JPH07117443A - Suspension control device - Google Patents

Abstract

PURPOSE:To perform good damping force control to correspond to change in the sprung resonance frequency resulting from change in the riding position of an occupant or the loading position of a cargo, etc. CONSTITUTION:When a vehicle is stopped the body of the vehicle is kept horizontal and load detected values WFL-WRR detected by load sensors are read (step S34), and the sprung mass m2FL-m2RR of each wheel is calculated according to the loads (step S35), and the sprung resonance frequency fUFL-fURR of each wheel is calculated according to the mass, and a control gain calculation map that represents the relationship between the frequencies of normal acceleration corresponding to the resonance frequencies and control gains is selected (step S37). The control gain calculation map selected when the vehicle is running is consulted to set a control gain to perform optimum damping force control, thereby restraining change in the position of the vehicle body.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a suspension control device for performing so-called semi-active control in which the control characteristics of a suspension are controlled based on the displacement speed of the vehicle body and the relative speed between the vehicle body and the wheels.

[0002]

2. Description of the Related Art As a conventional suspension control device, for example, Japanese Patent Application Laid-Open No. 61-163 previously proposed by the present applicant.
There is one disclosed in Japanese Patent No. 011. In this conventional example, a damping force variable shock absorber capable of switching a damping force into at least two levels of high and low by inputting a control signal, sprung speed measuring means for measuring sprung speed, and relative speed between sprung and unsprung. Code determination for determining whether or not the relative speed measuring means for measuring, the sign of the sprung speed measured by the sprung speed measuring means, and the sign of the relative speed measured by the relative speed measuring means match. Means and the sign of the sprung speed and the relative speed based on the judgment result of the sign judging means, the damping force variable shock absorber is set to a low damping force, and the signs of the sprung speed and the relative speed match. In some cases, the damping force variable shock absorber has a configuration including control signal output means for outputting a control signal for controlling the damping force variable shock absorber to a high damping force. The damping force of Soba appropriately controlled so that improve the steering stability and ride comfort.

[0003]

However, in the above-mentioned conventional suspension control device, the riding comfort and the steering stability can be improved by appropriately controlling the damping force of the variable damping force shock absorber. However, as a control mode, a semi-active control similar to the skyhook control is performed by simply controlling the damping force in two steps, high and low, according to the sign of the sprung velocity and the relative velocity. The control gain is controlled to a constant value without considering the frequency characteristic.

For this reason, when the control gain is set to a large value as shown in FIG. 15, the sprung resonance frequency region (about ₁ to 2 Hz) and the unsprung resonance frequency region (10) as shown by a curve L ₁ shown by a solid line. At about 12 Hz), the sprung power spectrum density can be suppressed to a low level to improve the vibration damping property, but the sprung mass can be spun even in the intermediate frequency range (4 to 8 Hz) between the sprung resonance frequency range and the unsprung resonance frequency range. power spectral density ride comfort during this period becomes higher is lowered, by setting smaller the control gain Conversely, as shown by the curve L ₂ of dashed lines shown, to reduce the sprung power spectral density in the intermediate frequency range However, the sprung mass power spectral density becomes high in the required sprung mass resonance frequency range, and good vibration damping effect cannot be exhibited. There is an unsolved problem that it is not possible to perform effective damping force control.

Therefore, the present invention has been made by paying attention to the unsolved problem of the above-mentioned conventional example, and by giving the control gain a frequency characteristic, the optimum damping is achieved regardless of the sprung mass change of the vehicle. It is an object of the present invention to provide a suspension control device that can perform force control.

[0006]

In order to achieve the above object, a suspension control device according to the present invention is interposed between a vehicle body side member and a wheel side member as shown in the basic configuration diagram of FIG. A suspension device capable of changing a vehicle body posture change suppression characteristic according to an input control signal, a vertical acceleration detecting means for detecting a vertical acceleration of the vehicle body at the position of the suspension device, a vehicle body side member and a wheel Relative displacement detection means for detecting relative displacement between the side members, and suppression of a change in posture of the vehicle body based on the vehicle body vertical acceleration detection value of the vertical acceleration detection means, the relative displacement detection value of the relative displacement detection means, and the control gain. In a suspension control device including a control unit that forms and outputs the control signal, the sprung mass is detected by detecting the wheel load at each wheel position when the vehicle is stopped. Sprung mass detection means, sprung resonance frequency calculation means for calculating sprung resonance frequency at each wheel position based on the sprung mass detected by the sprung mass detection means, and calculated by the sprung frequency calculation means Refer to map selection means for selecting a corresponding control gain calculation map based on the sprung resonance frequency and the control gain calculation map selected based on the frequency of the vehicle body vertical acceleration detection value while the vehicle is traveling. And a control gain setting means for setting a control gain.

[0007]

In the present invention, the sprung mass detection means detects the sprung mass based on the wheel load at each wheel position while the vehicle is stopped, and the sprung mass resonance is detected based on the detected sprung mass. The sprung resonance frequency at each wheel position is calculated by the frequency calculation means, and the control gain calculation map showing the relationship between the control gain corresponding to the sprung resonance frequency calculated by the map selection means and the frequency of the vehicle body vertical acceleration is selected. Then, this map for calculating the control gain is made to correspond to the sprung mass depending on the occupant and the load, and when the vehicle is subsequently in the running state, based on the frequency of the vehicle body vertical acceleration detection value at each wheel position. The optimum control gain according to the frequency characteristic is set by sequentially setting the control gain with reference to the control gain calculation map.

[0008]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 2 is a schematic configuration diagram showing an embodiment of the present invention, in which the damping force variable shock absorber 3 constitutes a suspension device between each of the wheels 1FL to 1RR and the vehicle body 2.
FL to 3RR are arranged, and a step motor 41 for switching the damping force of these damping force variable shock absorbers 3FL to 3RR
FL to 41RR are controlled by a control signal from the controller 4 described later, and air suspensions 70FL to 70RR for adjusting the vehicle height are provided around the damping force variable shock absorbers 3FL to 3RR. The air pressure is controlled by the controller 4.

Variable damping force shock absorber 3FL to 3RR
As shown in FIGS. 3 to 7, the piston is configured as a twin-tube type gas-filled strut type having a cylinder tube 7 composed of an outer cylinder 5 and an inner cylinder 6, and the inside of the inner cylinder 6 is in sliding contact with the piston. 8 define upper and lower pressure chambers 9U and 9L. 4 to 7, the piston 8 includes a cylindrical lower half body 11 having a central opening 10 formed in the inner peripheral surface of a seal member 9 that is slidably contacted with the inner cylinder 6 on the outer peripheral surface. It is composed of an upper half body 12 fitted in the lower half body 11.

In the lower half body 11, an expansion side oil passage 13 is formed so as to vertically penetrate therethrough, and a sealing member 9 is provided downward from the upper surface side.
Hole 14a having a diameter larger than that of the expansion-side oil passage 13 and extending from the outer peripheral surface of the cylindrical body 11 to the hole 14a.
The pressure side oil flow passage 14 formed by a hole portion 14b that is formed by communicating with the bottom portion of a, the annular grooves 15U and 15L formed at the upper and lower open ends of the central opening 10, and formed on the upper surface side. A long groove 16 that communicates with the annular groove 15U and the expansion-side oil passage 13 is formed, and a long groove 17 that is formed on the lower surface side and that communicates with the annular groove 15L is formed. The long groove 17 is closed by the expansion side disk valve 18, and the upper end side of the compression side oil flow path 14 is closed by the compression side disk valve 19.

The upper half body 12 has a small-diameter shaft portion 21 fitted in the central opening 10 of the lower half body 11 and the shaft portion 2.
1 and a large diameter shaft portion 22 having a diameter smaller than the inner diameter of the inner cylinder 6 integrally formed at the upper end of the small diameter shaft portion 21 and the lower end surface of the small diameter shaft portion 21 at the center position. Hole 23a reaching from the side to the middle portion of the large-diameter shaft portion 22 and a hole portion 23b having a smaller diameter than the hole portion 23a communicating with the upper end side of the hole portion 23a.
And a through hole 23 composed of a hole portion 23c having a larger diameter than this and communicating with the upper end side of the hole portion 23b is formed. Pair of through holes 24 penetrating to the inner peripheral surface side in the direction
a, 24b and 25a, 25b, and an arc-shaped groove 2 communicating with the upper end side of the hole portion 23a of the large-diameter shaft portion 22.
6 is formed, and an L-shaped pressure-side oil passage 27 that connects the arc-shaped groove 26 and the lower end surface is formed.
Is blocked by.

The lower half body 11 and the upper half body 12 are located below the lower half body 11 of the small diameter shaft portion 21 with the small diameter shaft portion 21 fitted in the central opening 10 of the lower half body 11. The nut 29 is screwed into the lower end portion projecting to the end and tightened with the nut to be integrally connected. Further, a cylindrical valve body 31 having a closed upper end which constitutes a variable throttle is rotatably disposed in the hole 23a of the upper half body 12. As shown in FIG. 4, the valve body 31 has an upper half body 1
2 has a through hole 32 that reaches the inner peripheral surface in the radial direction at a position facing the arcuate groove 26 of the large-diameter shaft portion 22, and the small-diameter shaft of the upper half body 12 is formed as shown in FIGS. A communication groove 33 is formed on the outer peripheral surface of the portion 21 corresponding to the space between the through holes 24a and 25a. Further, as shown in FIG. 6, the through hole 24b of the small diameter shaft portion 21 of the upper half body 12 is formed.
And 25b, an elongated hole 34 extending in the axial direction is formed on the outer peripheral surface corresponding to between the inner peripheral surface 25b and the inner peripheral surface 25b.
The positional relationship among the through hole 32, the communication groove 33, and the elongated hole 34 is selected so as to obtain a damping force characteristic with respect to the rotation angle of the valve body 31 shown in FIG. 8, that is, the step angle of step motors 41FL to 41RR described later. ing.

That is, at the position A in FIG. 8, which is the maximum rotation angle position in the clockwise direction, for example, as shown in FIG. 4, only the through hole 32 communicates with the arcuate groove 26, and therefore the piston 8 descends. For pressure side movement, lower pressure chamber 9L
From the pressure side oil flow path 14 to the upper pressure chamber 9U through an orifice formed by the opening end and the pressure side disk valve 19, and the pressure side flow path C1 shown by a broken line and the lower pressure chamber 9L.
Through the inner peripheral surface of the valve body 31, the through hole 32, the arc-shaped groove 2
6. A pressure side flow path C2, which is shown by a broken line, is formed toward the upper pressure chamber 9U through the orifice formed by the opening end of the pressure side oil flow path 27 and the pressure side disk valve 28, and the piston 8 rises. With respect to the extension side movement, the broken line extending from the upper pressure chamber 9U to the lower pressure chamber 9L through the elongated groove 16 and the extension side flow path 13 and the orifice formed by the open end and the extension side disk valve 18. Only the extension side flow path T1 shown in the figure is formed, and a high damping force that rapidly increases as the piston speed increases is generated on the extension side, and a low damping force slightly increases on the compression side as the piston speed increases. Generates damping force.

By rotating the valve body 31 counterclockwise from the position A, the communication groove 33 of the valve body 31 and the through holes 24a and 25a of the small-diameter shaft portion 21 communicate with each other, as shown in FIG. Therefore, the opening areas of the communication groove 33 and the through holes 24a and 25a gradually increase as the turning angle increases. Therefore, for the extension side movement of the piston 8, as shown in FIG. 5A, the long groove 16 and the annular groove 15 are arranged in parallel with the flow path T1.
U, the through hole 24a, the communication groove 33, the through hole 25a, the annular groove 15L, the long groove 17, and the lower pressure chamber 9 through the orifice formed by the long groove 17 and the pressure side disk valve 18.
Since the flow path T2 toward L is formed, the maximum value of the damping force gradually increases as the opening area between the communication groove 33 and the through holes 24a and 25a of the small diameter shaft portion 21 increases, as shown in FIG. As shown in FIG. 5 (b), the minimum damping force state is maintained in order to maintain the state in which the flow paths C 1 and C 2 are formed, as shown in FIG. 5B.

Further, when the valve body 31 is rotated counterclockwise to the vicinity of the position B, as shown in FIG. 6, the through holes 24b and 25b of the valve body 31 are communicated by the elongated hole 34. Become. Therefore, for the extension side movement of the piston 8, as shown in FIG. 6A, the elongated groove 16, the annular groove 15U, the through hole 24a, and the elongated hole 3 are arranged in parallel with the flow paths T1 and T2.
4, the flow path T3 that goes toward the lower pressure chamber 9L through the hole 23a
Is formed, and the extension side damping force becomes the minimum damping force state, and with respect to the pressure side movement of the piston 8, in addition to the flow paths C1 and C2, the hole portion 23a, the long hole 34, and the through hole 2 are formed.
4b, a flow path C reaching the long groove 16 through the annular groove 15U
3 and the hole 23a, the long hole 34, the through hole 25b, the annular groove 15L, the through hole 25a, the communication groove 33, the through hole 24a, and the annular groove 15U to form the flow path C4 that reaches the long groove 16. , The minimum damping force state is maintained, as shown in FIG.

Further, when the valve body 31 is rotated in the counterclockwise direction, the opening area between the elongated hole 34 and the through holes 24b and 25b becomes smaller, and the elongated hole 34 and the through hole 24 at the rotation angle θ _B2.
As shown in FIG. 7, the area between b and 25b is blocked, but the opening area between the through hole 32 and the arcuate groove 26 gradually decreases from the rotation angle θ _B2 . Therefore, between the rotation angle θ _B2 and the maximum counter-clockwise rotation angle θ _C , the flow paths T1 and T2 coexist with respect to the extension side movement of the piston 8, so that the minimum damping force state is set. On the contrary, with respect to the pressure side movement of the piston 8, the maximum damping force is gradually increased by gradually decreasing the opening area between the through hole 32 and the arcuate groove 26, and the valve body 31 is positioned. When reaching C, as shown in FIG. 7, the through-hole 32 and the arcuate groove 26 are cut off from each other, so that the lower pressure chamber 9 is prevented from moving toward the pressure side of the piston.
The flow path from L to the upper pressure chamber 9U is only the flow path C1 and is in the compression side high damping force state.

On the other hand, a cylindrical piston rod 35 is fitted in the hole portion 23c of the upper half body 12, and the upper end of the piston rod 35 is projected above the cylinder tube 7 as shown in FIG. , Its upper end side is the vehicle body side member 36
Is fixed to the bracket 37 attached to the bracket 37 by a nut 39 via rubber bushes 38U and 38L, and the step motors 41FL to 41RR are mounted on the upper end of the piston rod 35 via a bracket 40 so that the rotary shafts 41a to 41RR of the step motors 41FL to 41RR are rotated.
Is fixed in a downwardly projecting relationship, and the rotary shaft 41a and the valve element 31 described above are connected by a connecting rod 42 that is loosely inserted in the piston rod 35. In addition, 43 is a bumper rubber. The lower end of the cylinder tube 7 is connected to a wheel side member (not shown).

Further, the air suspensions 70FL to 70RR
3, each of the damping force variable shock absorbers 3FL to 3RR is interposed between the upper end side of the outer cylinder 5 of the cylinder tube 7 and the bracket 37 attached to the vehicle body side member 36, and the piston Elastic bodies 72FL to 72RR made of rubber or the like that surround the rod 35 in a vertically expandable and contractible manner to form an air chamber 71 therein, and the elastic bodies 72FL to 72FL.
The air compressor 75 communicated with the 72RR through the air pipes 73FL to 73RR and the electromagnetic supply / exhaust valves 74FL to 74RR interposed in the middle thereof, and between the electromagnetic supply / exhaust valves 74FL to 74RR and the air compressor 75 of the air pipes 73FL to 73RR. The electromagnetic supply / exhaust valves 74FL-7L are configured by the controller 4 which will be described later.
4RR and electromagnetic exhaust valves 76FL to 76RR are both closed to maintain the set vehicle height. From this state, electromagnetic exhaust valves 76FL to 76RR are closed and electromagnetic supply / exhaust valves 74FL to 74RR are opened. As a result, the high pressure air from the air compressor 75 is supplied into the air chamber 71 to increase the vehicle height, and conversely, both the electromagnetic supply / exhaust valves 74FL to 74RR and the electromagnetic exhaust valves 76FL to 76RR are opened. The high-pressure air in the air chamber 71 is exhausted to lower the vehicle height.

The input side of the controller 4 is shown in FIG.
As shown in, the vertical acceleration detection values X _2FL ″ to X 2 which are analog voltages that are positive in the upward direction and negative in the downward direction according to the vertical acceleration provided on the vehicle body side corresponding to each wheel position.
Vertical acceleration sensors 51FL to 51RR as vertical acceleration detecting means for outputting _2RR ″, and an inductance corresponding to relative displacement between a vehicle body side member and a wheel side member, for example, built in a cover of each damping force variable shock absorber 3FL to 3RR. Relative displacement detection value X which is analog voltage due to change
_DFL (= X _2FL -X _1FL ) ~ X _DRR (= X _2RR-
Stroke sensors 52FL to 52RR as relative displacement detecting means for outputting X _1RR ), a vehicle speed sensor 53 for detecting a vehicle speed, and an inclination in the front-rear direction of the vehicle body. For example, a potentiometer-type tilt angle sensor 54X that outputs a front-rear tilt detection value θ _X that is a negative analog voltage and the left-right tilt of the vehicle body are detected, and becomes positive when falling left and becomes negative when falling right. A potentiometer-type tilt angle sensor 54Y is also used to output a left / right tilt detection value θ _Y that is an analog voltage.
And a load cell for outputting an analog voltage according to the detected load, which is arranged between the piston rod 35 of each of the damping force variable shock absorbers 3FL to 3RR and the bracket 37 attached to the vehicle body side member 36. The load sensors 55FL to 55RR are connected to the step motors 41FL to 41RR for controlling the damping force of the damping force variable shock absorbers 3FL to 3RR on the output side, and the air suspension 70.
Electromagnetic supply / exhaust valve 74FL ~ 7 for controlling air pressure of FL ~ 70RR
4RR and electromagnetic exhaust valves 76FL to 76RR are connected.

The controller 4 includes a microcomputer 56 having at least an input interface circuit 56a, an output interface circuit 56b, an arithmetic processing unit 56c and a storage unit 56d, and a vertical acceleration sensor 51.
FL-51RR vertical acceleration detection values X _2FL ″ _{-X 2RR} ″ are converted into digital values and supplied to the input interface circuit 56 a. A / D converters 57FL-57RR and stroke sensors 52FL-52RR relative displacement detection values X Input interface circuit 56a by converting _{DFL to} X _DRR into digital values
And the A / D converters 58FL to 58RR supplied to the input interface circuit 56a for converting the tilt angle detection values θ _X and θ _Y of the tilt angle sensors 54X and 54Y into digital values. 60Y and load sensor 5
A / A which converts the load detection values W _{FL to} W _RR of 5 _{FL to} 55 _RR into digital values and supplies them to the input interface circuit 56a
D converters 60FL to 60RR and output interface circuit 5
The step control signals for the step motors 41FL to 41RR output from 6b are input, and the step control signals are converted into step pulses to drive the step motors 41FL to 41RR and output from the output interface circuit 56b. Opening / closing control signal controls the air supply / exhaust valves 74FL of the air suspensions 70FL to 70RR.
.About.74RR and drive circuits 61FL to 61RR and 62FL to 62RR for driving the exhaust valves 76FL to 76RR.

Here, the operation of the microcomputer 56
The processing device 56c is configured to load the load sensor 55FL while the vehicle is stopped.
~ 55RR load detection value W_FL~ W_RRRead the
Based on the sprung mass m of each wheel 11FL-11RR position
_2FL~ M_2RRFor each wheel position based on these
Sprung resonance frequency f_UFL~ F_URRIs calculated in advance
Frequency of a plurality of vehicle body vertical accelerations stored in the storage device 56d
Control gain C for number_SThe control gain calculation map
Sprung resonance frequency f calculated from_UFL~ F_{UR R}Corresponding to
Control gain calculation map to be selected and the vehicle starts running.
When started, the control game is based on the frequency of the vertical acceleration of the vehicle body.
Control gain C with reference to the in calculation map_SAnd calculate
This control gain C_SAnd the vertical acceleration sensors 51FL to 51
Vertical acceleration detection value X of the vehicle body input from RR_2FL″ 〜X
_2RR”Integrated vehicle vertical speed X _2FL’～ X_2RR'When,
Wheels input from the stroke sensors 52FL to 52RR
And relative displacement detection value X between the vehicle body_DFL(= X_2FL-X_1FL)
~ X_DRR(= X_2RR-X_{1R R}) Differentiated relative velocity X
_DFL’～ X_DRRSkyhook control based on
To determine the damping force coefficient C for
Target step of corresponding step motor 41FL to 41RR
Angle θ_TAnd the target step angle θ_TAnd the current status
Up angle θ_PCalculate the difference value between and
The control amount is output to the motor drive circuits 59FL to 59RR.

Further, the storage device 56d is the processing unit 5
A program required for the arithmetic processing of 6c is stored in advance, necessary values and arithmetic results in the arithmetic processing process are sequentially stored, and a plurality of control gain calculation maps corresponding to a plurality of sprung resonance frequencies are stored in advance. Is stored. Here, in the control gain calculation map, as shown in FIG. 13, the frequency of the vehicle body vertical acceleration is plotted on the horizontal axis and the control gain C _{S is} plotted on the vertical axis.
And the control gain Cs in the sprung resonance frequency range (about 1 to 2 Hz) and the unsprung resonance frequency range (about 10 to 12 Hz).
Is increased, the characteristic curve L _C is set such that the control gain C _S is reduced in the intermediate frequency range (4 to 8 Hz) between them, and a plurality of control gain calculation maps that differ according to the sprung resonance frequency are stored. Has been done.

Next, the operation of the above embodiment will be described with reference to FIGS. 10 to 12 showing an example of the processing procedure of the arithmetic processing unit 56c of the microcomputer 56. That is, FIG.
Process is executed as a timer interrupt process every predetermined time (for example, 20 msec). First, in step S1, the vehicle speed sensor 53
Of the vehicle speed detection value V, and then proceeds to step S2 to determine whether the vehicle speed detection value V is zero, that is, whether the vehicle is stopped. If the vehicle is stopped, the processing proceeds to step S3 and the control gain After the calculation map selection process is executed, the timer interrupt process is terminated and the process returns to the predetermined main program. When the vehicle speed detection value V is not zero, that is, when the vehicle is traveling, the process proceeds to step S4 and the damping is performed. After executing the force control process, the timer interrupt process is terminated and the process returns to the predetermined main program.

The control gain calculation map selection process is shown in FIG.
As shown in FIG. 1, first, in step S31, the inclination angle sensor
Inclination angle detection value θ of 54X and 54Y_XAnd θ_YRead
Then, the process proceeds to step S32 and the vehicle body is in the horizontal state.
Or not. This judgment is based on each tilt angle detection value
θ_XAnd θ_YDepending on whether or not is_X= Θ
_YWhen = 0, it is determined that the vehicle body is horizontal
Then, the process directly proceeds to step S34, where θ_x≠ 0 and / or
θ_YWhen ≠ 0, the process proceeds to step S33.

In step S33, the inclination detection value θ _X
Based on the inclination detection value θ _Y , the vehicle height is adjusted so that they become zero, the vehicle body is set in the horizontal state, and then the process proceeds to step S34. That is, first, the inclination detection value θ _X
, The front and rear air suspensions 70FL,
Of the 70FR, 70RL, and 70RR, for example, the control signals CS _1FL, CS _1FR or CS _1RL, CS _1RR for opening the electromagnetic supply / discharge valves 74FL, 74FR or 74RL, 74RR on the side where the vehicle height is low are output, and then inclination detection is performed. The left and right air suspensions 70FL, 70RL and 70FR, 70 so that the value θ _Y becomes zero.
Of the RRs, for example, the electromagnetic supply / discharge valves 74FL and 74RL on the side where the vehicle height is low
Or control signal CS _1FL, C for opening 74FR, 74RR
Output S _1RL or CS _1FR, CS _1RR to set the car body in a horizontal position.

In step S34, the load sensors 55FL ...
The load detection values W _{FL to} W _RR of 55 _RR are read, then the process proceeds to step S35, and the sprung masses m _{2FL to} m _2RR at each wheel position are calculated from the read load detection values W _{FL to} W _RR , and then Go to step S36, and calculate the sprung mass m
_{Based on 2i} (i = FL, FR, RL, RR), the following formula (1) is calculated to calculate the sprung resonance frequency f _Ui at each wheel position.

F _Ui = (1 / 2π) (K ₂ / m _2i ) ^1/2 (1) Next, the process proceeds to step S37 and the calculated sprung mass resonance frequencies f _{UFL to} f _{UR R} are set. Based on this, a control gain calculation map having peaks at the corresponding sprung resonance frequencies f _{UFL to} f _URR is selected. In this way, by selecting the control gain calculation map for each wheel position, it is possible to select the control gain calculation map corresponding to the current passengers and the weight of the load.

In the processing of FIG. 11, step S3
The processing of 1 to S36 corresponds to the sprung resonance frequency detecting means, and the processing of step S37 corresponds to the control gain calculating map selecting means. In addition, the damping force control process is shown in FIG.
As shown in 2, first the vertical acceleration detection value X _2i in step S41 "read and then the process proceeds to step S42, reading Nde vertical acceleration detection value X _2i" or obtaining the number of zero-cross per unit of time, The vertical acceleration frequencies f _{FL to} f _RR are calculated by measuring the time from when the zero crossing state occurs until when the next zero crossing state occurs.

Next, in step S43, the relative displacement detection values X _Di are read, and then in step S44, the vertical acceleration detection value X read in step S41.
_2i ″ is integrated by, for example, low-pass filtering to calculate the vehicle body vertical velocity X _2i ′, and these are stored in the storage device 5.
It is temporarily stored in a predetermined storage area 6d, and then step S4.
5, the relative displacement detection value X _Di read in step S43 is differentiated by, for example, high-pass filtering to calculate the relative velocity X _Di ′, which are stored in the storage device 56d.
After being temporarily stored in the predetermined storage area, the process proceeds to step S46.

In this step S46, step S42
Based on the frequencies f _{FL to} f _RR of the vertical acceleration detected in 1., the control gain C _S is calculated with reference to the control gain calculation map selected in the control gain calculation map selection process described above. Then, the process proceeds to step S47, in which the following formula (2) is calculated based on the vehicle body vertical velocity X _2i ′ and relative velocity X _Di ′ calculated in steps S44 and S45 and the control gain C _S calculated in step S46. The calculation is performed to calculate the damping coefficient C for performing the skyhook control.

C = C _S · X _2i ′ / X _Di ′ (2) Then, the process proceeds to step S48 and the above step S4 is performed.
It is determined whether or not the damping coefficient C calculated in step 7 is less than or equal to the minimum damping force C _{MIN in the} preset damping force variable shock absorber 3i. If C> C _MIN , the process proceeds to step S49 It is determined whether the vertical speed X _2i ′ is positive. If X _2i ′> 0, then step S50.
To the damping coefficient C calculated in step S47.
Is set on the extension side, the target step angle θ _T is calculated with reference to the regions of θ _A to θ _B1 of the control map corresponding to FIG. 8, and then the process proceeds to step S51.

In step S51, the storage device 56d
Current step angle θ stored in _PAnd target step angle
θ_TAnd the deviation is calculated as step control amount S
While updating and storing in a predetermined storage area of the storage device 56d,
The target step angle θ_TThe current step angle θ_PAs
Newly memorize, then move to step S52,
Step control stored in a predetermined storage area of the unit 56d
Output the quantity S to the motor drive circuit 59i and then interrupt the timer
The process is terminated and the process returns to the predetermined main program.

Further, the determination result of step S49 is X._2i′
When <0, the process proceeds to step S53 and the
The damping coefficient C calculated in step S47 is set on the compression side.
As shown in the control map corresponding to FIG._B2~ Θ_CArea of
Target step angle θ _TAfter calculating
Go to step S51. Furthermore, the determination in step S48
The result is C ≦ C_MINIf so, go to step S54.
Transition to θ in the control map corresponding to FIG._B1~ Θ_B2Territory of
Target step angle θ_TAfter calculating
The process proceeds to step S51.

In the processing of FIG. 12, step S4
The processing of 1, S42 and step S46 corresponds to the control gain setting means. Therefore, assuming that the vehicle is currently parked and the key switch is in the off state,
In this state, the controller 4 is in a non-energized state, and
Although the processes of 0 to 12 are not executed, when the occupant gets on the vehicle and the key switch is turned on from this state, the controller 4 is powered on, and the process shown in FIGS.
The process of No. 2 is started to be executed.

In this state, since the vehicle is stopped,
Since the vehicle speed detection value V of the vehicle speed sensor 53 is zero and the processing of FIG. 10 is executed, the process proceeds from step S2 to step S3, and the control gain calculation map selection processing of FIG. 11 is executed. At this time, when the vehicle is stopped on a horizontal road surface, since the vehicle body is in a horizontal state, both the inclination angle detection values θ _X and θ _{Y of} the inclination angle sensors 54X and 54Y become zero. Moving to S34, the load detection value W _FL of the load sensors 55FL to 55RR at each wheel position
W _RR is read and, based on this, the sprung mass m _{2FL to} m at each wheel position according to the occupant's riding position and the loading position of the load.
_2RR is calculated (step S35), and the sprung resonance frequencies f _{UFL to} f _URR are calculated accordingly (step S3).
6) Select a control gain calculation map corresponding to these sprung resonance frequencies f _{UFL to} f _URR (step S3).
7). As a result, it is possible to select the control gain calculation map that accurately corresponds to the sprung mass resonance frequency change at each wheel position in accordance with the riding position of the occupant and the loading position of the load.

When the vehicle is parked on a slope or a sidewalk and stopped, the inclination angle detection values θ _X and θ _Y are output from the inclination sensors 54X and 54Y corresponding to the inclination of the vehicle body. Therefore, in step S32, it is determined that the vehicle body is not in the horizontal state, the process proceeds to step S33, the air suspensions 70FL to 70RR are actuated, and the vehicle body is set in the horizontal state by the vehicle height adjusting function. Select the control gain calculation map.

By continuing this control gain calculation map selection processing until the vehicle starts, the control gain calculation map is selected according to the riding position of the occupant and the loading position of the load immediately before the start of traveling. . When the vehicle is slowly started from this stopped state to start traveling, the vehicle speed detection value V of the vehicle speed sensor V increases from zero by this, and when the processing of FIG. The process proceeds to S4 and the damping force control process of FIG. 12 is executed.

In this slow start state, since there is almost no vertical movement occurring in the vehicle body, the vertical acceleration detection values X _2FL " _{-X 2RR} " output from the vertical acceleration sensors 51FL-51RR become substantially zero. Therefore, when the process of FIG. 12 is executed, the frequency f of the vertical acceleration detected in step S42
_{FL to} f _RR also become substantially zero, and the vehicle body vertical velocities X _2FL ′ to X _2RR ′ calculated in step S44 also become substantially zero, and the control gain C _S calculated in step S46 is “as shown in FIG. Although it is set to a minimum value slightly smaller than 1 ″, the damping coefficient C calculated in step S47 becomes substantially zero. Therefore, the process _proceeds from step S48 to step S54, and the step angle within the range of the step angles θ _{B1 to} θ _B2 , which are the expansion-side and _compression- side minimum damping coefficients C _nMIN and C _aMIN , is set as the target step angle θ _T , This step motor 41F
The driving is performed so that the step angle of L to 41RR matches the target step angle θ _T. Therefore, the valve elements 31 of the damping force variable shock absorbers 3FL to 3RR are set to the position B shown in FIG. 6, whereby the damping coefficients C on the extension side and the _compression side of the piston 8 are the minimum damping coefficients C _nMIN and C _{aMIN, respectively.} Is set to. Therefore, in this state, even if vibrations due to fine unevenness of the road surface are input to the wheels, the vibrations are absorbed by the damping force variable shock absorbers 3FL to 3RR and are not transmitted to the vehicle body, and a good ride comfort can be secured. .

In this running condition on a good road, for example, the step of rising front
When passing a temporary step such as a difference, pass this step
If the vehicle body does not move up and down due to
_2FL’～ X_2RRSince ′ maintains zero, the minimum damping coefficient C
_aMINAnd C_nMINTo maintain the condition, the wheels should ride on the step
It can absorb the pushing force when raised, but
Riding on a relatively large step and absorbing the thrust force
If you can not move it, the vehicle body will also be displaced upward.
Therefore, the vertical acceleration X_2FL″ 〜X_2RR″ Is square
Direction, increasing the vehicle body vertical speed X_2FL’～ X_2RR′ Is also positive
Will increase. Thus, the vehicle body vertical velocity X
_2FL’～ X_2RRIf ′ increases in the positive direction, step S4
Since the damping coefficient C calculated in 7 has a large value,
Step S50, the step angle θ in FIG._A~ Θ_B1of
Target step angle θ according to the damping coefficient C in the region_TIs calculated
Damping force variable shock absorber 3FL to 3RR
The valve body 31 is switch-controlled as shown in FIG. This conclusion
As a result, the relative speed X_DFL’～ X_DRR′
Is negative, that is, the displacement speed X on the vehicle body side_2iDisplacement on the wheel side with respect to ′
Speed X_1iWhen ′ is fast and the piston 8 moves to the pressure side,
Is the minimum damping coefficient C on the compression side. _aMINKeeps the car
Vibration input to the wheel side can be absorbed, and from this state
By climbing over the step, the wheel side rise speed
If the speed becomes smaller than the rising speed of_DFL′ 〜
X_DRR′ Becomes positive and the piston 8 moves to the extension side
become. At this time, the damping coefficient C becomes a large value.
The damping effect that suppresses the rise of the car body is demonstrated,
When the lifting of the body stops, the vertical velocity X of the vehicle body_2FL′ 〜
X_2RRSince ′ becomes zero, the step
Pump motors 41FL to 41RR are rotated counterclockwise and positioned
It is returned to B, which reduces the compression side and the extension side to the minimum.
Extinction coefficient C_aMINAnd C_nMINControlled, then the car body descends
, The vehicle body vertical speed X_2FL’～ X
_2RR′ Is increased in the negative direction, the step S8
To step S12, refer to the control map of FIG.
And step angle θ_B2~ Θ_CDepending on the damping coefficient C in the range of
Target step angle θ_TBy calculating
Further, it is rotated counterclockwise to the rotation position shown in FIG.
It is rotated. Therefore, the vehicle body descends and the relative speed X
_DFL’～ X_DRR′ Becomes negative and the piston 8 moves to the pressure side.
In the state where the
The damping effect is exhibited.

At this time, when the vertical movement of the vehicle body causes the frequencies f _{FL to} f _RR of the vertical acceleration calculated in step S42 to increase and become close to, for example, the sprung resonance frequencies f _{UFL to} f _URR , accordingly step S46 is performed. The control gain C _S calculated in step S4 is set to a large value, for example, about “4”, and accordingly the damping coefficient C calculated in step S47 also becomes a large value, resulting in a larger damping effect. Can be demonstrated.

When the frequencies f _{FL to} f _RR of the vertical acceleration reach an intermediate frequency of about 4 to 8 Hz, which is higher than the sprung resonance frequency, the control gain C _S correspondingly decreases, and the frequencies f _{FL to} f _RR. Is increased to reach an unsprung resonance frequency of about 10 to 12 Hz, the control gain C _S is increased again,
In addition to exerting a large damping effect, the unsprung power spectrum density can be reduced as shown by the solid line in FIG. 14, and unsprung fluttering can be suppressed.

As a result, the optimum control gain C _S can be set in all frequency regions of the vertical acceleration of the vehicle body according to the change of the sprung resonance frequency, and a good vibration damping effect is exerted to provide a comfortable ride. Can be improved. On the contrary, when the wheels pass through the step on the front lower side, the relative speeds X _DFL ′ to X _DRR ′ increase in the positive direction due to the wheels rebounding first, but at this time, the vehicle body does not move vertically.
Since the vehicle body vertical velocities X _2FL ′ to X _2RR ′ are zero, the damping coefficients of the variable damping force shock absorbers 3FL to 3RR maintain the minimum damping coefficients _{Ca MIN} and C _{nM IN} to allow the wheels to descend, and thereafter, When the vehicle body starts to descend, the vehicle body vertical speed X
_{When 2FL} ′ to X _2RR ′ increase in the negative direction, the damping coefficient C becomes a large value, and the target step angle θ _{T in} the range of step angles θ _{B2 to} θ _C is calculated, and the valve body 31 is Since the piston 8 is rotated to the position shown in FIG. 7, a large damping force can be applied to the pressure side movement of the piston 8 to exert a large damping effect, and thereafter the vehicle body vertical speeds X _2FL ′ to X 2 can be obtained.
_{As 2RR} ′ becomes smaller and the damping coefficient C becomes smaller, the valve body 31 is rotated clockwise and returned to the position B side.
When the vehicle body vertical speeds X _2FL ′ to X _2RR ′ become zero, the valve body 3
1 becomes the position B, which is the minimum damping coefficient C _aMIN and C _nMIN . After that, when the vehicle body starts to rise by _swinging back, the vehicle body vertical velocities X _2FL ′ to X _2RR ′ increase in the positive direction and the relative velocities X _DFL ′ to X _DRR ′ become the positive direction, so that the damping coefficient C Step angle θ _A
The target step angle θ _T on the side is calculated, the valve body 31 is rotated clockwise to the position shown in FIG. 5, and a large damping force is applied to the extension side movement of the piston 8. It is possible to exert a vibration damping effect.

As described above, when passing a temporary step while traveling on a good road, a good vibration damping effect can be exhibited by skyhook control, and when traveling on a bad road. Also, the target step angle θ _T on the step angle θ _A side (or step angle θ _C side) is calculated by the positive (or negative) of the vehicle body vertical velocities X _2FL ′ to X _2RR ′, so that the vehicle body rises. When the relative speeds X _DFL ′ to X _DRR ′ are negative and the vehicle body is descending so that the relative speeds X _DFL ′ to X _DRR ′ are positive, the damping coefficient C is the minimum damping coefficient C.
_By controlling to _aMIN and _CnMIN , conversely, the vehicle body rises and the relative velocity X _DFL ′ to X _DRR ′ is positive, and the vehicle body descends and the relative velocity X.
When the damping direction is such that _DFL ′ to X _DRR ′ are negative, the damping coefficient C is set to the vertical velocity X _2FL ′ to X _2RR ′ and the relative velocity X.
Controlling to the optimum damping coefficient according to _DFL '~ X _DRR ',
A good ride comfort can be secured.

[0044] In the above embodiment, as the control gain calculation map, has been described which represents the control gain C _S for all frequencies in the frequency domain other than the sprung resonance frequency, the influence of the load weight change Therefore, a plurality of maps may be prepared only for the sprung resonance frequency portion. Further, in the above embodiment, the case where the damping force variable shock absorber is applied as the suspension device has been described, but the present invention is not limited to this, and an air spring whose spring constant can be changed can also be applied.

In the above embodiment, the valve body 31 for controlling the damping force is described as a rotary type, but the present invention is not limited to this. Different flow paths may be formed on the side of the step motors.
A pinion may be connected to the rotating shaft 41a of 1RR and a rack meshing with the pinion may be attached to the connecting rod 42 or an electromagnetic solenoid may be applied to control the sliding position of the valve element 31.

Further, in the above embodiment, the case where the posture change of the vehicle body due to the vibration input from the road surface is suppressed has been described, but the present invention is not limited to this, and the traveling state such as the turning state and the braking state of the vehicle is detected, The control for suppressing the change in the posture of the vehicle body due to this may be performed together. Furthermore, in the above embodiment, the microcomputer 56
However, the present invention is not limited to this, and may be configured by combining electronic circuits such as a function generator, an integrator, and a differentiator.

Furthermore, in the above embodiment, the potentiometer type tilt angle sensor is applied as the tilt angle sensor, but the present invention is not limited to this, and a torque balance type tilt angle sensor and other tilt angle sensors can be applied. Further, in the above embodiment, the case where the potentiometer is applied as the stroke sensor has been described, but the present invention is not limited to this, and an ultrasonic distance sensor that detects the relative distance between the vehicle body and the road surface, and a detection coil are used. Any relative displacement detecting means such as a displacement sensor that detects displacement by impedance change or inductance change can be applied.

Further, in the above embodiment, the vertical acceleration sensors 51FL to 51FL are located at the respective wheels 1FL to 1RR of the vehicle body 2.
Although the case where RR is provided has been described, any one vertical acceleration sensor may be omitted and the vertical acceleration at the omitted position may be estimated from the values of other vertical acceleration sensors. Furthermore, in the above embodiment, the case where the step motors 41FL to 41RR are subjected to open loop control has been described, but the present invention is not limited to this, and the closed angle control is performed by detecting the rotation angle of the step motor with an encoder or the like and feeding it back. You may do it.

[0049]

As described above, according to the suspension control device of the present invention, the damping force is calculated based on the relative displacement detection value of the relative displacement detection means, the vehicle body vertical acceleration detection value of the vertical acceleration detection means, and the control gain. In a suspension control device that controls the suspension characteristics of a suspension device such as a variable shock absorber, the sprung mass is detected by detecting the wheel load at each wheel position when the vehicle is stopped, and based on this sprung mass The control gain calculation map is selected and the control gain calculation map is set based on the frequency of the vertical acceleration of the vehicle when the vehicle is running, so the control gain is set. Even if the sprung resonance frequency at each wheel position changes depending on the mounting position of the object, the control gain that accurately corresponds to the change Since the control gain calculation map is used, the control gain can be finely set according to the frequency of the vertical acceleration of the vehicle body, and good damping force control is performed to improve riding comfort. The effect that can be obtained is obtained.

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram showing a basic configuration of the present invention.

FIG. 2 is a schematic configuration diagram showing an embodiment of the present invention.

FIG. 3 is a front view with a part in section showing an example of a damping force variable shock absorber.

FIG. 4 is an enlarged cross-sectional view showing a damping force adjusting mechanism in a maximum damping force state when the vehicle body is raised.

FIG. 5 is an enlarged cross-sectional view showing a damping force adjusting mechanism in an intermediate damping force state when the vehicle body is raised, (a) showing an extension side and (b) showing a pressure side hydraulic fluid path, respectively.

FIG. 6 is an enlarged cross-sectional view showing a damping force adjusting mechanism when there is no change in the vehicle body, where (a) shows an extension side and (b) shows a pressure side hydraulic fluid path, respectively.

FIG. 7 is an enlarged sectional view showing a damping force adjusting mechanism in a maximum damping force state when the vehicle body is descending, (a) showing an extension side and (b) showing a pressure side hydraulic fluid path, respectively.

FIG. 8 is an explanatory diagram showing damping force characteristics with respect to a step angle of a damping force variable shock absorber.

FIG. 9 is a block diagram showing an example of a controller.

FIG. 10 is a flowchart illustrating an example of a processing procedure of a controller.

FIG. 11 is a flowchart showing a control gain calculation map selection process of the controller.

FIG. 12 is a flowchart showing a damping force control process of the controller.

FIG. 13 is a diagram showing an example of a control gain calculation map showing the relationship between the frequency of vertical acceleration and the control gain.

FIG. 14 is a characteristic diagram showing the relationship between the frequency of vertical acceleration and the unsprung power spectrum density.

FIG. 15 is a characteristic diagram showing the relationship between the frequency of vertical acceleration and the sprung power spectrum density.

[Explanation of symbols]

 1FL to 1RR Wheels 2 Vehicles 3FL to 3RR Damping force variable shock absorber 4 Controller 41FL to 41RR Step motor 51FL to 51RR Vertical acceleration sensor 52FL to 52RR Stroke sensor 53 Vehicle speed sensor 54X Front-rear tilt angle sensor 54Y Left-right tilt angle sensor 55FL to 55RR Load sensor 56 Microcomputer 59FL to 59RR Motor drive circuit 70FL to 70RR Air suspension

Claims (1)

[Claims] 1. A suspension device capable of changing a vehicle body posture change suppressing characteristic according to an input control signal interposed between a vehicle body side member and a wheel side member, and a suspension device at a position of the suspension device of the vehicle body. A vertical acceleration detecting means for detecting a vertical acceleration of the vehicle body; a relative displacement detecting means for detecting a relative displacement between the vehicle body side member and the wheel side member; and a vehicle body vertical acceleration detection value and a relative displacement detecting means of the vertical acceleration detecting means. In a suspension control device comprising a control means for forming and outputting the control signal for suppressing a change in attitude of a vehicle body based on a relative displacement detection value and a control gain, a wheel load at each wheel position is detected when the vehicle is stopped. A sprung mass detection means for detecting sprung mass, and a sprung resonance frequency at each wheel position is calculated based on the sprung mass detected by the sprung mass detection means. Sprung resonance frequency calculating means,
Map selection means for selecting a corresponding control gain calculation map based on the sprung mass resonance frequency calculated by the sprung mass frequency calculation means, and selection based on the frequency of the vehicle body vertical acceleration detection value while the vehicle is traveling. And a control gain setting means for setting a control gain with reference to the control gain calculation map.