movement The hydraulic fluid is provided on one side of the spool valve in the direction and introduced from the discharge port, and the second pressure chamber has the first pressure with respect to the spool valve in the movement direction of the spool valve. The hydraulic fluid is introduced downstream of the variable orifice provided in the discharge passage in the direction of the flow of hydraulic fluid, and the spool valve operates the first pressure chamber. The control valve and the solenoid valve are controlled based on a difference between a hydraulic pressure of the hydraulic fluid and a hydraulic pressure of the hydraulic fluid in the second pressure chamber, and can control the hydraulic pressure of the hydraulic fluid in the first fluid pressure chamber. A solenoid portion and a valve portion, and the solenoid portion is driven and controlled in accordance with a change in energization amount to change the position of the valve portion in the variable orifice portion. A first region that is provided in the variable orifice portion and has a minimum flow area according to a position in the variable orifice portion according to a change in the position in the variable orifice portion, and a change in the position of the valve portion. Accordingly, the second region in which the flow path area changes and the position of the valve portion are located between the first region and the second region, and the flow accompanying the change in the position of the valve portion. And the valve portion having a third region in which the amount of change in the road area is smaller than that of the second region.

  In a preferred aspect of the pump device, the valve portion includes a variable orifice passage portion that is a through-hole through which hydraulic fluid passes, and the variable orifice passage portion includes the first region, the second region, or the An opening amount of the variable orifice passage portion with respect to the discharge passage is changeable according to a change in a position of the valve portion in the variable orifice portion, and the opening amount of the variable orifice passage portion in the third region is changed. The variable orifice passage portion is configured such that the amount of change in the overlap amount between the variable orifice passage portion and the discharge passage due to the change in the position of the valve portion in the variable orifice portion is smaller than that in the second region. Yes.

  In a preferred aspect of the pump device, the variable orifice passage portion in the third region is smaller in width in the direction perpendicular to the moving direction of the valve portion than the variable orifice passage portion in the second region. I am doing so.

  In a preferred aspect of the pump device, the variable orifice passage portion in the third region has a width of zero in a direction perpendicular to the moving direction of the valve portion.

  In another preferred aspect, the valve portion includes a fixed orifice passage portion and a variable orifice passage portion which are through holes through which the hydraulic fluid passes, and the fixed orifice passage portion is provided in the first region, The discharge orifice is always open, and the variable orifice passage is provided in a region combining the second region and the third region.

  In a preferred aspect of the pump apparatus, the variable orifice passage portion includes a first opening and a second opening, and the fixed orifice passage includes the first opening or the second opening. It is made to become smaller than the opening area.

  In another preferred embodiment, the valve portion is a spool valve.

  In another preferred embodiment, the solenoid valve includes a spring, and the solenoid portion changes the position of the valve portion in the variable orifice portion by urging the valve portion against the urging force of the spring. The spring has a non-linear characteristic.

  In another preferred aspect, the cam ring has a maximum at which the rotation axis of the drive shaft and the center of the inner periphery of the cam ring are maximized with respect to a position where the rotation axis of the drive shaft and the center of the inner periphery of the cam ring coincide with each other. It can move to a negative eccentric state opposite to the eccentric state, and when the valve portion is open to the discharge passage only in the first region, the negative eccentric state is set.

  In another preferred embodiment, the solenoid valve is configured such that the second region opens to the discharge passage when the energization amount to the solenoid portion is zero.

  In another preferred aspect, the pump device can supply hydraulic fluid to a power steering device mounted on a vehicle, and the power steering device has a power cylinder having a pair of hydraulic fluid chambers separated by pistons. Then, the steering force can be applied to the steered wheels by selecting the pair of hydraulic fluid chambers and supplying the hydraulic fluid, and the solenoid valve is driven and controlled by a controller. The signal regarding can be received.

  In a preferred aspect of the pump device, the controller controls the solenoid valve so that only the first region opens in the discharge passage based on a signal related to a foot brake, a side brake, a shift position, or a vehicle speed of the vehicle. The drive was controlled.

  In a preferred aspect of the pump device, the controller drives and controls the solenoid valve so that only the first region opens into the discharge passage when the steered shaft is at a stroke end.

  In a preferred aspect of the pump device, the controller drives and controls the solenoid valve so that a third region is opened in the discharge passage when the steering speed is not less than a predetermined vehicle speed.

  In a preferred aspect of the pump device, the controller drives and controls the solenoid valve according to a change in a load applied to a steered shaft of a steered wheel.

  In addition, this invention is not limited to the Example mentioned above, Various modifications are included. The above-described embodiments have been described in detail for easy understanding of the present invention, and are not necessarily limited to those having all the configurations described.

  10 .... Pump, 11 .... Pump housing, 11a ... Pump element receiving space, 12 .... Drive shaft, 13 .... Pump element, 13a ... Suction port, 13b ... Discharge port, 14 .... Cam ring, 15. Control valve, 15a, spool valve, 15b, first pressure chamber, 15c, second pressure chamber, 15d, coil spring, 16, solenoid valve, 16a, solenoid part, 16b, valve part, 16c ··· Spool valve, 16c1 ··· recess, 16c2 · · inclined portion, 17 · · rotor, 18 · · vane, 20 · · pump chamber, 21a · · first fluid pressure chamber, 21b · · second fluid pressure chamber 25 .... Discharge passage, 25a ... First discharge passage, 25b ... Second discharge passage, 25b1, ... Second discharge passage, 25b2 .... Second discharge passage, 26 .... Fixed orifice part, 27 .... Variable Oh Fist part 29 .. Fixed orifice part 30 .. Suction passage 33.. Steering angle sensor 34.. Vehicle speed sensor 36. Foot brake sensor 37 37 Side brake sensor 38 Shift position sensor 40..ECU, 76..Spring, 77..Variable orifice passage, 77a..Variable orifice passage, 77b..Variable orifice passage, 78..Merging portion.

Claims (15)

In the pump device,
A pump housing having a pump element accommodation space, a suction passage, a discharge passage, a suction port, and a discharge port;
The suction passage is connected to the suction port;
The discharge passage is connected to the discharge port;
The pump housing;
A drive shaft rotatably provided in the pump housing;
A rotor provided on the drive shaft and having a plurality of slits;
A plurality of vanes movably provided in each of the plurality of slits;
A cam ring, which is formed in an annular shape, is provided in the pump element accommodation space, forms a plurality of pump chambers together with the rotor and the plurality of vanes, and includes a first fluid in the pump element accommodation space. Forming a pressure chamber and a second fluid pressure chamber;
The suction port is open to a suction region in which the volume of the pump chamber increases with rotation of the drive shaft among the plurality of pump chambers,
The discharge port is open to a discharge region in which the volume of the pump chamber decreases with rotation of the drive shaft among the plurality of pump chambers,
The first fluid pressure chamber is a space provided on the radially outer side of the cam ring in the radial direction of the rotational axis of the drive shaft, and the eccentricity between the rotational axis of the drive shaft and the center of the inner peripheral edge of the cam ring. It is provided in the part where the volume decreases as the amount increases,
The second fluid pressure chamber is a space provided on the radially outer side of the cam ring in the radial direction of the rotation axis of the drive shaft, and the eccentricity between the rotation axis of the drive shaft and the center of the inner peripheral edge of the cam ring. It is provided in the part where the volume increases as the amount increases,
The cam ring is provided movably in the pump element accommodation space based on a pressure difference between the first fluid pressure chamber and the second fluid pressure chamber.
The cam ring;
A control valve comprising a spool valve, a first pressure chamber, and a second pressure chamber;
The first pressure chamber is provided on one side of the spool valve in the moving direction of the spool valve, and hydraulic fluid discharged from the discharge port is introduced,
The second pressure chamber is provided on the opposite side of the first pressure chamber with respect to the spool valve in the movement direction of the spool valve, and in the direction of the flow of hydraulic fluid than the variable orifice portion provided in the discharge passage. Downstream hydraulic fluid is introduced,
The spool valve is controlled based on the difference between the hydraulic pressure of the hydraulic fluid in the first pressure chamber and the hydraulic pressure of the hydraulic fluid in the second pressure chamber, and can control the hydraulic pressure of the hydraulic fluid in the first fluid pressure chamber. Is,
The control valve;
A solenoid valve comprising a solenoid part and a valve part;
The solenoid part is driven and controlled in accordance with a change in energization amount, and changes the position of the valve part in the variable orifice part,
The valve portion is provided in the variable orifice portion, and according to a change in the position of the variable orifice portion, a first region having a minimum flow path area and the flow in accordance with a change in the position of the valve portion. The change of the flow path area according to the change of the position of the second region where the path area changes, and the position of the valve portion is located between the first region and the second region. The valve portion having a third region whose amount is smaller than the second region;
A pump apparatus comprising: The pump device according to claim 1,
The valve portion includes a variable orifice passage portion that is a through hole through which hydraulic fluid passes,
The variable orifice passage portion is provided in the first region, the second region, or the third region, and is adapted to the discharge passage according to a change in the position of the valve portion in the variable orifice portion. The opening amount of the variable orifice passage portion can be changed,
The variable orifice passage portion in the third region has a change amount of an overlap amount between the variable orifice passage portion and the discharge passage in accordance with a change in the position of the valve portion in the variable orifice portion. Pump device characterized by being smaller than. The pump device according to claim 2,
The variable orifice passage portion in the third region is smaller in width in a direction perpendicular to the moving direction of the valve portion than the variable orifice passage portion in the second region. The pump device according to claim 3,
The variable orifice passage section in the third region has a width of zero in a direction perpendicular to the moving direction of the valve section. The pump device according to claim 1,
The valve portion includes a fixed orifice passage portion and a variable orifice passage portion which are through holes through which the hydraulic fluid passes,
The fixed orifice passage portion is provided in the first region, and is always open to the discharge passage,
The variable orifice passage portion is provided in a region where the second region and the third region are combined. The pump device according to claim 5,
The variable orifice passage includes a first opening and a second opening,
The pump device according to claim 1, wherein the fixed orifice passage portion is smaller than an opening area of the first opening or the second opening. The pump device according to claim 1,
The pump device, wherein the valve portion is a spool valve. The pump device according to claim 1,
The solenoid valve includes a spring,
The solenoid portion changes the position of the valve portion in the variable orifice portion by urging the valve portion against the urging force of the spring.
The pump device according to claim 1, wherein the spring has a non-linear characteristic. The pump device according to claim 1,
The cam ring is located on the opposite side of the maximum eccentric state where the rotation axis of the drive shaft and the center of the inner periphery of the cam ring are maximized with respect to the position where the rotation axis of the drive shaft and the center of the inner periphery of the cam ring coincide. A pump device that is movable to a negative eccentric state and is in the negative eccentric state when the valve portion is open to the discharge passage only in the first region. The pump device according to claim 1,
The pump device according to claim 1, wherein the solenoid valve has the second region opened to the discharge passage when the energization amount to the solenoid portion is zero. The pump device according to claim 1 can supply hydraulic fluid to a power steering device mounted on a vehicle.
The power steering device includes a power cylinder having a pair of hydraulic fluid chambers separated by pistons, and applies a steering force to the steered wheels by selecting the pair of hydraulic fluid chambers and supplying the hydraulic fluid. Is possible,
The solenoid valve is driven and controlled by a controller,
The pump device according to claim 1, wherein the controller is capable of receiving a signal related to a driving state of the vehicle. The pump device according to claim 11,
The controller drives and controls the solenoid valve so that only the first region opens in the discharge passage based on a signal related to a foot brake, a side brake, a shift position, or a vehicle speed of the vehicle. Pump device.   12. The pump device according to claim 11, wherein the controller drives and controls the solenoid valve so that only the first region opens in the discharge passage when the steered shaft is at a stroke end. Pump device to do. The pump device according to claim 11,
The controller is configured to control the drive of the solenoid valve so that a third region is opened to the discharge passage when the steering speed is not less than a predetermined vehicle speed. The pump device according to claim 11,
The pump device according to claim 1, wherein the controller drives and controls the solenoid valve in accordance with a change in a load applied to a steered shaft of a steered wheel.

Images