JP6701946B2 - Driving support control device using electric power steering mechanism and vehicle equipped with the same - Google Patents

Description

  The present invention relates to a driving assistance control device that uses an electric power steering mechanism that drives a motor based on a current command value to assist control of a steering system, and a vehicle equipped with the driving assistance control device. Particularly, a driver feels depending on vehicle conditions. The present invention relates to a driving assistance control device that uses an electric power steering mechanism that compensates for a reduction in sense of stability and reduces the load of steering operation during normal steering operation, and a vehicle equipped with the same.

  BACKGROUND ART An electric power steering device that assist-controls a steering system of a vehicle by a rotational force of a motor uses a driving force of a motor to transmit a steering assist force (assist force) to a steering shaft or a rack shaft by a transmission mechanism such as a gear or a belt via a reduction gear. ) Is added. Such a conventional electric power steering device performs feedback control of the motor current in order to accurately generate the torque of the steering assist force. The feedback control adjusts the motor applied voltage so that the difference between the current command value and the motor current detection value becomes small. The adjustment of the motor applied voltage is generally performed by adjusting the duty of PWM (pulse width modulation) control. I am adjusting.

  The general structure of the electric power steering apparatus will be described with reference to FIG. 6b, and further connected to steering wheels 8L, 8R via hub units 7a, 7b. Further, the column shaft 2 is provided with a torque sensor 10 for detecting the steering torque of the steering wheel 1 and a steering angle sensor 14 for detecting the steering angle θ, and a motor 20 for assisting the steering force of the steering wheel 1 is provided with a reduction gear 3. It is connected to the column shaft 2 via. Electric power is supplied from the battery 13 to the control unit (ECU) 30 that controls the electric power steering device, and an ignition key (IG) signal is input via the ignition key 11. The control unit 30 calculates the current command value of the assist (steering assistance) command based on the steering torque Ts detected by the torque sensor 10 and the vehicle speed V detected by the vehicle speed sensor 12, and compensates for the current command value. The current supplied to the motor 20 is controlled by the applied voltage control command value Vref.

  Note that the steering angle sensor 14 is not essential, and may not be provided, and the steering angle can be obtained from a rotation sensor such as a resolver connected to the motor 20.

  A CAN (Controller Area Network) 40 that exchanges various vehicle information is connected to the control unit 30, and the vehicle speed V can also be received from the CAN 40. Further, the control unit 30 can also be connected to a non-CAN 41 other than the CAN 40 that exchanges communication, analog/digital signals, radio waves, and the like.

  The control unit 30 is mainly composed of an MCU (including a CPU, an MPU, etc.), and a general function executed by a program inside the MCU is shown in FIG.

  The function and operation of the control unit 30 will be described with reference to FIG. 2. The steering torque Ts detected by the torque sensor 10 and the vehicle speed V detected by the vehicle speed sensor 12 (or from the CAN 40) are the current command value Iref1. It is input to the current command value calculation unit 31 for calculation. The current command value calculation unit 31 calculates a current command value Iref1, which is a control target value of the motor current supplied to the motor 20, using an assist map or the like based on the input steering torque Ts and vehicle speed V. The current command value Iref1 is input to the current limiting unit 33 via the addition unit 32A, and the current command value Irefm in which the maximum current is limited is input to the subtraction unit 32B, and the deviation ΔI (=I) from the fed back motor current value Im. Irefm-Im) is calculated, and the deviation ΔI is input to the PI (proportional integral) control unit 35 for improving the characteristic of the steering operation. The voltage control command value Vref whose characteristics have been improved by the PI control unit 35 is input to the PWM control unit 36, and the motor 20 is PWM-driven via the inverter 37 as a drive unit. The motor current value Im of the motor 20 is detected by the motor current detector 38 and fed back to the subtractor 32B. The inverter 37 uses a FET (Field Effect Transistor) as a drive element and is configured by a bridge circuit of FET.

  The compensating signal CM from the compensating signal generating unit 34 is added to the adding unit 32A, and the characteristic of the steering system system is compensated by adding the compensating signal CM to improve the convergence and the inertia characteristic. .. The compensation signal generation unit 34 adds the self-aligning torque (SAT) 34-3 and the inertia 34-2 by the addition unit 34-4, and further adds the convergence 34-1 to the addition result by the addition unit 34-5. However, the addition result of the addition unit 34-5 is used as the compensation signal CM.

  Such an electric power steering device can be positioned as a driving support control device and can be utilized as a device that assists the steering operation and reduces the steering operation load.

  Although the effect of reducing the steering driving load in the state where a normal road is normally steered (normal steering driving) is indirect, it is positioned as an important technology for traffic safety along with accident avoidance technology. ing. The steering operation in the normal steering operation is a steering control that follows a target trajectory recognized by the driver, and the steering operation action of the driver in this case is performed by recognition, judgment, and action. Therefore, in the reduction of the steering driving load, it is possible to respond to recognition, judgment, and action, and (1) a power assist characteristic that can reduce the physical burden accompanying the steering action, (2) a vehicle characteristic that facilitates steering control, and (3) a vehicle behavior. Target characteristics and realization of each of the three characteristics (hereinafter collectively referred to as "steering characteristics"), which are steering information characteristics that are easy to grasp from steering information (information related to steering such as steering torque). Means are being discussed. Here, the steering information characteristic refers to a characteristic from information indicating vehicle behavior such as yaw rate (vehicle behavior information) to steering information.

  The electric power steering device plays a role in many vehicles for reducing the physical burden associated with the steering action in (1).

  Regarding the vehicle characteristic (2) in which steering control is easy, discussion is made based on the stability of a characteristic equation which is a pole of a transmission characteristic in which an actual steering angle is input and a vehicle behavior is output. For example, in Non-Patent Document 1, a time constant condition for approximating response characteristics of yaw rate and lateral acceleration with a first-order lag characteristic to obtain a response characteristic with good feeling, and a time constant of steering torque characteristic with good feeling. And the time constant of the response characteristics are presented. In addition, as a means for realizing vehicle characteristics that facilitate steering control, for example, in Non-Patent Document 2, 4WS (four-wheel steering) is used to avoid the constraint that the tire and the steering wheel are mechanically coupled. From the viewpoint of steering response characteristics of yaw rate and lateral acceleration, a vehicle characteristic compensation method using state feedback and a compensation method using control in a reference model have been proposed.

  Regarding the characteristic of (3) that makes it easy to understand the vehicle behavior from the steering information, the relationship between the steering torque and the yaw rate or lateral acceleration is expressed as a Lissajous waveform (sometimes called a Lissajous figure, Lissajous curve, etc.), and its shape and subjective Discussions have been made based on the correlation with various steering feeling evaluations. For example, in Non-Patent Document 3, the relationship between the steering torque and the lateral acceleration is represented by a Lissajous waveform, and the driver's feeling is classified into a feeling of response, a feeling of steering wheel return, and a feeling of phase delay of the steering torque, and characteristics corresponding to each feeling. The value is suggested. In the discussion of means for realizing characteristics that make it easy to grasp vehicle behavior with steering information, the driver recognizes vehicle behavior not only with visual information but also with steering torque information. Analysis shows that there is a vehicle condition in which the driver does not feel a sense of stability even if the vehicle behavior is stable, because the characteristic equation of the vehicle characteristics related to vehicle stability is not included in the transfer characteristics. It is expected that there will be a problem that it is difficult to grasp the vehicle behavior from the steering torque information, and there are studies that suggest this (for example, Non-Patent Document 4).

Hiroaki Fujinami, two others, "Relationship between steering characteristics and driver feeling", Preprint 951 of the Society of Automotive Engineers of Japan, Automotive Engineering Society, May 1995, p. 181-184 Masato Abe, Hiroshi Osawa, "Technology for Improving Motor Performance of Motor Vehicles (Popular Edition)", Asakura Shoten, August 2008, p. 114-125 Hirofumi Sato, 2 others, "Consideration of steering response characteristics related to steering feeling", Proceedings of the Society of Automotive Engineers of Japan, Automotive Engineering Society, March 1990, Vol. 44, No. 3, p. 52-58 Daisuke Yamada, 2 others, "Effects of steering torque characteristics on human-vehicle system", Proceedings of the Society of Automotive Engineers of Japan, Automotive Engineering Society, March 2013, Volume 44, No. 2, p. 459-465

  However, in Non-Patent Document 1 and Non-Patent Document 2, it is easy to understand the technical content of (2) regarding the vehicle characteristics that facilitate steering control, and in Non-Patent Document 3 and Non-Patent Document 4, it is easy to grasp the vehicle behavior of (3) in steering information. Although the technical contents related to the characteristics are presented respectively, they are not the technologies that can deal with both (2) and (3). Regarding the problem (3) that there is a vehicle condition in which the driver does not feel a sense of stability even if the vehicle behavior is stable, the mechanism of occurrence has not been analyzed, and concrete measures have been discussed. Absent. In addition, since this problem is a problem of the interface between the vehicle and the driver, it is thought that it should be compensated by steering originally, but it has been dealt with as a problem of vehicle design in the past, so In particular, this problem has a large impact on vehicles with a low degree of freedom in vehicle design.

  The present invention has been made under the circumstances as described above, and an object of the present invention is to analyze the occurrence mechanism of a phenomenon that the driver does not feel a sense of stability even if the vehicle behavior is stable, and It is an object of the present invention to provide a driving assistance control device using an electric power steering mechanism that compensates for a reduction in the sense of stability that occurs to the driver and reduces the load of steering operation in normal steering operation, and a vehicle equipped with the same.

The present invention relates to a driving assistance control device using an electric power steering mechanism that drives a motor based on a current command value to assist control a steering system. The above object of the present invention is to provide a vehicle characteristic from a steering angle to a yaw rate. A vehicle zero point characteristic compensator for compensating the current command value so as to approximate the first-order lag characteristic, and the current command value so that the time constant of the steering information characteristic from the vehicle behavior information to the steering information becomes a desired value. A steering information characteristic compensating section for compensating and a vehicle damping characteristic compensating section for compensating the current command value by using the angle information so that the damping rate of the vehicle characteristic is approximately 1 are provided. For a vehicle whose cornering compliance Cr satisfies the following formula 1 with respect to the wheel base L, the distance Lf between the front axle center of gravity points, the distance Lr between the rear axle center of gravity points, the weight M, the yaw moment of inertia Ym, and the vehicle speed V: zero characteristic compensation unit vehicle, by compensating for more the current command value to the steering information characteristic compensation unit and said vehicle damping characteristic compensation unit is achieved by reducing the burden of steering operation of the vehicle.

Further, the above object of the present invention is to provide a vehicle zero point characteristic compensator for compensating the current command value so that the vehicle characteristic from the steering angle to the yaw rate approximates to the first-order lag characteristic, and steering from the vehicle behavior information to the steering information. and a steering information characteristic compensator time constants of information characteristic compensating the current command value to a desired value, the rear wheel cornering compliance Cr in normal steering operation, the wheel base L, between the front axle center-of-gravity point distance Lf, the distance Lr between the center of gravity points of the rear axles, the weight M, the yaw moment of inertia Ym, and the vehicle speed V, which satisfy the above-mentioned formula 1 and have a neutral steer characteristic, the vehicle zero-point characteristic compensator and the steering information. by compensating the more the current command value to the characteristic compensation section, it is achieved by reducing the burden of steering operation of the vehicle.

Further, according to the above-mentioned object of the present invention, the vehicle zero-point characteristic compensating unit has a characteristic that the torsion bar and the vehicle zero-point characteristic compensating unit have a pole for canceling the zero point of the vehicle characteristic, and is equivalent to the vehicle characteristic pole. By performing pole-zero cancellation so as to have a zero point of, by approximating the vehicle characteristic to a first-order lag characteristic, or by the vehicle behavior information being a yaw rate and the steering information being a steering torque, Alternatively the steering information characteristic compensator, by compensating the current command value using the self aligning torque, or the vehicle damping characteristic compensation unit, by using the steering angle as the angle information, some have the When the steering frequency is about 0.5 Hz or less, the normal steering operation is more effectively achieved.

Further, the above object is achieved by a vehicle equipped with a driving support control device using the electric power steering mechanism.

  According to the driving support control device using the electric power steering mechanism of the present invention, the vehicle characteristics are substantially reduced by utilizing the analysis result of the occurrence mechanism of the phenomenon that the driver does not feel the stability even if the vehicle behavior is stable. With a damping factor of 1, by compensating the current command value so that the time constant of the steering information characteristic becomes a desired value so as to approximate the first-order lag characteristic, the degree of freedom in vehicle design as in a small or medium-sized car It is possible to improve the sense of stability felt by the driver and reduce the load of the steering operation in the normal steering operation even in a vehicle with a small number of vehicles. Further, by mounting the above-mentioned driving support control device on a vehicle, it is possible to improve the sense of stability and travel with a reduced load.

It is a block diagram which shows the outline of an electric power steering device. It is a block diagram showing an example of composition of a control unit (ECU) of an electric power steering device. It is a block diagram which shows the simulation model which can reproduce a steering characteristic. It is a characteristic view which shows the power assist characteristic used with a simulation model. It is a Lissajous waveform which shows steering torque characteristics at the time of simulating steering operation using a simulation model. (A) when the vehicle speed is 80 km/hour and the steering frequency is 0.3 Hz, (B) when the vehicle speed is 80 km/hour and the steering frequency is 0.5 Hz, and (C) is when the vehicle speed is 120 km/hour and the steering frequency is Is 0.3 Hz, (D) is a Lissajous waveform when the vehicle speed is 120 km/hour and the steering frequency is 0.5 Hz. It is a bode diagram which shows the change of the gain and phase of a SAT transfer function with respect to a steering frequency. (A) is a Bode diagram showing a change in gain, and (B) is a Bode diagram showing a change in phase. It is a block diagram which shows the structural example (1st Embodiment) of this invention. It is an image figure showing the appearance of the torque generated between the road surface and steering. It is a block diagram showing a configuration example of state feedback control for adjusting a yaw rate attenuation rate. (A) is a block diagram showing a configuration example when a yaw rate is fed back, and (B) is a block diagram showing a configuration example equivalent to (A). It is a block diagram which shows the structural example for performing SAT estimation. (A) is a block diagram showing a configuration example when SAT estimation is performed using an external observer, and (B) is a block diagram showing a configuration example equivalent to (A) when the SAT estimated value and the actual SAT match. It is a figure. 5 is a flowchart showing an operation example (first embodiment) of the present invention. It is a schematic diagram showing a run course used in order to confirm the effect of the present invention. 5 is a graph showing changes in yaw rate and steering torque with and without the compensation of the present invention. (A) is a graph when compensation is not performed, and (B) is a graph when compensation is performed. It is a block diagram which shows the structural example (2nd Embodiment) of this invention. 7 is a flowchart showing an operation example (second embodiment) of the present invention. It is a block diagram which shows the outline of the vehicle carrying the driving assistance control device, 4WS, and variable gear ratio mechanism of the present invention. It is a block diagram which shows the structural example (3rd Embodiment) of this invention.

  In the present invention, among the steering characteristics, it is assumed that the power assist characteristics that can reduce the physical burden due to the steering action have already been realized, and the steering characteristics that facilitate steering control and the steering characteristics that facilitate the understanding of the vehicle behavior from the steering information. The information characteristic is realized by using an electric power steering mechanism (EPS). Among these, regarding the steering information characteristic, the driver also uses the steering torque information in the recognition of the vehicle behavior, so that the characteristic of the steering torque with respect to the vehicle behavior (hereinafter referred to as “steering torque characteristic”) is mainly targeted. There is.

  In achieving the above, the target characteristic of the steering torque characteristic is that it can be approximated by a first-order transmission characteristic having a time constant suitable for the driver characteristic in normal steering operation. This is because the steering torque information plays an important role in the on-center region and the off-center region, which are normal steering driving scenes, but the power assist amount is adjusted to increase outside the off-center region, and as a result, the steering torque with respect to the vehicle behavior is adjusted. The sensitivity is low, the relative importance of the steering torque information to the visual information and the steering angle information is low, and the steering torque characteristics are reproducible, and the vehicle behavior is simple to understand. Since it means that the desired amount of vehicle behavior occurs at the expected timing after the driver feels the steering torque, it is presumed to be appropriate as a target. It The yaw rate is used as the vehicle behavior information indicating the vehicle behavior. Further, as a vehicle characteristic, a characteristic of a yaw rate with respect to a steering angle (hereinafter, referred to as "yaw rate transmission characteristic") is used. The vehicle behavior information may use lateral acceleration instead of yaw rate.

  In order for the steering torque characteristics to reach the above target characteristics, it is necessary to take measures against the problem that there are vehicle conditions in which the driver does not feel a sense of stability even if the vehicle behavior is stable. Therefore, the Lissajous waveforms of yaw rate and steering torque are measured using a simulation model that can reproduce the steering characteristics, and the conditions under which the driver feels stable are derived based on the analysis results. The derivation of the condition will be described below.

The simulation model is modeled including a vehicle model, a steering mechanism model, and an EPS controller, and has a configuration as shown in FIG. In FIG. 3, θd is a steering angle, θg is a steering angle, θt is an actual steering angle, γ is a yaw rate, T _tor is a steering torque detection value, T _stg is a steering torque, T _sat is a self-aligning torque (SAT), K. _Tor is the torsion bar rigidity, STG(s) is the steering dynamic characteristic, g _tot is the total gear ratio, C ₁ (s) is the feed forward characteristic of the EPS controller, C ₂ (s) is the feedback characteristic of the EPS controller, and G ₂ is the feedback characteristic of the EPS controller. _θ (s) is the yaw rate characteristic with respect to the actual steering angle θt (hereinafter referred to as “actual steering angle yaw rate transfer characteristic”), G _γ (s) is the SAT characteristic with respect to the yaw rate γ, and the simulation model uses K _tor A torsion bar 100, a characteristic section 101 having C ₁ (s), a characteristic section 102 having C ₂ (s), a characteristic section 103 having STG(s), a gain section 104 having a gain of 1/g _tot , G _{It is} composed of a characteristic section 105 having _θ (s), a characteristic section 106 having G _γ (s), an addition/subtraction section 108, and subtraction sections 107 and 109. The characteristic unit 101 has the power assist characteristic as shown in FIG. 4 in addition to the function as the third-order phase compensator, and the steering torque detection value T _tor is converted into the power assist torque by the power assist characteristic. After that, phase compensation is performed. The power assist characteristic corresponds to the assist map used by the current command value calculation unit 31 in FIG. 2, but is a characteristic that does not depend on the vehicle speed here.

  The parameters of the vehicle model used in this simulation are shown in Table 1 below. The identified vehicle is a small passenger car with an engine displacement of 1.2 liters and a vehicle with understeer characteristics with a stability factor of 0.0017. is there.

The steering angle θd that changes in a sweep sine wave in the range of 10° to 90° at the vehicle speeds of 80 km/hour and 120 km/hour is input to the above simulation model at steering frequencies of 0.3 Hz and 0.5 Hz, respectively. FIG. 5 shows a Lissajous waveform of the steering torque characteristic (characteristic of the steering torque T _stg with respect to the yaw rate γ) in the case. The reason for using 0.3 Hz and 0.5 Hz as the steering frequency is as follows. That is, it is known that the steering operation controllability of following the target trajectory is important in the normal steering operation, and that the follow-up control with respect to the target trajectory has a frequency characteristic upper limit due to the ability adaptation of the driver. For example, according to the crossover theory, the driver corrects the vehicle dynamic characteristics and adapts to the follow-up control response of about 0.6 Hz, and there is also a report of setting it to 0.4 Hz. Therefore, it is assumed that the steering frequency in the normal steering operation is approximately 0.5 Hz or less, and 0.3 Hz and 0.5 Hz which are within the range are used.

  In FIG. 5, FIG. 5(A) shows a case where the vehicle speed is 80 km/hour and the steering frequency is 0.3 Hz, and FIG. 5(B) shows a case where the vehicle speed is 80 km/hour and the steering frequency is 0.5 Hz. 5) shows a Lissajous waveform when the vehicle speed is 120 km/hour and the steering frequency is 0.3 Hz, and FIG. 5D shows a Lissajous waveform when the vehicle speed is 120 km/hour and the steering frequency is 0.5 Hz. As can be seen from FIGS. 5A and 5B, at a vehicle speed of 80 km/hour, the hysteresis width at the steering frequency of 0.3 Hz is larger than that at the steering frequency of 0.3 Hz. Further increases with the amplitude of the yaw rate. Since the yaw rate in the off-center region is approximately 1 to 15°/sec, the steering torque gain with respect to the yaw rate in the range of ±10°/sec in the off-center region decreases as the amplitude of the yaw rate increases. Therefore, from this result, it is presumed that the yaw rate is difficult to be grasped by the steering torque. Further, looking at the simulation results at a vehicle speed of 120 km/hour shown in FIGS. 5C and 5D, at a vehicle speed of 120 km/hour, the hysteresis width near the yaw rate near zero becomes narrower than at a vehicle speed of 80 km/hour. It can be seen that the hysteresis width is further narrowed as the yaw rate amplitude and the steering frequency increase. It is a well known phenomenon that the hysteresis width becomes narrower at high speeds and the steering feel becomes unreliable, but the yaw rate attenuation rate in this simulation is about 0.7, which is a stable category. That is, even if the vehicle characteristics are stable and the vehicle behavior is stable, (a) the hysteresis width changes with respect to the yaw rate amplitude or the steering frequency, and (b) the hysteresis width near zero of the yaw rate according to the vehicle speed. It can be seen that when the phenomenon of "narrowing" occurs, the sense of stability felt by the driver may be lost. Therefore, the factors of these two phenomena will be described.

First, assuming that the dynamic characteristics of the steering are constant, when SAT is used as a substitute characteristic of steering torque and a planar two-wheel model using a linear tire model is used as a vehicle model, SAT T _sat (for yaw rate γ(s) The transfer function of s) (hereinafter referred to as “SAT transfer function”) is given by the following mathematical expression 2.

Here, ξ is a trail (caster rail), Yf(s) is a front tire lateral force, θt(s) is an actual steering angle, M is a vehicle weight, Ym is a yaw moment of inertia, Kr is a rear tire cornering power, and Lr is Distance between rear axle center of gravity points, L is wheel base, Cr is rear cornering compliance, T ₁ is time constant, and s is Laplace operator. Formula 2 is a basic formula of a two-wheel plane model (see, for example, Masato Abe, Hiroshi Osawa, “Technology for Improving Motor Performance of Motor Vehicles (Popular Edition)”, Asakura Shoten, August 2008, pp. 21-36), On the other hand, the following equations 3 and 4 that approximately represent the equations of motion in the lateral direction and the yaw direction, and SAT is the moment around the kingpin that is obtained by multiplying the front tire lateral force by the trail. It is calculated by applying the represented one.

Here, α(s) is lateral acceleration, Yr(s) is rear tire lateral force, and Lf is distance between front axle center of gravity points. The natural frequency ω _{zn of} the zero point, the damping ratio ζ _z of the zero point, and the natural frequency ω _{pn of the} pole of the SAT transfer function of the mathematical formula 2 are expressed by the following mathematical formulas 6, 7 and 8, respectively.

FIG. 6 is a Bode diagram showing the number 2 of the SAT transfer function. FIG. 6 shows changes in the gain and phase of the SAT transfer function when the steering frequency is changed at vehicle speeds of 60 km/hour, 80 km/hour, 100 km/hour and 120 km/hour, and FIG. The change in gain is shown, and FIG. 6B shows the change in phase. According to FIG. 6, in the region where the steering frequency is 0.5 Hz or less, the gain of the SAT transfer function increases as the vehicle speed increases, and the phase lead amount decreases. Then, at a vehicle speed of 120 km/hour, a phase delay characteristic is shown between 0.2 and 1 Hz, and as a result, the yaw rate amplitude and the steering frequency as shown in FIGS. It is presumed that there is a phenomenon that the hysteresis width is narrowed. This phase delay characteristic is a result of the natural frequency ω _{pn of the} pole expressed by the formula 8 being lower than the natural frequency ω _{zn of the} zero expressed by the formula 6. This is because the decrease in the amount of phase advance of the SAT transfer function due to the vehicle speed occurs even if the attenuation rate ζ _z of the zero point expressed by the equation 7 decreases according to the vehicle speed, but it does not lead to the phase delay characteristic. . As can be seen from Expression 6, the natural frequency ω _zn of the zero point is a constant obtained from the specifications of the vehicle, and becomes higher as the rear axle cornering compliance Cr becomes smaller. Further, as can be seen from Equation 8, the natural frequency ω _pn of the pole is an inverse function of the vehicle speed with the reciprocal of the rear-axle cornering compliance Cr as a coefficient, and decreases as the vehicle speed increases, and the rear-axle cornering compliance Cr Becomes larger, the frequency becomes lower than the natural frequency ω _zn of the zero point at a low vehicle speed, and the SAT transfer function partially has the phase delay characteristic. That is, even if the stability of the vehicle characteristics is ensured, the hysteresis width becomes narrower according to the vehicle speed, and the states (a) and (b) described above are obtained. Therefore, in order for the driver to feel a sense of stability, it is necessary to satisfy the following expression 9 in the normal steering operation, and for that purpose, the rear axle cornering compliance Cr may be a small value that satisfies the following expression 10. You will need it.

As a condition for the driver to feel a sense of stability, a condition that the rear axle cornering compliance Cr is satisfied has been derived. Therefore, assuming that this condition is satisfied, “the steering torque characteristic is the normal steering characteristic is the target characteristic of the steering torque characteristic described above. The conditions for achieving "" can be approximated by a first-order transfer characteristic having a time constant suitable for the driver characteristic in driving.

  First, the conditions under the assumption that the vehicle has the neutral steer characteristic will be described. The neutral steer characteristic can be defined as a characteristic in which the yaw rate attenuation rate becomes approximately 1 regardless of the vehicle speed, and under that definition, it is the characteristic of the yaw rate γ(s) with respect to the actual steering angle θt(s). It is assumed that a certain actual steering angle yaw rate transfer characteristic can be approximated to the first-order lag characteristic represented by the following equation 11 in the normal steering operation (steering frequency 0.5 Hz or less) (hereinafter, the assumption that this approximation can be made is “Assumption 1”). ")).

Here, τ _n is the time constant of the pole. Further, the transmission characteristic of the lateral acceleration α(s) with respect to the actual steering angle θt(s) (hereinafter, referred to as “actual steering angle lateral acceleration transmission characteristic”) is the same as the actual steering angle yaw rate transmission characteristic expressed by the following formula 12. It is assumed that the represented first-order phase characteristic can be approximated (hereinafter, the assumption that this can be approximated is “Assumption 2”).

Here, τ _α is the time constant of the zero point. Under the assumption 1 and the assumption 2, the SAT transfer function is obtained from the expressions 2, 5, 11, and 12, and the following expression 13 is obtained, and the steering torque characteristic substituted by the SAT becomes the primary transfer characteristic.

Here, τ _sat and τ ₁ are time constants, and τ _sat is calculated by the following _Expression 14.

τ ₁ is a time constant set for expressing the equation 13 by a proper transfer function (a transfer function in which the numerator order is less than or equal to the denominator order), and a value corresponding to 10 Hz is given.

If the steering torque characteristic becomes the first-order transmission characteristic, then the time constant τ _sat shown in _Formula 14 is set so as to match the driver characteristic (hereinafter, this setting is referred to as “time constant setting”). The steering torque characteristic can be made the target characteristic.

  As described above, when the rear-axle cornering compliance Cr in the normal steering operation is sufficiently small so as to satisfy the expression 10, the vehicle having the yaw rate attenuation factor of approximately 1 such as the neutral steer characteristic does not depend on the vehicle speed is assumed. If the assumption 2 is satisfied and the time constant can be set, the steering torque characteristic becomes the target characteristic.

  When the vehicle has the neutral steering characteristic, the steering torque characteristic can be set to the target characteristic if the above condition is satisfied. However, when the vehicle has the understeering characteristic, the yaw rate attenuation rate changes depending on the vehicle speed. Vehicle speed is limited. Therefore, for a vehicle with an understeer characteristic, it is necessary to compensate so that the yaw rate attenuation rate becomes approximately 1, but this compensation (hereinafter referred to as "attenuation rate compensation") is performed by adjusting the specifications of the vehicle. Is difficult. This is because when the stability factor K is expressed as a function of the rear-axis cornering compliance Cr, the following formula 15 is obtained. As can be seen from the formula 15, when the rear-axis cornering compliance Cr is set small, the stability factor K becomes large. This is because the understeer characteristic is strengthened.

Here, Kf is the front tire cornering power, and Cf is the front wheel cornering compliance. On the other hand, if the front wheel cornering compliance Cf is made small, there is a problem that the general vehicle becomes sensitive to shimmy and brake judder. Therefore, even with this method, it is difficult to perform damping rate compensation. Therefore, the damping function is compensated by using the steering function.

  In the steering function, in addition to the damping rate compensation, compensation that satisfies Assumption 1 and Assumption 2 (hereinafter referred to as “first-order approximation compensation”) and compensation for setting a time constant (hereinafter referred to as “time constant compensation”). Also). That is, the steering function performs damping rate compensation, first-order approximation compensation and time constant compensation when the vehicle has an understeer characteristic, and first-order approximation compensation and time constant compensation when the vehicle has a neutral steer characteristic. By performing these compensations, the steering torque characteristic becomes the target characteristic, and the actual steering angle yaw rate transfer characteristic becomes a simple characteristic with a damping rate of approximately 1 and is approximated to a first-order lag characteristic, and is linked to the actual steering angle. Since the yaw rate transfer characteristic, which is the characteristic of the yaw rate with respect to the changing steering angle, is also a simple characteristic, the steering characteristic is easily controlled.

  In the present invention, a compensator is provided for each of the above three compensations (attenuation rate compensation, first-order approximation compensation, time constant compensation). Specifically, a vehicle damping characteristic compensating section for damping rate compensation, a vehicle zero point characteristic compensating section for first-order approximation compensation, and a steering information characteristic compensating section for time constant compensation are prepared. However, if the vehicle has a neutral steer characteristic, the vehicle damping characteristic compensator is not necessary. Each compensation is performed by compensating the current command value. By performing such compensation, when the rear axle cornering compliance Cr in the normal steering operation is small, a vehicle characteristic that facilitates the steering control and a steering information characteristic that allows the vehicle behavior to be easily grasped by the steering information are realized. Even in a vehicle with a low degree of freedom in vehicle design, it is possible to improve the driver's sense of stability and reduce the load of steering operation.

  Embodiments of the present invention will be described below with reference to the drawings. In the following embodiments, it is assumed that the target vehicle has a small rear axle cornering compliance Cr in the normal steering operation and satisfies the expression (10).

  First, an embodiment (first embodiment) for a vehicle having an understeer characteristic will be described.

  FIG. 7 shows a configuration example of the first embodiment. FIG. 7 shows a part of a configuration example of the driving assistance control device according to the present invention. The compensation current command value Irefa output from the subtraction unit 151 is, for example, the addition unit 32A shown in FIG. 2 and the subsequent configuration is the same as the configuration shown in FIG. The current command value calculation unit 31 also has the same configuration as the current command value calculation unit 31 shown in FIG. Description of the same configuration as that of FIG. 2 is omitted.

The SAT estimation unit 110 estimates SAT T _sat from the steering torque Ts, the assist torque Tm, the motor angular velocity ωm, and the motor angular acceleration αm, and outputs it as a SAT estimated value T _se . The vehicle zero point characteristic compensating unit 120 receives the current command value Iref1, performs first-order approximation compensation, and outputs the compensation current command value Irefz. The steering information characteristic compensation unit 130 outputs a compensation signal Cs1 for time constant compensation using the SAT estimated value T _se . The vehicle damping characteristic compensating unit 140 outputs the compensation signal Cs2 for compensating the damping rate using the steering angle θg. The compensation current command value Irefz is compensated by the compensation signals Cs1 and Cs2 in the adder 150 and the subtractor 151, respectively, and is output as the compensation signal Irefa.

  Each part will be described in detail.

The SAT estimation unit 110 estimates SAT T _sat by a process similar to the process disclosed in Japanese Patent No. 4192442, for example. The outline of the processing will be described.

  The state of the torque generated from the road surface to the steering is as shown in FIG. 8, and the steering torque Ts is generated by the driver steering the steering wheel, and the assist torque Tm of the motor is generated in accordance with the steering torque Ts. To occur. As a result, the wheels are steered and SAT is generated as a reaction force. At that time, a torque that is a resistance to steering the steering wheel is generated by the inertia J and the friction (static friction) Fr of the motor. From the balance of these forces, the equation of motion as shown in the following Expression 16 is obtained.

Here, when Laplace transform is performed with the above-mentioned equation 16 as an initial value of zero and T _sat is solved, the following equation 17 is obtained.

As can be seen from the above equation 17, by obtaining the inertia J of the motor and the static friction Fr as constants in advance, the SAT T _sat can be estimated from the motor angular velocity ωm, the motor angular acceleration αm, the assist torque Tm, and the steering torque Ts. it can. The motor angular velocity ωm and the motor angular acceleration αm can be calculated from the rotation angle detected by the rotation angle sensor by connecting a rotation angle sensor (not shown) such as a resolver to the motor 20.

If the estimated SAT is fed back as it is, the steering may become too heavy and the steering feeling may not be improved. Therefore, the SAT estimated value T _se is signal-processed by using a filter having a frequency characteristic, and steering is performed. You may make it output only the necessary and sufficient information for improving a feeling. The SAT may be estimated by a method other than this method.

  The vehicle damping characteristic compensating section 140 calculates a compensation signal Cs2 for performing damping rate compensation, that is, compensation for setting the yaw rate damping rate to approximately 1.

The yaw rate attenuation rate of the understeer vehicle can be adjusted by state feedback control in which the yaw rate γ(s) is detected and is fed back to the actual steering angle θt(s) as shown in FIG. 9(A). . In FIG. 9A, K _STG is the steering system rigidity excluding the torsion bar, and C _γ (s) is the characteristic of the state feedback control. In order to realize this control by the EPS function, a feedback characteristic C ₂ (s) that equivalently replaces the configuration shown in FIG. 9A with the configuration controlled by the EPS shown in FIG. 9B is used. The configuration of FIG. 9B has the same configuration as a part of the simulation model shown in FIG. 3 with STG(s)=1/K _STG .

The actual steering angle yaw rate transfer characteristic G _θ (s) before the state feedback control is added is represented by the following _Expression 18.

Here, ζ _n is the yaw rate attenuation rate. T _yr is a time constant and is calculated by the following equation 19.

The characteristic C _γ (s) of the state feedback control that gives the yaw rate attenuation rate is given by the following Expression 20.

Here, Kd is a feedback gain.

  From Expressions 18 to 20, the characteristic in the configuration shown in FIG. 9A is as shown in Expression 21 below. As can be seen from Expression 21, since the attenuation rate corresponding to the feedback gain Kd is added, Expression 21 is added. The yaw rate attenuation rate can be set to approximately 1 by setting the feedback gain Kd such that is an approximate expression like the following Expression 22.

The feedback characteristic C ₂ (s) that gives a characteristic equivalent to that of the configuration shown in FIG. 9A is equal to the following equations 23 and 24 obtained from the respective configurations shown in FIGS. 9A and 9B. From the following equation 25 derived from the conditions, the following equation 26 is obtained.

Therefore, by using the feedback gain Kd that satisfies the approximate expression of the above Expression 22 as the feedback gain Kd in Expression 26, the yaw rate attenuation rate can be made approximately 1. Using this feedback characteristic C ₂ (s), the vehicle damping characteristic compensator 140 calculates the compensation signal Cs2 from the steering angle θg.

  The vehicle zero-point characteristic compensator 120 performs first-order approximation compensation, that is, compensation in which Assumption 1 and Assumption 2 are satisfied, but the assumption 2 is that the actual steering angle lateral acceleration transfer characteristic can be approximated to the first-order phase characteristic. Is presumed to be a valid assumption at a steering frequency of 0.5 Hz or less, which corresponds to the normal steering operation, and therefore the vehicle zero point characteristic compensating section 120 performs compensation satisfying Assumption 1. Since the natural frequency of the zero point of the actual steering angle lateral acceleration transfer characteristic is around 2 Hz, which is larger than the pole, without depending on the vehicle speed, the approximation to the first-order phase characteristic means that the second-order pole is 0.5 Hz or less. This is because it works in the direction of reducing the amount of phase delay, and is therefore assumed to be a reasonable approximation. Also in Assumption 1, the fact that the actual steering angle yaw rate transfer characteristic approximates to the first-order lag characteristic is the yaw rate transfer characteristic that is the characteristic of the yaw rate with respect to the steering angle that changes in conjunction with the actual steering angle (that is, the vehicle characteristic. ) Is equivalent to approximating the first-order lag characteristic, the yaw rate transfer characteristic is approximated to the first-order lag characteristic.

The approximation of the yaw rate transfer characteristic to the first-order delay characteristic is performed by canceling the natural frequency of the zero point to the pole-zero. This can be done by using a feedforward controller that has a pole that cancels the zero of the yaw rate transfer characteristic and has a zero that has the same time constant as the pole of the yaw rate transfer characteristic. In the EPS configuration, the characteristics of the closed loop including the torsion bar 100 and the characteristic unit 101 in the simulation model shown in FIG. 3 play the role of the feedforward controller. Therefore, the steering angle θg between the steering torque _{T stg} is assumed to be proportional below 0.5 Hz, when the proportionality coefficient _{K sat,} the following equation 28 so as to satisfy the following equation 27 Feedforward characteristics _C 1 of (s) By determining the time constants T ₃ and T ₄ , pole-zero cancellation can be performed.

Here, K _C is a steady gain and K _at is a power assist gain, and STG(s)=1/K _{STG in} FIG. The above assumption corresponds to the assumption that the actual steering angle and the SAT are proportional to each other at 0.5 Hz or less, and is therefore assumed to be a valid assumption.

  When the yaw rate transfer characteristic is obtained from the equations 22 and 28, it can be seen that the first-order lag characteristic is obtained as shown in the following equation 29 and the purpose is achieved.

The vehicle zero point characteristic compensator 120 compensates the current command value Iref1 using the feedforward characteristic C ₁ (s).

The steering information characteristic compensating unit 130 calculates a compensation signal Cs1 for performing time constant compensation, that is, compensation for setting the time constant τ _sat shown in _Expression 14 so as to match the driver characteristic. This is achieved by filtering the SAT estimate.

SAT estimation can be performed using an external observer, and the configuration when performing SAT estimation using an external observer is as shown in FIG. In FIG. 10A, C(s) is the characteristic of the control system including the torsion bar, F(s) is the filter characteristic, Q(s) is the characteristic of the disturbance observer compensator, and STGn(s) is the nominal dynamic characteristic. Is. If the SAT estimated value T _se in FIG. 10A and the actual SAT T _sat match, the configuration shown in FIG. 10A can be equivalently exchanged with the configuration shown in FIG. 10B. Then, the condition of the filter characteristic F(s) for the time constant τ _sat to become a desired time constant τ _r that _{matches the} driver characteristic becomes the following Expression 30, and when the Expression 13 is substituted into the Expression 30, the filter characteristics F(s) is the following Expression 31.

Further, since the estimated disturbance value only needs to match within a range of about 10 Hz or less, the characteristic Q(s) of the disturbance observer compensator is set to about 1 at 10 Hz or less. Therefore, the steering information characteristic compensation unit 130 calculates the compensation signal Cs1 from the SAT estimated value T _se using the filter characteristic F(s).

  With the above configuration, an operation example of the first embodiment will be described with reference to the flowchart of FIG.

  When the operation starts, the vehicle speed V, the steering torque Ts, the motor angular velocity ωm, the motor angular acceleration αm, the assist torque Tm, and the steering angle θg are input (step S10). Of these, the vehicle speed V is supplied to the current command value calculation unit 31, the steering torque Ts is supplied to the current command value calculation unit 31 and the SAT estimation unit 110, and the motor angular velocity ωm, the motor angular acceleration αm, and the assist torque Tm are supplied to the SAT estimation unit 110. The steering angle θg is input to the vehicle damping characteristic compensation unit 140. The motor angular acceleration αm is calculated by differentiating the motor angular velocity ωm.

  As in the case of the configuration shown in FIG. 2, the current command value calculation unit 31 calculates the current command value Iref1 from the vehicle speed V and the steering torque Ts (step S20), and outputs it to the vehicle zero point characteristic compensation unit 120.

The SAT estimation unit 110 calculates the SAT estimated value T _se , which is the estimated value of SAT T _sat , from Equation 17 using the steering torque Ts, the motor angular velocity ωm, the motor angular acceleration αm, and the assist torque Tm (step S30). The estimated value T _se is input to the steering information characteristic compensation unit 130. The operations of the current command value calculation unit 31 and the SAT estimation unit 110 may be switched in order or may be executed in parallel.

The vehicle zero-point characteristic compensating unit 120, to which the current command value Iref1 is input, converts the current command value Iref1 using Formula 26 and outputs it as the compensation current command value Irefz (step S40). An appropriate value is set in advance for the power assist gain K _at in Expression 26, and a value adjusted in advance to satisfy Expression 25 is set for the time constants T ₃ and T ₄ . The compensation current command value Irefz is input to the addition unit 150.

The steering information characteristic compensating unit 130, to which the SAT estimated value T _se is input, converts the SAT estimated value T _se using Formula 29 and outputs the compensation signal Cs1 (step S50). A previously determined value is set for the time constant τ _sat in the equation 29, and a desired value is set for the time constant τ _r . The compensation signal Cs1 is input to the adder 150. The operations of the vehicle zero point characteristic compensating unit 120 and the steering information characteristic compensating unit 130 may be switched in order or may be performed in parallel.

  The compensation current command value Irefz input to the adder 150 is compensated by adding the compensation signal Cs1 (step S60) and output as the compensation current command value Ireft. The compensation current command value Ireft is added and input to the subtraction unit 151.

The vehicle damping characteristic compensating unit 140, to which the steering angle θg is input, converts the steering angle θg using Equation 24 and outputs the compensation signal Cs2 (step S70). The values calculated in advance in G _θ (0), K _STG , time constant τ _n , and yaw rate damping ratio ζ _n in the equation 24 are set, and the approximation formula of the equation 22 is established in the feedback gain Kd. The value adjusted in advance is set. The compensation signal Cs2 is subtracted and input to the subtraction unit 151.

  The compensation current command value Ireft input to the subtraction unit 151 is compensated by subtracting the compensation signal Cs2 (step S80) and output as the compensation current command value Irefa.

  Since the steering angle and the motor angle have a proportional relationship, the vehicle damping characteristic compensator 140 may use the motor angle instead of the steering angle. Further, the steering information characteristic compensation unit 130 may use the steering torque instead of the SAT estimated value. In this case, it is necessary to perform conversion considering the dynamic characteristics of the steering.

  Here, the effect of the present embodiment will be described.

  In order to confirm the effects of the three types of compensation (attenuation rate compensation, first-order approximation compensation, and time constant compensation) performed in this embodiment, a vehicle with an understeer characteristic equipped with the exercise support control device of this embodiment is used. The lane change operation was performed only with or without compensation without changing the static power assist characteristics, and the convergence with respect to the target trajectory was compared. Convergence at the time of changing lanes is known because the steering operation load increases as convergence worsens and overshoot occurs. The specifications of the vehicle used are as shown in Table 2 below. The lane change running was performed at a vehicle speed of 80 km/hour, at which the highest yaw rate of about 15°/sec appeared in the off-center region, and a substitute characteristic of convergence to the target trajectory was obtained. The overshoot of the yaw rate was evaluated by giving the course constraint condition shown in FIG.

Before driving, the driver gave informed consent for the purpose of evaluation.

  The evaluation result is shown in FIG. FIG. 13A shows the result when no compensation was performed, and FIG. 13B shows the result when compensation was performed. The vertical axis represents yaw rate and steering torque, and the horizontal axis represents time. From FIG. 13, it can be seen that when no compensation is performed, an overshoot is observed immediately before the yaw rate converges, but when compensation is performed, no overshoot occurs. It is presumed that this is because the phase characteristic of the steering torque with respect to the yaw rate has an effect. Without compensation, it is difficult to recognize the yaw rate from the steering torque information because the phase difference between the yaw rate and the steering torque changes depending on the steering state, and steering is performed more than necessary, resulting in overshoot. It is supposed to be. On the other hand, when compensation is performed, the phase relationship between the steering torque and the yaw rate during steering is almost constant, and the driver could control the yaw rate with the steering torque information, and as a result, it is speculated that overshoot did not occur. To be done. Thus, it was confirmed that the three compensations according to the present embodiment suppressed the yaw rate overshoot and reduced the steering operation load.

  Next, an embodiment (second embodiment) for a vehicle having a neutral steer characteristic will be described.

  FIG. 14 shows a configuration example of the second embodiment. Similar to the case of FIG. 7, FIG. 14 shows a part of the configuration example of the driving assistance control device according to the present invention. As described above, in the vehicle having the neutral steer characteristic, the yaw rate attenuation rate is approximately 1, and it is not necessary to perform the attenuation rate compensation. Therefore, the second embodiment is different from the first embodiment shown in FIG. The vehicle damping characteristic compensator 140 is not provided. Therefore, the steering angle θg is not input, and the compensation current command value Ireft output from the addition unit 150 is input to the addition unit 32A illustrated in FIG. 2 instead of the compensation current command value Irefa.

  FIG. 15 shows an operation example of the second embodiment. Compared with the operation example of the first embodiment shown in FIG. 11, the operation in the vehicle damping characteristic compensation unit 140 (step S70) and the attenuation rate compensation (step S80) by the compensation signal Cs2 which is the operation result are unnecessary. Is becoming Other operations are the same as the operation example of the first embodiment.

  In the above-described embodiments (the first embodiment and the second embodiment), all the compensation is performed by using the EPS, but other systems and mechanisms mounted on the vehicle perform some compensation. It is also possible for the EPS to perform the remaining compensation. For example, the damping rate compensation and the first-order approximation compensation can be executed by 4WS (four-wheel steering), a variable gear ratio mechanism, SBW (Steer By Wire), etc., and the time constant compensation can be executed by SBW or the like. The 4WS is a system that steers not only the front two wheels but also the rear two wheels when changing the direction of the vehicle, and it is possible to suppress the phase change in the yaw rate response. And first-order approximation compensation can be performed. The variable gear ratio mechanism is a mechanism that appropriately controls the steering gear ratio according to the vehicle speed, and when used in combination with EPS, it becomes possible to automatically adjust the yaw rate transmission characteristic that is the vehicle characteristic, and the damping ratio Compensation and first-order approximation compensation can be performed. The SBW is a system for transmitting an operation of the steering wheel by an electric signal, and since it is possible to perform a steering system control with a high degree of freedom, it is possible to execute attenuation rate compensation, first-order approximation compensation and time constant compensation. However, since SBW is generally used as a substitute for EPS, when it is used together with EPS, the SBW is configured to execute only the compensation shared by the SBW. For example, as shown in FIG. 16, in addition to the driving assistance control device 70 according to the present invention, a 4WS mechanism is incorporated in the front wheel steering gear box 50 and the rear wheel steering gear box 51, and the variable gear ratio is set. In the vehicle in which the mechanism 60 is installed adjacent to the motor 20, the variable gear ratio mechanism 60 or 4WS performs the damping rate compensation and the first-order approximation compensation, and the driving support control device 70 performs only the time constant compensation. It is possible. The configuration example (third embodiment) of the embodiment of the present invention in this case is as shown in FIG. 17, which is different from the first embodiment shown in FIG. 7 in the vehicle zero point characteristic compensating section 120 and the vehicle damping characteristic compensating section. There is no 140. Then, the current command value Iref1 output from the current command value calculation unit 31 is compensated by adding the correction signal Cs1 output from the steering information characteristic compensation unit 130 in the addition unit 152, and the compensated current command value Ireft. Then, the compensation current command value Irefft' is output as', and is input to the adder 32A shown in FIG.

1 Handle 2 Column shaft (steering shaft, handle shaft)
10 torque sensor 12 vehicle speed sensor 20 motor 30 control unit (ECU)
31 current command value calculation unit 110 SAT estimation unit 120 vehicle zero point characteristic compensation unit 130 steering information characteristic compensation unit 140 vehicle damping characteristic compensation unit

Claims (8)

In a driving assistance control device using an electric power steering mechanism for assisting a steering system by driving a motor based on a current command value,
A vehicle zero point characteristic compensator for compensating the current command value so that the vehicle characteristic from the steering angle to the yaw rate approximates the first-order lag characteristic;
A steering information characteristic compensator for compensating the current command value so that the time constant of the steering information characteristic from the vehicle behavior information to the steering information becomes a desired value,
As the attenuation factor of said vehicle characteristics is substantially 1, and a vehicle damping characteristic compensation unit for compensating the current command value by using the angle information,
The rear wheel cornering compliance Cr in the normal steering operation is as follows with respect to the wheel base L, the distance between the front axle center of gravity points Lf, the distance between the rear axle center of gravity points Lr, the weight M, the yaw moment of inertia Ym, and the vehicle speed V.
The vehicle satisfying, the vehicle zero characteristic compensation unit, by compensating for more the current command value to the steering information characteristic compensation unit and said vehicle damping characteristic compensator, reducing the burden of steering operation of the vehicle A driving support control device using an electric power steering mechanism. In a driving assistance control device using an electric power steering mechanism for assisting a steering system by driving a motor based on a current command value,
A vehicle zero point characteristic compensator for compensating the current command value so that the vehicle characteristic from the steering angle to the yaw rate approximates the first-order lag characteristic;
And a steering information characteristic compensator time constants of the steering information characteristics of the vehicle motion information to steer information to compensate the current command value to a desired value,
The rear wheel cornering compliance Cr in the normal steering operation is as follows with respect to the wheel base L, the distance Lf between the front axle center of gravity points, the distance Lr between the rear axle center of gravity points, the weight M, the yaw moment of inertia Ym, and the vehicle speed V.
The filled, to a vehicle having a neutral steering characteristics by compensating the more the current command value to the vehicle zero characteristic compensation unit and the steering information characteristic compensation unit, to reduce the burden of steering operation of the vehicle A driving support control device using a characteristic electric power steering mechanism. The vehicle zero point characteristic compensation unit,
The characteristic by the torsion bar and the vehicle zero point characteristic compensator has a pole for canceling the zero point of the vehicle characteristic, and by performing pole zero cancellation so as to have a zero point equivalent to the pole of the vehicle characteristic,
The driving assistance control device using the electric power steering mechanism according to claim 1, wherein the vehicle characteristic is approximated to a first-order lag characteristic. The vehicle behavior information is a yaw rate,
The driving assistance control device using the electric power steering mechanism according to claim 1, wherein the steering information is steering torque.   The driving assistance control device using the electric power steering mechanism according to claim 4, wherein the steering information characteristic compensating unit compensates the current command value by using a self-aligning torque.   The driving assistance control device using the electric power steering mechanism according to claim 1, wherein the vehicle damping characteristic compensator uses a steering angle as the angle information. If steering frequency is less than approximately 0.5 Hz, the driving support control apparatus using an electric power steering mechanism according to any of claims 1 to 6, the normal steering operation. A vehicle equipped with a driving assistance control device using the electric power steering mechanism according to any one of claims 1 to 7 .