JP5970942B2 - Vehicle system vibration control device - Google Patents

Description

  The present invention relates to a vehicle system vibration control device that suppresses an estimated sprung behavior of a vehicle body during driving by correction control of driving torque.

  2. Description of the Related Art Conventionally, there is known a vehicle vibration suppression control device that adds disturbance torque estimated from wheel speed fluctuation to a driver input, estimates vehicle body vibration from a vehicle model, and controls wheel torque to suppress vehicle body vibration. (For example, refer to Patent Document 1).

JP 2009-127456 A

  However, the conventional apparatus described in Patent Document 1 has a problem in that the steering response is not expected to be improved in a turning scene because behavior due to steering is not included in the control target.

  The present invention has been made paying attention to the above problem, and provides a vehicle system vibration control device capable of ensuring the calculation accuracy of the turning resistance force regardless of the change of the driving condition in the turning scene where the driving condition changes. The purpose is to do.

In order to achieve the above object, the vehicle system vibration control device of the present invention is a wheel which is an input format to a vehicle model used when estimating the sprung behavior of the vehicle body based on sensing information from the vehicle acquired during traveling. an input converter for converting the dimension of the torque or force applied to the vehicle body vibration estimation unit that estimates a sprung mass behavior of the vehicle using the vehicle model and the torque or force applied to the wheel, the estimated result of the sprung mass behavior And a torque command value calculation unit for correcting the drive torque based on the above.
In this vehicle system vibration control device,
The input converter is
A first turning resistance calculation unit that calculates a first turning resistance using a steering angle signal, a vehicle speed signal, and a linear two-wheel model;
A second turning resistance force calculation unit for calculating a second turning resistance force using a three-dimensional map of a steering angle signal, a vehicle body speed signal, and a vehicle speed / steering angle / turning resistance force;
A switching processing unit for switching the turning resistance force to be output to the vehicle body vibration estimation unit based on the calculation results from the two turning resistance force calculation units and the running conditions;
Is provided.

For example, when one control logic that calculates the turning resistance force using the steering angle, the vehicle speed, and the linear two-wheel model is used, the calculation accuracy of the turning resistance force is ensured in a vehicle speed range of medium speed or higher where the model approximation is high. The However, the calculation accuracy of the turning resistance force is low in the low speed range where the model closeness decreases.
In contrast, the first turning resistance force calculation unit calculates the first turning resistance force using the steering angle signal, the vehicle body speed signal, and the linear two-wheel model, and the second turning resistance force calculation unit calculates the steering angle signal and A second turning resistance force is calculated using a three-dimensional map of the vehicle speed signal and the vehicle speed / steering angle / turning resistance force. Then, based on the two calculation results and the running condition, the turning resistance output to the vehicle body vibration estimation unit is switched in the switching processing unit. That is, the calculation method for ensuring the calculation accuracy of the turning resistance force can be assigned to the second turning resistance force calculation unit with respect to the traveling condition in which the calculation accuracy of the first turning resistance force calculation unit is low. it can.
As a result, in a turning scene where the driving condition changes, when the driving condition is such that the calculation accuracy by one of the turning resistance force calculation units is low, ensuring the calculation accuracy is shared by the other turning resistance force calculation unit. The calculation accuracy of the turning resistance force can be ensured regardless of the change of the rotation.

1 is an overall system configuration diagram showing an engine vehicle to which a vehicle system vibration control device of Embodiment 1 is applied. It is a control block diagram which shows the control program structure in the engine control module in the engine vehicle system of Example 1. FIG. It is a control block diagram which shows the vehicle system vibration control apparatus in the engine control module of Example 1. FIG. 6 is a schematic diagram showing that the tire is displaced in the front-rear direction when the suspension strokes in the description of the suspension stroke calculation unit of the first embodiment. It is a front-wheel tire characteristic view which shows an example of the stroke of the suspension in the suspension stroke calculation part which has in the input conversion part of Example 1, and the front-back direction displacement relation characteristic of a front-wheel tire. It is a rear-wheel tire characteristic figure which shows an example of the stroke of the suspension in the suspension stroke calculation part which has in the input conversion part of Example 1, and the longitudinal direction displacement relationship characteristic of a rear-wheel tire. It is a front-wheel turning resistance map figure which shows an example of the three-dimensional map which calculates the front-wheel turning resistance force in the 2nd turning resistance calculation part which has in the input conversion part of Example 1. FIG. It is a rear-wheel turning resistance map figure which shows an example of the three-dimensional map which calculates the rear-wheel turning resistance force in the 2nd turning resistance calculation part which has in the input conversion part of Example 1. FIG. FIG. 6 is a first weighting factor map diagram showing a transition characteristic of the rate of change of the weighting factor with respect to the vehicle speed when the switching processing unit included in the input conversion unit of the first embodiment changes from low speed to medium speed. FIG. 6 is a second weighting factor map diagram showing transition characteristics of the change rate of the weighting factor with respect to the vehicle speed when changing from medium speed to low speed in the switching processing unit included in the input conversion unit according to the first embodiment. FIG. 6 is a weighting factor variation characteristic diagram illustrating a variation characteristic in which weighting factor retention is added when a torque command value increases when the switching processing unit included in the input conversion unit of Example 1 changes from low speed to medium speed. It is a vehicle model figure which shows what represented the vehicle model which has in the vehicle body vibration estimation part of Example 1 graphically. It is a block diagram which shows the structure of the 1st-3rd regulator part which has in the torque command value calculation part of Example 1, the 1st-3rd tuning gain setting part, and an adder. It is a gain function explanatory drawing which shows the function which each regulator gain set to the 1st-3rd regulator part which has in the torque command value calculation part of Example 1 exhibits. It is explanatory drawing of the basic effect | action of vehicle structure vibration control, and shows the driving condition (a), the time chart (b) of an axle torque characteristic, and the time chart (c) of a pitch angular velocity characteristic. It is a flowchart which shows the flow of the vehicle structure vibration control process performed in the engine control module of Example 1. FIG. It is a principle explanatory view showing basic principles of “improvement of steering response”, “suppression of load fluctuation”, and “suppression of roll speed”, which are the effects aimed at the vehicle system vibration control of the first embodiment. It is a logic block diagram which shows the logic detail of the vehicle structure vibration control of Example 1. FIG. Pitch rate (no control) / steering input / control command value (= drive torque command value) / pitch rate (after control) representing the effect realized during steering in the engine vehicle equipped with the vehicle system vibration control device of the first embodiment It is a time chart showing contrast characteristics of yaw rate (after control) and roll rate (after control).

  Hereinafter, the best mode for realizing a vehicle system vibration control device of the present invention will be described based on Example 1 shown in the drawings.

First, the configuration will be described.
The configuration in the first embodiment includes the following: “overall system configuration”, “internal configuration of engine control module”, “input conversion unit configuration of vehicle system vibration control device”, “vehicle body vibration estimation unit configuration of vehicle system vibration control device”, “ The description will be divided into “the torque command value calculation unit configuration of the vehicle system vibration control device”.

[Overall system configuration]
FIG. 1 is an overall system configuration diagram illustrating an engine vehicle to which the vehicle system vibration control device of the first embodiment is applied. The overall system configuration will be described below with reference to FIG.
Here, “vehicle system vibration control” has a function of suppressing vehicle body vibration by appropriately controlling the drive torque by the vehicle actuator (“engine 106” in the first embodiment) in accordance with the vibration of the vehicle body. Refers to control. In the vehicle system vibration control of the first embodiment, the effect of improving the yaw response at the time of steering, the effect of improving the linearity at the time of steering, and the effect of suppressing the roll behavior are also obtained.

  As shown in FIG. 1, the engine vehicle to which the vehicle system vibration control device of the first embodiment is applied is a rear wheel drive vehicle by manual shift, and includes an engine control module (ECM) 101 and an engine 106. Yes.

  The engine control module 101 (hereinafter referred to as “ECM101”) performs drive torque control of the engine 106. This ECM101 is connected to the steering wheel 110 and signals from wheel speed sensors 103FR, 103FL, 103RR, 103RL connected to the left and right front wheels 102FR, 102FL (driven wheel) and the left and right rear wheels 102RR, 102RL (drive wheel). A signal from the steering angle sensor 111 is input. Further, a signal from the brake stroke sensor 104 that detects the driver operation amount to the brake pedal and a signal from the accelerator opening sensor 105 that detects the driver operation amount to the accelerator pedal are input. A torque command value for driving the engine 106 is calculated according to these input signals, and the torque command value is sent to the engine 106.

  The engine 106 generates a drive torque according to the torque command value from the ECM 101, and the generated drive torque is increased / decreased by the MT transmission 107 according to the shift operation of the driver. The drive torque changed by the MT transmission 107 is further changed by the shaft 108 and the differential gear 109 and transmitted to the left and right rear wheels 102RR and 102RL to drive the vehicle.

[Internal configuration of engine control module]
The vehicle structure vibration control device is configured in the form of a control program in the ECM 101, and FIG. 2 shows a block configuration representing the control program in the ECM 101. Hereinafter, the internal configuration of the ECM 101 will be described with reference to FIG.

  As shown in FIG. 2, the ECM 1101 includes a driver request torque calculation unit 201, a torque command value calculation unit 202, and a vehicle system vibration control device 203.

  The driver request torque calculation unit 201 inputs the brake operation amount information by the driver from the brake stroke sensor 104 and the accelerator operation amount information by the driver from the accelerator opening sensor 105, and calculates the driver request torque.

  The torque command value calculation unit 202 includes a torque command value obtained by adding the correction torque value from the vehicle system vibration control device 203 to the driver request torque from the driver request torque calculation unit 201, and other in-vehicle systems (for example, VDC and TCS). Etc.) is input. Based on the input information, a drive torque command value for the engine 106 is calculated.

  The vehicle system vibration control device 203 has a three-part configuration including an input conversion unit 204, a vehicle body vibration estimation unit 205, and a torque command value calculation unit 206. The input conversion unit 204 converts sensing information from the vehicle acquired during traveling into wheel input. The vehicle body vibration estimation unit 205 estimates the sprung behavior of the vehicle body using each wheel input from the input conversion unit 204 and the vehicle model. The torque command value calculation unit 206 suppresses the sprung behavior of the vehicle body based on the sprung behavior state quantity (bounce speed, bounce amount, pitch speed, pitch angle) of the vehicle body estimated by the vehicle body vibration estimation unit 205. The correction torque value is calculated.

[Configuration of input converter of vehicle system control device]
FIG. 3 shows a block configuration showing in detail the interior of the vehicle system vibration control device 203. Hereinafter, based on FIGS. 3 to 11, the configuration of the input conversion unit 204 in the three-part vehicle system vibration control device 203 will be described.

  The input conversion unit 204 converts sensing information from the vehicle into an input format (specifically, a dimension of torque or force applied to the wheels) to the vehicle model 307 used in the subsequent vehicle body vibration estimation unit 205. As shown in FIG. 3, the input conversion unit 204 includes a drive torque conversion unit 301, a high-pass filter 316, a suspension stroke calculation unit 302, a vertical force conversion unit 303, a vehicle body speed estimation unit 304, a first turning A resistance force calculation unit 305, a second turning resistance force calculation unit 306, and a switching processing unit 321 are included. In the input conversion unit 204, as input to the vehicle body vibration estimation unit 205, the drive shaft end torque Tw, the front wheel vertical force Ff and the rear wheel vertical force Fr, the front wheel turning resistance force Fcf, and the rear wheel turning resistance force Fcr , Is calculated.

<Calculation configuration of drive shaft end torque Tw>
The drive torque converter 301 adds the gear ratio to the driver request torque from the driver request torque calculator 201 and converts the engine end torque to the drive shaft end torque Tw. Here, the gear ratio is calculated from the ratio of the wheel speed (the average left and right rotational speed of the drive wheel) and the engine rotational speed. This gear ratio is the total gear ratio of the MT transmission 107 and the differential gear 109.

<Configuration for calculating front and rear wheel vertical forces Ff and Fr>
The high-pass filter 316 removes low-order steady components from the wheel speed signals from the wheel speed sensors 103FR, 103FL, 103RR, and 103RL. As the high-pass filter 316, a low-order filter having high stability and low calculation load is used.

The suspension stroke calculation unit 302 calculates the suspension stroke speed and the suspension stroke amount based on the wheel speed information after the high-pass filter process. When the suspension strokes, as shown in FIG. 4, the tire also has a displacement in the front-rear direction, and this relationship is determined by the geometry of the vehicle suspension. This is illustrated in FIGS. 5 and 6. FIG. By linearly approximating this relationship and assuming that the coefficient of vertical displacement relative to longitudinal displacement is KgeoF and KgeoR for the front and rear wheels, respectively, the vertical displacements Zf and Zr of the front and rear wheels are It becomes a relationship.
Zf = KgeoF xtf
Zr = KgeoR xtr
Differentiating the above equation yields the equation of tire longitudinal speed and vertical velocity, so the suspension stroke speed and suspension stroke amount are calculated using this relationship.

  In the vertical force conversion unit 303, the spring coefficient and the damping coefficient are added to the suspension stroke speed and the suspension stroke amount calculated by the suspension stroke calculation unit 302, respectively, and the sum is taken to obtain the front wheel vertical force Ff and the rear wheel. Convert to vertical force Fr.

<Calculation configuration of front and rear wheel turning resistance forces Fcf and Fcr>
The vehicle body speed estimation unit 304 outputs the wheel speed average value of the driven wheels 102FR and 102FL among the wheel speed information as the vehicle body speed V (= vehicle speed V).

In the first turning resistance calculation unit 305, the vehicle body speed V from the vehicle body speed estimation unit 304 and the steering angle θ from the steering angle sensor 111 are input, and the tire turning angle δ is calculated from the steering angle θ. The yaw rate γ and the vehicle body slip angle βv are calculated using the speed V, the steering angle θ, and the well-known linear two-wheel model. Then, based on the yaw rate γ, the vehicle body slip angle βv, and the tire turning angle δ, a first front wheel turning resistance force Fcf1 and a first rear wheel turning resistance force Fcr1 due to driver steering are calculated. That is, based on the yaw rate γ, the vehicle body slip angle βv, and the tire turning angle δ, the tire slip angles βf and βr of the front and rear wheels, which are tire side slip angles, are calculated using the following equations.
The front tire slip angle βf and the rear tire slip angle βr are
βf = βv + lf ・ γ / V-δ
βr = βv−lr ・ γ / V
It is calculated by the following formula. Here, lf and lr are distances from the center of gravity of the vehicle body to the front and rear axles.
Then, the tire lateral forces Fyf and Fyr of the front and rear wheels are calculated from the product of the tire slip angles βf and βr of the front and rear wheels and the cornering powers Cpf and Cpr of the front and rear wheels. Further, the first front wheel turning resistance force Fcf1 and the first rear wheel turning resistance force Fcr1 are calculated from the product of the tire slip angles βf, βr of the front and rear wheels and the tire lateral forces Fyf, Fyr of the front and rear wheels.

The second turning resistance calculation unit 306 receives the vehicle speed V from the vehicle speed estimation unit 304 and the steering angle θ from the steering angle sensor 111, and the vehicle speed / steering angle / turning resistance shown in FIGS. A second front wheel turning resistance force Fcf2 and a second rear wheel turning resistance force Fcr2 are calculated using a force three-dimensional map.
Here, the three-dimensional maps of the vehicle speed / steering angle / turning resistance shown in FIGS. 7 and 8 show that the turning resistance increases as the vehicle speed increases at the same steering angle, and the steering angle increases at the same vehicle speed. It is set for each of the front wheels (FIG. 7) and the rear wheels (FIG. 8) based on characteristics approximating the proportional relationship of lateral G (= lateral acceleration) so that the turning resistance force increases.

The switching processing unit 321 outputs the front wheel turning resistance force output to the vehicle body vibration estimation unit 205 based on the calculation results from the first turning resistance calculation unit 305 and the second turning resistance calculation unit 306 and the traveling condition (vehicle speed condition). Fcf and rear wheel turning resistance force Fcr are switched. That is, the first turning resistance force (the first front wheel turning resistance force Fcf1 and the first rear wheel) is used to determine the front wheel turning resistance force Fcf and the rear wheel turning resistance force Fcr when the vehicle speed is the medium speed or higher. The turning resistance force Fcr1) is used, and the second turning resistance force (second front wheel turning resistance force Fcf2 and second rear wheel turning resistance force Fcr2) is used to determine the turning resistance force when the vehicle speed is in the low speed range.
With respect to the switching vehicle speed, as shown in FIGS. 9 and 10, the low-speed to medium-speed switching vehicle speed V1 from the second turning resistance force to the first turning resistance force, and the first turning resistance force to the second turning resistance force. A hysteresis width ΔVH (= V1−V2) is provided between the transitional medium speed → low speed switching vehicle speed V2 (<V1).
Then, the switching processing unit 321 sets the value during the switching transition of the turning resistance force as a composite value F using a weighting coefficient for the first turning resistance force and the second turning resistance force, and according to the change in the vehicle speed V. A weighting switching configuration in which the turning resistance force changes smoothly is adopted.
That is, the turning resistance force (combined value F) is
F = k * Fb + (1-k) * Fa
In addition, Fa: 1st turning resistance force, Fb: 2nd turning resistance force, k: It calculates on the front-wheel side and the rear-wheel side, respectively using the formula of a weighting coefficient.
Further, as shown in FIGS. 9 and 10, the switching processing unit 321 sets the vehicle speed change gradient value of the weight coefficient k to a steeper vehicle speed change gradient value as the steering angular velocity is smaller, and becomes gentler as the steering angular velocity is larger. Use the value of the vehicle speed change gradient.
In addition, as shown in FIG. 11, the switching processing unit 321 turns when the absolute value of the torque command value calculated by the torque command value calculation unit 206 is smaller than the threshold (time t1 to t3, time t4 to). Switching of the resistance force is executed, and when the absolute value of the torque command value is equal to or greater than the threshold value (time t3 to t4), the turning resistance force is not changed.

[Configuration of vehicle vibration estimation unit of vehicle system vibration control device]
FIG. 3 shows a block configuration showing in detail the interior of the vehicle system vibration control device 203. Hereinafter, the configuration of the vehicle body vibration estimation unit 205 in the three-part vehicle system vibration control device 203 will be described with reference to FIGS. 3 and 12.

  The vehicle body vibration estimation unit 205 includes a vehicle model 307 (also referred to as “vibration model”), as shown in FIG. The vehicle model 307 is represented by a vertical motion equation and a pitching motion equation obtained by modeling an actual vehicle (vehicle body, front wheel suspension, rear wheel suspension, etc.) on which the system is mounted. Then, the “drive shaft end torque Tw”, “front and rear wheel vertical forces Ff and Fr”, and “front and rear wheel turning resistance forces Fcf and Fcr” calculated by the input conversion unit 204 are input to the vehicle model 307. Thereby, an estimated value by the vehicle model 307 of the sprung behavior state quantity (bounce speed, bounce quantity, pitch speed, pitch angle) of the vehicle body is calculated.

[Configuration of torque command value calculation unit of vehicle system vibration control device]
FIG. 3 shows a block configuration showing in detail the interior of the vehicle system vibration control device 203. Hereinafter, the configuration of the torque command value calculation unit 206 in the three-part vehicle system vibration control device 203 will be described with reference to FIGS. 3, 13, and 14.

  As shown in FIG. 3, the torque command value calculation unit 206 includes a first regulator unit 308, a second regulator unit 309, a third regulator unit 310, and a first tuning gain as a correction torque value generation processing configuration. A setting unit 317, a second tuning gain setting unit 318, a third tuning gain setting unit 319, and an adder 320 are provided. Further, as the actuator matching processing configuration of the corrected torque value, a limit processing unit 311, a band pass filter 312, a nonlinear gain amplification unit 313, a limit processing unit 314, and an engine torque conversion unit 315 are provided.

<Correction torque value generation processing configuration>
The first regulator unit 308 provides regulator gains F1 and F2 that suppress the sprung behavior to a minimum with respect to the “sprung behavior by torque input” that is the control target. As shown in FIG. 13, the first regulator unit 308 has a Trq-dZv gain F1 (bounce speed gain), a Trq-dSp gain F2 (pitch speed gain), as shown in FIG. ,give. As shown in FIG. 14, these regulator gains F1 and F2 contribute to the stabilization of the load. The Trq-dZv gain F1 suppresses the bounce speed, and the Trq-dSp gain F2 suppresses the pitch speed.

  The second regulator unit 309 provides regulator gains F3 to F6 that suppress the sprung behavior to the minimum with respect to the “sprung behavior due to disturbance” that is the control target. As shown in FIG. 13, the second regulator unit 309 has a Ws-SF gain F3 (front / rear balance gain) and a Ws-dSF gain F4 (front / rear balance change speed gain) as shown in FIG. Ws-dZv gain F5 (bounce speed gain) and Ws-dSp gain F6 (pitch speed gain) are given. As shown in FIG. 14, these regulator gains F3 to F6 contribute to the stabilization of the load. The Ws-SF gain F3 suppresses the longitudinal load change, and the Ws-dSF gain F4 indicates the longitudinal load change speed. The Ws-dZv gain F5 suppresses the bounce speed, and the Ws-dSp gain F6 suppresses the pitch speed.

  The third regulator unit 310 provides regulator gains F7 and F8 that improve behavior responsiveness due to steering with respect to the “sprung behavior due to steering” that is the object of control. As shown in FIG. 13, the third regulator unit 310 has a Str-dWf gain F7 (front wheel load change speed gain) and a Str-dWr gain F8 (rear wheel load change) as shown in FIG. Speed gain). As shown in FIG. 14, these regulator gains F7 and F8 contribute to the addition of a load. The Str-dWf gain F7 adds a front wheel load, and the Str-dWr gain F8 suppresses rear wheel load fluctuations. .

  The first tuning gain setting unit 317 sets a tuning gain K1 with respect to the Trq-dZv gain F1, as shown in FIG. 13, in order to perform weighting adjustment on the output from the first regulator unit 308, and Trq-dSp Set tuning gain K2 for gain F2. The tuning gains K1 and K2 are values in the positive direction that suppress vibrations, and are values included in the front and rear G fluctuation range that does not give a sense of incongruity. The tuning gains K1 and K2 enable gain correction by integration with the weighting coefficient when the weighting coefficient is determined according to the vehicle state, the traveling state, the driver selection, and the like with respect to the preset initial value.

  The second tuning gain setting unit 318 sets the tuning gain K3 with respect to the Ws-SF gain F3 as shown in FIG. 13 in order to perform weighting adjustment on the output from the second regulator unit 309, and Ws-dSF The tuning gain K4 is set for the gain F4, the tuning gain K5 is set for the Ws-dZv gain F5, and the tuning gain K6 is set for the Ws-dSp gain F6. The tuning gains K3 to K6 are values in the positive direction that suppress vibration and are included in the front-to-back G fluctuation range that does not give a sense of incongruity, like the tuning gains K1 and K2. The tuning gains K3 to K6 enable gain correction by integration with the weighting factors when the weighting factors are determined according to the vehicle state, the running state, the driver selection, and the like with respect to the preset initial values.

  The third tuning gain setting unit 319 sets the tuning gain K7 for the Str-dWf gain F7 as shown in FIG. 13 in order to perform weighting adjustment on the output from the third regulator unit 310, and the Str-dWr Set tuning gain K8 for gain F8. Unlike the tuning gains K1 to K6, the tuning gains K7 and K8 are set to values in the negative direction that promote vibration and values that are included in the front and rear G fluctuation range that does not give a sense of incongruity. The tuning gains K7 and K8 enable gain correction by integration with the weighting coefficient when the weighting coefficient is determined according to the vehicle state, the traveling state, the driver selection, etc. with respect to the preset initial value.

  The adder 320 performs regulator processing for each behavior to be controlled with respect to the sprung behavior state amount (bounce speed, bounce amount, pitch speed, pitch angle) of the vehicle body calculated by the vehicle body vibration estimation unit 205. Are integrated with the tuning gains K1 to K8, and the sum is calculated to calculate a correction torque value required for the control. This correction torque value is a value obtained by adding the correction torque value A based on the tuning gains K1 and K2, the correction torque value B based on the tuning gains K3 to K6, and the correction torque value C based on K7 and K8.

<Compensation processing configuration for correction torque value>
The limit processing unit 311 performs a maximum value limiting process of the absolute value of the correction torque value on the correction torque value from the adder 320 as a drive system resonance countermeasure, and a torque within a range that the driver does not feel as a G fluctuation. Restrict to.

  The band-pass filter 312 extracts the sprung vibration component of the vehicle body and removes the drive system resonance frequency component so as to suppress the sprung resonance as a countermeasure for the drive system resonance as in the limit processing unit 311. The reason for this is that in an actual vehicle, particularly an engine vehicle, when a vibration component is inadvertently added to the drive torque, vibration that interferes with the drive system resonance may be generated. In addition, an engine vehicle or the like is necessary because there is a possibility that the expected control effect cannot be sufficiently obtained because of poor response to the drive torque command and a dead zone.

  The non-linear gain amplifying unit 313 is used in the vicinity of the correction torque value positive / negative switching region (= actuator dead zone region) as a countermeasure against the response of the actuator (engine 106) to the correction torque value output from the bandpass filter 312. Amplify the correction torque value.

  The limit processing unit 314 performs a final limit process on the corrected torque value output from the nonlinear gain amplification unit 313 after the amplification process.

  The engine torque conversion unit 315 converts the corrected torque value after the limit processing from the limit processing unit 314 into an engine end torque value corresponding to the gear ratio, and outputs this as a final correction torque value.

Next, the operation will be described.
The functions of the vehicle system vibration control device of the first embodiment are as follows: “Basic system vibration control function”, “Car system vibration control processing function”, “Scenes and effects aiming at performance improvement by vehicle system vibration control”, “Car system vibration control” The vibration control logic and vehicle system vibration control effect ”,“ turning resistance force calculating action by switching according to vehicle speed ”, and“ turning resistance force calculating action by weighted switching transition ”will be described separately.

[Basic action of vehicle system vibration control]
It is necessary to understand in detail what mechanism controls the sprung behavior of the vehicle body in the vehicle system vibration control by the drive torque. Hereinafter, based on FIG. 15, the basic operation of the vehicle system vibration control that reflects this will be described.

First, the vehicle system vibration control is a control aimed at stabilizing the load and improving the turning performance by suppressing the change speed of the vehicle body behavior due to torque fluctuation or disturbance by correcting the engine torque.
Therefore, as a specific running situation, as shown in FIG. 15 (a), for example, a case where the vehicle starts and accelerates from a stop, enters a constant speed state, and then decelerates and stops.

  When starting and accelerating from the stop, the driving torque rapidly increases, so that a load movement occurs in which the wheel load of the rear wheel increases and the wheel load of the front wheel decreases, and the vehicle body behavior becomes a nose up in which the front side of the vehicle body is raised. At this time, as shown in FIGS. 15 (a) and 15 (b), if the driving torque to the rear wheel, which is the driving wheel, is reduced, a nose-down behavior occurs in which the front side of the vehicle body sinks like during deceleration, The nose-up due to load movement and the nose-down due to torque-down cancel each other, and the body behavior is stabilized.

  In a steady state where the vehicle enters a constant speed state after starting, control of correcting the driving torque is not performed because the vehicle body behavior is stable. After that, when the vehicle is decelerated and stopped by performing a brake operation or the like, a load movement occurs in which the wheel load of the rear wheel decreases and the wheel load of the front wheel increases due to a sudden decrease in the drive torque. It becomes a nose down where the front side of the body sinks. At this time, as shown in FIGS. 15 (a) and 15 (b), when the driving torque to the rear wheel, which is the driving wheel, is increased, a nose-up behavior in which the front side of the vehicle body is lifted as during acceleration occurs. The nose-down due to movement and the nose-up due to torque-up cancel each other, and the vehicle behavior becomes stable.

  Accordingly, when the change in the pitch angular velocity of the vehicle body is seen, as shown in FIG. 15C, the change in the pitch angular velocity of the vehicle body is smaller in the solid line characteristic of “with vibration suppression” than the dotted line characteristic of “without vibration suppression”. It will be suppressed.

[Car system vibration control processing action]
FIG. 16 is a flowchart showing the flow of the vehicle structure vibration control process executed by the engine control module 101 of the first embodiment. Hereinafter, the operation of the vehicle structure vibration control process will be described with reference to FIG.

  When the vehicle system vibration control process is started, the driver request torque is calculated by the driver request torque calculation unit 201 in step S1401. In the next step S1402, the drive torque converter 301 adds the gear ratio to the driver request torque, and converts the unit from the engine end torque to the drive shaft end torque Tw. In the next step S1403, the high-pass filter 316 performs filter processing for removing low-order steady components from the wheel speed signals of the wheel speed sensors 103FR, 103FL, 103RR, and 103RL. In the next step S1404, the suspension stroke calculation unit 302 calculates the suspension stroke speed and the suspension stroke amount based on the wheel speed information after the high-pass filter processing. In the next step S1405, the vertical stroke converting unit 303 converts the suspension stroke speed and the suspension stroke amount into the front and rear wheel vertical forces Ff and Fr. In the next step S1406, the steering angle is detected by the steering angle sensor 111. In the next step S1407, the vehicle body speed V is calculated by the vehicle body speed estimation unit 304. In the next step S1408, the first turning resistance calculating unit 305 calculates the yaw rate γ and the vehicle body slip angle βv (= vehicle body side slip angle). In the next step S1409, the first turning resistance calculation unit 305 calculates tire slip angles βf and βr (tire slip angles) of the front and rear wheels. In the next step S1410, the first turning resistance calculation unit 305 calculates the tire lateral forces Fyf and Fyr of the front and rear wheels. In the next step S1411, the first turning resistance calculation unit 305 calculates the first front and rear wheel turning resistance forces Fcf1, Fcr1. In the next step S1423, the second turning resistance calculation unit 306 calculates the second front and rear wheel turning resistance forces Fcf2 and Fcr2. In the next step S1424, the front and rear wheel turning resistance forces Fcf and Fcr are calculated in the switching process 321 by the weighted switching process for the vehicle speed change. The above processing is performed in the input conversion unit 204.

  In the next step S1412, the vehicle body vibration estimation unit 205 inputs the drive shaft end torque Tw, the front and rear wheel vertical forces Ff and Fr, and the front and rear wheel turning resistance forces Fcf and Fcr to the vehicle model 307, thereby Behavioral state quantities (bounce speed, bounce quantity, pitch speed, pitch angle) are calculated. In the next step S1413, the tuning gains K1 to K8 are corrected by the vehicle speed or the like. In the next step S <b> 1414, the first tuning gain setting unit 317 calculates a correction torque value A that suppresses vibration due to driver requested torque. In the next step S1415, the second tuning gain setting unit 318 calculates a correction torque value B that suppresses vibration due to disturbance. In the next step S1416, the third tuning gain setting unit 319 calculates a correction torque value C that amplifies fluctuations in the longitudinal load due to steering. In the next step S <b> 1417, a corrected torque value based on the sum of the corrected torque value A, the corrected torque value B, and the corrected torque value C is output.

In the next step S1418, the limit processing unit 311 performs drive system resonance countermeasure limit processing on the correction torque value. In the next step S1419, the bandpass filter 312 performs a filter process for removing the drive system resonance component on the correction torque value. In the next step S1420, nonlinear gain processing for amplifying the correction torque value in the vicinity of the positive / negative switching region is performed in the nonlinear gain amplifying unit 313. In the next step S1421, the limit processing unit 314 performs final limit processing on the corrected torque value after amplification processing. In the next step S1422, the engine torque conversion unit 315 converts the drive shaft end correction torque value into an engine end correction torque value, which is output as the final correction torque value.
The vehicle structure vibration control process that proceeds from step S1401 to step S1422 is repeated every predetermined control cycle.

[Scenes and effects aimed at improving performance through vehicle system vibration control]
The scene and effect aiming at performance improvement by the vehicle structure vibration control process of the first embodiment by the above-described vehicle structure vibration control process will be described based on FIG.

The scene aiming at performance improvement by the vehicle system vibration control of Example 1 and the effect are as follows:
(a) To obtain a stable linear turning performance with a gentle roll and good linearity in lane changes and scenes such as S-shaped roads.
(b) To obtain stable cruising performance of the vehicle due to the lack of correction steering and good pitch damping in scenes such as high-speed cruising.
It is in.

  To achieve the effect (a) above, it is necessary to “improve the steering response” and “suppress roll speed”, and to achieve the effect (b) above, it is necessary to “suppress load fluctuation”. It is. Hereinafter, the reason why these effects can be realized by the vehicle system vibration control will be described with reference to FIG.

  As shown in FIG. 17, “improvement of steering response” means that when the vehicle is decelerated = torque-down during steering, the front wheel load increases, the cornering power Cp of the front tire increases, and the tire lateral force increases. The steering response is improved. That is, the cornering power Cp is increased by increasing the wheel load at the time of steering by using a load-dependent characteristic that increases as the wheel load increases.

  As shown in FIG. 17, for example, when a nose-up behavior occurs, “deceleration of load fluctuation” causes a motion (nose-down) in an opposite phase to vehicle body vibration when deceleration = torque down. The load fluctuation is suppressed by canceling the load fluctuation. On the other hand, when nose-down behavior occurs, if acceleration = torque up is performed, motion in the opposite phase to the vehicle body vibration (nose-up) occurs, and load fluctuation is suppressed by offsetting the load fluctuation. The load fluctuation is suppressed both when the vibration (load fluctuation) is generated by the driver input and when the vibration (load fluctuation) is generated by the road surface disturbance. That is, when the pitch behavior due to the torque fluctuation and the road surface disturbance is estimated, “load fluctuation suppression” is realized with the driving torque having the opposite phase to the estimated pitch behavior.

  As shown in FIG. 17, “roll speed reduction” improves the linearity of yaw rate by the above-described “improvement of steering response” and “suppression of load fluctuation”. Therefore, the lateral G change is gentle in proportion to the yaw rate, the peak value of the roll rate is reduced, and the roll speed is suppressed. That is, as a result of combining “improvement of steering response” and “suppression of load fluctuation”, “suppression of roll speed” is realized.

  Therefore, at the time of steering, the yaw response is improved by actively promoting the nose-down behavior so that the front wheel load increases, and at the same time, the extra vibration component is suppressed, thereby ensuring the linearity. And since the sudden change of the horizontal G is suppressed by performing these controls simultaneously, the effect (a) aimed at by this control that can suppress the roll rate can be realized.

  On the other hand, when cruising on a straight road without steering, the pitch behavior due to torque fluctuation and road surface disturbance is estimated, and by applying a driving torque in the opposite phase to the estimated pitch behavior, load fluctuation is suppressed and the vehicle is stabilized. The effect (b) aimed by this control to obtain cruise performance can be realized.

[Vehicle structure vibration control logic and vehicle structure vibration control effect]
The vehicle structure vibration control logic and the vehicle structure vibration control effect of the first embodiment that achieve the performance improvement effect and the vehicle structure vibration control will be described with reference to FIGS. 18 and 19.

  First, as shown in FIG. 18, the vehicle system vibration control logic of the first embodiment includes a driver request torque (= drive shaft end torque Tw), front wheel vertical force Ff, rear wheel vertical force Fr, front wheel turning resistance force Fcf, rear The wheel turning resistance force Fcr is input to the vehicle model 307. Thereby, the bounce speed, the bounce amount, the pitch speed, and the pitch angle, which are the sprung behavior state quantities of the vehicle body, are calculated.

  Then, as shown in FIG. 18, each of the sprung behavior state quantities of the vehicle body is multiplied by regulator gains F1 to F8 for optimizing the bounce speed, the bounce amount, the pitch speed, and the pitch angle. Multiply the tuning gains K1 to K8.

  With the above processing, correction torque values A, B, and C are obtained for each of the “sprung behavior by torque input”, “sprung behavior by disturbance”, and “sprung behavior by steering”, which are control targets. Then, the correction torque values A, B, and C are added together to obtain a final correction torque value (= control torque in FIG. 18), and a drive torque command value for obtaining a drive torque obtained by adding the control torque to the driver request torque And output to the engine 106 of the actual vehicle.

Here, among the corrected torque values A, B, and C, the corrected torque value C corrects the driving torque so as to add the front wheel load during steering, and positively applies the wheel load to the left and right front wheels 102FR and 102FL. This is the correction torque value for getting on.
Therefore, at the time of steering, the yaw response is improved by actively promoting the nose-down behavior so that the front wheel load is increased by the correction torque value C, and at the same time, unnecessary vibration components are suppressed by the correction torque values A and B. This ensures linearity. That is, the effect (a) targeted by the present control for suppressing the roll rate is realized by adding the correction torque value C to the correction torque values A and B.

On the other hand, among the above correction torque values A, B, and C, the correction torque values A and B stabilize the longitudinal load fluctuation and reduce the vehicle body vibration regardless of the fluctuation of the driving torque and the road surface disturbance during traveling on the straight road. It is a correction torque value to suppress.
Therefore, during cruising on a straight road without steering, the pitch behavior, bounce behavior, and longitudinal load change due to torque fluctuation and road disturbance are estimated, and the estimated pitch behavior, bounce behavior, and longitudinal load change are estimated based on the corrected torque values A and B. By applying a driving torque having an opposite phase, pitch behavior, bounce behavior (up-down behavior), and change in front-rear load are suppressed. That is, the effect (b) targeted by the present control for obtaining a stable cruise performance of the vehicle is realized by the correction torque values A and B.

  Next, confirmation that the targeted effects (a) and (b) are realized by the vehicle system vibration control logic of the first embodiment will be described with reference to FIG. Note that FIG. 19 shows a contrast characteristic (solid line characteristic with control, dotted line characteristic without control) in a time series when steering from straight running.

In the vehicle system vibration control, as indicated by an arrow J in FIG. 19, a control command value (= drive torque command value) is output by (command torque for suppressing vehicle body vibration) + (command torque for controlling steering response). .
For this reason, in the straight traveling region up to time t1, as shown by the arrow E in FIG. 19, the pitch rate is suppressed as compared to the case without control, and the riding comfort is improved by the stable traveling performance of the vehicle. I understand that.

  Then, in the steering transition region after time t1, as shown by the arrow F in FIG. 19, it can be seen that the change in the pitch rate is suppressed and appropriate load movement is realized. As shown by an arrow G in FIG. 19, in the steering transient region, as shown by an arrow G in FIG. 19, it can be seen that the yaw rate rises earlier than in the case of no control, and the initial response is improved. Furthermore, in the steering transition region, in the late turning period, as shown by the arrow H in FIG. 19, it can be seen that the yaw rate changes more gently than in the case of no control, and the turning entanglement is suppressed.

  In the steering transition region (from the early turn to the late turn), the control for suppressing the change in the pitch rate and the control for suppressing the change in the yaw rate are performed at the same time. As shown by the arrow I of 19, it can be seen that the roll rate is suppressed as compared with the case of no control.

[Calculation of turning resistance by switching according to vehicle speed]
In order to realize the effect (a) targeted by the present control, it is assumed that the front and rear wheel turning resistance forces Fcf and Fcr are accurately calculated based on sensing information during traveling. Therefore, it is necessary to devise a method for ensuring the calculation accuracy of the front and rear wheel turning resistance forces Fcf and Fcr regardless of the traveling state. Hereinafter, the turning resistance calculation operation by switching according to the vehicle speed reflecting this will be described.

For example, a calculation example of a turning resistance force using one control logic that calculates a turning resistance force from a steering angle using a linear two-wheel model is used as a comparative example.
In the case of this comparative example, the calculation accuracy of the turning resistance force is ensured in a vehicle speed range of medium speed or higher where the approximation with the linear two-wheel model is high. However, in the low speed range where the approximation with the linear two-wheel model is lowered, the calculation accuracy of the turning resistance force is lowered. As a result, in a low-speed turning scene, a stable linear turning performance may not be realized due to a gentle roll and good linearity.

  On the other hand, in the first embodiment, both the first turning resistance calculation unit 305 and the second turning resistance calculation unit 306 that calculate the turning resistance force by different calculation methods based on the steering angle signal and the vehicle body speed signal, Based on the calculation results from the calculation units 305 and 306 and the traveling condition (vehicle speed condition), a switching processing unit 321 that switches between the front wheel turning resistance force Fcf and the rear wheel turning resistance force Fcr output to the vehicle body vibration estimation unit 205 is provided. Adopted.

  That is, in the first turning resistance calculation unit 305, the vehicle body speed V from the vehicle body speed estimation unit 304 and the steering angle θ from the steering angle sensor 111 are input, and the tire turning angle δ is calculated based on the steering angle. The yaw rate γ and the vehicle body slip angle βv are calculated using the angle and the equation of the linear two-wheel model. Then, based on the yaw rate γ, the vehicle body slip angle βv, and the tire turning angle δ, the first front wheel turning resistance force Fcf1 and the first rear wheel turning resistance force Fcr1 by the driver steering are calculated.

  On the other hand, in the second turning resistance calculation unit 306, the vehicle speed V from the vehicle speed estimation unit 304 and the steering angle θ from the steering angle sensor 111 are input, and the vehicle speed / steering angle / turning resistance shown in FIG. The second front wheel turning resistance force Fcf2 is calculated using the three-dimensional map, and the second rear wheel turning resistance force Fcr2 is calculated using the three-dimensional map of vehicle speed / steering angle / turning resistance force shown in FIG. .

  Here, the “linear two-wheel model” is a turning behavior model in which the left and right front wheels and the left and right rear wheels are regarded as one wheel and the linear characteristics in a steady turning state are simulated. However, the actual turning behavior of the vehicle is a complex behavior including non-linear characteristics. In particular, the yaw rate γ and the vehicle body are determined using a linear two-wheel model when turning in a low speed range where the steering angle and the vehicle speed vary greatly. When the slip angle βv is calculated, the calculated value may deviate from the actual value. On the other hand, the “three-dimensional map of vehicle speed / steering angle / turning resistance force” is a map based on characteristics approximating the proportional relationship of lateral G. For this reason, it is possible to follow changes in the steering angle and the vehicle speed when turning in a low speed region, and the calculation accuracy of the turning resistance force is increased as compared with the case of using a linear two-wheel model.

  Then, in the switching processing unit 321, based on the calculation results from the first turning resistance calculation unit 305 and the second turning resistance calculation unit 306 and the vehicle speed condition, the front wheel turning resistance force Fcf output to the vehicle body vibration estimation unit 205 and the rear The wheel turning resistance force Fcr is switched. In this switching process, the first turning resistance force (first front wheel turning resistance force Fcf1 and first rear wheel turning resistance force Fcr1) is determined in determining the front wheel turning resistance force Fcf and the rear wheel turning resistance force Fcr when the vehicle speed is medium or higher. ) Is used. On the other hand, the second turning resistance force (second front wheel turning resistance force Fcf2 and second rear wheel turning resistance force Fcr2) is used to determine the turning resistance force when the vehicle speed is in the low speed range.

  That is, the calculation method for ensuring the calculation accuracy of the turning resistance force is the second turning resistance force with respect to the low-speed turning that is the running condition in which the calculation accuracy of the turning resistance force is low in the calculation method by the first turning resistance force calculation unit 305. The calculation unit 306 can be assigned.

  As a result, in a scene where the vehicle speed condition changes, when the low-speed range condition in which the calculation accuracy by one of the first turning resistance calculation units 305 is low, the other second turning resistance calculation unit 306 ensures the calculation accuracy. By doing so, the calculation accuracy of the turning resistance force is ensured regardless of changes in the vehicle speed condition. Then, by outputting these turning resistance forces (front wheel turning resistance force Fcf and rear wheel turning resistance force Fcr) to the vehicle body vibration estimation unit 205, a linear with a sense of stability targeted by this control regardless of the vehicle speed condition. The turning performance (the above effect (a)) can be realized.

  In the first embodiment, with respect to the switching vehicle speed, as shown in FIG. 9 and FIG. 10, the low-speed to medium-speed switching vehicle speed V1 that changes from the second turning resistance force to the first turning resistance force and the first turning resistance force to the second A configuration is adopted in which a hysteresis width ΔVH (= V1−V2) is provided between the vehicle speed V2 (<V1) for switching from the medium speed to the low speed for transition to the turning resistance force.

  With this configuration, when the vehicle speed changes from the low speed to the medium speed, as shown in FIG. 9, the second turning resistance calculating unit 306 calculates the second speed until the vehicle speed V reaches the low speed → medium speed switching vehicle speed V1. The turning resistance force (second front wheel turning resistance force Fcf2 and second rear wheel turning resistance force Fcr2) is maintained. Thereafter, when the vehicle speed changes from the medium speed to the low speed, as shown in FIG. 10, the first turning resistance calculation unit 305 calculates until the vehicle speed V reaches the medium speed → low speed switching vehicle speed V2 (<V1). The first turning resistance force (first front wheel turning resistance force Fcf1 and second rear wheel turning resistance force Fcr1) is maintained.

In other words, assuming that the switching vehicle speed does not have the hysteresis width ΔVH, hunting in which the first turning resistance force and the second turning resistance force are frequently switched and selected in a turning scene in which the vehicle speed changes in the switching vehicle speed range. This causes fluctuations in the torque command value.
On the other hand, by providing the switching vehicle speed with a hysteresis width ΔVH, switching hunting between the first turning resistance force and the second turning resistance force is prevented in a turning scene where the vehicle speed changes in the switching vehicle speed range, and torque is reduced. The fluctuation of the command value is suppressed.

[Calculation of turning resistance by weight switching transition]
As described above, in the first embodiment, the first turning resistance force or the second turning resistance force is selected according to the vehicle speed. When switching between these two turning resistance forces, switching between ON / OFF before and after the switching vehicle speed may cause a sudden change in the value of the turning resistance input to the vehicle model. is necessary. Hereinafter, the turning resistance force calculating action by the weighted switching transition reflecting this will be described.

  In the first embodiment, the switching processing unit 321 sets the value during the transition of the turning resistance force as the combined value F using the weighting coefficient k for the first turning resistance force Fa and the second turning resistance force Fb, and the vehicle speed. A weighting switching configuration in which the turning resistance smoothly changes according to changes in V is adopted.

With this configuration, when the vehicle speed changes from the low speed to the medium speed, as shown in FIG. 9, when the vehicle speed V reaches the low speed → medium speed switching vehicle speed V1, the weighting coefficient k is increased from k = 1 as the vehicle speed V increases. It gradually decreases and reaches k = 0. That is, the weighted distribution gradually changes from the large distribution of the second turning resistance force Fb to the large distribution of the first turning resistance force Fa.
Similarly, when the vehicle speed changes from the medium speed to the low speed, when the vehicle speed V reaches the medium speed → low speed switching vehicle speed V2 (<V1) as shown in FIG. It gradually rises from k = 0 and reaches k = 1. That is, the weighted distribution gradually changes from the large distribution of the first turning resistance force Fa to the large distribution of the second turning resistance force Fb.

  That is, when the first turning resistance Fa and the second turning resistance Fb are configured to be switched ON / OFF, when the vehicle speed reaches the switching vehicle speed, the first turning resistance Fa and the second turning resistance Fb are The value is suddenly switched, causing the torque command value to fluctuate.

  On the other hand, since the first turning resistance force Fa and the second turning resistance force Fb are weighted and switched, when the turning scene reaches the switching vehicle speed V1 or V2 due to a change in the vehicle speed, the front and rear wheel turning resistance forces Fcf and Fcr Sudden changes are prevented, and fluctuations in the torque command value are suppressed.

  In the first embodiment, as shown in FIGS. 9 and 10, as the switching processing unit 321, the vehicle speed change gradient value of the weight coefficient k is set to a steeper vehicle speed change gradient value as the steering angular velocity is smaller, and the steering angular velocity is larger. The configuration is such that the value of the vehicle speed change gradient is moderate.

With this configuration, when the vehicle speed changes from the low speed to the medium speed, as shown in FIG. 9, when the vehicle speed V reaches the low speed → medium speed switching vehicle speed V1 and the steering angular speed is small, the weighting coefficient k increases the vehicle speed V. Therefore, it decreases at a steep slope from k = 1 and reaches k = 0. On the other hand, when the vehicle speed V reaches the low to medium speed switching vehicle speed V1 and the steering angular speed is large, the weighting coefficient k decreases from k = 1 with a gentle gradient and reaches k = 0 as the vehicle speed V increases. . That is, the weighted distribution transition speed from the second turning resistance force Fb to the first turning resistance force Fa changes depending on the magnitude of the steering angular velocity.
Similarly, when the vehicle speed changes from the medium speed to the low speed, as shown in FIG. 10, when the vehicle speed V reaches the medium speed → low speed switching vehicle speed V2 (<V1) and the steering angular speed is small, the weighting coefficient k becomes the vehicle speed. As V decreases, it rises at a steep slope from k = 0 and reaches k = 1. On the other hand, when the vehicle speed V reaches the low-to-medium speed switching vehicle speed V1 and the steering angular speed is large, the weighting coefficient k increases from k = 0 with a gentle gradient and reaches k = 1 as the vehicle speed V decreases. . That is, the weighted distribution transition speed from the first turning resistance force Fa to the second turning resistance force Fb changes depending on the magnitude of the steering angular velocity.

  That is, when the weighted distribution transition speed between the first turning resistance force Fa and the second turning resistance force Fb is set to a constant speed (for example, an intermediate speed with a large or small steering angular speed) regardless of the steering angular speed, the torque command value When the fluctuation width is small and the steering angular velocity is small, the weighted distribution transition speed becomes too slow, and the switching responsiveness decreases. On the other hand, when the fluctuation range of the torque command value is large and the steering angular velocity is large, the weighted distribution transition speed becomes too early, and the torque command value varies.

  On the other hand, by changing the weighted distribution transition speed (vehicle speed change gradient) between the first turning resistance force Fa and the second turning resistance force Fb according to the steering angular speed, the torque command value varies. When the steering angle speed is small and the width is small, the switching response is improved. On the other hand, when the fluctuation range of the torque command value is large and the steering angular velocity is large, the fluctuation of the torque command value is surely suppressed.

  In the first embodiment, as shown in FIG. 11, as the switching processing unit 321, when the absolute value of the torque command value calculated by the torque command value calculation unit 206 is smaller than the threshold (time t1 to t3, time t4 to) Execute the switching of the turning resistance force. In other words, when the absolute value of the torque command value is equal to or greater than the threshold (time t3 to t4), the turning resistance force is not switched.

With this configuration, for example, when the vehicle speed changes from low speed to medium speed, the absolute value of the torque command value is a threshold value from time t2 until time t3 when the vehicle speed V reaches the low speed → medium speed switching vehicle speed V1, as shown in FIG. Smaller than. However, when the absolute value of the torque command value becomes greater than or equal to the threshold value at time t3, the weighting coefficient k is maintained as it is between time t3 and t4 when the absolute value of the torque command value is equal to or greater than the threshold value, and the vehicle body vibration estimation unit 205 is notified. The values of the front and rear wheel turning resistance forces Fcf and Fcr to be output are calculated by a weighting factor k based on a fixed value. It should be noted that while the absolute value of the torque command value is equal to or greater than the threshold value, the weighting factor k is similarly held when the vehicle speed changes from medium speed to low speed.
In other words, when the turning resistance force is switched even when the absolute value of the torque command value is equal to or greater than the threshold value, the switching of the turning resistance force may further promote the fluctuation of the torque command value.
On the other hand, when the absolute value of the torque command value is equal to or greater than the threshold value, switching of the turning resistance force is not executed, so that it is possible to suppress the fluctuation of the torque command value by switching the turning resistance force.

Next, the effect will be described.
In the vehicle system vibration control device of the first embodiment, the effects listed below can be obtained.

(1) An input conversion unit 204 that converts sensing information from the vehicle acquired during traveling into wheel input; a vehicle body vibration estimation unit 205 that estimates the sprung behavior of the vehicle body using the wheel input and the vehicle model 307; In the vehicle system vibration control device, comprising a torque command value calculation unit 206 that corrects the drive torque based on the estimation result of the sprung behavior,
The input conversion unit 204 is
A first turning resistance calculation unit 305 that calculates a first turning resistance force (first front and rear wheel turning resistance forces Fcf1, Fcr1) using a steering angle signal, a vehicle body speed signal, and a linear two-wheel model;
Calculation of second turning resistance force (second front and rear wheel turning resistance force Fcf2, Fcr2) using a three-dimensional map of steering angle signal, vehicle speed signal, vehicle speed / steering angle / turning resistance force Part 306;
A switching processing unit 321 for switching the turning resistance force (front and rear wheel turning resistance force Fcf, Fcr) to be output to the vehicle body vibration estimation unit 205 based on the calculation results from the two turning resistance force calculation units 305 and 306 and traveling conditions;
(FIG. 3).
For this reason, it is possible to ensure the calculation accuracy of the turning resistance force (front and rear wheel turning resistance forces Fcf, Fcr) regardless of the change in the running condition in a turning scene in which the running conditions change.

(2) The switching processing unit 321 uses the first turning resistance force (the first turning force) in determining the turning resistance force (front and rear wheel turning resistance forces Fcf, Fcr) when the vehicle speed is a medium speed or higher with the traveling condition as the vehicle speed condition. The second turning resistance force (second front and rear wheel turning resistance force) is used to determine the turning resistance force (front and rear wheel turning resistance forces Fcf and Fcr) when the vehicle speed is low. Fcf2, Fcr2) are used (FIGS. 9 and 10).
For this reason, in addition to the effect of (1), it is possible to ensure the calculation accuracy of the turning resistance force (front and rear wheel turning resistance forces Fcf, Fcr) regardless of the change in the vehicle speed condition in the traveling scene where the vehicle speed condition changes.

(3) The switching processing unit 321 transitions from the second turning resistance force (second front and rear wheel turning resistance force Fcf2, Fcr2) to the first turning resistance force (first front and rear wheel turning resistance force Fcf1, Fcr1). Low-speed to medium-speed switching vehicle speed V1 and the transition from the first turning resistance force (first front and rear wheel turning resistance force Fcf1, Fcr1) to the second turning resistance force (second front and rear wheel turning resistance force Fcf2, Fcr2) A hysteresis width ΔVH is provided between the vehicle speed V2 and the vehicle speed V2 at which the vehicle speed is switched from low to high (FIGS. 9 and 10).
For this reason, in addition to the effect of (2), the first turning resistance force (first front and rear wheel turning resistance forces Fcf1, Fcr1) that causes the torque command value to fluctuate in a turning scene where the vehicle speed changes in the switching vehicle speed range. And switching hunting of the second turning resistance force (second front and rear wheel turning resistance force Fcf2, Fcr2) can be prevented.

(4) The switching processing unit 321 sets a value during the switching transition of the turning resistance force as a combined value F using a weighting coefficient k for the first turning resistance force Fa and the second turning resistance force Fb. A weighting switching configuration in which the turning resistance force (front and rear wheel turning resistance forces Fcf, Fcr) smoothly changes according to the change in the vehicle speed V (FIGS. 9 and 10).
For this reason, in addition to the effect of (2) or (3), the turning resistance force (front and rear wheel turning resistance forces Fcf, Fcr) that causes the torque command value to fluctuate when the turning scene reaches the switching vehicle speed V1 or V2 due to a change in the vehicle speed. ) Can be prevented.

(5) The switching processing unit 321 sets the value of the vehicle speed change gradient of the weighting factor k to a gentler value of the vehicle speed change gradient as the steering angular velocity increases (FIGS. 9 and 10).
For this reason, in addition to the effect of (4), it is possible to suppress a change in torque command value in a turning scene with a large steering angular velocity while improving switching responsiveness in a turning scene with a small steering angular velocity.

(6) The switching processing unit 321 performs switching of the turning resistance force when the absolute value of the torque command value calculated by the torque command value calculating unit 206 is smaller than the threshold value, and the absolute value of the torque command value is the threshold value. It is set as the structure which does not perform switching of turning resistance force at the above (FIG. 11).
For this reason, in addition to the effects (1) to (5), it is possible to suppress a change in the torque command value calculated by the torque command value calculating unit 206 from being promoted by switching the turning resistance force.

  As mentioned above, although the vehicle system vibration control device of the present invention has been described based on the first embodiment, the specific configuration is not limited to the first embodiment, and the invention according to each claim of the claims. Design changes and additions are permitted without departing from the gist of the present invention.

  In the first embodiment, the switching processing unit 321 sets the value during the transition of the turning resistance force as the combined value F using the weighting coefficient k for the first turning resistance force Fa and the second turning resistance force Fb, and the vehicle speed. An example of a weighted switching configuration in which the front and rear wheel turning resistance forces Fcf and Fcr change smoothly according to changes in V is shown. However, the switching processing unit may be an example of switching the turning resistance when the switching vehicle speed is reached. However, in this case, for example, it is preferable to gradually change the front and rear wheel turning resistance force toward the target turning resistance force after switching as time elapses.

  In the first embodiment, the torque command value calculation unit 206 includes a first tuning gain setting unit 317 for behavior due to torque input, a second tuning gain setting unit 318 for behavior due to disturbance, and third tuning for improving behavior response due to steering. An example including a gain setting unit 319 is shown. However, as long as the torque command value calculation unit includes at least the third tuning gain setting unit for improving the behavioral response due to steering, the two tuning tuning units may be provided even in the example including one tuning gain setting unit. Even an example provided with a gain setting unit may be an example provided with three or more tuning gain setting units.

  In the first embodiment, an example in which a state quantity represented by a bounce speed, a bounce amount, a pitch speed, and a pitch angle is used as the sprung behavior of the vehicle body estimated by the vehicle body vibration estimation unit 205 is shown. However, as the sprung behavior of the vehicle body estimated by the vehicle body vibration estimation unit, any one of pitch behavior, bounce behavior, wheel load fluctuation, or a combination of these behaviors may be used as the state quantity.

  In the first embodiment, an example in which the engine 106 is used as an actuator that outputs a control command value has been described. However, an actuator, such as a motor as a power source, a continuously variable transmission, a friction clutch, etc., is provided in the drive system and can control the drive torque transmitted to the drive wheels by an external command. It ’s fine.

  In the first embodiment, an example in which the sprung behavior of the vehicle body is estimated using the vehicle model 307 as the vehicle body vibration estimation unit 205 is shown. However, the vehicle body vibration estimation unit may be an example in which estimation is performed using one or a plurality of equations of motion corresponding to a vehicle model.

  In the first embodiment, an example of the MT transmission 107 that manually changes the transmission gear stage is shown as the transmission. However, the transmission may be an example of an automatic transmission that automatically changes the transmission gear stage and the gear ratio.

  In the first embodiment, the vehicle system vibration control device of the present invention is applied to an engine vehicle. However, the vehicle system vibration control device of the present invention can of course be applied to a hybrid vehicle or an electric vehicle. Furthermore, in the case of a hybrid vehicle, the response performance in the torque command value calculation unit of the vehicle system vibration control device may be switched between an engine travel mode and a motor travel mode with different actuators (power sources).

101 Engine control module (ECM)
102FR, 102FL Left and right front wheels (driven wheels)
102RR, 102RL Left and right rear wheels (drive wheels)
103FR, 103FL, 103RR, 103RL Wheel speed sensor
104 Brake stroke sensor
105 Accelerator position sensor
106 engine
107 MT transmission
108 shaft
109 Differential gear
110 Steering wheel
111 Steering angle sensor
201 Driver required torque calculation section
202 Torque command value calculator
203 Vehicle control system
204 Input converter
205 Body vibration estimation unit
206 Torque command value calculator
301 Drive torque converter
302 Suspension calculation part
303 Vertical force converter
304 Body speed estimation part
305 First turning resistance calculation unit
306 Second turning resistance calculation unit
307 Vehicle model
308 First regulator
309 Second regulator
310 Third regulator
311 Limit processing section
312 Bandpass filter
313 Nonlinear Gain Amplifier
314 Limit processing section
315 Engine torque converter
316 High pass filter
317 First tuning gain setting section
318 Second tuning gain setting section
319 Third tuning gain setting section
320 adder
321 Switching processor

Claims (6)

An input conversion unit that converts sensing information from the vehicle acquired during traveling into a dimension of torque or force applied to the wheel, which is an input format to the vehicle model used when estimating the sprung behavior of the vehicle body, and the wheel using the vehicle model and the torque or force applied to includes a vehicle body vibration estimation unit that estimates a sprung mass behavior of the vehicle body, the torque command value calculating unit for correcting the drive torque based on the estimated result of the sprung mass behavior, the In the vehicle system vibration control device,
The input converter is
A first turning resistance calculation unit that calculates a first turning resistance using a steering angle signal, a vehicle speed signal, and a linear two-wheel model;
A second turning resistance force calculation unit for calculating a second turning resistance force using a three-dimensional map of a steering angle signal, a vehicle body speed signal, and a vehicle speed / steering angle / turning resistance force;
A switching processing unit for switching the turning resistance force to be output to the vehicle body vibration estimation unit based on the calculation results from the two turning resistance force calculation units and the running conditions;
A vehicle system vibration control device comprising: In the vehicle system vibration control device according to claim 1,
The switching processing unit uses the first turning resistance force to determine the turning resistance force when the vehicle speed is a medium speed or higher, and the turning resistance force when the vehicle speed is in a low speed range. The vehicle structure vibration control device using the second turning resistance force. In the vehicle system vibration control device according to claim 2,
The switching processing unit includes a low-speed → medium speed switching vehicle speed that changes from the second turning resistance force to the first turning resistance force, and a medium-speed → low speed that changes from the first turning resistance force to the second turning resistance force. A vehicle system vibration control device characterized in that a hysteresis width is provided between the vehicle speed and the switching vehicle speed. In the vehicle system vibration control device according to claim 2 or 3,
The switching processing unit sets a value during the switching transition of the turning resistance force as a composite value using a weighting coefficient for the first turning resistance force and the second turning resistance force, and smoothly changes according to a change in vehicle speed. A vehicle system vibration control device characterized by having a weighted switching configuration in which a turning resistance force changes. In the vehicle system vibration control device according to claim 4,
The vehicle switching control device, wherein the switching processing unit sets a value of the vehicle speed change gradient of the weighting factor to a value of a gentler vehicle speed change gradient as the steering angular velocity is larger. In the vehicle system vibration control device according to any one of claims 1 to 5,
The switching processing unit switches the turning resistance force when the absolute value of the torque command value calculated by the torque command value calculating unit is smaller than a threshold value, and the turning resistance when the absolute value of the torque command value is equal to or larger than the threshold value. A vehicle system vibration control device characterized in that it does not perform force switching.