JP2012112529A - Differential pressure control valve - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a differential pressure control valve capable of controlling pressure with high accuracy by suppressing change in attraction force in association with the movement of a needle.SOLUTION: The differential pressure control valve 1 includes: a plunger 3 driven by a solenoid 2; a sleeve 4 for holding the plunger 3 slidably; a control piston 6 moved in association with movement of the plunger 3; and a guide 7 for holding the control piston 6 slidably. A magnet short structure is configured in such a manner that a columnar projection part 21 able to push-press a control piston 6 and a columnar recessed part 22 able to fit to the columnar projection part 21 are arranged. Further, the control piston 6 operated by the movement of the plunger 3 is equipped with a cone-shaped needle which closes a channel of a brake fluid by closing a hole arranged on a valve seat 12. As the shape of a needle tip part 6c of the control piston 6, the tip is cut to be flat (in a direction perpendicular to the axial direction). As a result, a pressure loss during opening the valve can be reduced.

Description

  The present invention relates to a differential pressure control valve that is disposed, for example, in a pipeline of a vehicle brake device of a vehicle and adjusts a differential pressure of a brake fluid pressure.

  2. Description of the Related Art Conventionally, in a vehicle brake device having a brake fluid pipe and a control valve that opens and closes the pipe, when a brake pedal is depressed, a predetermined braking force is obtained according to the depression amount and the depression force. It is set to be.

  Further, in this type of vehicle brake device, in order to increase the depressing force when depressing the brake pedal and increase the master cylinder pressure and thus the wheel cylinder pressure to improve the braking force, for example, a brake booster such as a vacuum booster is used. Force device is used.

  However, a vacuum booster or the like has a large physique and may be difficult to be mounted on the vehicle. Therefore, an alternative technique has been proposed. For example, in Japanese Patent Laid-Open No. 8-230634, when performing anti-skid control or traction control, the control of the return pump, the switching valve, and the suction valve is performed depending on the signal indicating the operation of the brake pedal. A technology that replaces the function of the device has been proposed.

  However, if a differential pressure control valve, which is an on-off type electromagnetic valve, is used as the switching valve in order to directly control the pressure on the wheel cylinder side, the suction force of the movable element is caused by the movement of the movable element in the electromagnetic valve. Since (the force for attracting the mover toward the stator) changes greatly, there is a problem that the differential pressure cannot be controlled stably.

  That is, in the conventional differential pressure control valve, when the mover moves, the distance (plunger stroke) between the mover and the stator changes as shown by the broken line in FIG. The amount of magnetic flux between them changes greatly, and the attractive force also changes greatly.

  Therefore, in the conventional differential pressure control valve, if the flow rate changes due to any cause, the position of the mover changes in order to keep the change in the flow rate, and thereby the suction force changes greatly. When this suction force changes, the differential pressure also changes as a result, so that there is a problem that the brake fluid pressure such as the wheel cylinder pressure cannot be controlled precisely.

  In addition, as shown by the broken line in FIG. 2A, the conventional differential pressure control valve does not generate a differential pressure when the applied current to the solenoid is equal to or less than a predetermined value. Therefore, there is a problem that precise brake fluid pressure cannot be controlled when the applied current is less than a predetermined value.

  This is because, as shown by the broken line in FIG. 2B, the relationship between the plunger stroke and the suction force is determined according to the applied current. For example, when the applied current is 2A, the plunger stroke is PS1, and the mover is movable. This is because, if it is in the range (actual use range), the mover suddenly moves so that the plunger stroke becomes PS0, and the suction force increases at once (for example, changes from 0 to K1).

  The present invention has been made in view of the above problems, and an object of the present invention is to provide a differential pressure control valve capable of performing highly accurate pressure control by suppressing a change in suction force associated with movement of a mover. .

  (1) In the first aspect of the present invention, in the differential pressure control valve that is disposed in the brake fluid conduit and adjusts the differential pressure of the brake fluid pressure, the differential pressure control valve is penetrated by the magnetic flux generated by the solenoid. A mover that is moved by the magnetic force, a stator that is disposed in a moving direction of the mover and through which the magnetic flux passes through the solenoid, a sleeve that restricts the maximum stroke of the mover, and the mover. By being pressed and seated on the valve seat, the valve element that regulates the minimum stroke value of the mover and the same current according to the magnitude of the current flowing to the solenoid (or the current on-off duty ratio). An electronic control unit that performs current control (or duty control) to generate a suction force of the solenoid, and changes in the amount of magnetic flux in the moving direction of the mover and the stator as the mover moves. And the maximum value and the minimum value of the stroke of the mover are set within a stroke range in which the attraction force of the solenoid is not changed by the magnetic short structure, and the electronic control device Is obtained from the balance between the force acting on the valve body due to the differential pressure of the brake fluid pressure and the suction force of the solenoid, and the current value ( Or a current value to be energized to the solenoid (or a duty ratio of the current) based on a proportional relationship with the duty ratio of the current) A conical needle for closing the brake fluid flow path by closing a hole provided in the seat is provided.

  The differential pressure control valve according to the invention can generate a suction force for moving the mover by flowing a solenoid current, and can set a differential pressure of the brake fluid pressure according to the suction force. In this type of differential pressure control valve, when the distance between the mover and the stator changes, the amount of magnetic flux in the moving direction of the mover changes. For example, when the mover and the stator approach each other, the amount of magnetic flux increases and attracts. Although the force increases, if the suction force changes in this way, a stable differential pressure cannot be generated, and the differential pressure cannot be adjusted when a low current is applied. Therefore, in the above invention, a magnetic short structure is provided in which the amount of magnetic flux in the moving direction is suppressed even when the mover moves.

  Thereby, for example, as shown by the solid line in FIG. 1, a flat characteristic is realized in the actual use region of the mover (plunger) so that the suction force does not change so much. In other words, when the distance between the mover and the stator is large, the suction force is not reduced, and even when the distance is small, the suction force is not increased rapidly (only one of them is Even if it is close to flat, it is effective in the corresponding area).

  As a result, the suction force can be prevented from changing greatly depending on the distance between the mover and the stator, so that a stable differential pressure can be generated regardless of the position of the mover, and precise brake control is realized. Can do.

  Further, in the above invention, for example, as shown by the solid line in FIG. 2, even if the current value (or current duty ratio) applied to the solenoid is changed, each current value (or current duty ratio) is movable. The suction force does not change much in the actual use area of the child. Therefore, the attractive force can be set according to the value (or duty ratio) of the applied current, so that there is an effect that the differential pressure can be precisely adjusted, particularly when the applied current is small.

  Furthermore, the above invention comprises a sleeve that regulates the maximum value of the stroke of the mover, and a valve body that regulates the minimum value of the mover by being pressed by the mover and seated on the valve seat. The maximum value and the minimum value of the child stroke are set within a stroke range in which the attraction force of the solenoid is not changed by the magnetic short structure.

  Therefore, even when the solenoid is energized and the mover is moved, the suction force does not change, so that the differential pressure control valve has a current value (or current) according to the current value (or current duty ratio) applied to the solenoid. The differential pressure can be precisely set from a low value to a high value.

  In the above-described invention, the electronic control device obtains a current flowing through the solenoid and the differential pressure of the brake fluid pressure, which is obtained from a balance between the force acting on the valve body due to the differential pressure of the brake fluid pressure and the suction force of the solenoid. A current value (or current duty ratio) to be energized to the solenoid is calculated based on a proportional relationship with the value (or current duty ratio).

  In the above invention, by increasing or decreasing the current value (or current duty ratio) energizing the solenoid, the suction force can be changed as shown by the solid line in FIG. Can be adjusted to.

Moreover, in the said invention, the valve body act | operated by operation | movement of a needle | mover is equipped with the conical needle | hook which closes the flow path of brake fluid by closing the hole provided in the valve seat.
(2) In invention of Claim 2, the pressure loss reduction structure which reduces the pressure loss at the time of valve opening is provided in the needle front-end | tip part of the valve body act | operated by operation | movement of a needle | mover.

For example, as shown in FIG. 26A, in the case of the shape of the needle tip of the conventional valve body having a sharp point and a sharp tip, pressure loss is likely to occur when the valve is opened.
Therefore, in the present invention, since the pressure loss reducing structure is provided at the needle tip portion, the pressure loss at the time of valve opening can be reduced. By reducing this pressure loss, there is an advantage that response delay at the time of brake release can be improved.

(3) In invention of Claim 3, a pressure-loss reduction structure is a structure which makes a needle front-end | tip part an obtuse angle.
The present invention exemplifies the invention of claim 2 and adopts a configuration in which the tip end portion of the needle is obtuse as shown in FIG. Thereby, the pressure loss at the time of valve opening can be reduced.

(4) In invention of Claim 4, a pressure loss reduction structure is the structure which cuts a needle front-end | tip part flatly.
The present invention exemplifies the invention of claim 2, and as a pressure loss reducing structure, for example, as shown in FIG. 26B, the needle tip is cut flat (for example, perpendicular to the valve body and the axial direction). The structure which cuts) is adopted. Thereby, the pressure loss at the time of valve opening can be reduced.

(5) In the invention of claim 5, the pressure loss reducing structure is configured to smoothly curve the tip of the needle.
The present invention exemplifies the invention of claim 2 and adopts, as the pressure loss reduction structure, a structure in which the needle tip portion is smoothly curved (and thus becomes obtuse) as shown in FIG. 26 (d), for example. is doing. Thereby, the pressure loss at the time of valve opening can be reduced.

It is explanatory drawing which shows the relationship between attraction | suction force and plunger stroke, comparing this invention with a prior art example. The present invention is compared with a conventional example, and (a) is an explanatory diagram showing the relationship between the generated differential pressure and the applied current, and (b) is an explanatory diagram showing the relationship between the suction force and the plunger stroke. It is explanatory drawing which shows the structure of the differential pressure control valve of Example 1. FIG. It is explanatory drawing which expands and shows the magnetic short structure of the differential pressure control valve of Example 1. FIG. It is a schematic block diagram which shows the brake device for vehicles in which the differential pressure control valve of Example 1 is used. It is explanatory drawing which shows the relationship between the W / C pressure by Example 1, and M / C pressure. It is explanatory drawing which shows the change of the magnetic flux in the magnetic short structure of the differential pressure control valve of Example 1. FIG. 3 is a flowchart illustrating pressure amplification assist brake control according to the first embodiment. It is a flowchart which shows the control at the time of failure of the brake booster of Example 1. FIG. It is a schematic block diagram which shows the brake device for vehicles of Example 2. FIG. 6 is a flowchart illustrating traction control according to the second embodiment. It is a schematic block diagram which shows the vehicle brake device of Example 3. It is explanatory drawing which shows the relationship between the W / C pressure and M / C pressure by the side of the rear wheel by Example 3. FIG. 10 is a flowchart illustrating control of braking force distribution according to a third embodiment. Another embodiment is shown, (a) is an explanatory view showing the magnetic short structure of the fourth embodiment, (b) is an explanatory view showing the magnetic short structure of the fifth embodiment, and (c) is a magnetic short structure of the sixth embodiment. (D) It is explanatory drawing which shows the magnetic short structure of Example 7. FIG. 10 is an explanatory view showing a magnetic short structure of Example 8. FIG. It is explanatory drawing which shows the change of the magnetic flux in the magnetic short structure of Example 8. It is explanatory drawing which shows the differential pressure control valve of Example 9. FIG. FIG. 10 is an explanatory view showing a differential pressure control valve of Example 10. It is explanatory drawing which shows the differential pressure control valve of Example 11. FIG. FIG. 20 is an explanatory view showing a differential pressure control valve of Example 12. It is explanatory drawing which shows the differential pressure control valve of Example 13. It is explanatory drawing which shows the characteristic of the target differential pressure control valve. It is explanatory drawing which shows the differential pressure control valve of Example 14. FIG. 20 is an explanatory view showing a differential pressure control valve of Example 15. The needle tip portion is shown, (a) is an explanatory view of a conventional needle tip portion, (b) is an explanatory view of a needle tip portion of Example 16, (c) is an explanatory view of a needle tip portion of another embodiment, (D) is the needle tip of yet another embodiment. It is explanatory drawing which shows the test | inspection method of the differential pressure control valve of Example 17. FIG. It is explanatory drawing which shows the test | inspection method of the differential pressure control valve of Example 17. FIG. FIG. 20 is an explanatory diagram illustrating a method for adjusting a differential pressure control valve according to an eighteenth embodiment. It is explanatory drawing which shows the structure for changing a suction-force characteristic. FIG. 20 is an explanatory view showing a differential pressure control valve of Example 19. It is explanatory drawing which shows the adjustment shim vicinity used for the differential pressure control valve of Example 19 from the control piston side. FIG. 20 is an explanatory view showing a differential pressure control valve of Example 20. 20 shows a ratchet mechanism of Example 20, wherein (a) is a side view showing the ratchet, (b) is a bottom view showing the ratchet, and (c) is a 360 ° unfolded state to show the engagement relationship between the ratchet and the valve seat. It is explanatory drawing which shows. It is explanatory drawing which shows the adjustment method of the differential pressure control valve of Example 21. FIG. 25 is an explanatory view showing a differential pressure control valve of Example 23. FIG. 20 is an explanatory diagram illustrating a method for adjusting a differential pressure control valve according to a twenty-fourth embodiment.

Hereinafter, preferred embodiments of a differential pressure control valve and a vehicle brake device according to the present invention will be described in detail with reference to the drawings by way of examples (examples).
Example 1
a) First, the differential pressure control valve will be described.

  As shown in FIG. 3, the differential pressure control valve 1 of this embodiment is an electromagnetic valve that is driven by energizing a solenoid 2. The differential pressure control valve 1 includes a mover (plunger) 3 driven by a solenoid 2, a sleeve 4 that covers the outer periphery of the plunger 3 and slidably holds the plunger 3, and is pressed as the plunger 3 moves. A rod-like control piston 6 that moves in this manner and a guide 7 that covers the outer periphery of the control piston 6 and slidably holds the control piston 6 are provided.

  Among them, the solenoid 2 changes the amount of magnetic flux (flux amount) indicated by a broken line in the figure in proportion to the magnitude of the applied current, and the yoke 8 through which the magnetic flux passes so as to cover the outer periphery thereof. Is arranged. The sleeve 4 is made of non-magnetic stainless steel, and the upper end side in FIG. 3 is closed and the lower end side is outside the upper end of the guide 7 in order to restrict the movement of the plunger 3 in the arrow B direction to a predetermined amount. It is fitted.

  The control piston 6 is a body valve (valve body) made of non-magnetic stainless steel, and has a conical shape positioned on the distal end side of the large diameter portion 6a on the upper end side, the small diameter portion 6b on the lower end side, and the small diameter portion 6b. It is comprised from the needle front-end | tip part 6c. And this needle front-end | tip part 6c closes the communicating hole 13 (it is a valve part orifice) of the cylindrical valve seat 12 internally fitted in the hollow part 11 of the guide 7, and closes the flow path of a brake fluid. It is configured. The control piston 6 is biased by a spring 14 in the direction of arrow B (that is, the valve opening direction).

  The guide 7 is a stator on the side where the plunger 3 is sucked. The hollow portion 11 on which the control piston 6 slides communicates with the pump 36 side and the wheel cylinder 32 side shown in FIG. 5 through the tip communication hole 16 at the lower end thereof, and with the master cylinder 31 through the side surface communication hole 17. (See FIG. 5).

  In particular, in this embodiment, as shown in FIG. 4, the magnetic short structure which is the main part of the plunger 3 is provided with a columnar convex portion 21 capable of pressing the control piston 6 on the lower end surface 3 a of the plunger 3 and a guide. A concave portion 22 is provided in a cylindrical shape on the upper end surface 7 a of the cylinder 7 so as to be fitted to the convex portion 21.

  As shown in FIG. 4 (a), when the solenoid 2 is not energized, the lower end surface 3a of the plunger 3 and the upper end surface 7a of the guide 7 are set so as to maintain a constant interval. The convex portion 21 and the concave portion 22 slightly overlap each other on the parallel side surfaces. In addition, the space | interval of the lower end surface 3a and the upper end surface 7a is the maximum value (PS1) of the plunger stroke called what is called a gap.

  As shown in FIG. 4B, when the solenoid 2 is energized, the plunger 3 moves downward (arrow A direction; valve closing direction) to move the control piston 6 in the same direction. As a result, the piston 6 is seated on the valve seat 12, and the movement of the plunger 3 in the direction of arrow A is set to stop. By this seating, the interval between the lower end surface 3a and the upper end surface 7a becomes the minimum value (PS0) of the plunger stroke.

b) Next, a vehicle brake device in which the differential pressure control valve 1 having the above-described configuration is used will be described. Here, the configuration of one of the four wheels will be described as an example.
As shown in FIG. 5, the differential pressure control valve 1 of this embodiment is arranged in a pipe line 33 connecting a master cylinder (M / C) 31 and a wheel cylinder (W / C) 32, and this difference A pump 36 is disposed in the pipeline 34 that bypasses the pressure control valve 1 such that the suction side is the master cylinder 31 side and the discharge side is the wheel cylinder 32 side. A brake pedal 38 is connected to the master cylinder 31 via a brake booster 37 that boosts the pedaling force.

  Therefore, this vehicle brake device increases the wheel cylinder pressure (W / C pressure) from the master cylinder pressure (M / C pressure) by operating the pump 36 with the differential pressure control valve 1 operated. Can do. Although not shown, this vehicle brake device detects various sensors, for example, a wheel speed sensor that detects a wheel speed, a vehicle speed sensor that detects a vehicle speed, and a master cylinder pressure by an electronic control unit (ECU). Master cylinder pressure sensor, wheel cylinder pressure sensor for detecting wheel cylinder pressure, negative pressure sensor for detecting negative pressure introduced into the brake booster 37, treading force sensor for detecting treading force of the brake pedal 38, load of loaded cargo, etc. On the basis of a signal from a load sensor or the like that detects this, various actuators such as a solenoid valve and a pump 36 are driven to perform various controls described later.

c) Next, the basic operation of the differential pressure control valve 1 will be described.
(1) When the wheel cylinder pressure is increased over the master cylinder pressure, that is, when a differential pressure is generated between the master cylinder pressure and the wheel cylinder pressure, the pump 36 is operated as shown in FIG. Thus, the brake fluid is sent from the master cylinder 31 side to the wheel cylinder 32 side, and the solenoid 2 of the differential pressure control valve 1 is energized to operate the differential pressure control valve 1.

  By energizing the solenoid 2, the plunger 3 moves downward (in the direction of arrow A) in FIG. 5 against the urging force of the spring 14 by electromagnetic force (attraction force F), and once sits on the valve seat 12. When this is seated, the following equation (1) is established.

A (P2-P1) = F (1)
However, A: pressure receiving area P2 of the communication hole of the valve seat; wheel cylinder pressure P1; master cylinder pressure. Accordingly, by energizing the solenoid 2 and operating the pump 36, the wheel cylinder pressure is expressed by the following equation (2). (P2) can be increased.

P2 = P1 + F / A (2)
Further, as will be described in detail later, an attractive force is generated in proportion to the current value (I) energizing the solenoid 2. Therefore, from the equation (2), as shown in FIG. 6, the relationship between the master cylinder pressure and the wheel cylinder pressure can be adjusted by changing the current value to change the suction force.

  (2) Next, the reason why the attractive force is generated in proportion to the current value energized to the solenoid 2 will be described. Usually, the magnetic flux passing through the plunger 3 and the guide 7 changes as the plunger 3 moves. That is, immediately after the solenoid 2 is energized, the plunger 3 and the guide 7 are slightly separated from each other, so that the amount of magnetic flux penetrating the plunger 3 and the guide 7 is small. Thereafter, as the plunger 3 approaches the guide 7, The amount of magnetic flux passing through the plunger 3 and the guide 7 increases.

  However, in the case of the present embodiment, the plunger 3 and the guide 7 are provided with the convex portion 21 and the concave portion 22 that are fitted to each other. The overlapping area on the side increases. Therefore, as shown in FIGS. 7A to 7C, even when the plunger 3 approaches the guide 7, the amount of magnetic flux in the movement direction of the plunger 3 (arrow A direction) does not increase. The amount of magnetic flux between the side surfaces of the convex portion 21 and the concave portion 22, that is, the direction perpendicular to the moving direction (radial direction) increases.

  That is, the change in the magnetic flux accompanying the movement of the plunger 3 only changes the amount of magnetic flux in the radial direction without changing the amount of magnetic flux in the moving direction. Therefore, even if the plunger 3 approaches the guide 7, the suction force for sucking the plunger 3 does not increase, and the plunger 3 is always sucked to the guide 7 side with a stable force. Therefore, since the control piston 6 is urged downward with a constant force, the differential pressure control valve 1 can always generate a constant differential pressure.

  This is shown by the solid line in FIG. 1, but in this embodiment, the suction force does not change even if the plunger stroke changes, so the relationship that always becomes a predetermined wheel cylinder pressure is maintained for a certain master cylinder pressure. be able to. Further, since the attraction force can be adjusted by the current value applied to the solenoid 2, by adjusting the current value, the master cylinder pressure and the wheel can be adjusted within a predetermined control range as shown in FIG. The relationship with the cylinder pressure can be arbitrarily set.

  In addition, as shown by the solid line in FIG. 2, the suction force can be set according to the current value even in the actual use region even when the current value is low. It can be carried out.

d) Next, control processing in the vehicle brake device using the above-described differential pressure control valve 1 will be described with reference to the flowcharts of FIGS.
(1) Here, the pressure amplification assist brake control (PAB) for increasing the braking force by increasing the wheel cylinder pressure from the master cylinder pressure will be described.

  First, at step 100 in FIG. 8, it is determined whether or not the brake pedal 38 is depressed and a stop switch (not shown) is turned on. If an affirmative determination is made here, the process proceeds to step 110, whereas if a negative determination is made, the present process is temporarily terminated.

  In step 110, it is determined whether or not a condition for performing the main pressure amplification assist brake control is satisfied. If an affirmative determination is made here, the process proceeds to step 120, whereas if a negative determination is made, the present process is temporarily terminated. Examples of the determination condition include whether or not the pedal stroke speed is a predetermined value (XAmm / sec) or more, or whether or not the pedaling force change speed is a predetermined value (XBkgf / sec) or more.

  In step 120, a pump motor (not shown) is turned on to operate the pump 36. In the subsequent step 130, the target hydraulic pressure is calculated from the pedal effort. In subsequent step 140, a current value to be applied to the differential pressure control valve 1 is calculated from the target hydraulic pressure. Alternatively, a current value (or duty ratio) corresponding to the current value is calculated.

  In the following step 150, the differential pressure control valve 1 is driven by controlling the current applied to the solenoid 2 based on the calculated current value (or duty ratio). In the following step 160, it is determined whether or not an end condition for ending this pressure amplification assist brake control is satisfied. Here, if a positive determination is made, the process proceeds to step 170, whereas if a negative determination is made, the process returns to step 130. Examples of the determination condition include whether or not the pedaling force decrease amount exceeds a predetermined value XC, or whether or not the pedaling force falls below a predetermined value XD.

  In step 170, the energization of the solenoid 2 is turned off (OFF), and the operation of the differential pressure control valve 1 is stopped. Thereby, the differential pressure control valve 1 is completely opened. In the subsequent step 180, the pump motor is turned off, the operation of the pump 36 is stopped, and the process is temporarily terminated.

Therefore, the wheel cylinder pressure can be made higher than the master cylinder pressure and the braking force can be increased by the control described above.
(2) Next, the control at the time of failure of the brake booster 37 that boosts the depression force of the brake pedal 38 will be described.

  First, the brake booster 37 is driven by the difference between the atmospheric chamber for introducing the atmosphere and the negative pressure chamber for introducing the negative pressure. In step 200, the pressure (internal pressure) of the negative pressure chamber is set to a predetermined value (− XEmmHg) is determined. If an affirmative determination is made here, it is determined that the required internal pressure has not been obtained, that is, the brake booster 37 has failed. On the other hand, if a negative determination is made, it is assumed that the brake booster 37 is normal, and the present process is temporarily terminated.

  In step 210, it is determined whether or not the brake pedal 38 is depressed and the stop switch is turned on. If an affirmative determination is made here, the process proceeds to step 220. If a negative determination is made, the present process is temporarily terminated. In step 220, the pump motor is turned on and the pump 36 is operated.

  In the following step 230, the target hydraulic pressure is calculated from the pedal effort. In subsequent step 240, a current value to be applied to the differential pressure control valve 1 is calculated from the target hydraulic pressure. Alternatively, the duty ratio of the current corresponding to the current value is calculated. In the subsequent step 250, the differential pressure control valve 1 is driven by controlling the current applied to the solenoid 2 based on the calculated current value (or duty ratio).

  In subsequent step ON 260, it is determined whether or not an end condition for ending the control when the brake booster 37 fails is satisfied. If an affirmative determination is made here, the process proceeds to step 270. If a negative determination is made, the process returns to step 220. Examples of the determination condition include whether or not the pedal effort is equal to or less than a predetermined value (XF).

  In step 270, the energization of the solenoid 2 is turned off, and the operation of the differential pressure control valve 1 is stopped. In the subsequent step 280, the pump motor is turned off, the operation of the pump 36 is stopped, and the process is temporarily terminated.

  Therefore, when the brake booster 37 fails due to the control described above, the braking force can be increased instead of the boosting action using the differential pressure of the brake booster 37 by increasing the wheel cylinder pressure above the master cylinder pressure. it can.

  (3) Further, using the above-described vehicle brake device, high G assist control, that is, control to perform a boosting action in a region where the pedal effort is high and the assist force of the brake booster 37 is used up. Can do.

  The high G assist control is substantially the same as the control shown in FIG. 8 and will not be described in detail. However, the control execution condition in step 110 and the control end condition in step 160 in FIG. 8 are different. That is, as the control execution condition, a condition such as whether the master cylinder pressure exceeds a predetermined value (XGbar) or whether the differential pressure in the brake booster 37 is lower than a predetermined value (XHmmHg) can be adopted.

  Further, as the control end condition, a condition such as whether the master cylinder pressure is lower than a predetermined value (XGbar) or whether the differential pressure in the brake booster 37 is higher than a predetermined value (XHmmHg) can be adopted. Thus, in this embodiment, since the plunger 3 and the guide 7 are provided with a magnetic short structure including the convex portion 21 and the concave portion 22 that are fitted to each other, the solenoid 3 is energized to move the plunger 3. But the suction does not change. Therefore, the differential pressure control valve 1 can accurately set the differential pressure from a low current value to a high value according to the current value applied to the solenoid 2.

  Further, since the differential pressure can be accurately set by the differential pressure control valve 1 in this way, in the vehicle brake device using the differential pressure control valve 1, the pressure amplification assist brake control and the brake booster 37 can be Control and high G assist control can be suitably performed.

(Example 2)
Next, the second embodiment will be described, but the description of the same parts as the first embodiment will be omitted or simplified. In this embodiment, the differential pressure control valve of the first embodiment is applied to another vehicle brake device.

  This vehicular brake device is a vehicular brake device that can perform turning control (VSC) for controlling vehicle behavior during turning in addition to well-known anti-skid control (ABS) and traction control (TRC). As shown in FIG. 10, this vehicle brake device has a tandem master cylinder 41, and a brake pedal 43 is connected to the master cylinder 41 via a brake booster 42.

  A master reservoir 46 is connected to the master cylinder 41, and a hydraulic control circuit 50 that is configured by two hydraulic systems of X piping (diagonal piping) and adjusts brake hydraulic pressure is connected. It is comprised from the 1st hydraulic piping 51a and the 2nd hydraulic piping 51b.

  In the hydraulic control circuit 50, a wheel cylinder 55 for the right front (FR) wheel and a wheel cylinder 56 for the left rear (RL) wheel are communicated with each other via the first hydraulic pipe 51a. Further, the wheel cylinder 57 of the right rear (RR) wheel and the wheel cylinder 58 of the left front (FL) wheel are communicated with each other via the second hydraulic pipe 51b.

  The first hydraulic pipe 51a includes a known pressure increase control valve 61 and pressure reduction control valve 65 for controlling the oil pressure of the wheel cylinder 55 for the FR wheel, and an increase for controlling the oil pressure of the wheel cylinder 56 for the RL wheel. A pressure control valve 62 and a pressure reduction control valve 66 are provided, and a pressure increase control valve 63 and a pressure reduction control valve 67 for controlling the hydraulic pressure of the wheel cylinder 57 of the RR wheel are provided in the second hydraulic pipe 51b. A pressure increase control valve 64 and a pressure reduction control valve 68 for controlling the hydraulic pressure of the wheel cylinder 58 are provided.

  Here, the first hydraulic pipe 51a will be described. A differential pressure control valve 71 similar to that of the first embodiment is disposed on the master cylinder 41 side from each pressure increase control valve 61, 62. When the differential pressure control valve 71 is used to drive the hydraulic pump 78 (with the SRC valve 74 open), the hydraulic pressure on the hydraulic path 75a side may be made higher by an arbitrary pressure than the master cylinder 41 side. It becomes possible.

  Further, the first hydraulic pipe 51a is provided with a reservoir 76 for temporarily storing brake oil discharged from the pressure reducing control valves 65 and 66, and a hydraulic pump 78 for pumping the brake oil to the hydraulic path 75a. ing. Note that a damper 87 for suppressing pulsation of the internal hydraulic pressure is provided in the brake oil discharge path from the hydraulic pump 78.

  The first hydraulic pipe 51a is provided with a hydraulic path 79a for supplying brake oil directly from the master cylinder 41 to the hydraulic pump 78 when the wheel cylinder pressure is increased. A cut valve (SRC valve) 74 that communicates and blocks the hydraulic path 79a is provided.

  On the other hand, similarly to the first hydraulic pipe 51a, the second hydraulic pipe 51b includes pressure increase control valves 63 and 64, pressure reduction control valves 67 and 68, a differential pressure control valve 72, a reservoir 77, a hydraulic pump 79, and a damper 88. The SRC valve 75 and the like are provided at the same location. The hydraulic pumps 78 and 79 are connected to an electric pump motor 81 and driven.

b) Next, the control processing in the present embodiment will be described based on the flowchart of FIG. Here, TRC control will be described.
In step 300 of FIG. 11, it is determined whether or not the brake switch 43 is depressed and a stop switch (not shown) is turned off. If an affirmative determination is made here, the process proceeds to step 310. If a negative determination is made, the present process is temporarily terminated.

  In step 310, it is determined whether or not the drive wheel wheel speed Va is greater than the rolling wheel wheel speed Vb, that is, whether or not the slip state is excessive. Here, if a positive determination is made, the process proceeds to step 320, whereas if a negative determination is made, the present process is once terminated. In step 320, the pump motor 81 is turned on to operate the pumps 78 and 79.

  In the subsequent step 330, a solenoid (not shown) is energized to drive the differential pressure control valves 71 and 72 and the SRC valves 74 and 75. In the subsequent step 340, the pressure increase control valves 61 to 64 and the pressure reduction control valves 65 to 68 are controlled to control the drive wheel wheel speed Va.

  In the following step 350, it is determined whether or not the drive wheel wheel speed Va is the same as the rolling wheel wheel speed Vb. If the determination is affirmative, the process proceeds to step 360. If the determination is negative, the process proceeds to step 340. In step 360, the pump motor 81 is turned off and the operations of the pumps 78 and 79 are stopped.

  In the subsequent step 370, the energization to the solenoid is turned off, the operations of the differential pressure control valves 71 and 72 and the SRC valves 74 and 75 are stopped, and the present process is temporarily terminated. In this way, in this embodiment, since the differential pressure control valves 71 and 72 are used in the above-described vehicle brake device, various kinds of control such as TRC control are performed by applying currents to be applied to the solenoids of the differential pressure control valves 71 and 72. By adjusting, it can be performed precisely.

(Example 3)
Next, the third embodiment will be described, but the description of the same parts as the first embodiment will be omitted or simplified. In the present embodiment, the brake fluid pressure on the front wheel side can be made higher than that on the rear wheel side by disposing the differential pressure control valve in the first embodiment in the vehicle brake device with the connecting direction reversed. .

  a) As shown in FIG. 12, in the vehicle brake device of this embodiment, the differential pressure control valve 91 is disposed in a conduit 94 that connects the master cylinder 92 and the wheel cylinder 93 on the rear wheel side. A wheel cylinder 94 on the front wheel side is connected to a pipeline 95 branched from the pipeline 94.

  In particular, in this embodiment, the leading end communication hole 97 (high pressure side) of the differential pressure control valve 91 is connected to the master cylinder 92 side and the front wheel wheel cylinder 94 side, while the side surface communication hole 98 ( The low pressure side) is connected to the wheel cylinder 93 side of the rear wheel.

b) Next, the operation of the vehicle brake device will be described.
When a solenoid (not shown) is energized and the differential pressure control valve 91 is operated, as shown in FIG. 13, the relationship between the master cylinder pressure and the wheel cylinder pressure on the rear wheel side according to the current value applied to the solenoid. Is set.

  Therefore, when the differential pressure control valve 91 is operated when the brake pedal 99 is depressed, the master cylinder pressure is directly transmitted to the wheel cylinder 94 on the front wheel side. However, as shown in FIG. The brake fluid pressure on the wheel side can be set arbitrarily low.

  For this reason, for example, during braking, the brake fluid pressure on the front wheel side can be increased from the rear wheel side, so that a so-called front wheel leading lock (when the wheel is locked) can be realized, and the locking vehicle can be braked more stably. Can do. Also, the difference between the brake fluid pressure on the rear wheel side and the front wheel side can be arbitrarily set depending on the magnitude of the current applied to the solenoid. By setting the differential pressure, the front / rear braking force distribution can be distributed such that the maximum braking force of the vehicle can be exerted, so that the stability of the vehicle during braking can be improved and the braking performance can be greatly improved.

c) Next, control processing in the vehicle brake device will be described based on the flowchart of FIG. In addition, this control is control which adjusts braking force distribution.
In step 400 of FIG. 14, for example, the loading amount of the vehicle is detected based on a signal from a load sensor.

  In the following step 410, it is determined whether or not a brake switch (not shown) is turned on. If an affirmative determination is made here, the process proceeds to step 420, whereas if a negative determination is made, this process is once terminated. In step 420, the pedaling force is detected based on the signal from the pedaling force sensor.

  In step 430, the target hydraulic pressure is calculated based on the pedal effort and the loaded state. In the following step 440, the current value (or its duty ratio) applied to the solenoid of the differential pressure control valve 91 is calculated based on the target hydraulic pressure and the pedal effort. In the subsequent step 450, the differential pressure control valve 91 is driven based on the calculated current value (or its duty ratio).

  In the subsequent step 460, it is determined whether or not the brake switch is on. If an affirmative determination is made here, the process returns to step 410, whereas if a negative determination is made, the process proceeds to step 470. In step 470, the solenoid is turned off, the driving of the differential pressure control valve 91 is stopped, and this process is temporarily terminated.

  That is, with such control, excellent braking performance can be exhibited with suitable braking force distribution. Further, by distributing the braking force in this way, a similar braking force can be realized with the same pedaling force even when the mounting load of the vehicle changes, for example.

Example 4
Next, the fourth embodiment will be described, but the description of the same parts as the first embodiment will be omitted or simplified. This embodiment differs from the first embodiment in the magnetic short structure of the differential pressure control valve.

  As shown in FIG. 15 (a), in the differential pressure control valve 101 of this embodiment, the convex portion 104 at the lower end of the plunger 103 fitted in the sleeve 102 is fitted in the plunger 103 and the guide 105. A convex taper portion 104a that is inclined from the moving direction of the control piston 106 (vertical direction in the figure) is formed. On the other hand, a concave taper portion 108a parallel to the convex taper portion 104a is formed in the concave portion 108 at the upper end of the guide 105. Thereby, the magnetic short structure shown in the figure is formed.

  Also in the differential pressure control valve 101 of the present embodiment, the amount of magnetic flux increases due to the downward movement of the plunger 103 in the figure, but the increase is between the convex taper portion 104a and the concave taper portion 108a. This is the amount of magnetic flux, and the amount of magnetic flux in the moving direction of the plunger 103 does not increase so much.

Accordingly, since the suction force does not increase due to the movement of the plunger 103, the differential pressure can be precisely set by the differential pressure control valve 101 as in the first embodiment.
(Example 5)
Next, the fifth embodiment will be described, but the description of the same parts as the first embodiment will be omitted or simplified.

  This embodiment differs from the first embodiment in the magnetic short structure of the differential pressure control valve. As shown in FIG. 15B, in the differential pressure control valve 111 of this embodiment, a cylindrical recess 114 is provided at the lower end of the plunger 113 fitted in the sleeve 112, and the upper end of the guide 115 is provided. Is provided with a cylindrical convex portion 116, and a control piston 117 is fitted into the guide 115 so as to penetrate the convex portion 116. Thereby, the magnetic short structure shown in the figure is formed.

  Also in the differential pressure control valve 111 of the present embodiment, the amount of magnetic flux increases due to the downward movement of the plunger 113 in the figure, but the increase is the amount of magnetic flux between the concave portion 114 and the convex portion 116. The amount of magnetic flux in the movement direction of the plunger 113 does not increase so much.

Therefore, since the suction force does not increase due to the movement of the plunger 113, the differential pressure can be set accurately by the differential pressure control valve 111 as in the first embodiment.
(Example 6)
Next, the sixth embodiment will be described, but the description of the same parts as the first embodiment will be omitted or simplified.

  This embodiment differs from the first embodiment in the magnetic short structure of the differential pressure control valve. As shown in FIG. 15C, in the differential pressure control valve 121 of this embodiment, the plunger 123 fitted in the sleeve 122 is urged upward in the figure by the spring 124. At the lower end, a columnar convex portion 125 having a conical tip is provided. On the other hand, the hollow portion 127 of the guide 126 is provided with a step 128 corresponding to the conical portion of the convex portion 125. Thereby, the magnetic short structure shown in the figure is formed.

  Also in the differential pressure control valve 121 of the present embodiment, the amount of magnetic flux increases due to the downward movement of the plunger 123 in the figure. The increase is the amount of magnetic flux between the tip portion of the convex portion 125 and the step 128. The amount of magnetic flux in the moving direction of the plunger 123 does not increase so much.

Therefore, since the suction force does not increase due to the movement of the plunger 123, the differential pressure can be precisely set by the differential pressure control valve 121 as in the first embodiment.
(Example 7)
Next, although Example 7 is demonstrated, description of the location similar to the said Example 1 is abbreviate | omitted or simplified.

  This embodiment differs from the first embodiment in the magnetic short structure of the differential pressure control valve. As shown in FIG. 15 (d), in the differential pressure control valve 131 of the present embodiment, the lower end of the plunger 133 fitted in the sleeve 132 is flat with no irregularities. On the other hand, the upper end of the guide 135 in which the control piston 134 is fitted is also flat.

  In particular, in the present embodiment, almost the entire sleeve 132 is a magnetic body, but a non-magnetic body is annularly arranged at a part of the lower end side. That is, when the plunger 133 moves upward, the position of the lower end of the plunger 133 and the position of the lower end of the non-magnetic material are set to substantially the same position. Thereby, the magnetic short structure shown in the figure is formed.

  Also in the differential pressure control valve 131 of the present embodiment, the amount of magnetic flux increases due to the downward movement of the plunger 133 in the figure. The increase is between the lower portion of the plunger 133 and the lower portion 132a of the magnetic body of the sleeve 132. The amount of magnetic flux in the movement direction of the plunger 133 does not increase so much.

Therefore, since the suction force does not increase due to the movement of the plunger 133, the differential pressure can be precisely set by the differential pressure control valve 131 as in the first embodiment.
(Example 8)
Next, Example 8 will be described, but the description of the same part as in Example 1 will be omitted or simplified.

  This embodiment differs from the first embodiment in the magnetic short structure of the differential pressure control valve. As shown in FIG. 16, in the differential pressure control valve 141 of the present embodiment, a solenoid 144 is disposed outside a sleeve 143 into which the plunger 142 is fitted, and a yoke 145 is provided so as to cover the outer periphery of the solenoid 144. Has been placed.

In particular, in the present embodiment, the upper portion of the yoke 145 and the upper position when the plunger 142 moves upward in the drawing are set to substantially the same position. Thereby, the magnetic short structure shown in the figure is formed. In the differential pressure control valve 141 of this embodiment, as shown in FIG. 17A, when the plunger is positioned in the upper part of the figure, the amount of magnetic flux passing through the yoke 145 and the plunger is large. As shown in FIG. 17B, when the plunger 142 moves downward in the figure, the amount of magnetic flux passing through the yoke 145 and the plunger 142 decreases.
Accordingly, the plunger 142 moves downward in FIG. 16, and normally, the amount of magnetic flux between the plunger 142 and the guide 146 should increase, but in this embodiment, as shown in FIG. 17 (b). Similarly, since the amount of magnetic flux in the magnetic short structure decreases, the total amount of magnetic flux indicated by the broken line in FIG. 16 does not change. Therefore, since the suction force does not increase, the differential pressure can be accurately set by the differential pressure control valve 141 as in the first embodiment.

Example 9
Next, Example 9 will be described, but description of the same parts as those in Example 1 will be omitted or simplified. In the present embodiment, an orifice is disposed as a pressure loss adding mechanism for adding pressure loss downstream of the differential pressure control valve.

a) First, the configuration of the differential pressure control valve of this embodiment will be described.
As shown in FIG. 18, the differential pressure control valve 151 of the present embodiment has a magnetic short structure similar to that of the first embodiment, and urges the plunger 152 that is a magnetic movable element and the plunger 152 by magnetic force. A solenoid 153, a control piston 154 that is a valve body, a guide 156 in which the control piston 154 is fitted, a valve seat 157 on which a needle tip 155 of the control piston 154 is seated, and the control piston 154 as an urging force of the solenoid 153 A spring (return spring) 158 or the like that urges upward in the figure against it is provided.

  Particularly in this embodiment, the plunger 152 and the control piston 154 are integrally connected by a coupling pin 159. Therefore, both 152,154 always move up and down integrally. In this embodiment, the pump side (and the wheel cylinder side) and the hollow portion 161 of the guide 156 in which the control piston 154 is disposed are connected by a main communication path 163 having an upstream orifice 162. Moreover, the wheel cylinder side and the master cylinder side are connected via a check valve 164.

  Further, the hollow portion 161 communicates with the master cylinder side via the downstream side communication passages 167 and 168, and the downstream side communication passages 167 and 168 have pressure loss on the downstream side of the differential pressure control valve 151. In order to apply (back pressure), downstream orifices 166 and 169 are provided.

  Filters 171 and 172 are disposed in the flow path leading to the main communication path 163 and the downstream communication paths 167 and 168, respectively. The guide 156 is fixed to the housing 174, and a frame 176 that forms part of a magnetic path is disposed between the housing 174 and the solenoid 153. Accordingly, the magnetic flux generated by the solenoid 153 forms a magnetic path that penetrates the yoke 177 and the frame body 176 as indicated by a broken line in FIG.

The solenoid 153 of the present embodiment is somewhat different from the cylindrical shape of the first embodiment, and is non-target on the left and right of the figure, but its function is almost the same.
b) Next, the effect of the differential pressure control valve 151 of the present embodiment will be described.

In the differential pressure control valve 151 of the present embodiment, downstream orifices 166 and 169 are provided on the downstream side (master cylinder side).
Therefore, for example, when using the above-described functions such as TRC, VSC, PAB, etc., when the brake pedal 43 is loosened, the differential pressure control valve 151 is used as a relief function (a function for releasing high pressure) of the pump upstream pressure. A slight back pressure is applied to the high-pressure wheel cylinder side by the downstream orifices 166 and 169 without immediately relieving the atmospheric pressure (master cylinder pressure).

  As a result, it is possible to suppress the occurrence of aeration, which is likely to occur when the brake fluid changes from high pressure to low pressure, so that the brake performance can be reliably maintained. In this embodiment, since the plunger 152 and the control piston 154 are integrated, the control piston 154 is unlikely to vibrate when the high-pressure wheel cylinder pressure is relieved. Therefore, also from this point, there is an advantage that aeration can be prevented.

(Example 10)
Next, although Example 10 is demonstrated, description of the same location as the said Example 9 is abbreviate | omitted or simplified. In the present embodiment, a one-way orifice provided with an auxiliary valve element is arranged as a pressure loss adding mechanism for adding pressure loss downstream of the differential pressure control valve.

  As shown in FIG. 19, the differential pressure control valve 181 of this embodiment includes a groove-shaped notch orifice 186, a spherical auxiliary, and a downstream communication passage 184 that communicates the hollow portion 183 of the guide 182 with the master cylinder side. A one-way orifice 189 including a valve body 187 and an auxiliary spring 188 is provided.

  The notch orifice 186 is normally open, and the auxiliary valve body 187 is always urged in the valve closing direction (arrow A) by the auxiliary spring 188. This auxiliary valve body 187 moves in the direction of arrow B when the differential pressure (large on the wheel cylinder side) increases, and opens the downstream communication path 184.

  The one-way orifice 189 opens the flow path largely when the auxiliary valve body 187 moves in the direction A in the figure when the brake pedal 43 is loosened when the differential pressure between the wheel cylinder side and the master cylinder side is large. The flow rate of the brake fluid can be ensured. As a result, it is possible to prevent a so-called dragging feeling (a phenomenon in which the braking force does not easily weaken even when the brake pedal 43 is released).

  On the other hand, when the differential pressure is small, the auxiliary valve body 187 does not move due to the urging force of the auxiliary spring 188, so that it functions as a normal orifice and can prevent aeration as in the ninth embodiment. . Further, since the amount of movement of the auxiliary valve body 187 (and hence its opening) varies depending on the magnitude of the differential pressure, in the case of a differential pressure where the auxiliary valve body 187 moves slightly, an orifice corresponding to the differential pressure is used. It has the function of.

In the present embodiment, since the unidirectional orifice 189 is provided, the inner diameter of the hollow portion 183 is reduced.
(Example 11)
Next, although Example 11 is demonstrated, description of the same location as the said Example 9 is abbreviate | omitted or simplified.

  In the present embodiment, a control sleeve provided with an orifice is disposed as a pressure loss applying mechanism for applying pressure loss on the tip side of the control piston. As shown in FIG. 20, in the differential pressure control valve 191 of this embodiment, a cylindrical control sleeve 193 is integrally fixed to the outer periphery on the front end side of the control piston 192, and this control sleeve 193 is controlled by the control sleeve 193. The periphery of the middle diameter portion 192a of the piston 192 and the spring 194 is covered. Further, a part of the tip of the control sleeve 193 is notched in a strip shape, and a plurality of notches 196 are provided.

  On the other hand, an annular convex portion 198 is provided on the upper surface of the valve seat 197 so as to cover the outer periphery of the tip of the control sleeve 193. The diameter of each member is set so that there is a slight gap between the inner peripheral surface of the annular convex portion 198 and the outer peripheral surface of the control sleeve 193.

  Therefore, when the control piston 192 is raised, the main communication passage 201 and the downstream communication passage 202 are arranged so that the flow rate of the brake fluid can be sufficiently secured through the notch 196 of the control sleeve 193. The flow path is open. On the other hand, when the control piston 192 is lowered, the gap between the inner peripheral surface of the annular convex portion 198 and the outer peripheral surface of the control sleeve 193 functions as an orifice. Back pressure can be applied on the downstream side of the pressure control valve 191. Thereby, generation | occurrence | production of aeration can be prevented.

  In particular, in this embodiment, when the brake pedal 43 is returned, the control piston 192 is raised, and the flow rate of the brake fluid can be secured. Therefore, there is an advantage that the drag phenomenon is small and the responsiveness is good. In this embodiment, the downstream orifice of the downstream communication path as in the ninth embodiment is not provided.

(Example 12)
Next, although Example 12 is demonstrated, description of the same location as the said Example 9 is abbreviate | omitted or simplified. In this embodiment, a separate control sleeve or the like is disposed on the tip side of the control piston as a pressure loss applying mechanism for applying pressure loss.

  As shown in FIG. 21, in the differential pressure control valve 211 of the present embodiment, a cylindrical control sleeve 213 separate from the control piston 212 is slidably disposed on the outer periphery on the front end side of the control piston 212. Has been. The control sleeve 213 is biased in the valve closing direction (downward in the drawing) by a spring (auxiliary spring) 214. Furthermore, a part of the tip of the control sleeve 213 is notched, and a notch 216 that functions as an orifice is provided.

  Further, a stopper 217 is disposed between the control sleeve 213 and the middle diameter portion 212 a of the control piston 212. The stopper 217 is urged upward in the figure by a spring (return spring) 218 and is seated on the step portion 212b, and restricts the downward movement of the control sleeve 213.

  On the other hand, on the upper surface of the valve seat 215, a convex portion 219 protruding in an annular shape is provided between the spring 218 and the tip of the control sleeve 213. The outer periphery of the convex portion 219 is inclined, and the lower end of the control sleeve 213 is seated on the outer periphery. In this embodiment, when the solenoid 221 is energized, the control piston 212 is lowered. At this time, the control sleeve 213 is also lowered by the biasing force of the spring 214. Before the valve closing of the control piston 212, the tip of the control sleeve 213 is seated on the outer periphery of the convex portion 219, and the function of the orifice is exhibited at the notch 216. As a result, a back pressure can be applied downstream of the differential pressure control valve 211, so that aeration can be prevented.

  Note that the front end of the control piston 212 is seated on the outer periphery of the convex portion 219 before the control piston 212 is closed, and the control piston 212 does not close when the control piston 212 controls the pressure regulation. This is to provide an orifice function at 216.

  On the other hand, when the solenoid 221 is de-energized, the control piston 212 and the control sleeve 213 are raised by the biasing force of the spring 218, so that the flow path of the brake fluid is sufficiently opened and the flow rate is ensured. Disappear. Thereby, when the brake pedal 43 is returned, the feeling of dragging of the brake can be eliminated and the responsiveness can be improved.

In this embodiment, the downstream orifice of the downstream communication path as in the ninth embodiment is not provided.
(Example 13)
Next, Example 13 will be described, but the description of the same part as in Example 9 will be omitted or simplified.

In this embodiment, there is provided a lateral force influence reducing mechanism for reducing the influence of the lateral force (side force) due to the magnetic force of the solenoid.
a) As shown in FIG. 22 (abbreviated hatching), the differential pressure control valve 231 of this embodiment has a concave portion (magnetic path escape portion) 233 for reducing the outer diameter of the plunger 232 at the upper portion of the plunger 232. It is provided all around. Further, a convex portion 236 that is slidably supported by the sleeve 234 is provided over the entire periphery of the plunger 232 above the concave portion 233. The formation position of the convex portion 236 is a region where there is almost no magnetic flux generated by the solenoid 237, that is, outside the magnetic path. That is, the magnetic path by the magnetic flux generated by the solenoid 237 is positioned below the recess 233 as shown.

b) Next, the function and effect of this embodiment will be described.
Since the sleeve 234 exists between the plunger 232 and the location where the magnetic flux is generated (the solenoid 237 and the yoke 238), the distance (side) between the member that generates the magnetic flux and the member that receives the urging force by the magnetic flux. There is a gap). When magnetic flux is generated, the plunger 232 receives not only the pressing force in the valve closing direction but also the lateral biasing force (lateral force = side force). However, this lateral force is not always stable when the side gap is present, so that when the current applied to the solenoid 237 increases or decreases, the lateral force depends on the applied current (upstream and downstream of the differential pressure control valve 231). There may be no fixed pressure difference between the two sides. That is, hysteresis occurs.

  Therefore, in this embodiment, in order to reduce the influence of the lateral force due to the magnetic force of the solenoid 237 that is the cause of this hysteresis, a convex portion 236 of the plunger 232 is provided outside the magnetic path, and this convex portion 236 is provided on the sleeve 234. The plunger 232 is slid by being supported.

  Accordingly, as shown in FIG. 23, when a predetermined current is applied to the solenoid 232, the determined differential pressure ΔP can always be uniquely set. Specifically, for example, if the current value is A, the differential pressure set by the differential pressure control valve 231 can be set to ΔPA, and if the current value B is larger than that, ΔP is larger than that. It can be PB.

  Although not described in the ninth embodiment, as shown in FIG. 18, an annular concave portion 152a is provided on the upper portion of the plunger 152, and a convex portion 152b similar to the present embodiment is provided above the concave portion 152a. Is provided outside the magnetic path. This convex portion 152b also has a function of reducing the influence of lateral force, as in the present embodiment.

(Example 14)
Next, although Example 14 is demonstrated, description of the location similar to the said Example 9 is abbreviate | omitted or simplified. Similarly to the thirteenth embodiment, this embodiment also includes a lateral force influence reducing mechanism that reduces the influence of a lateral force (side force) due to the magnetic force of the solenoid.

  As shown in FIG. 24, in the differential pressure control valve 241 of this embodiment, a notch 243 that reduces the outer diameter of the plunger 242 is provided over the entire circumference of the plunger 242. Further, a nonmagnetic ring 246 that is slidably supported by the sleeve 244 is fitted on the lower end of the notch 243.

  In the case of the present embodiment, the plunger 242 is positioned in the radial direction by the nonmagnetic ring 246, so that the influence of the lateral force due to the magnetic force of the solenoid 247 can be reduced. Thereby, the occurrence of hysteresis in the increase or decrease of the current applied to the solenoid 247 described above can be prevented, and a preferable characteristic as shown in FIG. 23 can be realized.

(Example 15)
Next, although Example 15 is demonstrated, description of the location similar to the said Example 9 is abbreviate | omitted or simplified. In the present embodiment as well, in the same manner as in the thirteenth and fourteenth embodiments, a lateral force influence reducing mechanism for reducing the influence of the lateral force (side force) due to the magnetic force of the solenoid is provided.

  As shown in FIG. 25, in the differential pressure control valve 251 of the present embodiment, the outer diameter of the plunger 252 becomes smaller as it goes upward, and a relief portion 253 is formed. The escape portion 253 has a function of stabilizing the movement of the plunger 252 with a certain clearance with the sleeve 255 against an inclination due to the clearance of the support portion 257 described later.

  A plurality of grooves 256 through which brake fluid can pass in the axial direction are formed on the outer periphery of the valve body 254 connected integrally with the plunger 252. In addition, a support portion 257 protruding in the outer peripheral direction of the valve body 254 is formed along the axial direction between the grooves 256, and the support portion 257 is supported by the guide 258 and slides. The valve body 254 is movable in the vertical direction.

  That is, in this embodiment, since the support portion 257 of the valve body 254 is slidably supported by the guide 258, the influence of the lateral force on the magnetic force of the solenoid 259 can be reduced, and thus the hysteresis described above can be eliminated. Can do. In addition, the structure in which the support portion 257 of the valve body 254 is slidably supported by the guide 258 prevents the occurrence of vibration in the vicinity of the seating position of the valve body 254, as in the ninth embodiment. Can be suppressed.

(Example 16)
Next, although Example 16 is demonstrated, description of the location similar to the said Example 9 is abbreviate | omitted or simplified. In the present embodiment, the shape of the needle tip is devised so as to reduce the pressure loss during valve opening.

  Conventionally, as shown in FIG. 26 (a), the tip of the needle has an acute angle, which causes pressure loss when the valve element is opened. Therefore, in this embodiment, as shown in FIG. 26B, the tip of the valve body 261 is cut flat (perpendicular to the axial direction) as the shape of the needle tip 262. Thereby, the pressure loss at the time of valve opening can be reduced.

  In another embodiment, as shown in FIG. 26C, the tip of the needle 272 of the valve body 271 has an obtuse conical shape. Thereby, the pressure loss at the time of valve opening can be reduced. Furthermore, in another Example, as shown in FIG.26 (d), as a shape of the needle front-end | tip part 282 of the valve body 281, the front-end | tip is made into the smooth curve (substantially spherical shape). Thereby, the pressure loss at the time of valve opening can be reduced.

(Example 17)
Next, Example 17 will be described. The present embodiment shows a method for inspecting a differential pressure control valve, and here, the differential pressure control valve of the ninth embodiment is taken as an example as a differential pressure control valve to be inspected.

  As shown in FIG. 27, as the unit 301 for inspection, a differential pressure control valve before assembly, in which a plunger 152, a control piston 154, a spring 158, a valve seat 157, a guide 156 and the like are assembled, is used. The inspection reference coil portion 302 is placed on the inspection unit 301 from the plunger 152 side (upward in the drawing). The inspection reference coil portion 302 is for generating magnetic force including a solenoid 153, a yoke 177, a frame body 176, a sleeve 160, and the like.

  Then, as shown in FIG. 28, the inspection unit 301 and the inspection reference coil unit 302 are combined, and the power source 303 is connected to the solenoid 153, and the ammeter 304 is arranged directly with the power source 303. On the other hand, the measuring device 306 is disposed in the main communication hole 163 on the master cylinder side (however, there is no filter 171). The measuring device 306 has a shape in which the rod 308 is extended from the tip of the conical portion 307, and the upper end of the rod 308 is in contact with the lower end of the needle tip 165. A load meter (not shown) is arranged below the measuring device 306 in order to measure the pressing force applied to the measuring device 306, that is, the pressing force applied to the rod 308 from the needle tip 165.

In this configuration, for example, the current applied to the solenoid 153 is changed, the load in that case (the pressing force of the control piston 154 by the magnetic force of the solenoid 153) is examined, and the desired solenoid attractive force characteristics are determined from the relationship between the current and the load. You can find out if
(Example 18)
Next, Example 18 will be described. The present embodiment shows a method for adjusting a differential pressure control valve using an adjustment shim. Here, the differential pressure control valve of the ninth embodiment is taken as an example of the differential pressure control valve to be inspected.

  For example, it is assumed that the measurement result (suction force characteristic of the solenoid 153) indicated by the solid line in FIG. 29 is obtained by the inspection method shown in Example 17. In this case, the target suction force characteristic is indicated by a broken line in FIG. Since the control piston 154 is biased in the valve opening direction by a spring (return spring) 158, when the current applied to the solenoid 153 is 0, the load F is the offset load of the spring 158 (however, the reverse) It is-) because it is a direction.

  Therefore, shim adjustment is performed in order to change the measured suction force characteristic of the solid line to the target suction force characteristic of the broken line. In this shim adjustment, as shown in FIG. 30, the adjustment shim 401 is placed (placed) between the spring 158 and the control piston 154 to adjust the offset load of the spring 158. For example, by offsetting the adjustment shim 401, the offset load is translated downward by Δt as shown in FIG. As a result, the solid-line attractive force characteristic can be changed to the broken-line attractive force characteristic.

  Since the offset load can be changed by changing the thickness of the adjustment shim, it is possible to change the suction force characteristic by moving it to an arbitrary position. Note that the load F (or FA) corresponds to the differential pressure ΔP (or ΔPA) × the sectional area S of the valve portion (upstream) orifice 162 when the actual differential pressure control valve is used.

(Example 19)
Next, Example 19 will be described. This embodiment shows another aspect of the adjustment method of the differential pressure control valve using the adjustment shim shown in the embodiment 18, and here, as the differential pressure control valve to be inspected, the embodiment 9 and A substantially similar differential pressure control valve is taken as an example.

  In this case, the effect of the adjustment is the same as that of the eighteenth embodiment, and the offset load is adjusted by changing the set length of the return spring. In this embodiment, the valve seat and the sleeve are press-fitted and after welding. It is different from Example 18 in that the adjustment is possible in the subsequent process in assembling.

  That is, in the eighteenth embodiment, the adjustment shim is inserted between the return spring and the stepped portion (spring receiving surface) of the body valve. In this configuration, however, the adjustment shim is attached before the valve seat or the sleeve is assembled. It is necessary to select, and it is not possible to cover the characteristic change due to subsequent assembly.

  Therefore, in this embodiment, as shown in FIGS. 31 and 32, the adjustment shim 501 is inserted from the outside through a hole (insertion hole) 509 provided in the guide 507 between the return spring 503 and the valve seat 505. It is possible. The adjustment shim 501 has a tip end side (left side in both figures) branched in two and a rear end side bent vertically (downward in FIG. 31). A substantially U-shaped cutout 513 is provided so that the portion 511 can enter. As a result, the lower end of the return spring 503 is received in contact with the upper surface of the adjustment shim 501.

  That is, in the present embodiment, the characteristic adjustment as the final valve configuration can be performed in the subsequent process in the assembly after the press-fitting and welding of the valve seat 505 and the sleeve 515, that is, in the final assembly process. Is possible. Note that the adjustment shim 501 is more preferably provided with a lock mechanism in order to prevent removal. 31 and 32, the locking mechanism is an engaging portion 517 on the rear end side of the adjustment shim 501 and a locking groove 519 in the insertion hole 509 of the guide 507.

(Example 20)
Next, Example 20 will be described. The present embodiment shows another aspect of the differential pressure control valve offset load adjusting method similar to that of the above-described embodiment 18 and embodiment 19, and here, as the differential pressure control valve to be inspected, the embodiment 9 described above. As an example, a differential pressure control valve that is almost the same as the above will be described.

  In this case, the effect of the adjustment is the same as that of the eighteenth embodiment and the nineteenth embodiment, and the offset load is adjusted by changing the set length of the return spring. The characteristic adjustment in the similar post-assembly process is performed by means different from that of the nineteenth embodiment.

  That is, in the eighteenth and nineteenth embodiments, the set load of the return spring is changed by shim adjustment. However, in this embodiment, the return spring is changed step by step by changing the combination of engagement of the ratchet mechanisms. The set load is changed.

  Specifically, as shown in FIG. 33, a disc-shaped ratchet 605 is disposed between the return spring 601 and the valve seat 603. As shown in FIG. 34, the ratchet 605 is provided with a spline-like outer peripheral tooth portion 605a and a ratchet-like lower surface tooth portion 605b, and a claw portion 607 on the valve seat 603 side facing the lower surface tooth portion 605b. Is provided.

  Then, a combination of engagement between the ratchet 605 and the claw portion 607 on the valve seat 603 side, that is, any one of (1), (2), and (3) in FIG. Thus, it is possible to change the position by rotating the adjustment hole 611 provided in the guide 609 shown in FIG. 33 by hooking the outer peripheral tooth portion 605a of the ratchet 605 with a screwdriver or the like.

  Thereby, the relative position of the ratchet 605 and the valve seat 603 changes, and as a result, the set length of the return spring 601 can be adjusted. Therefore, in this embodiment, as in the case of the nineteenth embodiment, the characteristic adjustment as the final valve configuration can be performed in the final assembly process, and the characteristic adjustment with higher accuracy can be performed. Further, there is an advantage that the configuration for adjustment is simple.

(Example 21)
Next, Example 21 will be described. The present embodiment shows a method for adjusting a differential pressure control valve using an adjustment plate. Here, the differential pressure control valve of the ninth embodiment will be described as an example of the differential pressure control valve to be inspected.

  For example, it is assumed that a measurement result (suction force characteristic) indicated by a solid line in FIG. 35 is obtained by the inspection method shown in Example 17. In this case, the target suction force characteristic is indicated by a broken line in FIG. Therefore, the magnetic aperture adjustment is performed in order to change the measured attractive force characteristic of the solid line to the target attractive force characteristic of the broken line.

  As shown in FIG. 30, the magnetic diaphragm adjustment is performed by replacing the plurality of plate-shaped adjustment plates 402 constituting the frame body 176 and changing the shape of the magnetic diaphragm portion 403, thereby changing the magnetic path. The gradient of the attractive force characteristic is adjusted by adjusting the state, that is, the magnetic flux passing through the frame 176.

  As the adjustment plate 402, a ring-shaped one can be adopted, and a plurality of adjustment plates 402 in which only a portion constituting the magnetic diaphragm portion 403 is cut out are prepared, selected, and used. 176 is configured. For example, if the adjustment plate 402 is changed to increase the magnetic aperture 403 (particularly, the cross-sectional area in the vertical direction in the figure), the magnetic path becomes smaller. As a result, even if the applied current is increased, the load is hardly increased. That is, when the magnetic diaphragm portion 403 is enlarged, the gradient of the attractive force characteristic is reduced.

  Therefore, since the offset load is not changed, the gradient of the attractive force characteristic is reduced by Δα as shown in FIG. As a result, the solid-line suction force characteristic can be changed to the broken-line suction force characteristic.

(Example 22)
Next, Example 22 will be described.
The present embodiment shows a method for adjusting a differential pressure control valve by adjusting an air gap between a mover and a stator on a magnetic circuit. Here, as a differential pressure control valve to be inspected, the ninth embodiment is described. The differential pressure control valve is taken as an example. For example, it is assumed that a measurement result (suction force characteristic) indicated by a solid line in FIG. 35 is obtained by the inspection method shown in Example 17. In this case, the target suction force characteristic is indicated by a broken line in FIG.

  Therefore, in the present embodiment, the air gap adjustment is performed in order to change the measured suction force characteristic of the solid line to the target suction force characteristic of the broken line. As shown in FIG. 30, the air gap adjustment is performed by adjusting the press-fitting amount of the valve seat 157 with respect to the guide 156, so that the mover (plunger 152) and the stator (guide 156) constituting the magnetic circuit are interposed. Thus, the gradient of the attractive force characteristic is adjusted by adjusting the state of the magnetic path, that is, the magnetic flux passing through the air gap.

  For example, if the amount of press-fitting of the valve seat 157 is changed (increased) to increase the air gap when the valve is closed, the generated magnetic force decreases. As a result, even if the applied current is increased, the load is hardly increased. That is, when the air gap is increased, the gradient of the suction force characteristic is decreased.

  Accordingly, since the offset load is not changed, if the air gap is increased, the gradient of the attractive force characteristic is decreased by Δα as shown in FIG. As a result, the solid-line suction force characteristic can be changed to the broken-line suction force characteristic. Contrary to the above, when the air gap is reduced, the gradient of the attractive force characteristic is increased.

(Example 23)
Next, Example 23 will be described. In this embodiment, the offset load change by the return spring set load adjustment shown in the embodiment 18 and the suction force characteristic change by the air gap adjustment shown in the embodiment 22 are adjusted by adjusting the press-fit amount of the valve seat. The method to realize simultaneously is shown.

  As a difference in configuration and effect with respect to the twenty-second embodiment, the receiving surface of the return spring 701 is changed to the guide 705 instead of the valve seat 703 as shown in FIG. Thus, by adjusting the press-fitting amount (seat press-fitting amount) of the valve seat 703, the air gap between the mover (plunger 707) and the stator (guide 705) can be changed, and the return spring 701 is set at the same time. The length (the valve closing side that is in the state of adjustment) can also be changed.

  Therefore, even in the region where the suction force change with respect to the air gap is small as shown in FIG. 1 (the actual use region in FIG. 1), the adjustment gain can be maintained by changing the set length. That is, in the actual use region of FIG. 1, even if the air gap is slightly changed, the suction force changes little and the adjustment effect is difficult to achieve. However, in this embodiment, the adjustment is accompanied by the offset load change at the same time. The adjustment effect is always secured.

In this embodiment, the configuration of the valve seat 703 can be simplified, and the drilling of the check valve 709 portion can be concentrated on the guide 705 side.
(Example 24)
Next, Example 24 will be described.

  This embodiment shows a method of adjusting a differential pressure control valve using an adjustment shim and an adjustment plate. Here, the differential pressure control valve of the ninth embodiment is taken as an example of the differential pressure control valve to be inspected. I will give you. In this example, the adjustments of Example 18 and Example 21 were performed together.

  For example, it is assumed that a measurement result (suction force characteristic) indicated by a solid line in FIG. 37 is obtained by the inspection method shown in Example 17. In this case, the target suction force characteristic is indicated by a broken line in FIG. Therefore, in order to change the measured attractive force characteristic of the solid line to the target attractive force characteristic of the broken line, magnetic aperture adjustment and shim adjustment are performed.

  First, the gradient of the attractive force characteristic indicated by the solid line is changed by adjusting the magnetic aperture to obtain the attractive force characteristic indicated by the alternate long and short dash line. Next, by the shim adjustment, the suction force characteristic indicated by the one-dot chain line is translated to obtain the suction force characteristic indicated by the broken line in the figure.

  As a result, even if the attractive force characteristic of the differential pressure control valve to be measured is greatly deviated from the target attractive force characteristic, it is possible to easily obtain the target attractive force characteristic by this magnetic aperture adjustment and shim adjustment. Can do. In the magnetic aperture adjustment, an example of reducing the gradient of the attractive force characteristic has been given. However, a reference magnetic aperture portion is set, and the magnetic aperture portion is enlarged to reduce the gradient, and the magnetic aperture portion is The gradient can be increased by decreasing the value.

  Similarly, with respect to shim adjustment, a reference adjustment shim is arranged, and by increasing the thickness of the adjustment shim, the offset load can be increased, and the suction force characteristic can be translated downward in FIG. By making the adjustment shim thinner, the offset load can be reduced and the suction force characteristic can be translated upward in FIG.

  Needless to say, the present invention is not limited to the above-described embodiments and can be carried out in various modes without departing from the technical scope of the present invention. For example, the differential pressure control valve of each embodiment described above can be used for the vehicle brake device of FIGS.

1, 71, 72, 91, 101, 111, 121, 131, 141, 151, 181, 191, 211, 231, 245, 251... Differential pressure control valve 2, 144, 153, 221, 237, 247, 259. Solenoid 3,103,113,123,133,142,152,232,242,252,707 ... plunger 4,102,122,132,160,234,244,515 ... sleeve 6,106,117,134,154 , 192, 212, 254, 261, 271, 281, 510... Control piston 7, 105, 115, 126, 135, 146, 156, 182, 258, 507, 609, 705, guide 8, 145, 177, 238,. Yoke 21, 104, 116, 125... (In magnetic short part) convex parts 22, 108, 11 ... recess (in the magnetic short section) 31,41,92 ... master cylinder 32,55,56,57,58,93,94 ... wheel cylinders 36,77,79 ... pump 261,271,281 ... valve body

Claims (5)

In the differential pressure control valve that is arranged in the brake fluid conduit and adjusts the differential pressure of the brake fluid pressure,
The differential pressure control valve includes a mover through which a magnetic flux generated by a solenoid passes and moves by the magnetic force;
A stator that is arranged in the moving direction of the mover and through which the magnetic flux by the solenoid penetrates;
A sleeve that regulates the maximum stroke of the mover;
A valve body that regulates a minimum stroke value of the mover by being pressed by the mover and seated on a valve seat;
An electronic control unit that performs current control so as to generate the attraction force of the solenoid according to the magnitude of the current flowing through the solenoid;
With
A magnetic short structure that suppresses fluctuations in the amount of magnetic flux in the moving direction of the mover and the stator along with the movement of the mover;
The maximum value and the minimum value of the stroke of the mover are set within a stroke range in which the attraction force of the solenoid is not changed by the magnetic short structure,
The electronic control unit energizes the solenoid and the differential pressure of the brake fluid pressure obtained from the balance between the force acting on the valve body due to the differential pressure of the brake fluid pressure and the suction force of the solenoid. Based on a proportional relationship with the current value to be calculated, the current value to be energized to the solenoid is calculated,
The differential pressure control valve according to claim 1, wherein the valve element that is actuated by the operation of the mover includes a conical needle that closes a passage of the brake fluid by closing a hole provided in the valve seat.   2. The differential pressure control valve according to claim 1, wherein a pressure loss reducing structure for reducing a pressure loss at the time of valve opening is provided at a needle tip portion of a valve body that is operated by an operation of the mover.   The differential pressure control valve according to claim 2, wherein the pressure loss reducing structure is configured to make the needle tip portion an obtuse angle.   The differential pressure control valve according to claim 2, wherein the pressure loss reducing structure is configured to cut a needle tip portion flatly.   The differential pressure control valve according to claim 2, wherein the pressure loss reducing structure is configured to smoothly curve the tip of the needle.

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