JP2008279964A - Brake control device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a brake control device in which a wheel cylinder pressure is stabilized even if the wheel cylinder pressure is high and the fluid pressure gradient is small. <P>SOLUTION: When the wheel cylinder pressure is pressurized by a pump, both a motor for driving the pump and a pressure reducing valve for reducing the pressure of the wheel cylinder are operated to control the wheel cylinder pressure. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a brake control device capable of increasing and decreasing a wheel cylinder hydraulic pressure so as to achieve a target deceleration.

  As a brake control device, a technique described in Patent Document 1 is known. The brake control device generates wheel cylinder pressure according to the operating force of the brake pedal, the state of the vehicle, etc., and applies a braking force to the vehicle. The brake fluid pressure control device includes two pumps for pressurizing brake fluid in the wheel cylinder, a pressure increasing valve and a pressure reducing valve provided for each wheel cylinder, and further, communication between the master cylinder and the wheel cylinder is shut off. And a shut-off valve. When this shut-off valve is closed, the wheel cylinder pressure is controlled by two pumps, a pressure increasing valve, and a pressure reducing valve.

Basically, pressure increase in the wheel cylinder is performed by supplying brake fluid from at least one pump to the wheel cylinder, opening the pressure increasing valve, and closing the pressure reducing valve. The wheel cylinder is depressurized by closing the pressure increasing valve and opening the pressure reducing valve. At this time, the pump is driven and controlled so that the pump pressure becomes the maximum value of all the command wheel cylinder pressures or a pressure obtained by adding a predetermined value to the maximum value.
Japanese Patent No. 3409721

  In the configuration of the prior art, when the wheel cylinder pressure is high and the hydraulic pressure gradient is small, or when the wheel cylinder pressure is maintained at a high pressure, the pump is driven at a low rotation speed while keeping the pump discharge pressure high. That is, it is necessary to control the rotational speed of the motor with low rotation and high torque. However, at the time of low rotation and high load, the frictional force of the pump becomes unstable, it is difficult to stabilize the rotational speed of the motor that drives the pump, and it is difficult to stabilize the wheel cylinder pressure.

  The present invention has been made paying attention to the above-mentioned problems, and the object of the present invention is to provide a brake control capable of stabilizing the wheel cylinder pressure even when the wheel cylinder pressure is high and the hydraulic pressure gradient is small. To provide an apparatus.

  In order to achieve the above object, in the present invention, a wheel cylinder provided on each of a plurality of wheels, a pump for pressurizing brake fluid in the wheel cylinder, a motor for rotationally driving the pump, and a wheel cylinder in the wheel cylinder A pressure reducing valve for reducing the pressure by causing the brake fluid to flow out by opening the valve, and a control means for operating and controlling both the motor and the pressure reducing valve so that the hydraulic pressure in the wheel cylinder becomes a target hydraulic pressure. Prepared.

  Therefore, even if the wheel cylinder pressure is high and the hydraulic pressure gradient is small, the wheel cylinder pressure can be stabilized.

  Hereinafter, the best embodiment of the present invention will be described with reference to the drawings.

[Brake-by-wire system configuration]
FIG. 1 is an overall block diagram of a brake-by-wire system composed of both a hydraulic type and an electric caliper type. The hydraulic brake-by-wire system is a system in which a pump is driven by an electric motor, and the hydraulic pressure in the wheel cylinder is controlled by the pump hydraulic pressure (or the hydraulic pressure accumulated in the accumulator), thereby generating a braking force. To tell.

  The electric caliper brake-by-wire system refers to a system that generates a braking force by controlling a pressing force of a brake pad against a brake rotor by an electric motor. In the first embodiment, a hydraulic brake-by-wire system is adopted for the two front wheels, and an electric caliper brake-by-wire system is adopted for the two rear wheels.

  The control unit CU calculates normal brake control according to the driver's operation, vehicle information such as anti-skid brake control (hereafter ABS), vehicle behavior control (hereafter VDC), inter-vehicle distance control, and obstacle avoidance control. Is used to calculate tire braking and vehicle behavior to calculate the braking force required for the vehicle (both front side and rear side) so that the required braking force target value on the front side is obtained. Controls wheel cylinder hydraulic pressure on the front side. The configuration of the hydraulic unit HU used for front-side wheel cylinder hydraulic pressure control will be described later.

  The control unit CU outputs the braking force target value required for the rear side to the left and right rear controller RCU. In the left and right rear controller RCU, the motor provided in the electric caliper EMB so as to obtain the braking force target value of each wheel To control the driving amount.

  The master cylinder M / C is provided with a stroke sensor S / Sen and a stroke simulator S / Sim. As the brake pedal BP is depressed, hydraulic pressure is generated in the master cylinder M / C, and the stroke signal S of the brake pedal BP is output to the control unit CU.

  The master cylinder pressure is supplied to the hydraulic unit HU via the oil passage A (FL, FR), and after the hydraulic unit HU is driven by the control unit CU to control the hydraulic pressure, the oil passage D (FL, FR, FR) and supplied to the front wheel cylinder W / C (FL, FR).

  The control unit CU calculates the regenerative braking force and the friction braking force in consideration of the limitation of the battery SOC (State Of Charge) and the limitation of the actuator response while achieving the required braking force. The regenerative braking device MGB generates braking force. For friction braking force, FL and FR wheel target hydraulic pressures P * fl and P * fr are calculated to drive the hydraulic unit HU, and the wheel cylinder W The braking force is generated by controlling the hydraulic pressure of / C (FL, FR).

  Further, the left and right rear controller RCU controls the braking force of the electric caliper EMB based on the target signal from the control unit CU.

  The hydraulic unit HU blocks communication between the master cylinder and the wheel cylinders W / C (FL, FR) during normal braking in the brake-by-wire system. On the other hand, the hydraulic pressure is supplied to the wheel cylinder W / C (FL, FR) by the pump P to generate a braking force.

  Then, the hydraulic pressure in the front wheel cylinder W / C (FL, FR) is reduced by appropriately driving a pressure reducing valve. When the brake-by-wire function fails, the master cylinder pressure is FL, and the hydraulic pressure is supplied to the FR wheel cylinder W / C (FL, FR) to obtain the braking force.

[Hydraulic circuit]
FIG. 2 is a hydraulic circuit diagram of the first embodiment. The discharge side of pump P is connected to FL and FR wheel cylinders W / C (FL, FR) via oil passage C (FL, FR) and oil passage D (FL, FR), respectively, and the suction side is an oil passage. Connect to reservoir RSV via B. The oil passage C (FL, FR) is connected to the oil passage B via the oil passage E (FL, FR).

  Also, the connection point I (FL) between the oil path C (FL) and the oil path E (FL) is connected to the master cylinder M / C via the oil path A (FL), and the oil path C (FR) and the oil path The connection point I (FR) of E (FR) is connected to the master cylinder M / C via the oil passage A (FR). Further, the connection point J of the oil passage C (FL, FR) is connected to the oil passage B through the oil passage G.

  Shut-off valve S. OFF / V (FL, FR) is a normally open solenoid valve that is provided on the oil passage A (FL, FR) to communicate / block the master cylinder M / C and connection point I (FL, FR). .

  The FL and FR wheel booster valves IN / V (FL and FR) are normally open proportional valves provided on the oil passage C (FL and FR), respectively. The discharge pressure of the pump P is proportionally controlled to control the FL and FR wheels. Supply to wheel cylinder W / C (FL, FR). A check valve C / V (FL, FR) for preventing back flow to the pump P side between the FL and FR wheel booster valves IN / V (FL, FR) on the oil passage C (FL, FR). ) Is provided.

  The FL and FR wheel pressure reducing valves OUT / V (FL, FR) are normally closed proportional valves and are provided on the oil passages E (FL, FR), respectively. A relief valve Ref / V is provided on the oil passage G connecting the connection point J and the oil passage B.

  Shut-off valve S. The oil passage A (FL, FR) between OFF / V (FL, FR) and the master cylinder M / C is provided with first and second M / C pressure sensors MC / Sen1, MC / Sen2, and M / Output C pressure PM1, PM2 to the control unit CU. In addition, FL and FR wheel hydraulic pressure sensors WC / Sen (FL, FR) are provided in the hydraulic unit HU and on the oil passage D (FL, FR), and the detected values Pfl and Pfr are sent to the control unit CU. Output.

[Configuration of pump P]
A gear pump is housed as a pump P in the hydraulic unit HU of the first embodiment. The configuration of this gear pump will be described below. FIG. 3 is a cross-sectional view showing the configuration of the pump P arranged in the hydraulic unit HU. For the sake of explanation, only the casing (pump housing 1a) has a cross section, and the pump assembly 3a housed in the casing is a side view.

  The pump housing 1a is provided with a cylindrical cylinder hole 1b for housing the pump assembly 3a. A drive shaft support hole 2a is provided on the bottom surface of the cylinder hole 1b. A bearing 20a is provided on the inner periphery of the drive shaft support hole 2a, and a drive shaft 10A described later is rotatably supported in the drive shaft support hole 2a.

  The inner peripheral surface of the cylinder hole 1b includes a positioning contact surface 101a and an inner wall 102a. The contact surface 101a is formed with higher accuracy in relation to the drive shaft support hole 2a than the inner wall 102a. A discharge port is provided in the radial direction of the pump housing 1a, and the cylinder hole 1b communicates with the outside. A suction port 4a communicating with the brake fluid suction passage is provided on the surface of the pump housing 1a that faces the surface including the drive shaft support hole 2a.

  As shown in FIG. 4, the pump assembly 3a is provided on a drive gear 10a provided on the drive shaft 10A, a driven gear 11a provided on the driven shaft 11A, and on both sides in the axial direction of the drive shaft 10A and the driven shaft 11A. A pair of side plates 7a and 8a and a seal block 12a are formed. A brushless motor M is connected to the drive shaft 10A. In addition, the pump assembly 3a of an Example shall be supported with respect to the casing only by the drive shaft 10A, and the driven shaft 11A is not supported.

  Next, the pump driving action will be described. When the drive shaft 10A is driven by the motor M, the driven gear 11a is driven via the drive gear 10a. By this action, a low-pressure fluid is introduced from the suction port 4a, and the high-pressure fluid is output to the high-pressure chamber 16a provided between the cylinder hole 1b, the oil seal 5a, and the pump assembly 3a. This high-pressure fluid is output from a discharge port provided in the pump housing 1a to a wheel cylinder (not shown).

  At this time, the outer periphery of the pump assembly 3a becomes high pressure by driving the drive shaft 10A, and the inside of the suction port 4a becomes low pressure. Therefore, each component (side plates 7a, 8a, seal block 12a, etc.) toward the low pressure portion. ) Aggregate. That is, at this time, each component takes a position where it comes into close contact while moving relatively. By such an action, the sealing performance of the gear pump is improved, and accordingly, the differential pressure between the low pressure part (suction port 4a) and the high pressure part (high pressure chamber 16a) can be secured more efficiently, and the pump performance is improved. It is.

[Normal brake in brake-by-wire control]
(When pressure is increased)
Shut-off valve S during normal brake pressure increase in brake-by-wire control. OFF / V (FL, FR) is closed, booster valve IN / V (FL, FR) is opened, decompressor valve OUT / V (FL, FR) is closed, motor M is driven, booster valve IN Increase the pressure by controlling the hydraulic pressure with / V (FL, FR).

(At reduced pressure)
During normal brake pressure reduction, the specified pressure increasing valve IN / V (FL, FR) is closed, the pressure reducing valve OUT / V (FL, FR) is opened, the hydraulic pressure is discharged to the reservoir RSV, and pressure reduction is performed. Note that the booster valve that performs full-opening control of the booster valve described later is not closed.

(When holding)
When holding the normal brake, the predetermined pressure increasing valve IN / V (FL, FR) and pressure reducing valve OUT / V (FL, FR) are closed to maintain the hydraulic pressure. As in the case of the pressure reduction, the pressure increasing valve that performs the pressure increasing valve full open control described later is not closed.

[Manual brake]
Normally open shut-off valve S. OFF / V and pressure increase valve IN / V (FL, FR) are opened, and normally closed pressure reducing valve OUT / V (FL, FR) is closed. Therefore, the master cylinder pressure Pm acts on the FL and FR wheel wheel cylinders W / C (FL, FR). This ensures a manual brake.

[Brake-by-wire booster valve fully open control]
If the oil passage volume from the pump discharge side to the booster valve is set small to reduce the size of the device, the oil passage volume that absorbs pump pulsation will be reduced when the booster valve is closed, and pressure fluctuation on the pump discharge side will be reduced. It becomes larger and the controllability deteriorates. Therefore, in Example 1 of the present application, the front pressure wheel IN / V_H having the higher target hydraulic pressure P * in the front wheel cylinder W / C (FL, FR) using the hydraulic brake is fully opened (the flow path area is maximized). To do.

  Since the hydraulic circuit of the first embodiment increases the pressure of the FL and FR wheel wheel cylinders W / C (FL, FR) with one pump P, when there is a pressure increasing wheel, the pump discharge pressure Pp is at least the high pressure side wheel cylinder. The target hydraulic pressure of P * _H must be exceeded. On the other hand, the target hydraulic pressure P * _L of the low pressure side wheel cylinder may be obtained by proportionally controlling the discharge pressure Pp by the pressure increasing valve IN / V.

  Therefore, the pressure increasing valve IN / V_H on the side with the higher target hydraulic pressure P * is fully opened, and the wheel cylinder W / C_H on the side with the higher target hydraulic pressure P * is directly controlled by the pump discharge pressure Pp. As a result, the discharge side of the pump P and the wheel cylinder W / C_H on the side with the higher target hydraulic pressure P * are in communication with each other, and the oil passage volume on the pump discharge side is increased to reduce the vibration of the hydraulic oil on the pump discharge side. To do. Further, since the wheel cylinder W / C_H on the side where the target hydraulic pressure P * is higher is directly controlled by the pump discharge pressure Pp, the discharge pressure Pp of the pump P may be the minimum necessary.

  Here, in order to feedback control the discharge pressure Pp of the pump P, it is necessary to provide a pump discharge pressure sensor for detecting the pump discharge pressure. In the first embodiment, since the pressure increasing valve IN / V_H on the high target hydraulic pressure P * _H side is fully opened, the pump discharge pressure Pp and the hydraulic pressure P_H on the high target hydraulic pressure P * _H side are substantially equal. Therefore, the pump discharge pressure sensor can be omitted by detecting the pump discharge pressure Pp with the hydraulic sensor WC / Sen_H on the high target hydraulic pressure side. That is, in the first embodiment, the hydraulic pressure of the wheel that maximizes the flow path area among the plurality of wheels is selected, and this hydraulic pressure is estimated as the pressure of the hydraulic pressure source, thereby reducing the discharge pressure of the hydraulic pressure source. The hydraulic pressure sensor to be detected is omitted.

  That is, in the pressure-increasing valve full-opening control in the brake-by-wire according to the first embodiment, a wheel that requires the maximum hydraulic pressure is selected, and a feedback loop is configured between the wheel cylinder pressure of the wheel and the pump P (hereinafter referred to as the following) The other wheels independently form a feedback loop between the wheel cylinder pressure and the pressure increasing valve IN / V (hereinafter referred to as a second feedback loop).

  Normally, when the opening degree of the booster valve IN / V changes, the hydraulic pressure downstream of the booster valve IN / V can be controlled, while the hydraulic pressure upstream of the booster valve IN / V also changes. Therefore, the influence of the second feedback loop may affect the second feedback loop of other wheels, and the control system may diverge. On the other hand, in the first embodiment, a stable control system is configured by acting so as to absorb the influence of the first feedback loop on the upstream side of the pressure increasing valve IN / V in the second feedback loop.

  As described above, the pump P according to the first embodiment employs the gear pump P including the seal block 12a. Friction is generated in the gear pump P because of the contact between the drive shaft 10A and the oil seal 5a, the contact between the drive shaft 10A and the bearing 20a, and the drive shaft 10A, the driven shaft 11A and the side plates 7a and 8a. The part and the part which the drive gear 10a, the driven gear 11a, and the side plates 7a and 8a contact can be considered.

  Of these, the portions where the drive gear 10a, the driven gear 11a and the side plates 7a and 8a are in contact with each other seem to be balanced in force between the side of the side plates 7a and 8a facing the gear and the side facing the high pressure chamber 16a. Therefore, the friction force is considered to be small.

  FIG. 16 is a diagram showing the oil film pressure distribution of the journal bearing. In general, it is known that in such a journal bearing, the oil film pressure is distributed as shown in FIG. When the distance between the shaft and the bearing becomes narrower, the pressure of the oil film at that portion increases, and a force that separates the distance between the shaft and the bearing acts to prevent the metal contact between the shaft and the bearing to lubricate. Here, P is the pressure, e is the amount of eccentricity of the rotating body, φ is the eccentric angle, W is the bearing load, r is the radius of the rotating body, and the like. In addition, since it is a general phenomenology, it does not explain in particular.

  FIG. 17 is a diagram showing the relationship between the rotational speed of the gear pump and the frictional force. The oil film pressure increases as the shaft rotates more rapidly. That is, the higher the rotation speed, the smaller the friction force. At a certain rotation speed or higher, the distances between the drive shaft 10A and the bearing 20a, the drive shaft 10A, the driven shaft 11A, and the side plates 7a and 8a are stabilized, so that the frictional force is substantially constant.

  In addition, the higher the pump discharge pressure, the stronger the shaft is pressed against the bearing, so the higher the pump discharge pressure, the greater the frictional force. When controlling the rotational speed of a gear pump having such characteristics, it is difficult to control because the frictional force increases rapidly in the region where the pump discharge pressure is high and the pump is low. Furthermore, if the motor is a brushless motor and the rotation angle is detected by a Hall element, the machine angle of the motor can only be distinguished every 60 °, so whether the motor is rotating or stopped when it rotates at a low speed. Judgment is difficult. That is, it can be said that it is very difficult to control the rotation speed of the pump P in the low rotation region.

  In addition, when the characteristics shown in FIG. 17 are stable, it may be possible to control by introducing a correction coefficient or the like. As a result of actually measuring the relationship between the rotational speed and the frictional force, this characteristic is obtained. Although it was a general trend, it was understood that the variation was large. That is, as shown by the arrows in FIG. 17, in the region where a high frictional force is generated at a low rotation, the variation in the frictional force itself is large, and the variation in the accuracy of the sensor at a low rotation is large. It was found that it is very difficult to control the rotational speed with high accuracy.

  Therefore, in the first embodiment, in order to avoid the use in the region where the frictional force is unstable, the calculated target motor rotation speed N * is set according to the discharge pressure in the motor target rotation speed calculation unit 114. When the pump rotational speed threshold value N0, which is the lowest rotational speed of the pump drive, falls below the target motor rotational speed N *, the rotational speed is increased so as to be the pump rotational speed threshold value N0.

  At this time, since the excess brake fluid is discharged from the pump P by the raised amount, a feedback control system for the pressure reducing valve OUT / V is configured to discharge the excess brake fluid from the pressure reducing valve OUT / V, and a desired foil is obtained. The cylinder pressure was achieved.

  Here, among the plurality of pressure reducing valves OUT / V, it becomes a problem from which pressure reducing valve OUT / V the excess brake fluid is discharged. In the brake control of the first embodiment, the booster valve IN / V corresponding to the maximum W / C target hydraulic pressure is fully opened, and feedback control is performed for the other booster valves IN / V. For this reason, even if the pressure reducing valve OUT / V that is paired with a pressure increasing valve IN / V that is not necessarily fully opened (hereinafter referred to as a non-full pressure increasing valve) is opened, Excess brake fluid does not flow all the way, and it cannot be said that the controllability is good.

  Therefore, it is decided to discharge excess brake fluid from the pressure reducing valve OUT / V corresponding to the pressure increasing valve IN / V that is fully open (hereinafter referred to as a fully open pressure increasing valve). The fully open booster valve IN / V has the lowest flow resistance, and excess brake fluid actively flows into this wheel cylinder. Further, the control system of the wheel cylinder is not particularly controlled because the pressure increasing valve IN / V is fully open, and the motor M is controlled to be the pump rotation speed threshold value N0. It can be said that the target value is fixed.

  Therefore, by configuring the feedback control system so that excess brake fluid is discharged from the pressure reducing valve OUT / V corresponding to this fully open pressure increasing valve IN / V, a plurality of target values are frequently changed to one system. Therefore, the controllability can be improved.

  The positioning when the feedback control system using the pressure reducing valve OUT / V is the third feedback loop will be described based on the relationship with the first feedback loop. The wheel cylinder system to which the fully open pressure increasing valve IN / V belongs constitutes a first feedback loop between the pump speed and the target hydraulic pressure. However, as described above, if the motor rotation speed threshold N0 is defined, the wheel cylinder pressure cannot be controlled by the motor rotation speed feedback. Therefore, hydraulic pressure feedback control is performed by the pressure reducing valve OUT / V in the wheel cylinder to which the fully open pressure increasing valve IN / V belongs.

  That is, even if the first feedback loop is activated, when the rotational speed of the motor M cannot be an element for controlling the wheel cylinder pressure, the third feedback loop of the pressure reducing valve OUT / V and the wheel cylinder pressure is switched. I can say that.

[Brake-by-wire control processing]
FIG. 5 is a flowchart showing the flow of the brake-by-wire control process. Hereinafter, each step will be described.

In step S10, it is determined whether the target hydraulic pressure mode is increased, held, or reduced, and the process proceeds to step S20.
In step S20A, correction processing for the pressure increase threshold value and the pressure reduction threshold value is executed based on the target hydraulic pressure mode determined in step S10.
In step S20, based on the pressure increase threshold value and pressure reduction threshold value set in step S20A and the wheel cylinder hydraulic pressure deviation, it is determined whether the W / C hydraulic pressure control mode is to be increased, held, or reduced.

In step S30, it is determined whether the pressure increasing valve control mode is fully opened, fully closed, or proportional control (intermediate opening), and the process proceeds to step S30A.
In step S30A, it is determined whether the pressure reducing valve control mode is fully open, fully closed, or proportional control (intermediate opening), and the process proceeds to step S40.

  In step S40, it is determined whether or not there is at least one W / C whose W / C hydraulic pressure control mode is increased. If YES, the process proceeds to step S50, and if NO, the process proceeds to step S60.

In step S50, a pump control process is executed, and the process proceeds to step S60.
In step S60, it is determined whether the pressure increasing valve control mode is proportional control. If YES, the process proceeds to step S70, and if NO, the control is terminated.

In step S70, a pressure increasing valve control process is executed, and the control is terminated.
In step S60A, it is determined whether the pressure reducing valve control mode is proportional control. If YES, the process proceeds to step S70A, and if NO, the control is terminated.
In step S70A, a pressure reducing valve control process is executed, and the control is terminated.

[Target hydraulic pressure mode decision process]
FIG. 6 is a flowchart showing the flow of the target hydraulic pressure mode determination process (FIG. 5: Step S10).
In step S11, it is determined whether or not the gradient ΔP * of the target hydraulic pressure P * is equal to or greater than the pressure increase command threshold. If YES, the process proceeds to step S13, and if NO, the process proceeds to step S12.
In step S12, it is determined whether or not the target hydraulic pressure gradient ΔP * ≦ the pressure reduction command threshold value. If YES, the process proceeds to step S14, and if NO, the process proceeds to step S15.

In step S13, the target hydraulic pressure mode is increased and the control is terminated.
In step S14, the target hydraulic pressure mode is reduced, and the control is terminated.
In step S15, the target hydraulic pressure mode is maintained and the control is terminated.

[Threshold correction processing]
FIG. 7 is a correction table showing the relationship of threshold correction processing (FIG. 5: step S20A). When the pressure is increased by the target hydraulic pressure mode determination process, the pressure increase threshold is corrected to be smaller than a preset reference pressure increase threshold, and the pressure reduction threshold is set to a preset reference pressure reduction. Correction is made to be larger than the threshold value. As a result, a threshold value that is easy to increase in pressure and difficult to decrease in pressure is set.

  Similarly, when the pressure is reduced by the target hydraulic pressure mode determination process, the pressure increase threshold is corrected to be larger than a preset reference pressure increase threshold, and the pressure reduction threshold is set in advance. Correction is made to be smaller than the reference decompression threshold. As a result, a threshold value that is easy to be depressurized and difficult to increase is set. Note that when the target hydraulic pressure mode is determined to be retained, the threshold value is not particularly corrected.

[W / C fluid pressure control mode decision processing]
FIG. 8 is a flowchart showing the flow of the W / C hydraulic pressure control mode determination process (FIG. 5: step S20). As the threshold value in this step, the increasing / decreasing threshold value corrected in step S20A is used.
In step S21, it is determined whether or not the deviation ΔP of the W / C target hydraulic pressure P * and the actual hydraulic pressure is equal to or greater than the pressure increase threshold. If YES, the process proceeds to step S23, and if NO, the process proceeds to step S22. .
In step S22, it is determined whether or not the deviation ΔP between the target hydraulic pressure P * of W / C and the actual hydraulic pressure is equal to or less than the pressure reduction threshold. If YES, the process proceeds to step S24, and if NO, the process proceeds to step S25.

In step S23, the W / C hydraulic pressure control mode is increased and the control is terminated.
In step S24, the W / C hydraulic pressure control mode is reduced, and the control is terminated.
In step S25, the W / C hydraulic pressure control mode is maintained and the control is terminated.

[Pressure boosting valve control mode decision processing (high-pressure side priority fully open)]
FIG. 9 is a flowchart showing the flow of the booster valve control mode determination process (FIG. 5: step S30).

  In step S31, it is determined whether or not the target W / C hydraulic pressure P * (for example, the FL wheel) is the maximum value among the respective W / C target hydraulic pressures P * fl and P * fr. If YES, the process proceeds to step S33. If NO, the process proceeds to step S32.

  In step S32, it is determined whether or not the W / C hydraulic pressure control mode is the pressure increasing mode. If YES, the process proceeds to step S34, and if NO, the process proceeds to step S35.

In step S33, the pressure increasing valve control mode is fully opened, and the control is terminated.
In step S34, the pressure increasing valve control mode is set to proportional control, and the control ends.
In step S35, the pressure increasing valve control mode is fully closed, and the control is terminated.

[Pressure reducing valve control mode determination process]
FIG. 10 is a flowchart showing the flow of the pressure reducing valve control mode determination process (FIG. 5: Step S30A).

In step S201, it is determined whether or not the W / C hydraulic pressure control mode is reduced pressure. If the pressure is reduced, the process proceeds to step S204. Otherwise, the process proceeds to step S202.
In step S202, it is determined whether or not the target rotational speed N * of the pump is equal to or smaller than the pump rotational speed threshold N0. If the target rotational speed N * is equal to or smaller than the threshold, the process proceeds to step S203.

In step S203, it is determined whether or not the pressure increasing valve control mode is fully opened. If it is fully open, the process proceeds to step S204. Otherwise, the process proceeds to step S205.
In step S204, the pressure reducing valve control mode is set to proportional control.
In step S205, the pressure reducing valve control mode is fully closed.

[Pump control processing block diagram]
FIG. 11 is a block diagram of the pump control process executed in step S50 of FIG. Pump control is performed by the pump control unit P. It shall be executed in CU.

  Pump control unit P. The CU includes a target amount calculation unit 110, a target pump pressure calculation unit 111, a W / C fluid amount deviation FB (feedback) calculation unit 112, a pump leak amount calculation unit 113, a motor target rotation number calculation unit 114, a loss torque calculation unit 115, It has a target rotation speed differential calculation unit 116 and a rotation speed deviation FB (feedback) calculation unit 117.

  The target amount calculation unit 110 uses the experimental data of the load stiffness of the caliper and its accessories as shown in FIG. 12, and based on the target hydraulic pressure P * (FL, FR) of each wheel FL, FR, The flow rate Qp * is calculated and output to the multiplication unit 122. Further, the target fluid amount Vw * (FL, FR) of each wheel cylinder W / C (FL to RR) is calculated and output to the adding unit 131. Further, the target hydraulic pressure P * _H of the pressure increasing valve IN / V system wheel cylinder that is fully open is output to the target pump pressure calculation unit 111.

  The target pump pressure calculation unit 111 calculates the target pump pressure Pp * based on the target hydraulic pressure P * _H of the pressure increasing valve IN / V system wheel cylinder that is fully open, the pump leak amount calculation unit 113, the loss torque calculation unit 115 And to the multiplication unit 121.

  The multiplication unit 121 multiplies the target pump pressure Pp * by the theoretical discharge amount Vth / 2π per rotation of the pump P, calculates the required theoretical torque Tth of the pump P required to output the target pump pressure Pp *, and adds it. Output to 134.

  The W / C fluid amount deviation FB calculation unit 112 calculates the target fluid amount Vw * (FL, FR) and the actual fluid amount Vw (FL, FR) of the wheel cylinder W / C (FL, FR) calculated by the addition unit 131. The feedback control calculation is performed using the deviation ΔVw (FL, FR) of the output, and the feedback component ΔVw (FB) is output to the adder 132.

  The pump leak amount calculation unit 113 calculates a pump leak amount Qpl based on an experimental value or the like, and outputs it to the addition unit 132.

  The adding unit 132 adds the pump leak amount Qpl, the fluid amount deviation FB component ΔVw (FB), and the product of the pump target flow rate Qp * and the inverse of the theoretical discharge amount Vth (calculated by the multiplying unit 122), and adds the motor target Output to rotation number calculation unit 114.

  Motor target rotation number calculation unit 114 calculates motor target rotation number N * based on the addition value calculated by addition unit 132, and outputs the result to loss torque calculation unit 115, target rotation number differentiation calculation unit 116, and addition unit 133.

  Further, in the motor target rotation speed calculation unit 114, motor rotation speed raising control is executed. Specifically, it is determined whether or not the target motor rotation speed N * calculated by the motor target rotation speed calculation unit 114 is less than or equal to the pump rotation speed threshold value N0. The number N * is replaced with the pump rotation speed threshold value N0 and output to the loss torque calculation section 115 and the target rotation speed differentiation calculation section 116, and the deviation ΔN0 between the target pump rotation speed N * and the pump rotation speed threshold value is described later. The output is output to the flow rate calculation unit 260 corresponding to the excessive pump rotation of the control unit.

  On the other hand, when the target motor rotation speed N * is equal to or higher than the pump rotation speed threshold value N0, it is not particularly necessary to increase the pump rotation speed, so the target pump rotation speed N * is converted to the loss torque calculation unit 115 and the target rotation speed differentiation calculation unit 116. Output as is. At this time, the pressure reducing valve control is not executed.

  The loss torque calculation unit 115 calculates the loss torque Tlo of the motor M from the experimental data and the like based on the motor target rotation speed N * and the target pump pressure Pp *, and outputs it to the addition unit 134.

  Target rotation speed differential calculation section 116 differentiates motor target rotation speed N * and outputs the result to inertia moment calculation section 123.

  The inertia moment calculation unit 123 calculates a torque necessary for acceleration / deceleration of the angular velocity of the motor M, and outputs the torque to the addition unit 135.

  A rotation speed deviation FB (feedback) calculation unit 117 performs feedback control calculation based on a deviation ΔN (calculated by the addition unit 133) between the target rotation speed N * of the motor M and the actual rotation speed N, and a feedback component ΔN of the rotation speed deviation ΔN. (FB) is output to the adding unit 135.

  The adding unit 134 adds the theoretical torque Tth and the loss torque Tlo of the motor M to calculate the load torque Td and outputs it to the adding unit 135.

  The adder 135 calculates the target torque T * of the motor M by adding the load torque Td of the motor M, the rotational speed deviation FB component ΔN (FB), and the torque necessary for acceleration / deceleration of the angular velocity of the motor M, and converts the current. To the unit 124.

  The current converter 124 converts the target torque T * into a target torque current and outputs it to the motor M to drive the pump P.

[Pressure increase valve control processing block diagram]
FIG. 13 is a block diagram of the pressure increasing valve control process executed in step S70 of FIG. FIG. 13 shows a case where P * Fr> P * fl, the FL wheel booster valve IN / V (FL) is proportionally controlled, and the FR wheel booster valve IN / V (FL) is fully opened.

  The booster valve control process includes a target amount calculator 150, a target pump pressure calculator 161, a wheel cylinder differential pressure deviation FB (feedback) calculator 162, a booster valve target current calculator 163, a current deviation FB (feedback) calculator 164, The booster valve voltage Duty calculation unit 165 is configured.

  Based on the target hydraulic pressure P * (FL, FR) of each wheel FL, FR, the target amount calculation unit 150 sets the target hydraulic pressure P * _H (here, P * _H = P * fr), FL wheel booster valve IN / V (FL) target flow rate Qvfl, and FL wheel target fluid pressure P * fl are output. The target flow rate Qv * fl of the FL wheel wheel cylinder is output as a value obtained by differentiating the value converted from the target hydraulic pressure P * fl into the target fluid amount.

  The FL wheel target hydraulic pressure P * fl is output to the adders 171 and 172. Further, the FL wheel booster valve target flow rate Qvfl is output to the booster valve target current calculator 163.

  The target pump pressure calculation unit 161 calculates the target pump pressure Pp * based on the fully opened target W / C pressure P * _H (P * _H = P * fr), and outputs Pp * to the addition unit 171. .

  The adder 171 calculates the difference between the target pump pressure Pp * and the FL wheel target hydraulic pressure P * fl, and outputs the difference as the FL wheel booster valve target differential pressure ΔPv * fl to the booster valve target current calculator 163.

  The adding unit 172 calculates a deviation ΔPwfl between the target hydraulic pressure P * fl of the FL wheel hydraulic pressure and the actual hydraulic pressure Pfl, and outputs it to the wheel cylinder differential pressure deviation FB calculating unit 162.

  The wheel cylinder differential pressure deviation FB calculation unit 162 feedback-controls the differential pressure deviation ΔPwfl and outputs the feedback component ΔPwfl (FB) of the differential pressure deviation ΔPwfl to the pressure increase valve target current calculation unit 163.

  The booster valve target current calculation unit 163 increases the FL wheel based on the FL wheel booster valve target differential pressure ΔPv * FL, the differential pressure deviation FB component ΔPwfl (FB), and the target flow rate Qvfl of the FL wheel booster valve IN / V (FL). The pressure valve target current I * fl is calculated and output to the pressure increasing valve voltage duty calculation unit 165 and the addition unit 173.

  Adder 173 calculates deviation ΔIfl (FB) between target current I * fl of FL wheel booster valve IN / V (FL) and actual current Ifl, and outputs the result to current deviation FB calculator 164.

  The current deviation FB calculation unit 164 outputs the feedback component ΔIfl (FB) of the current deviation ΔIfl of the FL wheel booster valve IN / V (FL) to the booster valve voltage duty calculation unit 165.

  The booster voltage Duty calculation unit 165 calculates the FL wheel booster valve IN / V (FL) based on the power monitor value from the power monitor 180, the FL wheel booster valve target current I * fl, and the current deviation feedback component ΔIfl (FB). The voltage Duty is calculated and the FL wheel booster valve IN / V (FL) is driven to proportionally control the FL wheel hydraulic pressure Pfl.

[Reducing valve proportional control]
FIG. 14 is a block diagram showing the pressure reducing valve control performed in step S70A.

  The target amount calculation unit 250 outputs each wheel target hydraulic pressure P * to the target differential pressure calculation unit 261 and the deviation calculation unit 272. Further, the fully-opened OUT / V target flow rate Qvout is calculated and output to the OUT / V target current calculation unit 263. The calculation of the target flow rate will be described later.

  When the target hydraulic pressure P * is determined, the differential pressure upstream and downstream of the pressure reducing valve OUT / V may be the target hydraulic pressure P *. Since the downstream of the pressure reducing valve OUT / V is almost atmospheric pressure, the target differential pressure is equal to the target hydraulic pressure. When the target differential pressure is large, it is necessary to reduce the amount of liquid flowing out from the pressure reducing valve OUT / V.Therefore, the target current value I * out is set to a small value, and when the target differential pressure is small, from the pressure reducing valve OUT / V The target current value I * out is set large because there is no need to reduce the amount of liquid flowing out.

(About target flow rate component)
The amount of fluid corresponding to the fluid pressure means that if this amount of fluid flows into the wheel cylinder system, this amount of pressure is generated. Since the change in the liquid amount is a flow rate, the target flow rate per unit time when flowing into the wheel cylinder system is calculated as a differential value of the liquid amount. Therefore, the target amount calculation unit 250 first converts the wheel cylinder target fluid pressure P * into the wheel cylinder target fluid amount using the experimental data of the load stiffness of the caliper and its accessories shown in FIG. Is the wheel cylinder target flow rate.

  Here, the pressure reducing valve OUT / V is an element that controls the amount of outflow from the wheel cylinder system. The target flow rate described above represents the flow rate that flows into the wheel cylinder system, and does not represent the flow rate that flows out from the pressure reducing valve OUT / V. Therefore, when the wheel cylinder target flow rate is positive, the pressure reducing valve OUT / V should not be opened because there is a pressure increase request, so 0 is output to the OUT / V target current calculation unit 263 as the pressure reducing valve target flow rate Qvout.

  On the other hand, when the wheel cylinder target flow rate is negative, since there is a pressure reduction request, the pressure reducing valve target flow rate Qvout is converted to an absolute value of the wheel cylinder target flow rate and a value obtained by reversing the sign is output. As a result, when the wheel cylinder target flow rate is negative, it is necessary to reduce the pressure by causing the brake fluid to flow out from the pressure reducing valve OUT / V, so the target current value to the pressure reducing valve OUT / V increases.

(About motor speed increase control processing)
Here, although the pressure reducing valve target flow rate Qvout is calculated from the target hydraulic pressure P * as described above, when the target rotational speed N * of the pump P is less than the pump rotational speed threshold N0, the pump P rotates excessively. Therefore, it is necessary to discharge the corresponding flow rate from the pressure reducing valve OUT / V. Therefore, the flow rate calculation unit 260 corresponding to the pump excessive rotation reads the target rotation speed N * of the pump and the pump rotation speed threshold N0, and calculates the flow rate Qpover corresponding to the pump excessive rotation by taking this deviation. Since this flow rate Qpover is the flow rate that should flow out from the pressure reducing valve OUT / V, the flow rate Qpover is added to the pressure reducing valve target flow rate Qvout for the pressure reducing valve belonging to the wheel cylinder whose pressure increasing valve is fully open, and the final target The flow rate Qv is output to the OUT / V target current calculation unit 263.

  Here, the pump rotation speed threshold N0 is a variable value. FIG. 15 is a map showing the relationship between the pump discharge pressure equivalent value and the pump rotation speed threshold N0. That is, when the pump discharge pressure is high, the frictional force variation of the gear pump P tends to increase accordingly, and in this case, stabilization is achieved by setting the pump rotation speed threshold N0 high. On the other hand, when the pump discharge pressure is low, the variation in the frictional force of the gear pump P tends to be small. In this case, the pump rotation speed threshold N0 is set low. Thereby, an appropriate pump drive state according to the situation can be achieved.

(About current feedback control)
When the target current value I * out is determined as described above, feedback control of the pressure reducing valve OUT / V is performed by feedback control with the current value detected from the solenoid of the pressure reducing valve OUT / V, and the wheel cylinder pressure is reduced. It is controlled appropriately.

[Time chart during booster valve control]
FIG. 18 is a time chart of comparison of FL and FR wheel hydraulic pressure, pressure increasing valve control mode, pressure reducing valve control mode, and motor rotation speed when the control of Embodiment 1 is performed. In the wheel cylinder hydraulic pressure, the thick solid line indicates the target hydraulic pressure P * fl of the FL wheel, the thick dotted line indicates the target hydraulic pressure P * fr of the FR wheel, and the thin solid line indicates the actual hydraulic pressure. In the booster valve control mode, the solid line represents the control mode of the FL wheel booster valve IN / V (FR), and the dotted line represents the control mode of the FR wheel booster valve IN / V (FL). In the pressure reducing valve control mode, the solid line represents the control mode of the FL wheel pressure reducing valve OUT / V (FL), and the dotted line represents the control mode of the FR wheel pressure reducing valve OUT / V (FR). In the motor rotation speed, the solid line represents the target motor rotation speed N *, and the alternate long and short dash line represents the actual motor rotation speed N.

(Time t1)
At time t1, each wheel pressure increase command is output, and the target hydraulic pressures P * fl and P * fr rise. Since both the FL and FR wheels are boosted and P * fl> P * fr, the high pressure side FL wheel booster valve IN / V (FL) is fully open and the low pressure side FR wheel booster valve IN / V (FR) Is proportional control. At this time, both pressure reducing valves are fully closed.

(Time t2)
At time t2, when the wheel cylinder hydraulic pressure becomes high and a certain level of hydraulic pressure can be secured with a small flow rate, the target motor rotational speed N * decreases (of course, the load torque required for the motor is high). Become). At this time, if the motor rotation speed threshold value N0 falls below, the motor rotation speed threshold value N0 is output instead of the target motor rotation speed N *, and therefore the actual motor rotation speed is substantially the same as the motor rotation speed threshold value N0.

  At the same time, in the FL wheel for which the booster valve full open control is selected, the proportional control is selected as the control of the FL wheel pressure reducing valve OUT / V (FL), and the pump excessive rotation speed is read. As a result, a flow rate corresponding to the excessive rotation speed of the pump flows out from the pressure reducing valve OUT / V (FL). On the other hand, the pressure increase for the FR wheel is continued, and the FR wheel pressure increase valve IN / V (FR) on the low pressure side is set to proportional control. Similarly, full-closed control is continued for the FR wheel pressure reducing valve OUT / V (FR).

  After that, when the FL wheel on the high pressure side shifts to the holding state, the check valve C / V (FL) is actually provided, so that the amount of liquid to be replenished is supplied to the pressure increasing valve IN / V of the FR wheel. Only the amount of brake fluid and the leak from the pump, etc. are sufficient. However, since these leaks are very small, it is necessary to maintain a high-load low-rotation state, so that stable control is very difficult. Therefore, an excessive amount of brake fluid is allowed to flow out from the pressure reducing valve OUT / V (FL) while rotating the motor to some extent to ensure a stable pump discharge pressure.

(Time t3)
At time t3, the FR wheel target hydraulic pressure P * fr reaches the target value, and the FR wheel target hydraulic pressure P * fr is maintained. Since the check valve C / V (FL) is actually provided, the amount of fluid to be replenished can be maintained at the wheel cylinder pressure unless the brake fluid is supplied from the pump P for the leak on the oil passage. Can not.

  However, since these leaks are much smaller than the leaks of the above-described pumps and the like, it is necessary to maintain a high load and low rotation state, so that stable control is very difficult. Therefore, in this scene as well, excessive brake fluid is allowed to flow out from the pressure reducing valve OUT / V (FL) while rotating the motor to some extent to ensure a stable pump discharge pressure.

  In the first embodiment, when both wheels have the same target hydraulic pressure, excessive brake fluid is continuously discharged only from the pressure reducing valve to which the pressure increasing valve that has been fully opened until now belongs. On the other hand, for example, both the pressure increasing valves IN / V of both wheels may be fully opened, and the discharge amount may be equally distributed to the wheels having the same target hydraulic pressure, and discharged from the pressure reducing valves belonging to the respective wheels having the maximum hydraulic pressure. Good. This makes it possible to further secure the volume on the pump discharge side and further stabilize the pump discharge pressure.

  Further, as in the first embodiment, if the configuration includes the check valve C / V, the pump may be driven intermittently to compensate for the leakage. At this time, when the condition that the target hydraulic pressures of both wheels are the same and is maintained is satisfied, the driving of the pump P is stopped, for example, the pump P is driven every predetermined time set in advance. . Thereby, a high load does not always act on the motor M, and power consumption can be suppressed. Note that if the torque of the motor M disappears in a state where a high pressure is applied between the pump P and the check valve, the pump P is rotated in the reverse direction, but there is no particular problem because of the check valve C / V.

(Time t4)
At time t4, for example, when the driver releases the brake pedal and the target hydraulic pressure P * becomes 0 at a time for both wheels, both the pressure increasing valves are fully closed and the pressure reducing valves are both fully open. Further, the target rotational speed of the motor M is also switched to zero.

  As described above, the effects listed below can be obtained in the first embodiment.

  (1) A wheel cylinder W / C provided on each of a plurality of wheels, a pump P for pressurizing brake fluid in the wheel cylinder W / C, a motor M for rotationally driving the pump P, and a wheel cylinder W / C There is a pressure reducing valve OUT / V that reduces the pressure by allowing the brake fluid to flow out by opening the valve, and both the motor M and the pressure reducing valve OUT / V are set so that the hydraulic pressure in the wheel cylinder becomes the target hydraulic pressure P *. It was decided to operate and control.

  Therefore, even if the wheel cylinder pressure is high and the hydraulic pressure gradient is small, the wheel cylinder pressure can be stabilized. Accordingly, it is possible to improve the hydraulic pressure control accuracy in the wheel cylinder and improve the control accuracy of vehicle behavior control and the like.

  (2) The target rotational speed N * of the pump P corresponding to the target hydraulic pressure P * is calculated, and when the target rotational speed N * is lower than the motor rotational speed threshold N0, the target rotational speed N * is converted to the motor rotational speed threshold N0. And excessive brake fluid corresponding to the deviation between the target rotational speed N * and the motor rotational speed threshold N0 is allowed to flow out from the pressure reducing valve OUT / V.

  Therefore, the pump rotational speed can be set to a rotational speed higher than the minimum rotational speed at which the rotational speed control accuracy required for the pump can be ensured, and stable motor rotational speed control can be easily achieved.

  (3) Since there is a check valve C / V that allows only the flow from the pump P side to the wheel cylinder W / C side between the pump P and the wheel cylinder W / C, the brake fluid for the excess number of revolutions of the pump The control to cause the pressure to flow out from the pressure reducing valve OUT / V may be executed only when the pressure in the wheel cylinder is increased.

  That is, since the brake fluid can be held in the wheel cylinder by the check valve C / V, the necessity of driving the pump P is low if the control mode is maintained. At this time, if only the leakage on the oil passage is compensated, the wheel cylinder pressure does not fluctuate only by driving the pump P intermittently. Thus, energy loss can be reduced by appropriately selecting a pump drive scene under reduced pressure, low rotation and high load.

  (4) Fully open control is performed in the wheel cylinder W / C so that the pressure increasing valve IN / V corresponding to the wheel cylinder W / C corresponding to the maximum target hydraulic pressure P * _H of each wheel is fully opened. It was decided to control the hydraulic pressure in the wheel cylinder corresponding to the pressure increasing valve IN / V that was fully opened by the control.

  The wheel cylinder pressure to which the pressure increasing valve IN / V that is fully open belongs is feedback controlled by the pump P. At this time, if the pump rotational speed is defined as the pump rotational speed threshold N0, feedback control cannot be performed. On the other hand, feedback control can be achieved with the pressure reducing valve, and stable control can be achieved.

  (5) Since the pump P of the first embodiment is a gear pump, the volume of each pump chamber is small, so that fine fluid pressure control can be achieved by motor rotation speed control.

  (6) The number of rotations of the motor M is detected by a Hall element. In order to improve the accuracy of the rotational speed control at the time of low rotation of the motor M, it is common to provide an expensive sensor with high resolution such as a resolver. On the other hand, even a low-cost sensor with low resolution such as a Hall element can be controlled with an inexpensive configuration because the motor speed is driven only at a speed higher than the motor speed threshold N0. The accuracy can be improved.

  (7) Since the pressure reducing valve OUT / V is an electromagnetic proportional control valve capable of continuously controlling the flow rate, the control accuracy can be improved.

  A second embodiment will be described with reference to FIGS. The basic configuration is the same as that of the first embodiment. In the first embodiment, only the front wheels are hydraulic brake-by-wire controlled, but in the second embodiment, all the four wheels are hydraulic brake-by-wire controlled.

  FIG. 19 is a system configuration diagram in the second embodiment, and FIG. 20 is a hydraulic circuit diagram. The brake hydraulic device is a hydraulic brake-by-wire system that normally increases the pressure of the wheel cylinders W / C (FL to RR) of all four wheels by one pump Main / P. The master cylinder M / C is a so-called tandem type, and is connected to the FL wheel wheel cylinder W / C (FL, FR) by manual circuits A (FL), A (FR).

  The master cylinder M / C is connected to the reservoir RSV, and each solenoid valve is driven by the control unit CU. The pump, which is the hydraulic pressure source, is provided with a main pump Main / P for normal use and a sub pump Sub / P for emergency use, which are driven by the main motor Main / M and the sub motor Sub / M based on commands from the control unit CU. The

  The manual circuit A (FL, FR) is provided with a shut-off valve S.OFF/V (FL, FR), which is a normally open solenoid valve (ON / OFF valve), and the first and second master cylinders M / C / M / C2 and FL / FR wheel cylinder W / C (FL, FR) are connected / disconnected.

  A stroke simulator S / Sim is provided on the manual circuit A (FR) and between the first master cylinder M / C and the shut-off valve S.OFF/V (FR). This stroke simulator S / Sim is connected to the manual circuit A (FR) via a cancel valve Can / V which is a normally closed solenoid valve (ON / OFF valve).

  When the FR shut-off valve S.OFF/V (FR) is closed and the cancel valve Can / V is opened, the hydraulic oil in the first master cylinder M / C is released as the brake pedal BP is depressed. Introduced to the stroke simulator S / Sim to secure the pedal stroke.

  The discharge sides of the main and sub pumps Main / P and Sub / P are connected to a pressure-increasing circuit C and connected to each wheel cylinder W / C (FL to RR) at a connection point I (FL to RR). On the other hand, the suction side of each pump Main / P, Sub / P is connected to a decompression circuit B.

  On this pressure increasing circuit C, pressure increasing valves IN / V (FL to RR), which are normally closed solenoid valves (proportional valves), are provided. Each pump Main / P, Sub / P and each wheel cylinder W / C (FL Switch between communication / blocking (~ RR).

  Each wheel cylinder W / C (FL to RR) is connected to the decompression circuit B at the connection point I (FL to RR). On this pressure reducing circuit B, a pressure reducing valve OUT / V (FL to RR) which is a normally closed solenoid valve (proportional valve) is provided, and communication / blocking between each wheel cylinder W / C (FL to RR) and the reservoir RSV is provided. Switch.

  A check valve C / V is provided on the discharge side of each pump Main / P and Sub / P, and hydraulic fluid flows backward from the pressure increasing circuit C to the pressure reducing circuit B via each pump Main / P and Sub / P. Avoid that. Further, the pressure increasing circuit C and the pressure reducing circuit B are connected via a relief valve Ref / V, and when the pressure in the pressure increasing circuit C becomes equal to or higher than a specified value, hydraulic fluid is released to the pressure reducing circuit B.

  The first and second master cylinder pressure sensors MC / Sen1 on the manual circuit A (FL, FR) and between the shutoff valve S.OFF/V (FL, FR) and the master cylinder M / C, respectively. , 2 are provided, and each wheel cylinder W / C (FL to RR) is provided with a hydraulic pressure sensor WC / Sen (FL to RR). A pump discharge pressure sensor P / Sen is provided on the pressure increasing circuit C.

  The control unit CU receives the detected first and second master cylinder pressures Pm1 and Pm2, the hydraulic pressures P (FL to RR), and the detection value of the stroke sensor S / Sen that detects the stroke of the brake pedal BP. The

  Based on these detected values, the control unit CU calculates the target hydraulic pressure P * (FL to RR) of each wheel FL to RR, and each motor Main / M, Sub / M and the pressure increasing valve IN / V (FL to RR). ), Driving the pressure reducing valve OUT / V (FL to RR). During normal braking, the shut-off valve S.OFF/V (FL, FR) is closed and the cancel valve Can / V is opened.

  The control unit CU also compares the target hydraulic pressure P * (FL to RR) and actual hydraulic pressure P (FL to RR) of each wheel cylinder W / C (FL to RR) to When the hydraulic pressure shows an abnormal response, an abnormal signal is output to the warning lamp WL. In addition, the wheel speed VSP is input to the control unit CU to determine whether the vehicle is running / stopped.

[Brake control]
(Normal pressure increase)
During normal pressure increase, cancel valve Can / V is opened and shut-off valve S. OFF / V (FL, FR) is shut off and the driver depresses the brake pedal BP by the stroke sensor S / Sen. Based on this detection value, each wheel cylinder W / C (FL to RR) is detected in the control unit CU. The target hydraulic pressure P * (FL to RR) is calculated.

  Further, the control unit CU drives the main motor Main / M or the sub motor Sub / M to apply the discharge pressure to the pressure increasing circuit C. Furthermore, each booster valve IN / V (FL to RR) is driven according to the calculated target hydraulic pressure P * (FL to RR), and hydraulic oil is supplied to each wheel cylinder W / C (FL to RR). Get braking force.

(At reduced pressure)
At the time of decompression, each decompression valve OUT / V (FL to RR) is driven by the control unit CU, and hydraulic oil is discharged from each wheel cylinder W / C (FL to RR) to the reservoir RSV via the decompression circuit B. .

(When holding)
At the time of holding, the specified pressure increasing valve IN / V (FL to RR) and each pressure reducing valve OUT / V (FL to RR) are closed, each wheel cylinder W / C (FL to RR) and pressure increasing and pressure reducing circuit Shut off C and B. The booster valve that performs full-opening control of the booster valve described later is not closed.

(Manual brake)
Normally open shut-off valve in case of system failure. OFF / V (FL, FR) is opened, each normally closed pressure increase valve IN / V (FL to RR) and FL, FR wheel pressure reducing valve OUT / V (FL, FR) are closed, RL, RR The ring pressure reducing valve OUT / V (RL, RR) is opened.

  As a result, the master cylinder M / C communicates with the FL and FR wheel cylinders W / C (FL, FR), and a manual brake is secured. On the other hand, the RL and RR wheel wheel cylinder pressures Prl and Prr are substantially zero to prevent locking.

  Even when all four wheels are controlled by the hydraulic unit HU, as in the first embodiment, the booster valves IN / V belonging to the wheel cylinders of the wheels with the maximum target hydraulic pressure P * are fully opened, and the other wheels By using proportional control for the pressure increasing valve IN / V, the same effects as those of the first embodiment can be obtained.

  In addition, the hydraulic pressure feedback control by the pump P is performed on the wheel cylinder W / C to which the pressure increasing valve IN / V that is fully opened control belongs. At this time, when the target rotational speed N * of the pump P falls below the pump rotational speed threshold value N0, the pump P is controlled to the pump rotational speed threshold value N0 as in the first embodiment, and the excess brake fluid is reduced to the pressure reducing valve OUT / V. By discharging from the above, the same effect as the first embodiment can be obtained.

[Other Examples]
As mentioned above, although this invention has been demonstrated based on Example 1, 2, this invention is not restricted to the said structure, Even if it applies to another structure, the same effect can be acquired.

  For example, in the second embodiment, the hydraulic pressure of the four wheels is controlled by one hydraulic pressure unit HU. However, even if the configuration includes separate hydraulic pressure units HU on the front wheel side and the rear wheel side, The same control as that in the first embodiment can be executed in the hydraulic unit HU.

  Also, rather than the front wheel side and the rear wheel side, a so-called X-pipe configuration comprising a hydraulic unit HU with the right front wheel and the left rear wheel as a set, and a hydraulic unit HU with the left front wheel and the right rear wheel as a set. Even in such a case, the same operational effects as those of the first embodiment can be obtained in each hydraulic unit HU.

  In the first embodiment, the booster valve on the side where the target hydraulic pressure is higher among the FL and FR wheels is fully opened. However, the booster valve during pressure increase may be fully opened. If the pressure is increasing on both the low pressure side and the high pressure side, the pressure increasing valve on the side with the higher target hydraulic pressure is fully opened. Further, when all the wheels are being held or depressurized, the pressure increasing valve on the side with the higher target hydraulic pressure is fully opened. When the high pressure side is in the holding / decompression state and the low pressure side is in the pressure increasing state, the low pressure side pressure increasing valve is fully opened and the high pressure side pressure increasing valve is fully closed.

  In Example 1, since the pressure increasing valve on the side where the target hydraulic pressure is high is fully opened, the check valve C / V is not necessarily required. However, when the pressure increasing valve during pressure increasing is fully opened as described above, It should be noted that a check valve C / V is necessary. Thus, by fully opening the pressure increasing valve during pressure increase, the pump P does not necessarily need to discharge the maximum target hydraulic pressure, and energy loss can be reduced.

  In the first embodiment, among the FL and FR wheels, the booster valve on the side where the target hydraulic pressure is high is always fully opened and the other wheels are controlled proportionally. However, when the side where the target hydraulic pressure is high is held or reduced, Since it is not necessary to further increase the pressure of the wheel, the low pressure side pressure increasing valve may be fully opened only when the pressure on the lower target hydraulic pressure is being increased. Also in this case, it is not necessary for the pump P to discharge the maximum target hydraulic pressure, and energy loss can be reduced.

  Further, in Examples 1 and 2, the pressure increasing valve IN / V is a normally open valve, but may be a normally closed valve. In this case, the wheel cylinder pressure can be maintained by de-energization without providing the check valve C / V. However, there is a concern that the energization amount may increase because one of the pressure increasing valves IN / V is fully opened during braking. Therefore, it is desirable to estimate and monitor the temperature and heat generation state of the pressure increasing valve IN / V.

1 is a system configuration diagram of a vehicle to which a brake control device of Embodiment 1 is applied. It is a figure which shows the structure of the hydraulic circuit and control unit in the system of Example 1. FIG. FIG. 3 is a cross-sectional view illustrating a configuration of a pump P disposed in the hydraulic unit HU according to the first embodiment. FIG. 3 is an external view illustrating a configuration of a pump P disposed in the hydraulic unit HU according to the first embodiment. It is a flowchart which shows the flow of the brake-by-wire control process of Example 1. 3 is a flowchart illustrating a flow of target hydraulic pressure mode determination processing according to the first embodiment. 3 is a correction table for correcting an increase / decrease threshold value according to the first embodiment. 3 is a flowchart illustrating a flow of W / C hydraulic pressure control mode determination processing according to the first embodiment. 3 is a flowchart illustrating a flow of a pressure increasing valve control mode determination process according to the first embodiment. 3 is a flowchart illustrating a flow of a pressure reducing valve control mode determination process according to the first embodiment. FIG. 3 is a block diagram of pump control processing according to the first embodiment. It is experimental data of the load rigidity of the caliper and its accessories, which is the relationship between the fluid pressure and the fluid volume. It is a block diagram of the pressure increase valve control process of Example 1. FIG. It is a block diagram showing pressure-reducing valve control including pump rotation speed raising control of Example 1. 3 is a map showing a relationship between a pump discharge pressure equivalent value and a pump rotation speed threshold value according to the first embodiment. It is a figure showing the oil film pressure distribution of a journal bearing. It is a figure showing the relationship between the rotation speed of the gear pump of Example 1, and frictional force. 6 is a time chart of comparison of FL, FR wheel hydraulic pressure, pressure-increasing valve control mode, pressure-reducing valve control mode, and motor rotation speed when the control of Example 1 is performed. It is a system configuration | structure figure of the vehicle to which the brake control apparatus of Example 2 was applied. It is a figure which shows the structure of the hydraulic circuit and control unit in the system of Example 2. FIG.

Explanation of symbols

3a Pump assembly 10A Drive shaft 10a Drive gear 11A Drive shaft 11a Drive gear 12a Seal block 13a Suction passage 16a High-pressure chamber 20a Bearing
EMB Electric caliper
RCU Left and right rear controller
MGB regenerative braking device
BP brake pedal
C / V check valve
Can / V cancel valve
CU control unit
HU hydraulic unit
IN / V Booster Valve M Motor
M / C master cylinder
MC / Sen1, MC / Sen2 Master cylinder pressure sensor
OUT / V Pressure reducing valve P Pump (gear pump)
P / Sen pump discharge pressure sensor
Ref / V relief valve
RSV reservoir
S Shutoff valve
S / Sen Stroke sensor
S / Sim stroke simulator
W / C wheel cylinder
WC / Sen hydraulic pressure sensor
WL warning lamp

Claims (4)

A wheel cylinder provided on each of a plurality of wheels;
A pump for pressurizing the brake fluid in the wheel cylinder;
A motor for rotating the pump;
A pressure reducing valve for reducing the pressure by allowing the brake fluid in the wheel cylinder to flow out by opening the valve;
Control means for operating and controlling both the motor and the pressure reducing valve so that the hydraulic pressure in the wheel cylinder becomes a target hydraulic pressure;
A brake fluid pressure control device comprising: In the brake fluid pressure control device according to claim 1,
Providing a target rotational speed calculating means for calculating a target rotational speed of the pump according to the target hydraulic pressure;
When the target rotational speed is lower than the predetermined rotational speed, the control means sets the target rotational speed to the predetermined rotational speed, and an excess brake fluid according to a deviation between the target rotational speed and the predetermined rotational speed. Is caused to flow out of the pressure reducing valve. In the brake fluid pressure control device according to claim 1 or 2,
Between the pump and the wheel cylinder, there is a check valve that allows only the flow from the pump side to the wheel cylinder side,
The brake fluid pressure control device is characterized in that the control means is executed when the pressure in the wheel cylinder is increased. In the brake fluid pressure control device according to any one of claims 1 to 3,
A pressure increasing valve for increasing the pressure by allowing brake fluid to flow into the wheel cylinder by opening the valve;
A fully open control means for fully opening the pressure increasing valve corresponding to the wheel cylinder corresponding to the maximum value of the target hydraulic pressure of each wheel;
Provided,
The brake hydraulic pressure control device, wherein the control means controls the hydraulic pressure in the wheel cylinder corresponding to the pressure increasing valve fully opened by the full opening control means.

Images