JP2010095210A - Vehicular suspension device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a vehicular suspension device capable of restraining the deterioration in ride comfort, even when damping force changing control cannot follow damping of vibration from a road surface. <P>SOLUTION: A suspension ECU 21 performs band-pass filter processing on a signal representing sprung acceleration x<SB>pb</SB>'' inputted from a sprung acceleration sensor 22 and a signal representing unsprung acceleration x<SB>pw</SB>'' inputted from an unsprung acceleration sensor 23 in a relatively high frequency side band, and acquires maximum signal amplitude (b) and maximum signal amplitude (a) passed by the band-pass filter processing. Next, the ECU 21 compares a value of the ratio of the maximum signal amplitude (b) to the maximum signal amplitude (a) with a performance target index α preset to "1" or less, and sets request damping force Freq to constant request damping force Fconst when a sprung-unsprung relative speed V is a reference relative speed Vo or more, since control followability is deteriorated, when a value of b/a is larger than α. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to a vehicle suspension apparatus including a damping force generating means whose damping coefficient is variably controlled.

  A vehicle suspension device generally includes a suspension spring and a shock absorber. The suspension spring absorbs vibration from the road surface. On the other hand, the shock absorber is interposed between the sprung member supported by the suspension spring and the unsprung member connected to the lower end of the suspension spring to support the suspension spring, and attenuates vibration from the road surface. To do. Conventionally, the damping force of the vibration by the shock absorber, specifically, the damping force with respect to the sprung-unsprung relative speed (the difference between the vertical speed of the sprung member and the vertical speed of the unsprung member). A suspension device in which the damping coefficient representing the size of the vehicle can be changed according to the traveling state of the vehicle is well known.

  As this type of suspension device, for example, in Patent Document 1 described below, various determinations can be made by accurately determining the road surface state based on the sprung vertical acceleration and controlling the damping force of the damper according to the determined road surface state. A road surface state determination method that can ensure a comfortable ride in the road surface state is shown. This conventional road surface condition determination method performs bandpass filtering on the vertical acceleration on the spring of the suspension device detected by the vertical acceleration sensor on the spring, and the effective value of vibration in the on-spring resonance frequency range of 0.7 Hz to 1.0 Hz, The effective value of vibration in the intermediate frequency region between 3.0 Hz and 8.0 Hz between the unsprung resonance frequency region and the unsprung resonance frequency region is calculated, and (effective value of vibration in the unsprung resonance frequency region) / (intermediate) By comparing the ratio expressed by the effective value of vibration in the frequency domain) with a threshold value, the road surface condition determination accuracy is improved. The damping force of the damper is controlled according to the determined road surface condition.

  Further, for example, Patent Document 2 below discloses a damping force variable shock absorber control device capable of performing smooth control with little sense of control discomfort. This conventional control device calculates an actuator signal instruction value based on the sprung speed and the sprung-unsprung relative speed, and controls the variable throttle valve based on the calculated value to adjust the damping force. . When the vehicle behavior is increased by the applied damping force, the damping force is corrected slightly lower.

Furthermore, for example, Patent Document 3 shown below discloses a vehicle suspension device that effectively reduces road surface vibration input that induces unsprung resonance. In this conventional vehicle suspension device, an acceleration signal on a spring is input to a feedback circuit, a reference signal corresponding to a road surface vibration input that induces an unsprung resonance is generated, and the reference signal and the unsprung acceleration signal are filtered. Input to the circuit. The filter circuit performs feed-forward control, and the unsprung acceleration signal is canceled by the control gain of the filter circuit.
JP 2006-82755 A JP-A-6-106944 Japanese Patent Laid-Open No. 6-106938

  By the way, in the vehicle suspension apparatus disclosed in Patent Documents 1 to 3, it is assumed that the damping force of the damper (shock absorber) can always be changed. However, when the sprung-unsprung relative speed increases, a larger damping force is required. In this case, the required damping force cannot be generated in a timely manner, that is, the control may not follow. Further, as the frequency of vibration of the sprung member increases, for example, a response delay of the actuator may occur, and in this case as well, the required damping force cannot be generated in a timely manner, that is, Control tracking may not be possible. As described above, in a situation where the control cannot follow, for example, there is a high possibility that a damping force for damping the vibration is applied with a delay, and as a result, there is a concern that the riding comfort may deteriorate.

  The present invention has been made to solve the above-described problems, and the object of the present invention is to follow the fact that the damping force change control attenuates the vibration from the road surface in the suspension device that is controlled to change the damping force. An object of the present invention is to provide a suspension device for a vehicle that can suppress the deterioration of the ride comfort even when there is not.

  In order to achieve the above object, a feature of the present invention is that it is disposed between an unsprung member and a sprung member of a vehicle, and generates a damping force that attenuates vibration of the sprung member with respect to the unsprung member. In a vehicle suspension apparatus comprising a damping force generating means and a damping force change control means for changing and controlling the damping force generated by the damping force generating means, the damping force changing means is arranged in the vertical direction of the unsprung member. Unsprung acceleration detecting means for detecting acceleration; sprung acceleration detecting means for detecting vertical acceleration of the sprung member; vertical speed of the unsprung member; and vertical speed of the sprung member; A sprung-unsprung relative speed detecting means for detecting a sprung-unsprung relative speed representing a relative difference between the sprung member and a required damping force required to damp vibration of the sprung member with respect to the unsprung member. And a ratio of the magnitude of the acceleration of the sprung member detected by the sprung acceleration detecting means to the magnitude of the acceleration of the unsprung member detected by the unsprung acceleration detecting means. Comparison determining means for determining whether or not the value is larger than a predetermined value set in advance to determine that the vibration of the sprung member is attenuated by the damping force generated by the damping force generating means. And the sprung-unsprung relative speed detected by the sprung-unsprung relative speed detecting means when the comparison determining unit determines that the ratio value is greater than the predetermined value. The required damping force required to attenuate the vibration of the sprung member with respect to the unsprung member when the absolute value of is equal to or higher than a preset reference relative speed is predetermined. The required damping force setting means for setting a constant constant value and the damping force generated by the damping force generating means are determined by the required damping force calculated by the required damping force calculating means or by the required damping force setting means. And a damping force changing unit that changes the required damping force to be a constant value. In this case, the predetermined value used by the comparison determination unit may be set in advance to a value of “1” or less.

  In this case, the required damping force calculation means calculates a variable damping coefficient that is a variable part of the damping coefficient of the damping force generation means based on the nonlinear H∞ control theory, and the calculated variable damping coefficient is calculated in advance. A required damping coefficient calculating means for calculating a requested damping coefficient by adding a fixed damping coefficient that is a fixed part of the set damping coefficient, and the spring is applied to the requested damping coefficient calculated by the requested damping coefficient calculating means. The required damping force may be calculated by multiplying the sprung-unsprung relative speed detected by the upper-unsprung relative speed detecting means.

  In addition, the maximum amplitude of the acceleration of the unsprung member after the comparison determination unit performs bandpass filtering on the signal representing the acceleration of the unsprung member detected by the unsprung acceleration detection unit with a band of a predetermined frequency. And a signal representing the acceleration of the sprung member detected by the sprung acceleration detecting unit after bandpass filtering is performed on a band of a predetermined frequency. Sprung acceleration maximum amplitude detecting means for detecting the maximum amplitude of acceleration, and the sprung acceleration maximum amplitude detecting means for the maximum amplitude of acceleration of the unsprung member detected by the unsprung acceleration maximum amplitude detecting means. The ratio value of the maximum amplitude of acceleration of the sprung member detected by the Of it may determine whether greater than the value.

  According to these, the comparison / determination means determines that the ratio of the magnitude of the acceleration of the sprung member to the magnitude of the acceleration of the unsprung member is “1” in order to determine that the vibration of the sprung member is attenuated. By determining whether or not the following value is larger than a predetermined value set in advance, it is determined whether or not the control for changing the damping force to attenuate the vibration of the sprung member is following. it can. In particular, by comparing and determining the ratio of the maximum amplitude of the acceleration of the sprung member to the maximum amplitude of the acceleration of the unsprung member subjected to the bandpass filter processing by a predetermined frequency band and the predetermined value, for example, a bandpass filter By performing the processing in the high frequency band, it is possible to more accurately determine the deterioration of the control followability with respect to the vibration of the sprung member in the high frequency range.

  The required damping force setting means is configured so that the sprung-unsprung relative speed is equal to or higher than the reference relative speed in a state in which the control followability with respect to the vibration of the sprung member (more preferably, vibration in the high-frequency region of the sprung member) is deteriorated. If the required damping force is a predetermined constant value, more specifically, the required damping force becomes smaller than the required damping force calculated by the required damping force calculation means, for example, based on the nonlinear H∞ control theory. Force can be set, and the damping force changing means can change to the set required damping force. Thereby, the damping force generation means can effectively suppress the deterioration of the riding comfort caused by the response delay or the like by generating the set predetermined damping force.

  Hereinafter, a vehicle suspension apparatus according to an embodiment of the present invention will be described in detail with reference to the drawings. FIG. 1 is an overall schematic diagram of the suspension device of the present embodiment. This suspension device absorbs and attenuates vertical vibrations of the vehicle, and includes a suspension mechanism 10 and an electric control device 20 that controls the operation of the suspension mechanism 10.

  The suspension mechanism 10 includes a suspension spring 11 and a shock absorber 12. One end (lower end) of the suspension spring 11 and the shock absorber 12 is connected to the unsprung member LA, and the other end (upper end) is connected to the sprung member HA. The suspension spring 11 absorbs vibration transmitted from the road surface via the tire W and the unsprung member LA. For example, a metal coil spring or an air spring is employed. The shock absorber 12 is arranged in parallel with the suspension spring 11 and attenuates the vibration. A knuckle connected to the wheel, a lower arm having one end connected to the knuckle, or the like corresponds to the unsprung member LA. The sprung member HA is a member supported by the suspension spring 11 and the shock absorber 12, and the vehicle body is also included in the sprung member HA.

  As shown in FIGS. 1 and 2, the shock absorber 12 includes a cylinder 12a, a piston 12b, and a piston rod 12c. The cylinder 12a is a cylindrical member in which a viscous fluid (for example, oil or the like) is sealed, and its lower end is connected to a lower arm that is an unsprung member LA. The piston 12b is disposed in the cylinder 12a and is configured to be movable in the axial direction in the internal space of the cylinder 12a. Thereby, the piston 12b divides the internal space of the cylinder 12a into an upper space 12a1 and a lower space 12a2. In addition, a communication passage 12b1 is formed in the piston 12b. The communication path 12b1 opens to an upper surface 12b2 facing the upper space 12a1 and a lower surface 12b3 facing the lower space 12a2, and connects the upper space 12a1 and the lower space 12a2. The piston rod 12c is a rod-shaped member, one end of which is connected to the piston 12b, and the other end is connected to a vehicle body that is a sprung member HA.

  In the shock absorber 12 configured as described above, when the unsprung member LA vibrates up and down due to road surface unevenness during traveling of the vehicle, this up-and-down vibration is transmitted from the unsprung member LA to the cylinder 12a of the shock absorber 12. The cylinder 12a also vibrates up and down. At this time, since the piston 12b is disposed in the cylinder 12a, the piston 12b is relatively displaced in the vertical direction by the vertical vibration of the cylinder 12a. And according to this relative displacement, viscous fluid generate | occur | produces when a viscous fluid distribute | circulates in the communicating path 12b1, This viscous resistance becomes a damping force with respect to a vertical vibration, and a vibration attenuate | damps.

  The suspension mechanism 10 includes a variable aperture mechanism 13. As shown in FIG. 2, the variable throttle mechanism 13 includes a rotary valve 13a, an actuator 13b, and a control rod 13c. The rotary valve 13a is provided in the communication passage 12b1 formed in the piston 12b, and changes the size of the flow passage cross-sectional area of at least a part of the communication passage 12b1, that is, the valve opening OP. The actuator 13b has, for example, a servo motor as a main component, and operates the rotary valve 13a via the control rod 13c. The control rod 13c is inserted into the piston rod 12c and connects the rotary valve 13a and the actuator 13b.

  With this configuration, in the variable throttle mechanism 13, the actuator 13 b operates stepwise (intermittently) to rotate the control rod 13 c by a predetermined rotation angle, so that the valve opening OP of the rotary valve 13 a has a plurality of stages (for example, , 9). As described above, the valve opening OP of the rotary valve 13a is changed stepwise, whereby the flow passage cross-sectional area of the communication passage 12b1 is changed stepwise, and as a result, the viscous fluid flows through the communication passage 12b1. Sometimes the resistance is also changed. Therefore, if the valve opening OP of the rotary valve 13a is changed stepwise, the damping coefficient indicating the magnitude of the damping force of the shock absorber 12 is also changed stepwise.

  As shown in FIG. 1, the electric control device 20 includes a suspension electronic control unit 21 (hereinafter simply referred to as a suspension ECU 21) for controlling the operation of the actuator 13b of the variable aperture mechanism 13. The suspension ECU 21 is a microcomputer whose main components are a CPU, a ROM, a RAM, and the like. Then, the suspension ECU 21 executes various programs including a damping force change control program, which will be described later, to control the driving of the actuator 13b and to generate a damping force F (more specifically, a damping coefficient) generated by the shock absorber 12. Change as appropriate. For this reason, as shown in FIG. 1, an unsprung acceleration sensor 22, an unsprung acceleration sensor 23, a stroke sensor 24, and a tire displacement sensor 25 are connected to the suspension ECU 21.

The sprung acceleration sensor 22 is assembled to the vehicle body as the sprung member HA, and detects the acceleration in the vertical direction of the sprung member HA with respect to the absolute space. Then, the sprung acceleration sensor 22 outputs a sprung acceleration x _pb ″ to the suspension ECU 21 as a signal corresponding to the detected vertical acceleration of the sprung member HA. The unsprung acceleration sensor 23 is assembled to a lower arm or the like as the unsprung member LA, and detects the acceleration in the vertical direction of the unsprung member LA with respect to the absolute space. Then, the unsprung acceleration sensor 23 outputs unsprung acceleration x _pw ″ to the suspension ECU 21 as a signal corresponding to the detected vertical acceleration of the unsprung member LA.

The stroke sensor 24 is disposed between the sprung member HA and the unsprung member LA, and detects a relative displacement amount between the sprung member HA and the unsprung member LA. Then, the stroke sensor 24 outputs, as signals according to the detected relative displacement amount, the vertical displacement amount x _pb from the reference position of the sprung member HA and the vertical displacement amount x _pw from the reference position of the unsprung member LA. The unsprung-unsprung relative displacement amount x _pw −x _pb that is the difference between the two is output to the suspension ECU 21. The tire displacement amount sensor 25 detects the displacement amount of the tire W. Then, the tire displacement amount sensor 25, the difference between the vertical displacement amount x _pw from the reference position of the displacement x _pr and the unsprung member LA from the reference position of the road surface as a signal corresponding to the displacement amount of the detected tire The unsprung relative displacement amount x _pr −x _pw is output to the suspension ECU 21.

  Here, the sprung acceleration sensor 22 and the unsprung acceleration sensor 23 detect the acceleration upward of the vehicle as a positive acceleration, and detect the acceleration downward of the vehicle as a negative acceleration. The stroke sensor 24 sets the displacement of the sprung member HA from the reference position upward to the vehicle to a positive displacement and the displacement to the vehicle downward to a negative displacement. The relative displacement is detected with the displacement as a positive displacement and the downward displacement of the vehicle as a negative displacement. Further, the tire displacement sensor 25 sets the displacement of the unsprung member LA from the reference position to the upper side of the vehicle as a positive displacement, the displacement toward the lower side of the vehicle as a negative displacement, and the displacement from the reference position of the road surface to the upper side. The relative displacement is detected with a positive displacement and a downward displacement as a negative displacement.

  Furthermore, as shown in FIG. 1, a drive circuit 26 for controlling the operation of the actuator 13b is connected to the suspension ECU 21. With this configuration, the suspension ECU 21 controls the drive circuit 26 to operate the actuator 13b to change the damping force F (damping coefficient), that is, the damping force characteristic of the shock absorber 12.

  Next, the operation of the suspension device configured as described above will be described. When the driver turns on an ignition switch (not shown), the suspension ECU 21 repeatedly executes the damping force change control program shown in FIG. 3 every predetermined short time. Hereinafter, the damping force change control program will be described in detail.

The execution of this damping force change control program is started in step S10. In the subsequent step S11, the suspension ECU 21 detects the detected values from the sprung acceleration sensor 22, the unsprung acceleration sensor 23, the stroke sensor 24, and the tire displacement sensor 25, specifically, the sprung acceleration x _pb ″. , _Unsprung acceleration x _pw ″, unsprung-unsprung relative displacement x _pw −x _pb and unsprung relative displacement x _pr −x _pw are input.

Subsequently, in step S12, the suspension ECU 21 integrates the sprung acceleration x _pb ″ and the unsprung acceleration x _pw ″ input in step S11 with respect to time, at the vertical speed of the sprung member HA. A certain sprung speed x _pb ′ and an unsprung speed x _pw ′ that is the speed in the vertical direction of the unsprung member LA are calculated. Further, the suspension ECU 21 uses the calculated sprung speed x _pb ′ and unsprung speed x _pw ′, and the sprung-unsprung relative speed V (= x _pw ′ −x _pb ′), which is the difference between these speeds. Calculate The sprung-unsprung relative velocity V can be calculated by differentiating the sprung-unsprung relative displacement amount x _pw -x _pb input in step S11 with time. When the suspension ECU 21 calculates the sprung speed x _pb ′, the unsprung speed x _pw ′, and the sprung-unsprung relative speed V, the process proceeds to step S13.

Here, the sprung speed x _pb ′ and the unsprung speed x _pw ′ are calculated as a positive speed when traveling upward and as a negative speed when traveling downward. Further, the sprung-unsprung relative speed V is calculated with a speed toward the direction in which the distance between the sprung member HA and the unsprung member LA is narrowed as a positive speed, and a speed toward the direction in which the distance is widened is a negative speed. Is calculated as

In step S13, the suspension ECU21 calculates the required damping coefficient C _H to determine the damping force F shock absorber 12 to be generated. In the present embodiment, the suspension ECU 21 calculates the required damping coefficient C _H using a well-known nonlinear H∞ control theory. Hereinafter, calculation of the required damping coefficient C _H using this nonlinear H∞ control theory will be briefly described.

The required reduction coefficient C _H calculated in this embodiment includes a variable attenuation coefficient C _V that is a variable part (nonlinear part) of an attenuation coefficient and a fixed attenuation coefficient C _S that is a fixed part (linear part) of the attenuation coefficient. Represented by the sum. Here, the variable attenuation coefficient C _V is calculated using the detection values obtained by the sensors 22 to 25 input in step S11 and the values calculated in step S12. The fixed damping coefficient C _S is determined in advance according to the specifications of the shock absorber 12. For example, the fixed damping coefficient C _S is about the middle value between the maximum value and the minimum value of the damping coefficient that can be realized by the shock absorber 12 and the variable throttle mechanism 30. The attenuation coefficient can be set.

In the calculation of the variable damping coefficient C _V , generally, a single-wheel model of a vehicle as shown in FIG. 4 is assumed, and the vertical displacement amount from the reference position of the sprung member HA is x _pb and the unsprung member. If the vertical displacement amount from the reference position of LA is x _pw and the vertical displacement amount from the reference position of the road surface is x _pr , the equations of motion of the sprung member HA and the unsprung member LA in the single wheel model of the wheel are as follows: It can be represented by (1) and formula (2).

However, the formula (1), in the formula (2), M _b represents the mass of the sprung member HA, M _w represents the weight of the unsprung member LA, K _s represents the spring constant of the suspension spring 11, K _t represents the spring constant of the tire W. In addition, the vertical displacement amount x _pb , the vertical displacement amount x _pw and the vertical displacement amount x _pr in the above formulas (1) and (2) are defined as positive displacements from the respective reference positions, and downward. The displacement of is a negative displacement.

ただし、前記（３）中におけるx_p，A_p，B_p1，B_p2(x_p)は下記式（４）により表される。

Further, when the control input u is the variable damping coefficient C _V , the disturbance w ₁ is the road surface vertical displacement speed x _pr ′, and this model is expressed in the state space, it is expressed by the following equation (3).

However, x _p , A _p , B _p1 , B _p2 (x _p ) in the above (3) are represented by the following formula (4).

Here, the objectives of improving the characteristics of the suspension device are the sprung speed x _pb ′ that greatly affects the vibration of the sprung member HA, the sprung acceleration x _pb ″ that greatly affects the riding comfort of the vehicle, and the unsprung member LA. Is to simultaneously suppress the sprung-unsprung relative velocity V that greatly affects the vibration of the spring. Therefore, the sprung speed x _pb ′, the sprung acceleration x _pb ″, and the sprung-unsprung relative speed V are used as the evaluation output z _p . Further, in the suspension system, on acceleration x _pb 'spring' and sprung - for unsprung relative displacement amount x _pw -x _pb is relatively detected easily, it sprung acceleration x _pb '' and the sprung - unsprung relative The displacement x _pw −x _pb is taken as the observation output y _p . Note that this observation output y _p includes observation noise w ₂ . And when these are expressed in a state space, they can be expressed by the following equations (5) and (6).

However, z _p in formula (5) and _{(6), y p, C} p1, D p12 (x p), C p2, D p21, D p22 (x p) , respectively, the following equation (7) It is represented by

Here, in the third term on the right side of the equation (3) representing the state space representation of the model, the coefficient B _p2 (x _p ) includes the state quantity x _p, and the control input u is included in this B _p2 (x _p ). Has been multiplied. Thus, the system is bi-linear system, the control input u is uncontrollable without effect near the origin of the state quantity x _p. In order to solve this problem, a non-linear H∞ state feedback control system using a non-linear or weight function is designed.

Now, in order to design a nonlinear H∞ state feedback control system, a generalized plant of a nonlinear H∞ state feedback control system as shown in FIG. 5 in which frequency weights are added to the evaluation output z _p and the control input u is assumed. Here, the frequency weight is a weight whose weight changes according to the frequency, and is a dynamic weight given by a transfer function. By using this frequency weight, it is possible to increase the weight of the frequency band for which the control performance is desired to be increased and reduce the weight for the frequency band where the control performance can be ignored.

In the generalized plant shown in FIG. 5, the evaluation output z _p and the control input u are multiplied by frequency weights W _s (s) and W _u (s), respectively, and further expressed by the following equation (8). Are multiplied by nonlinear weight functions a ₁ (x) and a ₂ (x) for the state quantity x satisfying the following condition.

Here, state space representations for the frequency weight W _s (s), the frequency weight W state quantity x _w of _s (s), the output z _w and the constant matrix A _w of the frequency weight _{W s (s), B w} , C _w, the D _w, it is represented by the following formula (9). The state space representation for the frequency weight W _u (s) is the state amount x _u of the frequency weight W _u (s), the output z _u and the constant matrix A _u of the frequency weight _{W u (s), B u} , C _{By u 1} and D _u , it is expressed by the following formula (10).

ただし、前記式（１１）において、状態量x、各係数行列A，B₁，B₂(x)，C₁₁，D₁₂₁(x)，C₁₂，D₁₂₂(x)は、下記式（１２）に示すように表現される。

And by using said Formula (9) and Formula (10), the state space expression represented by said Formula (3) is represented like the following formula (11).

However, in the above equation (11), the state quantity x, each coefficient matrix A, B ₁ , B ₂ (x), C ₁₁ , D ₁₂₁ (x), C ₁₂ , D ₁₂₂ (x) are expressed by the following equation (12 ).

Moreover, the state space expression represented by the said Formula (11) is represented like the following formula (14) by considering the conditions represented by the following formula (13).

Here, there is a coefficient matrix D ₁₂₂ ⁻¹ , there is a positive definite symmetric solution P that satisfies the Riccati equation expressed by the following equation (15) for a predetermined positive constant γ, and the weight function a ₁ (x ), A ₂ (x) satisfy the constraint condition of the following equation (16), the control input u (= k) where the closed-loop system is internally stable and the L2 gain representing robustness against disturbance is a positive constant γ or less. (x)) is expressed by the following equation (17).

ただし、前記式（１８）および式（１９）中のm₁(x)は、任意の正定関数である。 When weighting functions a ₁ (x) and a ₂ (x) satisfying the equation (16) are expressed as in the following equation (18), the control input u = k (x) is represented by the following equation (19).

However, m ₁ (x) in the equations (18) and (19) is an arbitrary positive definite function.

When the control input u, that is, the variable damping coefficient C _V is calculated based on the equation (19) derived as described above, the suspension ECU 21 adds the preset fixed damping coefficient C _S to the required damping coefficient. Calculate C _H. Here, when the damping force F of the shock absorber 12 is controlled using the required damping coefficient C _H calculated based on the nonlinear H∞ control theory, the sprung-unsprung relative speed as shown in FIG. An FV characteristic representing a change in damping force F with respect to a change in V is obtained. More specifically, the obtained FV characteristic is represented by a smooth Lissajous curve drawn in the first and third quadrants with the origin at the center. In this FV characteristic, when the absolute value of the sprung-unsprung relative speed V is large, the damping force F as a control target is increased and the control range is widened. When the absolute value is small, the damping force F changes linearly with respect to the relative speed V, and the control width becomes extremely small.

By the way, as described above, when the absolute value of the sprung-unsprung relative speed V is large, in other words, when the sprung member HA is vibrating at a relatively high frequency, the control range of the damping force F is widened. In this case, it is desirable to generate a large damping force F the shock absorber 12 by driving the actuator 13b to follow the vibration is determined by the required damping coefficient C _H.

However, in reality, due like calculation cycle response and required damping coefficient C _H of the actuator 13b, in particular, on the spring - the absolute value of the lower relative speed V spring is large (i.e., sprung member HA is In a situation where the vibration is vibrated at a relatively high frequency), the damping force change control that generates a large damping force F on the shock absorber 12 may not follow well. As described above, in a situation where the damping force change control cannot be satisfactorily followed, since a large damping force F for suppressing vibration (resonance) of the sprung member HA is not generated at an appropriate timing, the riding comfort is deteriorated. It may appear prominently. Therefore, the suspension ECU21, when calculating the required damping coefficient C _H at the step S13, the process proceeds to step S14.

  In step S14, the suspension ECU 21 does not follow the suppression of the vibration generated in the sprung member HA by the damping force change control. In other words, the suspension ECU 21 changes the damping force F of the shock absorber 12 to the sprung member HA. It is determined whether the generated vibration (resonance) is not suppressed. This will be specifically described below.

Now, assuming that the suppression (decrease) of vibration (resonance) generated in the sprung member HA is the target of ride comfort control, it is assumed that a good ride comfort is as shown in PSD (Power Spectrum Density) shown in FIG. This is a case where the sprung acceleration x _pb ″ of the sprung member HA decreases with the damping force change control. In other words, the magnitude of the sprung acceleration x _pb ″ is smaller than the magnitude of the unsprung acceleration x _pw ″ of the unsprung member LA, that is, the amplitude of the sprung acceleration x _pb ″ is reduced. In this case, it can be said that the damping force change control follows the suppression of the vibration generated in the sprung member HA.

Based on this, the suspension ECU 21 receives a signal representing the sprung acceleration x _pb ″ and a signal representing the unsprung acceleration x _pw ″ input from the sprung acceleration sensor 22 and the unsprung acceleration sensor 23 in step S11. It is used to determine whether the damping force change control is not able to follow the suppression of vibration generated in the sprung member HA. Specifically, the suspension ECU 21 first generates a signal representing the sprung acceleration x _pb ″ input from the sprung acceleration sensor 22 and a signal representing the unsprung acceleration x _pw ″ input from the unsprung acceleration sensor 23. A known band pass filter process (BPF process) is performed. In this bandpass filter processing, processing is performed by setting a band (band) on a relatively high frequency side (for example, about 5 to 6 Hz) in which deterioration of ride comfort is noticeable due to failure to follow control.

Next, the suspension ECU 21 represents a signal waveform representing the sprung acceleration x _pb ″ passed through the band-pass filter process (hereinafter, this passed signal waveform is referred to as a sprung signal waveform) and an unsprung acceleration x _pw ″. The maximum signal amplitude (that is, peak value) of each of the signal waveforms (hereinafter, this passed signal waveform is referred to as an unsprung signal waveform) is compared. That is, as shown in FIG. 8, the suspension ECU 21 acquires the maximum signal amplitude b of the sprung signal waveform and the maximum signal amplitude a of the unsprung signal waveform.

In the suspension ECU 21, the ratio b / a of the maximum signal amplitude b of the sprung signal waveform to the maximum signal amplitude a of the unsprung signal waveform is reduced, and the amplitude of the sprung acceleration x _pb ″ is reduced. In order to determine this, it is determined whether or not it is larger than a preset performance target index α. Here, the performance target index α is set to a value of “1” or less. That is, when the value of the ratio b / a is larger than the performance target index α, the suspension ECU 21 determines “Yes” because the amplitude of the sprung acceleration x _pb ″ is not reduced, and the process _proceeds to step S15. move on. On the other hand, when the value of the ratio b / a is equal to or less than the performance target index α, the amplitude of the sprung acceleration x _pb ″ is reduced, so that it is determined as “No” and the process proceeds to step S16.

In step S15, the suspension ECU 21 sets the required damping force Freq when the absolute value of the sprung-unsprung relative speed V is equal to or higher than a predetermined reference speed Vo to a predetermined constant damping force Fconst. To do. That is, the situation in which step S15 is executed is a situation where the amplitude of the sprung acceleration x _pb ″ is not reduced, and the absolute value of the sprung-unsprung relative speed V is equal to or higher than the reference speed Vo. In other words, the damping force change control cannot follow, and the shock absorber 12 cannot generate a large damping force F that suppresses vibration (resonance) of the sprung member HA at an appropriate timing. For this reason, as shown by a thick line in FIG. 6, the suspension ECU 21 can ensure the response of the actuator 13b and the ride quality is remarkably deteriorated when the absolute value of the sprung-unsprung relative speed V is equal to or higher than the reference speed Vo. A constant value Fconst, which is experimentally set to be small enough to be suppressed, is set as the required damping force Freq. As a result, when the amplitude of the sprung acceleration x _pb ″ is not reduced and the absolute value of the sprung-unsprung relative speed V is equal to or higher than the reference speed Vo, the shock absorber 12 is The damping force F closest to the constant value Fconst can be generated with good control followability.

In the process of step S15, even when the amplitude of the sprung acceleration x _pb ″ is not reduced, when the absolute value of the sprung-unsprung relative speed V is less than the reference speed Vo, suspension ECU21, the step S13 is calculated by the required damping coefficient C _H and sprung - calculating the required damping force Freq by multiplying the unsprung relative speed V. Then, the suspension ECU 21 proceeds to step S17.

On the other hand, in step S16, the suspension ECU21 computes the required damping force Freq with required damping coefficient C _H calculated at step S13. That is, the situation in which step S16 is executed is a situation where the amplitude of the sprung acceleration x _pb ″ is reduced, in other words, the damping force change control follows well, and the shock absorber 12 is appropriately This is a situation in which a damping force F that suppresses vibration (resonance) of the sprung member HA can be generated at an appropriate timing. Thus, the step S13 is calculated by the required damping coefficient C _H and sprung - calculating the required damping force Freq by multiplying the unsprung relative speed V. Then, the suspension ECU 21 proceeds to step S17.

  In step S17, for example, the suspension ECU 21 refers to a damping force characteristic table stored in advance in a ROM (not shown) to determine the flow path cross-sectional area of the communication path 12b1 formed in the piston 12b of the shock absorber 12. The valve opening OP to be changed is determined. Here, in this damping force characteristic table, the damping force of the shock absorber 12 with respect to the sprung-unsprung relative speed V using the number of stages of the valve opening OP that can be set by the shock absorber 12 and the variable throttle mechanism 13 as parameters. F data is stored. In addition, when this stored data is shown as the FV characteristic, it becomes a plurality of curves shown by middle thick solid lines in FIG. Then, the suspension ECU 21 selects the valve opening OP corresponding to the damping force F closest to the required damping force Freq calculated in step S15 or step S16 from among the damping forces stored in the damping force characteristic table. The number of stages is selected and determined, and the process proceeds to step S18.

  In step S18, the suspension ECU 21 outputs a signal corresponding to the number of stages of the valve opening OP determined in step S17 to the actuator 13b via the drive circuit 26. Then, the actuator 13b rotationally displaces the control rod 13c based on the output signal, and operates the rotary valve 13a so as to correspond to the determined number of stages of the valve opening OP. Thereby, the shock absorber 12 generates the damping force F closest to the required damping force Freq.

  As described above, when the operation of the actuator 13b is controlled, the suspension ECU 21 proceeds to step S19 and temporarily ends the execution of the damping force control program. Then, after a predetermined short period of time has elapsed, program execution is started again in step S10.

As can be understood from the above description, according to the present embodiment, the suspension ECU 21 has the maximum signal amplitude b of the sprung signal waveform with respect to the maximum signal amplitude a of the unsprung signal waveform subjected to the band-pass filter processing in a relatively high frequency band. By determining whether or not the ratio b / a is greater than the performance target index α set in advance to a value of “1” or less in order to determine that the vibration of the sprung member HA is damped. It is possible to determine whether the amplitude of the sprung acceleration x _pb ″ is reduced, that is, whether the damping force change control is following. Thereby, the deterioration of the control followability with respect to the vibration in the high frequency region of the sprung member HA can be determined more accurately.

The suspension ECU 21 determines that the required damping force Freq is a certain required damping force Fconst, more specifically, when the sprung-unsprung relative speed V is equal to or higher than the reference relative speed Vo in a state where the control followability is deteriorated. it can be set to a constant required damping force Fconst be smaller than the required damping force Freq is calculated using the required damping coefficient C _H that is calculated based on the nonlinear H∞ control theory at the step S13. As a result, the shock absorber 12 can effectively suppress the deterioration of the riding comfort due to the response delay or the like by generating the set required damping force Fconst.

  In carrying out the present invention, the present invention is not limited to the above embodiment, and various modifications can be made without departing from the object of the present invention.

  For example, in the above embodiment, the required damping coefficient Creq is calculated based on the nonlinear H∞ control theory, and the required damping force Freq is calculated. However, even when the required damping coefficient Creq and the required damping force Freq are calculated according to other well-known calculation methods (for example, a calculation method based on the Skyhook control theory), the absolute value of the sprung-unsprung relative velocity V As the value increases, the above-described followability may deteriorate. Therefore, the required damping coefficient Creq is calculated by a calculation method other than the calculation method based on the non-linear H∞ control theory, and the required damping force Freq of the shock absorber 12 is calculated based on the required damping coefficient Creq, and the damping force change control is performed. Even in this case, as in the above embodiment, it is possible to determine the control follow-up state and, if not following, set the required damping force Freq to a preset constant value Fconst and be able to implement it. Needless to say.

1 is an overall schematic diagram of a vehicle suspension apparatus according to an embodiment of the present invention. It is a figure which shows concretely the structure of the shock absorber of FIG. 1, and a variable aperture mechanism. It is a flowchart of the damping force change control program executed by the suspension ECU of FIG. It is a figure which shows the single-wheel model of the vehicle assumed in order to calculate a request | requirement damping coefficient based on nonlinear Hinfinity control theory. It is a block diagram which shows the generalized plant of a nonlinear Hinfinity state feedback control system. It is a figure which shows the FV characteristic at the time of calculating | requiring a request | requirement damping coefficient based on the nonlinear Hinfinity control theory. It is a graph which shows the relationship between a frequency and sprung acceleration PSD. It is a figure for demonstrating the maximum signal amplitude of a sprung acceleration and unsprung acceleration.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 ... Suspension mechanism, 11 ... Suspension spring, 12 ... Shock absorber, 12a ... Cylinder, 12b ... Piston, 12b1 ... Communication path, 12c ... Piston rod, 13 ... Variable throttle mechanism, 13a ... Rotary valve, 13b ... Actuator, 20 ... Electric control device 21 ... Suspension ECU 22 ... Spring acceleration sensor 23 ... Unsprung acceleration sensor 24 ... Stroke sensor 25 ... Tire displacement sensor 26 ... Drive circuit HA HA ... Spring member LA ... Unsprung Member, OP ... Valve opening

Claims (4)

A damping force generating means that is disposed between an unsprung member and a sprung member of the vehicle and generates a damping force that attenuates vibration of the sprung member with respect to the unsprung member, and the damping force generating means is generated. In a vehicle suspension apparatus comprising a damping force change control means for changing and controlling the damping force, the damping force changing means includes:
Unsprung acceleration detecting means for detecting vertical acceleration of the unsprung member;
Sprung acceleration detecting means for detecting the vertical acceleration of the sprung member;
A sprung-unsprung relative speed detecting means for detecting a sprung-unsprung relative speed representing a relative difference between a vertical speed of the unsprung member and a vertical speed of the sprung member;
A required damping force calculating means for calculating a required damping force required for damping the vibration of the sprung member with respect to the unsprung member;
The value of the ratio of the magnitude of the acceleration of the sprung member detected by the sprung acceleration detecting means to the magnitude of the acceleration of the unsprung member detected by the unsprung acceleration detecting means is the damping force generating means. Comparison determination means for determining whether or not the vibration of the sprung member is attenuated by the damping force generated by
The absolute value of the sprung-unsprung relative speed detected by the sprung-unsprung relative speed detecting means when the comparison determining unit determines that the value of the ratio is larger than the predetermined value. A request to set a required damping force required to attenuate the vibration of the sprung member against the unsprung member to a predetermined constant value when the value is equal to or higher than a preset reference relative speed. Damping force setting means;
The damping force generated by the damping force generating means is the required damping force calculated by the required damping force calculating means or the required damping force set to a predetermined constant value by the required damping force setting means. A suspension device for a vehicle, comprising: damping force changing means for changing. In the vehicle suspension device according to claim 1,
The required damping force calculation means is
Based on the non-linear H∞ control theory, a variable damping coefficient that is a variable part of the damping coefficient of the damping force generating means is calculated, and a fixed damping coefficient that is a fixed part of the damping coefficient set in advance to the calculated variable damping coefficient And a required attenuation coefficient calculating means for calculating the required attenuation coefficient by adding
Multiplying the required damping coefficient calculated by the required damping coefficient calculating means by the sprung-unsprung relative speed detecting means detected by the sprung-unsprung relative speed detecting means to calculate the required damping force; A vehicle suspension apparatus characterized by the above. In the vehicle suspension device according to claim 1,
The comparison judgment means
The maximum unsprung acceleration for detecting the maximum amplitude of the unsprung member acceleration after bandpass filtering the signal representing the unsprung member detected by the unsprung acceleration detecting means with a predetermined frequency band. Detection means;
Maximum sprung acceleration amplitude for detecting the maximum amplitude of the sprung member acceleration after bandpass filtering the signal representing the sprung member acceleration detected by the sprung acceleration detecting means with a predetermined frequency band Detection means,
The ratio of the maximum amplitude of the acceleration of the sprung member detected by the sprung acceleration maximum amplitude detection unit to the maximum amplitude of the acceleration of the unsprung member detected by the unsprung acceleration maximum amplitude detection unit is: It is determined whether it is larger than the predetermined value set beforehand, The suspension apparatus of the vehicle characterized by the above-mentioned.   The vehicle suspension apparatus according to claim 1, wherein the predetermined value used by the comparison determination unit is preset to a value equal to or less than “1”.

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