JP2009018649A - Damping force control device for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve calculation accuracy of corrected pitch moment, and to reduce a computing load. <P>SOLUTION: A target pitch angle θp* of a vehicle body is given by a function polynomial-approximated by using an actual roll angle θr of the vehicle body. The corrected pitch moment ΔMp is calculated based on a differential value (θp*dd-θpdd) between a value θp*dd obtained by second-order-differentiating the target pitch angle and a value θpdd obtained by second-order-differentiating an actual pitch angle, and a differential value (θp*-θp) between the target pitch angle θp* and the actual differential value θp. In that calculation, by using roll angle acceleration θrdd and pitch angle acceleration θpdd obtained by algebraically calculating a detection value of a sprung acceleration sensor, the single integral value and a double integral value of the roll angle acceleration, and a double integral value of the pitch angle acceleration, differential computing process is eliminated. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a vehicle damping force control device including a shock absorber capable of changing damping force.

  As a damping force control device for this type of vehicle, four wheels are suspended on a vehicle body by a suspension device including a shock absorber, and the damping force of each shock absorber can be individually changed according to the turning state of the vehicle. It has been known. For example, in the damping force control device proposed in Patent Document 1, an upward or downward force acts on the front wheel side and rear wheel side vehicle bodies in a direction that suppresses pitching generated in the vehicle body when the vehicle turns. The sum of the damping forces generated by the front wheel side shock absorber and the sum of the damping forces generated by the rear wheel side shock absorber are respectively controlled. Further, in this damping force control device, a target pitch angle is obtained so that the phases of the roll angle and the pitch angle generated in the vehicle body coincide with each other, and the vehicle body according to the difference between the target pitch angle and the detected pitch angle. Calculate the required correction pitch moment. The roll feeling is improved by controlling the damping force of the shock absorber based on the corrected pitch moment.

  In calculating the corrected pitch moment, first, the actual roll angle θr and pitch angle θp of the vehicle body are detected from the acceleration on the spring detected by the sprung acceleration sensor, and then the detected roll is detected. A target pitch angle θp * is obtained from the angle θr. As shown in FIG. 9, the target pitch angle θp * is obtained by referring to a target pitch angle calculation map in which the target pitch angle θp * is uniquely determined by the roll angle θr. The target pitch angle θp * has a characteristic that increases as the roll angle θr increases according to the target pitch angle calculation map.

Subsequently, a difference value (= θp * −θp) obtained by subtracting the actual pitch angle θp from the target pitch angle θp * is calculated as a corrected pitch angle Δθp, and then the corrected pitch angle obtained by second-order differentiation of the corrected pitch angle Δθp. The acceleration Δθpdd (= (d ² (Δθp) / dt ² )) is calculated.

Finally, a corrected pitch moment ΔMp necessary for correcting the pitch angle is calculated by the following equation.
ΔMp = I · Δθpdd + Kp · Δθp
Here, I represents the moment of inertia around the vehicle lateral axis passing through the center of gravity of the vehicle, and Kp represents a spring coefficient in consideration of pitch rigidity.
JP 2006-327312 A

  However, in the vehicle damping force control apparatus proposed in Patent Document 1, the second-order differential calculation of the corrected pitch angle Δθp (= θp * −θp) is performed in calculating the corrected pitch moment ΔMp. When such second-order differential calculation is performed, the amount of noise is greatly amplified, and the calculated control amount (corrected pitch moment ΔMp) tends to be vibrated. In addition, the calculation load increases by performing the filtering process twice, and the microcomputer is required to have high calculation capability. Therefore, depending on the vehicle, such damping force control cannot be performed.

  An object of the present invention is to cope with the above problem, and is to improve the calculation accuracy of the corrected pitch moment and reduce the calculation load.

  In order to achieve the above object, the present invention is characterized by a shock absorber that is interposed between a vehicle body and four wheels that are suspended by a suspension device on the vehicle body and that can individually change the damping force, and for each of the four wheels. A spring acceleration sensor for detecting the vertical acceleration on the spring, and a roll of the vehicle body about the longitudinal axis of the vehicle by calculating the roll angular acceleration from the detection value of the spring acceleration sensor and integrating the roll angular acceleration on the second floor. A roll angle detection means for detecting an angle, and a pitch angle acceleration is calculated from the detection value of the sprung acceleration sensor, and the pitch angle of the vehicle body around the left-right axis of the vehicle is detected by second-order integration of the pitch angle acceleration. A pitch angle detecting means; a target pitch angle calculating means for calculating a target pitch angle according to the detected roll angle; and a difference between the target pitch angle and the detected pitch angle. Correction moment calculation means for calculating a correction pitch moment required for the vehicle body, target damping force setting means for setting a target damping force required for the shock absorber according to the calculated correction pitch moment, and the target damping A damping force control device for a vehicle comprising damping force control means for controlling the damping force of the shock absorber according to the target damping force set by the force setting means, wherein the target pitch angle calculation means is a roll of the vehicle body A target pitch angle of the vehicle body corresponding to the detected roll angle is given by a function approximated by a polynomial using the angle, and the correction moment calculating means includes a target pitch angle given by a function using the roll angle. Differential value which is a difference value between a value obtained by differentiating twice and a value obtained by differentiating the detected pitch angle twice, and The corrected pitch moment is calculated based on the difference value between the target pitch angle and the detected pitch angle, and at least in calculating the differential difference value, the roll angular acceleration calculated from the detection value of the sprung acceleration sensor and The first-order integral value of the roll angular acceleration, the second-order integral value of the roll angular acceleration, and the pitch angular acceleration calculated from the detection value of the sprung acceleration sensor are used.

  In this invention, the target pitch angle calculation means calculates the target pitch angle according to the roll angle, and the correction moment calculation means calculates the correction pitch moment according to the difference between the target pitch angle and the detected pitch angle. The target damping force setting means sets a target damping force required for the shock absorber according to the corrected pitch moment, and the damping force control means controls the damping force of the shock absorber according to the set target damping force. Therefore, the phase of the roll angle and the pitch angle generated in the vehicle body can be matched, and the roll feeling can be improved. It should be noted that the damping force control performed according to the corrected pitch moment is not necessarily performed on the shock absorbers of all four wheels, and may be performed only on the front wheel side or only on the rear wheel side.

  The roll angle and the pitch angle are calculated using detection values of a sprung acceleration sensor provided on the sprung (vehicle body) for each of the four wheels. From the detection value of the sprung acceleration sensor, the roll angular acceleration and the pitch angular acceleration are obtained by calculation. Therefore, the roll angle detection means detects the roll angle (referred to as the actual roll angle) by integrating the roll angular acceleration into the second order, and the pitch angle detection means detects the pitch angle by integrating the pitch angular acceleration with the second order. (Referred to as the actual pitch angle).

  The target pitch angle calculation means gives the target pitch angle of the vehicle body corresponding to the actual roll angle by a function approximated by a polynomial equation using the roll angle of the vehicle body. The correction moment calculation means includes a differential difference value between a value obtained by differentiating the target pitch angle given by a function using the roll angle twice and a value obtained by differentiating the actual pitch angle twice, and the target pitch angle and the actual pitch angle. The corrected pitch moment is calculated on the basis of the difference value.

  In this case, the value obtained by differentiating the target pitch angle twice is obtained from the roll angular acceleration obtained by the algebraic calculation from the detection value of the sprung acceleration sensor, and the value obtained by integrating the roll angular acceleration with the first order integration and the second order integration. be able to. The value obtained by second-order differentiation of the actual pitch angle is the same as the pitch angular acceleration obtained by algebraic calculation from the detection value of the sprung acceleration sensor. Therefore, in the present invention, at least in calculating the differential difference value, the roll angular acceleration calculated from the detection value of the sprung acceleration sensor, the first-order integral value of the roll angular acceleration, and the second-order integral value of the roll angular acceleration. And the pitch angular acceleration calculated from the detection value of the sprung acceleration sensor, there is no need to perform differential calculation processing. As for the difference value between the target pitch angle and the actual pitch angle, the target pitch angle given by the function of the roll angle (second-order integral value of roll angular acceleration) calculated by the roll angle detection means, and the pitch angle detection means What is necessary is just to calculate the difference value with the pitch angle calculated by (2nd order integral value of pitch angular acceleration).

  As a result, noise generated when calculating the corrected pitch moment is reduced, and the accuracy of the final damping force control amount is improved. Accordingly, it is possible to accurately match the phase of the roll angle and the pitch angle generated in the vehicle body, thereby improving the roll feeling.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. FIG. 1 is a schematic diagram illustrating the entire damping force control apparatus for a vehicle according to the embodiment. This damping force control device includes a shock absorber 10 and a coil spring 20 between a vehicle body BD (sprung member) and left and right front and rear wheels FL, FR, RL, RR, respectively.

  The shock absorber 10 is interposed between an unsprung member LA that functions as a suspension device such as a lower arm and a knuckle connected to the left and right front and rear wheels FL to RR, and a vehicle body BD (sprung member). Is connected to the unsprung member LA at the lower end thereof, and is fixed to the vehicle body BD at the upper end of the piston rod 12 inserted into the cylinder 11 so as to be movable up and down. The coil spring 20 is provided in parallel with the shock absorber 10. The cylinder 11 is partitioned into upper and lower chambers R1 and R2 by a piston 13 that slides liquid-tightly on the inner peripheral surface thereof.

  A variable throttle mechanism 30 is assembled to the piston 13. The variable throttle mechanism 30 switches the degree of opening of the communication path for communicating between the upper and lower chambers R1 and R2 of the cylinder 11 in a plurality of stages by the operation of an actuator 31 constituting a part thereof. According to this switching step, when the opening degree of the communication path increases, the damping force of the shock absorber 10 is set to the soft side, and when the opening degree of the communication path decreases, the damping force of the shock absorber 10 is set to the hard side. It is like that. The damping coefficient of each shock absorber 10 is represented by Cfl, Cfr, Crl, and Crr, corresponding to the left and right front and rear wheels FL, FR, RL, and RR, respectively.

  Next, an electric control device that controls the operation of the actuator 31 will be described. The electric control device includes an electronic control unit 40. The electronic control unit 40 includes a microcomputer composed of a CPU, ROM, RAM, and the like as main components, and repeatedly executes the roll and pitching suppression control program shown in FIG. 3 every predetermined time after the ignition switch is turned on. Control the operation. The electronic control unit 40 is connected to sprung acceleration sensors 41fl, 41fr, 41rl, 41rr, vehicle height sensors 42fl, 42fr, 42rl, 42rr, a vehicle speed sensor 43, a yaw rate sensor 44, and a lateral acceleration sensor 45.

  The sprung acceleration sensors 41fl, 41fr, 41rl, and 41rr are assembled to the vehicle body BD corresponding to the left and right front and rear wheels FL, FR, RL, and RR, respectively. Accelerations Gzfl, Gzfr, Gzrl, Gzrr are detected. The sprung accelerations Gzfl to Gzrr detected by the sprung acceleration sensors 41fl to 41rr indicate that an upward acceleration is generated with respect to the vehicle when positive, and a downward acceleration with respect to the vehicle due to negative. It means that has occurred. The vehicle height sensors 42fl, 42fr, 42rl and 42rr are provided between the unsprung members LA corresponding to the left and right front and rear wheels FL, FR, RL and RR and the vehicle body BD, respectively. RR relative displacements (strokes) Xfl, Xfr, Xrl, and Xrr are detected. In the strokes Xfl to Xrr detected by the vehicle height sensors 42fl to 42rr, the direction in which the interval between the unsprung member LA and the vehicle body BD narrows is positive, and the direction in which the interval increases is negative.

  The vehicle speed sensor 43 detects the vehicle speed V. The yaw rate sensor 44 detects the yaw rate γ. A yaw rate γ detected by the yaw rate sensor 44 indicates that a positive angular velocity is generated around the vehicle vertical axis passing through the center of gravity of the vehicle, and a negative angular velocity is generated around the axis when negative. Represents that you are doing. The lateral acceleration sensor 45 detects the lateral acceleration Gy in the left-right direction of the vehicle. The lateral acceleration Gy detected by the lateral acceleration sensor 45 indicates that an acceleration in the right direction with respect to the vehicle is generated by positive, and an acceleration in the left direction with respect to the vehicle is generated by negative. Represents that.

  Here, before explaining the operation of the vehicle damping force control apparatus according to the embodiment of the present invention, the outline of the suspension system motion will be explained using the vehicle two-wheel model shown in FIG. FIG. 2A is a vehicle model representing the left and right wheels of an actual vehicle. In this actual vehicle model, a shock absorber 10 and a coil spring 20 are provided between the vehicle body BD and the left and right wheels L and R, respectively. It is intervened.

  On the other hand, FIG. 2B is a vehicle model representing the left and right wheels of a virtual vehicle. In this virtual vehicle model, coil springs 20 are respectively provided between the vehicle body BD and the left and right wheels L and R. For example, a lift absorber 110 for suppressing lifting of the vehicle body BD during vehicle turning between the virtual wheel VW traveling in the virtual position inside the turn and the vehicle body BD, and the vehicle body BD. And a roll-suppressing shock absorber 210 for suppressing the roll.

  According to this virtual vehicle model, an increase in vehicle height on the turning inner wheel (corresponding to the left wheel L in FIG. 2B) when the vehicle body BD is rolled is suppressed. Therefore, by adapting this virtual vehicle model to the actual vehicle model shown in FIG. 2A, the vehicle center of gravity O of the actual vehicle model is prevented from rising when the vehicle body BD is rolled, and maneuvering is performed. Can be improved.

Now, in the actual vehicle model shown in FIG. 2A, for example, it is assumed that the vehicle turns to the left. Here, the mass of the vehicle body BD is M, the spring constant of the coil spring 20 is K, the damping coefficient of the shock absorber 10 on the turning inner wheel (left wheel L) side is Cin, and the shock absorber on the turning outer wheel (right wheel R) side. The attenuation coefficient of 10 is Cout. Also, the vertical displacement and acceleration in the absolute space of the vehicle body BD are Xb and Xbdd, the stroke and stroke speed of the turning inner wheel (left wheel L) are Xin and Xind, respectively, and the stroke of the turning outer wheel (right wheel R) and If the stroke speeds are Xout and Xoutd, respectively, the equation of motion of the vehicle body BD in the vertical direction is expressed by the following equation 1.
M · Xbdd = K · Xin + K · Xout + Cin · Xind + Cout · Xoutd… Formula 1
Further, when the roll inertia moment of the vehicle is I, the wheel treads of the left and right wheels L and R are W, and the angular acceleration around the vehicle longitudinal axis passing through the vehicle center of gravity O is θdd, the vehicle longitudinal axis passing through the vehicle center of gravity O The surrounding equation of motion is expressed by the following equation 2.
I · θdd = W · (K · Xin-K · Xout + Cin · Xind-Cout · Xoutd) / 2 ... Formula 2

On the other hand, in the hypothetical vehicle model shown in FIG. 2B, the damping coefficient of the lift suppression shock absorber 110 is Cg, the roll suppression shock absorber 210 is C, and the vehicle center of gravity O and the lift suppression control are shown. Assuming that the distance from the shock absorber 110 is D, for example, the vertical motion equation of the vehicle body BD and the motion equation around the vehicle longitudinal axis passing through the vehicle center of gravity O when the vehicle turns to the left are expressed by the following equations 3 ~ Expressed using Equation 5.
M · Xbdd = K · Xin + K · Xout + T ... Formula 3
I ・ θdd = W ・ (K ・ Xin−K ・ Xout + C ・ Xind−C ・ Xoutd) / 2 + D ・ T
However, T = Cg · Xind · (W + 2D) / (2W) + Cg · Xoutd · (W−2D) / (2W)

From Expression 1 and Expression 3, the following Expression 6 is established.
Cin / Xind + Cout / Xoutd = T Equation 6
Further, from Expression 2 and Expression 4, the following Expression 7 is established.
Cin · Xind-Cout · Xoutd = C · Xind-C · Xoutd + 2D · T / W ... Equation 7
By adding both sides of Expression 6 and Expression 7, the damping coefficient Cin of the shock absorber 10 of the turning inner wheel (left wheel L) at the time of rolling is expressed using Expression 8 below.
Cin = T · (W + 2D) / (2W · Xind) + C · (1−Xoutd / Xind) / 2
Similarly, by subtracting both sides of Expression 6 and Expression 7, the damping coefficient Cout of the shock absorber 10 of the turning outer wheel (right wheel R) at the time of rolling is expressed using Expression 9 below.
Cout = T · (W−2D) / (2W · Xoutd) + C · (1−Xind / Xoutd) / 2 Equation 9
Using Equation 8 and Equation 9, the damping force generated by the shock absorber 10 of the turning inner wheel (left wheel L) and the shock absorber 10 of the turning outer wheel (right wheel R) during rolling is expressed by the following Equation 10 and Equation 11, respectively. It is expressed using.
Fin = Cin ・ Xind
= T ・ (W + 2D) / (2W) + C ・ (Xind−Xoutd) / 2
Fout = Cout ・ Xoutd
= T · (W−2D) / (2W) + C · (Xoutd−Xind) / 2 Equation 11

By applying the two-wheel model of the left and right wheels of the vehicle described above to the front and rear wheels of an actual vehicle, it corresponds to the turning inner front wheel, turning outer front wheel, turning inner rear wheel, and turning outer rear wheel during rolling. The damping coefficients Cfin, Cfout, Crin, and Crout of the shock absorber 10 are expressed using the following formulas 12 to 15, respectively. Here, the stroke speeds of the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel, and the turning outer rear wheel are Xfind, Xfoutd, Xrind, and Xroutd, respectively. In addition, the distance between the front wheel side vehicle center of gravity and the lift suppression shock absorber disposed on the front wheel side is Df, and the distance between the rear wheel side vehicle center of gravity and the rear wheel suppression shock absorber is Dr, The front wheel tread is Wf, and the rear wheel tread is Wr. Further, the damping coefficient of the roll suppressing shock absorber disposed on the front wheel side is Cf, and the damping coefficient of the roll suppressing shock absorber disposed on the rear wheel side is Cr.
Cfin = Tf · (Wf + 2Df) / (2Wf · Xfind) + Cf · (1−Xfoutd / Xfind) / 2
Cfout = Tf · (Wf−2Df) / (2Wf · Xfoutd) + Cf · (1−Xfind / Xfoutd) / 2 Equation 13
Crin = Tr · (Wr + 2Dr) / (2Wr · Xrind) + Cr · (1−Xroutd / Xrind) / 2
Crout = Tr · (Wr−2Dr) / (2Wr · Xroutd) + Cr · (1−Xrind / Xroutd) / 2 Equation 15
However, Tf represents the sum of the damping forces generated by the front wheel side shock absorber 10, and Tr represents the sum of the damping forces generated by the rear wheel side shock absorber 10. It is expressed using.
Tf = Cgf.Xfind. (Wf + 2Df) / (2Wf) + Cgf.Xfoutd. (Wf-2Df) / (2Wf) Equation 16
Tr = Cgr · Xrind · (Wr + 2Dr) / (2Wr) + Cgr · Xroutd · (Wr−2Dr) / (2Wr)
Here, Cgf represents the damping coefficient of the lifting suppression shock absorber disposed on the front wheel side, and Cgr represents the damping coefficient of the lifting suppression shock absorber disposed on the rear wheel side.

  Next, the operation of the embodiment configured as described above will be described. When the occupant operates the ignition key to turn on the ignition switch, the electronic control unit 40 starts to repeatedly execute the roll and pitching suppression control program of FIG. 3 every predetermined short time.

  In this roll and pitching suppression control program, the processing of step S11 to step S14 is repeatedly executed. First, the process from step S11 to step S14 will be briefly described.

  In step S11, the vehicle angle at the time of vehicle turning, that is, the roll angle and pitch angle of the vehicle body BD is detected by calculation (see FIGS. 4 and 5). In step S12, using the roll angle θr and the pitch angle θp obtained by calculation in step S11, the front wheel side target damping that is the sum of the damping forces of the shock absorbers 10 arranged corresponding to the front wheels FL and FR. The force is determined (see FIG. 6). In step S13, a rear wheel side target damping force, which is the sum of the damping forces of the shock absorbers 10 arranged corresponding to the rear wheels RL, RR, is determined (see FIG. 7).

  In step S14, the shock absorbers 10 arranged corresponding to the left and right front and rear wheels FL to RR using the front wheel side target damping force determined in step S12 and the rear wheel side target damping force determined in step S13. And the operation of each actuator 31 is controlled according to the calculated attenuation coefficient (see FIG. 8). After the process of step S14, the execution of the pitching suppression control program is terminated.

  Next, the processing of steps S11 to S14 will be specifically described. In the vehicle angle calculation in step S11, the roll angle calculation program shown in FIG. 4 and the pitch angle calculation program shown in FIG. 5 are executed. First, the roll angle calculation program will be described with reference to FIG.

When this roll angle calculation program is started, the sprung accelerations Gzfl, Gzfr, Gzrl, Gzrr detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr are input in step S21. Next, in step S22, the sprung accelerations Gozl and Gozr of the vehicle body left wheel side and right wheel side center of gravity are calculated using the following equations 18 and 19.
Gozl = (Gzfl·Lr + Gzrl·Lf) / L ... Equation 18
Gozr = (Gzfr · Lr + Gzrr · Lf) / L ... Equation 19
Here, L represents the wheel base of the vehicle, and Lf and Lr represent the center of gravity of the vehicle and the distance between the front axle and the rear axle, respectively.

Next, in step S23, the roll angular acceleration θrdd (= d (θr) ² / dt ² ) around the vehicle longitudinal axis passing through the center of gravity of the vehicle is calculated using the following equation 20.
θrdd = (Gozl−Gozr) / W Equation 20
Here, W represents a wheel tread of the vehicle. In step S24, the roll angle acceleration θrdd calculated using Expression 20 is integrated for the second time to calculate the roll angle θr. Hereinafter, the roll angle θr obtained by this calculation is referred to as an actual roll angle θr. Note that the actual roll angle θr indicates that the vehicle body BD is rolling rightward by positive, and that the vehicle body BD is rolled leftward by negative. After the process of step S24, the execution of this roll angle calculation program is once ended.

Next, the pitch angle calculation program will be described with reference to FIG. When the execution of the pitch angle calculation program is started, the sprung accelerations Gzfl, Gzfr, Gzrl, Gzrr detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr are input in step S31. Subsequently, in step S32, average values Gzf and Gzr of the sprung acceleration on the front wheel side and the rear wheel side of the vehicle body are calculated using the following equations 21 and 22.
Gzf = (Gzfl + Gzfr) / 2 Equation 21
Gzr = (Gzrl + Gzrr) / 2 Equation 22

Next, in step S33, the pitch angular acceleration θpdd (= d (θp) ² / dt ² ) of the vehicle is calculated using the following Equation 23.
θpdd = (Gzr−Gzf) / L Equation 23
Here, L represents the wheel base of the vehicle. Subsequently, in step S34, the pitch angle θp is calculated by second-order integration of the pitch angle acceleration θpdd calculated using Expression 23. Hereinafter, the pitch angle θp obtained by this calculation is referred to as an actual pitch angle θp. The actual pitch angle θp indicates that the vehicle is in a forward leaning posture by being positive. In the vehicle turning state, the actual pitch angle θp is always positive due to the vehicle structure and the like. After the process of step S34, the execution of this pitch angle calculation program is once terminated.

  Returning to the roll and pitching suppression control program of FIG. 3, the determination of the front wheel side target damping force in step S12 will be described. The front wheel side target damping force is calculated by calculating the damping force required for the front wheel side shock absorber 10 in order to suppress the pitching generated in the vehicle body BD during rolling, and the calculated damping force is set as the front wheel side target damping force. To do. Specifically, the electronic control unit 40 executes a front wheel side target damping force determination program shown in FIG.

  When the execution of the front wheel side target damping force determination program is started, a target pitch angle θp * corresponding to the actual roll angle θr obtained by executing the roll angle calculation program is calculated in step S41. This target pitch angle θp * is given by a function f (θr) approximated by a polynomial using the actual roll angle θr of the vehicle body, as shown in the following equation 24. As shown in FIG. 9, this approximate function has a characteristic that the target pitch angle θp * increases nonlinearly as the actual roll angle θr increases.

Subsequently, in step S42, to calculate a target pitch angle theta] p * of the second-order differential value ^{θp * dd (= d 2 (} θp *) / dt 2). In this case, the second-order differential value θp * dd of the target pitch angle θp * is obtained by second-order time differentiation of the approximate number f (θr) shown in Expression 24.
When the approximate function is used, the first-order differential value θp * d (= dθp * / dt) of the target pitch angle θp * is given by the following equation (25).

ここで、θｒdは、実ロール角θｒの１階微分値（ｄθｒ／ｄｔ）である。
従って、目標ピッチ角θｐ＊の２階微分値θｐ＊ddは次式２６にて与えられる。

ここで、θｒddは、実ロール角θｒの２階微分値（ｄ²（θｒ）／ｄｔ²）である。

Here, θrd is a first-order differential value (dθr / dt) of the actual roll angle θr.
Therefore, the second-order differential value θp * dd of the target pitch angle θp * is given by the following equation (26).

Here, θrdd is a second-order differential value (d ² (θr) / dt ² ) of the actual roll angle θr.

  As can be seen from Equation 26, the second-order differential value θp * dd of the target pitch angle θp * is obtained by using the actual roll angle θr, the first-order differential value θrd of the actual roll angle, and the second-order differential value θrdd of the actual roll angle. Can be sought.

  In this case, the second-order differential value θrdd of the actual roll angle is already calculated as the roll angular acceleration θrdd in step S23 when the roll angle calculation program (FIG. 4) is executed. The roll angular acceleration θrdd is a value obtained by algebraic calculation shown in Expression 20 with detection values (Gzfl, Gzfr, Gzrl, Gzrr) detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr. The actual roll angle θr is the final calculation result obtained by executing the roll angle calculation program, and is obtained by integrating the roll angular acceleration θrdd with the second-order time integration as shown in step S24. Further, the first-order differential value θrd of the actual roll angle is calculated in the calculation process of this step S24 (at the time integration calculation of the first floor).

  Therefore, in step S42, the actual roll angle θr calculated by the roll angle calculation program, the first-order differential value θrd of the actual roll angle, and the second-order differential value θrdd of the actual roll angle are set to 2 of the target pitch angle θp *. Calculate the derivative value θp * dd. Therefore, this calculation does not include differential calculation processing.

When the calculation in step S42 is completed, in step S43, the corrected pitch angle Δθp is calculated. The corrected pitch angle Δθp is a difference value obtained by subtracting the actual pitch angle θp from the target pitch angle θp *, and is calculated by the following equation (27).
Δθp = θp * −θp Equation 27
In this case, the target pitch angle θp * is a value calculated from the actual roll angle θr by the function f (θr) in step S41, and the actual pitch angle θp is calculated by executing the pitch angle calculation program shown in FIG. Value.

Subsequently, in step S44, a corrected pitch angular acceleration Δθpdd is calculated. This corrected pitch angular acceleration Δθpdd is a difference value (differential difference value) obtained by subtracting the second-order differential value θpdd of the actual pitch angle from the second-order differential value θp * dd of the target pitch angle θp *. Is calculated by
Δθpdd = θp * dd−θpdd Equation 28
In this case, the second-order differential value θp * dd of the target pitch angle θp * is the value calculated in step S42, and the second-order differential value θpdd of the actual pitch angle is calculated in step S33 of the pitch angle calculation program. Is the pitch angular acceleration θpdd.

Subsequently, in step S45, a corrected pitch moment ΔMp necessary for correcting the pitch angle of the vehicle body is calculated. This corrected pitch moment ΔMp is calculated by the following equation 29.
ΔMp = I · Δθpdd + Kp · Δθp Equation 29
Here, I represents the moment of inertia around the vehicle lateral axis passing through the center of gravity of the vehicle, and Kp represents a spring coefficient in consideration of pitch rigidity.

  Subsequently, in step S46, the corrected pitch moment ΔMp is divided by the distance Lf between the vehicle center of gravity and the front axle, and converted into a vertical force acting on the front wheel side vehicle body (ΔTf = ΔMp / Lf). Next, in step S47, the front wheel side target damping force Tf * is set to (Tf + ΔTf). Here, Tf is the total sum of the damping forces of the front wheel side shock absorber 10 that is currently set (see Equation 16).

  Next, in step S48, the front wheel side target damping force Tf * (= (Tf + ΔTf)) is set as the sum Tf of the damping force of the front wheel side shock absorber 10 used in the above equations 12 and 13. After the processing of step S48, the execution of the front wheel side target damping force determination program is terminated.

  Returning to the roll and pitching suppression control program of FIG. 3, the determination of the rear wheel side target damping force in step S13 will be described. The rear wheel side target damping force is calculated by calculating the damping force required for the rear wheel side shock absorber 10 to suppress the pitching generated in the vehicle body BD during rolling, and the calculated damping force is used as the rear wheel side target damping force. Set as force. Specifically, the electronic control unit 40 executes a rear wheel side target damping force determination program shown in FIG.

When the execution of the rear wheel side target damping force determination program is started, the vehicle speed V detected by the vehicle speed sensor 43, the yaw rate γ detected by the yaw rate sensor 44, and the lateral acceleration sensor 45 are detected in step S51. Input the lateral acceleration Gy. Next, in step S52, when the side slip angle of the vehicle center of gravity is β, the side slip angular velocity dβ / dt of the vehicle center of gravity is calculated using the following equation 30 from the equation of motion in the vehicle lateral direction at the vehicle center of gravity.
dβ / dt = (Gy / V) −γ Equation 30

Next, in step S53, the side slip angular velocity dβ / dt of the vehicle center of gravity calculated in step S52 is time integrated to calculate the side slip angle β of the vehicle center of gravity. Subsequently, in step S54, considering that the rear wheel has a speed component of the center of gravity of the vehicle and a speed component due to rotation around the center of gravity of the vehicle, the side slip angle θyr of the rear wheel is calculated using the following equation 31.
θyr = (γ · Lr / V) −β Equation 31
Here, Lr represents the distance between the center of gravity of the vehicle and the rear axle.

Next, in step S55, the estimated rear wheel lateral force Yr is calculated based on the following equation 32 using the rear wheel slip angle θyr calculated in step S54.
Yr = Cr · θyr / (TrS + 1) ... Equation 32
Here, Cr represents the cornering force of the rear wheel per unit side slip angle in the region of the side slip angle in which the cornering power at the rear wheel, that is, the cornering force generated at the rear wheel increases substantially in proportion to the side slip angle. Further, Tr is a time constant of the tire constituting the rear wheel, and takes into account a delay time in which the lateral force is generated after the elastic deformation by the tire.

Subsequently, in step S56, the estimated jackup force Jr acting on the rear wheel side vehicle body is calculated based on the following equation 33 using the estimated rear wheel lateral force Yr calculated in step S55.
Jr = Kjr · Yr ² ... Formula 33
Here, Kjr represents a jackup coefficient in consideration of a change in geometry of the rear wheel side unsprung member LA or the like.

  Next, in step S57, the rear wheel side target damping force Tr * is set to a force (-Jr) in a direction opposite to the rear wheel side estimated jackup force Jr. Next, in step S58, the rear wheel side target damping force Tr * (= −Jr) is set as the sum Tr of the damping force generated by the rear wheel side shock absorber 10 used in the above equations 14 and 15. After the process of step S58, the execution of the rear wheel side target damping force determination program is terminated.

  Returning to the roll and pitching suppression control program of FIG. 3, the damping force control of each shock absorber in step S14 will be described. The damping force control of each shock absorber is performed by adapting the above-described virtual vehicle model (see FIG. 2B) to the actual vehicle model (see FIG. 2A), so that the vehicle body BD at the time of turning of the vehicle The rise of the center of gravity of the vehicle is suppressed while suppressing the roll of the vehicle. Specifically, the electronic control unit 40 executes each shock absorber damping force control program shown in FIG.

  When the execution of each shock absorber damping force control program is started, the strokes of the left and right front and rear wheels FL, FR, RL, RR with respect to the vehicle body BD detected by the vehicle height sensors 42fl, 42fr, 42rl, 42rr in step S61. Enter Xi (i = fl, fr, rl, rr) respectively. Further, the lateral acceleration Gy of the vehicle detected by the lateral acceleration sensor 45 is input, and the process proceeds to step S62.

  In step S62, it is determined whether or not the absolute value | Gy | of the lateral acceleration Gy is larger than a predetermined threshold value Gyo, that is, whether or not the damping force control of each shock absorber 10 is necessary. First, a case where the vehicle is traveling straight ahead will be described. In this case, since the magnitude of the lateral acceleration Gy is almost “0”, “No” is determined in step S62, and in step S63, the damping coefficient Ci (i = fl, fr, rl) of each shock absorber 10 is determined. , Rr) is set to a predetermined value (for example, a soft-side attenuation coefficient) suitable for straight traveling. After the process of step S63, in step S70, the operation of the corresponding actuator 31 is controlled according to the attenuation coefficient Ci (i = fl, fr, rl, rr) set to the predetermined value. After the process of step S70, the execution of each shock absorber damping force control program is temporarily terminated.

  Next, a case where the vehicle starts to turn leftward, for example, will be described. In this case, if “Yes”, that is, if the absolute value | Gy | of the lateral acceleration Gy is determined to be larger than the predetermined threshold value Gyo in step S62, the lateral acceleration Gy is time-differentiated to obtain the differential value ΔGy in step S64. It is calculated, and it is determined whether or not the absolute value | ΔGy | of the differential value ΔGy is larger than a predetermined threshold value ΔGyo, that is, whether the roll angle of the vehicle body BD is increasing or decreasing. Since the vehicle is currently turning left and the roll angle of the vehicle body BD is increasing, it is determined in step S64 that “Yes”, that is, the absolute value | ΔGy | of the differential value ΔGy is larger than the predetermined threshold value ΔGyo. Then, the process proceeds to step S65.

  In step S65, the stroke Xi (i = fl, fr, rl, rr) of the left and right front and rear wheels FL, FR, RL, RR with respect to the vehicle body BD is time-differentiated, and the stroke speed Xid (i = fl, fr, rl) is obtained. , Rr). Next, in step S66, it is determined whether or not the lateral acceleration Gy is positive. Currently, since the vehicle is turning leftward, it is determined as “Yes” in Step S66, and the processes after Step S67 are executed.

  In step S67, the stroke speed Xfld of the left front wheel FL is set as the stroke speed Xfind of the turning inner front wheel, the stroke speed Xfrd of the right front wheel FR is set as the stroke speed Xfoutd of the turning outer front wheel, and the stroke speed of the left rear wheel RL. Xrld is set as the stroke speed Xrind of the turning inner rear wheel, and the stroke speed Xrrd of the right rear wheel RR is set as the stroke speed Xroutd of the turning outer rear wheel.

  Next, in step S68, the damping coefficient Cj (j = fin, fout) of the shock absorber 10 corresponding to the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel, and the turning outer rear wheel based on the above formulas 12 to 15. , Rin, rout). In this case, Tf used in the above formulas 12 and 13 is set to the front wheel side target damping force Tf * by the execution of the front wheel side target damping force determination program in FIG. Tr used in 15 is set to the rear wheel side target damping force Tr * by executing the rear wheel side target damping force determination program of FIG.

  Subsequently, in step S69, the damping coefficient Cfin of the shock absorber 10 for the turning inner front wheel is set as the damping coefficient Cfl of the shock absorber 10 for the left front wheel FL, and the damping coefficient Cfout of the shock absorber 10 for the turning outer front wheel is set to that of the right front wheel FR. The damping coefficient Cfr of the shock absorber 10 is set as the damping coefficient Cfr of the rear rear wheel, and the damping coefficient Crl of the shock absorber 10 of the left rear wheel RL is set. The coefficient Crout is set as the damping coefficient Crr of the shock absorber 10 for the right rear wheel RR.

  Next, in step S70, the operation of the corresponding actuator 31 is controlled according to the damping coefficient Ci (i = fl, fr, rl, rr) set in step S69. After the process of step S70, the execution of each shock absorber damping force control program is temporarily terminated. Thereafter, while the roll angle of the vehicle body BD is increasing, the processes of steps S61 to S62 and steps S64 to S70 are repeatedly executed.

  From this state, when the roll angle of the vehicle body BD reaches almost the maximum value and the increase of the roll angle is stopped, “No”, that is, the absolute value | ΔGy | of the differential value ΔGy of the lateral acceleration Gy from the predetermined threshold value ΔGyo in step S64. In step S71, the damping coefficient Ci (i = fl, fr, rl, rr) of each shock absorber 10 is set to a predetermined value suitable for turning (for example, hard-side damping). Coefficient). After the processing in step S71, the operation of the corresponding actuator 31 is controlled in step S70 according to the attenuation coefficient Ci (i = fl, fr, rl, rr) set to the predetermined value. After the process of step S70, the execution of each shock absorber damping force control program is temporarily terminated.

  If the roll angle of the vehicle body BD starts to decrease from the above-mentioned turning traveling to the straight traveling, “Yes”, that is, the absolute value | ΔGy | of the differential value ΔGy of the lateral acceleration Gy is set to a predetermined value again in step S64. It is determined that the threshold value ΔGyo is greater than the threshold value ΔGyo, and thereafter, the processing from step S65 to step S70 is executed as described above.

  On the other hand, when the vehicle starts to turn to the right, for example, from the straight running, the process of Steps S61 to S62, Steps S64, and S65 is executed, as in the case where the vehicle starts to turn to the left. In S66, "No", that is, it is determined that the vehicle is turning rightward, and the processing after Step S72 is executed. In step S72, the stroke speed Xfrd of the right front wheel FR is set as the stroke speed Xfind of the turning inner front wheel, the stroke speed Xfld of the left front wheel FL is set as the stroke speed Xfoutd of the turning outer front wheel, and the stroke speed of the right rear wheel RR. Xrrd is set as the stroke speed Xrind of the turning inner rear wheel, and the stroke speed Xrld of the left rear wheel RL is set as the stroke speed Xroutd of the turning outer rear wheel.

  In step S73, the damping coefficient Cj () of the shock absorber 10 corresponding to the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel, and the turning outer rear wheel is calculated based on the above formulas 12 to 15 in the same manner as in step S68. j = fin, fout, rin, rout) are calculated. In step S74, the damping coefficient Cj (j = fin, fout, rin, rout) calculated in step S73 is used as the damping coefficient Ci (i = fr, fl, rr, rl) of the shock absorber 10 corresponding to the right and left front and rear wheels. Set each. In step S70, the operation of the corresponding actuator 31 is controlled according to the damping coefficient Ci (i = fr, fl, rr, rl) set in step S74. After the process of step S70, the execution of each shock absorber damping force control program is temporarily terminated. Thereafter, while the roll angle of the vehicle body BD is increasing, the processes in steps S72 and subsequent steps are repeatedly executed through the processes in steps S61 to S62 and S64 to S66 as described above. Then, when the increase in the roll angle of the vehicle body BD is stopped, the processing after step S71 is executed after the processing in step S64. From this state, when the roll angle of the vehicle body BD starts to decrease, the processing after step S72 is executed after the processing of step S66 again.

  As can be understood from the above description of the operation, according to the above embodiment, the rear wheel side shock absorber is arranged so that the rear wheel side jack-up force Jr is canceled by the processing of steps S57 and S58 in FIG. 10 is set. For this reason, for example, as shown in FIG. 10, the lifting of the rear wheel side vehicle center of gravity Or is suppressed.

  Further, on the front wheel side of the vehicle body, the corrected pitch moment ΔMp is calculated so that the target pitch angle θp * is uniquely determined by the function of the actual roll angle θr, and the front wheel side shock absorber is processed by steps S46 and S47 in FIG. 10 is set to the total sum of damping forces (Tf + ΔTf). That is, for example, when the actual pitch angle θp is larger than the target pitch angle θp *, it is difficult to keep the actual pitch angle θp away from the target pitch angle θp * so that excessive forward tilt of the vehicle body BD is suppressed. A force is applied to the side vehicle center of gravity Of so that the sum of the inputs is upward (see FIG. 10). On the other hand, for example, when the actual pitch angle θp is smaller than the target pitch angle θp *, the actual pitch angle θp can be easily approached to the target pitch angle θp * to promote forward leaning of the vehicle body BD to a predetermined position. As described above, a force is applied to the front wheel side vehicle center of gravity Of so that the sum of the inputs is downward.

  Accordingly, when the vehicle turns, the roll of the vehicle body BD is suppressed, the rise of the vehicle center of gravity Oc is suppressed, and the pitching of the vehicle body BD is suppressed, and the vehicle body BD is tilted in a predetermined forward tilt posture. Therefore, the maneuverability and the grip feeling on the road surface are improved.

  In this embodiment, since the target pitch angle θp * is uniquely obtained from the actual roll angle θr, the phases of the actual roll angle θr and the actual pitch angle θp can be matched, and the actual roll angle θr And the actual pitch angle θp disappears, and the roll feeling is improved, that is, the smooth feeling during rolling is improved.

  Further, in the present embodiment, as shown in Expression 24, the target pitch angle θp * of the vehicle body corresponding to the actual roll angle θr is given by a function f (θr) approximated by a polynomial using the actual roll angle θr. I have to. Accordingly, in calculating the corrected pitch moment ΔMp, there is no need to perform differential calculation processing for the following reason.

  The corrected pitch moment ΔMp is given as ΔMp = I · Δθpdd + Kp · Δθp. The corrected pitch angular acceleration Δθpdd can be calculated as Δθpdd = θp * dd−θpdd, and the corrected pitch angle Δθp can be calculated as Δθp = θp * −θp.

The second-order differential value θp * dd of the target pitch angle θp * required to calculate the corrected pitch angular acceleration Δθpdd is the actual roll angle θr and the first-order differential value θrd of the actual roll angle, as can be seen from the above equation 26. (= Dθr / dt), the second-order differential value θrdd (= d ² (θr) / dt ² ) of the actual roll angle. The second-order differential value θrdd of the actual roll angle is calculated as the roll angular acceleration θrdd in step S23 when the roll angle calculation program (FIG. 4) is executed. This is a value obtained by algebraic calculation of detection values (Gzfl, Gzfr, Gzrl, Gzrr) detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr. The actual roll angle θr is obtained by integrating the roll angular acceleration θrdd in the second-order time in step S24 of the roll angle calculation program. The first-order differential value θrd of the actual roll angle is calculated in the calculation process (at the time of the first-floor time integration calculation) in step S24.

  The second-order differential value θpdd of the actual pitch angle θp is the pitch angle acceleration θpdd calculated in step S33 of the pitch angle calculation program, and this is also detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr. This is a value obtained by algebraic calculation of the detected value.

  On the other hand, the target pitch angle θp * required to calculate the corrected pitch angle Δθp is a value calculated from the actual roll angle θr by the function f (θr), and the other actual pitch angle θp is the pitch angle. It is a value calculated by executing the calculation program. The actual roll angle θr and the actual pitch angle θp are two acceleration values (roll angular acceleration, pitch angular acceleration) obtained by algebraic calculation of the detection values detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr. This is calculated by step integration (steps S24 and S34).

  Therefore, to calculate the corrected pitch moment ΔMp, it is only necessary to perform the algebra calculation of the detection values detected by the sprung acceleration sensors 41fl, 41fr, 41rl, 41rr, and the first-order integration and second-order integration of the algebra calculation results. Does not require differential processing. As a result, noise generated when calculating the corrected pitch moment ΔMp is reduced, and the accuracy of the final damping force control amount is improved. Thereby, it becomes possible to make the phase of the roll angle and pitch angle which generate | occur | produce in a vehicle body correspond precisely, and a roll feeling can be improved. In addition, the calculation load of the microcomputer that constitutes the main part of the electronic control unit 40 is reduced, and it is possible to perform such damping force control processing on various vehicles.

  The embodiments of the present invention have been described above. However, the present invention is not limited to the above-described embodiments and the like, and various modifications can be made without departing from the object of the present invention.

1 is a schematic diagram showing an entire vehicle damping force control apparatus according to an embodiment of the present invention. (A) is explanatory drawing for demonstrating the motion of a suspension system using the vehicle model showing the right-and-left wheel of the actual vehicle, (B) uses the vehicle model showing the left-right wheel of the virtual vehicle. FIG. 6 is an explanatory diagram for illustrating the motion of the suspension system. It is a flowchart which shows the roll and pitching suppression control program which are performed by the electronic control unit. It is a flowchart which shows the roll angle calculation program performed by the electronic control unit. It is a flowchart which shows the pitch angle calculation program performed by the electronic control unit. It is a flowchart which shows the front-wheel side target damping force determination program performed by the electronic control unit. It is a flowchart which shows the rear-wheel side target damping force determination program performed by the electronic control unit. It is a flowchart which shows the damping force control program of each shock absorber performed by the electronic control unit. It is a graph which shows the change characteristic of the target pitch angle with respect to an actual roll angle. It is explanatory drawing which shows the input of the up-down direction which acts on the front-wheel side and the rear-wheel side vehicle body by execution of a roll and a pitching suppression control program, respectively.

Explanation of symbols

BD ... vehicle body, LA ... unsprung member (suspension device), FL, FR, RL, RR ... left and right front and rear wheels, 10 ... shock absorber, 20 ... coil spring, 30 ... variable throttle mechanism, 31 ... actuator, 40 ... electronic control Unit: 41fl, 41fr, 41rl, 41rr ... sprung acceleration sensor, 42fl, 42fr, 42rl, 42rr ... vehicle height sensor, 43 ... vehicle speed sensor, 44 ... yaw rate sensor, 45 ... lateral acceleration sensor.

Claims (1)

A shock absorber that is interposed between the vehicle body and the four wheels that are suspended from the vehicle body by a suspension device and that can individually change the damping force;
A sprung acceleration sensor that detects the vertical acceleration on each spring;
Roll angle detection means for calculating the roll angular acceleration from the detection value of the sprung acceleration sensor and detecting the roll angle of the vehicle body about the longitudinal axis of the vehicle by integrating the roll angular acceleration in the second order;
Pitch angle detection means for calculating the pitch angular acceleration from the detection value of the sprung acceleration sensor, and detecting the pitch angle of the vehicle body around the left-right axis of the vehicle by integrating the pitch angular acceleration in the second order;
Target pitch angle calculating means for calculating a target pitch angle according to the detected roll angle;
Correction moment calculation means for calculating a correction pitch moment required for the vehicle body according to the difference between the target pitch angle and the detected pitch angle;
Target damping force setting means for setting a target damping force required for the shock absorber according to the calculated corrected pitch moment;
A damping force control device for a vehicle comprising: damping force control means for controlling the damping force of the shock absorber according to the target damping force set by the target damping force setting means;
The target pitch angle calculating means gives a target pitch angle of the vehicle body corresponding to the detected roll angle by a function approximated by a polynomial using the roll angle of the vehicle body,
The correction moment calculation means includes a differential difference value that is a difference value between a value obtained by differentiating the target pitch angle given by a function using the roll angle twice and a value obtained by differentiating the detected pitch angle twice. And calculating the corrected pitch moment based on the difference value between the target pitch angle and the detected pitch angle, and at least calculating the differential difference value, the roll calculated from the detection value of the sprung acceleration sensor The angular acceleration, the first-order integral value of the roll angular acceleration, the second-order integral value of the roll angular acceleration, and the pitch angular acceleration calculated from the detection value of the sprung acceleration sensor are used. Vehicle damping force control device.

Images