JPH1191329A - Ground load control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To suppress an unreasonable turn and improve turning performance temporarily exceeding a limit by generating proper steering reaction not exceeding the turn limit, generating the inertial force in the vertical direction on a vehicle body at the extension acceleration of an actuator, and increasing the ground load of tires with the reaction. SOLUTION: Signals are read from a lateral acceleration sensor 32, a steering angle sensor 30, and a vehicle speed sensor 29, the maximum allowable steering angle and the actual steering angle are compared by a road surface state detection section 24, and an excessive turn is judged by an excessive turn judgment section 28. Additional reaction is determined to allow a driver to sense it as an increase of steering reaction torque. When an excessive turn is further judged, it is considered as the steering for avoiding an obstacle, the expansion acceleration of an actuator 5 is controlled, and the ground load can be temporarily increased for each tire 1. The grip force of a steering wheel is improved, a steering limit is temporarily increased, and turning performance is improved.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a grounding load provided with grounding load control means capable of temporarily increasing a grounding load by generating acceleration in at least one of a sprung mass and an unsprung mass. The present invention relates to a control device, and more particularly to a contact load control device that can contribute to improvement of acceleration performance and turning performance.

[0002]

2. Description of the Related Art For example, in a region where a driver's steering amount is large when turning a vehicle on a low μ road, the behavior of an actual vehicle may not follow the steering amount. Because the torque is small, the steering reaction force is also small, and the driver may turn the steering wheel beyond the steering limit. To prevent this, an appropriate steering reaction force (active steering reaction force) is used to generate the auxiliary steering torque. Conventionally, a steering reaction force generating device generated by the above actuator has been proposed (for example, see Japanese Patent Application No. 8-255572).

[0003]

However, even in a vehicle provided with the active steering reaction force generating device,
It was possible for the driver to turn the steering wheel against the active steering reaction force, but this could not be prevented. Further, when the driver tries to turn the steering wheel even if the driver resists the above-mentioned active steering reaction force, the steering may be necessary to avoid obstacles or the like. It is desirable to reflect the behavior.

The grip force F of a tire is given by the product (F = μW) of the coefficient of friction μ between the tire and the road surface and the vertical load W applied to the contact surface of the tire. That is, it can be said that the grip force of the tire, which greatly affects the mobility of the vehicle, is proportional to the ground contact load if the friction coefficient between the tire and the road surface is constant.

The present invention has been devised to solve such problems imposed on the prior art, and its main object is to suppress an excessive increase in the steering angle beyond an appropriate steering range. Another object of the present invention is to provide a ground load control device capable of stably obtaining a turning performance temporarily exceeding a turning limit during steering for avoiding an obstacle or the like.

[0006]

In order to achieve such an object, according to the present invention, the vehicle speed, each wheel speed,
From the signals of the steering angle, the yaw rate, the lateral acceleration, etc., it is impossible to perform a desired turning as it is, that is, when it is determined that the vehicle is over-cutting, an appropriate steering reaction force is generated so as not to exceed the turning limit, and the steering is performed. If the turning force is further increased even after the reaction force has been generated, it is determined that this is avoidance steering of obstacles, etc., and a vertical inertial force is generated in the vehicle body by the extension acceleration of the actuator provided between the vehicle body and the axle, The reaction force increases the ground contact load of the tire, and temporarily achieves steering performance exceeding the turning limit.

According to this, during normal turning, steering exceeding the steering limit is suppressed by generating an appropriate steering reaction force, and when avoiding an obstacle or the like, a load exceeding the vehicle weight is temporarily applied to the contact surface of the tire. As a result, the generation limit of the grip force of the tire can be raised as desired, and the temporary turning performance can be improved.

[0008]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS The configuration of the present invention will be described below in detail with reference to the accompanying drawings.

FIG. 1 schematically shows a schematic configuration of a main part of an active suspension system of a vehicle to which the present invention is applied. The tire 1 is supported by the upper and lower suspension arms 2 and 3 so as to be vertically movable with respect to the vehicle body 4. A linear actuator 5 driven by hydraulic pressure is provided between the lower suspension arm 3 and the vehicle body 4.

The linear actuator 5 constituting the ground load control means is of a cylinder / piston type, and is supplied from a variable displacement hydraulic pump 9 to oil chambers 7 and 8 above and below a piston 6 inserted into the cylinder. By controlling the operating hydraulic pressure by the servo valve 10, a vertical thrust is generated on the piston rod 11, whereby the relative distance between the center (axle) of the tire 1 and the vehicle body 4 can be freely changed. I can do it.

The oil discharged from the pump 9 is stored in an accumulator 12 for removing the pump pulsation and securing the oil amount in a transient state. Are supplied via servo valves 10 provided individually.

An unload valve 13, an oil filter 14, a check valve 15, a pressure regulating valve 16, an oil cooler 17, and the like are connected to the hydraulic circuit, similarly to a known active suspension system.

The servo valve 10 provides a control signal issued from an electronic control unit (ECU) 18 to a solenoid 10a via a servo valve driver 19, so that the hydraulic pressure and direction applied to the hydraulic actuator 5 are continuously adjusted. The vehicle body 4 and the piston rod 1 are controlled.
1, a load sensor 20 provided between the vehicle body 4 and the lower suspension arm 3, a sprung acceleration sensor 22 for detecting a vertical acceleration on the vehicle body, and a vertical sensor on the tire side. The signal of the unsprung acceleration sensor 23 for detecting acceleration is controlled based on a signal processed by the first ECU 18 constituting the control means.

The ECU 18 includes a road surface condition detecting unit 27,
An overcutting determination unit 28 as an overcutting determination unit, a target load calculation unit 24, a stabilization calculation unit 25, and a displacement limit comparison calculation unit 26 are provided. Then, an estimated road surface friction coefficient μ generated by the road surface state detection unit operation 27 based on signals from the vehicle speed sensor 29, the steering angle sensor 30, the yaw rate sensor 31, the lateral acceleration sensor 32, and the steering torque sensor 33 provided on the driven wheels. And an over-cut prediction signal or over-cut detection signal output from the over-cut determination unit 28 based on the steering angle signal from the steering angle sensor 30, a signal from the sprung acceleration sensor 22 by the road surface state detection unit 28,
A temporary target load is determined by a target load calculating unit 24 with reference to a signal from the unsprung acceleration sensor 23, and a difference between this value and a signal of the load sensor 20 is processed by a stabilizing calculating unit 25. A command value given to the servo valve driver 19 is adjusted so that the control within the range of the stroke of the actuator 5 is performed by referring to the signal of the stroke sensor 21 in the limit comparison operation unit 26. Then, according to the adjusted command signal, the servo valve 10 is driven so that the target load and the actual load become equal, a stroke is generated in the actuator 5, and the vertical acceleration in a direction to increase the tire contact load is calculated.
It is generated in at least one of the sprung mass and the unsprung mass.

On the other hand, FIG. 2 schematically shows a schematic configuration of a main part of the electric power steering apparatus of the vehicle.
This electric power steering device also serves as an auxiliary steering force torque generator and an auxiliary steering reaction torque generator, and is connected to a steering shaft 42 integrally connected to a steering wheel 41 via a universal joint and a connection shaft. And a rack / pinion composed of a pinion 44 and a rack shaft 48 that can reciprocate in the vehicle width direction by meshing with the pinion 44 and that are connected at both ends to knuckle arms 47 of the left and right front wheels 1 via tie rods 45. A manual steering force generating means 49 composed of a mechanism, and an auxiliary steering force for reducing the steering force by the manual steering force generating means 49 are arranged coaxially with the rack shaft 48, and the rotational force is controlled by an axis. The motor 50 has a built-in ball screw mechanism 50a for converting the force into a force. The electric motor 5 on the basis of the value Tp · θs, etc.
The output of 0 is controlled by the second ECU 51 constituting the control means.

As shown in FIG. 3, the ECU 51 receives an auxiliary steering force torque control system and an auxiliary steering reaction force torque control system. The auxiliary steering force torque control system includes a steering torque sensor 33. Motor 5 based on the output Tp from
An output target value setting section 52 for setting an output torque to be generated to zero, and a motor output control section 53 for controlling the output of the motor 50 based on the output target value generated thereby.
And a motor current detector 54 that detects and feeds back a current flowing through the motor 50. The operation procedure is the same as that of the conventional power steering device, and therefore a detailed description is omitted. The auxiliary steering reaction torque control system includes an output value μ of the road surface state detection unit 27 and a steering angle sensor 3.
0 based on the output value θs
And an additional reaction force setting unit 55 for setting an additional reaction force to be applied to the traction control. The additional reaction force setting unit 55 is also connected to the overcut determination unit 28 so that signals are mutually transmitted and received.

Here, in the road surface state detecting section 27, a vehicle speed sensor 29, a steering angle sensor 30, a yaw rate sensor 31,
Based on the output value from the lateral acceleration sensor 32, the road friction coefficient μ is estimated by a known method with reference to a preset road friction coefficient determination map, for example, and a detailed description thereof will be omitted.

The road friction coefficient μ is estimated from various data when driving the electric motor 50,
Alternatively, a load cell or the like may be provided at an appropriate position between the tie rod 5 and the rack shaft 8 and the reaction force may be obtained from the directly detected reaction force of the rack shaft.

Next, the operation of the apparatus of the present invention will be described with reference to FIG.
Will be described along the flowchart of FIG. First, step 1
In step 4, the respective signals are read from the yaw rate sensor 31, the lateral acceleration sensor 32, the steering angle sensor 30, and the vehicle speed sensor 29, and the road surface condition detection unit 24 detects (estimates) the road surface friction coefficient μ, and determines a predetermined value. From the correspondence map between the road surface friction coefficient μ and the maximum allowable steering angle θmax, the maximum allowable steering angle θm corresponding to the road surface friction coefficient μ at that time.
ax is obtained, and this is compared with the actual steering angle θs, and the overcut determination unit 28 determines whether or not the actual steering angle θs is in an overcut region having a certain width (step 5). When the actual steering angle θs is not included in the over-cut region, the process returns to step 1. When the actual steering angle θs is included in the over-cut region, as shown in FIG.
The calculator 62 calculates the ratio between the maximum allowable steering angle θmax and the actual steering angle θs, and calculates the output θs of the steering angle sensor 30 from this ratio.
Is passed through a differentiator 63 to determine an additional reaction force C with reference to an additional reaction force map 64 determined in consideration of the steering angular velocity θs ′. This is subtracted from the output target value generated by the output target value setting section 51 of the auxiliary steering force control system (FIG. 3). This allows the driver to sense the subtraction of the auxiliary steering force applied to the steering wheel 41 as an increase in steering reaction torque (steering resistance).

Here, in the case of rapid steering with a high steering angular velocity θs', the steering energy due to the inertial force of the steering system becomes large, so that it is not possible to sufficiently suppress overcutting with only a constant additional reaction force. There can be. Therefore, the ratio θ
The relation value of the additional reaction force C with respect to s / θmax is set such that the higher the steering angular velocity θs ′, the higher the rate of increase of the additional reaction force C, so that the steering resistance applied to the steering wheel 41 is increased. An appropriate steering reaction force can be applied to the steering wheel 41.

Next, at step 7, it is determined whether or not the steering wheel is further turned even though the steering reaction torque (steering resistance) is increased. In this determination, the current steering angle and the immediately preceding steering angle may be simply compared. However, as shown by the symbol k in the additional reaction force map 64 of FIG. 5, the additional reaction force C exceeds the predetermined value k. Then, it may be determined that the steering wheel is turned further. If it is determined that the steering wheel is to be turned further by the additional reaction force C, the steering angular velocity θs ′ is considered, and the change in the vehicle behavior with respect to the steering angle becomes small (or no longer changes) at an appropriate timing. Contact load control can be performed.

If it is not determined in step 7 that the steering wheel is being turned further, the process returns to step 1, and if it is determined that the steering wheel is being turned further, it is determined that the steering operation is for avoiding an obstacle, and the process proceeds to step 8. Control to increase the ground contact load of the tire.

Next, the principle of temporarily increasing the contact load will be described. In the model of FIG. 6, M2: sprung mass M1: unsprung mass Z2: sprung coordinates Z1: unsprung coordinates Kt: tire spring constant Fz: actuator thrust, and if the downward direction is the positive direction, the sprung mass M2 and The equations of motion of the unsprung mass M1 are given by the following equations, respectively. However, the ^* mark in the equation represents the first derivative, and the ^** mark represents the second derivative. M2 · Z2 ^** =-Fz M1 · Z1 ^** + Kt · Z1 = Fz

Accordingly, the tire contact load W is given by the following equation. W = −Kt · Z1 = −Fz + M1 · Z1 ^** = M2 · Z2 ^** + M1 · Z1 ^**

That is, since the ground contact load W is the sum of the sprung inertia force and the unsprung inertial force, the expansion / contraction acceleration of the actuator 5 is controlled to at least one of the sprung mass and the unsprung mass. Is changed, the contact load W can be changed. Therefore, by controlling the extension acceleration of the actuator 5, it is possible to temporarily increase the contact load W for each tire. When a 1-ton thrust is generated in the actuator 5 with a suspension stroke of 200 mm, about 0.2
Can be activated for seconds.

For example, as shown in FIG. 7A, when the vehicle is turning, the ground contact load of the front wheels as the steered wheels is larger than that of the rear wheels. (Fig. 7)
(B)) The grip force of the steered wheels is improved, the steering limit is temporarily increased, and the turning performance is improved. Although the illustrated vehicle is a front-wheel drive vehicle, the same applies to a rear-wheel drive vehicle and a four-wheel drive vehicle, though the degree is different.

FIG. 7 conceptually shows the distribution of the contact load (= grip force) of the tire. The contact load in the range of the static load is represented by a solid circle, and the contact load increased by the stroke control of the actuator 5. Is represented by a two-dot chain line circle.

In general, a suspension spring for supporting the weight of the vehicle and a damper for generating damping force are used together in order to save energy consumption of the actuator (see FIG. 8). In this case, Ks: Assuming that the spring constant C is the damping coefficient of the damper, the equations of motion of the sprung mass M2 and the unsprung mass M1 are given by the following equations, respectively. M2 · Z2 ^** + C · (Z2 ^* −Z1 ^* ) + Ks · (Z2-Z1)
= −Fz M1 · Z1 ^** + C · (Z1 ^* −Z2 ^* ) + Ks · (Z1−Z2)
+ Kt.Z1 = Fz

Accordingly, the tire contact load W is given by the following equation. W = −Kt · Z1 = −Fz + M1 · Z1 ^** + C · (Z1 ^* −Z2 ^* ) + Ks · (Z1−Z2) = M2 · Z2 ^** + M1 · Z1 ^**

That is, it is understood that the ground load W can be changed by controlling the expansion and contraction acceleration of the actuator in the same manner as described above.

The actual inertial force of the vehicle is generated not only by the vertical motion but also by the rolling motion and the pitching motion. Here, the rotational movement around each axis passing through the center of gravity of the sprung mass is defined as roll rate: φ pitch rate: θ yaw rate: γ, and the center line in the front-rear direction and the center line in the left-right direction with respect to the center of gravity position The distance to the grounding center is Lf,
Lr, Tf / 2, and Tr / 2 (see FIG. 9), and the thrust of the actuator of each wheel is represented by Fz1 (front left), Fz2 (front right), Fz
Assuming that z3 (rear right) and Fz4 (rear left) and the direction of the force, moment, and coordinate system are as shown in FIG. 10, the rolling moment is: Mx = Tf / 2. (-Fz1 + Fz2) -Tf / 2・ (−Fz3
+ Fz4), and the pitching moment is as follows: My = Lf · (−Fz1−Fz2) −Lr · (−Fz3−Fz4)

Assuming that the rolling moment of inertia is Ix and the pitching moment of inertia is Iy, the rolling moment of inertia is Ixφ ^* = Mx = Tf / 2 · (−Fz1 + Fz2) −Tf / 2 · (−Fz3 + F
z4), and the pitching inertial force is Iyθ ^* = My = Lf · (−Fz1−Fz2) −Lr · (−Fz3−Fz4).

Further, the inertial force of the vertical motion is M2 · Z2 ^** = − Fz1−Fz2−Fz3−Fz4. By controlling at least one of these inertial forces, the rolling motion and the pitching motion are included. By controlling the contact load individually for each tire, the stability during turning is further improved. In addition, the conventional one distributes the load to four wheels, so the rolling inertia force,
The pitching inertia force and the inertia force of the vertical movement do not occur, and these values become zero.

In the above embodiment, the actuator using a hydraulically driven cylinder device as the actuator has been described. This can be achieved by using other electric thrust generating means such as a linear motor or a voice coil, or by using a cam. The same effect can be obtained even if the acceleration is generated using a mechanism or a spring means.

Further, the sensor used can be simplified without departing from the gist of the present invention. For example, sprung,
Since the position detection signal can also be obtained by integrating the output difference of the unsprung acceleration sensors into the second order, the stroke sensor can be eliminated, and the actual measured values of the sprung and unsprung weights and the sprung and sprung weights can be obtained. Since the force generated by the actuator can be obtained by calculating the output values of the lower acceleration sensors, the load sensor can be omitted.
Furthermore, a state estimator can be configured based on signals from the load sensor and the displacement sensor, and both sprung and unsprung accelerations can be obtained indirectly. Furthermore, it goes without saying that the ECU can be realized by any of digital, analog, and hybrid.

[0036]

As described above, according to the present invention, excessive turning of the steering wheel is detected at the time of turning, and an appropriate steering reaction force is generated based on the detection result. And one or both of the unsprung and unsprung accelerations are directly controlled by the thrust generated by the actuator to generate the inertia force of one or both of the unsprung and unsprung masses, and to respond to this. By applying the force to the contact surface, the contact load on the steered wheels is temporarily increased, the generation limit of the grip force of the tire is increased, and the vehicle can turn stably and efficiently. At this time, a sensor for detecting a steering state for generation of a steering reaction force, a determination unit and a sensor for predicting or detecting a steering limit for controlling a ground contact load,
By sharing the judgment means, the structure is simplified, and the number of parts is reduced.

[Brief description of the drawings]

FIG. 1 is a schematic system configuration diagram of an active suspension system for a vehicle to which the present invention is applied.

FIG. 2 is a diagram schematically showing a schematic configuration of a main part of an electric power steering device for a vehicle to which the present invention is applied.

FIG. 3 is a block diagram of an electric power steering device for a vehicle to which the present invention is applied.

FIG. 4 is a flowchart illustrating the operation of the present invention.

FIG. 5 is a block diagram of an additional reaction force setting unit.

FIG. 6 is a model diagram for explaining the principle of the present invention.

FIG. 7A is a conceptual ground load distribution diagram when a conventional front wheel drive vehicle turns, and FIG. 7B is a conceptual ground load distribution diagram when a front wheel drive vehicle to which the present invention is applied.

FIG. 8 is a model diagram of a general active suspension device.

FIG. 9 is an explanatory diagram showing a relationship between a vehicle center of gravity position and a contact position.

FIG. 10 is an explanatory diagram showing a relationship between a force, a moment, and a direction of a coordinate system.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Tire 2 Upper suspension arm 3 Lower suspension arm 4 Body 5 Actuator 6 Piston 7.8 Oil chamber 9 Hydraulic pump 10 Servo valve 11 Piston rod 12 Accumulator 13 Unload valve 14 Oil filter 15 Check valve 16 Pressure control valve 17 Oil cooler Reference Signs List 18 electronic control unit (ECU) 19 servo valve driver 20 load sensor 21 stroke sensor 22 sprung acceleration sensor 23 unsprung acceleration sensor 24 target load calculation unit 25 stabilization calculation unit 26 displacement limit comparison calculation unit 27 road surface state detection unit 28 cut Passing determination unit 29 Vehicle speed sensor 30 Steering angle sensor 31 Yaw rate sensor 32 Lateral acceleration sensor 33 Steering torque sensor 41 Steering wheel 42 Steering shaft 44 Pinion 45 Tie rod 47 Knuckle arm 48 Rack shaft 49 Manual steering force generation means 50 Motor 50a Ball screw mechanism 51 ECU 52 Output target value setting unit 53 Motor output control unit 54 Motor current detection unit 55 Additional reaction force setting unit 62 Computing unit 63 Differentiator 64 Additional counter Force map

Claims (2)

[Claims] An actuator for actively changing a vertical relative distance between a vehicle body and an axle is given a thrust to generate acceleration in at least one of a sprung mass and an unsprung mass, and the acceleration is generated. Ground contact load control means for applying a reaction force of inertia force of at least one of a sprung mass and an unsprung mass to a ground load acting between the tire and the road surface; a means for detecting a steering angle; Means for determining whether or not the steering angle is excessively turned, means for increasing the steering reaction torque, and steering when the steering angle is determined to be too large by the excessive turning determination means. Control means for increasing the contact load of each wheel by the contact load control means if the steering reaction torque is increased and the steering reaction torque is further increased even if the steering reaction torque is increased. Ground contact load control device characterized in that it comprises. 2. The ground load control according to claim 1, wherein the control means determines the turning amount after the steering reaction torque is increased from the increased additional reaction torque. apparatus.