JPH08207542A - Vehicle turning motion controller - Google Patents

Abstract

PURPOSE: To balance the turning characteristic of a vehicle at a high level by estimating the cornering force of front and rear wheels and the slipping angles thereof, obtaining a dynamic stability factor from the result of these estimations and controlling a steering characteristic so as to set the factor to a minimum value. CONSTITUTION: In an estimating means 25 provided in an electronic control unit 23, yaw angle acceleration, a yaw rate and vehicle lateral acceleration are calculated from the outputs of the lateral acceleration sensors 19a and 19b of front and rear wheels and a car body slipping angle is calculated from these calculated values and outputs from a steering angle sensor 21 and a car speed sensor 22. Followingly, the slipping angles of the front and rear wheels and the cornering force thereof are calculated. A dynamic stability factor is calculated from the obtained calculated values by a calculating means 26 and a steering characteristic is variably controlled by a steering characteristic variable control means 27 so as to set the factor to a minimum value. That is, when a yawing moment in the front wheel side is large, oversteering control is performed by determining the start of turning.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle turning motion control device, and more particularly to a technique for improving steering stability during turning.

[0002]

2. Description of the Related Art Steering characteristics such as understeer, oversteer, neutral steer, and reverse steer are factors that represent the turning characteristics of a vehicle. Generally, it is said that a neutral steer in which the turning radius does not change when the vehicle speed is changed while the steering wheel turning angle is constant is preferable.

The steering characteristics are influenced by the grip characteristics of the tire and the load distribution of the vehicle body, but the influence of the roll rigidity of the suspension occupies a considerable weight. JP-A-62-198511 proposes a technique for neutralizing the steering characteristics by positively changing the hydraulic or pneumatic actuators arranged between the wheel and each wheel by electronically controlling it. ing.

[0004]

In the above-mentioned known technique, the lateral acceleration sensors arranged in front of and behind the vehicle body detect the rate of change in yaw rate (yaw angular acceleration), and the turning of the vehicle is started depending on whether the yaw rate is positive or negative. Is determined to be the end of turning, and the front and rear roll rigidity distribution ratios are controlled based on the determination. According to this, since the start and the end of the turning are determined only by the yaw rate change rate, the yaw rate is increasing in both directions even though different controls must be performed at the start of turning and during the rapid deceleration during turning. However, it is impossible to determine which control is required at that time. In other words, in the above-mentioned known technology, since the control is performed using only the yaw rate obtained from the lateral acceleration sensor, there is the aspect that the motion state of the vehicle cannot be accurately evaluated, so the performance of the active suspension device can be fully utilized. It has the drawback of not breaking.

The present invention has been devised in order to eliminate such disadvantages of the prior art, and its main purpose is to:
It is an object of the present invention to provide a turning motion control device capable of accurately grasping the turning behavior of a vehicle and further improving the turning response and stability of the vehicle.

[0006]

According to the present invention, a turning state detecting means for detecting a turning state of a vehicle, a cornering force for front and rear wheels, and a cornering force for estimating slip angles of the front and rear wheels are provided. .Slip angle estimating means, dynamic stability factor calculating means for obtaining a dynamic stability factor from the estimated cornering force and slip angle, and steering characteristic variable for controlling steering characteristics so that the dynamic stability factor value is minimized. The present invention is achieved by providing a turning motion control device for a vehicle, the control device having a control means.

[0007]

According to this structure, since the relationship between the front and rear wheels slip angles is taken into consideration, the dynamic turning behavior can always be accurately grasped, and the turning characteristics of the vehicle can be further improved. It can be balanced in dimensions.

[0008]

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

The steering stability of the vehicle is determined by the front-rear balance of the lateral force (cornering power) generated by the tire. A stability factor (SF) is an amount for evaluating this balance, but this cannot represent nonlinearity or dynamic characteristics. Therefore, in the present invention, a dynamic stability factor (DSF), which is a dynamically expanded stability factor (SF), is introduced.

First, the stability factor (SF) is
It is expressed as follows using the cornering powers of the front wheels Wf and the rear wheels Wr.

[0011]

[Equation 1]

Here, the cornering power of the front and rear wheels is expressed by the following equation from the linearity at a minute slip angle by the cornering force of the front and rear wheels and the slip angle of the front and rear wheels.

[0013]

[Equation 2]

On the other hand, the extended cornering power before and after theoretical extension to the non-linear region is defined as follows.

[0015]

(Equation 3)

The extended stability factor (SF) 'in which the stability factor (SF) is redefined by using the extended cornering powers of the front and rear wheels is expressed as follows.

[0017]

[Equation 4]

Further, the cornering force of the front and rear wheels is 0.
In consideration of the continuity of the vehicle and the matching of the signs when turning left and right, the dynamic stability factor (DSF) is defined as the product of the extended stability factor (SF) ′ and the cornering force of the front and rear wheels. The stability factor (DSF) is expressed by the following equation.

[0019]

(Equation 5)

On the other hand, the turning behavior of the vehicle is expressed as follows by the yawing moment around the center of gravity generated by the front and rear tires and the magnitude of the slip angle.

[0021]

[Table 1]

From the above, if the cornering force and the slip angle of the front and rear wheels are known, the turning behavior of the vehicle can be accurately grasped. Therefore, from the estimated cornering force and slip angle of the front and rear wheels, the front and rear around the center of gravity can be obtained. It can be seen that the neutral steer characteristic can always be obtained if the yawing moment difference of is minimized, that is, the absolute value of the dynamic stability factor (DSF) is minimized.

Next, a method for estimating the cornering force and the slip angle of the front and rear wheels will be described. Assuming that the influence of the longitudinal force generated on the tire in the Y translation and Z axis rotation motion is ignored, the equation of motion is given by the following equation.

[0024]

(Equation 6)

Therefore, by detecting the lateral acceleration and the yaw angular acceleration, the cornering force of the front and rear wheels can be estimated from the following equation.

[0026]

(Equation 7)

Next, the slip angles of the front and rear wheels are given by the following equation.

[0028]

(Equation 8)

If lateral acceleration sensors are installed near the front and rear wheels of the vehicle body, the yaw angular acceleration, yaw rate, and vehicle lateral acceleration are given by the following equations.

[0030]

[Equation 9]

FIG. 2 shows a schematic structure of an active suspension device to which the present invention is applied. In this device, the suspension devices respectively arranged between the front, rear, left and right wheels of the four-wheeled vehicle and the vehicle body are individually controlled. However, since the suspension devices have the same configuration,
Only the suspension of the right front wheel Wfr will be described below as an example.

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 compression coil spring 5 and a hydraulic actuator 6 are provided between the lower suspension arm 3 and the vehicle body 4 in parallel.

The hydraulic actuator 6 is of a cylinder / piston type, and is provided with a multi-valve unit 10 and a servo valve 11 for a lower oil chamber 8 and an upper oil chamber 9 of the piston 7 inserted in the cylinder. The operating oil pressure is supplied from the variable displacement hydraulic pump 12 via the. The hydraulic oil pumped up from the tank 13 and discharged from the pump 12 is adjusted to a predetermined constant pressure by the multi-valve unit 10, and then the main accumulator 14
The pump pulsation is removed by and the hydraulic pressure controlled by the servo valve 11 is supplied to the piston upper chamber 9 and the piston lower chamber 8. Since there is an area difference between the pressure receiving surfaces on both sides of the piston 7 by the sectional integration of the piston rod 15, the piston lower chamber 8
And a piston upper chamber 9 and a pressure difference between the pressure receiving surface of the piston 7 acting on the piston rod 15 generate a thrust force, which causes the piston rod 15 to make a reciprocating linear motion in the vertical direction. The relative distance between the vehicle 1 and the vehicle body 4 is changed. Further, a second accumulator 16 is connected to a pipe line leading to the piston lower chamber 8 so as to supply an auxiliary hydraulic pressure so that the transient response of the actuator 6 is not insufficient.

The servo valve 11 is composed of a known 3-position 4-port linear solenoid valve so that the hydraulic pressure and flow rate applied to the hydraulic actuator 6 may be continuously changed in proportion to the current value applied to the solenoid. The load sensor 17 provided at the connecting portion between the vehicle body 4 and the piston rod 15, the vehicle body 4 and the lower suspension arm 3
A stroke sensor 18 provided between the vertical acceleration sensor 2 and a lateral acceleration sensor 19 provided at a proper position on the vehicle body.
0, the steering angle sensor 21, the vehicle speed sensor 22, and other various signals are processed by the electronic control unit 23 having a built-in computer to determine the vehicle behavior, and optimal control of the opening degree and direction is performed. Further, the electronic control unit 23 includes a hydraulic pressure sensor 2 provided at an appropriate position in the hydraulic circuit.
The signal of 4 is also input.

Now, in the estimation means 25 in the electronic control unit 23, the yaw angular acceleration, yaw rate,
And the lateral acceleration of the vehicle are obtained, and those values and the steering angle sensor 21
The vehicle body slip angle is obtained from the detected value of the vehicle speed sensor 22 and the detected value of the vehicle speed sensor 22, and the slip angles of the front and rear wheels are obtained based on them. Furthermore, the cornering forces of the front and rear wheels are obtained from the equation (7). These values are input to the (DSF) calculation means 26 to obtain the dynamic stability factor (DSF), and the steering characteristic variable control means 2 is set so that this value becomes the minimum.
The steering characteristic is variably controlled at 7 (see FIG. 3).

In this embodiment, the slip angles of the front and rear wheels and the cornering force of the front and rear wheels are estimated based on the signals of the front and rear lateral acceleration sensors 19.
As is clear from the equation (9), it is possible to obtain these from the yaw rate.

In the steering characteristic variable control means 27, the yawing moment on the front wheel side is larger (LfFf>
LrFr), that is, when the absolute value of the front wheel slip angle is larger when the yaw rate (γ) is increasing (βf> βr),
When it is determined that the vehicle is turning, oversteer control is performed. When the absolute value of the rear wheel slip angle is larger (βf <βr), it is determined that the tendency is spin and understeer control is performed. Also, when the yaw moment on the rear wheel side is larger (LfFf
<LrFr), that is, when the absolute value of the front wheel slip angle is larger when the yaw rate is in the decreasing direction (βf> βr), it is determined to be a drift-out tendency, and oversteer control is performed.
When the absolute value of the rear wheel slip angle is larger (βf <βr)
Determines that the turning has ended and performs understeer control.

When performing such steering characteristic variable control in the active suspension device, the front-rear distribution ratio of the moving amount by which the vertical load of the tire moves laterally during cornering, that is, the front-rear roll rigidity distribution ratio λ is set. This is done by intentionally changing it. That is, if the roll rigidity distribution ratio λ is increased to increase the load movement of the front wheel Wf, the slip angle of the front wheel Wf increases and the steering wheel tends to understeer. On the contrary, the roll rigidity distribution ratio λ is decreased to reduce the rear wheel Wr. If the load movement is increased, the slip angle of the front wheel Wf decreases, and oversteering tends to occur. In other words, if the front suspension is made relatively hard by controlling the hydraulic pressure supplied to the hydraulic actuators 6 provided on each wheel, on the other hand, if the rear suspension is made relatively hard, it becomes oversteer. .

In order to continuously change the roll rigidity front-rear distribution ratio λ by the lateral force in the steering characteristic variable control means 27,
Set λ as follows using appropriate parameters. λ = λ0-κ (DSF) where λ0: initial roll rigidity distribution ratio κ: control gain

In this way, understeer control is performed by increasing the roll rigidity distribution ratio above the initial value, and conversely, oversteer control is performed by decreasing the roll rigidity distribution below the initial value.

A comparison with the conventional vehicle by the dynamic stability factor control as described above will be described below.

FIGS. 4 and 5 show the characteristics of a conventional vehicle (dotted line / the same applies below) and a vehicle to which the present invention applies (solid line / the same applies below) when the throttle is turned off from the steady circular turn. . It can be seen from FIG. 4 that the vehicle to which the present invention is applied suppresses the generation of yaw rate to a high deceleration region at the time of deceleration due to the throttle off, and from FIG. 5, the vehicle to which the present invention is applied is accompanied by the throttle off. It can be seen that changes in yaw rate and lateral acceleration are suppressed during deceleration. That is,
Since the dynamic stability factor (DSF) becomes negative due to deceleration accompanying throttle-off, the peak yaw rate and yaw angular acceleration generated by increasing the roll rigidity distribution ratio λ are suppressed to a low level, and maneuverability in the limit region is improved. You can see that it will improve.

Regarding the additional steering response from the low-speed steady circular turning, as shown in FIG. 6, the vehicle to which the present invention is applied has a yaw rate gain of about 7 by reducing the roll rigidity distribution ratio λ. %, It can be seen that the understeer tendency in the low speed range is improved.

With respect to yaw rate damping at high speeds, as shown in FIG. 7, it can be seen that the gain resonance point is eliminated and the vehicle according to the present invention has a characteristic closer to neutral.

As for the steering amount and the vehicle trajectory at the time of lane change, the present invention is applied as shown in FIG. 8 due to the decrease of the roll rigidity distribution ratio λ at the beginning of turning and the increase of the roll rigidity distribution ratio λ at the end of turning. It can be seen that the overshoot is reduced and the drivability is improved in the car compared to the conventional car.

FIG. 9 shows, as another example to which the present invention is applied, a schematic configuration of drive wheels of a front-wheel drive vehicle. A transmission M is connected to an engine E mounted laterally on a vehicle body, and a transmission force is transmitted to a main differential D to an output shaft of the transmission M, that is, an input shaft 31 of a differential. Input gear 3
2 are provided.

The main differential device D includes the input gear 3 of the input shaft 31.
A ring gear 3 having an outer toothed gear 33 meshing with 2 on the outer circumference
4, an inner planetary gear 37 that meshes with the sun gear 35 that is coaxial with the ring gear 34, an outer planetary gear 36 that meshes with the ring gear 34, and an inner planetary gear 37 that meshes with the sun gear 35. It is composed of a carrier 38. Then, the planetary carrier 38 is connected to the right front wheel Wfr via the right drive shaft 39,
The sun gear 35 is connected to the left front wheel Wfl via the left drive shaft 40.

On the left drive shaft 40 side, a torque distribution device T for distributing the input torque from the ring gear 34 to the left and right drive shafts 39 and 40 at a predetermined ratio is provided. In this torque distributor T,
The planetary gear 42 provided on the planetary carrier 41 coupled to the left drive shaft 40 is
It meshes with a sun gear 43 which is rotatably supported at 0, and also meshes with a ring gear 44 arranged on the outer periphery of the planetary carrier 41. An external gear 45 formed integrally with the planetary carrier 38 of the main differential device D
And the external gear 46 formed on the ring gear 44 of the torque distribution device T are integrally formed as a pair of pinion 47.4.
8 meshed with each. In this way, the main differential device D and the torque distribution device T are connected to each other.

A known axial piston type variable displacement hydraulic pump 49 is driven by a pinion 50 that meshes with an external gear 33 formed integrally with a ring gear 34 of the main differential device D. The hydraulic pump 49 is connected to the hydraulic motor 53 via a pair of oil passages 51 and 52, and the pinion 5 provided on the output shaft of the hydraulic motor 53 is connected to the hydraulic motor 53.
4 meshes with an input gear 55 integrally formed with the sun gear 43 of the torque distributor T.

In the capacity adjusting lever 56 for driving the swash plate of the hydraulic pump 49, both ends of the steering gear 57, which moves in the left-right direction of the vehicle body in conjunction with the operation of the steering wheel, are provided with a pair of Bowden wires 58, 59. It is connected. As a result, the discharge amount of the hydraulic pump 49 becomes 0 when the steering wheel is in the neutral state, and when the steering wheel is steered in a certain direction, the steering angle and the rotation speed of the pinion 50 at that time, that is, the vehicle speed A large amount of pressure oil is discharged from the hydraulic pump 49 to either one of the oil passages 51 and 52.

By controlling the hydraulic pressure of the oil passages 51 and 52 by the hydraulic control means 60, the rotational speed of the hydraulic motor 53 can be set to an appropriate value.

The hydraulic control means 60 includes an engine rotation speed sensor 61 for detecting the rotation speed of the engine E and the engine E.
Based on a predetermined program, signals from an engine torque sensor 62 for detecting the torque of the vehicle, a steering angle sensor 63 for detecting the steering angle, a yaw rate sensor 64 for detecting the actual yaw rate of the vehicle, and a vehicle speed sensor 65 for detecting the vehicle speed are based on a predetermined program. It is controlled by the electronic control unit 66 which performs arithmetic processing.
The electronic control unit 66 includes an estimation unit, a (DSF) calculation unit, and a steering characteristic variable control unit, as in the above.

The drive shaft torque is obtained by obtaining the engine torque from the intake pressure and the engine rotation speed, and multiplying the engine torque by the gear ratio obtained from the engine rotation speed and the vehicle speed. When the drive shaft torque is obtained, the left / right wheel distribution torque to be generated by the hydraulic motor 53 is calculated by the left / right distribution torque calculating means provided in the steering characteristic variable control means. Thereby, the drive shaft torque can be distributed to the left and right wheels at a predetermined ratio.

When the vehicle is traveling straight ahead, the amount of oil discharged from the hydraulic pump 49 is 0, so the hydraulic motor 53 is held in a stopped state. Therefore, the sun gear 43 of the torque distribution device T connected to the pinion 54 of the hydraulic motor 53 via the input gear 55 is also fixed. At this time, the main differential D
The planetary carrier 38 and the planetary carrier 41 of the torque distributor T are a ring gear 44 and an external gear 4.
6, the pinion 48, the pinion 47, and the external gear 45 are interlockingly connected at a predetermined gear ratio, so that the rotational speeds of both planetary carriers 38 and 41, that is, the main differential D
The rotational speeds of the planetary carrier 38 and the sun gear 35 are forcibly matched, and the right front wheel Wfr and the left front wheel Wfl rotate at the same speed.

When a steering wheel (not shown) is steered, pressure oil is discharged from the hydraulic pump 49, and the hydraulic motor 53 rotates in a predetermined rotation direction and rotation speed.
Then, the sun gear 43 of the torque distribution device T rotates, and a predetermined difference occurs in the rotation speeds of the planetary carriers 38 and 41, that is, the rotation speeds of the planetary carrier 38 and the sun gear 35 of the main differential device D. Thus, the engine torque transmitted from the transmission M to the ring gear 34 of the main differential device D is distributed to the left and right wheels at a predetermined ratio determined by the rotation direction and rotation speed of the hydraulic motor 53. .

When the steering characteristic variable control is performed using the torque distribution device T, the torque distribution of the left and right wheels during cornering is intentionally changed. That is, if the torque of the wheels on the inside of the turning circle is made relatively large, there is an understeer tendency, and conversely, if the torque of the wheels on the outside of the turning circle is made relatively large, there is an oversteer tendency. In this way, the hydraulic control unit 60 changes the hydraulic pressure applied to the hydraulic motor 53 to change the torque distribution of the left and right drive wheels according to the driving condition of the vehicle at that time, whereby the maneuverability can be arbitrarily changed. it can.

[0057]

As described above, according to the present invention, various motion states of the vehicle can be accurately expressed and optimal control can be executed in accordance therewith, so that a dramatic improvement in motion performance can be achieved. it can. Further, this control also unifies the responsiveness of the vehicle, so that the drivability is also improved.

[Brief description of drawings]

FIG. 1 is an explanatory view of cornering force and slip angle.

FIG. 2 is a schematic configuration diagram of an active suspension device to which the present invention is applied.

FIG. 3 is a schematic control block diagram of the present invention.

FIG. 4 is a characteristic comparison diagram of a vehicle to which the present invention is applied and a conventional vehicle showing a relationship between lateral acceleration and yaw rate when the throttle is turned off from a steady circular turn.

FIG. 5 is a characteristic comparison diagram of a vehicle to which the present invention is applied and a conventional vehicle, showing the rate of change in yaw rate and lateral acceleration.

FIG. 6 is a characteristic comparison diagram showing a yaw rate response characteristic of a vehicle to which the invention is applied and a conventional vehicle.

FIG. 7 is a characteristic comparison diagram showing a yaw rate damping characteristic of a vehicle to which the invention is applied and a conventional vehicle.

FIG. 8 is a characteristic comparison diagram of a vehicle to which the present invention is applied and a conventional vehicle showing response characteristics at the time of lane change.

FIG. 9 is a schematic configuration diagram of drive wheels of a front-wheel drive vehicle to which the present invention is applied.

[Explanation of symbols]

 1 Wheel 2 Lower Suspension Arm 3 Upper Suspension Arm 4 Vehicle Body 5 Compression Coil Spring 6 Hydraulic Actuator 7 Piston 8 Lower Oil Chamber 9 Upper Oil Chamber 10 Multi-valve Unit 11 Servo Valve 12 Hydraulic Pump 13 Tank 14 Main Accumulator 15 Piston Rod 16 No. 2 Accumulator 17 Load sensor 18 Stroke sensor 19 Lateral acceleration sensor 20 Vertical acceleration sensor 21 Steering angle sensor 22 Vehicle speed sensor 23 Electronic control unit 24 Hydraulic pressure sensor 25 Estimating means 26 DSF computing means 27 Steering characteristic variable control means 31 Input shaft 32 Input gear 33 External gear 34 Ring gear 35 Sun gear 36 Outer planetary gear 37 Inner planetary gear 38 Planetary carrier 39/40 Drive shaft 41 Planetary carrier 42 planetary gear 43 sun gear 44 ring gear 45 external gear 46 external gear 47/48 pinion 49 hydraulic pump 50 pinion 51/52 oil passage 53 hydraulic motor 54 pinion 55 input gear 56 capacity adjusting lever 57 steering gear 58/59 Bowden wire 60 hydraulic Control means 61 Engine rotation speed sensor 62 Engine torque sensor 63 Steering angle sensor 64 Yaw rate sensor 65 Vehicle speed sensor 66 Electronic control unit E Engine M Transmission D Main differential device T Torque distribution device

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Tsukasa Fukusato 1-4-1, Chuo, Wako-shi, Saitama Inside Honda R & D Co., Ltd. (72) Inventor Masashige Osaki 1-4-1, Chuo, Wako, Saitama No. Incorporated in Honda R & D Co., Ltd. (72) Inventor Katsuji Watanabe 1-4-1 Chuo, Wako, Saitama Incorporated in Honda R & D Co., Ltd.

Claims (4)

[Claims] 1. A turning state detecting means for detecting a turning state of a vehicle, a cornering force / slip angle estimating means for respectively estimating a cornering force of front and rear wheels and a slip angle of front and rear wheels, and an estimated cornering force and slip angle. A turning stability control of a vehicle, comprising: a dynamic stability factor calculation means for obtaining a dynamic stability factor from the steering wheel; and a steering characteristic variable control means for controlling a steering characteristic so as to minimize the dynamic stability factor value. apparatus. 2. When the yaw moment on the front wheel side is larger, when the absolute value of the front wheel slip angle is larger, it is judged that turning has started and oversteer control is performed, and the absolute value of the rear wheel slip angle is larger. When the yaw moment on the rear wheel side is larger, the understeer control is performed, and when the absolute value of the front wheel slip angle is larger, it is determined to be the drift-out tendency and oversteer control is performed. 2. The vehicle turning motion control device according to claim 1, wherein when the absolute value of the rear wheel slip angle is larger, it is determined that the turning has ended and the understeer control is performed. 3. The understeer control is performed by increasing the front / rear roll rigidity distribution ratio above an initial value, and the oversteer control is performed by decreasing the front / rear roll rigidity distribution ratio below the initial value. 1. The vehicle turning motion control device according to 1. 4. The understeer control is performed by relatively increasing the driving force distribution inside the turning circle, and the oversteering control is performed by relatively increasing the driving force distribution outside the turning circle. The turning motion control device for a vehicle according to claim 1.