JP2013193716A - Vehicle controlling apparatus and method - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a vehicle controlling apparatus, capable of controlling a vehicle body attitude without making an occupant have the feeling of incompatibility.SOLUTION: A vehicle controlling apparatus includes: a plurality of actuators for performing sprung mass vibration suppression control; a stroke sensor for detecting the stroke speed of a shock absorber; and a plurality of actuator attitude control means for controlling the respective actuators so that the stroke speed becomes the stroke speed corresponding to a target sprung mass state.

Description

  The present invention relates to a control device that controls the state of a vehicle.

  As a technique related to a vehicle control device, a technique described in Patent Document 1 is disclosed. This publication discloses a technique for controlling a vibration state of a vehicle body using a driving force.

Japanese Patent Application Laid-Open No. 2011-223691

  However, it has been difficult to sufficiently suppress vibration only by vibration suppression control using a driving 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 suppressing vehicle vibration.

  In order to achieve the above object, in the vehicle control apparatus of the present invention, each actuator is controlled by a plurality of actuators that perform sprung mass damping control so that the stroke speed becomes a stroke speed corresponding to the target sprung state. It was.

  Therefore, by performing the sprung mass damping control with a plurality of actuators, the sprung behavior can be more effectively stabilized. In addition, since a feedback control system is configured in each actuator using a common value of stroke speed, even if individual control is performed without controlling mutual monitoring, the result of mutual monitoring is the result. Control (hereinafter, this control is described as emphasis control) is performed, and the vehicle posture can be converged in the stabilization direction.

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. FIG. 3 is a conceptual diagram illustrating a configuration of a stroke speed feedback control system according to the 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 the schematic showing a vehicle body vibration model. It is a control block diagram showing brake pitch control of Example 1. It is a figure showing the frequency characteristic of a stroke sensor. It is a control block diagram showing the frequency sensitive control in the sprung mass damping control of the first embodiment. It is a correlation diagram showing the human sensory characteristic in each frequency domain. It is a characteristic view showing the relationship between the vibration mixing ratio of the wing area | region by the frequency sensitive control of Example 1, and damping force. It is a figure showing the frequency characteristic detected by the stroke sensor in a certain running condition. 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. It is a time chart showing the attenuation coefficient change at the time of drive | working a wavy road surface and an uneven | corrugated road surface. 6 is a flowchart illustrating a mode selection process based on a running state in an attenuation coefficient arbitration unit according to the first embodiment.

[Example 1]
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). Right rear wheel brake: 20RR, left rear wheel brake: 20RL), and shock absorber 3 (hereinafter referred to as S / A) provided between each wheel and the vehicle body and capable of variably controlling the damping force. When displaying S / A corresponding to an individual wheel, right front wheel S / A: 3FR, left front wheel S / A: 3FL, right rear wheel S / A: 3RR, left rear wheel S / A: 3RL And).

  The engine 1 includes an engine controller (hereinafter also referred to as an engine control unit, which corresponds to a power source control means) 1a that controls torque output from the engine 1, and the engine controller 1a includes a throttle valve opening degree of the engine 1, 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.

  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.

  In addition to the selection of the orifice diameter, for example, 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. There is no particular limitation. 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.

  Further, a wheel speed sensor 5 for detecting the wheel speed of each wheel (hereinafter, when displaying the wheel speed corresponding to each wheel, 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, and master pressure sensor 11 for detecting master cylinder pressure. A brake switch 12 that outputs an on-state signal when the brake pedal is operated, an accelerator opening sensor 13 that detects the accelerator pedal opening, and the S / 3, a stroke sensor 14 for detecting the S / A3 stroke speed (hereinafter, when displaying a stroke speed corresponding to an individual wheel, right front wheel stroke speed: 14FR, left front wheel stroke speed: 14FL, Right rear wheel stroke speed: 14RR, left rear wheel stroke speed: 14RL). 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 the position of the center of gravity of the vehicle, and may be a configuration that can estimate various values at the position of the center of gravity even at other locations, and is not particularly limited. 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.

(Overall configuration of vehicle control device)
In the vehicle control apparatus of the first embodiment, three actuators are used to control the vibration state generated on the spring. At this time, since each control controls the sprung state, mutual interference becomes a problem. In addition, the elements that can be controlled by the engine 1, the elements that can be controlled by the brake 20, and the elements that can be controlled by the S / A 3 are different from each other, and how to control them in combination is a problem.
For example, the brake 20 can control a bounce motion and a pitch motion, but if both are performed, a feeling of deceleration is strong and a driver is likely to feel uncomfortable. In addition, S / A3 can control all of roll motion, bounce motion and pitch motion. However, when all control is performed by S / A3 in a wide range, the production cost of S / A3 is increased. Since the damping force tends to increase, high-frequency vibrations from the road surface side are easily input, and the driver is likely to feel uncomfortable. In other words, there is a trade-off that the control by the brake 20 does not cause deterioration of the high frequency vibration but increases the feeling of deceleration, and the control by the S / A3 does not cause the feeling of deceleration but invites the input of high frequency vibration. To do.

Therefore, in the vehicle control apparatus of the first embodiment, it is possible to comprehensively judge these problems and realize a control configuration that complements each other's weak points while taking advantage of the advantages as the respective control characteristics. Therefore, in order to realize a vehicle control apparatus that is inexpensive but has excellent vibration control capability, an overall control system was constructed mainly considering the points listed below.
(1) The control amount by S / A3 is suppressed by performing control by the engine 1 and the brake 20 in parallel.
(2) By limiting the control target motion of the brake 20 to the pitch motion, the feeling of deceleration in the control by the brake 20 is eliminated.
(3) 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.
(4) Skyhook control is performed by all actuators. At this time, when configuring the feedback control system in all the actuators, the skyhook control is realized by the configuration using the stroke sensor 14.
(5) When performing sprung control by S / A3, scalar control (frequency sensitive control) is newly introduced for the input of high frequency vibration that is difficult to cope with by vector control such as skyhook control.
(6) By appropriately selecting the control state realized by the S / A 3 according to the traveling state, an appropriate control state according to the traveling state is provided.
The above is the outline of the entire control system configured in the embodiment. Hereinafter, individual contents for realizing these will be sequentially described.

FIG. 2 is a control block diagram illustrating a control configuration of the vehicle control apparatus according to the first embodiment. In the first embodiment, the controller is composed of an engine controller 1a, a brake controller 2a, and an S / A controller 3a, and each controller constitutes a feedback control system using the stroke sensor 14.
In the first embodiment, 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.

(Engine controller configuration)
The engine controller 1 a mainly performs feedback control based on the stroke speed detected by the stroke sensor 14. The engine controller 1a includes a first traveling state estimation unit 100 that estimates stroke speed, bounce rate, roll rate, and pitch rate of each wheel used for skyhook control of a sprung mass damping control unit 101a, which will be described later, and an engine torque command. An engine attitude control unit 101 that calculates the engine attitude control amount, and an engine control unit 102 that controls the operating state of the engine 1 based on the calculated engine attitude control amount. 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, and 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. In this manner, by suppressing the bounce motion and the pitch motion by the engine 1, 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.

(Brake controller configuration)
The brake controller 2a is configured to estimate the stroke speed and pitch rate of each wheel based on the stroke speed detected by the stroke sensor 14, and based on the estimated stroke speed and pitch rate. A skyhook control unit 201 (details will be described later) that calculates a brake attitude control amount based on the skyhook control, and a brake control unit 202 that controls the braking torque of the brake 20 based on the calculated brake attitude control amount; Have In the first embodiment, 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 stroke speed. It may be used. In this way, by suppressing the pitch motion by the brake 20, in S / A3, the damping force control amount can be reduced, so that deterioration of high-frequency vibration can be avoided. Moreover, since S / A3 can concentrate on suppression of roll motion, it can suppress roll motion effectively.

(Configuration of S / A controller)
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 a traveling state based on (mainly a stroke speed detected by the stroke sensor 14), and a sprung damping that controls a 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 Based on the unsprung vibration suppression control amount output from the control unit 33 and the unsprung vibration suppression control amount output from the unsprung vibration suppression control unit 34, the S / A3 is set. It can determine the damping force, and a damping force control unit 35 for performing the damping force control of the S / A. In the first embodiment, the same estimation process is employed as the estimation process in the first travel state estimation unit 100, the second travel state estimation unit 200, and the third travel state estimation unit 32, but the process is estimated from the stroke speed. If so, other estimation processes may be used and there is no particular limitation.

  Here, in the first embodiment, the feedback control system using the stroke sensor 14 is configured in all the actuators. FIG. 3 is a conceptual diagram showing the configuration of the 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. At this time, when each actuator individually operates without mutually monitoring the operation state, control interference becomes a problem. However, since the influence of the control of each actuator appears as the stroke speed, by configuring a feedback control system based on the stroke speed, the influence of each actuator is monitored as a result, and control interference is avoided. Is. For example, when a certain sprung vibration is suppressed by the engine 1, even if the other actuators do not sense the control content performed in the engine 1, the brake 20 or the brake 20 or the like based on the stroke speed in which the influence is reflected. The S / A 3 performs control. In other words, since the feedback control system is configured using a common value of stroke speed, even if individual control is performed without controlling 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. Hereinafter, each feedback control system will be described sequentially.

(About the running state estimation unit)
First, the 1st, 2nd, 3rd driving state estimation part which is a common structure provided in each feedback control system is demonstrated. In the first embodiment, 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 vertical acceleration on the spring, and is not particularly limited.

  FIG. 4 is a control block diagram illustrating the configuration of the third traveling state estimation unit according to the first embodiment. In the third traveling state estimation unit 32 of the first embodiment, bounce rate 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 stroke speed detected by the stroke sensor 14. , Roll rate and pitch rate are calculated. First, the value of the stroke sensor 14 of each wheel is input to the stroke speed calculation unit 321, and the sprung speed is calculated from the stroke speed Vz_s of each wheel calculated by the stroke speed calculation unit 321.

  When the stroke speed calculation unit 321 calculates the stroke speeds Vz_sFL, Vz_sFR, Vz_sRL, and Vz_sRR for each wheel, the sprung speed calculation unit 322 calculates bounce rate, roll rate, and pitch rate for skyhook control.

(About the estimation model)
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. Here, in order to achieve the posture control on the spring by the skyhook control, it is necessary to feed back the sprung speed. The value that can be detected from the stroke sensor 14 is the stroke speed. 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. Hereinafter, the problem of the estimation model and the model configuration to be adopted will be described.

  FIG. 5 is a schematic diagram showing a vehicle body vibration model. FIG. 5A is a model of a vehicle (hereinafter referred to as a “convex vehicle”) having an S / A having a constant damping force, and FIG. 5B has an S / A having a variable damping force. This is a model for performing skyhook control. In FIG. 5, Ms represents the mass on the spring, Mu represents the mass under the spring, Ks represents the elastic coefficient of the coil spring, Cs represents the damping coefficient of S / A, and Ku represents the unsprung (tire). , Cu represents an unsprung (tire) damping coefficient, and Cv represents a variable damping coefficient. Z2 represents a position on the spring, z1 represents a position under the spring, and z0 represents a road surface position.

When the conveyor vehicle model shown in FIG. 5A is used, the equation of motion for the sprung is expressed as follows. Note that the first derivative (ie, speed) of z1 is represented by dz1, and the second derivative (ie, acceleration) is represented by ddz1.
(Estimation formula 1)
Ms · ddz2 = −Ks (z2−z1) −Cs (dz2−dz1)
When this relational expression is rearranged by Laplace transform, it is expressed as follows.
(Estimation formula 2)
dz2 = − (1 / Ms) · (1 / s ² ) · (Cs · s + Ks) (dz2−dz1)
Here, since dz2-dz1 is a stroke speed (Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR), the sprung speed can be calculated from the stroke speed. However, when the damping force is changed by the skyhook control, the estimation accuracy is remarkably lowered, and therefore, there is a problem that a large attitude control force (attenuating force change) cannot be given in the convex vehicle model.

Therefore, it is conceivable to use a vehicle model based on skyhook control as shown in FIG. 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.
(Estimation formula 3)
dz2 = − (1 / Ms) · (1 / s ² ) · {(Cs + Cv) · s + Ks} (dz2−dz1)
However,
When dz2 · (dz2−dz1) ≧ 0 Cv = Csky · {dz2 / (dz2−dz1)}
When dz2 · (dz2−dz1) <0, Cv = 0
That is, Cv has a discontinuous value.

  Now, if you want to estimate the sprung speed using a simple filter, in the semi-active skyhook model, when this model is viewed as a filter, each variable corresponds to a filter coefficient, and the pseudo-differential term { Since (Cs + Cv) · s + Ks} includes a discontinuous variable attenuation coefficient Cv, the filter response becomes unstable and appropriate estimation accuracy cannot be obtained. In particular, when the filter response becomes unstable, the phase shifts. If the correspondence between the phase of the sprung speed and the sign is broken, the skyhook control cannot be achieved. Therefore, even when a semi-active S / A3 is used, it is not dependent on the sign relationship between the sprung speed and the stroke speed, and the sprung is performed using an active skyhook model that can directly use stable Csky. The speed was estimated. When the active sky hook model is adopted and the sprung speed is obtained, it is expressed as follows.

(Estimation formula 4)
dz2 =-(1 / s). {1 / (s + Csky / Ms)}. {(Cs / Ms) s + (Ks / Ms)} (dz2-dz1)
In this case, discontinuity does not occur in the pseudo differential term {(Cs / Ms) s + (Ks / Ms)}, and the {1 / (s + Csky / Ms)} term can be configured by a low-pass filter. Therefore, the filter response is stable and appropriate estimation accuracy can be obtained. Here, even if the active sky hook model is adopted, only semi-active control is actually possible, so the controllable area is halved. Therefore, 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.

  From the above relationship, it can be understood that the sprung speed can be estimated if the stroke speed of each wheel is known. Next, since 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. Now, when 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. If the stroke amount bounce term is xsB, the roll term is xsR, the pitch term is xsP, the warp term is xsW, and the stroke amount corresponding to Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR is z_sFL, z_sFR, z_sRL, z_sRR, Holds.

以上の関係式から、xsB、xsR、xsP、xsWの微分ｄxsB等は以下の式で表される。
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） (Formula 1)

From the above relational expression, the differential dxsB of xsB, xsR, xsP, xsW, etc. is expressed by the following expression.
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)

Here, since the relationship between the sprung speed and the stroke speed is obtained from the estimation formula 4, among the estimation formula 4, − (1 / s) · {1 // (s + Csky / Ms)} · {(Cs / Ms) s + (Ks / Ms)} portion is denoted as G, and modal parameters (CskyB, CskyR, CskyP, CsB, CsR, CsP, Csky, Cs, and Ks corresponding to bounce term, roll term, and pitch term, respectively) KsB, KsR, KsP) are GB, GR, GP, bounce rate is dB, roll rate is dR, and pitch rate is dP, dB, dR, dP can be calculated as the following values.
dB = GB · dxsB
dR = GR · dxsR
dP = GP · dxsP
From the above, the state estimation on the spring in the actual vehicle can be achieved based on the stroke speed of each wheel.

(Spring control unit)
Next, the skyhook control configuration executed in the sprung mass damping control unit 101a, the skyhook control unit 201, and the sprung mass damping control unit 33 will be described. In the skyhook control, the sprung state estimated based on the stroke speed as described above is controlled to become the target sprung state. In other words, the stroke speed changes according to the sprung state, and when the sprung state such as bounce, roll, and pitch is controlled to the target sprung state, the detected stroke speed corresponds to the target sprung state. It controls so that it may become the stroke speed to perform.

[Configuration of Skyhook Control Unit]
The vehicle control apparatus according to the first embodiment includes the engine 1, the brake 20, and the S / A 3 as actuators for achieving sprung posture control. Among these, the sprung mass damping control unit 101a in the engine controller 1a has two bounce rate and pitch rate as control targets, and the skyhook control unit 201 in the brake controller 2a has pitch rate as control targets. In the skyhook control unit 33a in the controller 3a, three of bounce rate, roll rate, and pitch rate are controlled.

The amount of skyhook control in the bounce direction is
FB = CskyB · dB
The amount of skyhook control in the roll direction is
FR = CskyR · dR
The amount of skyhook control in the pitch direction is
FP = CskyP · dP
It becomes.
(Skyhook control amount FB in the bounce direction)
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. In addition, the skyhook control unit 33a calculates as a part of the S / A attitude control amount.
(Skyhook control amount FR in roll direction)
The sky hook control amount FR in the roll direction is calculated as part of the S / A attitude control amount in the sky hook control unit 33a.
(Pitch direction skyhook control amount FP)
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. In addition, the skyhook control unit 201 calculates the brake posture control amount. In addition, 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. Thus, 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. In the engine control unit 102, 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 the skyhook control unit 201 in order to prevent the driver from feeling uncomfortable like the engine 1 (the limit value will be described in detail later). Thereby, when 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. Output to the brake control unit 202. 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.

[Brake pitch control]
Here, the brake pitch control will be described. In general, it can be said that it is preferable to perform both of the brakes 20 because both bounce and pitch can be controlled. However, the bounce control by the brake 20 generates braking force at the same time for the four wheels. Therefore, despite the low control priority, the deceleration effect is strong for the difficulty of obtaining the control effect, and the driver tends to feel uncomfortable. there were. Therefore, the brake 20 has a configuration specialized for pitch control. FIG. 6 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, and the pitch rate is Vp. Then, the following relational expression is established.

BFf + BFr = m · a
m · a · Hcg = Mp
Mp = (BFf + BFr) · Hcg
Here, when the pitch rate Vp is positive, that is, when the braking force is applied when the front wheel side is depressed, the front wheel side is further depressed and the pitch motion is promoted. In this case, the braking force is not applied. On the other hand, when the pitch rate Vp is negative, that is, when the front wheel side is lifted, the braking pitch moment gives a braking force to suppress the front wheel side lift. This contributes to improving the sense of security and flatness by ensuring the driver's field of view and making it easier to see the front. From the above
When Vp> 0 (front wheel sinks) Mp = 0
When Vp ≦ 0 (front wheel lift) Mp = CskyP · Vp
The amount of control is given. Accordingly, since the braking torque is generated only when the vehicle body is lifted up on the front side, the generated deceleration can be reduced as compared with the case where the braking torque is generated in both the lifting and sinking. Moreover, since the actuator operation frequency is only half, a low-cost actuator can be employed.

  Based on the above relationship, 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.

[Deceleration feeling reduction processing]
Next, the deceleration feeling reduction process will be described. This process is a process corresponding to the restriction 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.

  In the jerk threshold limiting unit 3342d, 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 a 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.

[Frequency-sensitive control unit]
Next, frequency sensitive control processing in the sprung mass damping control unit will be described. In the first embodiment, the sprung speed control is achieved by estimating the sprung speed based on the detection value of the stroke sensor 14 and performing the skyhook control based on the estimated speed. However, when it is considered that the estimation accuracy cannot be sufficiently ensured by the stroke sensor 14, or when the traveling state and the driver's intention are positive, it is desirable to actively ensure a comfortable driving state (a comfortable ride feeling that is softer than the flat feeling of the vehicle body). In some cases. In such cases, 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 amount of vibration characteristics.

  FIG. 7 is a diagram showing stroke frequency characteristics of the stroke sensor. Here, 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. Of the frequency components of the stroke sensor 14, the region where the sprung resonance frequency component exists is sensed as if the occupant was thrown into the air by swinging the entire body of the occupant, in other words, gravitational acceleration acting on the occupant As a frequency region that brings about a sense that the frequency has decreased, the waving region (0.5 to 3 Hz) is used, and the region between the sprung resonance frequency component and the unsprung resonance frequency component is not a sensation that reduces gravitational acceleration. However, when riding fast, the feeling that the human body jumps in small increments, in other words, the frequency range that brings up and down movement that the whole body can follow is set as a frequency range (3 to 6 Hz), and unsprung resonance It is not a vertical movement until the mass of the human body follows the region where the frequency component exists, but a small vibration is transmitted to a part of the body such as the thigh of the occupant. It is defined as a bull region (6 to 23 Hz) as a simple frequency region.

FIG. 8 is a control block diagram showing 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 stroke 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 frequency amplitude (specifically, an area calculated by 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.

  Here, the setting of human sense weight will be described. FIG. 9 is a correlation diagram showing human sensory characteristics with respect to frequency. As shown in FIG. 9, in the waft region, which is a low frequency region, the sensitivity of the occupant is relatively low with respect to the frequency, and the sensitivity gradually increases as the frequency shifts to the high frequency region. Note that the high frequency region above the bull region becomes difficult to be transmitted to the occupant. From the above, 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, and the human sense weight Wb of the bull area is larger than Wf and Wh. Set to 0.38. Thereby, the correlation between the scalar quantity in each frequency band and the vibration actually propagated to the occupant can be further increased. These two weighting factors may be changed as appropriate according to the vehicle concept and the passenger's preference.

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 stroke sensor value of each wheel.

  The maximum value selection unit 357 selects the maximum value from the final scalar amounts calculated for 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. 10 is a characteristic diagram showing the relationship between the vibration mixing ratio in the wafer region and the damping force by the frequency sensitive control of the first embodiment. As shown in FIG. 10, the vibration level of sprung resonance is reduced by setting the damping force high when the ratio of the wing area is large. At this time, even if 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. On the other hand, when the ratio of the wing region is small, 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.

Here, an advantage of the frequency sensitive control when the frequency sensitive control is compared with the skyhook control will be described. FIG. 11 is a diagram showing frequency characteristics detected by the stroke sensor 14 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. When Skyhook control is performed while traveling on a road surface exhibiting such characteristics, 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.
On the other hand, when the control is based on the scalar quantity instead of the vector as in the frequency sensitive control, a low damping force is set because the ratio of the waft area is small on the road surface as shown in FIG. Become. As a result, even if the amplitude of the vibration in the bull region is large, the vibration transfer characteristic is sufficiently reduced, so that deterioration of high-frequency vibration can be avoided. From the above, high-frequency vibration can be suppressed by frequency-sensitive control based on the scalar amount in a region where control is difficult due to deterioration in phase estimation accuracy even if skyhook control is performed using an expensive sensor or the like. .

(About S / A side driver input controller)
Next, the S / A side driver input control unit will be described. The S / A side driver input control unit 31 calculates a driver input damping force control amount corresponding to the vehicle behavior that the driver wants to achieve based on signals from the steering angle sensor 7 and the wheel speed sensor 8, and a damping force control unit. 35 for output. 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.

(About roll control by S / A side driver input control)
Here, the roll suppression control performed by the S / A side driver input control will be described. FIG. 12 is a control block diagram illustrating a configuration of roll rate suppression control according to the first embodiment. The lateral acceleration estimation unit 31b1 estimates the lateral acceleration Yg based on the front wheel steering angle δf detected by the steering angle sensor 7 and the vehicle speed VSP detected by the vehicle speed sensor 8. The lateral acceleration Yg is calculated from the following equation based on the vehicle body plan view model.
Yg = (VSP ² / (1 + A · VSP ² )) · δf
Here, A is a predetermined value.

  The 90 ° phase advance component creation unit 31b2 differentiates the estimated lateral acceleration Yg and outputs a lateral acceleration differential value dYg. The first addition unit 31b4 adds the lateral acceleration Yg and the lateral acceleration differential value dYg. The 90 ° phase delay component creation unit 31b3 outputs a component F (Yg) obtained by delaying the phase of the estimated lateral acceleration Yg by 90 °. The second adder 31b5 adds F (Yg) to the value added by the first adder 31b4. The Hilbert transform unit 31b6 calculates a scalar quantity based on the envelope waveform of the added value. The gain multiplication unit 31b7 multiplies the scalar amount based on the envelope waveform by the gain, calculates a driver input attitude control amount for roll rate suppression control, and outputs the calculated value to the damping force control unit 35.

  FIG. 13 is a time chart showing the envelope waveform forming process of the roll rate suppressing control according to the first embodiment. When the driver starts steering at time t1, roll rate begins to gradually occur. At this time, the 90 ° phase advance component is added to form an envelope waveform, and 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. it can. Next, when the driver enters the steering holding state at time t2, the 90 ° phase advance component disappears, and this time, the phase delay component F (Yg) is added. At this time, even if the roll rate itself does not change much in the steady turning state, a roll rate resonance component corresponding to the roll back is generated after the roll once. 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.

  When the driver shifts from the steered state to the straight traveling state at time t3, the lateral acceleration Yg decreases and the roll rate converges to a small value. Again, since the damping force is firmly secured by the action of the 90 ° phase delay component F (Yg), instability due to the roll rate resonance component can be avoided.

(Unsprung vibration control unit)
Next, the configuration of the unsprung vibration suppression control unit will be described. As described in the conveyor vehicle of FIG. 5A, since the tire also has an elastic coefficient and a damping coefficient, a resonance frequency band exists. However, since the mass of the tire is smaller than the mass on the spring and the elastic coefficient is high, it exists on the higher frequency side than the resonance on the spring. Due to this unsprung resonance component, the tire may flutter under the unsprung mass, which may deteriorate the ground contact property. In addition, fluttering under the spring may cause discomfort to the occupant. Therefore, in order to suppress the flutter due to unsprung resonance, a damping force corresponding to the unsprung resonance component is set.

  FIG. 14 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 bandpass filter to the stroke speed. The unsprung resonance component is extracted from the region of approximately 10 to 20 Hz among the frequency components. In the envelope waveform shaping unit 342, the extracted unsprung resonance component is scalarized, and the envelope waveform is shaped using the EnvelopeFilter. The gain multiplying unit 343 multiplies the scalarized unsprung resonance component by a gain, calculates an unsprung damping damping force control amount, and outputs it to the damping force control unit 35. In the first embodiment, the unsprung resonance component is extracted by applying a bandpass filter to the stroke speed output from the deviation calculating unit 321b in the traveling state estimating unit 32. The unsprung resonance component may be extracted by estimating and calculating the unsprung speed in conjunction with the sprung speed.

(Configuration of damping force control unit)
Next, the configuration of the damping force control unit 35 will be described. FIG. 15 is a control block diagram illustrating a control configuration of the damping force control unit according to the first embodiment. In the equivalent viscosity damping coefficient conversion unit 35a, 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.

  In the damping coefficient arbitration unit 35b, the damping coefficient converted by the equivalent viscous damping coefficient conversion unit 35a (hereinafter, 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.

[Attenuation coefficient mediation section]
Next, the contents of arbitration by the attenuation coefficient arbitration unit 35b will be described. The vehicle control apparatus according to the first embodiment 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.

In the standard mode, 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.
In the sport mode, 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.
In the 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.
In the highway mode, 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.
Hereinafter, the adjustment of the attenuation coefficient in each mode will be described.

(Arbitration in standard mode)
FIG. 16 is a flowchart illustrating the attenuation coefficient arbitration process in the standard mode according to the first embodiment.
In 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.
In 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.
In step S3, it is determined whether or not the ratio of the bull area is equal to or greater than a predetermined value. If the ratio is greater than or equal to the predetermined value, there is a concern about deterioration of riding comfort due to high-frequency vibration. Set. On the other hand, if the ratio of the bull area is less than the predetermined value, even if the damping coefficient is set high, there is little fear of deterioration in riding comfort due to high-frequency vibration, so the routine proceeds to step S5 and k4 is set.

  As described above, in the standard mode, in principle, priority is given to unsprung vibration suppression control that suppresses unsprung resonance. However, when the damping force required by skyhook control is lower than the damping force required by unsprung vibration suppression control and the ratio of the bull area is large, the damping force of skyhook control is set and Avoid the deterioration of high-frequency vibration characteristics that accompanies the requirements. As a result, it is possible to obtain optimum damping characteristics according to the running state, and at the same time, it is possible to avoid a deterioration in riding comfort against high-frequency vibrations while achieving a flat feeling of the vehicle body.

(Mediation in sport mode)
FIG. 17 is a flowchart showing attenuation coefficient arbitration processing in the sport mode of the first embodiment.
In 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, and the damping force distribution ratio of each wheel is If xfr, xfl, xrr, xrl,
xfr = k1fr / (k1fr + k1fl + k1rr + k1rl)
xfl = k1fl / (k1fr + k1fl + k1rr + k1rl)
xrr = k1rr / (k1fr + k1fl + k1rr + k1rl)
xrl = k1rl / (k1fr + k1fl + k1rr + k1rl)
Is calculated by

In step S12, it is determined whether or not the damping force distribution ratio x is within a predetermined range (greater than α and smaller than β). 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.

In 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.
In 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.
In 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. In other words, 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.

  In 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.

  As described above, in the sport mode, priority is given to unsprung vibration suppression control that suppresses unsprung resonance in principle. However, 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. In addition, with respect to the movement that causes the posture change in the vehicle body posture while the damping force distribution ratio is maintained, the sky vehicle body posture can be maintained by selecting Skyhook control with select high.

(Mediation in Conford mode)
FIG. 18 is a flowchart illustrating the attenuation coefficient arbitration process in the comfort mode according to the first embodiment.
In 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.

  As described above, in the comfort mode, priority is given to unsprung resonance control that basically suppresses unsprung resonance. Originally 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. In the comfort mode, as in the standard mode, 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.

(Arbitration in highway mode)
FIG. 19 is a flowchart illustrating attenuation coefficient arbitration processing 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.
In 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.

  As described above, in the highway mode, the attenuation coefficient is adjusted using a value obtained by adding the S / A attitude attenuation coefficient k2 to the adjusted first attenuation coefficient k. Here, the operation will be described with reference to the drawings. FIG. 20 is a time chart showing a change in attenuation coefficient when traveling on a wavy road surface and an uneven road surface. For example, when trying to suppress the movement of the vehicle body to fluctuate due to slight road surface undulations when driving at high vehicle speeds, it is necessary to detect a slight change in stroke speed if trying to achieve it with only skyhook control Therefore, it is necessary to set the skyhook control gain to be quite high. In this case, the movement that fluctuates can be suppressed, but if the road surface is uneven, the control gain is too large and excessive damping force control may be performed. As a result, there is a concern about deterioration in ride comfort and vehicle body posture.

  On the other hand, since 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.

(Mode selection process)
Next, a mode selection process for selecting each travel mode will be described. FIG. 21 is a flowchart illustrating a mode selection process based on the running state in the attenuation coefficient arbitration unit of the first embodiment.
In 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. move on.
In 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. 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. 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.

  In other words, 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.

  On the other hand, when the vehicle is traveling at a high vehicle speed in a turning traveling state, by selecting the highway mode, control is performed by a value obtained by adding a damping coefficient, so that basically a high damping force can be obtained. As a result, even when the vehicle speed is high, unsprung resonance can be suppressed while positively securing the vehicle body posture during turning by driver input control. In addition, when driving at low vehicle speeds, 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.

  As for the mode selection process, the control example in which the driving state is detected and automatically switched is shown in the first embodiment. However, for example, 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.

As described above, Example 1 has the following effects.
(1) The engine 1, brake 20 and S / A3 (plural actuators) that perform sprung mass damping control, the stroke sensor 14 (stroke sensor) that detects the stroke speed of S / A3, and the stroke sensor 14 The engine 1, so that the sprung state (stroke speed: first traveling state estimation unit 100, second traveling state estimation unit 200, third traveling state estimation unit 32) becomes the target sprung state (corresponding stroke speed). An engine controller 1a that controls the brake 20 and the S / A 3 (each actuator), a brake controller 2a, and an S / A controller 3a (a plurality of actuator attitude control means) are provided.

  Therefore, because the feedback control system is configured in each actuator using a common value of stroke speed, even if individual control is performed without controlling mutual monitoring, the result of mutual monitoring is the result. Control (hereinafter, this control is described as emphasis control) is performed, and the vehicle posture can be converged in the stabilization direction. That is, when a plurality of actuators individually operate without monitoring their operating states, control interference becomes a problem. However, because the feedback control system based on the stroke speed is configured in each actuator, the influence of each actuator control appears as the stroke speed, and the influence of each control is consequently monitored and controlled. This is to avoid interference. For example, when a certain sprung vibration is suppressed by the engine 1, a stroke speed associated therewith is generated. Even if the other actuators do not sense the details of the control performed in the engine 1, the brake 20 and the S / A 3 perform the control based on the stroke speed reflecting the influence.

  In the first embodiment, the vehicle body posture control is performed by the skyhook control. However, the vehicle body posture control may be achieved by another vehicle body posture control. In the first embodiment, the pitch rate is controlled, but bounce rate or the like may be controlled. Further, in the first embodiment, the target posture is a flat posture, but for example, a nose dive-like vehicle body posture may be set as the target posture from the viewpoint of securing the driver's view during turning. Moreover, you may control not only on a sprung attitude | position but for the purpose of unsprung vibration suppression.

(2) The plurality of actuators are engine 1, brake 20 and S / A3, and the plurality of actuator attitude control means are engine controller 1a (power source attitude control means), brake controller 2a (brake attitude control means) and S / A controller 3a (damping force control means).
That is, when performing the sprung mass damping control, for example, when trying to cope with only the S / A3 having a high ability to control the posture on the spring without performing coordinated control with a plurality of actuators, It is necessary to increase the damping force to suppress the movement. However, if the damping force is increased, the transmission rate of vibration on the spring will increase, and if high-frequency vibration occurs due to fine irregularities on the road surface, there is a concern that the so-called high-frequency vibration characteristics will be deteriorated. The Therefore, by reducing the damping force control amount of the S / A 3 by the engine 1 and the brake 20 which are actuators that have nothing to do with the deterioration of the high-frequency vibration characteristics, the vibration transmissibility on the spring can be reduced. Deterioration of vibration characteristics can be avoided. Further, since the damping force control amount can be reduced by the engine 1 and the brake 20, the controllable area of the S / A 3 can be narrowed, and the vehicle body posture control can be achieved with an inexpensive configuration.

(3) The sprung mass damping control unit 101a (power source attitude control means) suppresses bounce motion and pitch motion of the vehicle body.
That is, when performing the sprung mass damping control, for example, when trying to cope with only the S / A3 having a high ability to control the posture on the spring without performing coordinated control with a plurality of actuators, It is necessary to increase the damping force to suppress the movement. However, if the damping force is increased, the transmission rate of vibration on the spring will increase, and if high-frequency vibration occurs due to fine irregularities on the road surface, there is a concern that the so-called high-frequency vibration characteristics will be deteriorated. The Therefore, the bounce movement and pitch movement of the vehicle body are suppressed by the engine 1, which is an actuator that has nothing to do with the deterioration of the high-frequency vibration characteristics, thereby reducing the damping force control amount of the S / A3. Thereby, the vibration transmissibility to a spring top can be reduced and the deterioration of a high frequency vibration can be avoided. Moreover, since S / A3 can concentrate on suppression of roll motion, it can suppress roll motion effectively.

  (4) The engine attitude control unit 101 (power source attitude control means) has a limit value that limits the engine attitude control amount to a predetermined value. In other words, the engine torque control amount is limited to be within a predetermined longitudinal acceleration range when converted to longitudinal acceleration. Therefore, the engine attitude control amount (engine torque control amount) is calculated based on the skyhook control amount FB in the bounce direction and the skyhook control amount FP in the pitch direction. The engine attitude control amount is output as the skyhook control amount in the bounce direction or pitch direction that can be achieved by the value. As a result, the vehicle body posture control can be achieved without causing the passenger to feel uncomfortable.

(5) The skyhook control unit 201 (brake posture control means) suppresses the pitch motion of the vehicle body.
In general, it can be said that it is preferable to perform both of the brakes 20 because both bounce and pitch can be controlled. However, the bounce control by the brake 20 has a tendency to feel uncomfortable for the driver, although the control priority is low, although the bounce control by the brake 20 is difficult to obtain the control effect. Therefore, the brake 20 is given priority over the suppression of the pitch motion over the suppression of the bounce motion. Thereby, a feeling of deceleration can be suppressed and an uncomfortable feeling to the occupant can be reduced.

  Here, in the first embodiment, when the pitch rate Vp is positive, that is, when the braking force is applied when the front wheel side is depressed, the front wheel side is further depressed and the pitch motion is promoted. Is not granted. On the other hand, when the pitch rate Vp is negative, that is, when the front wheel side is lifted, the braking pitch moment gives a braking force to suppress the front wheel side lift. This contributes to improving the sense of security and flatness by ensuring the driver's field of view and making it easier to see the front. Further, since the braking torque is generated only when the vehicle body is lifted on the front side, the generated deceleration can be reduced as compared with the case where the braking torque is generated for both lifting and sinking. Moreover, since the actuator operation frequency is only half, a low-cost actuator can be employed.

In addition, although the example specialized in pitch control was shown in Example 1, in controlling both pitch motion and bounce motion, suppression of pitch motion is performed preferentially, or the control amount is the control amount of the bounce motion. It is good also as controlling by multiplying the gain which becomes small. This is because the object of the present invention is satisfied if the pitch control is performed in preference to the bounce control.
In the first embodiment, the sky hook control is applied as the pitch control. However, any other control theory may be used as long as the braking torque is output to suppress the pitch rate.

(6) The skyhook control unit 201 (brake posture control means) has a limit value that limits the brake posture control amount to a predetermined value so that the rate of change of the vehicle body deceleration is less than or equal to a predetermined value.
Specifically, in the jerk threshold limiting unit 3342d, 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 Judgment is made as to whether the value is within the range of the longitudinal acceleration limit value. If any threshold is exceeded, the target deceleration is corrected to a value within the jerk threshold range, and the target deceleration is set to the limit value. If it exceeds, set it within the limit value. Thereby, the deceleration can be generated so as not to give the driver a sense of incongruity.

(7) The stroke sensor 14 and the sprung state detected by the stroke sensor 14 (stroke speed: the first traveling state estimation unit 100, the second traveling state estimation unit 200, the third traveling state estimation unit 32) are in the target sprung state. The engine controller 1a, the brake controller 2a, and the S / A controller 3a controller (controller) that control the engine 1, the brake 20, and the S / A3 (a plurality of actuators) are provided.
Therefore, because the feedback control system is configured in each actuator using a common value of stroke speed, even if individual control is performed without controlling mutual monitoring, the result of mutual monitoring is the result. Control (hereinafter, this control is described as emphasis control) is performed, and the vehicle posture can be converged in the stabilization direction.

(8) It has a stroke sensor 14, and the engine controller 1a, the brake controller 2a, and the S / A controller 3a controller (controller) detect the sprung state (stroke speed: first running state estimation unit 100, The engine 1, the brake 20, and the S / A 3 (plural actuators) are controlled so that the second traveling state estimation unit 200 and the third traveling state estimation unit 32) become the target sprung state (corresponding stroke speed).
Therefore, because the feedback control system is configured in each actuator using a common value of stroke speed, even if individual control is performed without controlling mutual monitoring, the result of mutual monitoring is the result. Control (hereinafter, this control is described as emphasis control) is performed, and the vehicle posture can be converged in the stabilization direction.

  Although the first embodiment has been described above, the present invention is not limited to the above-described configuration, and includes other configurations. For example, in the first embodiment, the engine 1, the brake 20, and the three actuators S / A 3 are provided as the actuator for performing the sprung mass damping control, but the configuration including only the engine 1 and the brake 20 is provided. It may be a configuration including only the engine 1 and S / A3, or a configuration including only the brake 20 and S / A3. In any combination, a plurality of actuators can operate in cooperation, and control can be stabilized.

1 Engine 1a Engine controller (engine control unit)
2 Brake control unit 2a Brake controller (brake control unit)
3 S / A (Damping force variable shock absorber)
3a S / A controller 5 Wheel speed sensor 6 Integrated sensor 7 Steering angle sensor 8 Vehicle speed sensor 14 Stroke sensor 20 Brake 31 Driver input control unit 32 Running state estimation unit 33 Sprung vibration suppression control unit 33a Skyhook control unit 33b Frequency response Control unit 34 Unsprung vibration suppression control unit 35 Damping force control unit 331 First target posture control amount calculation unit 332 Engine posture control amount calculation unit 333 Second target posture control amount calculation unit 334 Brake posture control amount calculation unit 335 Third target Attitude control amount calculation unit 336 Shock absorber attitude control amount calculation unit

Claims (8)

A plurality of actuators for sprung mass damping control;
A stroke sensor that detects the stroke speed of the shock absorber;
A plurality of actuator attitude control means for controlling each actuator so that the stroke speed detected by the stroke sensor becomes a stroke speed corresponding to a target sprung state;
A vehicle control device comprising: The vehicle control device according to claim 1,
The plurality of actuators are a vehicle power source, a friction brake, and a damping force variable shock absorber,
The vehicle control device, wherein the plurality of actuator control means are power source attitude control means, brake attitude control means, and damping force control means. The vehicle control device according to claim 2,
The power source attitude control means suppresses bounce motion and pitch motion of the vehicle body. The vehicle control device according to claim 2 or 3,
The power source attitude control means has a limit value for limiting the power source attitude control amount to a predetermined value. The vehicle control device according to any one of claims 2 to 4,
The friction brake posture control means suppresses the pitch motion of the vehicle body. The vehicle control device according to any one of claims 2 to 5,
The vehicle control apparatus according to claim 1, wherein the friction brake attitude control means has a limit value for limiting the brake attitude control amount to a predetermined value so that a rate of change of the vehicle body deceleration is not more than a predetermined value. A sensor for detecting the stroke speed of the shock absorber;
A controller that controls a plurality of actuators that perform sprung mass damping control so that the stroke speed detected by the sensor becomes a stroke speed corresponding to a target sprung state;
A vehicle control apparatus comprising: It has a sensor that detects the stroke amount of the shock absorber,
A controller for controlling a plurality of actuators that perform sprung mass damping control so that a stroke speed detected by the sensor becomes a stroke speed corresponding to a target sprung state.

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