JP2006069519A - Vehicle motion control device and vehicle motion control method - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve the steering stability of a vehicle by stabilizing the motion state of the vehicle, regardless of the traveling environment. <P>SOLUTION: A calculation part 10a calculates, based on a system matrix of equations of state, showing the state of motion of the vehicle, nonlinear terms having a slide angle βf for the front wheels and a slide angle βr for the rear wheels as variables, which are included in a polynominal expression showing at least one of a plurality of elements a11-a22 constituting the system matrix. A setting part 10b sets, based on the calculated nonlinear terms, a target value r* of driving force distribution ratio r or load distribution ratio r' to the respective wheels. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a vehicle motion control apparatus and method.

2. Description of the Related Art Conventionally, a technique for controlling a motion state of a vehicle by controlling a driving force distribution ratio (or load distribution ratio) for each wheel is known. For example, in a driving situation such as cornering, it is possible to improve the operability by controlling the driving force distribution ratio so that the motion state of the vehicle is appropriate. One such technique is a device that controls the motion state of a vehicle using a wheel frictional force utilization rate (see, for example, Patent Document 1). In this device, the wheel friction force utilization factor of each wheel is obtained, and the wheel state of each wheel is controlled so that the wheel friction force utilization factor approaches the target wheel friction force utilization factor. In addition, in the state equation representing the vehicle motion state, based on the knowledge that each element constituting the system matrix of this state equation affects the vehicle motion state, focusing on this element, the vehicle motion state is determined. There is also a method of controlling (see, for example, Patent Document 2).
Japanese Patent No. 3132190 JP-A-11-102499

  However, the motion state of the vehicle is closely related to the motion of the wheels provided on the vehicle, and there is a problem that the operability is lowered due to the nonlinearity of the wheels. Therefore, in order to obtain a stable vehicle response, it is preferable to control the motion state of the vehicle by paying attention to the nonlinearity of the wheels.

  An object of the present invention is to provide a novel method for controlling the motion state of a vehicle.

  Another object of the present invention is to improve the maneuverability of the vehicle.

  In order to solve this problem, the first invention provides a vehicle motion control device. The control device includes a calculation unit and a setting unit. The calculation unit is based on the system matrix of the state equation representing the motion state of the vehicle, and includes the slip angle of the front wheel and the rear wheel included in the polynomial representing at least one of the plurality of elements constituting the system matrix. Calculate a nonlinear term with the slip angle as a variable. The setting unit sets a target value of the driving force distribution ratio or a target value of the load distribution ratio for each of the wheels based on the calculated nonlinear term.

  Here, in the first invention, the vehicle motion control device estimates the friction coefficient between the wheel and the road surface based on the detection unit that detects the vertical force acting on each of the wheels and the state of the vehicle. And an estimation unit that estimates the sum of the driving forces applied to each of the wheels as a total driving force. In this case, the calculation unit calculates and sets a plurality of nonlinear terms corresponding to different driving force distribution ratios based on a parameter group including the vertical force of each wheel, the friction coefficient, and the total driving force. The unit preferably sets, as a target value, a driving force distribution ratio in which the absolute value of the nonlinear term becomes smaller than the current value based on the plurality of calculated nonlinear terms. Furthermore, the setting unit may set the driving force distribution ratio that minimizes the absolute value of the nonlinear term as the target value.

  In the first aspect of the invention, the vehicle motion control device calculates the friction coefficient between the wheel and the road surface based on the state of the vehicle and the detection unit that detects the vertical force and the longitudinal force acting on each of the wheels. You may further have an estimation part to estimate. In this case, the calculation unit calculates a plurality of nonlinear terms corresponding to different load distribution ratios based on a group of parameters including the vertical force of each wheel, the longitudinal force of each wheel, and the friction coefficient. The setting unit preferably sets, as the target value, a load distribution ratio in which the absolute value of the nonlinear term becomes smaller than the current value based on the plurality of calculated nonlinear terms. Furthermore, the setting unit may set a load distribution ratio that minimizes the absolute value of the nonlinear term as a target value.

  In the first invention, the estimation unit further estimates the slip angle of the front wheel and the slip angle of the rear wheel based on the slip angle of the vehicle, and the calculation unit calculates the slip angle of the front wheel and the slip angle of the rear wheel. It is preferable to calculate the nonlinear term based on the angle and the parameter group. Further, it is desirable that the detection unit further detects a lateral force acting on each of the wheels, and the calculation unit calculates a nonlinear term based on each lateral force of the wheel and a parameter group. In this case, the calculation unit calculates the maximum value of the driving force applied to the wheel based on the vertical force of the wheel and the friction coefficient for each of the wheels, and determines the maximum value of the driving force of each wheel. The nonlinear term may be calculated based on the lateral force of each wheel and the parameter group. Further, the calculation unit calculates the maximum value of the driving force applied to the wheel based on the vertical force and the friction coefficient of the wheel for each of the wheels, and the maximum value of the driving force of each wheel, It is preferable to calculate the nonlinear term based on the parameter group.

  In the first invention, the calculation unit preferably calculates a nonlinear term in a polynomial that is a sum of diagonal elements in the system matrix.

  The second invention provides a vehicle motion control device. The control device includes a calculation unit and a setting unit. In the state plane expressed by the vector field with the state quantity representing the motion state of the vehicle as the axis, the calculation unit converts the divergence as the characteristic value indicating the tendency of the vector in the vector field into a state equation representing the motion state of the vehicle. Calculate based on Based on the calculated divergence, the setting unit sets the target value or the load distribution ratio of the driving force distribution ratio for each of the wheels so that the divergence is smaller than the current value or the divergence is 0 or less. Set the target value.

  The third invention provides a vehicle motion control device. The control device includes a calculation unit and a setting unit. The calculation unit calculates attenuation, which is a characteristic value indicating the convergence of the vehicle as the vibration system, based on a state equation representing the motion state of the vehicle. The setting unit sets a target value of the driving force distribution ratio or a target value of the load distribution ratio for each of the wheels based on the calculated attenuation so that the attenuation is larger than the current value.

  Here, in the second invention or the third invention, the vehicle motion control device detects the vertical force acting on each of the wheels, and the friction between the wheels and the road surface based on the state of the vehicle. The apparatus may further include an estimation unit that estimates the coefficient and estimates the total driving force applied to each wheel as the total driving force. In this case, the calculation unit configures a system matrix of a state equation for each of a plurality of different driving force distribution ratios based on a parameter group including the vertical force of each wheel, the friction coefficient, and the total driving force. A non-linear term that includes the slip angle of the front wheel and the slip angle of the rear wheel included in the polynomial that is the sum of the diagonal elements to be calculated, and the setting unit drives based on each of the calculated non-linear terms It is preferable to set a target value of the force distribution ratio.

  In the second invention or the third invention, an estimation for estimating a friction coefficient between the wheel and the road surface based on the state of the vehicle and the detection unit that detects the vertical force and the longitudinal force acting on each of the wheels. May further have a part. In this case, the calculation unit calculates the system matrix of the state equation for each of a plurality of different load distribution ratios based on a parameter group including the vertical force of each wheel, the longitudinal force of each wheel, and the friction coefficient. A non-linear term that includes the slip angle of the front wheel and the slip angle of the rear wheel, included in the polynomial that is the sum of the diagonal elements that constitute, is calculated, and the setting unit is based on each of the calculated non-linear terms. It is desirable to set a target value for the load distribution ratio.

  The fourth invention provides a vehicle motion control method. In this control method, as a first step, based on a system matrix of a state equation representing a motion state of a vehicle, a front wheel included in a polynomial representing at least one element of a plurality of elements constituting the system matrix. A non-linear term is calculated with the slip angle and the slip angle of the rear wheel as variables. As a second step, the target value of the driving force distribution ratio or the target value of the load distribution ratio for each of the wheels is set based on the calculated nonlinear term.

  Here, in the fourth invention, the vehicle motion control method estimates the friction coefficient between the wheel and the road surface based on the third step of detecting the vertical force acting on each of the wheels and the state of the vehicle. In addition, a fourth step of estimating the sum of the driving forces applied to each of the wheels as a total driving force may be further included. In this case, the first step calculates a plurality of nonlinear terms corresponding to different driving force distribution ratios based on a parameter group including the vertical force of each wheel, the friction coefficient, and the total driving force. The second step is a step of setting, as a target value, a driving force distribution ratio in which the absolute value of the nonlinear term becomes smaller than the current value based on the plurality of calculated nonlinear terms. preferable.

  In the fourth invention, the vehicle transportation control method includes a third step of detecting the vertical force and the longitudinal force acting on each of the wheels, and the friction between the wheels and the road surface based on the state of the vehicle. And a fourth step of estimating a coefficient. In this case, the first step is to calculate a plurality of nonlinear terms corresponding to different load distribution ratios based on a group of parameters including the vertical force of each wheel, the longitudinal force of each wheel, and the friction coefficient. The second step is a step of setting, as a target value, a load distribution ratio in which the absolute value of the nonlinear term becomes smaller than the current value based on the plurality of calculated nonlinear terms. Is desirable.

  The fifth invention provides a vehicle motion control method. In this control method, as a first step, a divergence as a characteristic value indicating a vector tendency in a vector field is expressed on a state surface expressed by a vector field with a state quantity representing a motion state of the vehicle as an axis. Calculation is made based on a state equation representing a motion state. As a second step, based on the calculated divergence, the target value or load of the driving force distribution ratio for each of the wheels so that the divergence is smaller than the current value or the divergence is 0 or less. Set a target value for the distribution ratio.

  The sixth invention provides a vehicle motion control method. In this control method, as a first step, attenuation, which is a characteristic value indicating the convergence of the vehicle as the vibration system, is calculated based on a state equation representing the motion state of the vehicle. As a second step, based on the calculated attenuation, the target value of the driving force distribution ratio or the target value of the load distribution ratio for each of the wheels is set so that the attenuation is greater than the current value.

  According to the present invention, the driving force distribution ratio or the load distribution ratio is set based on the state equation representing the motion state of the vehicle, and this can be set in consideration of the nonlinearity of the wheels. Therefore, by controlling the driving force distribution ratio or the load distribution ratio for each wheel based on the set target value, for example, it is possible to suppress the vehicle response due to the non-linearity of the wheel. As a result, the operability of the vehicle can be improved.

  Prior to the description of the system configuration and system processing related to the vehicle motion control apparatus according to the present embodiment, first, the system matrix in the state equation will be described in order to clarify the control concept. Since the braking force can be considered as a reverse component (minus component) of the driving force, the term driving force is used in the present embodiment to include the braking force.

FIG. 1 is an explanatory diagram of a vehicle model. The vehicle model shown in the figure is a two-wheel model in which the motion state of the vehicle is expressed by two front and rear wheels. In this vehicle model, the motion state of the vehicle is expressed by a rotational motion (yaw motion) around the vertical axis (Z-axis) and a translational motion in the lateral direction (Y-axis direction). When the front wheels are being steered (the rear wheels are parallel to the X-axis direction), the motion state of the vehicle is represented by the state equation shown below.

Here, βb is the slip angle of the vehicle body (hereinafter referred to as “vehicle slip angle”), γ is the yaw rate, and Δf is the steering angle of the front wheels. Further, in this specification, a symbol “′” is attached to each state quantity to represent a time differential value of the state quantity. For example, βb ′ indicates a vehicle slip angular velocity that is a time differential value of the vehicle slip angle βb, and γ ′ indicates a yaw angular acceleration that is a time differential value of the yaw rate γ. The equation of state shown in Equation 1 can be represented by the block diagram shown in FIG. This block diagram includes a block having an integral indicated by “1 / s” as an element, and shows a time-series change of this state equation (vehicle motion state). In the equation, each matrix having a11 to a22 and b1 and b2 as elements is called a system matrix of the state equation, and each element a11 to b2 of the system matrix can be expressed by each equation in Equation 2. it can.

  In the equation, V is the vehicle speed, M is the vehicle mass, lf is the distance between the front wheel axis and the center of gravity, lr is the distance between the rear wheel axis and the center of gravity, Iz is the moment of inertia of the vehicle around the Z axis, and ka is This is the cornering power considering the non-linearity of the wheel. In the present specification, the wheels (front and rear wheels) indicated by the respective state quantities are explicitly distinguished by using alphabets “f” representing front wheels and “r” representing rear wheels. For example, kf_a indicates the cornering power ka of the front wheel, and kr_a indicates the cornering power ka of the rear wheel. When a four-wheeled vehicle is assumed, the state quantity of the front wheels (or rear wheels) indicated by the alphabet “f” (or “r”) is, for example, the average value of the state quantities of the left and right front wheels (or left and right rear wheels). Can be considered.

  The cornering power ka is the cornering force for a slight change in the wheel slip angle (hereinafter referred to as “wheel slip angle”) βw (the wheel traveling direction of the frictional force generated on the wheel contact surface when turning at a certain wheel slip angle βw). The rate of change in the component force generated in the direction perpendicular to. That is, the cornering power ka is the inclination (differential value) of the cornering force at a certain wheel slip angle βw. The cornering power ka is a parameter that has a great influence on the operability of the vehicle. When this value is large, the response of the behavior change is fast, and when this value is small, the response of the behavior change is slow. .

  FIG. 3 is an explanatory diagram of the acting force acting on the wheel. As the acting force acting on the wheel, in addition to the above-described cornering force, there are a longitudinal force Fx, a lateral force Fy, a vertical force Fz, and the like. When turning at a wheel slip angle βw, the component force generated in the direction parallel to the wheel center plane among the friction force generated on the ground contact surface is the longitudinal force Fx, and the component force generated in the direction perpendicular to the wheel center plane Is the lateral force Fy. The vertical force Fz is a load in the vertical direction, so-called vertical load.

  Of these forces listed as acting forces, the cornering force and the lateral force Fy can be treated as similar forces. Although these forces are not strictly coincident values, they tend to approximate within the range of the wheel slip angle βw that the vehicle can take. In this specification, the cornering force and the lateral force Fy are regarded as substantially the same value, and the cornering power ka is considered based on the lateral force Fy. Therefore, the relationship between the lateral force Fy and the cornering power ka will be considered below.

The lateral force Fy is a quadratic expression of the wheel slip angle βw by using a tire model indicating the dynamic characteristics of the wheel, for example, a fiala tire model (secondary approximation model) that can take into account the influence of nonlinearity. Can be expressed as

In this case, the cornering power ka of the wheel satisfies the relationship shown below. In other words, the cornering power ka is a change rate (differential value) of the lateral force Fy with respect to a minute change in the wheel slip angle βw.

In Equations 3 and 4, the coefficient k is a constant that can be obtained experimentally, and is referred to as reference cornering power. The reference cornering power k changes depending on the coefficient of friction μ between the road surface and the wheel and the vertical force Fx, and indicates the characteristics of the wheel. Specifically, when this value k is large, it means that the rigidity of the wheel is high, and when this value k is small, it means that the rigidity of the wheel is low. The reference cornering power k is defined as the rate of change (differential value) of the lateral force Fy when the wheel slip angle βw is “0”.

On the other hand, the maximum lateral force value Fymax that can be taken by the lateral force Fy is uniquely calculated based on the vertical force Fz, the longitudinal force Fx, and the friction coefficient μ.

In consideration of the non-linearity of the wheels, the elements a11 to b2 can be rewritten as Equation 7.

In the equation, kf is an equivalent cornering power in consideration of elastic deformation of the suspension with respect to the reference cornering power k of the front wheel, and kr is an equivalent cornering power in consideration of elastic deformation of the suspension with respect to the reference cornering power k of the rear wheel. βf is a front wheel slip angle, for example, an average value of left and right front wheel slip angles βw (hereinafter referred to as “average front wheel slip angle”), and βr is a rear wheel slip angle, for example, left and right rear wheel slip. The average value of the angle βw (hereinafter referred to as “average rear wheel slip angle”). These slip angles βf and βr can be expressed by the following mathematical expressions.

  Elements a11 and a22 are parameters that affect the stability (ease of convergence of behavior) of the vehicle, and are called diagonal elements of the system matrix. In particular, the element a11 autonomously stabilizes the lateral movement, and the element a22 autonomously stabilizes the yaw movement. Elements a12 and a21 are parameters that affect the improvement of vehicle responsiveness (ease of vibration of behavior), and are called coupled elements of the system matrix. When the elements a12 and a21 are relatively small with respect to the elements a11 and a22, the stability of the vehicle during high speed traveling is improved. On the other hand, when the elements a12 and a21 are relatively larger than the elements a11 and a22, the response of the vehicle to steering is improved. On the other hand, the elements b1 and b2 are vehicle behavior gains with respect to the driver's steering, and can be adjusted by the steering gear ratio or the like.

  In the present embodiment, a vehicle in which the motion characteristics of the vehicle do not change greatly even when the driving environment changes is considered as one ideal, and the main control purpose is to suppress changes in the elements a11 to a22 of the elements a11 to b2. And When the elements a11 to a22 are analyzed analytically, the individual elements a11 to a22 are, as shown in Expression 7, a term depending on vehicle mass M, vehicle specifications such as a wheel base, and equivalent cornering powers kf and kr , The average front and rear wheel slip angles βf and βr, the vertical force Fz, and the term depending on the friction coefficient μ. In other words, the individual elements a11 to a22 are represented by the sum of a linear term that changes due to the linearity of the wheel and a nonlinear term that changes due to the nonlinearity of the wheel. When only the linear term acts and the elements a11 to a22 change, the wheel motion has a linear property, so that there is no particular problem from the viewpoint of maneuverability. However, when the elements a11 to a22 change due to the action of a nonlinear term, the wheel motion has a nonlinear property, which may reduce the maneuverability. Based on the knowledge that the change in the nonlinear term is due to the driving force distribution ratio for the front and rear wheels, the absolute value of the nonlinear term is set to the current value (current Smaller than the absolute value of the value, and preferably the minimum (“0”). By minimizing the nonlinear term, the influence of the nonlinear term in the elements a11 to a22 is reduced, and the change in the elements a11 to s22 is suppressed.

The total driving force that is the sum of the driving forces given to each wheel is Fa, and the driving force distribution ratio is r. Front wheel front / rear force Ff_x, for example, the average value of front / rear front wheel front / rear force Fx (hereinafter referred to as “front wheel average front / rear force”) and rear wheel front / rear force Fr_x, Hereinafter, “the average longitudinal force of the rear wheel”) can be expressed by the following mathematical formula.

Minimization of Nonlinear Term of Element a11 FIG. 4 is an explanatory diagram showing the tendency of the driving force distribution ratio r that minimizes the nonlinear term of the element a11. The nonlinear term of the element a11 is a linear sum of the average front wheel slip angle βf (absolute value) and the average rear wheel slip angle βr (absolute value). Hereinafter, for simplicity of explanation, the average rear wheel slip angle βr is expressed by s times the average front wheel slip angle βf (hereinafter, this s is referred to as “wheel slip angle ratio”). The figure shows the vehicle acceleration Acc (total driving force Fa / vehicle mass M) at a predetermined wheel slip angle ratio (s = 0.1, 1.0, 10) under the condition that the friction coefficient is constant (for example, μ = 0.9). The correspondence relationship with the driving force distribution ratio r is illustrated (the same applies to FIGS. 6 and 7 described later). As shown in Equations 6, 7, and 9, the tendency shown in the figure is the average value of the vertical force Fz of the left and right front wheels as the friction coefficient μ, the total driving force Fa, and the front wheel vertical force Ff_z (hereinafter referred to as “average front wheel vertical For each wheel slip angle ratio s based on a group of parameters including an average value of the vertical force Fz of the left and right rear wheels as the vertical force Fr_z of the rear wheels (hereinafter referred to as “average vertical force of the rear wheels”). Can be specified. For example, a plurality of nonlinear terms corresponding to different driving force distribution ratios r are calculated, and a driving force distribution ratio r that minimizes the absolute value of the nonlinear terms is specified based on the calculated nonlinear terms. That's right. When the wheel slip angle ratio s is “1.0”, the driving force distribution ratio r is approximately in the range of “0.5” to “0.8”, and is deviated from the front wheels. On the other hand, when the wheel slip angle ratio s is smaller than “1.0”, the rear wheel wheel tends to be deviated compared to the case where the wheel slip angle ratio s is “1.0”, and the wheel slip angle ratio s is smaller than “1.0”. When the wheel slip angle ratio is large, the wheel slip angle ratio s tends to deviate from the front wheels as compared with the case where the wheel slip angle ratio s is “1.0”.

Minimization of Nonlinear Terms of Elements a12 and a21 FIG. 5 is a diagram showing the tendency of the driving force distribution ratio r that minimizes the coefficients of the nonlinear terms of the elements a12 and a21. The nonlinear terms of the elements a12 and a21 are linear sums of the average front wheel slip angle βf (absolute value) and the average rear wheel slip angle βr (absolute value), and are represented by the same polynomial. Similar to the element a11, the tendency shown in the figure can be uniquely specified based on the parameter group described above. Specifically, when the wheel slip angle ratio s is “1.0”, the driving force distribution ratio r is approximately “0.7” or more, which tends to be more deviated from the front wheels than the element a11. On the other hand, when the wheel slip angle ratio s is smaller than “1.0”, there is a tendency for the rear wheel to be deviated as compared with the case where the wheel slip angle ratio s is “1.0”. If it is large, the solution does not exist between “0.0” and “1.0”.

Minimization of Nonlinear Term of Element a22 FIG. 6 is an explanatory diagram showing the tendency of the driving force distribution ratio r that minimizes the coefficient of the nonlinear term of the element a22. The nonlinear term of the element a22 is also a linear sum of the average front wheel slip angle βf (absolute value) and the average rear wheel slip angle βr (absolute value). Similar to the elements a11 to a12, the tendency shown in the figure can be uniquely identified based on the friction coefficient μ, the total driving force Fa, and the average vertical forces Ff_z and Fr_z of the front and rear wheels. Specifically, compared to the driving force distribution ratio r of the element a11, the front wheel tendency is overall.

Of these elements a11 to a22, which element is focused on to minimize the nonlinear term is arbitrary, and can be appropriately selected from a11, a22, and a12 (or a21). Note that it is difficult to accurately measure the average front wheel slip angle βf and the average rear wheel slip angle βr included in the nonlinear term, so that these values βf and βr are not reflected in the control. difficult. Accordingly, in the range where the wheel slip angle βw is small, the integrated value of the reference cornering power (equivalent cornering power) k and the wheel slip angle βw is based on the knowledge that it corresponds to the lateral force Fy, and is included in the nonlinear term. The front wheel slip angle βf and the average rear wheel slip angle βr are replaced with the front and rear wheel lateral forces Ff_y and Fr_y, respectively. In other words, the nonlinear terms in the elements a11 to a22 are the average values of the parameter groups μ, Fa, Ff_z, Fr_z and the lateral force Fy of the left and right front wheels as the lateral force Ff_y of the front wheels (hereinafter referred to as “average lateral of the front wheels”). Force ”) and an average value of the lateral force Fy of the left and right rear wheels as the lateral force Fr_y of the rear wheel (hereinafter referred to as“ average lateral force of the rear wheel ”).

  The equation shows only the nonlinear term of the element a11, but the same replacement can be made for the nonlinear terms related to the other elements a12 to a22. However, since the basic concept is the same, the description thereof is omitted, but the driving force distribution ratio r that minimizes the nonlinear term of each element a11 to a22 is the friction coefficient μ and the average vertical force Ff_z of the front and rear wheels. , Fr_z, front and rear wheel average lateral forces Ff_y, Fr_y, and total driving force Fa can be calculated.

  Based on the above description of the concept, the system configuration of the vehicle motion control apparatus according to the present embodiment will be described. FIG. 7 is an explanatory diagram of a vehicle to which the vehicle motion control device 1 according to the present embodiment is applied. This vehicle is a four-wheel drive vehicle driven by four front and rear wheels. Power from a crankshaft (not shown) of the engine 2 that is a drive source is respectively transmitted to a front wheel side and a rear wheel side drive shaft (axle) 5 via an automatic transmission 3 and a center differential device 4 described later. Communicated. When power is transmitted to the drive shaft 5, rotational torque is applied to each of the front wheel 6f and the rear wheel 6r, and the front and rear wheels 6f and 6r rotate, whereby a driving force is applied to the front and rear wheels 6f and 6r. In this specification, when the front wheel 6f and the rear wheel 6r are collectively referred to, the term “wheel 6” is simply used.

  The center differential device 4 is a compound planetary gear type differential limiting device. A hydraulic multi-plate clutch 4c is provided between two output members of the center differential device 4, that is, an output member (carrier 4a) on the front wheel side and an output member (sun gear 4b) on the rear wheel side. . In the hydraulic multi-plate clutch 4c, its own engagement state can be adjusted by adjusting the hydraulic pressure (for example, increasing pressure, holding and reducing pressure) by the solenoid valve 4d. In a state where the hydraulic multi-plate clutch 4c is released, the driving shafts 5 of the front and rear wheels are different from each other, so that the driving force distribution ratio r becomes a value depending on the preset device 4 (for example, r = 0.35). . On the other hand, in the state where the hydraulic multi-plate clutch 4c is completely engaged, the differential of the drive shaft 5 of the front and rear wheels is limited, so that the driving force distribution ratio r in the front-rear direct connection state is obtained. That is, the driving force distribution ratio r is variably set according to the engaged state of the hydraulic multi-plate clutch 4c.

  FIG. 8 is a block diagram showing the overall configuration of the vehicle motion control device 1. The vehicle motion control device 1 is mainly configured by a control unit 10. As the control unit 10, a microcomputer mainly including a CPU, a ROM, a RAM, and an input / output interface can be used. The control unit 10 performs calculations related to vehicle motion control in accordance with a control program stored in the ROM. If the microcomputer as this control part 10 is understood functionally, the control part 10 has the calculation part 10a and the setting part 10b. In the present embodiment, the calculation unit 10a calculates a nonlinear term representing an element of this matrix (in the present embodiment, an element a11) based on the system matrix. The setting unit 10b calculates a target value r * of the driving force distribution ratio r based on the calculated nonlinear term. In order to perform such calculation, the control unit 10 receives detection signals from various sensors including the detection unit 11 and various calculation values estimated by the estimation unit 12 (not shown in FIG. 7). (In this embodiment, the friction coefficient μ and the total driving force Fa) are also input.

  The detection unit 11 is a sensor that detects an acting force acting on the wheel 6. In FIG. 8, only one block corresponding to the detection unit 11 is shown for convenience of explanation, but this detection unit 11 is provided on each of the four wheels 6. The individual detection units 11 can detect the longitudinal force Fx, the lateral force Fy, and the vertical force Fz as individual forces as acting forces. The detection unit 11 is mainly configured by a strain gauge and a signal processing circuit that processes an electrical signal output from the strain gauge and generates a detection signal corresponding to the acting force. Based on the knowledge that the stress generated in the drive shaft (axle) 5 is proportional to the acting force, the detecting unit 11 directly detects the acting force by embedding a strain gauge in the axle 5. The specific configuration of the detection unit 11 is disclosed in, for example, Japanese Patent Application Laid-Open Nos. 04-331336 and 10-318862, and should be referred to if necessary. As will be described later, the estimation unit 12 estimates the friction coefficient μ and the total driving force Fa based on detection results from various sensors (not shown).

  FIG. 9 is a flowchart showing a vehicle control procedure according to the present embodiment. The process shown in this flowchart is called at predetermined intervals and executed by the vehicle motion control device 1. First, in step 1, various detection signals are read. Examples of the detection value read in step 1 include the acting force (lateral force Fy, vertical force Fz) at each wheel 6.

  In step 2, the friction coefficient μ and the total driving force Fa are estimated. First, as a method for estimating the friction coefficient μ, for example, a method using a vehicle motion model obtained by modeling the behavior of a vehicle based on the vehicle motion theory is well known. In this method, based on the actual vehicle motion state (for example, slip angle), for example, the vehicle motion model motion state assuming a high μ road and the vehicle motion model motion state assuming a low μ road are obtained. By comparison, the current friction coefficient μ is estimated. Details of the method for estimating the friction coefficient μ are disclosed in, for example, Japanese Patent Application Laid-Open No. 2000-071968, and should be referred to if necessary. In addition to this, for example, as disclosed in JP 2003-237558 A, the friction coefficient μ may be estimated based on the speed difference between the two wheels 6 and the acceleration. Further, for example, as disclosed in Japanese Patent Application Laid-Open No. 2002-27882, the friction coefficient μ is estimated in consideration of a detection result obtained by detecting a road surface condition obtained from a camera in a vehicle motion state. May be. Thus, in the present embodiment, a method for estimating the friction coefficient based on the state of the vehicle can be widely used. Here, as described above, the vehicle state includes various states such as a vehicle motion state such as a slip angle and a yaw rate, a wheel speed, and a road surface in contact with the wheel 6.

  On the other hand, the total driving force Fa is calculated based on the output torque of the engine 2 corresponding to the engine speed and the output torque of the automatic transmission 3 corresponding to the shift position and the like. Specifically, first, the speed ratio of the torque converter is calculated based on the turbine speed of the torque converter in the automatic transmission 3 and the engine speed. Accordingly, a map showing the correspondence between the speed ratio of the torque converter and the pump capacity coefficient is referred to, and the pump capacity coefficient corresponding to the calculated speed ratio is specified. On the other hand, the speed ratio of the torque converter is based on a map showing the correspondence between the torque converter input-side torque ratio and the torque converter output-side torque ratio (hereinafter referred to as “torque ratio”). A torque ratio corresponding to the ratio is identified.

  Next, pump torque is calculated based on the specified pump capacity coefficient and the engine speed, and turbine torque output from the turbine is calculated based on the pump torque and the torque ratio. Then, the output torque of the automatic transmission 3 is calculated by multiplying the turbine torque by a gear ratio corresponding to the current gear position. Finally, the drive torque is calculated by multiplying the output torque of the automatic transmission 3 by the final gear ratio of the automatic transmission 3. Thereby, the total driving force Fa is calculated based on the wheel radius and the driving torque.

  In step 3, a target value r * of the driving force distribution ratio r is calculated. The driving force distribution ratio r that minimizes the nonlinear term of the element a11 can be calculated according to the above-described Expression 10. Specifically, the driving force distribution ratio r is changed stepwise from “0.00” to “1.00”, the friction coefficient μ, the total driving force Fa, the average vertical forces Ff_z, Fr_z of the front and rear wheels 6f, 6r, Based on the average lateral forces Ff_y and Fr_y of the front and rear wheels 6f and 6r, the values of the nonlinear terms are sequentially calculated. The driving force distribution ratio r having the smallest value among the calculated nonlinear terms is set as the target value r *. The average lateral forces Ff_y and Fr_y of the front and rear wheels 6f and 6r and the average vertical forces Ff_z and Fr_z of the front and rear wheels 6f and 6r are unambiguous on the basis of the detected values (lateral force Fy or vertical force Fz) of the respective wheels 6. Can be identified.

  In step 4, the current value rc of the driving force distribution ratio r and the target value r * are compared, and it is determined whether or not the difference between them is larger than the determination value rth. The determination value rth is set in advance through experiments and simulations as a maximum value of the difference (absolute value) between the current value rc and the target value r * that can be regarded as requiring no control from the viewpoint of suppressing control hunting. Has been. If an affirmative determination is made in step 4, that is, if the difference between the two is larger than the determination value rth (| rc−r * |> rth), the process proceeds to step 5 following step 4. On the other hand, if a negative determination is made in step 4, that is, if the difference between the two is equal to or less than the determination value rth (| rc−r * | <rth), step 5 is skipped and the routine is exited.

  In step 5, by referring to a map or the like in which the correspondence relationship between the driving force distribution ratio r and the engagement state (engagement torque) is described, a control signal (instruction for engagement torque) corresponding to the set target value r * is obtained. Value) Sigr is output to the solenoid valve 4d, and this routine is exited. As a result, the duty ratio of the solenoid valve 4d is controlled according to the control signal Sigr, and the engagement state of the hydraulic multi-plate clutch 4c is adjusted. As a result, the driving force distribution ratio r is changed from the current value rc to the target value r *.

  Thus, according to this embodiment, based on the value of the nonlinear term of the element a11 of the system matrix, the driving force distribution ratio r that minimizes this value (absolute value) is set as the target value r *. Based on the set target value r *, the center differential device 4 (specifically, the solenoid valve 4d) that variably sets the driving force distribution ratio r between the front wheels 6f and the rear wheels 6r is controlled. . Thereby, since the nonlinear term of the element a11 is minimized and the nonlinear element acting on the wheel 6 is suppressed, a stable vehicle responsiveness can be obtained. Further, by minimizing the non-linear term, a change in the element a11 having a high correlation with the convergence property of the vehicle slip angle is suppressed, so that a stable vehicle response can be realized regardless of the driving environment.

  The detection unit 11 directly detects the acting force acting on the wheel 6. For this reason, for example, even if it is a driving situation such as limit cornering or a driving situation such as a road surface with a low friction coefficient, the acting force can be specified with high accuracy. As a result, it is possible to improve the calculation accuracy of the nonlinear term of the element a11, so that it is possible to more effectively control the motion state of the vehicle.

  In the above-described embodiment, for the purpose of minimizing the nonlinear term, the driving force distribution ratio r that minimizes the value is uniquely used as the target value r *. However, in terms of control stability, it is sufficient to set the target value r * so that the nonlinear term (absolute value) is smaller than the current value. For example, the current driving force distribution ratio r is changed by a step value. Even in this method, since the non-linear element acting on the wheel 6 is suppressed as compared with the current state, stable vehicle responsiveness can be obtained. Such a modification related to the minimization of the nonlinear term can be similarly applied to embodiments described later.

  In the above-described embodiment, from the viewpoint that it is difficult to actually measure the average front wheel slip angle βf and the average rear wheel slip angle βr, the driving force is changed after replacing these with the average lateral forces Ff_y and Fr_y of the front and rear wheels. A distribution ratio r was set. However, as shown in the nonlinear term of Equation 7, the nonlinear term may be calculated based on the average front and rear wheel slip angles βf and βr and the parameter groups μ, Fa, Ff_z, and Fr_z described above. In this case, the average front and rear wheel slip angles βf and βr can be uniquely calculated using Equation 8 by detecting the vehicle body slip angular velocity βb ′ and calculating the integral value by the estimation unit 12.

(Second Embodiment)
In the first embodiment, the value of the nonlinear term is calculated while changing the driving force distribution ratio r stepwise from “0.00” to “1.00” according to Equation 10, and the smallest driving force among these values is calculated. The distribution ratio r is determined as the target value r *. However, in Expression 10, the square root has a negative value, and the driving force distribution ratio r may not be calculated. Therefore, in actual control, it is preferable to calculate the driving force distribution ratio r based on the following equation.

In the equation, Ff_Lim is a maximum value of driving force applied to the front wheel 6f per wheel (hereinafter referred to as “front wheel maximum value”), and is an integrated value of the friction coefficient μ and the vertical force Ff_z per wheel of the front wheel 6f. It is. Fr_Lim is the maximum value of driving force applied to the rear wheel 6r per wheel (hereinafter referred to as “rear wheel maximum value”), and is an integration of the friction coefficient μ and the vertical force Ff_z per wheel of the rear wheel 6r. Value. On the other hand, Ff_div and Fr_div are expressed by the following relational expressions. Since Ff_Lim and Fr_Lim are amounts obtained separately for the left and right wheels, “Σ” shown in the following expression represents the sum of the front two wheels or the sum of the rear two wheels.

  This formula is a calculation method for calculating the driving force distribution ratio r based on the formula 10 described above. The equation shows only the nonlinear term of the element a11, but the same replacement can be made for the nonlinear terms related to the other elements a12 to a22. However, since the basic concept is the same, the description thereof is omitted, but the nonlinear terms of the individual elements a11 to a22 are the front wheel maximum value Ff_Lim, the rear wheel maximum value Fr_Lim, and the average lateral force Ff_y of the front and rear wheels. , Fr_y and the parameter groups μ, Fa, Ff_z, Fr_z described above. According to this method, the same effects as those of the first embodiment described above can be obtained, and the occurrence of a situation where there is no solution of the driving force distribution ratio r can be suppressed, so that reliability in control is improved. be able to.

(Third embodiment)
In the second embodiment described above, Formula 11 is developed by using the average lateral forces Ff_y and Fr_y of the front and rear wheels by steering as inputs. However, when it is considered that disturbances such as crosswinds and road surface irregularities are applied as acting force (especially average lateral force Fz), the nonlinear terms in elements a11 to a22 are minimized in accordance with the above-described concept even during straight running. Can be realized. However, unlike the case of steering, it is difficult to predict the average lateral forces Ff_y and Fr_y during straight traveling, and therefore it is assumed that the average lateral forces Ff_y and Fr_y of the front and rear wheels are equivalent. In this case, the nonlinear term of the element a11 can be expressed by the following equation.

  The equation shows only the nonlinear term of the element a11, but the same replacement can be made for the nonlinear terms related to the other elements a12 to a22. However, since the basic concept is the same, description thereof is omitted, but the nonlinear terms of the individual elements a11 to a22 are the front wheel maximum value Ff_Lim, the rear wheel maximum value Fr_Lim, and the parameter groups μ, Fa, It can be calculated based on Ff_z and Fr_z.

(Fourth embodiment)
FIG. 10 is an explanatory diagram of a state surface. For example, an analysis technique called a state plane is effective for a nonlinear control system in which the control input changes discontinuously, such as a vehicle motion state influenced by the nonlinearity of the wheels. In a vehicle in which a motion state is represented by a two-wheel model-based state equation, the state surface has a vector field S (β ′ (βb, γ), γ ′ (βb, γ)). In the vector field S, it is considered that the state surface (vehicle behavior) tends to diverge as the vector goes outward, and the state surface (vehicle behavior) tends to non-diverge as the vector goes inward. . When the state surface has a non-divergence tendency, even if a disturbance is input, the divergence of the vehicle behavior is suppressed, which means that a stable vehicle response is obtained.

  In such a vector field, the vector tendency can be quantitatively evaluated by a characteristic value called divergence ▽ S. This divergence ▽ S is the sum of a value obtained by partial differentiation of the vehicle slip angular velocity βb 'with respect to the vehicle slip angle βb and a value obtained by partial differentiation of the yaw angular acceleration γ' with respect to the yaw rate γ. When the divergence ▽ S is a positive value, the vector tends to face outward, and when the divergence ▽ S is a negative value, the vector tends to face inward. Therefore, by minimizing the divergence ▽ S of the vector field S, the vector tendency becomes inward, and thus it is possible to suppress the divergence of the vehicle behavior.

In this embodiment, based on such a concept, the driving force distribution ratio r is controlled so as to minimize the divergence ＳS of the vector field S describing the state surface. In this embodiment, the term “minimization of divergence ▽ S” is not only used to mean the smallest value in the range that this value ▽ S can take, but at least more than the current value. Also used to mean smaller. When the divergence ＳS is calculated based on the equation of state shown in Equations 1 and 2 described above, the divergence ＳS satisfies the relationship shown below.

As shown in the equation, the divergence ＳS of the vector field S is represented by the sum of a linear term and a nonlinear term, as in the first embodiment. Further, as shown in the equation, this divergence ▽ S is equal to the negative component of the sum of the diagonal elements a11 and a22 of the system matrix. Since the linear term in the sum of the diagonal elements a11 and a22 is always positive, in order to minimize the divergence ▽ S, the nonlinear term having a minus sign may be minimized on the basis of the absolute value (nonlinear term). Minimization). The driving force distribution ratio r for minimizing the nonlinear term is calculated based on the following equation in which Equation 9 is substituted for the nonlinear term shown in Equation 14.

  Specifically, the driving force distribution ratio r is changed stepwise between “0.00” and “1.00”, and the friction coefficient μ, the total driving force Fa, the average vertical forces Ff_z and Fr_z of the front and rear wheels, and the average front and rear wheels. Based on the slip angles βf and βr, the value of the nonlinear term is sequentially calculated. The driving force distribution ratio r having the smallest value among the calculated nonlinear terms is set as the target value r *. Note that when performing such sequential calculation, it is difficult to accurately measure the average front and rear wheel slip angles βf and βr. Therefore, as described in the first embodiment, these slip angles βf and βr may be replaced with average lateral forces Ff_y and Fr_y of front and rear wheels that can be detected, respectively.

  As described above, according to the present embodiment, the target value r * of the driving force distribution ratio r is set so that the divergence ▽ S of the vector field S representing the state surface is minimized from the viewpoint of the state surface of the vehicle motion. Is done. Therefore, the vehicle response (spin etc.) resulting from the non-linearity of the wheels is suppressed, and the operability can be improved.

  In this embodiment, a nonlinear term in the divergence ＳS is calculated, and the driving force distribution ratio r is set based on the calculated nonlinear term so that the nonlinear term becomes smaller than the current value. However, minimization of this nonlinear term can be equated with minimization of divergence 散 S. Therefore, the above-described series of control procedures can be rephrased as procedures shown below. As a first step, the divergence ▽ S is calculated in the calculation unit 10a. As a second step, the driving force distribution ratio r is set in the setting unit 10b based on the calculated divergence ＳS so that the divergence ＳS becomes smaller than the current value. When the divergence ▽ S is “0” or less, the state surface tends to be non-divergent. Therefore, the divergence ▽ It is also possible to set the driving force distribution ratio r at which S is “0” or less as the target value r *. That is, when the constraint condition that the current value ▽ S is negative is satisfied, the driving force distribution ratio r is set so that the value becomes large within a range where the divergence ▽ S can be "0" or less. It may be r *.

  Further, in each embodiment, the state quantities of the vehicle slip angle βb and the yaw rate γ are used for the vector field axis representing the state plane in relation to the state equation based on the two-wheel model. However, as long as the state quantity representing the motion state of the vehicle such as the lateral force Fz, the yaw moment, the vehicle slip angle βb, and the vehicle slip angular velocity βb ′ is used, parameters that can be taken as the axes can be arbitrarily set.

(Fifth embodiment)
A vehicle in which a motion state is expressed by a state equation based on a two-wheel model can be considered by considering it as a vibration system with one degree of freedom that is a second-order lag system. This vibration system can be modeled by free vibration in consideration of damping, and its characteristics are expressed by two state quantities of an angular natural frequency ωn and a damping ratio ζ. Referring to the state equation indicating the motion state of the one-degree-of-freedom vibration system, the angular natural frequency ωn and the damping ratio ζ are expressed by the following equations.

Here, m is mass, c is a damping coefficient, and k is a spring constant. The angular natural frequency ωn is the square root of a division value obtained by dividing the spring constant k by the mass m, and the damping ratio ζ is a value obtained by doubling the square root of the integrated value of the mass m and the spring constant k. Divide value divided by. The integrated value obtained by integrating the angular natural frequency ωn and the damping ratio ζ is a characteristic value indicating the convergence of the vibration system. In this embodiment, attention is paid to this integrated value (hereinafter, this value is simply referred to as “damping”). Called). When equating such a vibration system and the motion of the vehicle (vehicle system), these values ωn, ζ, ωn · ζ are expressed by the following formula based on the state equation (Formula 1) representing the motion state of the vehicle. Calculated.

  The angular natural frequency ωn of the vehicle system is the square root of the subtraction value obtained by subtracting the product of the coupled elements a12 and a21 from the product of the diagonal elements a11 and a22 of the system matrix, and the damping ratio ζ of the vehicle system is This is a divided value obtained by dividing the sum of the diagonal elements a11 and a22 by a value obtained by doubling the natural angular frequency ωn. The vehicle system attenuation ωn · ζ is a value obtained by halving the sum of the diagonal elements a11 and a22, that is, an average value of the diagonal elements a11 and a22. This damping ωn · ζ means that the larger the value, the better the convergence of the vibration system with respect to the disturbance. In other words, the greater the attenuation ωn · ζ, the more the vehicle stability can be improved.

  In the present embodiment, based on such a concept, the driving force distribution ratio r is controlled so as to maximize the attenuation ωn · ζ. In the present embodiment, the term “maximization of attenuation ωn · ζ” is not only used to mean the largest value in the range that this value ωn · ζ can take, but at least the current It is also used in the sense that it is larger than the value (that is, it suppresses the decrease in attenuation ωn · ζ). Since the linear term in the sum of the diagonal elements a11 and a22 is always positive, in order to maximize the attenuation ωn · ζ, the nonlinear term having a minus sign may be minimized on an absolute value basis (nonlinearity). Term minimization). The driving force distribution ratio r for minimizing the nonlinear term can be determined based on the above-described equation 15 as in the case of minimizing the divergence ▽ S.

  As described above, according to the present embodiment, the motion state of the vehicle is regarded as a one-degree-of-freedom vibration system that is a second-order lag system, and from the viewpoint of improving the convergence of the system, the damping ωn · ζ is maximized. The driving force distribution ratio r is determined. As a result, the stability of the vehicle can be improved, so that the operability can be improved.

  In this embodiment, a nonlinear term is calculated, and the driving force distribution ratio r is set based on the calculated nonlinear term so that the nonlinear term becomes smaller than the current value. However, minimization of this nonlinear term can be regarded as maximization of attenuation ωn · ζ. Therefore, the above-described series of control procedures can be rephrased as procedures shown below. As a first step, attenuation ωn · ζ is calculated in the calculation unit 10a. As a second step, in the setting unit 10b, the driving force distribution ratio r is set based on the calculated attenuation ωn · ζ so that the attenuation ωn · ζ becomes smaller than the current value.

(Sixth embodiment)
In the first to third embodiments described above, the nonlinear term is minimized by focusing on at least one of the elements a11 to a22. However, focusing on two or more elements a11 to a22. The nonlinear term may be minimized. For example, focusing on two elements, element a11 and element a22, the nonlinear term that is the sum of these elements a11 and a22 (diagonal elements) is minimized. Minimization of the nonlinear term in the sum of the elements a11 and a22 can be realized by appropriately determining the driving force distribution ratio r based on the above-described Expression 15. In this way, by focusing on the plurality of elements a11 to a22 as well as the single elements a11 to a22, for example, the effects shown in the fourth or fifth embodiment can be achieved.

(Seventh embodiment)
In the first to sixth embodiments described above, the method for controlling the driving force distribution ratio r between the front and rear wheels has been described. However, the present invention is not limited to this, and the driving force distribution ratio for each wheel (four wheels in this embodiment) including the left and right wheels may be controlled. Hereinafter, the driving force distribution control for each wheel will be described based on the control concept according to the fourth embodiment, but the basic concept is the same for the other embodiments, and thus the description thereof is omitted here. To do. Hereinafter, for convenience of explanation, each state quantity is indicated using alphabets of “fl” representing the left front wheel, “fr” representing the right front wheel, “rl” representing the left rear wheel, and “rr” representing the right rear wheel. Explicitly distinguish the wheels. For example, kfl_a and kfr_a are the cornering powers ka of the left and right front wheels, and krl_a and krr_a are the cornering powers ka of the left and right rear wheels (the same applies to the other state quantities).

The front wheel cornering power kf_a is the average value of the left and right front wheel cornering powers kfl_a and kfr_a, and the rear wheel cornering power kr_a is the average value of the left and right rear wheel cornering powers krl_a and krr_a. The cornering powers kfl_a to krr_a of each wheel can be calculated based on Equation 3 which is a calculation formula for the cornering power ka per wheel. When the divergence ＳS (Equation 14) is calculated in consideration of individual wheels, this value ＳS becomes the following equation.

In the equation, Ffl_ymax to Frr_ymax are the lateral force maximum values Fymax of the left and right front and rear wheels. Here, the driving force distribution ratio for the front and rear wheels (hereinafter referred to as “front / rear distribution ratio”) is Rfr, the driving force distribution ratio for the left and right wheels of the front wheel (hereinafter referred to as “front left / right distribution ratio”) is Rlr_f, and the rear wheel is determined for the left and right wheels. The driving force distribution ratio (hereinafter referred to as “rear left / right distribution ratio”) is Rlr_r. The relationship shown below is established between the driving force distribution ratios Rfr to Rlr_r and the longitudinal forces Ffl_x to Frr_x of each wheel via the total driving force Fa.

In order to minimize the divergence ＳS, a nonlinear term having a minus sign may be minimized on an absolute value basis (nonlinear term minimization). In Formula 18, each lateral force maximum value Ffl_ymax to Frr_ymax can be calculated based on Formula 6, which is a formula for calculating the lateral force maximum value Fymax per wheel. When the nonlinear terms in Equation 18 are arranged based on Equation 19, the distribution ratios Rfr, Rlr_f, Rlr_r that minimize the nonlinear terms are calculated based on the following equations.

Specifically, the distribution ratios Rfr, Rlr_f, and Rlr_r are individually changed stepwise from “0.00” to “1.00”, and the friction coefficient μ, the total driving force Fa, and the vertical force Ffl_z of each wheel. Based on ~ Frr_z and average front and rear wheel slip angles βf and βr, the values of the nonlinear terms are sequentially calculated. The combination of the distribution ratios Rfr, Rlr_f, Rlr_r having the smallest value among the calculated nonlinear terms is set as the target values Rfr *, Rlr_f *, Rlr_r * of the distribution ratios Rfr, Rlr_f, Rlr_r. The In the equation, the average front wheel slip angle βf and the average rear wheel slip angle βr may be replaced by the lateral forces Ffl_y to Frr_y of the respective wheels. In this case, the nonlinear term of the divergence ▽ S is as follows.

  Thus, according to the present embodiment, the driving force distribution ratios Rfr *, Rlr_f *, and Rlr_r * to each wheel are determined so that the divergence ▽ S of the vector field representing the state plane of the vehicle motion is minimized. . By appropriately allocating the driving force transmitted to each wheel, it is possible to realize a more stable vehicle response and improve the maneuverability. By setting the front left / right distribution ratio Rlr_f and the rear left / right distribution ratio Rlr_r to “0.50” as fixed values, it can be regarded as a simple driving force distribution for the front and rear wheels. For a method relating to the distribution of driving force to the left and right wheels, refer to, for example, if disclosed in Japanese Patent Laid-Open No. 2000-168385.

(Eighth embodiment)
In each of the above-described embodiments, the nonlinear term is minimized by adjusting the driving force distribution ratio for each wheel. However, since the change of the nonlinear term is also caused by the load distribution ratio r ′ for each wheel, the nonlinear term may be minimized by this load distribution control. The basic control concept is the same as the driving force distribution control described above, and in the present embodiment, the difference will be described.

The total load, which is the sum of the loads on each wheel, is W, and the load distribution ratio for the front and rear wheels is r ′. In this case, the average vertical force Ff_z of the front wheels and the average vertical force Fr_x of the rear wheels can be expressed by the following mathematical expressions.

The load distribution ratio r ′ that minimizes the nonlinear term in each element a11 to a22 is unambiguous on the basis of Equations 6, 7, and 22 by changing Equation 9 described in the first embodiment to Equation 22. Can be specified. For example, the nonlinear term of the element a11 is expressed by the following expression.

  Here, the total load W can be uniquely specified as the sum of the vertical forces (detected values) Fz of the respective wheels. The load distribution ratio r ′ that minimizes the nonlinear term can be uniquely calculated based on the average front and rear wheel slip angles βf and βr, the friction coefficient μ, the total load W, and the front and rear wheel average front and rear forces Ff_x and Fr_x. .

Further, such load distribution control is not only applicable to the method of the first embodiment, but is also applicable to other embodiments as well. For example, in the fourth to sixth embodiments, the load distribution ratio r ′ for minimizing the nonlinear term that is the sum of the elements a11 and a22 is based on the following equation in which the equation 22 is substituted into the nonlinear term shown in the equation 14. Calculated.

  Thus, according to the present embodiment, the same effects as those of the above-described embodiments can be obtained, and the control method can be diverse. Thereby, this control can be applied to a wide variety of vehicles. The load distribution to each wheel can be dynamically changed by using an electronically controlled suspension device, for example.

  In each of the above-described embodiments, the four-wheel drive vehicle using the engine 2 as a drive source has been described. However, a motor may be used as the drive source. In this case, as with the engine 2, the four wheels may be driven by a single motor via a differential device, or a motor may be provided for each wheel or front and rear wheels. Even with this method, paying attention to the nonlinear terms of the elements of the system matrix, the differential of the differential device is controlled, or the output of the motor is directly controlled, so that an appropriate driving force distribution ratio r is obtained. Achieve. Thereby, there can exist an effect similar to each embodiment mentioned above. Further, in order to perform control during braking, output control of the engine 2 and an anti-lock brake system can be used according to the driving force distribution ratio r.

  Furthermore, in each embodiment mentioned above, although the detection part 11 is a structure which detects the acting force which acts on three directions, this invention is not limited to this, It acts on the required component force direction. It is sufficient if the acting force can be detected. Further, it may be a six-component force meter that detects not only three-component components but also six-component forces including moments around these three directions. Even in such a configuration, there is no problem as a matter of course, since at least the required acting force can be detected. In addition, about the method of detecting the six component force which acts on the wheel 6, since it is disclosed by Unexamined-Japanese-Patent No. 2002-039744 and Unexamined-Japanese-Patent No. 2002-022579, please refer if necessary.

  Moreover, although the case where the detection part 11 was embed | buried under the axle shaft 5 was demonstrated, this invention is not limited to this, Other variations are also considered. In terms of detecting the acting force, for example, the detection unit 11 may be provided on a member that holds the wheel 6, for example, a hub or a hub carrier. In addition, since the method of providing the detection unit 11 in the hub is disclosed in Japanese Patent Application Laid-Open No. 2003-104139, refer to it if necessary.

  In each of the embodiments described above, the longitudinal force Fx, the lateral force Fy, and the vertical force Fz are detected by a sensor that directly detects an acting force acting in three directions, but the present invention is not limited to this. For example, the lateral force Fy may be specified by estimating the cornering force (ie, indirectly detected), and the vertical force Fz may be specified by estimating the vertical load (ie, Indirectly detected). In this estimation method, a wheel speed sensor for detecting the wheel speed of each wheel, a lateral acceleration sensor and a longitudinal acceleration sensor provided at the center of gravity of the vehicle, and a yaw rate sensor are used.

  First, a cornering force estimation method will be described. In the detection values obtained from various sensors, the wheel speeds of the wheels are Vfl_s, Vfr_s, Vrl_s, Vrr_s, yaw rate γ, lateral acceleration y ″ s, and longitudinal acceleration x ″ s. The center of gravity height is h, the tread width is d, and the gravitational acceleration is g.

These detection values are subjected to digital filter processing for the purpose of removing disturbance components caused by road surface cant and body roll in addition to low frequency noise and high frequency noise. Therefore, the output values (wheel speed) of the filtered wheel speed sensor are Vfl_f, Vfr_f, Vrl_f, and Vrr_f, and the vehicle body speed estimated based on these values (for example, the average value of the respective wheel speeds) is V. ^{^,} the vehicle acceleration obtained by differentiating calculating the vehicle speed V ^{^} and x '' ^{^.} The output value (yaw rate) of the filtered yaw rate sensor is γf, and the yaw angular acceleration γ ′ ^{^} obtained by differentiating the yaw rate γf is assumed. Further, the output value (lateral acceleration) of the filtered lateral acceleration sensor is y ″ ^{^} .

When the cornering force of the left and right front wheels is Ffl ^{^} , Ffr ^{^} , and the cornering force of the left and right rear wheels is Frl ^{^} , Frr ^{^} , each value can be estimated using the following equation based on the equation of motion.

Therefore, the estimated value Ff ^{^} of the average cornering force of the front wheels and the estimated value Fr ^{^} of the average cornering force of the rear wheels can be uniquely calculated based on the following equations.

When the cornering force acting on the vehicle body is Fc, the following relationship is established between the cornering force Fc and the lateral force Fy acting on the wheel 6.

Second, a method for estimating the vertical load will be described. When the estimated vertical load values of the left and right front wheels are Wfl ^{^} and Wfr ^{^} and the estimated vertical load values of the left and right rear wheels are Wrl ^{^} and Wrr ^{^} , these estimated values are the acceleration motion in the vertical direction, roll axis and pitch. If the acceleration motion around the axis is ignored, it can be calculated based on the following equation.

Therefore, the estimated value Ff ^{^} of the average cornering force of the front wheels and the estimated value Fr ^{^} of the average cornering force of the rear wheels can be uniquely calculated based on the following equations.

In this equation, ΔWlong is a load movement amount due to longitudinal acceleration, and ΔWlat is a load movement amount due to lateral acceleration, and each value can be uniquely calculated based on the following equation.

The longitudinal acceleration ax is the output value x ″ s of the longitudinal acceleration sensor or the vehicle body acceleration x ″ ^{^} calculated from the wheel speed, and the lateral acceleration ay is the output value y ″ ^{^} of the lateral acceleration sensor. In the above calculation, the acceleration motion in the vertical direction on the spring and the rotational motion about the roll axis and the pitch axis are ignored, but the cornering force and the vertical load are estimated in consideration of these. Is also possible.

Explanatory drawing showing the vehicle model Block diagram showing the state equation of vehicle motion Explanatory drawing of acting force acting on wheel The figure which shows the tendency of the driving force distribution ratio which minimizes the nonlinear term of element a11 The figure which shows the tendency of driving force distribution ratio which minimizes the coefficient of the nonlinear term of element a12 and a21 The figure which shows the tendency of the driving force distribution ratio which minimizes the coefficient of the nonlinear term of element a22 The block diagram which showed the whole structure of the vehicle motion control apparatus concerning this embodiment Block diagram of vehicle motion control device A flowchart showing a procedure of vehicle control according to the present embodiment. State drawing

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Vehicle motion control apparatus 2 Engine 3 Automatic transmission 4 Center differential apparatus 4a Carrier 4b Sun gear 4c Hydraulic multi-plate clutch 5 Drive shaft 6 Wheel 6f Front wheel 6r Rear wheel 7 Solenoid valve 10 Control part 10a Calculation part 10b Setting part 11 Detection part 12 Estimator

Claims (19)

In the vehicle motion control device,
Based on the system matrix of the state equation representing the motion state of the vehicle, the slip angle of the front wheel and the slip angle of the rear wheel included in the polynomial representing at least one element of the plurality of elements constituting the system matrix are calculated. A calculation unit for calculating a nonlinear term as a variable;
A vehicle motion control device comprising: a setting unit configured to set a target value of a driving force distribution ratio or a target value of a load distribution ratio for each of the wheels based on the calculated nonlinear term. A detection unit for detecting a vertical force acting on each of the wheels;
Based on the state of the vehicle, the friction coefficient between the wheel and the road surface is estimated, and further includes an estimation unit that estimates the total driving force applied to each of the wheels as a total driving force,
The calculation unit calculates a plurality of non-linear terms corresponding to different driving force distribution ratios based on a group of parameters including the vertical force of each wheel, the friction coefficient, and the total driving force. And
The setting unit is configured to set, as the target value, the driving force distribution ratio in which an absolute value of the nonlinear term becomes smaller than a current value based on the plurality of calculated nonlinear terms. Item 4. The vehicle motion control device according to Item 1.   3. The vehicle motion control device according to claim 2, wherein the setting unit sets the driving force distribution ratio that minimizes the absolute value of the nonlinear term as the target value. 4. A detection unit for detecting a vertical force and a longitudinal force acting on each of the wheels;
An estimation unit that estimates a friction coefficient between the wheel and the road surface based on the state of the vehicle;
The calculation unit includes a plurality of nonlinear terms corresponding to the different load distribution ratios based on a parameter group including each vertical force of the wheel, each longitudinal force of the wheel, and the friction coefficient. To calculate
The setting unit sets the load distribution ratio in which an absolute value of the nonlinear term is smaller than a current value as the target value based on the plurality of calculated nonlinear terms. A vehicle motion control device according to claim 1.   5. The vehicle motion control device according to claim 4, wherein the setting unit sets, as the target value, the load distribution ratio at which an absolute value of the nonlinear term is minimized. The estimation unit further estimates the slip angle of the front wheel and the slip angle of the rear wheel based on the slip angle of the vehicle,
5. The vehicle motion according to claim 2, wherein the calculation unit calculates the nonlinear term based on a slip angle of the front wheel, a slip angle of the rear wheel, and the parameter group. Control device. The detection unit further detects a lateral force acting on each of the wheels,
5. The vehicle motion control device according to claim 2, wherein the calculation unit calculates the nonlinear term based on a lateral force of each of the wheels and the parameter group.   The calculation unit calculates the maximum value of the driving force applied to the wheel based on the vertical force of the wheel and the friction coefficient for each of the wheels, and determines the driving force of each of the wheels. The vehicle motion control device according to claim 7, wherein the nonlinear term is calculated based on a maximum value, a lateral force of each of the wheels, and the parameter group.   The calculation unit calculates the maximum value of the driving force applied to the wheel based on the vertical force of the wheel and the friction coefficient for each of the wheels, and determines the driving force of each of the wheels. 5. The vehicle motion control device according to claim 2, wherein the nonlinear term is calculated based on a maximum value and the parameter group.   The vehicle motion control device according to claim 1, wherein the calculation unit calculates the nonlinear term in a polynomial that is a sum of diagonal elements in the system matrix. In the vehicle motion control device,
Based on a state equation representing the motion state of the vehicle, the divergence as a characteristic value indicating the tendency of the vector in the vector field in a state surface represented by a vector field having a state quantity representing the motion state of the vehicle as an axis A calculation unit for calculating,
Based on the calculated divergence, the target value of the driving force distribution ratio or the load distribution ratio of each of the wheels is set so that the divergence is smaller than a current value or the divergence is 0 or less. A vehicle motion control device comprising: a setting unit for setting a target value. In the vehicle motion control device,
A calculation unit that calculates attenuation, which is a characteristic value indicating convergence of a vehicle as a vibration system, based on a state equation representing a motion state of the vehicle;
A setting unit configured to set a target value of a driving force distribution ratio or a target value of a load distribution ratio for each of the wheels based on the calculated attenuation so that the attenuation is larger than a current value. A vehicle motion control device. A detection unit for detecting a vertical force acting on each of the wheels;
Based on the state of the vehicle, the friction coefficient between the wheel and the road surface is estimated, and further includes an estimation unit that estimates the total driving force applied to each of the wheels as a total driving force,
The calculation unit is a system of the state equation for each of a plurality of different driving force distribution ratios based on a parameter group including each vertical force of the wheel, the friction coefficient, and the total driving force. Calculate a nonlinear term in the polynomial that is the sum of the diagonal elements that make up the matrix, with the slip angle of the front wheel and the slip angle of the rear wheel as variables,
The vehicle motion control device according to claim 11 or 12, wherein the setting unit sets a target value of the driving force distribution ratio based on each of the calculated nonlinear terms. A detection unit for detecting a vertical force and a longitudinal force acting on each of the wheels;
An estimation unit that estimates a friction coefficient between the wheel and the road surface based on the state of the vehicle;
The calculation unit is configured to calculate the state equation for each of a plurality of different load distribution ratios based on a parameter group including each vertical force of the wheel, each longitudinal force of the wheel, and the friction coefficient. A non-linear term that includes the slip angle of the front wheel and the slip angle of the rear wheel included in the polynomial that is the sum of the diagonal elements constituting the system matrix of
The vehicle motion control device according to claim 11 or 12, wherein the setting unit sets a target value of the load distribution ratio based on each of the calculated nonlinear terms. In the vehicle motion control method,
Based on the system matrix of the state equation representing the motion state of the vehicle, the slip angle of the front wheel and the slip angle of the rear wheel included in the polynomial representing at least one element of the plurality of elements constituting the system matrix are calculated. A first step of calculating a nonlinear term as a variable;
And a second step of setting a target value of a driving force distribution ratio or a target value of a load distribution ratio for each of the wheels based on the calculated nonlinear term. A third step of detecting a vertical force acting on each of the wheels;
A fourth step of estimating a friction coefficient between the wheel and the road surface based on a state of the vehicle, and estimating a sum of driving forces applied to each of the wheels as a total driving force; ,
The first step includes a plurality of nonlinear terms corresponding to different driving force distribution ratios based on a group of parameters including the vertical force of each wheel, the friction coefficient, and the total driving force. Is the step of calculating
The second step is a step of setting, as the target value, the driving force distribution ratio in which an absolute value of the nonlinear term becomes smaller than a current value based on the plurality of calculated nonlinear terms. The vehicle motion control method according to claim 15. A third step of detecting a vertical force and a longitudinal force acting on each of the wheels;
A fourth step of estimating a friction coefficient between the wheel and the road surface based on the state of the vehicle;
The first step is based on a group of parameters including each vertical force of the wheel, each longitudinal force of the wheel, and the friction coefficient. Calculating a nonlinear term,
The second step is a step of setting, as the target value, the load distribution ratio in which the absolute value of the nonlinear term becomes smaller than a current value based on the calculated nonlinear terms. The vehicle motion control apparatus according to claim 15, wherein the apparatus is a vehicle motion control apparatus. In the vehicle motion control method,
Based on a state equation representing the motion state of the vehicle, the divergence as a characteristic value indicating the tendency of the vector in the vector field in a state surface represented by a vector field having a state quantity representing the motion state of the vehicle as an axis A first step of calculating;
Based on the calculated divergence, the target value of the driving force distribution ratio or the load distribution ratio of each of the wheels is set so that the divergence is smaller than a current value or the divergence is 0 or less. And a second step of setting a target value. In the vehicle motion control method,
A first step of calculating attenuation, which is a characteristic value indicating convergence of a vehicle as a vibration system, based on a state equation representing a motion state of the vehicle;
A second step of setting a target value of a driving force distribution ratio or a target value of a load distribution ratio for each of the wheels based on the calculated attenuation so that the attenuation is larger than a current value. A vehicle motion control method.

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