WO2013133059A1 - Appareil de commande de véhicule, et procédé de commande de véhicule - Google Patents

Appareil de commande de véhicule, et procédé de commande de véhicule Download PDF

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Publication number
WO2013133059A1
WO2013133059A1 PCT/JP2013/054681 JP2013054681W WO2013133059A1 WO 2013133059 A1 WO2013133059 A1 WO 2013133059A1 JP 2013054681 W JP2013054681 W JP 2013054681W WO 2013133059 A1 WO2013133059 A1 WO 2013133059A1
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WIPO (PCT)
Prior art keywords
control
damping force
vehicle
sprung
unsprung
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PCT/JP2013/054681
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English (en)
Japanese (ja)
Inventor
宏信 菊池
平原　道人
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日産自動車株式会社
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Priority to JP2014503768A priority Critical patent/JP5979221B2/ja
Publication of WO2013133059A1 publication Critical patent/WO2013133059A1/fr

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    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G17/00—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load
    • B60G17/015—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G17/00—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load
    • B60G17/015—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements
    • B60G17/018—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements characterised by the use of a specific signal treatment or control method
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G17/00—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load
    • B60G17/06—Characteristics of dampers, e.g. mechanical dampers
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2400/00—Indexing codes relating to detected, measured or calculated conditions or factors
    • B60G2400/10—Acceleration; Deceleration
    • B60G2400/102—Acceleration; Deceleration vertical
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2400/00—Indexing codes relating to detected, measured or calculated conditions or factors
    • B60G2400/20—Speed
    • B60G2400/208—Speed of wheel rotation
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2400/00—Indexing codes relating to detected, measured or calculated conditions or factors
    • B60G2400/90—Other conditions or factors
    • B60G2400/91—Frequency
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2500/00—Indexing codes relating to the regulated action or device
    • B60G2500/10—Damping action or damper

Definitions

  • the present invention relates to a control device that controls the state of a vehicle.
  • Patent Document 1 A technique described in Patent Document 1 is disclosed as a technique related to a vehicle control device.
  • This publication discloses a technique for setting a damping force in accordance with the unsprung speed in a suspension control device capable of changing the damping force.
  • the present invention has been made paying attention to the above problem, and an object of the present invention is to provide a vehicle control device capable of ensuring the grounding property of wheels while ensuring the ride comfort.
  • the vehicle control apparatus calculates the damping force control amount of the damping force variable shock absorber according to the unsprung vertical speed, and the amplitude of an arbitrary frequency band of the sprung vertical acceleration.
  • the damping force control amount is reduced and corrected.
  • FIG. 1 is a system schematic diagram illustrating a vehicle control apparatus according to a first embodiment.
  • FIG. 2 is a control block diagram illustrating a control configuration of the vehicle control device according to the first embodiment.
  • 1 is a conceptual diagram illustrating a configuration of a wheel speed feedback control system according to a first embodiment.
  • FIG. 3 is a control block diagram illustrating a configuration of a traveling state estimation unit according to the first embodiment. It is a control block diagram showing the control content in the stroke speed calculating part of Example 1.
  • FIG. 3 is a block diagram illustrating a configuration of a reference wheel speed calculation unit according to the first embodiment. It is the schematic showing a vehicle body vibration model. It is a control block diagram showing brake pitch control of Example 1.
  • FIG. 3 is a control block diagram illustrating a configuration of roll rate suppression control according to the first embodiment. 3 is a time chart illustrating an envelope waveform forming process of roll rate suppression control according to the first embodiment.
  • FIG. 3 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the first embodiment.
  • FIG. 3 is a control block diagram illustrating a control configuration of a damping force control unit according to the first embodiment.
  • 6 is a flowchart illustrating attenuation coefficient arbitration processing in a standard mode according to the first embodiment.
  • 6 is a flowchart illustrating an attenuation coefficient arbitration process in the sport mode according to the first embodiment.
  • 6 is a flowchart illustrating attenuation coefficient arbitration processing in the comfort mode according to the first embodiment.
  • 6 is a flowchart illustrating attenuation coefficient arbitration processing in a highway mode according to the first exemplary embodiment.
  • FIG. 10 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to a fifth embodiment.
  • FIG. 1 is a system schematic diagram illustrating a vehicle control apparatus according to the first embodiment.
  • the vehicle includes an engine 1 that is a power source and a brake 20 that generates braking torque due to friction force on each wheel (hereinafter, when displaying brakes corresponding to individual wheels, right front wheel brake: 20FR, left front wheel brake: 20FL).
  • S / A shock absorber 3
  • the engine 1 includes an engine controller (hereinafter also referred to as an engine control unit, which corresponds to power source control means) 1a that controls torque output from the engine 1, and the engine controller 1a is configured to By controlling the fuel injection amount, ignition timing, etc., the desired engine operating state (engine speed and engine output torque) is controlled. Further, the brake 20 generates a braking torque based on the hydraulic pressure supplied from the brake control unit 2 that can control the brake hydraulic pressure of each wheel according to the traveling state.
  • the brake control unit 2 includes a brake controller (hereinafter also referred to as a brake control unit) 2a for controlling a braking torque generated by the brake 20, and a master cylinder pressure generated by a driver's brake pedal operation or a built-in motor.
  • a pump pressure generated by the drive pump is used as a hydraulic pressure source, and a desired hydraulic pressure is generated in the brake 20 of each wheel by opening and closing operations of a plurality of solenoid valves.
  • the S / A3 is a damping force generator that attenuates the elastic motion of a coil spring provided between a vehicle unsprung (axle, wheel, etc.) and a sprung (vehicle body, etc.). It is configured to be variable.
  • the S / A 3 includes a cylinder in which fluid is sealed, a piston that strokes in the cylinder, and an orifice that controls fluid movement between fluid chambers formed above and below the piston. Furthermore, orifices having a plurality of types of orifice diameters are formed in the piston, and an orifice corresponding to a control command is selected from the plurality of types of orifices when the S / A actuator is operated. Thereby, the damping force according to the orifice diameter can be generated. For example, if the orifice diameter is small, the movement of the piston is easily restricted, so that the damping force is high. If the orifice diameter is large, the movement of the piston is difficult to be restricted, and thus the damping force is small.
  • an electromagnetic control valve is arranged on the communication path connecting fluids formed above and below the piston, and the damping force is set by controlling the opening / closing amount of the electromagnetic control valve.
  • the S / A 3 has an S / A controller 3a (corresponding to damping force control means) that controls the damping force of the S / A 3, and controls the damping force by operating the orifice diameter by the S / A actuator.
  • a wheel speed sensor 5 for detecting the wheel speed of each wheel (hereinafter, when displaying wheel speeds corresponding to individual wheels, right front wheel speed: 5FR, left front wheel speed: 5FL, right rear wheel speed: 5RR). , Left rear wheel speed: 5RL)), an integrated sensor 6 for detecting longitudinal acceleration, yaw rate and lateral acceleration acting on the center of gravity of the vehicle, and a steering angle which is a steering operation amount of the driver is detected.
  • Steering angle sensor 7 vehicle speed sensor 8 for detecting vehicle speed
  • engine torque sensor 9 for detecting engine torque
  • engine speed sensor 10 for detecting engine speed
  • master pressure sensor 11 for detecting master cylinder pressure.
  • a brake switch 12 that outputs an on-state signal when a brake pedal operation is performed, and an accelerator opening sensor 13 that detects an accelerator pedal opening.
  • the signals from these various sensors are input to the engine controller 1a, the brake controller 2a, and the S / A controller 3a as necessary.
  • the arrangement of the integrated sensor 6 may be at the center of gravity of the vehicle, or may be any place other than that as long as various values at the center of gravity can be estimated. Moreover, it is not necessary to be an integral type, and a configuration in which yaw rate, longitudinal acceleration, and lateral acceleration are individually detected may be employed.
  • the control amount by the engine 1 and the brake 20 is limited and output from the control amount that can be actually output, thereby reducing the burden on the S / A 3 and accompanying the control of the engine 1 and the brake 20. Suppresses discomfort that occurs.
  • Skyhook control is performed by all actuators. At this time, without using a stroke sensor or a sprung vertical acceleration sensor generally required for skyhook control, the skyhook control can be performed with an inexpensive configuration using wheel speed sensors mounted on all vehicles. Realize.
  • scalar control frequency sensitive control
  • FIG. 2 is a control block diagram illustrating a control configuration of the vehicle control apparatus according to the first embodiment.
  • the controller includes an engine controller 1a, a brake controller 2a, and an S / A controller 3a, and each controller constitutes a wheel speed feedback control system.
  • the configuration including three controllers as the controller is shown, but each controller may be configured by one integrated controller without any particular limitation.
  • the configuration including the three controllers in the first embodiment is that the engine controller and the brake controller in the existing vehicle are used as they are to form the engine control unit 1a and the brake control unit 2a, and the S / A controller 3a is separately mounted. Thus, it is assumed that the vehicle control apparatus of the first embodiment is realized.
  • the engine controller 1a mainly uses the wheel speed detected by the wheel speed sensor 5 to determine the stroke speed, bounce rate, roll rate and pitch of each wheel used for the skyhook control of the sprung mass damping control unit 101a described later.
  • a first running state estimation unit 100 that estimates the rate, an engine posture control unit 101 that calculates an engine posture control amount that is an engine torque command, and controls the operating state of the engine 1 based on the calculated engine posture control amount.
  • An engine control unit 102 The details of the estimation process of the first traveling state estimation unit 100 will be described later.
  • the engine attitude control unit 101 includes a sprung mass damping control unit 101a that calculates a sprung control amount that suppresses bounce motion and pitch motion by skyhook control, and ground load variation suppression that suppresses ground load variation of front and rear wheels.
  • a ground load control unit 101b that calculates a control amount
  • an engine-side driver input control unit 101c that calculates a yaw response control amount corresponding to a vehicle behavior that the driver wants to achieve based on signals from the steering angle sensor 7 and the vehicle speed sensor 8 And have.
  • the engine attitude control unit 101 calculates an engine attitude control amount that minimizes the control amount calculated by each of these control units by optimal control (LQR), and determines the final engine attitude control amount for the engine control unit 102. Output.
  • LQR optimal control
  • the S / A 3 can reduce the damping force control amount, and therefore, deterioration of the high frequency vibration can be avoided. Moreover, since S / A3 can concentrate on suppression of roll motion, it can suppress roll motion effectively.
  • the brake controller 2a Based on the wheel speed detected by the wheel speed sensor 5, the brake controller 2a estimates the stroke speed and pitch rate of each wheel, and the like based on the estimated stroke speed and pitch rate.
  • Skyhook control unit 201 (details will be described later) that calculates a brake attitude control amount based on skyhook control, and brake control unit 202 that controls the braking torque of brake 20 based on the calculated brake attitude control amount And have.
  • the same estimation process is adopted as the estimation process in the first traveling state estimation unit 100 and the second traveling state estimation unit 200, but other estimation processes are performed as long as the process is estimated from the wheel speed. It may be used.
  • the S / A controller 3a includes a driver input control unit 31 that performs driver input control for achieving a desired vehicle posture based on a driver's operation (steering operation, accelerator operation, brake pedal operation, etc.), and detection values of various sensors.
  • a third traveling state estimation unit 32 that estimates the traveling state based on (mainly the wheel speed sensor value of the wheel speed sensor 5), and a sprung mass damping that controls the vibration state on the spring based on the estimated traveling state
  • a control unit 33 an unsprung vibration suppression control unit 34 that controls the unsprung vibration state based on the estimated traveling state, a shock absorber attitude control amount output from the driver input control unit 31, and a sprung mass damping
  • a damping force to be set in the S / A 3 is determined.
  • a damping force control unit 35 for performing the damping force control of the A.
  • the same estimation process is adopted as the estimation process in the first traveling state estimation unit 100, the second traveling state estimation unit 200, and the third traveling state estimation unit 32, but the process is estimated from the wheel speed. If so, other estimation processes may be used and there is no particular limitation.
  • FIG. 3 is a conceptual diagram showing the configuration of the wheel speed feedback control system of the first embodiment.
  • the engine 1, the brake 20 and the S / A 3 individually constitute an engine feedback control system, a brake feedback control system, and an S / A feedback control system.
  • control interference becomes a problem.
  • the effects of the control of each actuator appear as wheel speed fluctuations, by configuring the wheel speed feedback control system, the effect of each actuator is monitored as a result, and control interference is avoided. It is. For example, if a certain sprung vibration is suppressed by the engine 1, the wheel speed fluctuation
  • the brake 20 and the S / A 3 perform control based on the wheel speed in which the influence is reflected.
  • the feedback control system is configured using a common value of wheel speed, even if individual control is performed without controllable mutual monitoring, as a result, control based on mutual monitoring (below) This control is described as emphasis control), and the vehicle posture can be converged in the stabilization direction.
  • each feedback control system will be described sequentially.
  • the 1st, 2nd, 3rd driving state estimation part which is a common structure provided in each feedback control system is demonstrated.
  • the same estimation process is adopted as the estimation process in the first traveling state estimation unit 100, the second traveling state estimation unit 200, and the third traveling state estimation unit 32. Therefore, since the process in each estimation part is common, the estimation process in the 3rd driving state estimation part 32 is demonstrated as a representative.
  • Each of the running state estimation units may be provided with a separate estimation model as long as it is a state estimation using the wheel speed, and is not particularly limited.
  • FIG. 4 is a control block diagram showing the configuration of the third traveling state estimation unit of the first embodiment.
  • the stroke of each wheel used for the skyhook control of the sprung mass damping control unit 33 to be described later is basically based on the wheel speed detected by the wheel speed sensor 5. Calculate speed, bounce rate, roll rate and pitch rate. First, the value of the wheel speed sensor 5 of each wheel is input to the stroke speed calculation unit 321, and the sprung speed is calculated from the stroke speed of each wheel calculated by the stroke speed calculation unit 321.
  • FIG. 5 is a control block diagram showing the control contents in the stroke speed calculation unit of the first embodiment.
  • the stroke speed calculation unit 321 is individually provided for each wheel, and the control block diagram shown in FIG. 5 is a control block diagram focusing on a certain wheel.
  • the value of the wheel speed sensor 5, the front wheel steering angle ⁇ f detected by the steering angle sensor 7, and the rear wheel steering angle ⁇ r (actual rear wheel steering if a rear wheel steering device is provided).
  • the reference wheel speed calculation unit 300 that calculates a reference wheel speed based on the vehicle body lateral speed and the actual yaw rate detected by the integrated sensor 6, and the angle may be appropriately set to 0 in other cases.
  • a tire rotation vibration frequency calculation unit 321a that calculates the tire rotation vibration frequency based on the calculated reference wheel speed, and a deviation calculation unit 321b that calculates a deviation (wheel speed fluctuation) between the reference wheel speed and the wheel speed sensor value.
  • a GEO conversion unit 321c that converts the deviation calculated by the deviation calculation unit 321b into a suspension stroke amount, a stroke speed calibration unit 321d that calibrates the converted stroke amount to a stroke speed,
  • a band elimination filter corresponding to the frequency calculated by the tire rotation vibration frequency calculation unit 321a is applied to the value calibrated by the roke speed calibration unit 321d to remove the tire rotation primary vibration component and calculate the final stroke speed.
  • a signal processing unit 321e that calculates the tire rotation vibration frequency based on the calculated reference wheel speed
  • a deviation calculation unit 321b that calculates a deviation (wheel speed fluctuation) between the reference wheel speed and the wheel speed sensor value.
  • a GEO conversion unit 321c that converts the deviation calculated by the deviation calculation unit 321b
  • FIG. 6 is a block diagram illustrating a configuration of a reference wheel speed calculation unit according to the first embodiment.
  • the reference wheel speed refers to a value obtained by removing various disturbances from each wheel speed.
  • the difference between the wheel speed sensor value and the reference wheel speed is a value related to a component that fluctuates according to the stroke generated by the bounce behavior, roll behavior, pitch behavior, or unsprung vertical vibration of the vehicle body.
  • the stroke speed is estimated based on this difference.
  • the plane motion component extraction unit 301 calculates the first wheel speed V0 that is the reference wheel speed of each wheel based on the vehicle body plan view model with the wheel speed sensor value as an input.
  • the wheel speed sensor value detected by the wheel speed sensor 5 is ⁇ (rad / s)
  • the front wheel actual steering angle detected by the steering angle sensor 7 is ⁇ f (rad)
  • the rear wheel actual steering angle is ⁇ r (rad )
  • the vehicle body lateral speed is Vx
  • the yaw rate detected by the integrated sensor 6 is ⁇ (rad / s)
  • the vehicle speed estimated from the calculated reference wheel speed ⁇ 0 is V (m / s)
  • the reference to be calculated Wheel speed is VFL, VFR, VRL, VRR
  • front wheel tread is Tf
  • rear wheel tread is Tr
  • distance from vehicle center of gravity to front wheel is Lf
  • distance from vehicle center of gravity to rear wheel is Lr.
  • VFL (V-Tf / 2 ⁇ ⁇ ) cos ⁇ f + (Vx + Lf ⁇ ⁇ ) sin ⁇ f
  • VFR (V + Tf / 2 ⁇ ⁇ ) cos ⁇ f + (Vx + Lf ⁇ ⁇ ) sin ⁇ f
  • VRL (V ⁇ Tr / 2 ⁇ ⁇ ) cos ⁇ r + (Vx ⁇ Lr ⁇ ⁇ ) sin ⁇ r
  • VRR (V + Tr / 2 ⁇ ⁇ ) cos ⁇ r + (Vx-Lr ⁇ ⁇ ) sin ⁇ r
  • V is described as V0FL, V0FR, V0RL, V0RR (corresponding to the first wheel speed) as a value corresponding to each wheel.
  • V0FL ⁇ VFL-Lf ⁇ ⁇ sin ⁇ f ⁇ / cos ⁇ f + Tf / 2 ⁇ ⁇
  • V0FR ⁇ VFR-Lf ⁇ ⁇ sin ⁇ f ⁇ / cos ⁇ f-Tf / 2 ⁇ ⁇
  • V0RL ⁇ VRL + Lr ⁇ ⁇ sin ⁇ r ⁇ / cos ⁇ r + Tr / 2 ⁇ ⁇
  • V0RR ⁇ VRR + Lf ⁇ ⁇ sin ⁇ f ⁇ / cos ⁇ r-Tr / 2 ⁇ ⁇
  • the roll disturbance removing unit 302 calculates the second wheel speeds V0F and V0R as the reference wheel speeds for the front and rear wheels based on the vehicle body front view model with the first wheel speed V0 as an input.
  • the vehicle body front view model removes the wheel speed difference caused by the roll motion that occurs around the roll rotation center on the vertical line passing through the center of gravity of the vehicle when the vehicle is viewed from the front. Is done.
  • V0F (V0FL + V0FR) / 2
  • V0R (V0RL + V0RR) / 2
  • the second wheel speeds V0F and V0R from which disturbance based on the roll is removed are obtained.
  • the pitch disturbance removal unit 303 calculates the third wheel speeds VbFL, VbFR, VbRL, and VbRR, which are the reference wheel speeds for all the wheels, based on the vehicle side view model, with the second wheel speeds V0F and V0R as inputs.
  • the vehicle body side view model is to remove the wheel speed difference caused by the pitch motion generated around the pitch rotation center on the vertical line passing through the center of gravity of the vehicle when the vehicle is viewed from the lateral direction. It is expressed by the following formula.
  • the sprung speed calculation unit 322 calculates the bounce rate, roll rate, and pitch rate for skyhook control. Calculated.
  • Skyhook control is to achieve a flat running state by setting a damping force based on the relationship between the S / A3 stroke speed and the sprung speed, and controlling the posture on the sprung.
  • the value that can be detected from the wheel speed sensor 5 is the stroke speed, and since the vertical acceleration sensor or the like is not provided on the spring, the sprung speed needs to be estimated using an estimation model.
  • the problem of the estimation model and the model configuration to be adopted will be described.
  • FIG. 7 is a schematic diagram showing a vehicle body vibration model.
  • FIG. 7A is a model of a vehicle (hereinafter referred to as a conveyor vehicle) having an S / A with a constant damping force
  • FIG. 7B has an S / A having a variable damping force.
  • Ms represents the mass on the spring
  • Mu represents the mass below the spring
  • Ks represents the elastic coefficient of the coil spring
  • Cs represents the damping coefficient of S / A
  • Ku represents the unsprung (tire).
  • Cu represents an unsprung (tire) damping coefficient
  • Cv represents a variable damping coefficient
  • Z2 represents a position on the spring
  • z1 represents a position under the spring
  • z0 represents a road surface position.
  • Changing the damping force basically means changing the force that limits the piston moving speed of S / A 3 in accordance with the suspension stroke. Since the semi-active S / A3 that cannot positively move the piston in the desired direction is used, when the semi-active skyhook model is employed and the sprung speed is obtained, it is expressed as follows.
  • the magnitude of the estimated sprung speed is smaller than the actual value in the frequency band below the sprung resonance, but the most important in skyhook control is the phase. If the correspondence between the phase and the sign can be maintained, the skyhook can be maintained. Since control is achieved and the magnitude of the sprung speed can be adjusted by other factors, there is no problem.
  • the sprung speed can be estimated if the stroke speed of each wheel is known.
  • the actual vehicle is four wheels instead of one wheel, it is considered to estimate the state of the spring by mode decomposition into roll rate, pitch rate and bounce rate using the stroke speed of each wheel. To do.
  • the above three components are calculated from the stroke speed of the four wheels, one corresponding component is insufficient, and the solution becomes indefinite. Therefore, a war plate representing the movement of the diagonal wheels is introduced.
  • the stroke amount bounce term is xsB
  • the roll term is xsR
  • the pitch term is xsP
  • the warp term is xsW
  • the stroke amount corresponding to Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR is z_sFL, z_sFR, z_sRL, z_sRR, Holds.
  • dxsB 1/4 (Vz_sFL + Vz_sFR + Vz_sRL + Vz_sRR)
  • dxsR 1/4 (Vz_sFL-Vz_sFR + Vz_sRL-Vz_sRR)
  • dxsP 1/4 (-Vz_sFL-Vz_sFR + Vz_sRL + Vz_sRR)
  • dxsW 1/4 (-Vz_sFL + Vz_sFR + Vz_sRL-Vz_sRR)
  • the vehicle control apparatus includes the engine 1, the brake 20, and the S / A 3 as actuators for achieving sprung posture control.
  • the sprung mass damping control unit 101a in the engine controller 1a has two bounce rate and pitch rate as control targets
  • the skyhook control unit 201 in the brake controller 2a has pitch rate as control targets.
  • the skyhook control unit 33a in the controller 3a three of bounce rate, roll rate, and pitch rate are controlled.
  • the bounce direction skyhook control amount FB is calculated as a part of the engine attitude control amount in the sprung mass damping control unit 101a.
  • the skyhook control unit 33a calculates as a part of the S / A attitude control amount.
  • the skyhook control amount FR in roll direction is calculated as part of the S / A attitude control amount in the sky hook control unit 33a.
  • the sky hook control amount FP in the pitch direction is calculated as a part of the engine attitude control amount in the sprung mass damping control unit 101a.
  • the skyhook control unit 201 calculates the brake posture control amount.
  • the skyhook control unit 33a calculates as a part of the S / A attitude control amount.
  • the engine attitude control unit 101 is set with a limit value for limiting the engine torque control amount according to the engine attitude control amount so as not to give the driver a sense of incongruity.
  • the engine torque control amount is limited to be within a predetermined longitudinal acceleration range when converted to longitudinal acceleration. Therefore, when the engine attitude control amount (engine torque control amount) is calculated based on FB or FP and a value equal to or greater than the limit value is calculated, bounce rate or pitch rate skyhook control that can be achieved by the limit value
  • the engine attitude control amount is output as a quantity.
  • an engine torque control amount is calculated based on the engine attitude control amount corresponding to the limit value and is output to the engine 1.
  • a limit value for limiting the braking torque control amount is set in order to prevent the driver from feeling uncomfortable as in the case of the engine 1 (details of the limit value will be described later).
  • the braking torque control amount is converted into the longitudinal acceleration, the braking torque control amount is limited to be within a predetermined longitudinal acceleration range (a limit value obtained from the occupant's uncomfortable feeling, the life of the actuator, etc.). Therefore, when the brake attitude control amount is calculated based on the FP and a value equal to or greater than the limit value is calculated, a pitch rate suppression amount (hereinafter referred to as a brake attitude control amount) that can be achieved by the limit value.
  • a pitch rate suppression amount hereinafter referred to as a brake attitude control amount
  • the brake control unit 202 calculates a braking torque control amount (or deceleration) based on the brake attitude control amount corresponding to the limit value, and outputs the calculated braking torque control amount to the brake 20.
  • FIG. 8 is a control block diagram showing the brake pitch control of the first embodiment.
  • the vehicle body mass is m
  • the front wheel braking force is BFf
  • the rear wheel braking force is BFr
  • the height between the vehicle center of gravity and the road surface is Hcg
  • the vehicle acceleration is a
  • the pitch moment is Mp
  • the pitch rate is Vp.
  • the brake attitude control amount calculation unit 334 is composed of the following control blocks.
  • the dead zone processing code determination unit 3341 determines the sign of the input pitch rate Vp, and when it is positive, it outputs 0 to the deceleration reduction processing unit 3342 because control is unnecessary, and when it is negative, it determines that control is possible.
  • the pitch rate signal is output to the deceleration reduction processing unit 3342.
  • the deceleration feeling reduction process is a process corresponding to the limit by the limit value performed in the brake attitude control amount calculation unit 334.
  • the square processor 3342a squares the pitch rate signal. This inverts the sign and smoothes the rise of the control force.
  • the pitch rate square decay moment calculation unit 3342b calculates the pitch moment Mp by multiplying the squared pitch rate by the skyhook gain CskyP of the pitch term considering the square process.
  • the target deceleration calculating unit 3342c calculates the target deceleration by dividing the pitch moment Mp by the mass m and the height Hcg between the vehicle center of gravity and the road surface.
  • the calculated rate of change of the target deceleration that is, whether the jerk is within a preset range of the deceleration jerk threshold and the extraction jerk threshold, and the target deceleration is the longitudinal acceleration limit value. Judgment is made whether or not it is within the range. If any threshold is exceeded, the target deceleration is corrected to a value within the jerk threshold range, and if the target deceleration exceeds the limit value, the limit is set. Set within the value. Thereby, the deceleration can be generated so as not to give the driver a sense of incongruity.
  • the target pitch moment conversion unit 3343 calculates the target pitch moment by multiplying the target deceleration limited by the jerk threshold limiting unit 3342d by the mass m and the height Hcg, and outputs the target pitch moment to the brake control unit 2a.
  • the sprung speed is estimated based on the detection value of the wheel speed sensor 5 and the skyhook control is performed based on the estimated sprung speed control.
  • a comfortable driving state (a comfortable ride feeling softer than the vehicle body flatness) is guaranteed.
  • vector control where the relationship (phase, etc.) of the sign of stroke speed and sprung speed is important, such as skyhook control, may make it difficult to achieve proper control due to a slight phase shift. Therefore, we decided to introduce frequency-sensitive control, which is sprung mass damping control according to the scalar quantity of vibration characteristics.
  • FIG. 9 is a diagram in which the wheel speed frequency characteristic detected by the wheel speed sensor and the stroke frequency characteristic of a stroke sensor not mounted in the embodiment are simultaneously written.
  • the frequency characteristic is a characteristic in which the vertical axis represents the magnitude of the amplitude with respect to the frequency as a scalar quantity. Comparing the frequency component of the wheel speed sensor 5 with the frequency component of the stroke sensor, it can be understood that substantially the same scalar amount is taken from the sprung resonance frequency component to the unsprung resonance frequency component. Therefore, the damping force is set based on this frequency characteristic among the detection values of the wheel speed sensor 5.
  • the area where the sprung resonance frequency component exists is felt as if the occupant was thrown into the air by swinging the entire body of the occupant, in other words, the feeling that the gravitational acceleration acting on the occupant was reduced.
  • the frequency region that brings about the waving region (0.5 to 3 Hz), and the region between the sprung resonance frequency component and the unsprung resonance frequency component is not a feeling that gravitational acceleration decreases.
  • the feeling that the human body jumps in small increments when performing (trot), in other words, the frequency range that brings up and down movement that the whole body can follow is the leopard region (3 to 6 Hz), and the region where the unsprung resonance frequency component exists Is not a vertical movement until the mass of the human body follows, but a bull region (6 to 6) is used as a frequency region where vibration is transmitted to a part of the body such as the occupant's thigh. 23 Hz).
  • FIG. 10 is a control block diagram illustrating frequency sensitive control in the sprung mass damping control of the first embodiment.
  • the band elimination filter 350 cuts noise other than the vibration component used for the main control from the wheel speed sensor value.
  • the predetermined frequency domain dividing unit 351 divides the frequency band into a wide area, a horizontal area, and a bull area.
  • the Hilbert transform processing unit 352 performs Hilbert transform on each divided frequency band, and converts it into a scalar quantity based on the amplitude of the frequency (specifically, an area calculated from the amplitude and the frequency band).
  • the vehicle vibration system weight setting unit 353 sets weights at which vibrations in the frequency bands of the fur region, the leopard region, and the bull region are actually propagated to the vehicle.
  • the human sense weight setting unit 354 sets weights at which vibrations in the frequency bands of the fur region, the leopard region, and the bull region are propagated to the occupant.
  • FIG. 11 is a correlation diagram showing human sensory characteristics with respect to frequency.
  • the occupant's sensitivity is relatively low with respect to the frequency, and the sensitivity gradually increases as the frequency shifts to the high frequency region.
  • the high frequency region above the bull region becomes difficult to be transmitted to the occupant.
  • the human sense weight Wf of the wafe area is set to 0.17
  • the human sense weight Wh of the leopard area is set to 0.34 which is larger than Wf
  • the human sense weight Wb of the bull area is larger than Wf and Wh. Set to 0.38.
  • the weight determining unit 355 calculates the ratio of the weight of each frequency band to the weight of each frequency band. If the weight of the wing area is a, the weight of the leopard area is b, and the weight of the bull area is c, the weight coefficient of the wing area is (a / (a + b + c)), and the weight coefficient of the leap area is (b / (a + b + c). )), And the weighting factor of the bull area is (c / (a + b + c)).
  • the scalar amount calculation unit 356 multiplies the scalar amount of each frequency band calculated by the Hilbert transform processing unit 352 by the weight calculated by the weight determination unit 355, and outputs a final scalar amount. The processing so far is performed on the wheel speed sensor value of each wheel.
  • the maximum value selection unit 357 selects the maximum value from the final scalar amounts calculated for each of the four wheels. Note that 0.01 in the lower part is set to avoid the denominator becoming 0 because the sum of the maximum values is used as the denominator in the subsequent processing.
  • the ratio calculation unit 358 calculates the ratio using the sum of the scalar value maximum values in each frequency band as the denominator and the scalar value maximum value in the frequency band corresponding to the waving region as the numerator. In other words, the mixing ratio (hereinafter simply referred to as the ratio) of the wafer region included in all vibration components is calculated.
  • the sprung resonance filter 359 performs filter processing of about 1.2 Hz of the sprung resonance frequency with respect to the calculated ratio, and extracts a sprung resonance frequency band component representing a waft region from the calculated ratio. In other words, since the wing area exists at about 1.2 Hz, the ratio of this area is considered to change at about 1.2 Hz. Then, the finally extracted ratio is output to the damping force control unit 35, and a frequency sensitive damping force control amount corresponding to the ratio is output.
  • FIG. 12 is a characteristic diagram showing the relationship between the vibration mixing ratio in the waft region and the damping force by the frequency sensitive control of the first embodiment.
  • the vibration level of sprung resonance is reduced by setting the damping force high when the ratio of the wing area is large.
  • the damping force is set high, the ratio of the leopard area and the bull area is small, so that high frequency vibration or vibration that moves with the leopard is not transmitted to the occupant.
  • the damping force is set low, so that the vibration transmission characteristic more than the sprung resonance is reduced, the high frequency vibration is suppressed, and a smooth riding comfort is obtained.
  • FIG. 13 is a diagram showing the wheel speed frequency characteristics detected by the wheel speed sensor 5 under a certain traveling condition. This is a characteristic that appears particularly when traveling on a road surface in which small unevenness such as a stone pavement continues.
  • the damping force is determined by the value of the amplitude peak in Skyhook control. There is a problem that a very high damping force is set at an incorrect timing and high-frequency vibration is deteriorated.
  • the S / A side driver input control unit 31 calculates a driver input damping force control amount corresponding to a vehicle behavior that the driver wants to achieve based on signals from the steering angle sensor 7 and the vehicle speed sensor 8, and a damping force control unit 35. Output for. For example, when the driver is turning, if the nose side of the vehicle is lifted, the driver's field of view easily deviates from the road surface. In this case, the four-wheel damping force is used as a driver input damping force to prevent the nose from rising. Output as a controlled variable. In addition, a driver input damping force control amount that suppresses a roll generated during turning is output.
  • FIG. 14 is a control block diagram illustrating the configuration of roll rate suppression control according to the first embodiment.
  • the lateral acceleration estimation unit 31b1 the front wheel rudder angle ⁇ f detected by the rudder angle sensor 7 and the rear wheel rudder angle ⁇ r (the actual rear wheel rudder angle if a rear wheel steering device is provided, and 0 in other cases as appropriate)
  • the lateral acceleration Yg is estimated based on the vehicle speed VSP detected by the vehicle speed sensor 8. This lateral acceleration Yg is calculated by the following equation using the yaw rate estimated value ⁇ .
  • Yg VSP ⁇ ⁇
  • the yaw rate estimated value ⁇ is calculated by the following equation.
  • the 90 ° phase advance component creation unit 31b2 differentiates the estimated lateral acceleration Yg and outputs a lateral acceleration differential value dYg.
  • the 90 ° phase delay component creation unit 31b3 outputs a component F (dYg) obtained by delaying the phase of the lateral acceleration differential value dYg by 90 °.
  • the component F (dYg) is obtained by returning the phase of the component from which the low-frequency region has been removed by the 90 ° phase advance component creation unit 31b2 to the phase of the lateral acceleration Yg. It is a transient component of acceleration Yg.
  • the 90 ° phase delay component creation unit 31b4 outputs a component F (Yg) obtained by delaying the phase of the estimated lateral acceleration Yg by 90 °.
  • the gain multiplication unit 31b5 multiplies the lateral acceleration Yg, the lateral acceleration differential value dYg, the lateral acceleration DC cut component F (dYg), and the 90 ° phase delay component F (Yg) by a gain. Each gain is set based on a roll rate transfer function with respect to the steering angle. Each gain may be adjusted according to four control modes described later.
  • the square calculator 31b6 squares and outputs each component multiplied by the gain.
  • the combining unit 31b7 adds the values output from the square calculation unit 31b6.
  • the gain multiplication unit 31b8 multiplies the square value of each added component by the gain and outputs the result.
  • the square root calculation unit 31b9 calculates a driver input attitude control amount for roll rate suppression control by calculating the square root of the value output from the gain multiplication unit 31b7, and outputs the calculated value to the damping force control unit 35.
  • 90 ° phase advance component creation unit 31b2, 90 ° phase lag component creation unit 31b3, 90 ° phase lag component creation unit 31b4, gain multiplication unit 31b5, square operation unit 31b6, synthesis unit 31b7, gain multiplication unit 31b8, square root operation unit 31b9 Corresponds to the Hilbert transform unit 31b10 that generates an envelope waveform using the Hilbert transform.
  • FIG. 15 is a time chart showing an envelope waveform forming process of the roll rate suppressing control according to the first embodiment.
  • the driver starts steering at time t1
  • roll rate begins to gradually occur.
  • the 90 ° phase advance component dYg is added to form an envelope waveform
  • the driver input attitude control amount is calculated based on the scalar amount based on the envelope waveform, thereby suppressing the occurrence of roll rate in the initial stage of steering.
  • Can do Furthermore, by adding the lateral acceleration DC cut component F (dYg) to form an envelope waveform, it effectively suppresses the roll rate that occurs in a transitional state when the driver starts or ends steering. Can do.
  • phase delay component F (Yg) If the phase delay component F (Yg) is not added, the damping force from the time t2 to the time t3 is set to a small value, which may cause the vehicle behavior to become unstable due to the roll rate resonance component. In order to suppress this roll rate resonance component, a 90 ° phase delay component F (Yg) is added.
  • the problems that may occur when setting the damping force according to the unsprung resonance component will be described in more detail.
  • the frequency band in which the vibration transmissibility rises includes the resonance frequency region of the human body of 3 to 6 Hz (leak region).
  • the increase in the vibration transmissibility in this frequency region causes the occupant's body to jump in small increments.
  • Such a vibration that is, vibration in the horizontal region
  • the unsprung vibration suppression control amount is set to It was decided to perform control to suppress.
  • FIG. 16 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the first embodiment.
  • the unsprung resonance component extraction unit 341 extracts a unsprung resonance component by applying a band-pass filter to the wheel speed fluctuation output from the deviation calculation unit 321b in the traveling state estimation unit 32.
  • the unsprung resonance component is extracted from the region of approximately 10 to 20 Hz of the wheel speed frequency component.
  • the envelope waveform shaping unit 342 the extracted unsprung resonance component is scalarized, and the envelope waveform is shaped using the EnvelopeFilter.
  • the gain multiplication unit 343 multiplies the scalarized unsprung resonance component by a gain, calculates an unsprung vibration damping damping force control amount, and outputs it to the control amount setting unit 346.
  • the occupant vibration estimation unit 344 reads the sprung acceleration at the suspension attachment point from the state amount estimated by the third traveling state estimation unit 32, and represents a scalar amount Wb * g representing the vibration state felt by the occupant based on human sensory characteristics. And Wh * g. Specifically, the first envelope filter 344a performs 3-6 Hz bandpass filter processing of the sprung vertical acceleration to form an envelope waveform, and the weight multiplication unit 344a1 multiplies the human sense weight Wh of the leopard region described in FIG. The scalar amount Wh * g is output to the correction gain calculation unit 345.
  • a 6-23 Hz bandpass filter process of sprung vertical acceleration is performed by the second envelope filter 344b to form an envelope waveform, and the weight multiplication unit 344b1 multiplies the human sense weight Wb of the bull region described in FIG.
  • the amount Wb * g is output to the correction gain calculator 345.
  • the correction gain calculation unit 345 selects 1 or 0 as the correction gain based on the magnitude relationship between Wb * g and Wh * g, and sets the correction gain K * g that has undergone smoothing processing at the time of change as a control amount setting unit. Output to 346. Specifically, when Wb * g> Wh * g in the gain selection unit 345a, since the vibration in the horizontal region is not particularly deteriorated, 1 is selected and output to the smoothing unit 345b. In this case, it is determined that the vibration component in the leopard region deteriorates and the vibration component increases at 3 to 6 Hz where the human body vertical resonance exists, and 0 is selected and output to the smoothing unit 345b.
  • the smoothing unit 345b smoothes the sudden change of the values of 1 and 0 selected by the gain selection unit 345a and outputs the correction gain K * g to the control amount setting unit 346.
  • the control amount setting unit 346 outputs a value obtained by multiplying the unsprung vibration damping damping force control amount by the correction gain K * g to the damping force control unit 35.
  • the unsprung resonance component is extracted by applying a bandpass filter to the wheel speed fluctuation output from the deviation calculating section 321b in the running state estimating section 32.
  • the unsprung resonance component may be extracted by applying a bandpass filter to the driving force, or the unsprung resonance component may be extracted by the running state estimation unit 32 by estimating and calculating the unsprung speed along with the sprung speed. Good.
  • FIG. 17 is a control block diagram illustrating a control configuration of the damping force control unit according to the first embodiment.
  • the driver input damping force control amount output from the driver input control unit 31 the S / A attitude control amount output from the skyhook control unit 33a, and the frequency sensitive control unit 33b output
  • the frequency sensitive damping force control amount, the unsprung damping damping force control amount output from the unsprung damping control unit 34, and the stroke speed calculated by the running state estimation unit 32 are input, and these values are equivalent. Convert to viscous damping coefficient.
  • each damping coefficient is referred to as driver input damping coefficient k1, S / A attitude damping coefficient k2, frequency sensitive damping coefficient k3, unsprung). (Which is described as damping damping coefficient k4)), which arbitration is performed based on which damping coefficient is controlled, and a final damping coefficient is output.
  • the control signal converter 35c converts the control signal (command current value) for S / A3 based on the attenuation coefficient and stroke speed adjusted by the attenuation coefficient adjuster 35b, and outputs the control signal to S / A3.
  • the vehicle control apparatus has four control modes. First, the standard mode assuming a state where an appropriate turning state can be obtained while driving in a general urban area, and second, a state where a stable turning state can be obtained while actively driving a winding road etc. In sport mode, thirdly, comfort mode that assumes a state of driving with priority on ride comfort, such as when starting at a low vehicle speed, and fourthly, highway mode that assumes a state of traveling at high vehicle speed on highways with many straight lines is there.
  • sport mode thirdly, comfort mode that assumes a state of driving with priority on ride comfort, such as when starting at a low vehicle speed
  • highway mode that assumes a state of traveling at high vehicle speed on highways with many straight lines is there.
  • priority is given to unsprung vibration suppression control by the unsprung vibration suppression control unit 34 while performing skyhook control by the skyhook control unit 33a.
  • priority is given to driver input control by the driver input control unit 31, and skyhook control by the skyhook control unit 33a and unsprung vibration suppression control by the unsprung vibration suppression control unit 34 are performed.
  • comfort mode the control for giving priority to the unsprung vibration damping control by the unsprung vibration damping control unit 34 is performed while performing the frequency sensitive control by the frequency sensitive control unit 33b.
  • priority is given to driver input control by the driver input control unit 31, and control for adding the amount of unsprung vibration suppression control by the unsprung vibration control unit 34 to skyhook control by the skyhook control unit 33a is performed. To do.
  • the adjustment of the attenuation coefficient in each mode will be described.
  • FIG. 18 is a flowchart illustrating the attenuation coefficient arbitration process in the standard mode according to the first embodiment.
  • step S1 it is determined whether or not the S / A attitude damping coefficient k2 is larger than the unsprung damping damping coefficient k4. If larger, the process proceeds to step S4 and k2 is set as the damping coefficient.
  • step S2 a scalar amount ratio of the bull region is calculated based on the scalar amounts of the fur region, the leopard region, and the bull region described in the frequency response control unit 33b.
  • step S3 it is determined whether or not the ratio of the bull area is equal to or greater than a predetermined value.
  • the routine proceeds to step S5 and k4 is set.
  • FIG. 19 is a flowchart showing attenuation coefficient arbitration processing in the sport mode of the first embodiment.
  • step S11 the four-wheel damping force distribution ratio is calculated based on the four-wheel driver input damping coefficient k1 set by the driver input control.
  • the right front wheel driver input damping coefficient is k1fr
  • the left front wheel driver input damping coefficient is k1fl
  • the right rear wheel driver input damping coefficient is k1rr
  • the left rear wheel driver input damping coefficient is k1rl
  • xfl k1fl / (k1fr + k1fl + k1rr + k1rl)
  • xrr k1rr / (k1fr + k1fl + k1rr + k1rl)
  • xrl k1rl / (k1fr + k1fl + k1rr + k1rl)
  • xrl k
  • step S12 it is determined whether or not the damping force distribution ratio x is within a predetermined range (greater than ⁇ and smaller than ⁇ ). If it is within the predetermined range, it is determined that the distribution to each wheel is substantially equal, and the process proceeds to step S13. If any one is out of the predetermined range, the process proceeds to step S16. In step S13, it is determined whether or not the unsprung damping damping coefficient k4 is larger than the driver input damping coefficient k1. If it is determined that the unsprung damping damping coefficient k4 is larger, the process proceeds to step S15 and k4 is set as the first damping coefficient k. On the other hand, if it is determined that the unsprung damping damping coefficient k4 is equal to or less than the driver input damping coefficient k1, the process proceeds to step S14, and k1 is set as the first damping coefficient k.
  • step S16 it is determined whether or not the unsprung damping damping coefficient k4 is the maximum value max that S / A3 can be set. If it is determined that the maximum value is max, the process proceeds to step S17, and otherwise, the process proceeds to step S18. move on.
  • step S17 the maximum value of the four-wheel driver input damping coefficient k1 is the unsprung damping damping coefficient k4, and the damping coefficient that satisfies the damping force distribution ratio is calculated as the first damping coefficient k. In other words, a value that maximizes the damping coefficient while satisfying the damping force distribution rate is calculated.
  • step S18 a damping coefficient that satisfies the damping force distribution ratio in a range where all the four-wheel driver input damping coefficients k1 are equal to or greater than k4 is calculated as the first damping coefficient k.
  • a value that satisfies the damping force distribution ratio set by the driver input control and also satisfies the requirements of the unsprung vibration suppression control side is calculated.
  • step S19 it is determined whether or not the first attenuation coefficient k set in each of the above steps is smaller than the S / A attitude attenuation coefficient k2 set by skyhook control. Since the damping coefficient requested on the side is larger, the process proceeds to step S20 and k2 is set. On the other hand, if it is determined that k is equal to or greater than k2, the process proceeds to step S21 and k is set.
  • the damping force distribution rate required from the driver input control side is closely related to the vehicle body posture, and particularly because it is closely related to the driver's line-of-sight change due to the roll mode.
  • the highest priority is to secure the damping force distribution ratio.
  • the sky vehicle body posture can be maintained by selecting Skyhook control with select high.
  • FIG. 20 is a flowchart illustrating the attenuation coefficient arbitration process in the comfort mode according to the first embodiment.
  • step S30 it is determined whether or not the frequency sensitive damping coefficient k3 is larger than the unsprung damping damping coefficient k4. If it is determined that the frequency sensitive damping coefficient k3 is larger, the process proceeds to step S32 and the frequency sensitive damping coefficient k3 is set. On the other hand, if it is determined that the frequency sensitive damping coefficient k3 is equal to or less than the unsprung damping damping coefficient k4, the process proceeds to step S32 to set the unsprung damping damping coefficient k4.
  • the comfort mode priority is given to unsprung resonance control that basically suppresses unsprung resonance.
  • frequency sensitive control was performed as sprung mass damping control, and the optimum damping coefficient was set according to the road surface condition, so it was possible to achieve control that ensured riding comfort and lack of grounding feeling due to fluttering under the spring. Can be avoided by unsprung vibration suppression control.
  • the attenuation coefficient may be switched according to the bull ratio of the frequency scalar quantity. As a result, the ride comfort can be further ensured in the super comfort mode.
  • FIG. 21 is a flowchart illustrating the attenuation coefficient arbitration process in the highway mode according to the first embodiment. Since steps S11 to S18 are the same as the arbitration process in the sport mode, the description thereof is omitted.
  • step S40 the S / A attitude attenuation coefficient k2 by the skyhook control is added to the first attenuation coefficient k that has been adjusted up to step S18, and is output.
  • FIG. 22 is a time chart showing a change in attenuation coefficient when traveling on a wavy road surface and an uneven road surface.
  • the first damping coefficient k is always set as in the highway mode, a certain amount of damping force is always secured, and the vehicle body fluctuates even when the damping coefficient by the skyhook control is small. Such movement can be suppressed. Further, since it is not necessary to increase the skyhook control gain, it is possible to appropriately deal with road surface irregularities by using a normal control gain. In addition, since the skyhook control is performed with the first damping coefficient k set, unlike the damping coefficient limit, the damping coefficient decreasing process can be performed in the semi-active control region, and at the time of high-speed traveling It is possible to ensure a stable vehicle posture.
  • FIG. 23 is a flowchart illustrating a mode selection process based on the running state in the attenuation coefficient arbitration unit of the first embodiment.
  • step S50 it is determined whether or not the vehicle is in the straight traveling state based on the value of the steering angle sensor 7. If it is determined that the vehicle is traveling straight, the process proceeds to step S51. If it is determined that the vehicle is turning, the process proceeds to step S54. move on.
  • step S51 it is determined based on the value of the vehicle speed sensor 8 whether or not the vehicle speed is equal to or higher than a predetermined vehicle speed VSP1 representing a high vehicle speed state.
  • step S52 If it is determined that the vehicle speed is VSP1 or higher, the process proceeds to step S52 and the standard mode is selected. On the other hand, if it is determined that it is less than VSP1, the process proceeds to step S53 and the comfort mode is selected. In step S54, based on the value of the vehicle speed sensor 8, it is determined whether or not the vehicle speed is equal to or higher than a predetermined vehicle speed VSP1 representing a high vehicle speed state. If it is determined that the vehicle speed is VSP1 or higher, the process proceeds to step S55 and the highway mode is selected. On the other hand, if it is determined that it is less than VSP1, the process proceeds to step S56 to select the sport mode.
  • the standard mode when driving at a high vehicle speed in a straight running state, the standard mode is selected to stabilize the vehicle body posture by skyhook control and to suppress the high frequency vibration such as leopard and bull. In addition, unsprung resonance can be suppressed. Further, when traveling at a low vehicle speed, by selecting the comfort mode, it is possible to suppress unsprung resonance while suppressing the input of vibrations such as leopard and bull to the occupant as much as possible.
  • the highway mode is selected, so that it is controlled by the value obtained by adding the damping coefficient, so that basically a high damping force can be obtained.
  • the sport mode is selected, so that the vehicle posture during turning is positively secured by driver input control, and unsprung resonance is suppressed while skyhook control is performed as appropriate. Can travel in a stable vehicle posture.
  • the control example in which the driving state is detected and automatically switched is shown in the first embodiment.
  • a changeover switch that can be operated by the driver is provided to select the driving mode. You may control to. As a result, ride comfort and turning performance according to the driving intention of the driver can be obtained.
  • Example 1 has the following effects.
  • S / A3 (damping force variable shock absorber) capable of changing damping force
  • third traveling state estimating unit 32 unsprung vertical speed detecting means for detecting unsprung vertical speed, detecting sprung vertical acceleration The unsprung acceleration detection means) and the unsprung damping damping force control amount (damping force control amount) of S / A3 according to the unsprung vertical speed estimated by the third running state estimation unit 32.
  • the vibration control unit 34 (damping force control amount calculating means) and the magnitude of the amplitude in any frequency band of the sprung vertical acceleration are defined as the frequency scalar quantity, it corresponds to the amplitude in the higher frequency band than the human body resonance frequency band.
  • correction gain K * g is switched from 1 to 0.
  • Correction gain calculator 345 (damping force control The amount and the reduced correction for correcting section), with a. Therefore, even when the unsprung vibration suppression control amount, which is a damping force set according to the unsprung vertical speed, is set to be large, the component of the human body resonance frequency band is large among the sprung vertical acceleration components. In some cases, by reducing and correcting the damping force, the grounding property of the wheel can be ensured while ensuring the riding comfort.
  • the third traveling state estimation unit 32 estimates the sprung vertical speed and the sprung vertical acceleration based on the detected wheel speed sensor value. Therefore, it is not necessary to provide an expensive sensor such as a sprung vertical acceleration sensor or a stroke sensor, and the number of parts can be reduced by estimating the sprung state from the wheel speed sensor 5 that is generally mounted on any vehicle. In addition, the cost can be reduced and the vehicle mountability can be improved.
  • the correction gain calculation unit 345 calculates Wh * g by multiplying the scalar quantity corresponding to the amplitude of the human body resonance frequency band by Wh (first human sense weight coefficient), and is higher than the human body resonance frequency band.
  • Wb * g is calculated by multiplying the scalar quantity corresponding to the amplitude of the frequency band by Wb (second human sensory weight coefficient) larger than Wh. Therefore, it is possible to reflect the sense of human feeling in actual control and to secure a ride comfort.
  • FIG. 24 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the second embodiment.
  • the sprung vertical acceleration at the suspension attachment point was input to the first envelope filter 344a and the second envelope filter 344b.
  • the second embodiment instead of the suspension attachment point, it is converted into the sprung vertical acceleration at the seat position, and further converted into the sprung vertical acceleration considering the human body response, and this value Ghi is converted into the first and second envelope filter 344a. , 344b is different.
  • the sprung vertical acceleration detecting means is a means for detecting the vertical acceleration at the attachment point of the damping force variable shock absorber.
  • the boarding position vertical acceleration at the boarding position of the occupant is estimated, and the damping force control amount is reduced and corrected based on the boarding position vertical acceleration. Therefore, the damping force can be set according to the vertical acceleration component felt by the occupant, and an appropriate damping force with higher accuracy can be set.
  • FIG. 25 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the third embodiment.
  • Example 2 the sprung vertical acceleration considering the seat position and the human body response was used.
  • the third embodiment is different in that the seat transfer function of the seat is further considered.
  • the correction gain calculation unit 345 estimates the boarding position vertical acceleration in consideration of the elastic vibration of the seat at the occupant's boarding position, and reduces and corrects the damping force control amount based on the boarding position vertical acceleration. Therefore, the vertical acceleration component felt by the occupant can be estimated with higher accuracy, and an appropriate damping force with higher accuracy can be set.
  • FIG. 26 is a characteristic diagram illustrating the value of the human sense weight Wb in the bull area according to the fourth embodiment.
  • the human sense weight Wb of the bull area is set as a constant value.
  • the fourth embodiment is different in that the weight increases in accordance with the lateral acceleration. That is, when Wb is large, a higher value is set as the damping force, and when Wb is small, a smaller value is set as the damping force.
  • comfort is the most important during straight traveling, a relatively small Wb is set and the damping force is set small in a region where the lateral acceleration is small.
  • FIG. 27 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the fifth embodiment.
  • the correction gain is set in accordance with the magnitude relationship between the horizontal and vertical components of the sprung vertical acceleration.
  • the bull region frequency band accounts for the ratio of the frequency band of each frequency band calculated by the weight determination unit 355 in the frequency sensitive control unit 33b. The difference is that the correction gain is set according to the weighting factor (c / (a + b + c)) of the region.
  • the damping force set in the unsprung vibration suppression control is small. That is, when there are many components in the wing region and the leopard region, it is preferable to give priority to the damping force characteristic set in the skyhook control unit 33a of the sprung mass damping control unit 33. In this case, if the damping force by the unsprung vibration suppression control becomes a large value, the requested damping force on the unsprung vibration control side becomes larger than the requested damping force on the skyhook control side, and than the damping force by the skyhook control side. The damping force for the unsprung vibration suppression control is selected. Then, the ability to suppress the fluffy movement on the upper side of the spring is not sufficiently exhibited. Therefore, when the ratio occupied by the frequency band of the bull region is small, the correction gain is reduced from the viewpoint of securing the performance by the skyhook control.

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  • Mechanical Engineering (AREA)
  • Vehicle Body Suspensions (AREA)

Abstract

L'invention concerne un appareil de commande de véhicule, caractérisé en ce qu'une valeur de commande de force d'amortissement pour un amortisseur à force d'amortissement variable est calculée conformément à une vitesse verticale non suspendue. Lorsque l'importance de l'amplitude de l'accélération verticale suspendue dans une bande de fréquence donnée est exprimée en tant que quantité scalaire de fréquence, si la quantité scalaire de fréquence d'une bande à haute fréquence est supérieure à la quantité scalaire de fréquence d'une bande à basse fréquence, la valeur de commande de force d'amortissement est diminuée en vue d'une correction.
PCT/JP2013/054681 2012-03-09 2013-02-25 Appareil de commande de véhicule, et procédé de commande de véhicule WO2013133059A1 (fr)

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Cited By (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2017149297A (ja) * 2016-02-25 2017-08-31 日立オートモティブシステムズ株式会社 車両制御装置
JP2017206161A (ja) * 2016-05-19 2017-11-24 トヨタ自動車株式会社 車両用減衰力制御装置
WO2019027010A1 (fr) * 2017-08-04 2019-02-07 クラリオン株式会社 Dispositif d'alarme pour véhicule et procédé d'alarme pour véhicule
KR102031921B1 (ko) * 2018-08-23 2019-10-14 경북대학교 산학협력단 파라미터의 불확실성을 이용한 차량-인체 모델의 승차감 최적화 장치 및 방법, 상기 방법을 수행하기 위한 기록 매체
CN112659995A (zh) * 2019-10-15 2021-04-16 丰田自动车株式会社 车辆的减振控制装置
JP2022088717A (ja) * 2020-12-03 2022-06-15 本田技研工業株式会社 電動サスペンション装置
WO2022168683A1 (fr) * 2021-02-03 2022-08-11 株式会社アイシン Dispositif d'estimation de quantité d'état de véhicule
CN117124789A (zh) * 2023-10-26 2023-11-28 成都创一博通科技有限公司 基于人工智能和大数据的车辆悬架控制方法和控制系统

Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH06247121A (ja) * 1993-02-24 1994-09-06 Unisia Jecs Corp 車両懸架装置
JPH08216641A (ja) * 1995-02-09 1996-08-27 Nippondenso Co Ltd サスペンションの減衰力制御装置
JPH08238915A (ja) * 1995-03-06 1996-09-17 Nippondenso Co Ltd 車体挙動制御装置
JPH1120443A (ja) * 1997-06-27 1999-01-26 Mitsubishi Motors Corp 車両用サスペンション制御装置
JP2001047835A (ja) * 1999-08-11 2001-02-20 Nissan Motor Co Ltd 能動型サスペンション
JP2008183971A (ja) * 2007-01-29 2008-08-14 Toyota Motor Corp 車両状態推定装置及びショックアブソーバ制御装置
JP2008275336A (ja) * 2007-04-25 2008-11-13 Toyota Central R&D Labs Inc 評価装置及びプログラム
JP2011225040A (ja) * 2010-04-16 2011-11-10 Nissan Motor Co Ltd サスペンション制御装置
JP2012001076A (ja) * 2010-06-16 2012-01-05 Nissan Motor Co Ltd 車両の減衰力制御装置及び減衰力制御方法

Family Cites Families (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH05286326A (ja) * 1992-04-07 1993-11-02 Mitsubishi Motors Corp 減衰力制御装置
JP3430321B2 (ja) * 1994-03-18 2003-07-28 トキコ株式会社 減衰力調整式油圧緩衝器
JPH0858338A (ja) * 1994-08-25 1996-03-05 Nippondenso Co Ltd 車両減衰力制御装置
JPH08230432A (ja) * 1995-02-23 1996-09-10 Toyota Motor Corp 車両の減衰力制御装置
JP4356544B2 (ja) * 2004-07-13 2009-11-04 日産自動車株式会社 ショックアブソーバ

Patent Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPH06247121A (ja) * 1993-02-24 1994-09-06 Unisia Jecs Corp 車両懸架装置
JPH08216641A (ja) * 1995-02-09 1996-08-27 Nippondenso Co Ltd サスペンションの減衰力制御装置
JPH08238915A (ja) * 1995-03-06 1996-09-17 Nippondenso Co Ltd 車体挙動制御装置
JPH1120443A (ja) * 1997-06-27 1999-01-26 Mitsubishi Motors Corp 車両用サスペンション制御装置
JP2001047835A (ja) * 1999-08-11 2001-02-20 Nissan Motor Co Ltd 能動型サスペンション
JP2008183971A (ja) * 2007-01-29 2008-08-14 Toyota Motor Corp 車両状態推定装置及びショックアブソーバ制御装置
JP2008275336A (ja) * 2007-04-25 2008-11-13 Toyota Central R&D Labs Inc 評価装置及びプログラム
JP2011225040A (ja) * 2010-04-16 2011-11-10 Nissan Motor Co Ltd サスペンション制御装置
JP2012001076A (ja) * 2010-06-16 2012-01-05 Nissan Motor Co Ltd 車両の減衰力制御装置及び減衰力制御方法

Cited By (10)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2017149297A (ja) * 2016-02-25 2017-08-31 日立オートモティブシステムズ株式会社 車両制御装置
JP2017206161A (ja) * 2016-05-19 2017-11-24 トヨタ自動車株式会社 車両用減衰力制御装置
WO2019027010A1 (fr) * 2017-08-04 2019-02-07 クラリオン株式会社 Dispositif d'alarme pour véhicule et procédé d'alarme pour véhicule
EP3647149A4 (fr) * 2017-08-04 2020-08-12 Clarion Co., Ltd. Dispositif d'alarme pour véhicule et procédé d'alarme pour véhicule
KR102031921B1 (ko) * 2018-08-23 2019-10-14 경북대학교 산학협력단 파라미터의 불확실성을 이용한 차량-인체 모델의 승차감 최적화 장치 및 방법, 상기 방법을 수행하기 위한 기록 매체
CN112659995A (zh) * 2019-10-15 2021-04-16 丰田自动车株式会社 车辆的减振控制装置
JP2022088717A (ja) * 2020-12-03 2022-06-15 本田技研工業株式会社 電動サスペンション装置
WO2022168683A1 (fr) * 2021-02-03 2022-08-11 株式会社アイシン Dispositif d'estimation de quantité d'état de véhicule
CN117124789A (zh) * 2023-10-26 2023-11-28 成都创一博通科技有限公司 基于人工智能和大数据的车辆悬架控制方法和控制系统
CN117124789B (zh) * 2023-10-26 2023-12-22 成都创一博通科技有限公司 基于人工智能和大数据的车辆悬架控制方法和控制系统

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