JP2014133514A - Vehicle control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To suppress a drive force control using front and rear wheel load shift amounts from disturbing other functions applied to a vehicle.SOLUTION: A basis-request-torque calculation unit 11 calculates a basic request torque Ton the basis of an operator's accelerator pedal operation. A target-load-shift-amount calculation unit 19 calculates a target load shift amount Δ so that a vehicle can stably travel. A spring upper-vibration-model computing unit 121 estimates an estimated load shift amount yof the vehicle. A vibration-control correction unit 20 corrects the basic request torque Tso that the estimated load shift amount ycan follow up the target load shift amount Δ, and calculates a corrected request torque T. A target-load-shift-amount correction unit 21 and a request-torque arbitration unit 22 prohibit control over a drive force of the vehicle using the corrected request torque Tif preset prohibition conditions including a condition that at least one of a fuel cut control, a traction control, and a slide slip prevention control is executed are satisfied.

Description

  The present invention relates to a vehicle control device that controls driving force of a vehicle.

  As a method for maintaining the vehicle's rotational angular velocity (i.e., yaw rate) at an appropriate value while turning and obtaining a favorable vehicle turning performance, conventionally, the front and rear wheel loads have been changed in accordance with changes in driving force or braking force. There has been proposed a method for obtaining a desired vehicle steering characteristic by utilizing the fact that the cornering power of the tire is changed by the movement of the front and rear wheels and the generated force of the front and rear wheels is changed.

  In contrast, the applicant of the present application calculates a target value of the front and rear wheel load movement amount based on the vehicle speed V, the turning radius ρ, the steering angle δ, and the like so that the front and rear wheel load movement amount of the vehicle follows the target value. A technique for improving the vehicle turning performance by controlling the driving force has already been proposed (see, for example, Patent Document 1).

JP 2005-256636 A

  However, the technique described in Patent Document 1 is applied to improve the driving performance of the vehicle depending on the vehicle state by controlling the driving force so that the front and rear wheel load movement amount of the vehicle follows the target value. There was a risk of inhibiting other functions.

  This invention is made in view of such a problem, and provides the technique which suppresses that the other function applied to the vehicle is inhibited by the driving force control using the front-and-rear wheel load movement amount. Objective.

  The vehicle control apparatus according to claim 1, wherein the basic driving force calculation means is a driving force requested by the driver to the vehicle based on an accelerator pedal operation by the driver. Calculate the driving force. Further, the target load movement amount calculation means uses the load movement amount between the front and rear wheels of the vehicle as the front and rear wheel load movement amount, and the target load movement that is the front and rear wheel load movement amount for the vehicle to travel stably. Calculate the amount. Further, the load movement amount estimation means estimates the front and rear wheel load movement amount of the vehicle. Then, the driving force correcting means corrects the basic required driving force so that the front and rear wheel load moving amount estimated by the load moving amount estimating means follows the target load moving amount, and calculates the corrected required driving force.

  Thus, the vehicle control apparatus according to claim 1 stabilizes the steering characteristics when the vehicle travels by correcting the basic required driving force so that the front and rear wheel load movement amount of the vehicle follows the target load movement amount. Can be achieved.

  Further, in the vehicle control device according to claim 1, the prohibiting unit includes in advance that the other driving force control that is a driving force control based on a control condition other than the front and rear wheel load movement amount is executed. When the set prohibition condition is satisfied, it is prohibited to control the driving force of the vehicle using the corrected required driving force calculated by the driving force correcting means.

  Thereby, it is possible to prohibit the driving force control by the driving force correcting means when the driving force control based on the control condition other than the front and rear wheel load movement amount is being executed. For this reason, the driving force control by the driving force correcting means can suppress the occurrence of a situation where the driving force control (other driving force control) based on the control conditions other than the front and rear wheel load movement amount is hindered.

It is a block diagram which shows the structure of the electronic control apparatus 1 of 1st Embodiment. 3 is a block diagram illustrating a configuration of a running resistance disturbance estimation unit 14. FIG. 3 is a block diagram showing a configuration of a target load movement amount calculation unit 19. FIG. It is a figure explaining the parameter used with a sprung vibration model. 3 is a block diagram showing a configuration of a target load movement amount correction unit 21. FIG. 3 is a block diagram showing a configuration of a required torque arbitration unit 22. FIG.

Embodiments of the present invention will be described below with reference to the drawings.
As shown in FIG. 1, an electronic control device 1 of this embodiment is mounted on a vehicle and controls an engine 2 of the vehicle.

  The electronic control device 1 inputs signals from the accelerator stroke sensor 3, the intake air amount sensor 4, the crank angle sensor 5, the wheel speed sensor 6, the steering angle sensor 7, the vehicle speed sensor 8, and the navigation device 9.

The accelerator stroke sensor 3 outputs a signal corresponding to the depression amount of the accelerator pedal by the driver.
The intake air amount sensor 4 outputs a signal corresponding to the intake air amount to the engine 2.

The crank angle sensor 5 outputs a pulse signal that causes an edge at every predetermined angle according to the rotation of the crankshaft of the engine 2.
The wheel speed sensor 6 is attached to each of the left front wheel, the right front wheel, the left rear wheel, and the right rear wheel, and outputs a pulse signal in which an edge is generated at every predetermined angle according to the rotation of each wheel shaft.

The steering angle sensor 7 outputs a signal corresponding to the steering angle of the steering wheel of the vehicle.
The vehicle speed sensor 8 outputs a pulse signal in which an edge is generated at every predetermined angle according to the rotation of the drive shaft of the vehicle.

  The navigation device 9 acquires road map data from a map storage medium in which road map data and various types of information are recorded, and based on a GPS signal received via a GPS (Global Positioning System) antenna (not shown) or the like. It is configured to detect the current position and execute route guidance from the current position to the destination. The road map data includes various data such as road position, road type (highway, toll road, general road, etc.), road shape, road width, road name, number of lanes, road gradient, and the like. The navigation device 9 includes a yaw rate sensor (not shown) that detects the yaw rate of the vehicle.

  The electronic control device 1 includes a basic required torque calculation unit 11, an estimated driving wheel torque calculation unit 12, a wheel speed calculation unit 13, a running resistance disturbance estimation unit 14, a steering angle calculation unit 15, a vehicle speed calculation unit 16, and a road gradient acquisition unit. 17, a virtual turning radius calculation unit 18, a target load movement amount calculation unit 19, a vibration suppression correction unit 20, a target load movement amount correction unit 21, and a required torque arbitration unit 22.

The basic required torque calculation unit 11 calculates the accelerator pedal depression amount based on the signal from the accelerator stroke sensor 3, and further calculates the torque T _{w_tgt} applied to the drive shaft of the vehicle based on the accelerator pedal depression amount. The accelerator pedal depression amount corresponds to the torque request by the driver, and the torque T _{w_tgt} is a basis for torque correction by the vibration suppression correction unit 20. Therefore, hereinafter, the torque calculated by the basic required torque calculation unit 11 is referred to as basic required torque T _{w_tgt} .

The estimated drive wheel torque calculation unit 12 first calculates the intake air amount based on the signal from the intake air amount sensor 4 and also determines the rotational speed Ne (hereinafter referred to as the engine 2) of the engine 2 based on the signal from the crank angle sensor 5. Rotational speed) is calculated. Then, the estimated drive wheel torque calculation unit 12 calculates the torque T _{w_est} applied to the drive shaft of the vehicle based on the calculated intake air amount and the engine rotation speed. The estimated drive wheel torque calculation unit 12 estimates the torque generated in the engine 2 and applied to the drive shaft based on the intake air amount and the engine rotation speed. Hereinafter, this torque is referred to as estimated drive wheel torque T _{w_est} .

Based on the signal from the wheel speed sensor 6, the wheel speed calculation unit 13 determines the wheel speed V _{fl of} the left front wheel (hereinafter referred to as the left front wheel speed V _fl ), the wheel speed V _{fr of} the right front wheel (hereinafter, the right front wheel speed V). _fr ), wheel speed V _{rl of} the left rear wheel (hereinafter referred to as left rear wheel speed V _rl ) and wheel speed V _{rr of} the right rear wheel (hereinafter referred to as right rear wheel speed V _rr ).

  As shown in FIG. 2, the running resistance disturbance estimation unit 14 includes average processing units 31, 32, 33, subtractors 34, 35, differentiators 36, 37, low-pass filters 38, 39, and amplifiers 40, 41. With.

The average processing unit 31 calculates an average value of the left front wheel speed V _fl and the right front wheel speed V _fr .
The average processing unit 32 calculates an average value of the left rear wheel speed V _rl and the right rear wheel speed V _rr .
The average processing unit 33 calculates an average value of the left front wheel speed V _fl , the right front wheel speed V _fr , the left rear wheel speed V _rl, and the right rear wheel speed V _rr .

  The subtractor 34 calculates a value obtained by subtracting the calculated value of the average processing unit 33 from the calculated value of the average processing unit 31. The subtractor 35 calculates a value obtained by subtracting the calculated value of the average processing unit 33 from the calculated value of the average processing unit 32.

Differentiators 36 and 37 calculate differential values of the calculated values of the subtractors 34 and 35, respectively.
The low-pass filters 38 and 39 remove high-frequency components from the calculated values of the differentiators 36 and 37, respectively, and output them.

The amplifiers 40 and 41 respectively output the signals input from the low-pass filters 38 and 39 by multiplying the vehicle mass by M.
Therefore, the running resistance disturbance estimation unit 14 estimates the difference between the average value of the wheel speeds of the front wheels and the average value of the wheel speeds of the four wheels multiplied by the vehicle weight as the running resistance disturbance of the front wheels. The difference between the average value of the wheel speeds of the rear wheels and the average value of the wheel speeds of the four wheels multiplied by the vehicle weight is estimated as the disturbance resistance of the rear wheels.

As shown in FIG. 1, the steering angle calculation unit 15 calculates a steering wheel steering angle δ _n based on a signal from the steering angle sensor 7.
The vehicle speed calculation unit 16 calculates a traveling speed V of the vehicle (hereinafter referred to as a vehicle speed V) based on a signal from the vehicle speed sensor 8.

The road gradient acquisition unit 17 acquires the road gradient φ at the current position of the vehicle from the navigation device 9.
Based on the yaw rate γ acquired from the navigation device 9 and the vehicle speed V acquired from the vehicle speed calculation unit 16, the virtual turning radius calculation unit 18 is a virtual turning radius suitable for the vehicle to travel (hereinafter referred to as virtual turning radius). Ρ) (referred to as radius) is calculated using the following equation (1).

ρ = V / γ (1)
As shown in FIG. 3, the target load movement amount calculation unit 19 includes a cornering power calculation unit 51, a steering angle dependent parameter calculation unit 52, and a multiplier 53.

  The cornering power calculation unit 51 includes a front and rear wheel static load calculation unit 61, a road gradient influence term calculation unit 62, a static ground load change amount calculation unit 63, multipliers 64, 65, and 68, adders 66 and 67, and an amplifier 69. Prepare.

The front and rear wheel static load calculation unit 61 includes a constant output device 71 and amplifiers 72 and 73.
Constant output device 71 outputs a fixed value indicating the static load W _o of the vehicle. The static load _Wo of the vehicle is expressed by the following equation (2), where M is the mass of the vehicle and g is the acceleration of gravity.

W _o = M × g (2)
The amplifier 72 has the wheel base of the vehicle as L (see FIG. 4), the distance between the center of gravity of the vehicle and the rear wheel shaft as L _r (see FIG. 4), and the signal inputted from the constant output device 71 (L _r / L) Multiply and output. This output value corresponds to the static load _Wfo of the front wheel and is expressed by the following equation (3).

W _fo = (L _r / L) W _o (3)
The amplifier 73 sets the distance between the center of gravity of the vehicle and the front wheel shaft to L _f (see FIG. 4), and multiplies the signal input from the constant output device 71 by (L _f / L) and outputs it. This output value corresponds to the static load W _ro of the rear wheel and is expressed by the following expression (4).

W _ro = (L _f / L) W _o (4)
Accordingly longitudinal HanawaShizu load calculating section 61 outputs a value indicating wheel static load W _fo and Kowasei load W _ro.

The road gradient influence term calculation unit 62 includes a cosine function calculator 81, a sine function calculator 82, amplifiers 83 and 84, a subtracter 85, and an adder 86.
The cosine function calculator 81 inputs the road gradient φ from the road gradient acquisition unit 17 and performs a cosine function (cos) calculation. That is, the cosine function calculator 81 outputs cosφ.

The sine function calculator 82 inputs the road gradient φ from the road gradient acquisition unit 17 and performs a sine function (sin) calculation. That is, the sine function calculator 82 outputs sinφ.
The amplifier 83 sets the distance between the road surface and the center of gravity of the vehicle (hereinafter referred to as the height of the center of gravity) as h _cg (see FIG. 4), and the signal input from the sine function calculator 82 is multiplied by (h _cg / L _r ). And output. The amplifier 84 multiplies the signal input from the sine function calculator 82 by (h _cg / L _f ) and outputs the result.

The subtracter 85 calculates a value obtained by subtracting the output value of the amplifier 83 from the output value of the cosine function calculator 81. That is, the subtractor 85 outputs {cos φ− (h _cg / L _r ) sin φ}.
The adder 86 calculates a value obtained by adding the output value of the cosine function calculator 81 and the output value of the amplifier 83. That is, the adder 86 outputs {cos φ + (h _cg / L _f ) sin φ}.

Therefore, the road gradient influence term calculation unit 62 outputs {cosφ− (h _cg / L _r ) sinφ} and {cosφ + (h _cg / L _f ) sinφ}.
The static ground load change amount calculation unit 63 includes a sprung vibration model steady solution calculation unit 91 and front and rear wheel load change amount calculation units 92.

The sprung vibration model steady solution calculation unit 91 uses the sprung vibration model to calculate the vertical displacement xν of the vehicle body (see FIG. 4) and the pitch angle θ _p (see FIG. 4) around the pitching center of the vehicle body. For each, a value in the steady state is calculated.

  In the sprung vibration model, as shown in FIG. 4, the front wheel and the vehicle body of the vehicle and the rear wheel and the vehicle body are connected by a suspension in which a predetermined spring constant and a predetermined damping coefficient are set. Assuming that the vehicle vibrates with pitching, the vehicle state of the vehicle is expressed by a state equation.

And the state equation about the vertical displacement xν and the pitch angle θ _p of the vehicle body is expressed by the following equation (5). Here, xν ′ and xν ″ represent the first and second derivatives of xν, respectively. θ _p ′ and θ _p ″ indicate the first and second derivatives of θ _p , respectively. The _{a 1 ~ 4, b 1 ~} 4, p 1 ~ 3 are constants that are set in advance. ΔF _df and ΔF _dr are changes in translational force acting on the front wheel shaft and the rear wheel shaft, respectively. [Delta] T _w is the change amount of the estimated drive wheel torque _T w_est.

When the vehicle is in a steady state, xν ′, xν ″, θ _p ′, and θ _p ″ are 0. Therefore, in equation (5), xν ′, xν ″, θ _p ′, θ _p substituting 0 into '', the following equation (6), as shown in (7), is obtained and the stationary solution theta _{p_s} of stationary solutions Xnyu _{_s} and theta _p of Xnyu.

Sprung vibration model constant solution calculating section 91 thus has the formula (6) with (7), the stationary solution Xnyu _{_s,} calculates the theta _{p_s.}
Front and rear wheel load variation calculating unit 92, Equation (8), with (9), calculates the front wheel static ground load change amount [Delta] W _{f_s,} the rear wheel static ground load change amount [Delta] W _{r_s.} Here, K _sf is the spring constant of the suspension on the front wheel side, and K _sr is the spring constant of the suspension on the rear wheel side (see FIG. 4).

Thus static ground load change amount calculating unit 63, as shown in FIG. 3, and outputs a static ground load change amount [Delta] W _{r_s} front wheel static ground load change amount [Delta] W _{f_s} and rear wheels.
The multiplier 64 calculates a multiplication value of the output value from the amplifier 72 and the output value from the subtracter 85. That is, the multiplier 64 outputs W _fo {cos φ− (h _cg / L _r ) sin φ}. This value corresponds to the ground load applied to the front wheels along the direction perpendicular to the road surface traveling on the road surface having the road gradient φ.

The multiplier 65 calculates a multiplication value of the output value from the amplifier 73 and the output value from the adder 86. That is, the multiplier 65 outputs W _ro {cos φ + (h _cg / L _f ) sin φ}. This value corresponds to the ground contact load applied to the rear wheel along the road surface vertical direction while traveling on the road surface of road gradient φ.

The adder 66 calculates an addition value between the output value from the multiplier 64 and the static ground load change amount ΔW _{f — s} from the static ground load change amount calculation unit 63. This value is the total value W _f of the front wheel static ground load with the load movement caused by the steady driving torque added to the ground load of the front wheel corresponding to the road gradient φ in the steady state (hereinafter, the front wheel static load). corresponds to) that the ground load W _f.

The adder 67 calculates an addition value between the output value from the multiplier 65 and the static ground load change amount ΔW _{r — s} from the static ground load change amount calculation unit 63. This value is a total value W _r (hereinafter, referred to as “rear wheel static ground load”) in which the load movement caused by the steady driving torque is added to the ground load of the rear wheel corresponding to the road gradient φ in the steady state. This corresponds to the rear wheel static ground load _Wr ).

Multiplier 68 calculates a multiplication value of front wheel static ground load W _f from adder 66 and rear wheel static ground load W _r from adder 67.
Amplifier 69 outputs a signal inputted from the multiplier 68 C _w ² multiplied by. C _w is a load dependent coefficient of cornering power. That is, the amplifier 69 outputs (C _w ² × W _f × W _r ).

The front wheel cornering power _Cpf and the rear wheel cornering power _Cpr are expressed by the following equations (10) and (11), respectively.
C _pf = C _w W _f (10)
C _pr = C _w W _r (11)
Accordingly, the cornering power calculation unit 51 outputs a value C _pf C _pr obtained by multiplying the front wheel cornering power C _pf by the rear wheel cornering power C _pr .

  Next, the steering angle dependent parameter calculation unit 52 includes amplifiers 101, 104, and 110, an absolute value calculator 102, a multiplier 103, a constant output unit 105, a subtractor 106, a square calculator 107, a zero division preventer 108, and a divider. 109.

The amplifier 101 sets the steering gear ratio to R _s , multiplies the steering wheel steering angle δ _n input from the steering angle calculation unit 15 by (1 / R _s ), and outputs the result as the steering angle δ of the front wheels.
The absolute value calculator 102 calculates and outputs the absolute value of the output value from the amplifier 101.

The multiplier 103 calculates a multiplication value of the absolute value of the steering angle δ and the virtual turning radius ρ from the virtual turning radius calculation unit 18.
The amplifier 104 multiplies the output value from the multiplier 103 by (1 / L) and outputs the result.

The constant output unit 105 outputs a preset constant 1.
The subtractor 106 calculates a value obtained by subtracting the output value from the constant output unit 105 from the output value from the amplifier 104. That is, the subtractor 106 outputs (ρ | δ | / L−1).

The square calculator 107 receives the vehicle speed V from the vehicle speed calculator 16 and performs a square calculation. That is, the square calculator 107 outputs the V ^2.
The zero division preventer 108 performs zero division prevention on the output value V ² from the square calculator 107.

The divider 109 outputs a value obtained by dividing the output value from the subtractor 106 by the output value from the zero division preventer 108.
The amplifier 110 multiplies the output value from the divider 109 by a coefficient k expressed by the following equation (12) and outputs the result.

k = −2L ² / MC _w (12)
Therefore, the steering angle dependent parameter calculation unit 52 outputs {2L ² (1-ρ | δ | / L) / MC _w V ² }.

The multiplier 53 calculates and outputs a multiplication value of the output value from the cornering power calculation unit 51 and the output value from the steering angle dependent parameter calculation unit 52.
As described above, the target load movement amount calculation unit 19 outputs a value represented by the following expression (13) as the target load movement amount Δ.

Δ = (2L ² C _pf C _pr / MC _w V ² ) × (1−ρ | δ | / L) (13)
The turning radius ρ of the vehicle is expressed by the following formula (14).

Equation (14) shows that the value of the turning radius with respect to increase / decrease in vehicle speed with a constant rudder angle depends on the value of (L _f C _pf −L _r C _pr ). Specifically, when (L _f C _pf −L _r C _pr ) is smaller than 0, the turning radius ρ increases as the vehicle speed V increases, that is, understeer, and (L _f C _pf −L _r When C _pr ) is greater than 0, the turning radius ρ decreases as the vehicle speed V increases, that is, oversteer. That is, the steering characteristic can be controlled using (L _f C _pf −L _r C _pr ).

Further, (L _f C _pf −L _r C _pr ) is expressed by the following equation (15) using equations (10) and (11).
_{(L f C pf -L r C} pr) = C w (L f W f -L r W r) ··· (15)
Therefore, (L _f W _f −L _r W _r ) in the equation (15) is expressed by the following equation (16).

As can be understood by comparing Expression (13) and Expression (16), the target load movement amount calculation unit 19 calculates (L _f W _f −L _r W _r ) as the target load movement amount Δ. ing.
Next, as shown in FIG. 1, the vibration suppression correction unit 20 includes a sprung vibration model calculation unit 121, subtractors 122, 127, 130, an integrator 123, amplifiers 124, 125, 128, 129, and an adder / subtractor 126. .

The sprung vibration model calculation unit 121 calculates the state quantity x = (xν, xν ′, θ _p , θ _p ′) by the above equation (5). The sprung vibration model calculation unit 121 calculates the estimated load movement amount y _s by the following equation (17) using the state quantity x calculated by the above equation (5). Here, K _sf is the spring constant of the suspension on the front wheel side, K _sr is the spring constant of the suspension on the rear wheel side, C _sf is the damping coefficient of the suspension on the front wheel side, and C _sr is the damping coefficient of the suspension on the rear wheel side. (See FIG. 4).

Subtractor 122 calculates the output value from the target load shift amount correcting unit 21, a value obtained by subtracting the estimated load shift amount y _s from sprung vibration model calculating unit 121, i.e. the model output error.
The integrator 123 integrates the output error from the subtractor 122. The amplifier 124 multiplies the integral value input from the integrator 123 by a preset integral gain K _i and outputs it. Therefore, the vibration damping correction unit 20 is a type 1 servo system.

The amplifier 125 multiplies the state quantity x output from the sprung vibration model calculation unit 121 by a preset state feedback gain K _s and outputs the result.
The adder / subtractor 126 adds the output value from the amplifier 124 and the running resistance disturbance from the running resistance disturbance estimation unit 14. Further, the adder / subtractor 126 calculates a value obtained by subtracting the output value from the amplifier 125 from the added value, and outputs the subtracted value to the sprung vibration model calculation unit 121 and the subtractor 127. The output value of the adder / subtractor 126 is an absolute value of driving torque (hereinafter referred to as corrected driving torque) corrected in view of the running resistance disturbance and the state quantity x estimated to follow the target load movement amount.

The subtractor 127 subtracts the correction drive torque from the adder / subtractor 126 from the torque T _{w_est} from the estimated drive wheel torque calculation unit 12 to calculate a relative drive torque correction value.
The amplifier 128 sets the reduction ratio (final gear ratio) in the final reduction gear as R _d , and outputs the output value from the subtractor 127 multiplied by (1 / R _d ). The amplifier 129 _multiplies the basic required torque T _{w — tgt} from the basic required torque calculation unit 11 by (1 / R _d ) and outputs the result.

The subtractor 130 calculates a value obtained by subtracting the output value from the amplifier 128 from the output value from the amplifier 129. The output of the subtracter 130 is the corrected required torque, and hereinafter referred to as the corrected required torque T _req .

Accordingly, the vibration damping correction unit 20 outputs the corrected required torque T _req .
The target load movement amount correction unit 21 includes a fuel cut correction factor calculation unit 141 and a multiplier 142 as shown in FIG.

The fuel cut correction factor calculation unit 141 is a fuel cut control state that indicates the target load movement amount Δ from the target load movement amount calculation unit 19 and whether or not the electronic control device 1 is performing fuel cut control on the engine 2. A fuel cut correction factor F _c is calculated based on the flag. Specifically, the target load movement amount correction unit 21 corrects the fuel cut when the fuel cut control state flag is 1 and the target load changes in the post-load direction, that is, the target load movement amount Δ decreases. Set factor F _c to zero. On the other hand, the fuel cut correction factor F _c is set to 1 except the above, that is, except when the fuel cut control state flag is 1 and the target load movement amount Δ is decreasing.

The multiplier 142 calculates and outputs a multiplication value of the target load movement amount Δ from the target load movement amount calculation unit 19 and the fuel cut correction factor F _c from the fuel cut correction factor calculation unit 141.
Therefore, the target load movement amount correction unit 21 outputs the target load movement amount Δ when the fuel cut correction factor F _c becomes 1, while it outputs 0 when the fuel cut correction factor F _c becomes zero.

The required torque arbitration unit 22 includes a first torque selection unit 151 and a second torque selection unit 152, as shown in FIG.
The first torque selection unit 151 determines the corrected request torque T based on the corrected request torque T _req from the vibration suppression correction unit 20, the traction control request torque T _trc and the traction control state flag from the traction control device 161. Select _req or traction control required torque T _trc and output. The traction control device 161 is mounted on the vehicle, and performs control for suppressing the torque of the engine 2 in order to prevent, for example, tire slipping during acceleration of the vehicle. Then, the traction control device 161 provides an electronic control device with a traction control request torque T _trc required for the engine 2 for traction control and a traction control state flag indicating whether or not the traction control device 161 is performing traction control. Output to 1.

Specifically, the first torque selection unit 151 selects the traction control request torque T _trc when the traction control state flag is 1 and the corrected request torque T _req is greater than the traction control request torque T _trc . On the other hand, except for the above case, that is, when the traction control state flag is 1 and the corrected required torque T _req is larger than the traction control required torque T _trc , the corrected required torque T _req is selected. Hereinafter, the torque selected and output by the first torque selection unit 151 is referred to as first selection torque.

Based on the first selected torque from the first torque selecting unit 151, the skid prevention request torque T _vsc and the skid prevention control state flag from the skid prevention control device 162, the second torque selection unit 152 _{Alternatively} , the skid prevention request torque T _vsc is selected and output.

The skid prevention control device 162 is mounted on the vehicle and performs control for preventing the skid of the vehicle when the vehicle is turning. Then, the skid prevention control device 162 indicates the skid prevention required torque T _vsc required for the engine 2 for the skid prevention control, and whether or not the skid prevention control device 162 performs the skid prevention control. The flag is output to the electronic control unit 1.

Specifically, the second torque selection unit 152 selects the skid prevention request torque T _vsc when the skid prevention control state flag is 1, and selects the first selected torque when the skid prevention control state flag is 0. select. That is, the required torque arbitration unit 22 is performing the skid prevention control.
It is configured not to accept a request torque by other control.

  Then, the electronic control unit 1 includes a throttle motor (not shown) that changes the opening of the throttle valve based on the required torque output from the required torque adjuster 22, and an ignition plug (not shown) for igniting the fuel in each cylinder. ) And various actuators such as an injector (not shown) for injecting fuel into each cylinder, and the engine 2 is operated.

In the electronic control device 1 configured as described above, the basic required torque calculation unit 11 calculates the basic required torque T _{w_tgt} based on the accelerator pedal operation by the driver. The target load movement amount calculation unit 19 calculates the target load movement amount Δ so that the vehicle travels stably, and the sprung vibration model calculation unit 121 estimates the estimated load movement amount y _s of the vehicle. The vibration damping correcting unit 20, the estimated load shift amount y _s is correcting the basic request torque T _{W_tgt} so as to follow the target load shift amount delta, calculates the corrected requested torque T _req.

In this way, the electronic control unit 1 can stabilize the steering characteristics when the vehicle is traveling.
Further, in the electronic control unit 1, the target load movement amount correction unit 21 and the required torque arbitration unit 22 are set in advance including a condition that at least one of fuel cut control, traction control, and skid prevention control is executed. When the condition is satisfied, it is prohibited to control the driving force of the vehicle using the corrected required torque T _req calculated by the vibration suppression correction unit 20.

Thereby, the occurrence of a situation where the fuel cut control, the traction control, and the skid prevention control are hindered by the driving force control by the vibration suppression correction unit 20 can be suppressed.
Specifically, after the first torque selection unit 151 is executing traction control (the traction control state flag is 1) and the corrected required torque T _req is greater than the traction control required torque T _trc , the corrected The traction control required torque T _trc is selected instead of the required torque T _req . As a result, control using the corrected required torque T _req calculated by the vibration suppression correction unit 20 is prohibited.

This is because when the traction control is activated due to the tire idling and the traction control required torque T _{trc for reducing} the driving force is output from the traction control device 161, the corrected required torque T _req is the traction. This is because if the corrected required torque T _req is selected when the control required torque T _{trc is} greater than the required control torque T _trc , the driving force does not decrease sufficiently until gripping, and the effect of traction control may not be exhibited. However, since a torque lower than the corrected required torque T _req is required for the engine 2, there is a possibility that the vehicle stabilization function cannot be sufficiently exhibited. However, since the traction control is a control related to safety, the traction control is given priority.

Further, when the corrected required torque T _req is equal to or less than the traction control required torque T _trc , the corrected required torque T _req is selected. Thereby, the torque required for the engine 2 becomes lower than the torque required for the traction control. However, the function of gripping the tire can be performed, and the vehicle stabilization function is also exhibited by executing the driving force control using the corrected required torque T _req calculated by the vibration damping correction unit 20.

Further, when the side slip prevention control is being executed (side slip prevention control state flag is 1), the second torque selection unit 152 selects the side slip prevention request torque T _vsc instead of the corrected request torque T _req . As a result, control using the corrected required torque T _req calculated by the vibration suppression correction unit 20 is prohibited.

This is because when the braking force (brake) and the driving force are appropriately controlled by the skid prevention control device 162 in order to turn the vehicle in the target traveling direction, the vehicle stabilization function can be used in any situation. This is because when the corrected required torque T _req is applied, the skid prevention function may not work properly.

The target load movement amount correction unit 21 sets the target load movement amount Δ to 0 when the fuel cut control is being executed (the fuel cut control state flag is 1) and the target load movement amount Δ is decreasing. Set to. As a result, control using the corrected required torque T _req calculated by the vibration suppression correction unit 20 is prohibited.

The fuel cut control is executed when the accelerator is turned off. When the steering angle is returned during the fuel cut period in normal forward travel, it is desirable to bring the steering characteristics in the understeer direction. Therefore, the target load changes in the post-load direction, that is, the target load movement amount Δ decreases, and the vibration correction controller 20 calculates the post-correction required torque T _req so that the driving force increases. Ends. Thereafter, when the vehicle goes straight and the correction by the vibration suppression correction unit 20 is completed, the fuel cut control is executed again. Here, when the steering angle is turned back, the corrected required torque T _req is calculated so that the driving force increases again, so that the fuel cut control ends. In this way, the fuel cut control may be hunted. Therefore, when the fuel cut control is being executed and the target load movement amount Δ is decreasing, the target load movement amount Δ is set to zero. This prevents hunting of fuel cut control. If the target load movement amount Δ does not decrease even during the fuel cut control, the corrected required torque is output by outputting the target load movement amount Δ input from the target load movement amount calculation unit 19. Occurrence of a situation where control using T _req is prohibited uselessly is suppressed.

In the embodiment described above, the electronic control device 1 is the vehicle control device in the present invention, the basic required torque calculation unit 11 is the basic driving force calculation means in the present invention, and the target load movement amount calculation unit 19 is the target load movement amount in the present invention. The calculation means, the sprung vibration model calculation section 121 is the load movement amount estimation means in the present invention, the vibration suppression correction section 20 is the driving force correction means, the target load movement amount correction section 21 and the required torque adjustment section 22 in the present invention. The basic required torque T _{w_tgt} is the basic required driving force in the present invention, the corrected required torque T _req is the corrected required driving force in the present invention, the traction control, the skid prevention control and the fuel cut control are the other driving in the present invention. The force control and traction control required torque T _trc is the traction control required driving force in the present invention.

  As mentioned above, although one Embodiment of this invention was described, this invention is not limited to the said embodiment, As long as it belongs to the technical scope of this invention, a various form can be taken.

  DESCRIPTION OF SYMBOLS 1 ... Electronic controller, 11 ... Basic required torque calculation part, 19 ... Target load movement amount calculation part, 20 ... Damping correction part, 21 ... Target load movement amount correction part, 22 ... Request torque mediation part, 121 ... On spring Vibration model calculator

Claims (4)

A vehicle control device (1) mounted on a vehicle for controlling the driving force of the vehicle,
Basic driving force calculating means (11) for calculating a basic required driving force which is a driving force required by the driver for the vehicle based on an accelerator pedal operation by the driver;
A target load for calculating a target load movement amount, which is the front and rear wheel load movement amount for the vehicle to travel stably, with the load movement amount between the front and rear wheels of the vehicle as a front and rear wheel load movement amount. A movement amount calculating means (19);
Load movement amount estimation means (121) for estimating the front and rear wheel load movement amount of the vehicle;
Driving force correction means for correcting the basic required driving force and calculating the corrected required driving force so that the front and rear wheel load moving amount estimated by the load moving amount estimating means follows the target load moving amount ( 20)
The driving force correction is performed when a preset prohibition condition is satisfied that includes a condition that another driving force control that is a control of the driving force based on a control condition other than the front and rear wheel load movement amount is executed. A vehicle control apparatus comprising: prohibiting means (21, 22) for prohibiting control of the driving force of the vehicle using the corrected required driving force calculated by a means. The other driving force control controls the driving force of the vehicle by traction control, calculates a traction control required driving force that is a driving force required for the vehicle for the traction control,
The prohibition condition is
The vehicle control device according to claim 1, wherein the traction control is being executed, and the corrected required driving force is larger than the traction control required driving force. The other driving force control controls the driving force of the vehicle by a skid prevention control,
The prohibition condition is
The vehicle control device according to claim 1 or 2, wherein the skid prevention control is being executed. The other driving force control controls the driving force of the vehicle by fuel cut control,
The prohibition condition is
The vehicle control device according to any one of claims 1 to 3, wherein the fuel cut control is being executed, and the target load movement amount is decreasing.

Images