Shock absorber damping coefficient control device

Abstract

To provide a damping coefficient control apparatus for a shock absorber, improved so as to be capable of satisfactorily reducing longitudinal force acting on a wheel due to vertical input from a road surface, even if an equivalent inertia moment of a driving system fluctuates due to a change of a gear ratio of a transmission.SOLUTION: A damping coefficient control apparatus for a shock absorber is applied to a vehicle 20 which includes: a wheel 12 supported rotatably about a rotational axis 24 and having a tire 26; and a driving system for driving the wheel by an engine 28 as a driving source via a transmission 30. The damping coefficient control apparatus includes: a variable damping coefficient shock absorber 18 arranged between a wheel carrier and a vehicle body; and a control device 44 for controlling the shock absorber. The control device acquires information on an index indicating an equivalent inertia moment IP of the driving system, and controls a damping coefficient C of the shock absorber such that the damping coefficient C of the shock absorber becomes larger as the equivalent inertia moment indicated by the index becomes smaller.SELECTED DRAWING: Figure 1

Description

[Technical Field]

[0001] The present invention relates to a damping coefficient control device for a shock absorber of a vehicle such as an automobile. [Background technology]

[0002] There is known a suspension that improves the ride comfort of a vehicle by reducing the longitudinal forces acting on the wheels due to vertical inputs from the road surface. For example, Patent Document 1 listed below describes a suspension in which suspension arms are arranged so that the rotation axis of the wheel, as viewed in the lateral direction of the vehicle, follows a trajectory that is tilted backward as the wheel moves up and down, and shock absorbers are arranged to be tilted forward.

[0003] With this type of suspension, the longitudinal forces acting on the wheels due to vertical inputs from the road surface can be at least partially offset by the longitudinal forces generated by the backward tilt of the trajectory and the longitudinal forces generated by the damping force of the shock absorber, thereby reducing the longitudinal forces on the wheels. [Prior art documents] [Patent documents]

[0004] [Patent Document 1] Japanese Patent Application Laid-Open No. 2009-40349 Summary of the Invention

[0005] [Problem to be solved by the invention] In a vehicle with a drivetrain in which the wheels are rotated by a drive source via a transmission, longitudinal vibrations caused by longitudinal forces at the wheels increase when the torsional resonance frequency of the drivetrain approaches the vertical resonance frequency of the suspension. As will be explained in detail later, the torsional resonance frequency of the drivetrain is determined by the moment of inertia and torsional rigidity of the drivetrain, and the moment of inertia of the drivetrain varies with changes in the gear ratio of the transmission. Therefore, the longitudinal forces acting on the wheels due to vertical inputs from the road surface vary with changes in the moment of inertia of the drivetrain, and therefore with changes in the gear ratio of the transmission.

[0006] The suspension described in Patent Document 1 cannot deal with fluctuations in the longitudinal forces acting on the wheels due to fluctuations in the moment of inertia of the drivetrain, and therefore cannot effectively reduce the longitudinal forces acting on the wheels due to vertical inputs from the road surface, regardless of fluctuations in the moment of inertia of the drivetrain that accompany changes in the gear ratio of the transmission.

[0007] The present invention provides an improved shock absorber damping coefficient control device that can effectively reduce the longitudinal forces acting on the wheels due to vertical inputs from the road surface, even if the equivalent moment of inertia of the drivetrain fluctuates as the gear ratio of the transmission changes.

[0008] [Means for solving the problems and effects of the invention] According to the present invention, there is provided a shock absorber damping coefficient control device (10) applied to a vehicle (20) including a wheel (12) rotatably supported around a rotation axis (24) by a wheel carrier (22) and having a tire, a drive system (14) that rotates and drives the wheel via a transmission (30) by a drive source (engine 28), and a suspension arm (16) arranged between the wheel carrier and a vehicle body (34).

[0009] The damping coefficient control device includes a variable damping coefficient shock absorber (18) disposed between the wheel carrier and the vehicle body, and a control device (44) for controlling the shock absorber, and the control device calculates an equivalent moment of inertia (I) of the drive train (14). PThe system is configured to obtain information on an index (speed ratio Rt or gear stage St) indicating the equivalent moment of inertia, and control the damping coefficient (C) of the shock absorber so that the smaller the equivalent moment of inertia indicated by the index, the larger the damping coefficient (C) of the shock absorber.

[0010] As will be explained in detail later, one way to reduce the longitudinal forces acting on the wheels due to vertical inputs from the road surface to the wheels is to match the vertical resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft. The anti-resonance frequency of the drive shaft varies depending on the equivalent moment of inertia of the drive train, and the vertical resonance frequencies of the unsprung mass vary depending on the damping coefficient of the shock absorber. The damping coefficient of the shock absorber required to match the vertical resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft increases as the equivalent moment of inertia of the drive train decreases.

[0011] According to the above configuration, the damping coefficient of the shock absorber is controlled so that the smaller the equivalent moment of inertia indicated by the index, the larger the damping coefficient of the shock absorber. Therefore, the upper and lower resonance frequencies of the unsprung mass can be changed in accordance with the change in the anti-resonance frequency of the drive shaft that accompanies the change in the equivalent moment of inertia of the drive train. Therefore, even if the equivalent moment of inertia of the drive train changes, the longitudinal forces acting on the wheels can be effectively reduced compared to when the damping coefficient of the shock absorber is not changed in accordance with the change in the equivalent moment of inertia of the drive train.

[0012] [Mode of the Invention] In one aspect of the present invention, the wheel carrier (22) and the wheel (12) constitute an unsprung mass, and the control device (44) stores a relationship between a target damping coefficient (Ct) of the shock absorber (18) and an index (speed ratio Rt or gear St) required to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft (30), and is configured to calculate the target damping coefficient from the relationship based on the index, and control the damping coefficient so that the damping coefficient (C) becomes the target damping coefficient.

[0013] In another aspect of the present invention, the relationship between the target damping coefficient (Ct) of the shock absorber (18) and the index (speed ratio Rt or speed St) is determined by the target damping coefficient of the shock absorber and the equivalent moment of inertia (I) required to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft. P ) is set based on the relationship with

[0014] Furthermore, in another aspect of the present invention, the index is the gear ratio (Rt) of the transmission (30), and the above relationship is a relationship between the target damping coefficient (Ct) and the gear ratio of the transmission, which is set so that the target damping coefficient (Ct) increases as the gear ratio of the transmission decreases.

[0015] Furthermore, in another aspect of the present invention, the transmission (30) is a multi-speed transmission, the index is the gear stage (St) of the transmission, and the above relationship is the relationship between the target damping coefficient (Ct) and the gear stage of the transmission, which is set so that the target damping coefficient (Ct) increases as the gear stage of the transmission becomes higher.

[0016] Other objects, other features and attendant advantages of the present invention will be readily apparent from the following description of the preferred embodiments of the present invention which will be given with reference to the accompanying drawings. [Brief explanation of the drawings]

[0017] [Figure 1] 1 is a schematic diagram showing an embodiment of a damping coefficient control device for a shock absorber according to the present invention; [Figure 2] FIG. 2 is an explanatory diagram showing wheels driven by a drive train. [Figure 3] 1A is a diagram showing an example of the relationship between the vertical displacement Z of the wheel and the longitudinal force FX1; FIG. 1B is a diagram showing an example of the relationship between the vertical displacement Z of the wheel and the longitudinal force FX2; and FIG. 1C is a diagram showing an example of the relationship between the vertical displacement Z of the wheel and the longitudinal force FX3. [Figure 4] FIG. 10 is a diagram showing a longitudinal force FXT acting on a wheel due to the inclination of the road surface. [Figure 5] FIG. 10 is a diagram showing longitudinal forces FXS acting on the wheels due to wheel slippage. [Figure 6] FIG. 10 is a diagram showing the relationship between the gain of h′(s) and the frequency (upper part) and the relationship between the phase of h′(s) and the frequency (lower part). [Figure 7] 1A shows a single-wheel model of a vehicle in which the upper support damper is omitted, and FIG. 1B shows a single-wheel model of a vehicle in which the suspension spring and upper support spring are integrated into the suspension spring. [Figure 8] FIG. 10 is a diagram showing an example of the relationship between M×(2πf) 2 and frequency f (thick line) and the relationship between the sum of spring constants K′+KZ and frequency f (thin line). [Figure 9] FIG. 4 is a diagram illustrating an example of the relationship between unsprung longitudinal acceleration Gx and frequency f. [Figure 10] FIG. 10 is a diagram illustrating an example of the relationship between the equivalent moment of inertia IP of the drive train, which determines the resonance frequency ωD, and the damping coefficient C of the shock absorber. [Figure 11] 1A is a diagram showing an example of the relationship between the gear ratio Rt and the target damping coefficient Ct of the shock absorber, and FIG. 1B is a diagram showing an example of the relationship between the gear position St and the target damping coefficient Ct of the shock absorber. [Figure 12] 4 is a flowchart corresponding to a damping coefficient control program according to the embodiment. DETAILED DESCRIPTION OF THE INVENTION

[0018] [Embodiment] DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, embodiments of the present invention will be described in detail with reference to the accompanying drawings.

[0019] As shown in FIGS. 1 and 2, a damping coefficient control device 10 according to an embodiment of the present invention is applied to a vehicle 20 having a wheel 12, a drivetrain 14, and a suspension arm 16, and includes a variable damping coefficient shock absorber 18. The vehicle 20 may be an autonomous vehicle. The wheel 12 is rotatably supported about a rotation axis 24 by a wheel carrier 22 and has an elastically deformable tire 26, as is well known. The drivetrain 14 is configured to rotate the wheel 12 via a transmission 30 and a drive shaft 32 using an engine 28 as a drive source. The suspension arm 16 and the shock absorber 18 are disposed between the wheel carrier 22 and a vehicle body 34.

[0020] 1 shows only one suspension arm, the suspension arm may include multiple arms, links, etc., and suspension arm 16 indicates a suspension arm equivalent to multiple arms, links, etc. when viewed from the side of vehicle 20. The drive source may be any drive source known in the art other than an engine. Furthermore, in the embodiment, transmission 30 is a continuously variable transmission, but may also be a multi-stage transmission such as a gear transmission.

[0021] The suspension arm 16 is disposed between the wheel carrier 22 and the vehicle body 34 so that, when viewed laterally of the vehicle 20, the rotation axis 24 describes a locus 36 that is tilted backward at an angle γ with respect to the vertical as the wheel 12 moves up and down. The end of the suspension arm 16 on the vehicle body 34 side is positioned higher than the end on the wheel carrier 22 side. When the wheel 12 is in a neutral position for vertical displacement, the inclination angle of the suspension arm 16, i.e., the angle that the line connecting both ends of the suspension arm 16 makes with the horizontal, is β. Note that the locus 36 does not have to be a straight line, and the backward inclination angle γ of the locus 36 may be considered to be the same as the inclination angle β within the range of vertical displacement of the wheel 12 as the vehicle 20 moves.

[0022] The shock absorber 18 is disposed between the wheel carrier 22 and the vehicle body 34 in a forward tilted state, and the angle of forward tilt, i.e., the angle that the main axis of the shock absorber (not shown) makes with the vertical direction, is α. The shock absorber 18 is connected at its upper end to the vehicle body 34 via an upper support 38, and at its lower end to the wheel carrier 22. The upper support 38 functions as a spring 38A and a damper 38B. As is well known, the shock absorber 18 generates a damping force corresponding to the product of the relative speed between the wheel 12 and the vehicle body 34 and a damping coefficient C, and the damping force changes as the damping coefficient C changes.

[0023] The angle γ of the backward tilt of the trajectory 36 and the angle α of the forward tilt of the shock absorber are determined by the longitudinal force F acting on the wheel 12 at the contact point P due to the vertical input from the road surface 40. X 1 is the longitudinal force F generated by the backward tilt of the trajectory 36 X 2 and the longitudinal force F generated by the damping force of the shock absorber X 3.

[0024] For example, FIG. 3A shows the relationship between the vertical displacement Z and the longitudinal force F of the wheel 12 as the unsprung part. X 1 and the longitudinal force F X 1 is a positive value when the vertical displacement Z is positive (upward displacement). FIG. 3(B) shows the relationship between the vertical displacement Z of the wheel 12 and the longitudinal force F X 2 and the longitudinal force F X 2 is a negative value when the vertical displacement Z is positive, and is proportional to the vertical displacement Z. FIG. 3(C) shows the relationship between the vertical displacement Z of the wheel 12 and the longitudinal force F X 3 shows an example of the relationship between the longitudinal force F X 3 is a differential value of the vertical displacement Z, that is, when the upward displacement speed of the wheel 12 is positive, it becomes a negative value and is proportional to the displacement speed.

[0025] Fore-and-aft force F X 2 and F X The sum of these three is the value shown by the dashed line in Figure 3(A). X 1, the longitudinal force F X 2 and FX 3(A) and 3(C), the arrows indicate the longitudinal force F X 1 and F X This shows the direction of change in 3.

[0026] Also disposed between the wheel carrier 22 and the vehicle body 34 are suspension springs 42. In the illustrated embodiment, the suspension springs 42 are coil springs, but may be any spring known in the art, such as air springs.

[0027] The damping coefficient control device 10 further includes a control device 44 that controls the shock absorber 18. The control device 44 includes a drive system control ECU 48 and a damping coefficient control ECU 50. ECU stands for Electronic Control Unit, which includes a microcomputer as its main component.

[0028] The microcomputer of each ECU includes a CPU, ROM, RAM, a read / write nonvolatile memory (N / M), an interface (I / F), etc. The CPU realizes various functions by executing instructions (programs, routines) stored in the ROM. Furthermore, these ECUs and sensors such as a vehicle speed sensor (not shown) are connected to each other via a CAN (Controller Area Network) 52 so as to be able to communicate with each other.

[0029] The drivetrain control ECU 48 controls the output torque T of the drivetrain 14, and therefore the drive torque applied to the wheels 12, by controlling the output of the engine 28 and the gear ratio (gear ratio) Rt of the transmission 30. The drivetrain control ECU 48 outputs a signal indicating the gear ratio Rt of the transmission 30 to the damping coefficient control ECU 50 via the CAN 52. The gear ratio Rt is calculated based on the equivalent moment of inertia I of the drivetrain 14. P is an index showing the equivalent moment of inertia I Pis the product of the sum of the equivalent moments of inertia of the individual components of the transmission 30 and the square of the speed ratio Rt, and is therefore proportional to the square of the speed ratio Rt.

[0030] As mentioned above, the longitudinal force F X 2 and F X The sum of the three gives the longitudinal force F X By canceling out 1, the longitudinal force F X The reduction of the longitudinal force F is the principle of the longitudinal force reduction described in the aforementioned Patent Document 1. In the embodiment of the present invention, the damping coefficient C of the shock absorber 18 is controlled to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft, thereby reducing the longitudinal force F X In the embodiment, the longitudinal force F X 2 and F X The sum of the three gives the longitudinal force F X Preferably, 1 is at least partially cancelled out.

[0031] As will be explained in detail later, the wheel carrier 22 and the wheel 12 constitute an unsprung mass, and the damping coefficient control ECU 50 controls the damping coefficient C of the shock absorber 18 so that the upper and lower resonance frequencies of the unsprung mass match the anti-resonance frequency of the drive shaft 32. The anti-resonance frequency of the drive shaft 32 is determined by the equivalent moment of inertia I of the drive train 14. P Therefore, the damping coefficient control ECU 50 changes in accordance with the change in the equivalent moment of inertia I P The information on the gear ratio Rt as an index showing the equivalent moment of inertia is acquired, and the damping coefficient is controlled so that the damping coefficient C of the shock absorber 18 increases as the equivalent moment of inertia shown by the index decreases.

[0032] [Principle of damping coefficient control adopted in the embodiment of the present invention] To facilitate understanding of the present invention and its embodiments, the principle of damping coefficient control in the present invention will be described.

[0033] A longitudinal force F acting on the wheel 12 at the contact point P due to vertical input from the road surface 40 X1 is a longitudinal force F acting on the wheel 12 due to the inclination of the road surface 40. XT (FIG. 4) and the longitudinal force F acting on the wheel due to the slip of the wheel 12. XS (Figure 5)

[0034] As shown in FIG. 4, the inclination angle of the road surface 40 is θ X The ground load of the wheel 12 is defined as W. The steady component of the vehicle speed is defined as U, the fluctuating component of the longitudinal speed of the wheel 12 is defined as ΔU, and the vertical displacement of the road surface 40 is defined as Z0. The longitudinal force F XT is expressed by the following equation (1).

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[0035] As shown in Figures 2 and 5, the driving stiffness is X The slip ratio of the wheel 12 is S X The radius (steady-state component) of the tire 26 is r0, and the fluctuation component of the rolling radius of the tire 26 is ηΔr. Here, η is the ratio of the rolling radius fluctuation amount to the vertical deformation amount of the tire. The steady-state component and fluctuation component of the rotational angular velocity of the wheel 12 are ω0 and Δω, respectively. The longitudinal force F acting on the wheel at the contact point P due to the driving torque applied to the wheel 12 is XW The front and rear spring constants of the tire 26 are K X and the Laplace operator is s. The longitudinal force F XS teeth 、 It is expressed by the following formula (2).

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[0036] The rotational angular velocities of the drive train 14 and the wheels 12 are respectively ω P and ω T (=ω0+Δω), and the torsional rigidity of the drive shaft 32 is K P Then, the equation of motion in the rotational direction of the drive system 14 is given by the following equation (3).

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[0037] The equivalent moment of inertia of the wheel 12 is I T Then, the equation of motion in the rotational direction of the wheel is given by the following equation (4).

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[0038] From the above equations (3) and (4), the longitudinal force F XW By substituting the above equation (2), the longitudinal force F acting on the wheel 12 is calculated. X 1 is expressed by the following equation (5).

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[0039] In equation (5), h'(s) is a term of dynamic characteristics, and the term in curly brackets is a term of steady characteristics. The first and third terms of the steady characteristics are the longitudinal force F X The influence on 1 is smaller than that of the second term. Therefore, if the first and third terms are omitted, h'(s) can be expressed by the following equation (6).

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[0040] In addition, in equation (6), ω E is the resonance frequency of the rotation of the drive system 14 expressed by the following equation (7), and ω W is the resonance frequency of the rotation of the wheel 12, which is expressed by the following equation (8): D is the torsional resonance frequency of the drive shaft 32, i.e., the anti-resonance frequency, expressed by the following equation (9), and ζ W is the damping ratio in the rotational direction of the wheel 12, expressed by the following equation (10).

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[0041] The upper part of Fig. 6 shows an example of the relationship between the gain of h'(s) and the frequency, and the lower part of Fig. 6 shows an example of the relationship between the phase of h'(s) and the frequency. In Fig. 6, the solid line, the dashed line, and the dashed dotted line respectively represent the equivalent moment of inertia I P The values ​​are shown for 0, 0.43 and 0.60.

[0042] From Figure 6, the equivalent moment of inertia I P As the frequency band of 10 to 20 Hz increases, the gain of h'(s) increases and the phase of h'(s) lags. D It can be seen that it becomes a minimum at .

[0043] FIG. 7(A) shows a single-wheel model of the vehicle 20, omitting the damper 38B (see FIG. 1) of the upper support 38. In FIG. 7(A), Z0 and Z respectively represent the vertical displacement of the road surface 40 and the vertical displacement of the unsprung mass 54. U is the spring constant of the upper support 38 (spring 38A), and K C is the spring constant of the suspension spring 42, and K Z is the vertical spring constant of the tire 26. Furthermore, C is the variable damping coefficient of the shock absorber 18, and M is the mass of the unsprung mass 54.

[0044] Fig. 7(B) shows a single-wheel model of the vehicle 20 in which the suspension spring 42 and the spring 38A of the upper support 38 are integrated into a suspension spring 42'. In Fig. 7(B), K' is the equivalent spring constant of the suspension spring 42', and C' is the equivalent damping coefficient of the shock absorber 18. The equivalent spring constant K' and the equivalent damping coefficient C' are expressed by the following equations (11) and (12), respectively.

number

[0045] From FIG. 7(B) and the above formula (11), the vertical resonance frequency of the unsprung mass 54 is a frequency f that satisfies the following formula (13).

number

[0046] Figure 8 shows M×(2πf) 2 and frequency f (thick solid line), and the sum of spring constants K'+K Z 8 shows an example of the relationship (thin lines) between the damping coefficient C (Nsec / m) and the frequency f. In Fig. 8, among the thin lines, the solid line, dashed line, dashed line, and dashed line indicate values ​​when the damping coefficient C (Nsec / m) is 1859, 3492, 5341, and 10072, respectively.

[0047] In Fig. 8, circles indicate the intersections of the thick solid line and multiple thin lines, and therefore the points where the above formula (13) is established, and the frequencies at each intersection indicate the upper and lower resonance frequencies of the unsprung mass 54. It can be seen from Fig. 8 that the upper and lower resonance frequencies of the unsprung mass 54 change depending on the damping coefficient C, and in particular, become higher as the damping coefficient C increases.

[0048] Torsional rigidity K of drive shaft 32 P is the anti-resonance frequency ω expressed by the above equation (10). D is set to a stiffness that provides the upper and lower resonance frequencies of the unsprung mass 54. When the damping coefficient C is 5341, the upper and lower resonance frequencies of the unsprung mass 54 are 14 Hz.

[0049] 9 shows an example of the relationship between the longitudinal acceleration Gx of the unsprung mass 54 and the frequency f. In FIG. 9, the solid line indicates the anti-resonance frequency f D =ω D / The dashed line indicates the value when 2π is 14Hz, and the anti-resonance frequency f D The figure shows the value when the anti-resonance frequency f D is a frequency that satisfies the above equation (13), the anti-resonance frequency f DIt can be seen that the longitudinal acceleration Gx of the unsprung mass 54 is reduced compared to when the anti-resonance frequency f does not satisfy the formula (13), and thus the longitudinal force acting on the unsprung mass can be reduced. In other words, the longitudinal acceleration Gx of the unsprung mass 54 and the anti-resonance frequency f D It can be seen that if these match, the longitudinal force acting on the unsprung mass can be reduced compared to when these do not match.

[0050] The anti-resonance frequency ω expressed by the above formula (10) D is the equivalent moment of inertia I of the drivetrain 14 P Therefore, as shown in Table 1 below, the equivalent moment of inertia I P The values ​​in Table 1 depend on the equivalent moment of inertia I T is 3kgm 2 and the torsional stiffness K of the drive shaft 32 P This is the value when the strain is 8000Nm / rad. [Table 1]

[0051] Frequency f is the anti-resonance frequency f D When this is the case, the following equation (14) is established from the above equation (13): From equation (14), the damping coefficient C of the shock absorber 18 is expressed by the following equation (15).

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[0052] Figure 10 shows the anti-resonance frequency ω D Determine the equivalent moment of inertia I of the drivetrain 14 P and the damping coefficient C of the shock absorber 18. From FIG. 10, it can be seen that the damping coefficient C of the shock absorber 18 that satisfies the above formula (13) is P It can be seen that the smaller is, the larger is.

[0053] FIG. 11(A) shows the equivalent moment of inertia IP 11A shows an example of the relationship between the gear ratio Rt and the target damping coefficient Ct of the shock absorber 18, based on the relationship between the gear ratio Rt and the gear ratio Rt of the transmission 30. As shown in FIG. 11A, the target damping coefficient Ct increases as the gear ratio Rt decreases.

[0054] In this embodiment, the ROM of the damping coefficient control ECU 50 stores a damping coefficient control program and a map corresponding to Fig. 11(A), i.e., a map of the relationship between the gear ratio Rt and the target damping coefficient Ct. The program corresponds to the flowchart shown in Fig. 12, and the damping coefficient control is executed in accordance with this flowchart.

[0055] <Damping coefficient control program according to the embodiment> Next, the damping coefficient control in this embodiment will be described with reference to the flowchart shown in Fig. 12. The damping coefficient control according to the flowchart shown in Fig. 12 is repeatedly executed at predetermined time intervals by the CPU of the damping coefficient control ECU 50 when an ignition switch (not shown) is on.

[0056] First, in step S10, the CPU reads a signal indicating the gear ratio Rt of the transmission 30 from the drivetrain control ECU 48. As described above, the gear ratio Rt is calculated by multiplying the equivalent moment of inertia I of the drivetrain 14 by P It is an indicator that shows the following.

[0057] In step S20, the CPU determines the target damping coefficient Ct by referring to a map corresponding to FIG. 11(A) based on the gear ratio Rt.

[0058] In step S30, the CPU controls the shock absorber 18 so that the damping coefficient C of the shock absorber 18 becomes the target damping coefficient Ct.

[0059] As can be seen from the above explanation, in order to reduce the longitudinal force acting on the wheel 12 due to the vertical input from the road surface 40 to the wheel 12, it is necessary to adjust the vertical resonance frequency of the unsprung mass and the anti-resonance frequency ω of the drive shaft 32. DThe anti-resonance frequency of the drive shaft is determined by the equivalent moment of inertia I P The upper and lower resonance frequencies of the unsprung mass change depending on the damping coefficient C of the shock absorber 18. The damping coefficient of the shock absorber, which is used to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft, becomes larger as the equivalent moment of inertia of the drive system becomes smaller.

[0060] According to the embodiment, the equivalent moment of inertia I indicated by the indicator (speed ratio Rt or speed stage St) P The damping coefficient of the shock absorber is controlled so that the smaller is , the larger is the damping coefficient C of the shock absorber 18. Therefore, the upper and lower resonance frequencies of the unsprung mass can be changed in accordance with the change in the anti-resonance frequency of the drive shaft 32 that accompanies a change in the equivalent moment of inertia of the drive train 24. Therefore, even if the equivalent moment of inertia of the drive train changes, the longitudinal forces acting on the wheels can be effectively reduced compared to when the damping coefficient of the shock absorber is not changed in accordance with the change in the equivalent moment of inertia, and thereby the longitudinal forces input from the wheels to the vehicle body can be effectively reduced.

[0061] In particular, according to the embodiment, the control device 44 adjusts the upper and lower resonance frequencies of the unsprung mass to the anti-resonance frequency ω of the drive shaft 30. D The controller 44 stores the relationship (FIG. 11) between the target damping coefficient Ct of the shock absorber 18 and the index (speed ratio Rt or gear position St) required to match the target damping coefficient Ct. Furthermore, the controller 44 obtains the target damping coefficient Ct from the above relationship based on the index, and controls the damping coefficient C so that the damping coefficient C becomes the target damping coefficient.

[0062] Therefore, by determining the target damping coefficient Ct from the above relationship based on the index and controlling the damping coefficient so that the damping coefficient C becomes the target damping coefficient, the upper and lower resonance frequencies of the unsprung mass can be matched with the anti-resonance frequency of the drive shaft 30, and the longitudinal forces acting on the wheels can be effectively reduced.

[0063] According to the embodiment, the relationship between the target damping coefficient Ct of the shock absorber 18 and the index (speed ratio Rt or gear St) is determined by the target damping coefficient of the shock absorber and the equivalent moment of inertia I required to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft. P It is set based on the relationship with

[0064] Therefore, the relationship between the target damping coefficient Ct of the shock absorber 18 and the index (speed ratio Rt or gear St) can be set to a relationship required to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft.

[0065] Furthermore, according to the embodiment, the index is the gear ratio Rt of the transmission 30, and the above relationship is the relationship between the target damping coefficient and the gear ratio of the transmission, which is set so that the target damping coefficient Ct becomes larger as the gear ratio of the transmission becomes smaller.

[0066] Therefore, by calculating the target damping coefficient Ct from the above relationship based on the gear ratio Rt of the transmission 30 and controlling the damping coefficient so that the damping coefficient C becomes the target damping coefficient, the longitudinal forces acting on the wheels can be effectively reduced.

[0067] Although the present invention has been described in detail above with reference to specific embodiments, it will be apparent to those skilled in the art that the present invention is not limited to the above-described embodiments, and that various other embodiments are possible within the scope of the present invention.

[0068] For example, in the above embodiment, the equivalent moment of inertia I of the drivetrain 14 is P The relationship between the index indicating this and the target damping coefficient Ct is the relationship between the gear ratio Rt of the transmission 30 and the target damping coefficient Ct (FIG. 11(A)). However, if the transmission 30 is a multi-stage transmission, the relationship between the index and the target damping coefficient Ct may be the relationship between the gear stage St of the transmission 30 and the target damping coefficient Ct (variation example). In this case, the relationship between the gear stage St and the target damping coefficient Ct is set so that the target damping coefficient Ct increases as the gear stage increases (FIG. 11(B)).

[0069] In Fig. 11(B), the black circles indicate examples of gear positions. In a modified example, a signal indicating the gear position St is read in step S10 of Fig. 12, and the target damping coefficient Ct is calculated in step S20 by referring to a map corresponding to Fig. 11(B) based on the gear position St.

[0070] The gear ratio Rt of the transmission 30 becomes smaller as the gear stage St of the transmission 30 becomes higher. Therefore, the equivalent moment of inertia I P becomes smaller as the gear stage of the transmission becomes higher. According to a modified example, the target damping coefficient Ct becomes larger as the gear stage St of the transmission 30 becomes higher. Therefore, the equivalent moment of inertia I of the drivetrain 14 changes with the change in the gear stage of the transmission. P Even if the equivalent moment of inertia fluctuates, the damping coefficient of the shock absorber can be controlled so that the damping coefficient C increases as the equivalent moment of inertia decreases.

[0071] In the above-described embodiment, the suspension arm 16 is disposed between the wheel carrier 22 and the vehicle body 34 so that, when viewed laterally of the vehicle 20, the rotation axis 24 describes a locus 36 that is tilted rearward at an angle γ with respect to the vertical direction as the wheel 12 moves up and down. The shock absorber 18 is disposed between the wheel carrier 22 and the vehicle body 34 in a forward-tilting state. However, the suspension arm 16 may be disposed so that the rotation axis 24 describes a locus that is not tilted rearward, and the shock absorber 18 may be disposed so that it is not tilted forward. It is preferable that the suspension arm 16 and the shock absorber 18 are disposed so that at least one of the backward tilt of the locus and the forward tilt of the shock absorber is satisfied. [Explanation of symbols]

[0072] 10... Vehicle height control device, 12... Wheel, 14... Drive system, 16... Suspension arm, 18... Shock absorber, 20... Vehicle, 22... Wheel carrier, 24... Rotation axis, 26... Tire, 28... Engine, 30... Transmission, 34... Vehicle body, 36... Trajectory, 40... Road surface, 42... Suspension spring, 44... Control device, 50... Damping coefficient control ECU

Claims

1. A damping coefficient control device for a shock absorber applied to a vehicle including: a wheel having a tire and supported rotatably around a rotation axis by a wheel carrier; a drive system that rotationally drives the wheel by a drive source via a transmission and a drive shaft; and a suspension arm disposed between the wheel carrier and a vehicle body, The damping coefficient control device includes a variable damping coefficient shock absorber arranged between the wheel carrier and the vehicle body, and a control device that controls the shock absorber, wherein the control device is configured to obtain information on an index that indicates an equivalent moment of inertia of the drive system, and to control the damping coefficient of the shock absorber so that the smaller the equivalent moment of inertia indicated by the index, the larger the damping coefficient of the shock absorber.

2. 2. The shock absorber damping coefficient control device according to claim 1, wherein the wheel carrier and the wheel constitute an unsprung mass, and the control device stores a relationship between the target damping coefficient of the shock absorber and the index, which is necessary to match the upper and lower resonance frequencies of the unsprung mass with the anti-resonance frequency of the drive shaft, and is configured to determine the target damping coefficient from the relationship based on the index, and control the damping coefficient so that the damping coefficient becomes the target damping coefficient.

3. 3. The shock absorber damping coefficient control device according to claim 2, wherein the relationship between the target damping coefficient of the shock absorber and the index is set based on the relationship between the target damping coefficient of the shock absorber and the equivalent moment of inertia, which is required to match the upper and lower resonance frequencies of the unsprung portion with the anti-resonance frequency of the drive shaft.

4. 3. The shock absorber damping coefficient control device according to claim 2, wherein the index is a gear ratio of the transmission, and the relationship is a relationship between the target damping coefficient and the gear ratio of the transmission, the target damping coefficient being set so that the smaller the gear ratio of the transmission, the larger the target damping coefficient.

5. 3. The shock absorber damping coefficient control device according to claim 2, wherein the transmission is a multi-stage transmission, the index is a gear stage of the transmission, and the relationship is a relationship between the target damping coefficient and the gear stage of the transmission such that the target damping coefficient increases as the gear stage of the transmission is higher.

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