JP7543785B2 - Vehicle brake control device - Google Patents

Description

This disclosure relates to a vehicle braking control device.

Patent Document 1 describes the purpose of providing a brake fluid pressure control device that can "prevent the spool from sticking and vibrating by constantly monitoring the operating state of the spool of the fluid pressure control valve and adjusting the dither, and can also deal with individual differences in the fluid pressure control valve," and that "in response to the operation of the brake pedal 1, the controller 6 operates the fluid pressure control valve 9 to control the brake fluid pressure supplied to the brake wheel cylinder 10. In controlling this brake fluid pressure, the controller 6 determines the micro-vibration amplitude of the brake fluid pressure in the brake wheel cylinder 10 from the fluid pressure detection value supplied from the brake fluid pressure sensor 11, and compares the micro-vibration amplitude with a predetermined threshold value to estimate the operating state of the spool 9c. Based on the estimation result, a dither is set to eliminate the sticking or vibration state of the spool 9c, and control is performed so that a drive current superimposed on the target current is supplied to the proportional solenoid 9d."

In the "dither control" described in Patent Document 1, a dither waveform (for example, an upward dither rising to the right with a constant gradient and a downward dither falling to the right repeated at a predetermined dither cycle) is superimposed on the target current of the solenoid valve, so that the brake fluid pressure supplied to the wheel cylinder is oscillated with a small amplitude. This suppresses the sticking and vibration of the solenoid valve spool, thereby avoiding a decrease in responsiveness of the brake operation caused by this sticking, and an unnecessary decrease in fluid pressure in the brake fluid pressure source caused by vibration of the fluid pressure control valve spool.

Patent document 2 states that the objective of the device is to "provide a brake control device capable of suppressing vibrations that may occur in a solenoid valve," and that "the brake control device includes a wheel cylinder that is supplied with brake fluid to apply a braking force to the wheels, a solenoid valve that is connected to the wheel cylinder via a flow path and whose opening is adjusted by current control, and a control unit that controls the supply of a pulse signal that is pulse-width modulated to the solenoid valve. When the control unit predicts the occurrence of self-excited vibration in the solenoid valve, it increases the frequency of the pulse signal that is supplied to the solenoid valve above a predetermined normal frequency." If the frequency of the control signal input to the solenoid valve is increased as in the device of Patent document 2, the control unit (drive circuit provided in the control unit) that generates the control signal will generate more heat, and there is a concern that the durability of the device will decrease.

The applicant has developed a vehicle hydraulic pressure control device as described in Patent Document 3, in order to solve the problems with the device described in Patent Document 2 and suppress the occurrence of abnormal noise caused by self-excited vibration generated in the solenoid valve while suppressing a decrease in the durability of the device. In the self-excited vibration suppression process of Patent Document 3, when a predetermined suppression control permission condition is met while both the supply pump of the pressure adjustment unit and the differential pressure adjustment valve 321 are operating, a valve body that is separated from the valve seat is brought into contact with the valve seat. Then, when the valve body comes into contact with the valve seat, the valve body is moved away from the valve seat.

In solenoid valves, in addition to the problem of spool sticking and vibration (problems in Patent Document 1), there is also the problem of hydraulic pressure fluctuations (problems in Patent Documents 2 and 3). In addition to the self-excited vibrations mentioned above, some of these hydraulic pressure fluctuations are caused by changes in the fluid force (force caused by the flow of fluid) Fb of the brake fluid BF. In particular, these hydraulic pressure fluctuations are highly likely to occur in configurations in which linear solenoid valves UI (also called "proportional valves") are used in normally open inlet valves provided to independently adjust the actual brake hydraulic pressure Pw in each wheel cylinder CW.

This will be explained in detail below with reference to the schematic diagram of FIG. 5. The linear inlet valve UI is disposed between the wheel cylinder provided on each wheel WH and the pressurizing unit. Here, the "pressurizing unit" corresponds to the master cylinder CM in a configuration in which the brake-by-wire type is not adopted, and corresponds to a fluid unit powered by an electric motor in a configuration in which the brake-by-wire type is adopted. The linear inlet valve UI is supplied with brake fluid BF from a fluid pump HP driven by an electric motor MT (the flow of brake fluid BF is shown by the upward arrow). The inlet valve UI is then used to quietly and smoothly control (adjust) the brake fluid pressure Pw for each wheel independently in antilock brake control, vehicle stability control, traction control, etc.

The left diagram (a) of the central axis Jv of the inlet valve UI shows a state in which the normally open inlet valve UI is fully energized (i.e., energized enough to maintain a closed state), with the spherical tip Vt of the valve body VT abutting against the valve seat Vz (conical surface) provided on the retaining member HJ, and the inlet valve UI being closed. On the other hand, the right diagram (b) of the central axis Jv shows a state in which the amount of energization has been reduced from full energization, resulting in a gap Li (called the "valve opening amount") between the tip Vt and the valve seat Vz. In the situation in the right diagram (b), the flow of brake fluid BF is restricted by the valve opening amount Li. Due to the orifice effect according to the valve opening amount Li, a hydraulic pressure difference (also called "differential pressure") Qa occurs between the hydraulic pressure (called "first hydraulic pressure") Pa on the upstream side of the valve seat Vz and the hydraulic pressure (called "second hydraulic pressure") Pb on the downstream side of the valve seat Vz in the flow of brake fluid BF. Note that since this situation is "Pa ≧ Pb", it can also be said that "due to the valve opening amount Li (i.e., the orifice), the first hydraulic pressure Pa is increased by the hydraulic pressure difference Qa from the second hydraulic pressure Pb (i.e., "Pa = Pb + Qa").

The valve opening amount Li is determined by the balance between the fluid force Fb of the brake fluid BF acting on the valve body VT (more precisely, the force "Fb + Fs" obtained by adding the elastic force Fs to the fluid force Fb) and the thrust force Fa (the suction force generated by the solenoid SD) generated by the solenoid SD (coil CL + plunger PL). Since the fluid force Fb fluctuates slightly due to disturbances such as the pulsation of the fluid pump HP, when the amount of electricity supplied to the solenoid SD is constant, the valve opening amount Li also changes slightly. For example, in order to maintain a constant hydraulic pressure difference Qa, the suction force Fa of the solenoid SD must be adjusted with high response in response to changes in the fluid force Fb, the valve body VT must be returned to the required position, and the valve opening amount Li must be appropriately adjusted. In particular, when a large flow rate and a high hydraulic pressure difference Qa (target differential pressure Qt) are required, the valve opening amount Li is small, so there is a high demand for high response from the solenoid SD.

In order to improve the responsiveness of the solenoid SD, its output needs to be increased. However, increasing the output of the solenoid SD leads to an increase in size of the inlet valve UI (particularly the solenoid SD) and the drive circuit DR that drives the inlet valve UI. For this reason, it is desirable for the vehicle brake control device SC to be able to appropriately control the hydraulic pressure difference Qa (and thus the brake hydraulic pressure Pw) without increasing the size of the inlet valve UI, etc.

JP 11-099918 A JP 2011-084147 A JP 2017-065664 A

The object of the present invention is to provide a vehicle brake control device that can reduce hydraulic pressure fluctuations without increasing the size of the linear inlet valve device.

The vehicle brake control device of the present invention adjusts the brake fluid pressure (Pw) in the wheel cylinder (CW) of the vehicle to control the braking force (Fx) of the wheels (WH), and includes a "fluid pump (HP) driven by an electric motor (MT)", a "linear inlet valve (UI) composed of a solenoid (SD), a valve body (VT) driven by the solenoid (SD), and a valve seat (Vz) that can abut against the valve body (VT), in which a fluid force (Fb) of the brake fluid (BF) discharged by the fluid pump (HP) acts on the valve body (VT)", and a "controller (ECU) that adjusts the energization state of the solenoid (SD) to control the inlet valve (UI)".

In the vehicle braking control device of the present invention, the controller (ECU) calculates the standard of the energized state as a standard amount of energization (Is), periodically increases the standard amount of energization (Is) only in the direction in which the valve body (VT) approaches the valve seat (Vz) to calculate a target amount of energization (It), and controls the energized state based on the target amount of energization (It).

A fluid force Fb of the brake fluid BF discharged by the fluid pump HP acts on the valve body VT. When the fluid force Fb increases due to the pulsation of the fluid pump HP, the valve body VT is moved in a direction away from the valve seat Vz (i.e., in the direction in which the fluid force Fb and the elastic force Fs act), and the brake fluid pressure Pw increases. One way to suppress this increase in the brake fluid pressure Pw (i.e., a way to reduce the fluid pressure fluctuations caused by the fluid force Fb) is to instantly move the valve body VT closer to the valve seat Vz. However, to achieve this, it is necessary to increase the output of the solenoid SD, which raises concerns about the device becoming larger.

According to the above configuration, the standard current flow amount Is, which is the standard for the current flow state of the solenoid SD, is periodically increased only in the direction in which the valve body VT approaches the valve seat Vz (i.e., the direction in which the valve body VT opposes the elastic force Fs and the fluid force Fb, and in which the suction force Fa increases). In other words, the solenoid SD is energized in anticipation of the increase in the fluid force Fb caused by the pulsation of the fluid pump HP. This makes it possible to reduce hydraulic pressure fluctuations caused by the fluid force Fb without increasing the size of the devices related to the inlet valve UI (solenoid SD, drive circuit DR, etc.).

In the vehicle brake control device according to the present invention, the inlet valve (UI) adjusts the hydraulic pressure difference (Qt, Qa) between a first hydraulic pressure (Pa) which is the hydraulic pressure upstream of the valve seat (Vz) and a second hydraulic pressure (Pb) which is the hydraulic pressure downstream of the valve seat (Vz) in the flow of the brake fluid (BF) discharged by the fluid pump (HP). The controller (ECU) increases the adjustment current amount (Ic) by which the standard current amount (Is) is increased to the target current amount (It) as the hydraulic pressure difference (Qt, Qa) increases. The controller (ECU) also permits an increase adjustment in which the standard current amount (Is) is increased to calculate the target current amount (It) when the brake fluid pressure (Pw) is increased, and prohibits it when the brake fluid pressure (Pw) is decreased.

The fluid force Fb exerts a large influence on the valve body VT when the pressure difference is large. Furthermore, when the inlet valve UI is closed, the influence of the fluid force Fb can be ignored. With the above configuration, it is possible to adjust the amount of current passing through the inlet valve UI just enough to counteract the fluid force Fb, but only when necessary.

1 is an overall configuration diagram for explaining an embodiment of a vehicle brake control device SC; FIG. 4 is a cross-sectional view illustrating an inlet valve UI. FIG. 4 is a functional block diagram for explaining energization control of an inlet valve UI. FIG. 4 is a time series diagram for explaining a mode selection process in antilock brake control. FIG. 1 is a schematic diagram for explaining a problem.

<Symbols for components, etc., and suffixes at the end of symbols>
In the following description, components, elements, signals, characteristics, etc. with the same symbol, such as "CW", have the same function. The suffixes "1" and "2" attached to the symbols relating to two braking systems are generic symbols indicating which system they relate to, with "1" indicating one braking system (also called "first braking system BK1") and "2" indicating the other braking system (also called "second braking system BK2"). For example, in the two pressure chambers (also called "hydraulic pressure chambers") of a tandem type master cylinder CM, the one connected to the first braking system BK1 is the first hydraulic pressure chamber Rm1, and the one connected to the second braking system BK2 is the second hydraulic pressure chamber Rm2. The suffixes "1" and "2" may be omitted. When the suffixes "1" and "2" are omitted, the symbol represents a generic name. For example, "Rm" represents a hydraulic pressure chamber.

In the connection path HS, the side closer to the master cylinder CM (or the side farther from the wheel cylinder CW) is referred to as the "upper part," and the side closer to the wheel cylinder CW is referred to as the "lower part." In the flow of the brake fluid BF discharged from the fluid pump HP, the side closer to the fluid pump HP (particularly the discharge part Ht) is referred to as the upstream side (upstream part), and the farther side is referred to as the downstream side (downstream part).

<Embodiments of the vehicle brake control device SC>
An embodiment of the brake control device SC according to the present invention will be described with reference to the overall configuration diagram of FIG. 1. In this embodiment, in two fluid paths (first and second brake systems BK1 and BK2), in the first brake system BK1, the first hydraulic chamber Rm1 is connected to the wheel cylinders of the right front wheel and the left rear wheel (referred to as the "first wheel cylinder CW1"). In the second brake system BK2, the second hydraulic chamber Rm2 is connected to the wheel cylinders of the left front wheel and the right rear wheel (referred to as the "second wheel cylinder CW2"). That is, a diagonal type (also called "X type") is adopted as the two fluid paths. Here, the "fluid path" is a path for moving the brake fluid BF, which is a working fluid, and corresponds to the brake piping, the flow path of the fluid unit HU, the hose, etc.

A vehicle equipped with a brake control device SC is provided with a brake operating member BP, a rotating member KT, a wheel cylinder CW, a master reservoir RV, a master cylinder CM, a brake operating amount sensor BA, a wheel speed sensor VW, a steering angle sensor SA, a yaw rate sensor YR, a deceleration sensor GX, and a lateral acceleration sensor GY.

The brake operating member (e.g., brake pedal) BP is a member that the driver operates to decelerate the vehicle. By operating the brake operating member BP, the braking torque Tq of the wheel WH is adjusted and a braking force Fx is generated on the wheel WH. Specifically, a rotating member (e.g., brake disc) KT is fixed to the wheel WH of the vehicle. A brake caliper is then positioned to clamp the rotating member KT.

The brake caliper is provided with a wheel cylinder CW. When the pressure Pw of the brake fluid BF in the wheel cylinder CW ("wheel cylinder fluid pressure", also called "braking fluid pressure") is increased, a friction member (e.g., a brake pad) is pressed against the rotating member KT. Since the rotating member KT and the wheel WH are fixed so that they rotate together, the frictional force that is generated at this time generates a braking torque Tq on the wheel WH. The braking torque Tq then generates a braking force Fx on the wheel WH.

The master reservoir (also called the "atmospheric pressure reservoir") RV is a tank for hydraulic fluid, and brake fluid BF is stored inside. The interior of the master cylinder CM is divided into two hydraulic chambers Rm1, Rm2 (=Rm) by two pistons PG, PH. In other words, a tandem type is used for the master cylinder CM. The primary piston PG and secondary piston PH are moved in conjunction with the brake operating member BP. When the brake operating member BP is not operated, the hydraulic chamber Rm of the master cylinder CM and the master reservoir RV are in communication. When there is a shortage of brake fluid BF in the first and second brake systems BK1, BK2 (=BK), brake fluid BF is replenished from the master reservoir RV to the hydraulic chamber Rm.

When the brake operating member BP is operated, the pistons PG, PH in the master cylinder CM are moved forward, and communication between the hydraulic chamber Rm and the master reservoir RV is blocked. When the operation of the brake operating member BP is further increased, the pistons PG, PH are moved further forward, the volume of the hydraulic chamber Rm decreases, and the brake fluid (working fluid) BF is discharged (pressurized) from the master cylinder CM. When the operation of the brake operating member BP is decreased, the pistons PG, PH are moved backward, opposite to the forward direction, the volume of the hydraulic chamber Rm increases, and the brake fluid BF is returned toward the master cylinder CM.

The first and second hydraulic chambers Rm1, Rm2 (=Rm) of the tandem master cylinder CM and the first and second wheel cylinders CW1, CW2 (=CW) are connected by the first and second connection passages HS1, HS2 (=HS). The connection passage HS is a fluid passage that connects the master cylinder CM and the wheel cylinder CW. The first and second connection passages HS1, HS2 are branched into two at the bottom of the first and second pressure regulating valves UM1, UM2 (=UM) and are connected to the first and second wheel cylinders CW1, CW2.

The braking operation amount sensor BA detects the amount of operation Ba of the brake operating member (brake pedal) BP by the driver. Specifically, at least one of the master cylinder hydraulic pressure sensor PM (=PM1, PM2) that detects the hydraulic pressure (called "master cylinder hydraulic pressure") Pm (=Pm1, Pm2) in the hydraulic pressure chamber Rm, the operation displacement sensor SP that detects the operation displacement Sp of the brake operating member BP, and the operation force sensor FP that detects the operation force Fp of the brake operating member BP is adopted as the braking operation amount sensor BA. In other words, the operation amount sensor BA is a collective term for the master cylinder hydraulic pressure sensor PM, the operation displacement sensor SP, and the operation force sensor FP, and the braking operation amount Ba is a collective term for the master cylinder hydraulic pressure Pm, the operation displacement Sp, and the operation force Fp.

The wheel speed sensor VW detects the wheel speed Vw, which is the rotational speed of each wheel WH. For example, the wheel speed sensor VW is provided on the wheel WH. The wheel speed Vw is used to calculate the vehicle speed Vx.

The steering angle sensor SA detects the steering angle Sa of the steering wheel SW. The steering wheel SW is a member that the driver operates to turn the vehicle. By operating the steering wheel SW, the steering angle of the steered wheels (e.g., the front wheels) is adjusted, and a turning force Fy is generated on the steered wheels. For example, the steering angle sensor SA is provided on the steering wheel SW.

The yaw rate sensor YR detects the actual yaw rate Yr of the vehicle. The deceleration sensor GX detects the actual deceleration Gx of the vehicle. The lateral acceleration sensor GY detects the actual lateral acceleration Gy of the vehicle.

The detection values (Vw, etc.) of each sensor (VW, etc.) are input to the controller ECU. Based on these signals, the controller ECU adjusts the brake fluid pressure Pw independently in the wheel cylinder CW of each wheel WH, and performs anti-lock brake control to suppress the tendency of the wheel WH to lock, vehicle stability control to ensure vehicle stability, etc. This independent adjustment of the brake fluid pressure Pw for each wheel WH is called "independent hydraulic control." In other words, independent hydraulic control is a general term for anti-lock brake control (also known as "ABS"), vehicle stability control (also known as "ESC", also known as "sideslip prevention control"), traction control, etc.

<Driving assistance system>
A vehicle is provided with a driving assistance system that automatically stops the vehicle via a braking control device SC in place of or in assistance to the driver. This control is called "automatic braking control". The driving assistance system includes a distance sensor OB and a driving assistance controller ECJ. The distance sensor OB detects a distance (relative distance) Ob between an object (another vehicle, a fixed object, a person, a bicycle, a stop line, a sign, a signal, etc.) in front of the vehicle and the vehicle itself. For example, an image sensor, a radar sensor, an ultrasonic sensor, etc. are used as the distance sensor OB. Alternatively, information from a GPS (Global Positioning System) installed in the vehicle may be referred to map information to calculate the relative distance Ob. The relative distance Ob is input to the driving assistance controller ECJ.

The driving assistance controller ECJ calculates the required deceleration Gs for automatically stopping the vehicle based on the relative distance Ob. The required deceleration Gs is a target value of the vehicle deceleration for executing automatic braking control. The required deceleration Gs is transmitted to the braking controller ECU of the braking control device SC via the communication bus BS.

<Brake control device SC>
The brake control system SC is composed of a hydraulic unit HU and a brake controller ECU (also simply referred to as a "controller").

The fluid unit HU is provided in the first and second connection paths HS1, HS2 (fluid paths). That is, the first and second hydraulic chambers Rm1, Rm2 are connected to the first and second wheel cylinders CW1, CW2 via the fluid unit HU. The fluid unit HU includes the first and second pressure regulating valves UM1, UM2, the first and second master cylinder hydraulic pressure sensors PM1, PM2, the first and second hydraulic pumps HP1, HP2, the electric motor MT, the first and second pressure regulating reservoirs RC1, RC2, the first and second regulated hydraulic pressure sensors PP1, PP2, the first and second inlet valves UI1, UI2, and the first and second outlet valves VO1, VO2.

The first and second connection paths HS1 and HS2 (=HS) are provided with first and second pressure regulating valves UM1 and UM2 (=UM). The pressure regulating valve UM is a normally open proportional valve whose valve opening amount (lift amount) Lm is continuously controlled according to the amount of electricity (current value). The proportional valve is also called a "differential pressure valve" because it can adjust the hydraulic pressure difference (differential pressure) before and after it in the fluid path.

First and second master cylinder pressure sensors PM1, PM2 (=PM) are provided above the first and second pressure regulating valves UM1, UM2 (=UM) to detect the hydraulic pressures (master cylinder hydraulic pressures) Pm1, Pm2 (=Pm) in the first and second hydraulic chambers Rm1, Rm2. The master cylinder hydraulic pressure sensor PM corresponds to the operation amount sensor BA, and the master cylinder hydraulic pressure Pm corresponds to the operation amount Ba. Note that since the first and second master cylinder hydraulic pressures Pm1, Pm2 are substantially the same, it is possible to omit one of the first and second master cylinder hydraulic pressure sensors PM1, PM2.

First and second return paths HK1, HK2 (=HK) are provided to connect the upper part of the pressure regulating valve UM to the lower part of the pressure regulating valve UM. First and second fluid pumps HP1, HP2 (=HP) and first and second pressure regulating reservoirs RC1, RC2 (=RC) are provided in the return path HK (fluid path).

The fluid pump HP is driven by an electric motor MT. Specifically, the fluid pump HP is rotated by the electric motor MT driven by the controller ECU. The fluid pump HP sucks in the brake fluid BF from the upper part of the pressure regulating valve UM (a part in the connection path HS between the master cylinder CM and the pressure regulating valve UM) and discharges the brake fluid BF to the lower part of the pressure regulating valve UM (a part in the connection path HS between the pressure regulating valve UM and the wheel cylinder CW). When the electric motor MT is driven, the first and second refluxes KN1 and KN2 (=KN) of the brake fluid BF are generated in the return path HK as shown by the dashed arrows (flow of "HP → UM → RC → HP"). "Return" means that the brake fluid BF circulates and returns to its original flow. A check valve (also called a "check valve") is provided in the return path HK to prevent the brake fluid BF from flowing back.

The pressure regulating valve UM throttles the return flow KN, generating a pressure difference Qm between the hydraulic pressure at the top of the pressure regulating valve UM (i.e., the master cylinder hydraulic pressure Pm) and the hydraulic pressure at the bottom Pp. Specifically, the controller ECU energizes the normally open pressure regulating valve UM, reducing its valve opening Lm and adjusting the hydraulic pressure Pp at the bottom of the pressure regulating valve UM to be increased from the master cylinder hydraulic pressure Pm. Here, the hydraulic pressures adjusted by the pressure regulating valve UM are referred to as the "first and second adjusted hydraulic pressures Pp1 and Pp2."

First and second adjusted hydraulic pressure sensors PP1, PP2 (=PP) are provided in the connection passage HS to detect first and second adjusted hydraulic pressures Pp1, Pp2 (=Pp). Since there is a correlation between the valve opening amount Lm of the pressure regulating valve UM and the supplied power, the adjusted hydraulic pressure Pp can be adjusted in a predetermined relationship with the amount of electricity (e.g., current value) supplied to the pressure regulating valve UM. Therefore, the adjusted hydraulic pressure sensor PP may be omitted.

The first and second connection passages HS1 and HS2 are each branched into two and connected to the first and second wheel cylinders CW1 and CW2. The connection passage HS is provided with first and second inlet valves UI1 and UI2 (=UI). A normally open proportional valve is used as the inlet valve UI. That is, like the pressure regulating valve UM, the inlet valve UI is an electromagnetic valve whose valve opening amount (lift amount) Li can be continuously controlled according to the amount of current flow (current value).

The configuration of the connection passage HS from the branch point downward (the side closer to the wheel cylinder CW) is the same for each wheel cylinder CW. The connection passage HS is connected to the first and second pressure reduction passages HG1, HG2 (=HG) below the inlet valve UI (i.e., between the inlet valve UI and the wheel cylinder CW). The pressure reduction passage HG (fluid passage) is connected to the first and second pressure regulating reservoirs RC1, RC2 (=RC). The pressure reduction passage HG is provided with first and second outlet valves VO1, VO2 (=VO). A normally closed type on/off valve is used as the outlet valve VO.

To perform independent hydraulic control (independent adjustment of the brake fluid pressure Pw at each wheel WH), each wheel WH (i.e., wheel cylinder CW) is provided with a pair of an inlet valve UI and an outlet valve VO (shown by a dashed line). When the brake fluid pressure Pw is reduced by the independent hydraulic control, the inlet valve UI is closed and the outlet valve VO is opened. The inflow of brake fluid BF from the inlet valve UI is prevented, and the brake fluid BF in the wheel cylinder CW flows out to the pressure regulating reservoir RC, reducing the brake fluid pressure Pw. When the brake fluid pressure Pw is increased, the inlet valve UI is opened and the outlet valve VO is closed. The outflow of brake fluid BF to the pressure regulating reservoir RC is prevented, and the regulated hydraulic pressure Pp (or the master cylinder hydraulic pressure Pm) is introduced into the wheel cylinder CW, increasing the brake fluid pressure Pw. When the brake fluid pressure Pw is kept constant, the inlet valve UI and the outlet valve VO are both closed. In other words, by controlling the solenoid valves UI and VO, the brake fluid pressure Pw can be adjusted independently for each wheel cylinder CW.

The brake controller (also called "electronic control unit") ECU is composed of an electric circuit board on which a microprocessor MP, a drive circuit DR, etc. are mounted, and a control algorithm programmed into the microprocessor MP. The controller ECU is networked with other controllers (ECJ, etc.) via an on-board communication bus BS so as to share signals (detected values, calculated values, etc.). For example, the brake controller ECU is connected to the driving assistance controller ECJ through the communication bus BS. The brake controller ECU transmits the vehicle speed Vx to the driving assistance controller ECJ. Meanwhile, the driving assistance controller ECJ transmits a required deceleration Gs for executing automatic braking control to the brake controller ECU.

The wheel speed Vw and other information are input to the brake controller ECU (electronic control unit). The brake controller ECU controls the electric motor MT and the solenoid valves UM, UI, and VO. Specifically, based on a control algorithm in the microprocessor MP, drive signals Ua, Ui, and Vo for controlling the various solenoid valves UM, UI, and VO are calculated. Similarly, a drive signal Mt for controlling the electric motor MT is calculated.

<Details of inlet valve UI>
The inlet valve UI will be described in detail with reference to the cross-sectional view of Fig. 2. The inlet valve UI is a normally open proportional valve. In the inlet valve UI, the opening amount (lift amount) Li is decreased as the amount of electricity supplied thereto is increased by the braking controller ECU. The inlet valve UI is composed of a solenoid SD, a valve body VT, a guide member GD, a holding member HJ, and a spring member SB.

The solenoid SD is composed of a fixed coil CL and a plunger (movable iron core) PL. The fixed coil CL is fixed to the housing (e.g., guide member GD) of the inlet valve UI. A valve body VT is fixed to the plunger PL. The tip Vt of the valve body VT is formed (machined) into a spherical shape. In the inlet valve UI, the gap (i.e., the valve opening amount) Li between the spherical tip Vt and the valve seat Vz, which is formed (machined) into a conical shape on the holding member HJ described below, is linearly controlled according to the amount of electricity (current value) passed through the fixed coil CL. This linear control utilizes the force Fa (downward thrust in the figure, called "suction force") that draws the plunger PL into the fixed coil CL when current is passed through the fixed coil CL.

The guide member GD is provided with two holes with different diameters. The smaller of the two holes is called the "guide hole Ag" and the larger of the two holes is called the "sealing hole Af". The valve body VT is inserted into the guide hole Ag of the guide member GD so that it can move smoothly along its central axis Jv. The sealing hole Af of the guide member GD, which is located on the opposite side to the plunger PL, is sealed by a holding member HJ. Specifically, the holding member HJ is pressed into the cylindrical inner periphery of the sealing hole Af.

The valve chamber Rz is formed by the inner periphery of the sealing hole Af of the guide member GD, the end face of the holding member HJ, and the valve body VT. A conical valve seat Vz is formed on the end face of the holding member HJ on the side of the valve chamber Rz. An inlet hole Ai is provided in the center of the valve seat Vz of the holding member HJ. The inlet hole Ai is connected to the fluid pump HP (particularly the discharge part Ht) via the return flow path HK. A check valve is provided in the holding member HJ on the side of the fluid pump HP from the side of the valve chamber Rz so that the brake fluid BF can move.

A spring member SB (e.g., a compression coil spring) is provided between the retaining member HJ and the valve body VT to press the valve body VT toward the plunger PL. The spring member SB presses the valve body VT toward the plunger PL with an elastic force Fs (upward thrust in the figure). Here, the plunger PL, valve body VT (tip Vt), spring member SB, valve seat Vz, and inlet hole Ai are arranged coaxially on the central axis Jv. Therefore, the suction force Fa and the elastic force Fs oppose each other on the central axis Jv.

An outflow hole Ao is provided on the inner periphery of the sealing hole Af that forms the valve chamber Rz. The outflow hole Ao is connected to the wheel cylinder CW via a connection passage HS. In the connection passage HS, the section between the outflow hole Ao and the wheel cylinder CW is connected to the fluid pump HP (particularly, the suction section Hs) via a pressure reduction passage HG and an outlet valve VO. In the pressure reduction passage HG, a pressure regulating reservoir RC is connected between the outlet valve VO and the suction section Hs.

When independent hydraulic control is not performed, the electric motor MT and the fixed coil CL are de-energized. Therefore, no suction force Fa is generated, and the valve body VT is pressed toward the solenoid SD (plunger PL, fixed coil CL) by the elastic force Fs. Therefore, the tip Vt of the valve body VT is separated from the valve seat Vz (i.e., the inlet valve UI is open).

When independent hydraulic control begins, the electric motor MT is driven and brake fluid BF is discharged from the fluid pump HP. When it is necessary to reduce or maintain the brake fluid pressure Pw, electricity is applied to the fixed coil CL and the suction force Fa is increased. The suction force Fa causes the tip Vt of the valve body VT to abut against the valve seat Vz, and the inlet valve UI is closed. At this time, in addition to the elastic force Fs, a fluid force Fb (upward thrust in the figure) due to the brake fluid BF acts on the valve body VT. Therefore, when closing the inlet valve UI, the amount of electricity required to generate a suction force Fa (> "Fb + Fs") that is sufficient to counter the elastic force Fs and the fluid force Fb is supplied to the fixed coil CL.

When it is necessary to increase the brake fluid pressure Pw, the amount of electricity supplied to the fixed coil CL is reduced compared to when the inlet valve UI is closed. This reduces the suction force Fa, and when the suction force Fa and the thrust force "Fb + Fs" are in balance, the valve opening amount Li, which is the gap between the tip end Vt and the valve seat Vz, is determined. Due to the orifice effect of the valve opening amount Li, a hydraulic pressure difference (differential pressure) Qa is generated between the upstream hydraulic pressure Pa (the first hydraulic pressure, the hydraulic pressure in the inlet hole Ai) and the downstream hydraulic pressure Pb (the second hydraulic pressure, the hydraulic pressure in the valve chamber Rz) in the return flow KN of the brake fluid BF (i.e., "Qa = Pa - Pb"). Therefore, the differential pressure Qa (actual value) is the hydraulic pressure difference between the adjustment hydraulic pressure Pp (or the master cylinder hydraulic pressure Pm) (= Pa) and the brake hydraulic pressure Pw (= Pb). The actual differential pressure Qa is adjusted based on the target differential pressure Qt (= Pt - Pp) in independent hydraulic control.

The brake fluid BF discharged by the fluid pump HP contains the pulsation (fluctuations in hydraulic pressure) of the fluid pump HP. Therefore, the fluid force Fb contains a fluctuating component caused by the pulsation of the fluid pump HP. In order to keep the valve opening amount Li constant, the fluctuating component of this fluid force Fb needs to be compensated for. This is particularly necessary when the opening area of the valve seat Vz is set large so as to ensure the flow rate of the inlet valve UI. For example, one method of compensation is to increase the spring constant of the spring member SB to relatively reduce the influence of the fluctuating component of the fluid force Fb. However, this method increases the elastic force Fs, so that it is necessary to increase the suction force Fa to counteract this. As a result, the solenoid SD and the members related thereto are enlarged and the power consumption is increased. In the brake control device SC according to the present invention, the adjustment of the amount of current as described below is adopted to perform the compensation. This adjustment effectively reduces hydraulic pressure fluctuations caused by the fluctuating components of the fluid force Fb without increasing the size of the solenoid SD.

<Differences between the inlet valve UI and the pressure regulating valve UM>
Both the inlet valve UI and the pressure regulating valve UM are normally-open proportional valves (differential pressure valves), but differ in the following respects. In the inlet valve UI, the inlet hole Ai of the holding member HJ is connected to the discharge part Ht of the fluid pump HP, and the tip part Vt of the valve body VT is brought close to the valve seat Vz of the holding member HJ to adjust the valve opening amount Li. Therefore, a fluid force Fb acts on the valve body VT of the inlet valve UI. On the other hand, in the pressure regulating valve UM, the master cylinder CM is connected to the inlet hole, and the discharge part Ht of the fluid pump HP is connected to the outlet hole (see Figures 1 and 2 of Patent Document 3). Therefore, the fluid force of the brake fluid BF pumped from the master cylinder CM acts on the valve body of the pressure regulating valve UM, but the fluid force of the brake fluid BF discharged by the fluid pump HP does not act on the valve body of the pressure regulating valve UM. Since the brake fluid BF from the master cylinder CM is not affected by pulsation, the pressure regulating valve UM does not have the problem of hydraulic pressure fluctuations caused by the fluid force of the brake fluid BF discharged by the fluid pump HP.

<Electrical control of inlet valve UI>
The energization control of the inlet valve UI will be described with reference to the functional block diagram of FIG. 3. The inlet valve UI is provided in the connection passage HS below the pressure regulating valve UM (connection passage HS between the pressure regulating valve UM and the wheel cylinder CW). The inlet valve UI and the outlet valve VO adjust the hydraulic pressure (braking hydraulic pressure) Pw for each wheel cylinder CW independently. In other words, the inlet valve UI and the outlet valve VO enable independent hydraulic pressure control. The energization control in the inlet valve UI is performed by an independent hydraulic pressure control block DC, a standard energization amount calculation block IS, an adjustment energization amount calculation block IC, and a target energization amount calculation block IT. The calculation processes of the independent hydraulic pressure control block DC to the target energization amount calculation block IT are algorithmic processes programmed in the microprocessor MP. The drive circuit DR is controlled based on the target energization amount It, which is the output from the target energization amount calculation block IT. The microprocessor MP and the drive circuit DR are included in a controller ECU.

In the independent hydraulic control block DC, the control mode Md of the inlet valve UI is determined based on the braking operation amount Ba, etc., so that independent hydraulic control is performed, and the target value Qt of the hydraulic pressure difference to be formed by the inlet valve UI is calculated. Here, the independent hydraulic control includes well-known antilock brake control, traction control, and vehicle stability control. In order to perform these controls, the independent hydraulic control block DC receives the wheel speed Vw detected by the wheel speed sensor VW, the steering angle Sa detected by the steering angle sensor SA, the yaw rate Yr detected by the yaw rate sensor YR, the deceleration Gx detected by the deceleration sensor GX, and the lateral acceleration Gy detected by the lateral acceleration sensor Gy.

The independent hydraulic control block DC is composed of a vehicle speed calculation block VX, a slip state quantity calculation block SL, a turning state quantity calculation block SJ, a target differential pressure calculation block QT, and a control mode selection block MD.

The vehicle body speed calculation block VX calculates the vehicle body speed Vx based on the wheel speeds Vw of the vehicle's four wheels WH and a known method. For example, when braking, the vehicle body speed Vx is calculated based on the fastest of the four wheel speeds Vw. When not braking, including when coasting, the vehicle body speed Vx is calculated based on the slowest of the four wheel speeds Vw.

In the slip state amount calculation block SL, a slip state amount Sl representing the degree of slip state of the wheel WH is calculated based on the wheel speed Vw and the vehicle speed Vx. Specifically, the slip state amount Sl is calculated as the "time change amount dV of the wheel speed Vw (wheel acceleration, time differential value of the wheel speed Vw)" and the "deviation between the vehicle speed Vx and the wheel speed Vw". Furthermore, as the deviation, the deceleration slip Sg is used for anti-lock brake control, and the acceleration slip Sk is used for traction control. The deceleration slip Sg is determined as the value obtained by subtracting the wheel speed Vw from the vehicle speed Vx (i.e., "Sg = Vx - Vw"). The deceleration slip Sg represents the tendency of the wheel WH to lock, and when the wheel WH is completely locked, "Sg = Vx". Additionally, the acceleration slip Sk is determined as the wheel speed Vw minus the vehicle speed Vx (i.e., "Sk = Vw - Vx"). The acceleration slip Sk represents the tendency of the wheels to spin (overspin).

In the turning state quantity calculation block SJ, a turning state quantity Sj that indicates the degree of turning stability of the vehicle (vehicle body) is calculated based on the vehicle body speed Vx, yaw rate Yr, lateral acceleration Gy, steering angle Sa, and a known calculation method. Specifically, the yaw rate deviation hY, sideslip angle β, and sideslip angular velocity dβ are calculated as the turning state quantity Sj. Here, the sideslip angle β is the angle between the direction of the vehicle (vehicle body) and its traveling direction. The sideslip angular velocity dβ is the time change amount (time differential value) of the sideslip angle β. The yaw rate deviation hY is calculated as the deviation between the standard yaw rate Ys calculated based on the steering angle Sa and the vehicle body speed Vx and the actual yaw rate Yr detected by the yaw rate sensor YR.

In the control mode selection block MD and the target differential pressure calculation block QT, the control mode Md and the target differential pressure Qt are calculated based on at least one of the calculation results Sl (wheel acceleration dV, etc.) of the slip state quantity calculation block SL and the calculation results Sj (yaw rate deviation hY, etc.) of the turning state quantity calculation block SJ. The control mode Md indicates an increase or decrease in the brake fluid pressure Pw, and is also a general term for the pressure reduction mode Mg, the pressure holding mode Mh, and the pressure increase mode Mz described below. The target differential pressure Qt is the target value of the actual differential pressure Qa generated by the inlet valve UI. Specifically, the target hydraulic pressure Pt for the brake fluid pressure Pw of each wheel cylinder CW is determined, and the target differential pressure Qt is determined based on the target hydraulic pressure Pt. This is based on the fact that the master cylinder hydraulic pressure Pm and the adjustment hydraulic pressure Pp are detected (or estimated) and are known.

In anti-lock brake control, which suppresses deceleration slip of the wheels (i.e., the tendency of the wheels WH to lock), the target hydraulic pressure Pt (resulting in braking hydraulic pressure Pw) is increased or decreased based on the deceleration slip Sg and the wheel acceleration dV. Specifically, the hydraulic pressures Pt (target value) and Pw (actual value) are decreased under the condition that the wheel acceleration dV is less than a predetermined acceleration dx (a constant with a negative sign, a threshold value related to the wheel acceleration dV in the anti-lock brake control) and the deceleration slip Sg is equal to or greater than a first predetermined slip sx (a constant with a positive sign, a threshold value related to the deceleration slip Sg). This situation is called the "reduction mode Mg" among the control modes Md. After the target hydraulic pressure Pt is decreased, the target hydraulic pressure Pt is maintained constant for a predetermined time tx. This situation is called the "maintenance mode Mh" among the control modes Md. After a predetermined time tx has elapsed, the target hydraulic pressure Pt is increased so that the deceleration slip Sg approaches a second predetermined slip st (a preset constant smaller than the first predetermined slip sx). This state is called the "pressure increase mode Mz" among the control modes Md. In other words, in the pressure increase mode Mz, the hydraulic pressures Pt (target value) and Pw (actual value) are adjusted so that the wheel speed Vw approaches and matches the target speed Vs ("=Vx-st", a target value related to the wheel speed Vw) determined based on the vehicle speed Vx. In the control mode selection block MD, one of the pressure reduction mode Mg, the pressure holding mode Mh, and the pressure increase mode Mz is selected from among the various modes of the control mode Md based on the target hydraulic pressure Pt.

In traction control, which suppresses wheel acceleration slip (i.e., overspinning of the wheels WH), the output of the vehicle's power source (e.g., the engine) is adjusted and the target hydraulic pressure Pt (resulting in the brake hydraulic pressure Pw) is increased or decreased based on the acceleration slip Sk and the wheel acceleration dV. The hydraulic pressures Pt and Pw are increased under the condition that the wheel acceleration dV is equal to or greater than a predetermined acceleration dy (a positive constant that is a threshold value related to the wheel acceleration dV in traction control) and the acceleration slip Sk is equal to or greater than a predetermined slip sk (a positive constant that is a threshold value related to the acceleration slip Sk). When the wheel acceleration dV and the acceleration slip Sk fall within a predetermined range, the hydraulic pressures Pt and Pw are maintained and then decreased. As with antilock brake control, in traction control, a pressure increase mode Mz in which the hydraulic pressures Pt and Pw are increased, a pressure retention mode Mh in which the hydraulic pressures Pt and Pw are maintained, and a pressure reduction mode Mg in which the hydraulic pressures Pt and Pw are decreased are calculated (determined).

In vehicle stability control, which stabilizes the vehicle (i.e., suppresses excessive oversteer and understeer), the target hydraulic pressure Pt of each wheel cylinder CW is calculated based on the turning state quantity Sj, and the brake hydraulic pressure Pw is controlled based on the target hydraulic pressure Pt. For example, when oversteer is occurring, the target hydraulic pressure Pt of the front wheel located on the outside in the turning direction of the vehicle is calculated as "Pt = Ka x β + Kb x dβ". Here, "Ka, Kb" are predetermined coefficients (constants) that are set in advance. In vehicle stability control, a pressure increase mode Mz in which the hydraulic pressures Pt and Pw are increased, a pressure retention mode Mh in which the hydraulic pressures Pt and Pw are maintained, and a pressure reduction mode Mg in which the hydraulic pressures Pt and Pw are reduced are calculated (determined).

In the target differential pressure calculation block QT, the target differential pressure Qt is calculated based on the target hydraulic pressure Pt. Specifically, the target differential pressure Qt is determined by subtracting the adjusted hydraulic pressure Pp (hydraulic pressure adjusted by the pressure regulating valve UM) from the target hydraulic pressure Pt (i.e., "Qt = Pt - Pp"). When antilock brake control is executed due to the driver's operation of the brake operating member BP, no current is applied to the pressure regulating valve UM and the adjusted hydraulic pressure Pp and the master cylinder hydraulic pressure Pm are equal, so the target differential pressure Qt is determined by subtracting the master cylinder hydraulic pressure Pm from the target hydraulic pressure Pt (i.e., "Qt = Pt - Pm").

In the standard current flow calculation block IS, the standard current flow Is is calculated based on the target differential pressure Qt (target value of hydraulic pressure difference) and a preset calculation map Zis. The standard current flow Is is a reference for the current flow state of the inlet valve UI (particularly the solenoid SD). In other words, the standard current flow Is is a target value of the current flow to the inlet valve UI that determines the opening amount Li of the inlet valve UI. The standard current flow Is is calculated to monotonically increase in accordance with an increase in the target differential pressure Qt according to the calculation map Zis. An upper limit value is set for the standard current flow Is. The upper limit value is a preset constant and corresponds to the current flow at which the normally open inlet valve UI is reliably closed.

In the adjustment current amount calculation block IC, the adjustment current amount Ic is calculated based on the calculation map Zic. The adjustment current amount Ic is a target value of the current amount for reducing the fluctuation of the brake fluid pressure Pw caused by the fluid force Fb of the brake fluid BF discharged from the fluid pump HP. In the calculation map Zic, the adjustment current amount Ic is calculated so that it increases periodically only in the direction in which the valve body VT approaches the valve seat Vz. Specifically, in the calculation map Zic, the adjustment current amount Ic is calculated as a rectangular wave that is periodically repeated with the passage of time T. A "rectangular wave" is a wave in which the temporal change in vibration repeats at two constant values ("0" and value ip), high and low, and is also called a "square wave". The shape of the rectangular wave is determined by the period ts, the value ip (also called the "magnitude of the adjustment current amount Ic"), and the time tp (also called the "maintenance time") for maintaining the value ip. Here, the adjusted current flow amount Ic (particularly, "Ic = ip") increases the thrust (attraction force) Fa in the direction in which the valve body VT approaches the valve seat Vz (i.e., the direction against the elastic force Fs and the fluid force Fb). Therefore, the adjusted current flow amount Ic can be said to be "the current flow amount that increases the suction force Fa only in the direction in which the valve body VT approaches the valve seat Vz."

The adjusted current flow amount Ic can be calculated in conjunction with the independent hydraulic control. For example, when independent hydraulic control is being performed, the adjusted current flow amount Ic is calculated, but when independent hydraulic control is not being performed, the adjusted current flow amount Ic is not calculated. This is because, when independent hydraulic control is being performed, the electric motor MT is driven, brake fluid BF is discharged from the fluid pump HP, and the inlet valve UI is driven.

The calculation of the adjusted current Ic may also be linked to the driving of the electric motor MT. For example, when the electric motor MT is not driven, the adjusted current Ic is not calculated, and the calculation of the adjusted current Ic is performed only when the electric motor MT is driven. This is based on the fact that the fluctuation in the brake fluid pressure Pw due to the fluid force Fb of the brake fluid BF is caused by the hydraulic pressure fluctuation (pulsation) of the brake fluid BF discharged by the fluid pump HP.

The calculation map Zic of the adjusted current flow amount Ic can be adjusted based on the target differential pressure Qt. For example, as shown in the calculation map Zip in the left diagram of the blow-out section (Z), the adjustment current flow amount Ic is calculated to be larger as the target differential pressure Qt increases. Also, as shown in the calculation map Ztp in the right diagram of the blow-out section (Z), the maintenance time tp may be calculated to be larger (longer) as the target differential pressure Qt increases.

In the target current calculation block IT, the target current It is calculated based on the standard current calculation block IS and the adjusted current Ic. The target current It is the final target value of the current of the inlet valve UI. Specifically, in the target current calculation block IT, the standard current Is (reference current state of the inlet valve UI) is superimposed with the adjusted current Ic (compensation for hydraulic pressure fluctuations caused by fluid forces) to calculate the target current It (i.e., "It = Is + Ic"). The target current It is sent to the drive circuit DR. Then, in the drive circuit DR, the current Ia of the inlet valve UI is actually controlled based on the target current It.

In addition to the elastic force Fs generated by the spring member SB, the fluid force Fb of the brake fluid BF discharged by the fluid pump HP acts on the valve body VT. When the fluid force Fb increases due to the pulsation of the fluid pump HP while the amount of current Ia flowing through the fixed coil CL is kept constant, the valve body VT moves in a direction away from the valve seat Vz (i.e., in the direction in which the fluid force Fb and the elastic force Fs act). The movement of the valve body VT increases the valve opening Li, so the actual differential pressure Qa decreases. At this time, in order to instantly move the valve body VT closer to the valve seat Vz so as to maintain the valve opening Li constant, a corresponding amount of power is required, and a large solenoid SD is required. Conversely, in a situation in which the fluid force Fb decreases due to the pulsation of the fluid pump HP, the amount of current flowing through the solenoid SD only needs to be reduced, so there is no need for high responsiveness.

In the brake control device SC, the standard of the energized state of the solenoid SD is calculated as a standard energization amount Is according to the target differential pressure Qt. Then, the target energization amount It is calculated so that this standard energization amount Is is periodically increased only in the direction in which the valve body VT approaches the valve seat Vz (i.e., the direction against the elastic force Fs and the fluid force Fb, and the direction in which the suction force Fa increases). In other words, the energization amount Ic corresponding to the increase in the fluid force Fb due to the pulsation of the fluid pump HP is predicted. Then, by predicting this in advance, the target energization amount It is calculated. Therefore, the hydraulic pressure fluctuation caused by the fluid force Fb can be reduced without increasing the size of the device related to the linear type inlet valve UI (solenoid SD, drive circuit DR, etc.). Note that the adjusted energization amount Ic is not for seating the valve body VT on the valve seat Vz, but for suppressing the movement amount of the valve body VT against the fluid force Fb. Therefore, even if the adjusted current Ic is superimposed, the inlet valve UI is not closed. In other words, the magnitude ip of the adjusted current Ic is an amount of current that does not cause the valve body VT (particularly the tip portion Vt) to come into contact with the valve seat Vz.

The situation in which the fluid force Fb is most likely to affect the valve body VT is when the valve opening Li is narrowed and the target differential pressure Qt (and thus the actual differential pressure Qa) is adjusted to a large value. In the brake control device SC, the inlet valve UI adjusts the adjustment current flow Ic based on the hydraulic pressure difference Qt (target value) and Qa (actual value) between the first hydraulic pressure Pa, which is the hydraulic pressure upstream of the valve seat Vz, and the second hydraulic pressure Pb, which is the hydraulic pressure downstream of the valve seat Vz, in the flow of the brake fluid BF discharged by the fluid pump HP. Specifically, the larger the hydraulic pressure difference Qt (target value) and Qa (actual value), the larger the adjustment current flow Ic is calculated to be. This allows the adjustment current flow Ic to be suitably adjusted based on the hydraulic pressure differences Qt and Qa. The adjusted current flow amount Ic is determined by the "magnitude ip of the adjusted current flow amount Ic (the magnitude in the current flow axis direction in the rectangle)" and the "maintenance time tp (the magnitude in the time axis direction during which the magnitude ip is maintained in the rectangle)." Therefore, "calculating the adjusted current flow amount Ic to be large" corresponds to at least one of "increasing the magnitude ip of the adjusted current flow amount Ic" and "increasing the maintenance time tp."

For example, the superposition of the adjusted current Ic in the target current calculation block IT is performed (permitted) only when independent hydraulic control is being performed. That is, when independent hydraulic control is not being performed, the superposition of the adjusted current Ic is prohibited. Alternatively, the superposition of the adjusted current Ic may be performed (permitted) only when the fluid pump HP driven by the electric motor MT is discharging brake fluid BF. That is, when the electric motor MT is not being driven, the superposition of the adjusted current Ic is prohibited. Since the superposition of the adjusted current Ic is performed only when necessary, the reliability of the current control of the inlet valve UI can be improved.

The control mode Md is taken into consideration when calculating the target current It. In the target current calculation block IT, the adjustment current Ic is added to the standard current Is only when the control mode Md is the pressure increase mode Mz (i.e., when the brake fluid pressure Pw is increased). In other words, when the control mode Md is the holding mode Mh or the pressure reduction mode Mg (i.e., when the brake fluid pressure Pw is not increased), the adjustment current Ic is not added to the standard current Is, and the standard current Is is output as it is as the target current It. This is because in the holding mode Mh and the pressure reduction mode Mg, the inlet valve UI is closed, so that fluctuations in the brake fluid pressure Pw due to the fluid force Fb cannot occur. In the holding mode Mh and the pressure reduction mode Mg, the calculation (superimposition) of the adjusted current flow amount Ic is prohibited, so that the adjusted current flow amount Ic is not superimposed when it is not necessary, thereby improving the reliability of the current flow amount control of the inlet valve UI.

The above-mentioned permission/prohibition of superimposition of the adjustment current amount Ic may be performed in the adjustment current amount calculation block IC. That is, in the adjustment current amount calculation block IC, the adjustment current amount Ic is calculated based on the calculation map Zic only when superimposition of the adjustment current amount Ic is permitted (when the independent hydraulic control is being executed, when the electric motor MT is being driven, and in any one of the pressure increase mode Mz). On the other hand, when superimposition of the adjustment current amount Ic is prohibited (when the independent hydraulic control is not being executed, when the electric motor MT is stopped, in the holding mode Mh, and in the pressure reduction mode Mg), the adjustment current amount Ic is calculated to be "0" in the adjustment current amount calculation block IC. Then, in the target current amount calculation block IT, the adjustment current amount Ic is superimposed on the standard current amount Is to determine the target current amount It. In either configuration, the increase adjustment in which the standard current flow amount Is is increased to calculate the target current flow amount It is permitted when necessary, but is prohibited when not necessary. This provides the above-mentioned effects.

In the above embodiment, three modes, namely, pressure increase mode Mz, pressure hold mode Mh, and pressure reduction mode Mg, are adopted as the control mode Md. In the control mode Md, the pressure hold mode Mh may be omitted. In this case, the independent hydraulic control is configured with the pressure increase mode Mz in which the brake hydraulic pressure Pw is increased, and the pressure reduction mode Mg in which the brake hydraulic pressure Pw is decreased. The increase adjustment according to the adjustment current flow amount Ic is permitted in the pressure increase mode Mz in which the inlet valve UI is open, but is prohibited in the pressure reduction mode Mg in which the inlet valve UI is closed. Even if the brake control device SC has a control mode Md configured with two modes, the same effect as above can be achieved.

<Selection process of control mode Md in anti-lock brake control>
An example of the selection process of the control mode Md in the anti-lock brake control will be described with reference to the time series diagram of Fig. 4 (changes in the state quantities Pt and Pw with respect to time T). In the diagram, the actual brake fluid pressure Pw is controlled to match the target fluid pressure Pt, so the diagram of the brake fluid pressure Pw and the diagram of the target fluid pressure Pt overlap. As described above, the holding mode Mh is omitted from the diagram. In the example, the anti-lock brake control is being executed, the electric motor MT is being driven, and the brake fluid BF is being discharged from the fluid pump HP.

Up until time t0, the control mode Md is in the pressure increase mode Mz, and the brake fluid pressure Pw is increased. At time t0, the wheel acceleration dV and the deceleration slip Sg reach the antilock brake control thresholds dx (threshold related to wheel acceleration) and sx (threshold related to deceleration slip), and the control mode Md transitions from the pressure increase mode Mz to the pressure reduction mode Mg. At time t0, the inlet valve UI is closed and the outlet valve VO is opened, thereby decreasing the brake fluid pressure Pw from the value pa to the value pb at time t1. Here, the time from time t0 until time t1 when the pressure reduction mode Mg is executed (referred to as the "pressure reduction time") is set in advance as a predetermined time (constant).

At time t1, the pressure reduction mode Mg is ended and the pressure increase mode Mz is started. In the first stage of the pressure increase mode Mz, the pressure reduction amount dp by the pressure reduction mode Mg is increased by a predetermined ratio Gp. This pressure increase mode Mz is called the "first pressure increase mode Mzf". Specifically, in the first pressure increase mode Mzf, the difference (pressure reduction amount) dp (= pa-pb) between the value pa before the pressure reduction and the value pb after the pressure reduction is multiplied by a predetermined ratio Gp (for example, 50%) to calculate the target pressure increase amount up from the value pb (i.e., "up = Gp x dp"). Then, in the first pressure increase mode Mzf, the target hydraulic pressure pc to be reached by the brake hydraulic pressure Pw is determined (i.e., "pb + up = pb + Gp x dp"). The target hydraulic pressure Pt is calculated so that the brake hydraulic pressure Pw reaches the target hydraulic pressure pc at the hydraulic pressure gradient Kf (referred to as the "first gradient"). Here, the first gradient Kf is a preset value (constant).

At time t2, the brake fluid pressure Pw reaches the target fluid pressure pc. At time t2, the first pressure-increasing mode Mzf is ended, and a new pressure-increasing mode Mzs (the latter part of the pressure-increasing mode Mz) is started. The latter part Mzs of the pressure-increasing mode Mz is called the "second pressure-increasing mode." In the second pressure-increasing mode Mzs, the fluid pressures Pt and Pw are increased at a fluid pressure gradient Ks (referred to as the "second gradient") that is smaller than the first gradient Kf. Specifically, in the second pressure-increasing mode Mzs, the fluid pressures Pt and Pw are increased so that the wheel speed Vw coincides with the target speed Vs (=Vx-st) that is set based on the vehicle speed Vx. Here, the second gradient Ks in the second pressure-increasing mode Mzs is a value smaller than the first gradient Kf and is a predetermined value (constant) that is set in advance.

At time t3, the wheel acceleration dV and the deceleration slip Sg again reach the anti-lock brake control thresholds dx, sx, and the control mode Md transitions from the pressure increase mode Mz to the pressure decrease mode Mg. The control mode Md transitions from the pressure increase mode Mz to the pressure decrease mode Mg, and the brake fluid pressure Pw is reduced. In the anti-lock brake control, such an increase and decrease in the brake fluid pressure Pw is repeated to avoid the tendency of the wheel WH to lock (excessive deceleration slip Sg). In the second pressure increase mode Mzs, the brake fluid pressure Pw is increased more gradually than in the first pressure increase mode Mzf.

In the above embodiment, the increase adjustment based on the adjustment current Ic is not performed in the pressure reduction mode Mg, but is performed only in the pressure increase mode Mz. In addition, in the first mode Mzf of the pressure increase mode Mz (the front part of the pressure increase mode Mz), the increase adjustment according to the adjustment current Ic may be prohibited, and in the second mode Mzs of the pressure increase mode Mz (the back part of the pressure increase mode Mz), the increase adjustment may be permitted. This is because in the first pressure increase mode Mzf based on the hydraulic pressures Pt and Pw before the start of the pressure increase mode Mz, the brake fluid BF to the wheel cylinder CW is important, and the valve opening amount Li is large. Therefore, in the first pressure increase mode Mzf, the adjustment current Ic is not added (superimposed) to the standard current Is, or the adjustment current Ic is calculated to be "0". On the other hand, in the second pressure increase mode Mzs based on the vehicle speed Vx, the valve opening amount Li is smaller than that in the first pressure increase mode Mzf. Therefore, in the second pressure-boosting mode Mzs, the adjusted current Ic is added (superimposed) to the standard current Is to calculate the target current It. This makes it possible to appropriately suppress fluctuations in the brake fluid pressure Pw caused by the fluid force Fb of the brake fluid BF from the fluid pump HP.

<Other embodiments>
In the above embodiment, a diagonal type fluid path is used as the two fluid paths. Alternatively, a front-rear type (also called "II type") fluid path may be used as the two fluid paths. In this case, the first hydraulic pressure chamber Rm1 of the master cylinder CM is connected to the wheel cylinders CW of the left and right front wheels, and the second hydraulic pressure chamber Rm2 is connected to the wheel cylinders CW of the left and right rear wheels. With this configuration, the same effects as those described above can be achieved.

In the above embodiment, a configuration in which a brake pad is used as the friction member MS and a brake disc is used as the rotating member KT (a so-called disk-type brake configuration) has been exemplified. Alternatively, a configuration in which a brake lining is used as the friction member MS and a brake drum is used as the rotating member KT (a so-called drum-type brake configuration) may be used. Even with a drum-type brake, the same effects as with a disk-type brake can be achieved.

SC...Brake control device, BP...Brake operating member, CM...Master cylinder, CW...Wheel cylinder, HS...Connection path, BA...Operation amount sensor, HU...Fluid unit, ECU...Brake controller (electronic control unit), MT...Electric motor, HP...Fluid pump, UM...Pressure regulating valve (proportional valve), UI...Inlet valve (proportional valve), VO...Outlet valve (ON/OFF valve), SD...Solenoid, CL...Fixed coil, PL...Plunger, VT...Valve body, HJ...Retaining member, GD...Guide member, Vz...Valve seat, Vt...Tip, SB...Spring member, Pm...Master cylinder hydraulic pressure , Pp...adjusted fluid pressure, Pw...wheel cylinder fluid pressure (brake fluid pressure), Pa...first fluid pressure, Pb...second fluid pressure, Qa...actual differential pressure (fluid pressure difference between adjusted fluid pressure Pp and wheel cylinder fluid pressure Pw), Qt...target differential pressure (target value of actual differential pressure Qa), Ia...amount of current, Md...control mode, Mz...pressure increase mode, Mh...holding mode, Mg...pressure reduction mode, Is...standard amount of current, Ic...adjusted amount of current, It...target amount of current, Fa...suction force (thrust force by solenoid SD), Fb...fluid force (force by brake fluid BF discharged by fluid pump HP), Fs...elastic force (thrust force by spring member SB).

Claims (3)

A vehicle brake control device that adjusts brake fluid pressure in a wheel cylinder of a vehicle to control a braking force of a wheel, comprising:
a fluid pump driven by an electric motor;
a linear inlet valve including a solenoid, a valve body driven by the solenoid, and a valve seat capable of contacting the valve body, in which a fluid force of the brake fluid discharged by the fluid pump acts on the valve body;
a controller that adjusts a current supply state of the solenoid to control the inlet valve;
Equipped with
The inlet valve is normally open,
The controller:
The reference current flow state is calculated as a standard current flow amount,
The standard energization amount is periodically increased only in a direction in which the valve body approaches the valve seat to calculate a target energization amount;
controlling the energization state based on the target energization amount;
When the size of the gap between the valve body and the valve seat is defined as the valve opening amount,
sequentially controlling the inlet valve in a first pressure-increasing mode in which the brake fluid pressure is increased and a second pressure-increasing mode in which the brake fluid pressure is increased with the valve opening amount smaller than that in the first pressure-increasing mode;
An increasing adjustment for periodically increasing the standard energization amount to calculate the target energization amount;
When the brake fluid pressure is increased in the second pressure increasing mode, the brake fluid pressure is permitted to increase.
A vehicle brake control device that prohibits the brake fluid pressure from being increased in the first pressure increase mode . 2. The vehicle brake control device according to claim 1,
The controller:
When the brake fluid pressure is to be increased after being decreased, the inlet valve is controlled in sequence in the first pressure increase mode in which the brake fluid pressure is increased to a target pressure corresponding to a decrease in the brake fluid pressure when the brake fluid pressure is decreased, and in the second pressure increase mode.
A vehicle brake control device which prohibits the increasing adjustment when the brake fluid pressure is to be decreased . 3. The vehicle brake control device according to claim 1 ,
The inlet valve is
a hydraulic pressure difference between a first hydraulic pressure, which is a hydraulic pressure on an upstream side of the valve seat, and a second hydraulic pressure, which is a hydraulic pressure on a downstream side of the valve seat, in a flow of brake fluid discharged by the fluid pump;
The controller:
A vehicle brake control device that increases an adjustment current amount, which is increased from the standard current amount to the target current amount, as the hydraulic pressure difference increases.