JP6449697B2 - Vehicle control apparatus and vehicle control method - Google Patents

Description

  The present invention relates to a vehicle control device and a vehicle control method.

  Conventionally, various techniques for controlling turning of a vehicle are known. For example, in Patent Document 1 below, when the road surface μ is higher than a reference value, the braking force control of each wheel is performed according to the calculated longitudinal force based on the ground contact load of the wheel, and the road surface μ is lower than the reference value. It is described that the braking force control of each wheel is performed according to the calculated target slip ratio based on the wheel slip angle.

JP 2007-237888 A

  However, in the technique described in Patent Document 1, either one of the braking force control of each wheel based on the ground load and the driving force control based on the slip angle of the wheel is selectively performed. It is difficult to perform optimal control according to the conditions.

  Therefore, the present invention has been made in view of the above problems, and the object of the present invention is to optimally perform control according to the skid angle and control according to the ground load based on the road surface condition. It is an object of the present invention to provide a new and improved vehicle control apparatus and vehicle control method capable of greatly improving the turning performance.

  In order to solve the above-described problem, according to an aspect of the present invention, a skid angle obtaining unit that obtains a skid angle of each wheel and a first torque correction coefficient for each wheel based on the skid angle are calculated. A torque correction coefficient calculating section; a ground load estimating section for estimating a ground load of each wheel; a second torque correction coefficient calculating section for calculating a second torque correction coefficient for each wheel based on the ground load; Based on the parameter representing the coefficient, when the friction coefficient is small, the distribution of the first torque correction coefficient is increased, and when the friction coefficient is large, the distribution of the second torque correction coefficient is increased. A vehicle control apparatus comprising: a torque correction coefficient calculation unit that calculates a torque correction coefficient for each wheel from two torque correction coefficients; and a correction processing unit that corrects a drive target torque for each wheel based on the torque correction coefficient. It is provided.

  The side slip angle acquisition unit is configured to calculate the first side slip angle when the difference is small based on the difference between the first side slip angle of each wheel obtained from the vehicle model and the second side slip angle of each wheel obtained from the sensor. If the distribution is increased and the difference is large, the distribution of the second side slip angle may be increased, and the side slip angle of each wheel may be calculated from the first and second side slip angles.

  The side slip angle acquisition unit acquires the first side slip angle of each wheel from the side slip angle of the vehicle center of gravity position obtained from the vehicle model, and the second side of each wheel from the side slip angle of the vehicle center of gravity position obtained from the sensor. The side slip angle may be acquired.

  The first torque correction coefficient calculation unit increases the first torque correction coefficient as the side slip angle increases with respect to the wheel on the outer side of the turn, and increases the first torque as the side slip angle increases with respect to the wheel on the inner side of the turn. The correction coefficient may be reduced.

  The second torque correction coefficient calculation unit may increase the second torque correction coefficient as the ground load increases.

  The ground load estimating unit may estimate the ground load of each wheel based on at least the vehicle weight, the longitudinal acceleration, and the lateral acceleration.

  Further, the torque correction coefficient calculation unit may calculate the torque correction coefficient using a difference between a side slip angle of a vehicle center of gravity position obtained from a vehicle model and a side slip angle of a vehicle center of gravity position obtained from a sensor as the parameter. good.

  Further, a drive target torque calculation unit that calculates the drive target torque of each wheel based on the vehicle speed, the steering angle, and the yaw rate may be provided.

  In order to solve the above problem, according to another aspect of the present invention, a step of obtaining a side slip angle of each wheel, a step of calculating a first torque correction coefficient of each wheel based on the side slip angle, and , A step of estimating a contact load of each wheel, a step of calculating a second torque correction coefficient of each wheel based on the contact load, and a parameter indicating a friction coefficient of a road surface when the friction coefficient is small Calculating the torque correction coefficient of each wheel from the first and second torque correction coefficients by increasing the distribution of the first torque correction coefficient and increasing the distribution of the second torque correction coefficient if the friction coefficient is large. And a step of correcting the drive target torque of each wheel based on the torque correction coefficient.

  As described above, according to the present invention, the turning performance can be greatly improved by optimally performing the control according to the skid angle and the control according to the ground contact load based on the road surface condition.

It is a mimetic diagram showing a vehicle concerning one embodiment of the present invention. It is a schematic diagram which shows turning control by steering. It is a schematic diagram which shows the structure of a control apparatus. It is a schematic diagram which shows the map which calculates weighting gain (tau) f. It is a schematic diagram which shows the map which calculates weighting gain (tau) r. It is a schematic diagram which shows the map which calculates 1st correction coefficient Coef1 of the wheel inside turning. It is a schematic diagram which shows the map which calculates 1st correction coefficient Coef1 of the wheel inside turning. It is a schematic diagram which shows the map which calculates 1st correction coefficient Coef1 of the wheel of the turning outer side. It is a schematic diagram which shows the map which calculates 1st correction coefficient Coef1 of the wheel of the turning outer side. It is a schematic diagram showing a map defining the relationship between the ground load FzFr (In), FzFr (Out), FzRr (In), FzRr (Out) and the second torque correction coefficient Coef2. It is a schematic diagram which shows the map which calculates weighting gain (tau). It is a flowchart which shows the process of this embodiment. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented. It is a schematic diagram for demonstrating the effect acquired when the control of this embodiment is implemented.

  Exemplary embodiments of the present invention will be described below in detail with reference to the accompanying drawings. In addition, in this specification and drawing, about the component which has the substantially same function structure, duplication description is abbreviate | omitted by attaching | subjecting the same code | symbol.

  First, with reference to FIG. 1, the structure of the vehicle 1000 which concerns on one Embodiment of this invention is demonstrated. FIG. 1 is a schematic diagram showing a vehicle 1000 according to the present embodiment. As shown in FIG. 1, a vehicle 1000 includes driving force generators (motors) 108, 110, 112, driving front wheels 100 and 102, rear wheels 104 and 106, front wheels 100 and 102, and rear wheels 104 and 106. 114, wheel speed sensors 116, 118, 120, 122 for detecting the respective wheel speeds of the front wheels 100, 102 and the rear wheels 104, 106, a steering wheel 124, a steering angle sensor 130, a power steering mechanism 140, a yaw rate sensor 150, a lateral An acceleration sensor 160 and a control device (controller) 200 are included.

  The vehicle 1000 according to the present embodiment is provided with motors 108, 110, 112, and 114 for driving the front wheels 100 and 102 and the rear wheels 104 and 106, respectively. Therefore, the driving torque can be controlled by each of the front wheels 100 and 102 and the rear wheels 104 and 106. Accordingly, the yaw rate can be generated by the torque vectoring control by driving each of the front wheels 100 and 102 and the rear wheels 104 and 106 independently of the yaw rate generation by the steering of the front wheels 100 and 102. In particular, in this embodiment, the yaw rate is generated independently of the steering system by controlling the torque of the rear wheels 104 and 106 individually. The driving torque of the rear wheels 104 and 106 is controlled by controlling the motors 112 and 114 corresponding to the rear wheels 104 and 106 based on a command from the control device 200.

  The power steering mechanism 140 controls the steering angle of the front wheels 100 and 102 by torque control or angle control according to the operation of the steering wheel 124 by the driver. The steering angle sensor 130 detects the steering angle θH input by the driver operating the steering wheel 124. The yaw rate sensor 150 detects the actual yaw rate γ of the vehicle 1000. Wheel speed sensors 116, 118, 120 and 122 detect vehicle speed V of vehicle 1000.

  Note that the present embodiment is not limited to this embodiment, and the motors 108 and 102 for driving the front wheels 100 and 102 are not provided, and only the rear wheels 104 and 106 are independently driven by the motors 112 and 114. The generated vehicle may be used. Further, the present embodiment is not limited to torque vectoring based on driving force control, and can also be realized in a 4WS system for controlling the steering angle of the rear wheels.

  FIG. 2 is a schematic diagram showing turning control (steering control) by steering performed by the vehicle 1000 according to the present embodiment. In the turning control by the steering, the turning of the vehicle 1000 is supported by generating a driving force difference between the rear wheels 104 and 106 according to the operation of the steering wheel 130 by the driver. In the example shown in FIG. 2, the vehicle 1000 is turning to the left by the steering of the driver (driver). Further, due to the difference in driving force between the rear wheels 104 and 106, a forward driving force is generated on the right rear wheel 106, and the driving force on the left rear wheel 104 is suppressed with respect to the right rear wheel 106 or backwards. By generating a driving force, a driving force difference is generated on the left and right, and a moment is generated in a direction that supports a counterclockwise turn.

  In the present embodiment, with respect to vehicle driving force control, turning support control is performed by using the left / right difference of the drive target torque MotTrqTgt obtained by referring to the vehicle speed, the steering amount, the sensor detection value of the vehicle behavior, and the like. At this time, a first torque correction coefficient Coef1 based on the side slip angle corresponding to the tire front-rear position and a second torque correction coefficient Coef2 based on the ground contact load estimated from the longitudinal acceleration and the lateral acceleration acting on the vehicle are calculated. Then, the torque correction coefficient TrqAdjustCoef is calculated by allocating the first torque correction coefficient and the second torque correction coefficient according to an index based on the side slip angle (difference between the vehicle model value and the calculated value) of the vehicle center of gravity position, and driving target A product obtained by multiplying the torque and the torque correction coefficient is calculated as a drive request torque MotTrqReq. And by controlling the driving force of each wheel individually according to the driving environment such as the change of road surface μ and the gradient by the required driving torque MotTrqReq, it is possible to reduce the driver's discomfort with respect to the vehicle behavior and ensure the vehicle response performance. Let As a result, the driving force of each wheel can be individually controlled in accordance with the driving environment such as the change in road surface μ and the gradient, and both reduction of the driver's uncomfortable feeling associated with driving and ensuring of vehicle response performance can be achieved. Details will be described below.

  FIG. 3 is a schematic diagram illustrating the configuration of the control device 200. The control device 200 includes an in-vehicle sensor 210, a drive target torque calculation unit 220, a skid angle calculation unit 230, a ground load estimation unit 240, a torque correction gain calculation unit 250, and a drive target torque correction processing unit 260.

  The skid angle calculation unit 230 includes a vehicle model 232, a skid angle calculation unit 234 that calculates a skid angle based on a sensor detection value, and a skid angle reference value calculation unit (β reference value (β_ref) calculation that calculates a reference value of the skid angle. Part) 236.

  The torque correction gain calculation unit 250 includes a first torque correction coefficient calculation unit 252, a second torque correction coefficient calculation unit 254, and a torque correction gain calculation unit 256.

  The drive target torque calculation unit 220 uses a drive target torque (MotTrqTgtFL) to be applied to each wheel 100, 102, 104, 106 from parameters indicating vehicle motion detected by the in-vehicle sensor 210, such as the vehicle speed V, the steering angle θH, and the yaw rate γ. , MotTrqTgtFR, MotTrqTgtRL, and MotTrqTgtRR). It should be noted that the control target torque may be any calculation means, such as a state quantity calculated by steering control by steering, vehicle control by an external recognition unit to be described later, or an instruction value given by a map.

  The vehicle model 232 refers to a vehicle model (the following equations (1) and (2)) simulating the behavior of the vehicle based on the vehicle speed V and the steering angle θH, and the equations (1) and (2) The side slip angle model value β_mdl of the vehicle center of gravity position is calculated with reference to Further, the yaw rate model value γ_mdl is calculated as γ in the equations (1) and (2). The side slip angle model value β_mdl calculated here is a side slip angle at the center of gravity of the vehicle. The skid angle model value β_mdl of the vehicle center of gravity position is input to the skid angle reference value calculation unit 236.

  In equations (1) to (2), I is the yaw inertia of the vehicle, β is the side slip angle of the vehicle, and corresponds to the side slip angle model value β_mdl. Γ corresponds to the yaw rate model value γ_mdl. Further, δ is a tire steering angle, and can be calculated by dividing the steering angle θH by the steering gear ratio. Regarding the cornering force, not only the parameters Kf and Kr corresponding to the actual vehicle, but also the cornering forces KfTgt and KrTgt that change according to the control target value of the yaw rate may be used. When the cornering forces KfTgt and KrTgt are used, the skid angle model value β_mdl and the yaw rate model value γ_mdl can be calculated in consideration of characteristics when vehicle control intervenes.

  Further, the vehicle model 232 refers to the vehicle speed V, the steering angle θH, and the vehicle specifications, and calculates the skid angle model values βf_mdl and βr_mdl corresponding to the front and rear tire positions based on the following formulas (3) and (4). calculate. Here, βf_mdl is a side slip angle model value corresponding to the front wheel position, and βr_mdl is a side slip angle model value corresponding to the rear wheel position.

  Further, the side slip angle calculation unit 234 calculates a side slip angle β_clc corresponding to the position of the center of gravity of the vehicle from the following equation (5) based on the lateral acceleration Gy acquired from the acceleration sensor 160, the actual yaw rate γ, and the vehicle speed V. . The skid angle β_clc corresponding to the position of the center of gravity of the vehicle is input to the skid angle reference value calculation unit 236.

  Further, the β calculation unit 234 refers to the skid angle β_clc, the actual yaw rate γ, the vehicle speed V, and the steering angle θH, and based on the following equations (6) and (7), the skid angle βf_clc corresponding to the front and rear tire positions is obtained. , Βr_clc is calculated. As described above, the side slip angle β_clc is calculated based on the lateral acceleration Gy acquired from the acceleration sensor 160. Accordingly, βf_clc is a side slip angle corresponding to the front wheel position in a state quantity calculated from the sensor value, and βr_clc is a side slip angle corresponding to the rear wheel position in a state quantity calculated from the sensor value.

  Further, the side slip angle reference value calculation unit 236 calculates the front wheel side slip angle deviation βf_diff from the difference between the side slip angle model value βf_mdl and the side slip angle βf_clc calculated from the sensor value based on the following equation (8), and calculates the front wheel position This is an index for discriminating the degree of deviation between the vehicle model and the actual vehicle behavior at the side slip angle.

  Further, the side slip angle reference value calculation unit 236 calculates the weighting gain τf using a map having the front wheel side slip angle deviation βf_diff as an input. FIG. 4 is a schematic diagram showing a map for calculating the weighting gain τf. In FIG. 4, a region A1 is a region where the vehicle model and the actual vehicle behavior match (high μ, normal region), and is a region where βf_ref = βf_mdl. The region A3 is a region where the vehicle model and the actual vehicle behavior do not match (low μ, limit region), and is a region where βf_ref = βf_clc. The region A2 is a region where the normal region transitions to the limit region, and is a region where the distribution of βf_mdl and βf_clc is calculated according to the slope of the map.

  As shown in FIG. 4, when | βf_diff | is within a predetermined threshold value TH1, it is determined that the reliability of the vehicle model 232 is high and τf = 1, and | βf_diff | is greater than the predetermined threshold value TH2. Is larger, it is determined that the reliability of the vehicle model 232 is low, and τf = 0 is set. On the other hand, in the range of TH1 <| βf_diff | <TH2, τf is linearly interpolated between 0 and 1 in accordance with the degree of deviation (model reliability) between the sideslip angle model value βf_mdl and the sideslip angle βf_clc calculated from the sensor value. To do.

  Also, the side slip angle reference value calculation unit 236 allocates the side slip angle model value βf_mdl and the side slip angle βf_clc calculated from the sensor value based on the following equation (9) using the weighting gain τf, and refers to the front wheel side slip angle. Calculated as the value βf_ref.

  Further, the side slip angle reference value calculation unit 236 calculates the rear wheel side slip angle deviation βr_diff from the difference between the side slip angle model value βr_mdl and the side slip angle βr_clc calculated from the sensor value based on the following equation (10). This is an index for determining the degree of deviation between the vehicle model and the actual vehicle behavior at the side slip angle of the position.

  Further, the side slip angle reference value calculation unit 236 calculates the weighting gain τr using a map that receives the rear wheel side slip angle deviation βr_diff. FIG. 5 is a schematic diagram showing a map for calculating the weighting gain τr. In FIG. 5, a region A4 is a region where the vehicle model and the actual vehicle behavior match (high μ, normal region), and is a region where βr_ref = βr_mdl. Region A6 is a region where the vehicle model and the actual vehicle behavior do not match (low μ, limit region), and is a region where βr_ref = βr_clc. The region A5 is a region that transitions from the normal region to the limit region, and is a region that calculates the distribution of βr_mdl and βr_clc according to the slope of the map. As shown in FIG. 5, when | βr_diff | is within the predetermined threshold value TH1, it is determined that the reliability of the vehicle model 232 is high and τr = 1, and | βr_diff | is greater than the predetermined threshold value TH2. Is larger, it is determined that the reliability of the vehicle model 232 is low, and τr = 0 is set. On the other hand, in the range of TH1 <| βf_diff | <TH2, the weighting gain τr between 0 and 1 is set in accordance with the degree of deviation (model reliability) between the sideslip angle model value βr_mdl and the sideslip angle βr_clc calculated from the sensor value. Perform linear interpolation.

  Also, the side slip angle reference value calculation unit 236 distributes the side slip angle model value βr_mdl and the side slip angle βr_clc calculated from the sensor value based on the following equation (11) using the weighting gain τr, and calculates the rear wheel side slip angle. Calculated as a reference value βr_ref.

  The reference value βf_ref of the front wheel side slip angle and the reference value βr_ref of the rear wheel side slip angle are input to the first torque correction coefficient calculation unit 252. The first torque correction coefficient calculation unit 252 calculates the first torque correction coefficients (Coef1FL, Coef1FR) of the two front wheels (FL wheel, FR wheel) from the reference value βf_ref from the map. Further, the first torque correction coefficient calculation unit 252 calculates the first torque correction coefficients (Coef1RL, Coef1RR) of the two rear wheels (RL wheel, RR wheel) from the map from the reference value βr_ref.

  When calculating the first torque correction coefficient Coef1, the first torque correction coefficient calculation unit 252 calculates the first torque correction coefficient Coef1 using different maps depending on whether the vehicle is turning outside or turning inside. 6A and 6B are schematic diagrams illustrating maps for calculating the first torque correction coefficient Coef1 of the wheel on the inner side of the turn. FIG. 6A is a map for calculating the first torque correction coefficient Coef1_Fr (In) of the wheel on the inside of the turn from the reference value βf_ref of the front wheel side slip angle, and FIG. 6B shows the wheel on the inside of the turn from the reference value βr_ref of the rear wheel side slip angle. This is a map for calculating the first torque correction coefficient Coef1_Rr (In). FIGS. 7A and 7B are schematic diagrams showing maps for calculating the first torque correction coefficient Coef1 of the wheels on the outer side of the turn. FIG. 7A is a map for calculating the first torque correction coefficient Coef1_Fr (Out) of the outer wheel from the reference value βf_ref of the front wheel side slip angle, and FIG. 7B shows the outer wheel of the turn from the reference value βr_ref of the rear wheel side slip angle. This is a map for calculating the first torque correction coefficient Coef1_Rr (Out).

  According to the maps shown in FIGS. 6A and 6B, the control is performed so that the motor torque of the wheels on the inner side of the turn decreases as the reference values βf_ref and βr_ref of the sideslip angle increase. Further, according to the maps shown in FIGS. 7A and 7B, control is performed so that the motor torque of the wheels on the outer side of the turn increases as the reference values βf_ref and βr_ref of the sideslip angle increase. According to the maps shown in FIG. 6 and FIG. 7, the control amount on the inside of the turn having a smaller ground load than that on the outside of the turn is suppressed, and the control amount on the outside of the turn having a large ground load is increased while preventing the idling of the wheels. By doing so, turning of the vehicle can be supported while avoiding a state where the limit of the tire friction circle is exceeded (wheels are idling). More specifically, because the wheel outside the turn can use more friction circles than the wheel inside the turn, increasing the control amount outside the turn more than the control amount inside the turn makes it possible to Without changing the torque, idling can be avoided and turning of the vehicle can be supported. Further, by using different maps on the outside and inside of the turn, the left and right first torque correction coefficients Coef1 can be calculated based on the sideslip angle without performing a complicated calculation.

  The first torque correction coefficient calculation unit 252 determines which of the left and right wheels is outside the turn or inside the turn based on the detected value of the lateral acceleration Gy, the steering angle θH, and the like, and turns based on this The map of FIG. 6 is applied to the inner wheels, and the map of FIG. 7 is applied to the outer wheels. Therefore, the first torque correction coefficient calculation unit 252 can calculate the first torque correction coefficient Coef1 according to the reference values βf_ref and βr_ref of the sideslip angle according to the inside of the turn or the outside of the turn.

  The yaw rate reference value may be calculated by weighting the yaw rate model value and the yaw rate sensor value based on the degree of divergence between the yaw rate model value and the yaw rate sensor value by the same method as the calculation of the reference value of the skid angle. . In this case, the first torque correction coefficient Coef1 and the second torque correction coefficient Coef2 may be calculated by applying the reference value of the yaw rate to the maps of FIGS.

  The ground load estimation unit 240 uses the longitudinal acceleration (Gx), lateral acceleration (Gy), and specifications acquired from the sensors mounted on the vehicle to calculate each wheel from the following equations (12) to (15). Estimate the contact load that occurs in

  In Expressions (12) to (15), FzFr (In) is a grounding load inside the turning of the front wheel, and FzFr (Out) is a grounding load outside the turning of the front wheel. Further, FzRr (In) is a grounding load on the inner side of the rear wheel and FzRr (Out) is a grounding load on the outer side of the rear wheel.

  Kf_φ is a calculation parameter (coefficient for multiplying front and lateral acceleration), Kr_φ is a calculation parameter (coefficient for multiplying rear and lateral acceleration), and Kh_φ is a parameter for calculation (coefficient for multiplying longitudinal acceleration). .

  The above are parameters calculated according to the turning situation, and FzFr (In) and FrFr (Out) are set as the ground load FzFL of the left front wheel or the ground load FzFR of the right front wheel, and FzRr (In) and FzRr. (Out) is set as the ground load FzRL of the left rear wheel or the ground load FzRR of the right rear wheel.

  The operation parameters Kf_φ, Kr_φ, and Kh_φ are factors related to ground load estimation and vehicle specifications, and are calculated from the following equations (16) to (18).

In the equations (12) to (18), the constants are as follows.
RsdFr: front roll stiffness distribution RsdRr: rear roll stiffness distribution hs: roll arm length (length from roll center to center of gravity)
hf: Front roll center height hr: Rear roll center height hg: Body center of gravity height tf: Front tread length tr: Rear tread length W: Vehicle weight Wf: Vehicle front axle weight Wr: Vehicle rear axle weight

  The contact loads FzFL, FzFR, FzRL, and FzRR of each wheel estimated by the contact load estimation unit 240 are input to the second torque correction coefficient calculation unit 254. FIG. 8 is a schematic diagram showing a map that defines the relationship between the ground loads FzFL, FzFR, FzRL, FzRR and the second torque correction coefficient Coef2.

  The second torque correction coefficient calculation unit 254 calculates a second torque correction coefficient (Coef2) for correcting the torque of each wheel from the map of FIG. 8 based on the ground loads FzFL, FzFR, FzRL, FzRR of each wheel. . Here, Coef2FL is calculated from FzFL, Coef2FR is calculated from FzFR, Coef2RL is calculated from FzRL, and Coef2RR is calculated from FzRR. As shown in FIG. 8, the value of the second torque correction coefficient Coef2 increases as the ground load increases. Therefore, according to the second torque correction coefficient Coef2, the drive target torque can be set according to the ground load.

  The torque correction gain calculation unit 256 calculates a deviation β_diff between the sideslip angle β_mdl calculated by the vehicle model and the sideslip angle β_clc calculated from the sensor detection value based on the following equation (19). The deviation β_diff is an index for determining the degree of divergence between the vehicle model and the actual vehicle behavior at the side slip angle of the vehicle gravity center position.

  The torque correction gain calculation unit 256 refers to the deviation β_diff and calculates the weighting gain τ from the map of FIG. Specifically, when β_diff is within the interval specified by a predetermined threshold (TH1_M to TH1_P), the distribution of Coef2 is set to 100%. Further, when the deviation β_diff is positive and not less than a predetermined threshold (β_diff> TH2_P), or the deviation β_diff is negative and not more than the predetermined threshold (β_diff ≦ TH2_M), the distribution of Coef1 is performed. 100%. When TH2_M <β_diff <TH1_M, or TH1_P <β_diff ≦ TH2_P, Coef1 and Coef2 are distributed according to the degree of divergence between β_mdl and β_clc to calculate the weighting gain τ. The threshold used in the weighting process of FIG. 9 can be set to any value as long as the vehicle control is within the range. In FIG. 9, a region A7 is a region where the vehicle model and the actual vehicle behavior match (high μ, normal region), and a region A9 is a region where the vehicle model and the actual vehicle behavior do not match (low μ, limit region). The region A8 is a region where a transition is made from the normal region to the limit region.

  Then, the torque correction gain calculation unit 256 calculates the first torque correction coefficient (Coef1) calculated from the side slip angle of the front and rear wheels and the ground contact load of the front and rear wheels based on the following equation (20) using the weighting gain τ. The second torque correction coefficient (Coef2) is allocated, and the torque correction coefficient (TrqAdjustCoef) used in the control is calculated. The weighting process according to Expression (20) is performed for each of the FL wheel, the FR wheel, the RL wheel, and the RR wheel. In the calculation of the torque correction coefficient TrqAdjustCoef, state quantities other than the longitudinal acceleration, lateral acceleration, and sideslip angle (for example, vertical acceleration acting on the vehicle body and tire) may be used.

  The first torque correction coefficient Coef1 calculated from the side slip angle of the front and rear wheels and the second torque correction coefficient Coef2 calculated from the ground contact load of each wheel are distributed by the weighting gain τ and designated as the torque correction coefficient used in the control. As described above, the first torque correction coefficient Coef1 is a value calculated according to the reference values βf_ref and βr_ref of the sideslip angle depending on the inside or outside of the turn. On the other hand, the second torque correction coefficient Coef2 is a value calculated according to the contact load. At low μ, slipping of each wheel is likely to occur, and the transmission of driving force to the road surface also decreases, so it is not always effective to perform motor torque correction control with the ground load estimated using the vehicle model. . Therefore, at low μ, by performing control based on the first torque correction coefficient Coef1 calculated based on the side slip angle of the tire front and rear wheels, the turning characteristics of the vehicle according to the road surface condition compared to the case of controlling only with the ground load. Can be secured. On the other hand, when the vehicle is at high μ, the vehicle model is highly reliable, and the contact load estimating unit 240 can estimate a highly reliable contact load. Therefore, at high μ, control based on the second torque correction coefficient Coef2 is performed. By doing so, natural turning according to the contact load of each wheel becomes possible.

  The drive request torque correction processing unit 260 calculates the drive target torque (MotTrqTgtFL, MotTrqTgtFR, MotTrqTgtRL, MotTrqTgtRR) calculated by the drive target torque calculation unit 220 and the torque correction coefficient TrqAdjustCoef obtained by the weighting process according to the following equation (21). To calculate the motor required torque MotTrqReq. This torque correction process is performed for each of the FL wheel, the FR wheel, the RL wheel, and the RR wheel.

  The motors 108, 110, 112, and 114 of each wheel are controlled based on the motor required torques MotTrqReqFL, MotTrqReqFR, MotTrqReqRL, and MotTrqReqRR of each wheel.

  FIG. 10 is a flowchart showing the processing of this embodiment. First, in step S10, the drive target torque calculator 220 calculates the drive target torque. In the next step S12, a skid angle (reference value) is calculated. In the next step S14, the contact load of each wheel is estimated. In the next step S16, a torque correction gain (first torque correction coefficient, second torque correction coefficient) is calculated. In the next step S18, a drive target torque correction process is performed.

  Next, the effect obtained when the control of this embodiment is implemented will be described based on FIGS. First, turning at high μ will be described. FIG. 11 is a characteristic diagram showing changes in the vehicle speed V as input data. FIG. 12 is a characteristic diagram showing a change in steering angle (steering wheel angle) as input data.

  FIGS. 13 and 14 show the characteristics of the motor torque of each wheel and the left and right torque difference in comparison with the control of the present embodiment and the conventional control when FIGS. 11 and 12 are input given by the driver. FIG. Here, the conventional control indicates a control for adding a yaw moment to the vehicle by adjusting the left and right torque distribution according to the turning state of the vehicle while equally distributing the driving torque in the front-rear direction.

  FIG. 13 is a characteristic diagram showing the motor torque of the left front wheel (FL wheel), the motor torque of the right front wheel (FR wheel), and the difference between the two. Here, the characteristic of the thick line indicates the control of the present embodiment, and the characteristic of the thin line indicates the conventional control.

  In the conventional control shown in FIG. 13, the torque of the left front wheel rises as the steering angle increases in a state where torque control according to the ground load is not performed.

  On the other hand, in the control of the present embodiment, at the time of high μ, the ratio of the second torque correction coefficient Coef2 is high, and control according to the ground load is performed. Accordingly, as shown in FIG. 13, it can be seen that the motor torque of the right front wheel (FR wheel) is increased with respect to the left front wheel (FL wheel). That is, the left and right torque distribution is switched by the side slip angle corresponding to the tire position and the ground contact load, and the motor torque of the left front wheel (FL wheel) is reduced compared to the conventional control, and the motor torque of the right front wheel (FR wheel) is reduced. Increases, and a motor torque difference proportional to the additional yaw moment is generated between the left front wheel (FL wheel) and the right front wheel (FR wheel). As a result, as shown in the characteristic of the motor torque difference between the left front wheel (FL wheel) and the right front wheel (FR wheel) in FIG. 13, in this embodiment, the transition of the torque output difference at the initial stage of steering is more gradual than in the conventional control. On the other hand, even when the steering angle and the vehicle speed are in a steady state, the amount of overshoot related to the torque output difference is reduced, so that the transition of the yaw moment applied to the vehicle along with the turning becomes gentle. Accordingly, it is possible to perform a turn with a natural feeling in the initial stage of steering, and to stabilize the behavior during steady steering.

  Similarly for the rear wheel, as shown in FIG. 14, the motor torque of the right rear wheel (RR wheel) is increased with respect to the left rear wheel (RL wheel), and motor torque control according to the ground load is performed. Yes. That is, the left and right torque distribution is switched by the side slip angle corresponding to the tire position and the ground contact load, and the motor torque of the left rear wheel (RL wheel) is reduced with respect to the conventional control, and the right rear wheel (RR wheel) is reduced. As the motor torque increases, a motor torque difference proportional to the additional yaw moment is generated between the left rear wheel (RL wheel) and the right rear wheel (RR wheel). As a result, as shown in the characteristic of the motor torque difference between the left rear wheel (RL wheel) and the right rear wheel (RR wheel), in this embodiment, the torque output difference at the initial stage of steering is suppressed compared to the conventional control. . Therefore, the feeling of being moved at the rear of the vehicle that occurs at the beginning of steering can be reduced, and the amount of overshoot when the steering angle and vehicle speed reach a steady state can be reduced, so that the vehicle can turn with a natural feeling. The vehicle behavior during turning can be stabilized.

  FIG. 15 is a characteristic diagram showing the left-right torque difference between the front and rear, the motor torque difference characteristic (FR-FL) between the left front wheel (FL wheel) and the right front wheel (FR wheel), and the left rear wheel (RL wheel). The graph shows the characteristic (RR-RL) of the motor torque difference between the right rear wheel and the right rear wheel (RR wheel). As shown in FIG. 15, in the conventional control, the front and rear torques are controlled at the same timing. On the other hand, in the control of this embodiment, a phase difference corresponding to the ground load and the side slip angle is attached to the front and rear torques. Therefore, by distributing the driving force according to the state of each wheel, the phenomenon that the rear wheel turns around the front wheel can be suppressed, and the overshoot of the motor torque (left-right difference) can be reduced compared to the conventional control. .

  FIG. 16 is a characteristic diagram showing changes in the yaw rate in the examples of FIGS. FIG. 17 is a characteristic diagram showing changes in lateral acceleration in the examples of FIGS. As shown in FIG. 16, in the conventional control, the yaw rate changes almost linearly with respect to changes in the vehicle speed and the steering angle, but at the initial stage of steering, driving force distribution control proportional to steering is performed for both the front and rear wheels. As a result, a behavior such that the rear part of the vehicle turns around the turning center occurs, which leads to a driver's uncomfortable feeling. On the other hand, in the control of this embodiment, since the motor torque is corrected by the ground load calculated independently for each wheel, compared to the conventional control, the drive is performed according to the transition of the front and rear load distribution and the left and right load distribution. Force distribution is performed and the transition of the yaw rate in the initial stage of steering becomes gradual. For this reason, the behavior that occurs in proportion to the initial stage of steering as seen in the prior art is alleviated, and the feeling that the rear part of the vehicle is moved is reduced, so that the driver's uncomfortable feeling is also reduced.

  Similarly, as shown in FIG. 17, in the conventional control, the lateral acceleration changes almost linearly with respect to the transition of the vehicle speed and the steering angle. For this reason, at the initial stage of steering, the driving force distribution control proportional to the steering is performed on both the front wheels and the rear wheels, so that the behavior of the rear part of the vehicle turning around the turning center occurs, and the driver feels uncomfortable. It leads to. On the other hand, in the control of this embodiment, since the motor torque is corrected by the ground load calculated independently for each wheel, compared to the conventional control, the drive is performed according to the transition of the front and rear load distribution and the left and right load distribution. Force distribution is performed and the transition of the yaw rate in the initial stage of steering becomes gradual. For this reason, since the feeling that the vehicle rear portion is moved in proportion to the initial stage of steering as seen in the prior art is reduced, the driver's uncomfortable feeling is also reduced. Therefore, according to the control of the present embodiment, natural turning according to the ground contact load of each wheel is possible in turning at high μ.

  Next, turning at low μ will be described. Changes in vehicle speed and steering angle (steering wheel angle) as input data are the same as in FIGS. 11 and 12.

  FIGS. 18 and 19 are characteristic diagrams showing the motor torque of each wheel and the difference between the left and right torques in comparison with the control of this embodiment and the conventional control when FIGS. 11 and 12 are used as input data. is there. Here, the conventional control indicates a control for adding a yaw moment to the vehicle by adjusting the left and right torque distribution according to the turning state of the vehicle while equally distributing the driving torque in the front-rear direction.

  FIG. 18 is a characteristic diagram showing the motor torque of the left front wheel (FL wheel), the motor torque of the right front wheel (FR wheel), and the difference between the two. Here, the characteristic of the thick line indicates the control of the present embodiment, and the characteristic of the thin line indicates the conventional control.

  On the other hand, in the control of the present embodiment, since the ratio of the first torque correction coefficient Coef1 is high at low μ, the left-right torque distribution according to the ground load is changed to the left-right torque distribution according to the side slip angle of the front and rear tire positions. Control is performed to increase the motor torque outside the turn and decrease the motor torque inside the turn. Therefore, as shown in FIG. 18, it can be seen that control is performed to increase the motor torque of the right front wheel (FR wheel) with respect to the left front wheel (FL wheel).

  Similarly, for the rear wheels, as shown in FIG. 19, control is performed to increase the motor torque outside the turn and decrease the motor torque inside the turn. By setting the motor torque of each wheel as shown in FIG. 18 and FIG. 19, it is possible to realize a driving force distribution according to the state of each wheel on the front, rear, left and right according to the ground load and the side slip angle.

  FIG. 20 is a characteristic diagram showing the left-right difference between the front and rear torques, the characteristic of the motor torque difference between the left front wheel (FL wheel) and the right front wheel (FR wheel) (FR-FL), and the left rear wheel (RL wheel). The graph shows the characteristic (RR-RL) of the motor torque difference between the right rear wheel and the right rear wheel (RR wheel). As shown in FIG. 20, in the conventional control, the front and rear torques are controlled at the same timing. On the other hand, in the control of this embodiment, a phase difference corresponding to the ground load and the side slip angle is attached to the front and rear torques. Therefore, by distributing the driving force according to the state of each wheel, the phenomenon that the rear wheel turns around the front wheel can be suppressed, and the overshoot of the motor torque (left-right difference) can be reduced compared to the conventional control. .

  FIG. 21 is a characteristic diagram showing a change in yaw rate in the examples of FIGS. FIG. 22 is a characteristic diagram showing changes in lateral acceleration in the examples of FIGS. As shown in FIG. 21, in the conventional control, the change in the yaw rate in the initial stage of the steering becomes linear, while the yaw rate vibrates during the steering, and the change in the yaw rate becomes dull before the steady steering is started, and the driver feels uncomfortable. These are caused by the distribution of the driving force proportional to the steering in spite of the decrease in μ and the transmission of the driving force from the motor to the road surface. On the other hand, in the control of the present embodiment, the yaw rate gradually increases until the steady steering (after time 6) is performed by distributing the driving force of each wheel according to the ground load or the front and rear sideslip angle, but until immediately before the steady steering is entered. Yaw rate vibration is suppressed. Therefore, according to the control of the present embodiment, the response performance of the vehicle can be ensured while appropriately distributing the driving force of each wheel in a low μ environment.

  Similarly, as shown in FIG. 22, in the conventional control, the change in the lateral acceleration at the initial stage of steering becomes linear, while the change in the lateral acceleration becomes dull in the middle of steady steering, the vehicle behavior changes, and the driver feels uncomfortable. Will occur. On the other hand, in the control of the present embodiment, the lateral acceleration is gradually increased until the steady steering is entered, compared with the conventional control, and the responsiveness of the vehicle can be ensured in a low μ environment. Therefore, the driver's uncomfortable feeling is also reduced.

  Next, the effect of this embodiment will be described based on the values of the first torque correction coefficient Coef1, the second torque correction coefficient Coef2, and the weighting coefficient τ based on FIGS. 23 and 24 are characteristic diagrams showing the motor torque of each wheel and the difference between the left and right torques when the same input as in FIGS. 11 and 12 is given at high μ (for example, μ = 1.0).

  FIG. 23 is a characteristic diagram showing the motor torque of the left front wheel (FL wheel), the motor torque of the right front wheel (FR wheel), and the difference between the two. FIG. 24 is a characteristic diagram showing the motor torque of the left rear wheel (RL wheel), the motor torque of the right rear wheel (RR wheel), and the difference between the two. FIG. 25 is a characteristic diagram showing the difference between the left and right torques (the sum of the front and rear wheels) and the weighting gain β_diffτ in the case of FIGS. FIG. 26 is a characteristic diagram showing characteristics of the yaw rate, lateral acceleration, and lateral movement in the case of FIGS. 23 to 26, the solid line indicates the characteristics of the present embodiment, the one-dot chain line indicates the characteristics when the first torque correction coefficient Coef1 is 100%, and the two-dot chain line indicates the second torque correction coefficient when the second torque correction coefficient is 100%. The broken line indicates the characteristic of the conventional control.

  As shown in FIG. 25, according to the control of the present embodiment, the value of the weighting gain τ decreases from before the steady steering (after time 6) is entered. This is because a difference occurs between β_mdl and β_clc because β_mdl is calculated using a vehicle model that does not involve control.

  As shown in FIGS. 23 and 25, at the time of high μ, the control characteristic (solid line) of the present embodiment is when the second torque correction coefficient Coef2 is set to 100% (two-dot chain line) until around time 4.5. ). Even after time 4.5, compared to the conventional control, the motor torque on the left wheel is lower than that on the conventional control, and on the right wheel, the motor torque is higher than that on the conventional control. It can be seen that the left and right motor torques are controlled accordingly.

  As shown in FIGS. 23 and 25, after time 4.5, motor torque overshoot is suppressed more than when the second torque correction coefficient Coef2 is set to 100% (two-dot chain line). As a result, as shown in FIG. 26, the yaw rate and lateral acceleration also increase gradually from the start of steering to steady steering. Therefore, it is understood that natural feeling turning can be realized by optimally allocating the first torque correction coefficient and the second torque correction coefficient based on the weighting coefficient τ.

  27 and 28 are characteristic diagrams showing the motor torque of each wheel and the difference between the left and right torques when FIG. 11 and FIG. 12 are used as input data at low μ (for example, μ = 0.3). FIG. 27 is a characteristic diagram showing the motor torque of the left front wheel (FL wheel), the motor torque of the right front wheel (FR wheel), and the difference between the two. FIG. 28 is a characteristic diagram showing the motor torque of the left rear wheel (RL wheel), the motor torque of the right rear wheel (RR wheel), and the difference between the two. FIG. 29 is a characteristic diagram showing the difference between the left and right torques (the sum of the front and rear wheels) and the weighting gain τ in the case of FIGS. FIG. 30 is a characteristic diagram showing characteristics of the yaw rate, lateral acceleration, and lateral movement in the case of FIGS. 27 to 30, the solid line indicates the characteristics of the present embodiment, the dashed-dotted line indicates the characteristics when the first torque correction coefficient Coef1 is 100%, and the dashed-two dotted line indicates the characteristics when the second torque correction coefficient is 100%. The broken line indicates the characteristic of the conventional control.

  As shown in FIG. 29, according to the control of the present embodiment, the value of the weighting gain τ decreases and the weighting of the first torque correction coefficient Coef1 increases before the steady steering (after time 6) starts. This is because a difference occurs between β_mdl and β_clc by calculating β_mdl with a vehicle model that does not involve control and a vehicle model that assumes a high μ road surface.

  As shown in FIGS. 27 and 28, at the time of low μ, the control characteristic (solid line) of the present embodiment is when the second torque correction coefficient Coef2 is set to 100% until the time around 4.5 (two-dot chain line). ). Even after time 4.5, the motor torque is suppressed on the left wheel compared to the conventional control compared to the conventional control, and torque distribution according to the turning situation is performed to the left wheel and the right wheel compared to the conventional control. For example, when turning left, it can be seen that the motor torque is distributed by the first torque correction coefficient Coef1 set according to the side slip angle.

  Further, as shown in FIGS. 27 and 28, the torque distribution between the FL wheel and the FR wheel is reduced after time 4.5 as compared with the case where the first torque correction coefficient Coef1 is set to 100% (two-dot chain line). On the other hand, while the amount of overshoot is reduced, the drive torque distribution amount to the RL wheel and the RR wheel is increased, so that the drive force distribution according to the load distribution can be performed, and the vehicle is moved from the initial stage of steering. The vehicle behavior up to steady steering can be made smooth while reducing the feeling. Therefore, it is understood that natural feeling turning can be realized by optimally allocating the first torque correction coefficient and the second torque correction coefficient based on the weighting coefficient τ.

  In addition, from the environmental information detected by the external recognition unit composed of a stereo camera or the like (“curvature of road surface”, “lateral deviation between own vehicle center line (front / rear axis) and target travel path”, etc.) State quantities such as a target yaw rate to be applied and a target lateral acceleration may be calculated, and a drive target torque to be applied to each wheel may be instructed.

  In addition, when the ground load estimation unit 240 estimates the ground load of each wheel, information detected by a device related to braking / driving, such as wheel speed, accelerator opening, and brake pedaling force, instead of acquiring longitudinal acceleration and lateral acceleration from the sensor. The longitudinal acceleration may be estimated from the above, and the lateral acceleration may be estimated from parameters relating to the vehicle motion such as the vehicle speed V and the steering angle θH.

  Further, the longitudinal acceleration after the foreseeing time elapses may be estimated from the transition of the target vehicle speed output from the driving force control unit such as a cruise control system, and may be used as the longitudinal acceleration used by the ground load estimation unit 240. The lateral acceleration after the foreseeing time may be estimated from the surrounding environment information output from the external environment recognition unit constituted by:

  As described above, according to the present embodiment, the torque correction coefficient TrqAdjustCoef can be calculated by optimally allocating the first torque correction coefficient and the second torque correction coefficient based on the parameter (β_ref) representing the road surface condition. . Therefore, it is possible to greatly improve the turning performance at both high μ and low μ.

  The preferred embodiments of the present invention have been described in detail above with reference to the accompanying drawings, but the present invention is not limited to such examples. It is obvious that a person having ordinary knowledge in the technical field to which the present invention pertains can come up with various changes or modifications within the scope of the technical idea described in the claims. Of course, it is understood that these also belong to the technical scope of the present invention.

200 Control Device 220 Drive Target Torque Calculation Unit 230 Side Slip Angle Calculation Unit 240 Ground Load Estimation Unit 252 First Torque Correction Coefficient Calculation Unit 254 Second Torque Correction Coefficient Calculation Unit 256 Torque Correction Gain Calculation Unit 260 Drive Required Torque Correction Processing Unit

Claims (9)

A skid angle obtaining unit for obtaining a skid angle of each wheel;
A first torque correction coefficient calculating unit that calculates a first torque correction coefficient of each wheel based on the sideslip angle;
A contact load estimation unit for estimating the contact load of each wheel;
A second torque correction coefficient calculating unit that calculates a second torque correction coefficient of each wheel based on the ground load;
Based on the parameter representing the friction coefficient of the road surface, the distribution of the first torque correction coefficient is increased when the friction coefficient is small, and the distribution of the second torque correction coefficient is increased when the friction coefficient is large. A torque correction coefficient calculation unit for calculating a torque correction coefficient of each wheel from the first and second torque correction coefficients;
A correction processing unit for correcting the drive target torque of each wheel based on the torque correction coefficient;
A vehicle control device comprising:   The side slip angle acquisition unit is configured to calculate the first side slip angle when the difference is small based on the difference between the first side slip angle of each wheel obtained from the vehicle model and the second side slip angle of each wheel obtained from the sensor. The distribution is increased, and when the difference is large, the distribution of the second side slip angle is increased, and the side slip angle of each wheel is calculated from the first and second side slip angles. The vehicle control device according to claim 1.   The side-slip angle acquisition unit acquires the first side-slip angle of each wheel from the side-slip angle of the vehicle center-of-gravity position obtained from a vehicle model, and the second side-slip angle of each wheel from the side-slip angle of the vehicle center-of-gravity position obtained from a sensor. The vehicle control device according to claim 2, wherein:   The first torque correction coefficient calculation unit increases the first torque correction coefficient as the side slip angle increases with respect to the wheel on the outer side of the turn, and increases the first torque correction coefficient as the side slip angle increases with respect to the wheel on the inner side of the turn. The vehicle control device according to claim 1, wherein the vehicle control device is made smaller.   5. The vehicle control device according to claim 1, wherein the second torque correction coefficient calculation unit increases the second torque correction coefficient as the ground load increases. 6.   2. The vehicle control device according to claim 1, wherein the ground load estimation unit estimates a ground load of each wheel based on at least a vehicle weight, a longitudinal acceleration, and a lateral acceleration.   The torque correction coefficient calculation unit calculates the torque correction coefficient using a difference between a skid angle of a vehicle center of gravity position obtained from a vehicle model and a skid angle of a vehicle center of gravity position obtained from a sensor as the parameter. Item 7. The vehicle control device according to any one of Items 1 to 6.   The vehicle control device according to claim 1, further comprising a drive target torque calculation unit that calculates the drive target torque of each wheel based on a vehicle speed, a steering angle, and a yaw rate. Obtaining the sideslip angle of each wheel;
Calculating a first torque correction coefficient for each wheel based on the sideslip angle;
Estimating the contact load of each wheel;
Calculating a second torque correction coefficient for each wheel based on the ground load;
Based on the parameter representing the friction coefficient of the road surface, the distribution of the first torque correction coefficient is increased when the friction coefficient is small, and the distribution of the second torque correction coefficient is increased when the friction coefficient is large. Calculating a torque correction coefficient for each wheel from the first and second torque correction coefficients;
Correcting the drive target torque of each wheel based on the torque correction coefficient;
A vehicle control method comprising:

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