JP2017077735A - Brake control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a brake control device capable of accurately estimating pressure on a discharge side of a liquid pressure source without adding an additive member.SOLUTION: A brake control device comprises: a liquid pressure source for sucking a brake liquid in a master cylinder or a reservoir, discharging the brake liquid to a wheel cylinder for boosting the wheel cylinder liquid pressure; a motor for driving the liquid pressure source; a motor drive part for on/off driving the motor; a motor actual rotation number estimation part 21 for estimating the actual rotation number of the motor on the basis of inter terminal voltage of the motor; a motor model rotation number estimation part 24 for estimating the model rotation number of the motor from a motor model which is preset on the basis of characteristics of the motor; and a liquid pressure source discharge side pressure estimation part 23 for estimating a load applied to a motor, on the basis of a difference between the model rotation number estimated by the motor model rotation number estimation part and the actual rotation number estimated by the motor actual rotation number estimation part, and estimating pressure on a discharge side of the liquid pressure source on the basis of the estimated load.SELECTED DRAWING: Figure 2

Description

  The present invention relates to a brake control device.

  Patent Document 1 discloses a brake control device that increases the wheel cylinder hydraulic pressure by sucking brake fluid in a master cylinder or a reservoir with a hydraulic pressure source and discharging the brake fluid to the wheel cylinder.

JP 2004-231071 A

In the above prior art, if the pressure on the discharge side of the hydraulic pressure source is known, the control accuracy of the wheel cylinder hydraulic pressure can be improved. Here, it is conceivable to provide a pressure sensor in order to know the pressure on the discharge side of the hydraulic pressure source, but this increases the cost.
The objective of this invention is providing the brake control apparatus which can estimate the pressure of the discharge side of a hydraulic pressure source accurately, without adding an additional member.

  In the present invention, the load applied to the motor is estimated based on the difference between the model rotational speed estimated from the motor characteristics and the actual rotational speed estimated from the voltage between the terminals of the motor, and the discharge of the hydraulic pressure source is performed based on the estimated load. Estimate the side pressure.

  Therefore, the pressure on the discharge side of the hydraulic pressure source can be accurately estimated without adding an additional member.

It is a block diagram of the brake control apparatus of Example 1. FIG. 4 is a control block diagram of a voltage conversion motor rotation speed estimation unit 21, a motor OFF time calculation unit 22, and a pump discharge side pressure estimation unit 23 in the brake control unit BCU. FIG. 4 is a control block diagram of a motor model rotation speed estimation unit 24 in the brake control unit BCU. It is a flowchart which shows the flow of the pump discharge side pressure calculation process in brake control unit BCU. It is a time chart of the motor rotation speed of Example 1, motor electromotive force, and a motor estimated load. 3 is a time chart of a pump discharge side pressure according to the first embodiment.

[Example 1]
First, the configuration will be described.
[System configuration]
FIG. 1 is a configuration diagram of the brake control device according to the first embodiment.
The hydraulic control unit HU adjusts the braking force applied to each wheel of the vehicle. Based on commands from the brake control unit BCU, the wheel cylinder W / C (RL) for the left rear wheel and the wheel cylinder for the right front wheel Increase / decrease or maintain the hydraulic pressures of W / C (FR), left front wheel wheel cylinder W / C (FL), and right rear wheel wheel cylinder W / C (RR).
The hydraulic control unit HU has a piping structure called X piping, which is composed of two systems, a P system and an S system. By adopting X piping, even if one piping system fails, the other piping system can be used to generate half of the braking force during normal operation. In addition, P attached to the end of the code | symbol of each site | part described in FIG. 1 shows P system | strain, S shows S system | strain, RL, FR, FL, RR is a right rear wheel, a left front wheel, a left front wheel, and a right rear. Indicates that it corresponds to a ring. In the following description, the description of P, S or RL, FR, FL, RR is omitted when the P, S system or each wheel is not distinguished.

The hydraulic control unit HU according to the first embodiment uses a closed hydraulic circuit. Here, the “closed hydraulic circuit” refers to a hydraulic circuit that returns the brake fluid supplied to the wheel cylinder W / C to the reservoir tank RSV via the master cylinder M / C. By the way, the hydraulic circuit that can return the brake fluid supplied to the wheel cylinder W / C directly to the reservoir tank RSV without passing through the master cylinder M / C is called “Open hydraulic circuit”. That's it.
The brake pedal BP is connected to the master cylinder M / C via the input rod IR. The pedal depression force input to the brake pedal BP is boosted by the brake booster BB. The master cylinder M / C generates brake fluid pressure according to the output of the brake booster BB.
The wheel cylinder W / C (RL) for the left rear wheel RL and the wheel cylinder W / C (FR) for the right front wheel FR are connected to the S system, and the wheel cylinder W / C (for the left front wheel FL is connected to the P system. FL) and the wheel cylinder W / C (RR) of the right rear wheel RR are connected. The P system and the S system are provided with oil pumps PP and PS. The oil pumps PP and PS are driven by one motor M. The motor M is a direct current motor. The oil pumps PP and PS are plunger pumps.

Master cylinder M / C and wheel cylinder W / C are connected by pipe line 1 and pipe line 2. The pipeline 2S is branched into pipelines 2RL and 2FR, the pipeline 2RL is connected to the wheel cylinder W / C (RL), and the pipeline 2FR is connected to the wheel cylinder W / C (FR). The pipe line 2P branches into pipe lines 2FL and 2RR, the pipe line 2FL is connected to the wheel cylinder W / C (FL), and the pipe line 2RR is connected to the wheel cylinder W / C (RR).
On the pipeline 1, a gate-out valve 3 which is a normally open type proportional control valve is provided. A pressure sensor 17 for detecting the master cylinder hydraulic pressure is provided at a position closer to the master cylinder side than the gate-out valve 3P of the pipe line 1P of the P system. On the pipeline 1, a pipeline 4 is provided in parallel with the gate-out valve 3. A check valve 5 is provided on the pipeline 4. The check valve 5 allows the flow of brake fluid from the master cylinder M / C to the wheel cylinder W / C and prohibits the flow in the opposite direction.
A solenoid-in valve 6 which is a normally open proportional control valve corresponding to each wheel cylinder W / C is provided on the pipe line 2. On the pipe 2, a pipe 7 is provided in parallel with the solenoid-in valve 6. A check valve 8 is provided on the pipeline 7. The check valve 8 allows the brake fluid to flow in the direction from the wheel cylinder W / C toward the master cylinder M / C, and prohibits the flow in the opposite direction.

The discharge side of the oil pump P and the pipeline 2 are connected by a pipeline 9. A discharge valve 10 is provided on the pipeline 9. The discharge valve 10 allows the flow of brake fluid in the direction from the oil pump P toward the pipe 2 and prohibits the flow in the opposite direction.
The position on the master cylinder side of the gate line 1 relative to the gate-out valve 3 of the pipe line 1 and the suction side of the oil pump P are connected by a pipe line 11 and a pipe line 12. A pressure regulating reservoir 13 is provided between the pipe line 11 and the pipe line 12.
A position on the wheel cylinder side of the solenoid valve 6 in the pipeline 2 and the pressure adjusting reservoir 13 are connected by a pipeline 14. The pipeline 14S branches to pipelines 14RL and 14FR, and the pipeline 14P branches to pipelines 14FL and 14RR and is connected to the corresponding wheel cylinder W / C.
A solenoid-out valve 15 that is a normally closed solenoid valve is provided on the pipeline 14.
The pressure adjustment reservoir 13 includes a pressure-sensitive check valve 16. When the pressure in the pipe line 11 exceeds a predetermined pressure, the check valve 16 prohibits the flow of brake fluid into the pressure regulating reservoir 13 so that a high pressure is applied to the suction side of the oil pump P. Is prevented. The check valve 16 is opened regardless of the pressure in the pipe 11 when the oil pump P is activated and the pressure in the pipe 12 becomes low, and the brake fluid to the pressure regulating reservoir 13 is Allow inflow.

[Brake control]
The brake control unit BCU performs anti-lock brake (ABS) control as brake control. When ABS control detects that a wheel has become locked during brake operation by the driver, the wheel cylinder hydraulic pressure is reduced, held, and increased in order to generate the maximum braking force while preventing the wheel from locking. It is a control that repeats the pressure. At the time of ABS pressure reduction control, the solenoid cylinder valve 6 is closed and the solenoid valve 15 is opened from the state shown in FIG. 1 to release the brake fluid of the wheel cylinder W / C to the pressure regulating reservoir 13 to reduce the wheel cylinder fluid pressure. In the ABS holding control, the wheel cylinder hydraulic pressure is held by closing both the solenoid-in valve 6 and the solenoid-out valve 15. In ABS pressure increase control, the solenoid in valve 6 is controlled in the opening direction and the solenoid out valve 15 is closed, and the brake fluid is supplied from the master cylinder M / C to the wheel cylinder W / C to increase the wheel cylinder hydraulic pressure. .
Further, when the hydraulic pressure control unit HU of the first embodiment detects that an oversteer tendency or an understeer tendency has become strong at the time of turning of the vehicle by operating each valve and the oil pump P as a brake control, a predetermined control is performed. Vehicle behavior stability control that stabilizes the vehicle behavior by controlling the wheel cylinder hydraulic pressure of the target wheel, and the wheel cylinder W / C generates a pressure higher than the pressure actually generated in the master cylinder M / C when the driver operates the brake Automatic brake control such as brake assist control to be performed and control to automatically generate a braking force according to the relative relationship with the preceding vehicle by auto cruise control can be performed.

The brake control unit BCU determines the target of the wheel cylinder W / C based on signals from the pressure sensor 17 and other in-vehicle sensors (wheel speed sensor, steering angle sensor, yaw rate sensor, lateral acceleration sensor, etc.) in each brake control described above. Generate hydraulic pressure. Then, each valve and the motor M of the hydraulic pressure control unit HU are driven so that the wheel cylinder hydraulic pressure matches the target hydraulic pressure. The gate-out valve 3 and the solenoid-in valve 6 are PWM controlled with a constant control cycle, and the motor M is ON / OFF controlled with a constant control cycle. The control period of the motor M is set to be sufficiently longer than the PWM control. The solenoid out valve 15 is ON / OFF controlled. The brake control unit BCU includes a motor driving unit 20 that drives the motor M ON / OFF.
Here, in the automatic brake control described above, the desired differential pressure is generated upstream and downstream of the gate-out valve 3 by closing or controlling the gate-out valve 3 in the closing direction (proportional control), and the wheel cylinder hydraulic pressure is set as the target. Match with hydraulic pressure. For this reason, in order to set the differential pressure to a desired value, it is necessary to know the actual differential pressure. Here, although the hydraulic pressure control unit HU of the first embodiment is provided with the pressure sensor 17 on the upstream side of the gate-out valve 3, the downstream side of the gate-out valve 3, that is, the discharge side of the oil pump P Since no pressure sensor is provided, the pressure on the discharge side of the oil pump P (hereinafter referred to as pump discharge side pressure) is estimated based on other parameters. In the first embodiment, the pump discharge side pressure is estimated as described below with the aim of accurately estimating the pump discharge side pressure without adding an additional member such as a pressure sensor. Even when proportional control of the gate-out valve 3 is not performed, it is important for improving the controllability of the wheel cylinder hydraulic pressure that the pump discharge side pressure can be accurately estimated. Although it is possible to estimate the pressure on the discharge side of the hydraulic pressure source from the deceleration generated in the vehicle, there is a problem of a decrease in estimation accuracy.

[Estimation of pump discharge pressure]
FIG. 2 shows a voltage-converted motor rotation speed estimation section (motor actual rotation speed estimation section) 21, a motor OFF time calculation section (drive cycle calculation section) 22 and a pump discharge side pressure estimation section (hydraulic pressure source discharge side pressure) in the brake control unit BCU. FIG. 7 is a control block diagram of an estimation unit 23.
(Voltage conversion motor speed estimation part)
The voltage conversion motor rotation speed estimation unit 21 estimates the voltage conversion motor rotation speed ω _raw from the voltage across the terminals of the motor M (hereinafter referred to as motor end voltage) E. The voltage conversion motor rotation speed estimation unit 21 calculates the voltage conversion motor rotation speed ω _raw by multiplying the motor end voltage E by 1 / K _e . K _e is the counter electromotive force constant. In a DC motor, since the motor rotational speed × the counter electromotive force constant = the motor electromotive force, the motor rotational speed can be obtained if the motor electromotive force and the counter electromotive force constant are known. Here, the motor M is ON / OFF controlled, and the motor electromotive force cannot be obtained when the motor M is ON. However, when the motor M is OFF, the motor end voltage is equal to the motor electromotive force. The motor speed can be obtained from the power constant. The voltage conversion motor rotation speed ω _raw is output to the pump discharge side pressure estimation section 23 and the motor model rotation speed estimation section (motor model rotation speed estimation section) 24 (see FIG. 3).

(Motor OFF time calculation part)
The motor OFF time calculation unit 22 calculates the motor OFF time T _off from the motor drive signal. The motor OFF time calculation unit 22 includes an ON → OFF timing detection unit 22a, an OFF → ON timing detection unit 22b, and a motor OFF time calculation unit 22c. The ON → OFF timing detection unit 22a detects a motor ON → OFF timing f _{ON_OFF} that is a timing at which the motor M is switched from ON to OFF from the motor drive signal. The OFF → ON timing detection unit 22b detects a motor OFF → ON timing f _{OFF_ON} that is a timing at which the motor M is switched from OFF to ON from the motor drive signal. The motor OFF time calculation unit 22c calculates a motor OFF time T _OFF that is an OFF time of the motor M from the motor ON → OFF timing f _{ON_OFF} and the motor OFF → ON timing f _{OFF_ON} . Motor OFF time T _off is outputted to the pump discharge side pressure estimating unit 23, the motor OFF → ON timing f _{OFF_ON} is outputted to the pump discharge side pressure estimating unit 23 and the motor model rotational speed estimation unit 24.

(Pump discharge side pressure estimation part)
The pump discharge side pressure estimation unit 23 is based on the voltage conversion motor rotational speed ω _raw , the estimated motor model rotational speed ω estimated by the motor model rotational speed estimation unit 24, the motor OFF time T _off, and the motor OFF → ON timing f _{OFF_ON.} Thus, the pump discharge side pressure P _est is estimated.
The rotation speed difference calculation block 23a calculates a rotation speed difference Δω by subtracting the voltage conversion motor rotation speed ω _raw from the motor model rotation speed estimation unit 24.
The rotation speed difference change rate calculation block 23b calculates a rotation speed difference change rate Δω / T _off obtained by dividing the rotation speed difference Δω by the motor OFF time T _off .
The motor estimated load correction value change rate calculation block 23c calculates the motor estimated load correction value change rate ΔT _{load_core} by multiplying the rotation speed difference change rate Δω / T _off by the motor estimated load correction coefficient C _corr .
The correction output determination block 23d outputs the motor estimated load correction value change rate ΔT _{load_corr} calculated by the motor estimated load correction value change rate calculation block 23c when the motor OFF → ON timing f _{OFF_ON} is input. _Otherwise , 0 is output as the motor estimated load correction value ΔT _{load_corr} .
Estimated motor load selection block 23e is the estimated motor load correction value T _{Load_corr} by adding the previous value of the corrected output decision block estimated motor load correction value change rate is output from 23d ΔT _{load_corr} and the estimated motor load correction value T _{Load_corr} Calculate.
The previous value calculation block 23f calculates the previous value of the motor estimated load correction value _{Tload_corr} .
The estimated motor load calculation block 23g adds the preset motor load base constant and the estimated motor load correction value T _{load_corr} to calculate the estimated motor load T _load . The estimated motor load T _load is output to the subsequent pressure conversion block 23h and the motor model rotational speed estimation unit 24.
The pressure conversion block 23h calculates the pump discharge side pressure P _est by converting the estimated motor load T _load into a pressure.

(Motor model speed estimation part)
FIG. 3 is a control block diagram of the motor model rotation speed estimation unit 24 in the brake control unit BCU.
The motor model rotational speed estimation unit 24 is preset according to the characteristics of the motor M based on the motor end voltage E, the voltage conversion motor rotational speed ω _raw , the motor OFF → ON timing f _{OFF_ON,} and the motor estimated load T _load . A motor model estimated rotational speed ω is estimated from the motor model.
The motor model rotational speed estimation unit 24 includes a motor rotational speed estimation model 25 and a motor current estimation model 26 as motor models. The motor rotation speed estimation model 25 receives the estimated motor load T _load , the OFF → ON timing f _{OFF_ON} , the voltage conversion motor rotation speed ω _raw, and the estimated motor current I, and outputs the estimated motor model rotation speed ω. Motor current estimation model 26, and inputs the induced electromotive force E _m of the motor end voltage E and the motor M, and outputs the estimated motor current I

First, the motor rotation speed estimation model 25 will be described.
The motor rotational speed estimation model 25 constitutes a model for obtaining the motor model estimated rotational speed ω from the following relational expression representing the mechanical characteristics of the motor M.
T = K _t * I
T _rot = J * dω / dt
T _rot = T-T _f -T _load
= K _t * I-ω * K _f -T _load
Where T is the motor torque, K _t is the motor torque constant, I is the motor estimated current, T _rot is the rotational torque, J is the motor inertia moment, dω / dt is the motor angular acceleration, T _f is the rotational speed dependent torque, T _load is an estimated motor load, and K _f is a motor friction constant.
The rotation speed switching block (motor rotation speed correction unit) 25a outputs the voltage conversion motor rotation speed ω _raw as the motor model estimated rotation speed when the motor OFF → ON timing f _{OFF_ON} is input, and otherwise the input motor The model estimated rotation speed ω is output as it is.
Motor torque calculation block 25b calculates the motor torque T is multiplied by a motor torque constant K _t to the estimated motor current I.
Gain multiplication block 25c calculates the rotational speed-dependent torque T _f is multiplied by a motor friction constant K _f to the motor model estimation rotational speed omega.
The rotational torque calculation block 25d calculates the rotational torque T _rot by subtracting the motor estimated load T _load and the rotational speed dependent torque T _f from the motor torque T.
The motor angular acceleration calculation block 25e calculates the motor angular acceleration dω / dt by dividing the rotational torque T _rot by the motor moment of inertia J.
The motor rotation speed calculation block 25f calculates the motor model estimated rotation speed ω by integrating the motor angular acceleration dω / dt. The motor model estimated rotation speed ω is output to the back electromotive force calculation block 25g and the pump discharge side pressure estimation unit 23 in the subsequent stage.
Counter electromotive force calculation block 25g calculates the counter electromotive force (induced electromotive force) E _m by multiplying the back electromotive force constant K _e to the motor model estimation rotational speed omega.

Next, the motor current estimation model 26 will be described.
The motor current estimation model 26 constitutes a model for obtaining the motor estimated current I from the following relational expression representing the electrical characteristics of the motor M.
E = E _R + E _L + E _m
= RI + L * dI / dt + K _e * ω
Here, E _R is the voltage drop due to the motor resistance, E _L is a voltage drop due to inductance of the motor, the E _m is the counter-electromotive force, dI / dt is the motor current change rate, K _e is a counter electromotive force constant.
The motor resistance voltage calculation unit 26a calculates the motor resistance voltage E _R by multiplying the estimated motor current I by the internal resistance R of the motor M.
Brush voltage drop calculation block 26b calculates a voltage drop E _L due to the inductance of the motor by subtracting the voltage drop E _R by resistance from the motor end voltage E between the back electromotive force E _m.
Motor current change rate calculating unit 26c is a voltage drop E _L of the brush is divided by the self-inductance L and calculates the motor current change rate dI / dt.
The motor current calculation block 26d calculates the motor estimated current I by integrating the motor current change rate dI / dt.

[Pump discharge side pressure calculation processing]
FIG. 4 is a flowchart showing a flow of pump discharge side pressure calculation processing in the brake control unit BCU. Each step will be described below.
In step S1, the motor model rotational speed estimation unit 24 estimates the motor model estimated rotational speed ω.
In step S2, the voltage conversion motor rotation speed estimation unit 21 estimates the voltage conversion motor rotation speed ω _raw .
In step S3, the motor OFF → ON timing f _{OFF_ON} is detected by the OFF → ON timing detection unit 22b of the motor OFF time calculation unit 22.
In step S4, the ON → OFF timing detection unit 22a of the motor OFF time calculation unit 22 detects the motor ON → OFF timing f _{ON_OFF} .
In step S5, the motor OFF time calculation unit 22c of the motor OFF time calculation unit 22 calculates the motor OFF time T _off.
In step S6, the correction output determination block 23d of the pump discharge side pressure estimation unit 23 determines whether or not the motor OFF → ON timing f _{OFF_ON} has been input. If YES, the process proceeds to step S7. If NO, the process proceeds to step S8.
In step S7, the motor estimated load correction value T _{load_corr} is calculated in the motor estimated load selection block 23e of the pump discharge side pressure estimating unit 23. Since the motor OFF → ON timing, the motor estimated load correction value T _{load_corr} is a value _{obtained by adding} the motor estimated load correction value change rate ΔT _{load_corr} and the previous value of the motor estimated load correction value T _{load_corr} .
In step S8, a motor estimated load correction value T _{load_corr} is calculated in the motor estimated load selection block 23e of the pump discharge side pressure estimating unit 23. Since it is not the motor OFF → ON timing, the motor estimated load correction value T _{load_corr} is the previous value.
In step S9, in the rotation speed switching block 25a of the motor rotation speed estimation model 25, the voltage conversion motor rotation speed ω _raw is output as the motor model estimated rotation speed.
In step S10, the estimated motor load T _load is calculated in the estimated motor load calculation block 23g of the pump discharge side pressure estimation unit 23.
In step S11, the pump discharge side pressure P _est is calculated in the pressure conversion block 23h of the pump discharge side pressure estimation unit 23.

Next, the operation will be described.
[Pump discharge side pressure estimation action]
The motor electromotive force of the DC motor can be obtained by multiplying the motor rotational speed and the counter electromotive force constant. In other words, if the motor electromotive force and the electromotive force constant are known, the motor rotation speed can be known. In general motor ON / OFF control, a switch is provided in a part of a circuit, and the motor is controlled by turning it ON / OFF. Although the motor electromotive force cannot be known when the switch is turned on, the motor rotational speed can be known because the motor electromotive force is known by measuring the voltage between the motor terminals when the switch is turned off.
Therefore, in the first embodiment, the motor model rotational speed estimation unit 24 uses a motor M plant model (motor rotational speed estimation model 25 and motor current estimation model 26) having a parameter of the motor estimated load T _load , and the motor rotational speed. (Motor model estimated rotation speed ω) is estimated. Further, the voltage conversion motor rotation speed estimation unit 21 calculates the actual motor rotation speed (voltage conversion motor rotation speed ω _raw ) while the switch is OFF.
In the pump discharge side pressure estimation unit 23, the difference (rotational speed difference Δω) between the motor model estimated rotational speed ω and the voltage-converted motor rotational speed ω _raw before the motor M is turned OFF → ON is estimated as an estimation error of the motor estimated load T _load. Then, the motor estimated load correction value change rate ΔT _{load_corr} is calculated by multiplying the rotational speed difference Δω by the motor OFF time T _off and the motor estimated load correction coefficient C _corr , and the estimated motor load in the model according to ΔT _{load_corr Increase or decrease the} correction value T _{load_corr} . That is, by correcting the motor estimated load T _load so as to eliminate the rotation speed difference Δω, the motor estimated load T _load can be brought close to the actual motor load. Further, the motor model estimated rotational speed ω is matched with the voltage conversion motor rotational speed ω _{raw at the} timing of motor OFF → ON.
In the oil pump P, the motor torque changes in proportion to the pump discharge pressure. Since the pump discharge pressure is proportional to the pump discharge side pressure P _est , the pump discharge side pressure is calculated by substituting the estimated motor load T _load obtained by adding the motor estimated load correction value T _{load_corr} to the motor load base constant into the pressure conversion formula. P _est can be estimated.

FIG. 5 is a time chart of the motor rotation speed, the motor electromotive force, and the motor estimated load according to the first embodiment.
As described above, when the motor is ON, the motor end voltage is not equal to the motor electromotive force, but while the motor is OFF, the motor end voltage is equal to the motor electromotive force. The number ω _raw ) almost coincides with the actual motor speed.
The motor model estimated rotation speed ω estimated from the motor model is processed so as to match the voltage-converted motor rotation speed ω _{raw at} each motor OFF → ON timing, but if the estimated motor load T _load is different from the actual motor load, the motor During OFF, a rotational speed difference Δω is generated between ω and ω _raw . Therefore, the estimated motor load T _load is corrected by the estimated motor load correction value T _{load_corr} corresponding to Δω. When ω> ω _raw , that is, when Δω is a positive value, the motor estimated load T _load is smaller than the actual motor load. _Therefore , the motor is estimated by the positive motor estimated load correction value T _{load_corr} corresponding to Δω. The load T _load is increased and corrected. On the other hand, when ω <ω _raw , that is, when Δω is a negative value, the estimated motor load T _load is larger than the actual motor load. _Therefore, the negative estimated motor load correction value T _{load_corr} corresponding to Δω Reduce the motor estimated load T _load . As a result, the estimation accuracy of the estimated motor load T _load is improved, so that the pump discharge side pressure P _est obtained by converting the estimated motor load T _load into a pressure can be accurately estimated. Moreover, the cost increase by addition of additional members, such as a pressure sensor, can be suppressed. As shown in FIG. 6, in the first embodiment, the pump discharge side pressure P _est can be made to follow the actual pump discharge side pressure with a slight delay.

The pump discharge side pressure estimation unit 23 calculates the rotational speed difference change rate Δω / T _off by dividing the rotational speed difference Δω between the motor model estimated rotational speed ω and the voltage conversion motor rotational speed ω _raw by the motor OFF time T _off. The motor estimated load correction value change rate ΔT _{load_corr} is calculated by multiplying the rotational speed difference change rate Δω / T _off by the motor estimated load correction coefficient C _corr . By dividing the rotational speed difference Δω by the motor OFF time T _off , variations in the rotational speed difference Δω during the motor OFF time can be suppressed, so that the estimation accuracy of the motor estimated load T _load can be improved.
When the motor OFF → ON timing f _{OFF_ON} is input, the rotation speed switching block 25a outputs the voltage conversion motor rotation speed ω _raw as the motor model estimated rotation speed ω. When the motor is ON, the motor end voltage is not equal to the motor electromotive force. However, since the motor end voltage is equal to the motor electromotive force at the motor OFF → ON timing, the motor model estimated rotational speed ω matches the actual motor rotational speed. The estimation accuracy of the model rotational speed ω can be improved. Since the motor end voltage is equal to the motor electromotive force even when the motor is OFF, the actual rotational speed may be corrected whenever the motor is OFF, but the timing immediately before turning ON the motor M is the largest and the difference of Δω coming out. Therefore, by correcting the motor model estimated rotation speed ω with the voltage-converted motor rotation speed ω _raw at the motor OFF → ON timing, it is possible to suppress a decrease in estimation accuracy of the motor estimated load T _load .

Next, the effect will be described.
In Example 1, the following effects are exhibited.
(1) A hydraulic pressure source (oil pump P) for increasing the hydraulic pressure of the wheel cylinder by sucking the brake fluid in the master cylinder M / C or the pressure regulating reservoir 13 and discharging it to the wheel cylinder W / C; A motor M that drives the pressure source, a motor drive unit 20 that drives the motor M ON / OFF, and a voltage-converted motor rotational speed estimator that estimates the voltage-converted motor rotational speed ω _raw based on the motor end voltage generated at the time of the motor MOFF 21, a motor model rotational speed estimation unit 24 that estimates a motor model estimated rotational speed ω from a preset motor model based on the characteristics of the motor M, a motor model estimated rotational speed ω, a voltage-converted motor rotational speed ω _raw , And a pump discharge side pressure estimation unit 23 that estimates a motor estimated load T _load based on the rotation speed difference Δω and estimates a pump discharge side pressure P _est based on the motor estimated load T _load .
Therefore, it is possible to accurately estimate the pump discharge side pressure P _est without adding an additional member.
(2) A motor OFF time calculation unit 22 that calculates the motor OFF time T _off is provided, and the pump discharge side pressure estimation unit 23 calculates a motor estimated load T _load based on the rotation speed difference Δω and the motor OFF time T _off .
Therefore, the estimation accuracy of the motor estimated load T _load can be improved.
(3) A rotational speed switching block 25a for correcting the motor model estimated rotational speed ω estimated by the motor model rotational speed estimating unit 24 based on the voltage converted motor rotational speed ω _raw estimated by the voltage converted motor rotational speed estimating unit 21 Prepared.
Therefore, the estimation accuracy of the motor model rotation speed ω can be improved.
(4) The hydraulic pressure source is the oil pump P.
Therefore, since the pressure at the time of pump-up can be estimated with high accuracy, the control accuracy of brake control can be improved.

[Other Examples]
As mentioned above, although the form for implementing this invention was demonstrated based on the Example, the concrete structure of this invention is not limited to the structure shown in the Example, and is the range which does not deviate from the summary of invention. Any design changes are included in the present invention.
The present invention can accurately estimate the pump discharge side pressure without providing a pressure sensor, but is also useful in a hydraulic control unit including a pressure sensor for detecting the pump discharge side pressure. In other words, by using the sensor value of the pressure sensor and using the pump discharge side pressure obtained by the estimation of the present invention when the pressure sensor fails, the brake control can be continued even when the pressure sensor fails, and the reliability of the device It can contribute to improvement.
It is good also as a structure which provided the rotation speed sensor as a motor actual rotation speed estimation part which estimates the actual rotation speed of a motor.

M motor
M / C master cylinder
P Oil pump (hydraulic pressure source)
RSV reservoir tank (reservoir)
W / C wheel cylinder
20 Motor drive
21 Voltage conversion motor speed estimator (motor actual speed estimator)
22 Motor OFF time calculation unit (drive cycle calculation unit)
23 Pump discharge side pressure estimation part (hydraulic pressure source discharge side pressure estimation part)
24 Motor model rotation speed estimation section (Motor model rotation speed estimation section)
25a Speed change block (Motor speed correction part)

Claims (4)

A fluid pressure source for increasing the fluid pressure of the wheel cylinder by sucking the brake fluid in the master cylinder or reservoir and discharging it to the wheel cylinder;
A motor for driving the hydraulic pressure source;
A motor drive unit for driving the motor ON / OFF;
A motor actual rotational speed estimation unit that estimates an actual rotational speed of the motor based on a voltage between terminals of the motor;
A motor model rotational speed estimation unit that estimates the model rotational speed of the motor from a preset motor model based on the characteristics of the motor;
A load applied to the motor is estimated based on a difference between the model rotational speed estimated by the motor model rotational speed estimation unit and the actual rotational speed estimated by the motor actual rotational speed estimation unit, and the liquid is calculated based on the estimated load. A hydraulic pressure source discharge side pressure estimation unit for estimating the pressure on the discharge side of the pressure source;
A brake control device comprising: The brake control device according to claim 1, wherein
A drive cycle calculation unit that calculates the ON / OFF drive cycle of the motor,
The hydraulic pressure source discharge side pressure estimation unit calculates a load applied to the motor based on the difference and the driving cycle calculated by the driving cycle calculation unit. The brake control device according to claim 1 or 2,
A brake control device comprising: a motor rotation number correction unit that corrects the model rotation number estimated by the motor model rotation number estimation unit based on the actual rotation number estimated by the motor actual rotation number estimation unit. The brake control device according to any one of claims 1 to 3,
The brake control device, wherein the hydraulic pressure source is a pump.

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