JP6172071B2 - Vehicle control device - Google Patents

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.

  On the other hand, the applicant of the present invention drives the stability factor, which is the controlling factor, so that a desired steer characteristic with respect to the target turning radius can be obtained using a theoretical relationship between the steering characteristic of the vehicle and the turning radius. There has already been proposed a control system that adjusts by front and rear wheel load movement accompanying force manipulation (see, for example, Patent Document 1). In this control system, in the calculation process of the target stability factor, the target value is determined by the product of the target turning radius ρ and the steering angle δ and the vehicle speed V based on a strict theoretical formula.

JP 2005-256636 A

  However, the technique described in Patent Document 1 requires the product calculation of the target turning radius ρ and the steering angle δ. For this reason, in the straight traveling state, the calculation accuracy of the product of ρ = ∞ and δ = 0 greatly affects the accuracy of the front and rear wheel load movement amount target value. Due to a slight shift in the phase between the two, for example, the target value in the understeer direction is calculated at the timing of the steering wheel turning operation at the beginning of the turn, and the yaw response is delayed. As a result, a sufficient rear wheel load cannot be obtained by calculation and problems such as instability of the vehicle body posture during acceleration occur.

  The present invention has been made in view of such problems, and an object thereof is to achieve both responsiveness and stability of vehicle control using front and rear wheel load movement amounts.

  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 that is the front and rear wheel load movement amount for the vehicle to travel stably. While calculating the movement amount, the load movement amount estimating means estimates the front and rear wheel load movement amount of the vehicle. 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 calculated by the target load moving amount calculating means. .

When the front wheel cornering power is C _pf , the rear wheel cornering power is C _pr , the distance between the vehicle center of gravity and the front wheel shaft is L _f , and the distance between the vehicle center of gravity and the rear wheel shaft is L _r The wheel load movement amount is expressed as (L _f C _pf −L _r C _pr ).

  Therefore, when the front and rear wheel load movement amount is smaller than 0, the load is applied to the rear wheel rather than the front wheel, so that the turning radius is increased (understeer). On the other hand, when the front and rear wheel load movement amount is larger than 0, the load is applied to the front wheel rather than the rear wheel, so the turning radius becomes smaller (oversteer). Thus, the steering characteristics can be controlled using the front and rear wheel load movement amount.

  Then, the target load movement amount is set so that the vehicle travels stably, and the basic required driving force is corrected so that the front and rear wheel load movement amount of the vehicle follows the target load movement amount. Steering characteristics can be stabilized.

  Note that the applicant of the present application has a steering power that includes both a steering amount by the driver's steering operation and a steering speed index having information capable of discriminating the steering directionality (that is, the cutting operation and the switching operation). It was found that it has utility as a physical quantity that reflects the driver's steering intention.

For example, the product δ (dδ / dt) of the steering angle δ and the time change rate (dδ / dt) of the steering angle is, as is clear from the following equations (18) and (19), That is, it can be regarded as a physical quantity corresponding to the steering power. Here, G _s is a constant corresponding to the steering system torsional rigidity. ^{_{Further, (G s δ 2/2}} ) is the steering potential energy.

The target load movement amount calculating means calculates the target load movement amount based on a physical quantity corresponding to the steering power of the driver.

  As a result, when calculating the target load movement amount, it is not necessary to use the turning radius ρ as information capable of discriminating the steering directionality, and the product calculation of the turning radius ρ and the steering angle δ becomes unnecessary. For this reason, the calculation accuracy of the product of the turning radius ρ = ∞ and the steering angle δ = 0 does not affect the calculation accuracy of the target load movement amount, and the responsiveness of vehicle control using the front and rear wheel load movement amounts And stability.

The target load movement amount calculation means determines whether the target load movement amount is positive or negative, and changes the magnitude of the target load movement amount based on the positive / negative determination result.
As a result, the basic required driving force can be corrected independently of each other between the case where the target load movement amount is larger than 0 and the case where the target load movement amount is smaller than 0, and the front and rear wheel load movement amounts are used. The responsiveness and stability of vehicle control can be further improved.

1 is a block diagram showing a configuration of an electronic control device 1. FIG. 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 18. FIG. It is a figure explaining the parameter used with a sprung vibration model. It is a graph which shows the specific example of target load movement amount (DELTA) and correction | amendment drive torque. Factor p _+, p _- is a graph showing a specific example of a change in the target load shift amount Δ by. It is a figure explaining the parameter used with the energy which contributes to the acceleration and load movement of the front-back direction of a vehicle body. It is a block diagram which shows the structure of the target load movement amount calculation part 18 of another embodiment. It is a graph which shows the relationship between coefficient p _<+> , p _{< - >} and vehicle speed.

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. In addition to the detection method described above, the vehicle speed sensor 8 includes the left and right average values of the rolling wheels among the detection values of the four wheel speed sensors, and the average value of the four wheel speed detection values in the case of an all-wheel drive vehicle. It may be substituted by a method such as calculating.

  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 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 target load movement amount calculation unit 18 and a vibration suppression correction unit 19 are provided.

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. To do. The accelerator pedal depression amount corresponds to the torque demand by the driver, and the torque T _{w — tgt} is a basis for torque correction by the vibration damping correction unit 19. 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 a 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 averaging processing units 21, 22, 23, subtractors 24, 25, differentiators 26, 27, low-pass filters 28, 29, and amplifiers 30, 31. With.

The average processing unit 21 calculates an average value of the left front wheel speed V _fl and the right front wheel speed V _fr .
The average processing unit 22 calculates an average value of the left rear wheel speed V _rl and the right rear wheel speed V _rr .
The average processing unit 23 calculates average values 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 24 calculates a value obtained by subtracting the calculated value of the average processing unit 23 from the calculated value of the average processing unit 21. The subtracter 25 calculates a value obtained by subtracting the calculated value of the average processing unit 23 from the calculated value of the average processing unit 22.

Differentiators 26 and 27 calculate differential values of the calculated values of the subtractors 24 and 25, respectively.
The low pass filters 28 and 29 remove the high frequency components from the calculated values of the differentiators 26 and 27, respectively, and output the result.

The amplifiers 30 and 31 respectively output the signals input from the low-pass filters 28 and 29 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.
As shown in FIG. 3, the target load movement amount calculation unit 18 includes a cornering power calculation unit 41, a steering angle dependent parameter calculation unit 42, and a multiplier 43.

  The cornering power calculation unit 41 includes a front and rear wheel static load calculation unit 51, an influence term calculation unit 52 due to road gradient, a static ground load change amount calculation unit 53, multipliers 54, 55, and 58, adders 56 and 57, and an amplifier 59. Is provided.

The front and rear wheel static load calculation unit 51 includes a constant output device 61 and amplifiers 62 and 63.
Constant output device 61 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 (1), where M is the mass of the vehicle and g is the acceleration of gravity.

W _o = M × g (1)
The amplifier 62 assumes that the wheel base of the vehicle is L (see FIG. 4), the distance between the center of gravity of the vehicle and the rear wheel shaft is L _r (see FIG. 4), and the signal input from the constant output device 61 (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 (2).

W _fo = (L _r / L) W _o (2)
The amplifier 63 multiplies the signal input from the constant output device 61 by (L _f / L) and outputs the distance between the center of gravity of the vehicle and the front wheel shaft as L _f (see FIG. 4). This output value corresponds to the static load W _ro of the rear wheel and is expressed by the following expression (3).

W _ro = (L _f / L) W _o (3)
Accordingly longitudinal HanawaShizu load calculating section 51 outputs a value indicating wheel static load W _fo and Kowasei load W _ro.

The road term influence term calculation unit 52 includes a cosine function calculator 71, a sine function calculator 72, amplifiers 73 and 74, a subtractor 75, and an adder 76.
The cosine function calculator 71 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 71 outputs cosφ.

The sine function calculator 72 receives the road gradient φ from the road gradient acquisition unit 17 and performs a sine function (sin) calculation. That is, the sine function calculator 72 outputs sinφ.
The amplifier 73 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 72 is multiplied by (h _cg / L _r ). And output. The amplifier 74 multiplies the signal input from the sine function calculator 72 by (h _cg / L _f ) and outputs the result.

The subtractor 75 calculates a value obtained by subtracting the output value of the amplifier 73 from the output value of the cosine function calculator 71. That is, the subtractor 75 outputs {cos φ− (h _cg / L _r ) sin φ}.
The adder 76 calculates a value obtained by adding the output value of the cosine function calculator 71 and the output value of the amplifier 73. That is, the adder 76 outputs {cos φ + (h _cg / L _f ) sin φ}.

Therefore, the influence term calculation unit 52 due to the road gradient outputs {cosφ− (h _cg / L _r ) sinφ} and {cosφ + (h _cg / L _f ) sinφ}.
The static ground load change amount calculation unit 53 includes a sprung vibration model steady solution calculation unit 81 and front and rear wheel load change amount calculation units 82.

The sprung vibration model steady solution calculation unit 81 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.

The state equation for the vertical displacement xν and the pitch angle θ _p of the vehicle body is expressed by the following equation (4). 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. ΔT _w is a change amount of the estimated driving wheel torque T _{w — est} .

When the vehicle is in a steady state, xν ′, xν ″, θ _p ′, θ _p ″ is 0. Therefore, in equation (4), xν ′, xν ″, θ _p ′, θ _p substituting 0 into '', the following equation (5), as shown in (6), is obtained and the stationary solution theta _p _ _s of stationary solutions Xnyu_ _s and theta _p of Xnyu.

Therefore, the sprung vibration model steady solution calculation unit 81 calculates the steady solution xν_ _s , θ _p _ _s using the equations (5) and (6).

Front and rear wheel load change amount calculating unit 82, Equation (7), using (8), and _ _s front wheel static ground load change amount [Delta] W _f, of the rear wheel static ground load change amount [Delta] W _r _ _s calculate. 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).

Therefore, as shown in FIG. 3, the static ground load change amount calculation unit 53 outputs the front wheel static ground load change amount ΔW _{f — s} and the rear wheel static ground load change amount ΔW _{r — s} .

The multiplier 54 calculates a multiplication value of the output value from the amplifier 62 and the output value from the subtractor 75. That is, the multiplier 54 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 55 calculates a multiplication value of the output value from the amplifier 63 and the output value from the adder 76. That is, the multiplier 55 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 56 calculates an addition value between the output value from the multiplier 54 and the static ground load change amount ΔW _{f — s} from the static ground load change amount calculation unit 53. 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 57 calculates an addition value between the output value from the multiplier 55 and the static ground load change amount ΔW _{r — s} from the static ground load change amount calculation unit 53. 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 ).

The multiplier 58 calculates a multiplication value of the front wheel static ground load W _f from the adder 56 and the rear wheel static ground load W _r from the adder 57.
Amplifier 59 outputs a signal inputted from the multiplier 58 C _w ² multiplied by. C _w is a load dependent coefficient of cornering power. That is, the amplifier 59 outputs (C _w ² × W _f × W _r ).

The cornering power C _{pf for the} front wheels and the cornering power C _{pr for the} rear wheels are expressed by the following equations (9) and (10), respectively.
C _pf = C _w W _f (9)
C _pr = C _w W _r (10)
Accordingly, the cornering power calculation unit 41 outputs a value C _pf C _pr obtained by multiplying the cornering power C _{pf of the} front wheels and the cornering power C _{pr of the} rear wheels.

  Next, the steering angle dependent parameter calculation unit 42 includes amplifiers 91, 94, 98, 100, an absolute value calculator 92, a calculation factor calculation unit 93, a constant output unit 95, an adder 96, a differentiator 97, a multiplier 99, and a positive / negative sign. An amplifying unit 130 is provided.

The amplifier 91 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 92 calculates and outputs the absolute value of the output value from the amplifier 91.

  The calculation factor calculation unit 93 includes a two-dimensional map M1 in which the value of the calculation factor D is set in advance using the absolute value of the steering angle δ and the vehicle speed V as parameters. Then, the calculation factor calculation unit 93 calculates the calculation factor D based on the absolute value of the steering angle δ input from the absolute value calculator 92 and the vehicle speed V input from the vehicle speed calculation unit 16. The calculation factor D set in the two-dimensional map M1 is set so as to increase as the absolute value of the steering angle δ increases and decrease as the vehicle speed V increases.

Amplifier 94, the operation factor D from the calculation factor calculation unit 93, and outputs the coefficient k ₁ factor that is set in advance.
The constant output unit 95 outputs a preset coefficient c ₁ .

The adder 96 calculates an addition value between the output value from the amplifier 94 and the output value from the constant output unit 95. That is, the adder 96 outputs (k ₁ × D + c ₁ ).
The differentiator 97 calculates a differential value of the absolute value of the steering angle δ input from the absolute value calculator 92.

The amplifier 98 multiplies the output value from the differentiator 97 by a preset coefficient k ₂ and outputs the result.
The multiplier 99 calculates a multiplication value of the output value from the adder 96 and the output value from the amplifier 98.

The positive / negative amplification unit 130 includes amplifiers 131 and 132, a positive / negative determination unit 133, and a changeover switch 134.
The amplifier 131 multiplies the output value from the multiplier 99 by a preset coefficient p ₊ and outputs the result.

The amplifier 132 multiplies the output value from the multiplier 99 by a preset coefficient p ₋ and outputs the result.
The positive / negative determining unit 133 determines whether the output value from the multiplier 99 is positive or negative. The positive / negative determining unit 133 outputs a positive instruction signal when it determines that the output value is positive, and outputs a negative instruction signal when it determines that the output value is negative.

  The changeover switch 134 receives the output values from the amplifiers 131 and 132, selects and outputs one of the output values of the amplifiers 131 and 132 based on the positive instruction signal and the negative instruction signal from the positive / negative determination unit 133. . Specifically, the changeover switch 134 selects and outputs the output value of the amplifier 131 when a positive instruction signal is input from the positive / negative determination unit 133, and when the negative instruction signal is input from the positive / negative determination unit 133. The output value of the amplifier 132 is selected and output.

Therefore, the positive / negative amplifying unit 130 has p ₊ × {(k ₁ × D + c ₁ ) × k ₂ × (d | δ | / dt)} or p ₋ × {(k ₁ × D + c ₁ ) × k ₂ × (d | δ | / dt)} is output. Hereinafter, the coefficient p ₊ and the coefficient p ₋ are collectively expressed as a coefficient p ±.

The amplifier 100 multiplies the output value from the positive / negative amplification unit 130 by the coefficient k expressed by the following equation (11) and outputs the result.
_{^{k = 2L 2 / MC w ···}} (11)
Therefore, the steering angle dependent parameter calculation unit 42 outputs k × p ± × {(k ₁ × D + c ₁ ) × k ₂ × (d | δ | / dt)}.

The multiplier 43 calculates and outputs a multiplication value of the output value from the cornering power calculation unit 41 and the output value from the steering angle dependent parameter calculation unit 42.
As described above, the target load movement amount calculation unit 18 outputs a value represented by the following expression (12) as the target load movement amount Δ.

Δ = (2L ² C _pf C _pr / MC _w )
× p ± × {(k ₁ × D + c ₁ ) + k ₂ × (d | δ | / dt)} (12)
The turning radius ρ of the vehicle is expressed by the following equation (13).

Equation (13) shows that the value of the turning radius with respect to the increase and decrease of the 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 (14) using equations (9) and (10).
_{(L f C pf -L r C} pr) = C w (L f W f -L r W r) ··· (14)
Therefore, (L _f W _f −L _r W _r ) in the equation (14) is expressed by the following equation (15).

As can be understood by comparing Expression (12) and Expression (15), the target load movement amount calculation unit 18 calculates (L _f W _f −L _r W _r ) as the target load movement amount Δ. ing. Specifically, the target load movement amount calculation unit 18 converts the portion of p ± × {(k ₁ × D + c ₁ ) + k ₂ × (d | δ | / dt)} in the equation (12) into the equation (15). The target load movement amount Δ is calculated by assuming that {(1-ρδ / L) / V ² }.

  Next, the vibration damping correction unit 19 includes a sprung vibration model calculation unit 111, subtractors 112, 117, 120, an integrator 113, amplifiers 114, 115, 118, 119, and an adder / subtractor 116, as shown in FIG. .

The sprung vibration model calculation unit 111 calculates the state quantity x = (xν, xν ′, θ _p , θ _p ′) by the above equation (4). The sprung vibration model calculation unit 111 calculates the estimated load movement amount y _s by the following equation (16) using the state quantity x calculated by the above equation (4). 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 112 calculates the target load shift amount Δ from the target load movement amount calculating section 18, a value obtained by subtracting the estimated load shift amount y _s from sprung vibration model calculating unit 111, i.e. the model output error. The integrator 113 integrates the output error from the subtractor 112. The amplifier 114 multiplies the integral value of the output error input from the integrator 113 by a preset integral gain K _i and outputs it. Therefore, the vibration damping correction unit 19 is a type 1 servo system.

The amplifier 115 multiplies the state quantity x output from the sprung vibration model calculation unit 111 by a preset state feedback gain K _s and outputs the result.
The adder / subtractor 116 adds the output value from the amplifier 114 and the running resistance disturbance from the running resistance disturbance estimation unit 14. Further, the adder / subtractor 116 calculates a value obtained by subtracting the output value from the amplifier 115 from the added value, and outputs the subtracted value to the sprung vibration model calculation unit 111 and the subtractor 117. The output value of the adder / subtractor 116 is a correction driving torque that suppresses vibration of the vehicle (hereinafter referred to as a correction driving torque).

The subtractor 117 calculates a value obtained by subtracting the correction drive torque from the adder / subtractor 116 from the torque T _{w — est} from the estimated drive wheel torque calculation unit 12.
The amplifier 118 multiplies the output value from the subtractor 117 by (1 / R _d ) and outputs the reduction ratio (final gear ratio) at the final reduction gear as R _d . The amplifier 119 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 120 calculates a value obtained by subtracting the output value from the amplifier 118 from the output value from the amplifier 119.
The output of the subtractor 120 is the corrected required torque, and the electronic control unit 1 uses the throttle motor (not shown) that changes the opening of the throttle valve based on the corrected required torque, and the fuel in each cylinder. The engine 2 is operated by controlling various actuators such as an ignition plug (not shown) for igniting and an injector (not shown) for injecting fuel into each cylinder.

Next, specific examples of the target load movement amount Δ and the correction driving torque calculated by the electronic control unit 1 will be described with reference to FIG.
A case where the driver performs an operation of turning the steering wheel left and right as when traveling on an S-shaped curve will be described (see graph G1).

  The graph G1 first turns the rudder to the left (see arrow AL1), then turns the rudder back (see arrow AL2), further turns the rudder to the right (see arrow AL3), and then turns the rudder back ( This operation is shown in FIG.

In this case, when the steering angle becomes a value close to 0 by the operation of turning the rudder back, the turning radius suddenly increases (see peaks PK1, PK2, and PK3 in the graph G2).
In the conventional method, the change in the target load movement amount Δ from the understeer to the oversteer is delayed with respect to the turning operation of the rudder (see the curve L1 of the graph G3 and the arrow AL5).

  On the other hand, in the electronic control device 1 of the present embodiment, the target load movement amount Δ changes from oversteer to understeer with good response to the steering turning back operation (see the curve L2 and the arrow AL6 in the graph G3), and After the switching operation, the target load movement amount Δ changes from understeer to oversteer with good response (see the curve L2 and the arrow AL7 in the graph G3). As a result, the correction drive torque increases following the rudder turning-back operation (load shift to the rear wheel side), changes from oversteer to understeer (see curve L3 and arrow AL8 in graph G4), and the rudder Following the cutting operation, the correction drive torque decreases (load moves to the front wheel side), and changes from understeer to oversteer (see curve L3 and arrow AL9 in graph G4). In this way, it is possible to realize both improvement of the vehicle yaw response at the beginning of turning and stabilization of the vehicle posture at the time of acceleration after the end of turning.

On the other hand, in the conventional method, the increase in torque is delayed with respect to the steering turning back operation (see curve L4 and arrow AL10 in graph G4).
Next, specific examples of the target load movement amount Δ1 when p ₊ = p ₋ = 1.0 and the target load movement amount Δ2 when p ₊ = 1.0 and p ₋ = 0.3 are shown. This will be described with reference to FIG.

A case where the driver performs an operation of turning the steering wheel back and forth as when traveling on an S-curve will be described (see steering angle δ in FIG. 6).
First, the target load movement amounts Δ1 and Δ2 change from oversteer to understeer in response to a rudder switchback operation (see arrow AL21). At this time, the minimum value of the target load movement amount Δ1 is about −4 × 10 ⁶ (see the peak PK21). For this reason, the target load movement amount Δ1 not only enables load movement to the rear wheels, but also affects the longitudinal acceleration of the vehicle body. On the other hand, the minimum value of the target load movement amount Δ2 at this time is about −1 × 10 ⁶ (see the peak PK22). Therefore, the target load movement amount Δ2 enables load movement to the rear wheel, but does not affect the longitudinal acceleration of the vehicle body.

Next, in response to the turning operation of the rudder (see arrow AL22), the target load movement amounts Δ1 and Δ2 change from understeer to oversteer. The maximum value of the target load movement amounts Δ1, Δ2 at this time is about + 1.5 × 10 ⁶ (see peak PK23). For this reason, the target load movement amounts Δ1 and Δ2 enable the load movement to the front wheels, but do not affect the longitudinal acceleration of the vehicle body.

By changing the coefficient p _{− in} this way, the target load movement amount on the understeer side can be adjusted without affecting the target load movement amount on the oversteer side.
As shown in FIG. 7, the change amount of the drive wheel torque is ΔT _w , the rear wheel contact load change amount is ΔW _r , the rear wheel radius is r _t , the vehicle wheelbase is L, and the vehicle mass is M. The relationship of the following equation (17) is established, where α is the acceleration of the vehicle.

The first term on the right side in Expression (17) is energy that contributes to acceleration in the longitudinal direction of the vehicle body, and the second term is energy that contributes to load movement of the vehicle body. That is, the driving torque contributes to both the longitudinal acceleration of the vehicle body and the load movement of the vehicle body.

For this reason, the coefficient p ₊ and the coefficient p ₋ of the positive / negative amplifying unit 130 indicate that the increase / decrease amount of the driving torque applied to the vehicle body based on the target load movement amount Δ contributes to the load movement of the vehicle body, while It is set so as not to contribute as much as possible to acceleration.

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. Further, the target load movement amount calculation unit 18 calculates the target load movement amount Δ so that the vehicle travels stably, and the sprung vibration model calculation unit 111 estimates the estimated load movement amount y _s of the vehicle. The vibration damping correcting unit 19, the estimated load shift amount y _s corrects the basic request torque T _w _ _tgt so as to follow the target load shift amount delta.

In this way, the electronic control unit 1 can stabilize the steering characteristics when the vehicle is traveling.
Then, the applicant of the present application, the steering power including both the steering amount and the steering speed index having information capable of discriminating the steering direction (that is, the cutting operation and the switching operation) by the steering operation of the driver, It was found that it has utility as a physical quantity that reflects the driver's steering intention.

For example, the product δ (dδ / dt) of the steering angle δ and the time change rate (dδ / dt) of the steering angle is, as is clear from the following equations (18) and (19), That is, it can be regarded as a physical quantity corresponding to the steering power. Here, G _s is a constant corresponding to the steering system torsional rigidity. ^{_{Further, (G s δ 2/2}} ) is the steering potential energy.

Then, the target load movement amount calculation unit 18 uses, as the steering power, a value obtained by multiplying the calculation factor D, the relationship with the steering angle δ set in advance in the two-dimensional map M1, by the time change rate of the steering angle δ, A target load movement amount Δ is calculated.

  As a result, when calculating the target load movement amount Δ, it is not necessary to use the turning radius ρ as information capable of discriminating the steering direction, and the product calculation of the turning radius ρ and the steering angle δ becomes unnecessary. For this reason, the calculation accuracy of the product of the turning radius ρ = ∞ and the steering angle δ = 0 does not affect the calculation accuracy of the target load movement amount Δ, and the vehicle control steering using the front and rear wheel load movement amounts is prevented. Both responsiveness and stability can be achieved.

The target load movement amount calculation unit 18 determines whether the target load movement amount Δ is positive or negative, and changes the magnitude of the target load movement amount Δ based on the positive / negative determination result.
As a result, it is possible to correct the basic required torque T _{w _tgt} independently of each other when the target load movement amount Δ is larger than 0 and when the target load movement amount Δ is smaller than 0. The responsiveness and stability of vehicle control using the movement amount can be further improved.

Further, the target load movement amount calculation unit 18 uses k × {(k ₁ × D + c ₁ ) × k ₂ × (d | δ | / dt)} as a basic target for calculating the target load movement amount Δ. When the basic target load movement amount is positive when calculated as the load movement amount, a value obtained by multiplying a preset coefficient p ₊ and the basic target load movement amount is set as the target load movement amount Δ. Further, when the basic target load movement amount is negative, the target load movement amount calculation unit 18 sets a value obtained by multiplying the preset coefficient p ₋ and the basic target load movement amount as the target load movement amount Δ.

As a result, the magnitude of the target load movement amount Δ can be changed based on the result of positive / negative judgment by a simple method of multiplying the coefficient p ₊ or the coefficient p ₋ and the basic target load movement amount. The calculation processing load can be reduced.

The coefficient p ₊ and the coefficient p ₋ are adjusted by the coefficient p ₊ and the coefficient p ₋ while the increase / decrease amount of the drive torque applied to the vehicle based on the target load movement amount Δ contributes to the load movement of the vehicle. It is set so as to suppress the contribution to the longitudinal acceleration of the vehicle as compared to the case where it is not broken.

Thereby, it is possible to suppress the occurrence of a situation in which the vehicle is accelerated in the front-rear direction due to vehicle control using the front-rear wheel load movement amount.
Further, the target load movement amount Δ includes a value obtained by multiplying a calculation factor D whose relationship with the steering angle δ is preset in the two-dimensional map M1 by a time change rate of the steering angle δ. For this reason, it is not necessary to calculate the target load movement amount Δ based on a strict theoretical formula as shown in the equation (15), and the calculation processing load for calculating the target load movement amount Δ can be reduced. Accordingly, the target load movement amount Δ can be calculated without using a microcomputer having a high calculation processing capability, for example, using a microcomputer that does not have a floating point calculation function.

  Further, as shown in the equation (15), the front and rear wheel load movement amount varies depending on the vehicle speed V. The calculation factor D set in the two-dimensional map M1 changes according to the vehicle speed V. Therefore, the target load movement amount calculation unit 18 can set an appropriate target load movement amount Δ according to the vehicle speed V.

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 18 is the target load movement amount in the present invention. The calculation means, the sprung vibration model calculation section 111 is the load movement amount estimation means in the present invention, the vibration suppression correction section 19 is the driving force correction means in the present invention, and the basic required torque T _{w _tgt} is the basic required driving force in the present invention. is there.

The coefficient p ₊ is a first adjustment coefficient in the present invention, and the coefficient p ₋ is a second adjustment coefficient 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.

  For example, in the above-described embodiment, the target load movement amount Δ is calculated using the multiplication value of the calculation factor D that changes according to the steering angle δ and the differential value of the steering angle δ. However, as long as the power has the same physical dimension (N · m / s), a physical quantity resulting from steering may be used in addition to the steering angle δ. For example, the target load movement amount Δ may be calculated using a multiplication value of the steering torque (N · m) and the steering angular velocity (rad / s).

  In the above embodiment, the calculation factor calculation unit 93 calculates the calculation factor D by the two-dimensional map M1 using the absolute value of the steering angle δ and the vehicle speed V as parameters. You may comprise as shown in.

  As shown in FIG. 8, the calculation factor calculation unit 93 includes a one-dimensional map M2 in which the value of the calculation factor a is set in advance using the vehicle speed V as a parameter. The calculation factor a set in the one-dimensional map M2 is set so as to decrease as the vehicle speed V increases. Then, the calculation factor calculation unit 93 calculates -exp (-aδ) using the absolute value of the steering angle δ input from the absolute value calculator 92 and the calculation factor a set using the one-dimensional map M2. Then, this value is output as the calculation factor D.

  Thereby, it is not necessary to calculate the target load movement amount Δ based on a strict theoretical formula as shown in the equation (15), and the calculation processing load for calculating the target load movement amount Δ can be reduced.

In the above embodiment, the coefficient p ₊ and the coefficient p ₋ are fixed values. For example, as shown in FIG. 9, the coefficient p ₊ and the coefficient p ₋ are changed according to the vehicle speed V. Also good.
In FIG. 9, the coefficient p ₊ and the coefficient p ₋ are set to decrease as the vehicle speed V increases. This is because the calculation factor D used for calculating the target load movement amount Δ is set so as to decrease as the vehicle speed V increases.

  DESCRIPTION OF SYMBOLS 1 ... Electronic controller, 11 ... Basic required torque calculation part, 18 ... Target load movement amount calculation part, 19 ... Damping correction part, 111 ... Sprung vibration model calculation part

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. Movement amount calculation means (18);
Load movement amount estimation means (111) for estimating the front and rear wheel load movement amount of the vehicle;
Driving force that 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 calculated by the target load moving amount calculating means. Correction means (19),
The target load movement amount calculating means includes
The target load movement amount is calculated based on a physical quantity corresponding to the steering power of the driver, and whether the target load movement amount is positive or negative is determined, and the target load is determined based on the positive / negative determination result. A vehicle control device characterized in that the amount of movement is changed. The target load movement amount calculating means includes
Based on a physical quantity corresponding to the steering power of the driver, a basic target load movement amount serving as a basis for calculating the target load movement amount is calculated. Further, when the basic target load movement amount is positive, A value obtained by multiplying a preset first adjustment coefficient and the basic target load movement amount is set as the target load movement amount, and when the basic target load movement amount is negative, a preset second adjustment is performed. The vehicle control device according to claim 1, wherein a value obtained by multiplying a coefficient and the basic target load movement amount is set as the target load movement amount. The first adjustment coefficient and the second adjustment coefficient are such that the increase / decrease amount of the driving force applied to the vehicle based on the target load movement amount contributes to the load movement of the vehicle, 3. The vehicle control device according to claim 2, wherein the vehicle control device is set so as to suppress contribution to acceleration in the front-rear direction of the vehicle as compared with a case where adjustment by the second adjustment coefficient is not performed. The vehicle control device according to claim 2 or 3, wherein the first adjustment coefficient and the second adjustment coefficient change according to a traveling speed of the vehicle.