JPH08230504A - Drive force regulator for vehicle - Google Patents

Abstract

PURPOSE: To improve a turning characteristic according to the yawing rate of a vehicle and ensure the stability of control by providing a speed changing means to slide a movable sheave on the basis of an increase or decrease amount of change gear ratio from the change gear ratio determined by a change gear ratio determining means. CONSTITUTION: Feed-back control is performed to bring an actual yawing rate γr detected by a yawing rate sensor 41 close to a theoretical yawing rate γo obtained from a steering angle δf detected by a steering angle sensor 40 and car body speed V (Vf R/Vf L). Since movable sheaves 23, 24 of first and second pulleys are slid on the basis of a change gear ratio increase amount Δρ obtained from the feed-back control, moving torque ΔT is regulated between left and right wheels according to the change gear ratio increase or decrease amount Δρand a vehicle is turned with a satisfactory responce property according to the moving torque to improve turning characteristics.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicular drive force adjusting device, and more particularly to a vehicular drive force adjusting device capable of controlling a yaw moment around a center of gravity by controlling a drive force difference between right and left wheels of a vehicle. .

[0002]

2. Description of the Related Art Improvements in drive system elements have been made for the purpose of improving the dynamic performance of vehicles. The drive system requires a mechanism for appropriately distributing the driving force of each wheel in accordance with the driver's will and running condition, which change from moment to moment, and a control device thereof has also been proposed. According to this type of device, 4
By optimizing the slip ratio and the sideslip angle of the wheel tires, the tire grip force can be effectively used in a well-balanced manner, and the vehicle dynamic performance can be improved. In particular, a drive system including a left / right torque distribution mechanism capable of positively operating the torque difference between the left and right wheels is known, and one example thereof is shown in FIGS.
It was shown to.

In the left / right torque distribution mechanism shown in FIG. 26, the input torque Ti from the engine is divided by the differential, and the divided torque is transmitted to the left / right driving wheels Wh, respectively.
The left and right axles S _L and S _R are connected to each other via a slip clutch Cs whose slip ratio is controlled by a controller CU. In this case, of the left and right wheels S _L , S _R , the torque is transmitted to the slower wheel than the faster wheel, so that the torque is bypassed.
Wheel slip is regulated. Controlled LS like this
An example of a drive system provided with D (Limited Slip Differential) is disclosed in Japanese Patent Application Laid-Open No. 2-415555.

In the left / right torque distribution mechanism shown in FIG. 27, the input torque Ti from the engine is divided by a differential, and the divided torque is transmitted to the left / right drive wheels Wh, respectively.
The left and right axles S _L and S _R are respectively provided with a torque adjusting brake Bd whose braking force is controlled by the controller CU. In this case, of the left and right wheels, an arbitrary torque difference can be generated in the left and right wheels regardless of the traveling conditions. However, since it is accompanied by deceleration due to the braking force, it can be said that control during deceleration or limit travel is realistic.

In the left / right torque distribution mechanism shown in FIG. 28, the central axle Sc is driven by the input torque Ti from the engine,
End axles S _L and S _R are connected to both ends of the central axle Sc via slip clutches Cs whose slip ratio is controlled by a controller CU. In this case, an arbitrary torque difference can be generated regardless of the traveling conditions. However,
Under the condition that the acceleration / deceleration is not affected, the torque difference that can be generated is limited by the input torque. An example of a drive system for independently driving the left and right wheels in this way is disclosed in Japanese Patent Application Laid-Open No. 3-254703.

By the way, as shown in FIG. 29, the yaw moment M when the vehicle is turning, which can be controlled arbitrarily, is expressed by the equation (1).

Here, the wheel width is tr, the left and right wheels (driving wheels)
The divided torque was defined as T, the torque difference between the left and right sides was ΔT, and the tire radius was r. M = 1/2 × tr × ΔT / r (1) Next, FIG. 3 shows a two-wheel model that makes a steady circular turn with a turning radius R.
0 to 31. This model is a two-degree-of-freedom model capable of lateral displacement of the center of gravity and change of the yaw angle. As constraints, the traveling direction speed is constant and the tire characteristics are in a linear range (lateral slip angle β-turning force CF shown in FIG. 33. In the characteristics, it is assumed to be in the e1 region in which the turning coefficient C _f 1 is held), and the influence of the driving / braking force and the load movement are not considered.

Here, it is assumed that the yaw moment M as expressed by the equation (1) depending on the torque difference between the left and right wheels can be arbitrarily controlled. Further, the balance equation of the force acting on the center of gravity at the time of steady circle turning is shown in equation (2), and the balance equation of moment is shown in equation (3). Where C _f is the turning coefficient of the front wheels, C _r is the turning coefficient of the rear wheels, β of the vehicle is the sideslip angle, β _f is the sideslip angle of the front wheels, β _r is the sideslip angle of the rear wheels, γ is the yaw rate, and m is The vehicle mass and L indicate the wheel bases, respectively.

MV (γ + dβ / dt) = C _f × β _f + C _r × β _r ... (2) L × dγ / dt = C _f × β _f × L _f −C _r × β _r × L _r + M (3) When making a steady circular turn at the vehicle speed V, the front wheel steering angle δ _f is expressed by the formula (4) from the turning radius R and the side slip angles β _f and β _{r of the} front and rear wheels. . δ _f = β _f −β _r + L / R (4) Here, when formulas (2) to (4) are introduced by introducing the stability factor K shown in formula (5), Equation (6) is obtained.

Here, the stability factor K is used for evaluation of steering characteristics, and is a value determined by the configuration of a vehicle suspension system, tire characteristics, road surface conditions, and the like. _{K = - (C f × L} f -C r × L r) / (C f × C r × L) × (m / L) ······· (5) δ f = M · (C f + C _r ) / (C _f · C _r · L) + (1 + KV ² ) · L / R ···· (6) Here, in the extremely low speed traveling state geometrically required during steady circle turning. The steering angle ratio δ _f / δ _f0, which is the ratio of the front wheel steering angle δ _f during acceleration to the front wheel steering angle δ _f0 (= L / R), is expressed by equation (7).

Δ _f / δ _f0 = (1 + KV ² ) + M (C _f + C _r ) R / C _f · C _r · (L _f + L _r ) ² ··· (7) This steering angle ratio δ The relationship between _f / δ _f0 and the square value (V ² ) of the vehicle speed can be expressed, for example, as a diagram shown in FIG. 32, which serves as an evaluation index of turning characteristics. Here, the inclination rate of the same line diagram is expressed as the stability factor K, the hatched area E becomes the controllable area, and it is possible to arbitrarily set the steering characteristics when the vehicle makes a steady circular turn by the yaw moment control. Recognize. Considering the above points, as the next step, it is desired to further improve the motion performance during turning of the vehicle. In this case, the function of controlling the yaw moment around the center of gravity G during turning of the vehicle is required. To be done.

However, in the conventional left / right torque distribution mechanism shown in FIG. 26, the torque of the turning inner wheel can be increased, but the torque of the outer wheel cannot be increased, the yaw moment cannot be controlled, and the clutch is disengaged. Since it is used by slip coupling, there is energy loss, which is a problem. FIG.
Since the left-right torque distribution mechanism shown in (1) uses a brake, energy loss is large and the yaw moment cannot be controlled, which is a problem. Since the right and left torque distribution mechanism shown in FIG. 28 also frequently uses the clutch in slip coupling, energy loss is large and the yaw moment cannot be controlled, which is a problem.

The left and right torque distribution mechanism by the present applicant is disclosed in the specification and drawings of Japanese Patent Application No. 4-64233. In this case, the controller controls the slip ratio of the slip clutch (wet multi-plate clutch) to increase the slip ratio of the slip clutch on the inside of the turn and decrease the slip ratio of the slip clutch on the outside of the turn (direct coupling when the vehicle turns). ), A large input torque from the engine is transmitted to the drive wheels on the outer side of the turn, and a large yaw moment in the direction in which the vehicle body is about to turn is given, so that the turn can be performed smoothly. In this case, an arbitrary torque difference can be generated regardless of traveling conditions. However, under the condition that does not affect the acceleration / deceleration, the torque difference that can be generated is limited by the input torque. Moreover, all the input torque of the engine is transmitted to the wheels via the left and right slip clutches, and the durability of the slip clutches and the like is likely to occur.

Therefore, there is a demand for a mechanism capable of improving the turning characteristic by controlling the yaw moment, and having a small energy loss particularly when the moving torque is transmitted from one wheel to the other wheel and having a durability. Among these, an example of a durable mechanism with less energy loss when transmitting moving torque from one wheel to the other wheel is disclosed in the specification and drawings of Japanese Patent Application No. 1-324257 filed by the present applicant. To be done.

Here, a left and right torque distribution mechanism is disclosed in which the left and right axles are arranged so as to sandwich the diff, a continuously variable transmission is arranged at a position bypassing the diff, and the left and right axles are connected by the continuously variable transmission. Here, the first pulley of the right axle and the second pulley of the left axle are connected by an endless belt, the movable sheaves of both pulleys are slid by the speed changing means to change the winding radius of the endless belt, and between the left and right axles. Disclosed is a configuration in which the transmission gear ratio is variable. In this case, the input torque is split into two by the differential, the split torque is transmitted to the left and right wheels, and then the gear ratio of the continuously variable transmission is increased / decreased as necessary, and only the torque difference between the left and right axles is used as the moving torque. It is possible to obtain a durable left and right torque distribution mechanism that can be transmitted to the opposite axle, has little energy loss of the continuously variable transmission.

[0016]

However, in the left-right torque distribution mechanism using such a conventional continuously variable transmission, the gear ratio of the continuously variable transmission is changed according to the turning characteristics of the vehicle body, particularly, the yaw rate of the vehicle. There is no disclosure of how to control the vehicle, and this technology poses a problem because the turning characteristics of the vehicle cannot be further improved. Further, when the feedback control is simply performed as the gear ratio corresponding to the target yaw rate of the continuously variable transmission, the actual gear ratio of the continuously variable transmission due to the turning of the vehicle is not taken into consideration (the straight traveling state becomes a reference. ), As a result, the control tends to be unstable, which is a problem.

An object of the present invention is to provide a vehicle driving force adjusting device which can further improve the turning characteristics according to the yaw rate of the vehicle and can secure the stability of control.

[0018]

In order to achieve the above-mentioned object, the invention of claim 1 is interposed between a first rotating shaft and a second rotating shaft and connected to the first rotating shaft. A first pulley having a movable sheave, a second pulley having a movable sheave connected to the second rotating shaft, the first pulley and the second pulley.
The winding radius of the endless belt that connects the pulley and the endless belt that slides the movable sheave of at least one of the first pulley and the second pulley is changed to change the first rotating shaft and the first rotating shaft. In a vehicular drive force adjusting device provided with a continuously variable transmission including a speed changing means for changing a speed ratio with a second rotating shaft, the speed changing means detects a rotational speed on the side of the first rotating shaft. A first rotation speed detection means, a second rotation speed detection means for detecting the rotation speed on the second rotation shaft side, a rotation speed on the first rotation shaft side detected by the first rotation speed detection means, and the first rotation speed detection means. Based on the traveling state of the vehicle, a gear ratio calculating means for calculating a gear ratio of the continuously variable transmission corresponding to a rotation speed difference from the rotation speed on the second rotating shaft side detected by the two rotation speed detecting means. , By the gear ratio calculation means A gear ratio determining means for determining the calculated gear ratio increase / decrease amount from the gear ratio, and a gear shift for sliding the movable sheave based on the gear ratio increase / decrease amount from the gear ratio determined by the gear ratio determining means. And an operating means.

According to a second aspect of the invention, in the vehicle driving force adjusting apparatus according to the first aspect, a yaw rate detecting means for detecting a yaw rate of the vehicle, a steering angle detecting means for detecting a steering angle of the vehicle, and a vehicle body of the vehicle. A vehicle speed detecting means for detecting a speed, wherein the gear ratio determining means detects the actual yaw rate detected by the yaw rate detecting means by the steering angle detected by the steering angle detecting means and the vehicle body speed detecting means. It is characterized in that the gear ratio increase / decrease amount from the gear ratio calculated by the gear ratio calculating means is determined by performing feedback control so as to approach the theoretical yaw rate obtained from the calculated vehicle speed.

[0020]

According to the invention of claim 1, the vehicular driving force adjusting device provided with the continuously variable transmission includes the speed changing means and the speed changing operation means, and in particular, the speed changing means includes the speed ratio calculating means, the speed ratio determining means and the speed changing operation. And means. Here, the speed ratio difference between the rotation speed on the side of the first rotation axis detected by the first rotation speed detection means by the speed change ratio calculation means and the rotation speed on the side of the second rotation axis detected by the second rotation speed detection means. The speed ratio of the continuously variable transmission is calculated, and the speed ratio determining means determines the speed ratio increase / decrease amount from the speed ratio calculated by the speed ratio calculating means based on the running state of the vehicle. Since the movable sheave of the first pulley and the movable sheave of the second pulley are slid on the basis of the amount of increase / decrease in the gear ratio determined by the gear ratio determining means, the torque movement according to the amount of increase / decrease in the gear ratio based on the running state is moved to the left or right. Go between the wheels and turn the vehicle.

According to a second aspect of the present invention, in the vehicle drive force adjusting apparatus according to the first aspect, in particular, the yaw rate detecting means, the steering angle detecting means, and the vehicle body speed detecting means are provided, and further, the gear ratio determining means. However, in order to bring the actual yaw rate detected by the yaw rate detecting means close to the theoretical yaw rate obtained from the steering angle detected by the steering angle detecting means and the vehicle body speed detected by the vehicle body speed detecting means, feedback control is performed. The movable sheaves of the first and second pulleys are slid with the gear ratio increase amount obtained by the yaw rate feedback control, and torque is moved between the left and right wheels according to the gear ratio increase / decrease amount to turn the vehicle.

[0022]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS The vehicle driving force adjusting device shown in FIGS. 1 and 2 is mounted over a power transmission system and a steering system of a passenger vehicle. Here, a center differential 2 is arranged at the center of the vehicle body 1, and the center differential 2 is rotated by an engine side (not shown) to divide and transmit the rotational force to the front and rear propeller shafts 3, 4. The front propeller shaft 3 transmits the input torque to the front left and right wheels 6, 7 via the front differential 5, and the rear peller shaft 4 transmits the left and right axles 9, 10 and the rear left and right wheels 11, 1 via the rear differential 8.
Transmit the input torque to 2. Furthermore, the left and right axles 9 on the rear wheel side,
10 is connected by a continuously variable transmission 13 bypassing the rear differential 8.

The left axle 9 has a first gear 14 integrally attached to its middle portion, and the right axle 10 has a first gear 14 of the continuously variable transmission 13.
The fixed sheave 16 (see FIG. 5) of the pulley 15 is integrally attached. The detour shaft 1 is arranged in parallel with the left and right axles 9 and 10.
7 are opposite. The second gear 18 is attached to one end of the bypass shaft 17, and the fixed sheave 21 (see FIG. 5) of the second pulley 20 of the continuously variable transmission 13 is integrally attached to the other end of the bypass shaft 17. The second gear 18 meshes with the first gear 14 via an idle gear 19, and the second pulley 20 is connected to the first pulley 15 via an endless belt 22. Here, the first gear 14 and the second gear 18 have the same shape and rotate at the same speed in the same direction. The left and right axles 9 and 10 are pivotally supported by an axle housing (not shown), and a bulging portion (not shown) of the axle housing has a bypass shaft 17 and an idle gear 1.
9. The continuously variable transmission 13 and the like are also pivotally supported.

As shown in FIGS. 1 and 5, the continuously variable transmission 13
Is equipped with a first pulley 15 and a second pulley 20, and a steel endless belt 22 is stretched over the first pulley 15 and the second pulley 20. Both of the pulleys 15 and 20 are divided into two parts, and the movable sheaves 23 and 24 are fixed sheaves 16,
It is inserted in the 21 side so that relative rotation is impossible and a relative space | interval is separable. The movable sheaves 23 and 24 and the fixed sheaves 16 and 21
Between the first cylinder 25 and the second cylinder 2 as hydraulic actuators for operating the relative distance between the two pulleys.
And 6 are formed.

A pair of first and second rotational speed detecting means for detecting the rotational speeds V _RR and V _RL of both the first pulley 15 (rear right wheel 12) and the second pulley 20 (rear left wheel 11). The rotation sensors 27 and 28 are mounted as means for detecting the actual gear ratio γr (= V _RR / V _RL ). In this case, the movable sheave 23 is brought closer to the fixed sheave 16 of the first pulley 15 to increase the winding diameter of the first pulley, and the movable sheave 24 is moved away from the fixed sheave 21 of the second pulley 20 to reduce the winding diameter. As a result, the actual gear ratio ρr (= V _RR / V
_RL ) is increased, the rear left wheel is rotated higher than the rear right wheel, torque is transmitted from the rear right wheel to the rear left wheel, and the reverse operation is performed to achieve a low gear ratio, and the rear right wheel is higher than the rear left wheel. Rotates and transmits torque from the rear left wheel to the rear right wheel. Continuously variable transmission 1
The hydraulic circuit 29 for controlling the No. 3 will be described with reference to FIG.

The oil pump 30 is driven by the engine side (not shown), and the hydraulic pressure discharged from the oil pump 30 is adjusted by the regulator valve 31 to an appropriate pressure, so-called line pressure. This regulator valve 31
The duty is controlled by the electromagnetic control valve 33 driven by the duty ratio set in the CVT ECU 32 according to the driving state of the vehicle. The line pressure regulated by the regulator valve 31 is connected to the cylinders 25 and 26 of the first and second pulleys 15, and the first and second speed ratio control valves 3 are connected.
4,35. Each gear ratio control valve 34, 35 receives a pilot pressure from the first and second electromagnetic control valves 36, 37 to perform a pressure adjusting operation. An output of a predetermined duty ratio is input from the CVTECU 32 to the first and second electromagnetic control valves 36 and 37.

Here, the CVTECU 32 sets the duty ratio so that the gear ratio ρr of the first and second pulleys 15 and 20 becomes a value set according to the operating state of the vehicle, and the outputs are used to output the first and second duty ratios. First and second via the electromagnetic control valves 36, 37
The gear ratio control valves 34 and 35 are driven to set the desired hydraulic pressure to the first,
Supply into each cylinder 25, 26 of the second pulley.

The line pressure is the modulator valve 38.
The hydraulic pressure is regulated by the first and second electromagnetic control valves 36, 3
It is supplied to 7th grade.

The CVTECU 32 is connected to an engine controller (not shown) and exchanges signals with each other. Moreover, the CVTECU 32 responds to the input port (not shown) of both rotation speeds V _RR and V _RL of the first pulley 15 (rear right wheel 12) and the second pulley 20 (rear left wheel 11) and the steering wheel angle of the steering wheel 39. The steering angle δ _f information is input from the steering angle sensor 40 serving as the steering angle detecting means, and the yaw rate γr (rad / s) information is input from the yaw rate sensor 41 serving as the yaw rate detecting means for the vehicle body weight center position G. Rotational speed V _fR of the front left and right wheels, which are the driving wheels,
V _fL is input from the pair of rotation sensors 42 and 43, and a control signal is output to the electromagnetic control valve 33 and the first and second electromagnetic control valves 36 and 37 of the hydraulic circuit 29 through an output port (not shown). ing.

A main part of the CVTECU 32 is constituted by a microcomputer, and a built-in memory circuit has a turning moment setting map m1 and sheave pressing force F shown in FIG.
_{The A} and F _B setting maps m2 and the control program of the flowchart of FIG. 12 are stored. Here, a block diagram showing the configuration of the vehicle drive force adjusting apparatus of FIG. 1 by function is shown in FIG.

In this vehicle drive force adjusting apparatus, the gear ratio ρr of the first pulley 15, the second pulley 20, the endless belt 22, the right axle 10 and the bypass shaft 17 (which rotates in the same manner as the left axle 9). A continuously variable transmission 13 including a transmission means A for varying (= V _RR / V _RL ) is provided, and in particular, the transmission means A is a first rotation speed detection means for detecting a rotation speed V _RR of the right axle 10. Of the rotation sensor 27 and the rotation speed V _RL of the left axle 9 side
Rotation sensor 2 as second rotation speed detection means for detecting
8, the gear ratio calculation means A1 for calculating the gear ratio ρr of the continuously variable transmission 13 corresponding to the difference in rotation speed between the rotation speed V _RR and the rotation speed V _RL, and the gear ratio ρr based on the running state of the vehicle.
, That is, the gear ratio increase / decrease amount ± Δ for correcting the gear ratio ρr
Gear ratio determining means A2 for determining ρ, and gear ratio determining means A
The gear shift operation means A3 for sliding the movable sheaves 23, 24 based on the target gear ratio ρo (= ρr + Δρ) determined by 2.

Further, this vehicle driving force adjusting device includes a yaw rate sensor 41 for detecting a yaw rate γr of the vehicle,
The steering angle sensor 40 for detecting the steering angle of the vehicle and the average value of the rotational speeds V _fR and V _fL of the front left and right wheels are calculated as the vehicle body speed V (=
And a pair of rotation sensors 42 and 43 as a vehicle body speed detecting means for detecting as V _fR / V _fL ). Moreover, in particular, the gear ratio determining means A2 calculates the theoretical yaw rate γo from the steering angle δ _f and the vehicle body speed V (A21), and performs feedback control processing so that the actual yaw rate γr approaches the theoretical yaw rate γo (A22). This feedback control sets the gear ratio increase / decrease amount Δρ (A23). In addition,
The speed change means A determines the target speed ratio ρ from the speed ratio increase / decrease amount Δρ and the speed ratio ρr calculated by the speed ratio calculation means A1.
o (= ρr + Δρ) is determined (A24).

CVT of such a vehicle driving force adjusting apparatus
The control function performed by the ECU 32 will be described with reference to the block diagram of FIG. The CVTECU 32 includes a pair of rotation sensors 42, 4
The rotation speeds V _fR and V _{fL of the} front left and right wheels are taken in from 3, and the average value is calculated as the vehicle body speed V (= V _fR / V _fL ). The current steering angle δ _f is fetched from the steering angle sensor 40, and the theoretical yaw rate γo is expressed by the equation (8) with this and the vehicle body speed V as variables and the wheel base L and the stability factor K as constants. It is calculated.

Γo = V × δ _f / L × (1 + K · V ² ) ... (8) The current yaw rate γr is taken in from the yaw rate sensor 41, and the deviation between this and the theoretical yaw rate γo Δ
γ is required. Here, the feedback correction coefficient Y _FB is calculated in order to perform feedback control so that the actual yaw rate γr approaches the theoretical yaw rate γo. In this case, the proportional term YP (Δγ) corresponding to the deviation Δγ, the deviation Δ
An integral term ΣYI (Δγ) corresponding to γ and time integration is calculated appropriately, and these values are added in the feedback region and used as the feedback correction coefficient Y _FB for the PI control shown in FIG. 11. Note that FIG. 10 schematically shows the change characteristic of the steering angle δ _f with time at this time.

The theoretical yaw rate .gamma. O is modified by the ratio of the feedback correction coefficient Y _FB, i.e., calculates the (1 + Y _FB) multiplied by setting the yaw rate _{γc (= (1 + Y FB} ) × γo). The set yaw rate γc is calculated as the yaw moment M by multiplying the moment conversion coefficient mα (Kg · m · s / rad). This moment conversion coefficient m
α is set so that the yaw moment for left turn is (+) and the yaw moment for right turn is (−), and this value is set in advance for each vehicle by actual measurement or the like. FIG. 4 shows the gear ratio setting map m1.

The gear ratio setting map m1 is set so that the gear ratio of the continuously variable transmission 13 is increased or decreased from 1.0 (zero yaw moment) to positively generate the yaw moment M on the vehicle body. Here, the gear ratio increase / decrease amount Δρ is set as (> 1) when the left turning moment (+ M is set) is applied to the vehicle body, and the right turning moment (−M) is set.
Is set), the gear ratio increase / decrease amount Δρ is set as (<1). This gear ratio setting map m1 is in a region where the slip ratio of the vehicle is proportional to the drive torque (a region where the drive torque is not excessively large) and the difference between the slip ratios of the left and right wheels is the gear ratio ρr of the continuously variable transmission 13. Is set and the yaw moment M of the vehicle is set in consideration of each vehicle characteristic based on the point corresponding to the drive torque difference between the left and right wheels.

On the other hand, the first pulley 15 (same rotation as the right axle 10) by the detection signals of the pair of rotation sensors 27, 28.
And the actual gear ratio ρr (= V _RR / V _RL ) between the second pulley 20 (the same rotation as the left axle 10 and the bypass shaft 17) is calculated. The gear ratio ρr and the gear ratio increase / decrease amount Δρ obtained from the gear ratio setting map m1 are added to calculate the target gear ratio ρo. Further, in FIG. 4, the sheave pressing force setting map m
2 is shown. This sheave pressing force setting map m2 is the fixed sheave 16 required to increase or decrease the gear ratio of the continuously variable transmission 13 from 1.0 (when the winding diameters around the first pulley 15 and the second pulley 20 are the same). , 21 are set to the sheave pressing forces F _A and F _B of the movable sheaves 23 and 24.

Here, the target gear ratio ρo is 1.0 and the first
The sheave pressing forces F _A and F _B of the second pulleys 15 and 20 are set to the minimum value F0. In this case, the first and second pulleys 15 and 2
In the case of 0, the endless belt 22 is wound in a slip-free state, but the minimum value F0 is set so that the state in which torque is not transmitted can be maintained. On the other hand, the target gear ratio ρo is 1.0
In the case of a left turn exceeding 1, the winding diameter of the first pulley 15 needs to increase and the winding diameter of the second pulley 20 needs to decrease. Here, the pressing force F _A of the first pulley 15 is set to increase, The pressing force F _B of the second pulley 20 is held at the minimum value F0. On the contrary, in the case of the right turn in which the target gear ratio ρo is less than 1.0, the winding diameter of the first pulley 15 needs to be decreased and the winding diameter of the second pulley 20 needs to be increased. The pressing force F _A of the pulley 15 is kept at the minimum value F0,
The pressing force F _B of the second pulley 20 is set to increase.

The sheave pressing forces F _A and F _B set by the sheave pressing force setting map m2 as described above generate the hydraulic pressure required to generate the pressing forces by the first cylinder 25 of both pulleys.
And is supplied to the second cylinder 26. Here, the duty drive section 321 that constitutes the shift operation means A3 of the CVTECU 32 calculates the respective duty ratios DU1 and DU2 that can achieve the hydraulic pressures P _A and P _B of the first cylinder 25 and the second cylinder 26, and outputs the first with the same output. , The first and second speed ratio control valve 3 via the second electromagnetic control valves 36 and 37.
4, 35 are driven to drive the first cylinder 25 and the second cylinder 2
The hydraulic pressures P _A and P _B are supplied to 6 to increase or decrease the winding diameters of the first pulley 15 and the second pulley 20 to achieve the target gear ratio ρo. Next, the operation of the vehicle drive force adjusting apparatus of FIG. 1 will be described.

The front, rear, left and right wheels 6, 7, 11, 12 are driven by transmitting the output rotation from the engine through the center differential 2, the front and rear propeller shafts 3, 4, the front and rear differentials 5, 8, etc. The driving force input to the wheels 11 and 12 is adjusted by the moving torque ΔT corresponding to the gear ratio ρr of the continuously variable transmission 13. The CVT ECU 32 includes a steering angle sensor 40, a yaw rate sensor 41, and a pair of rotation sensors 4 for detecting vehicle speed.
2, 43, a pair of rotation sensors 2 for detecting the actual gear ratio ρr
The output values of 7, 28, etc. are sequentially taken in according to the running state of the vehicle, and the gear ratio of the continuously variable transmission 13 is adjusted via the first and second electromagnetic control valves 36, 37 in the hydraulic circuit 29 based on the same value. Then, the moving torque ΔT to the rear left and right wheels 11, 12 is adjusted.

The CVTECU 32 feeds back from the actual yaw rate γr and the target yaw rate γo according to a control program as shown in FIG.
The I control is used to eliminate it. Below, FIG.
The operation will be described according to the control program shown in FIG.

First, when control is started by turning on a main switch (not shown), the steering angle δ _f , yaw rate γ r, front left and right wheel speeds V _fR , V _fL , and the first sensor
Rear left and right wheel speeds V _RR and V _RL , which are the speeds of the pulley 15 and the second pulley 20, are taken in (step s1). In steps s2 and s3, the average value of the front left and right wheel speeds V _fR and V _fL is calculated as the vehicle body speed V (= V _fR / V _fL ), and the average value of the rear left and right wheel speeds is actually calculated by the continuously variable transmission 13. Gear ratio ρr
It is calculated as (= V _RR / V _RL ). In step s4, the vehicle speed V (= V _fR / V _fL ) and the current steering angle δ _f are used as variables, the wheel base L and the stability factor K are used as constants, and the theoretical yaw rate γo is calculated from the equation (8). .

Γo = V × δ _f / L × (1 + K · V ² ) ... (8) In step s5, the deviation Δγ between the current yaw rate γr and the theoretical yaw rate γo is obtained, and then the deviation Δγ. The proportional term YP (Δγ) according to the above, the deviation Δγ, and the integral term ΣYI (Δγ) according to the time integration are obtained, and the feedback correction coefficient Y _FB is obtained. Here, the feedback correction coefficient Y _FB is calculated in order to perform feedback control so that the actual yaw rate γr approaches the theoretical yaw rate γo. In this case, the proportional term YP (Δγ) corresponding to the deviation Δγ, the deviation Δ
γ and the integral term ΣYI (Δγ) corresponding to the time integration are appropriately calculated, these values are added to calculate the feedback correction coefficient Y _FB , and the theoretical yaw rate γo is corrected by the ratio of the feedback correction coefficient Y _FB. (1+
Y _FB ) is multiplied to set yaw rate γc (= (1+
Y _FB ) × γo) is calculated.

In this case, the PI control shown in FIG. 11 is performed along the displacement of the steering angle δ _{f in} FIG. 10, that is, when the actual yaw rate γr is corrected to the theoretical yaw rate γo, the deviation Δγ is feedback-corrected. Since the correction is performed by the coefficient Y _FB and the correction target is set to the set yaw rate γc, the control is performed, so that the control with good responsiveness can be executed. Theoretical yaw rate γ
o is corrected by the ratio of the feedback correction coefficient Y _FB ,
That is, (1 + Y _FB ) is multiplied to set yaw rate γc (=
(1 + Y _FB ) × γo) is calculated. Step s6
When the process proceeds to s7, the set yaw rate γc is multiplied by the preset moment conversion coefficient mα to calculate the yaw moment M.
Is calculated. Next, along the gear ratio setting map m1,
A gear ratio increase / decrease amount Δρ corresponding to the yaw moment M is calculated.

Here, the gear ratio increase / decrease amount Δρ is set to (> 1) for the left yaw moment (set to + M), and the gear ratio increase / decrease amount Δρ is set to (<1) for the right yaw moment (set to −M). . When the process proceeds to steps s8, s9, and s10, the actual speed ratio ρr (= V _RR / V _RL ) and the speed ratio increase / decrease amount Δρ are added to calculate the target speed ratio ρo. Next, the sheave pressing forces F _A and F _B corresponding to the target gear ratio ρo are set along the sheave pressing force setting map m2. Further, the sheave pressing forces F _A and F _B are converted into respective duty ratios DU _A and DU _B capable of achieving the oil pressures P _A and P _B equivalent to the pressing forces. In step s11, the first and second electromagnetic control valves 36 and 37 are driven with the duty ratios DU _A and DU _B , and one control cycle ends.

Here, when the vehicle is traveling straight ahead, as shown in FIG. 7, the target speed ratio ρo is set to 1.0, and the sheave pressing forces F _A and F _{B of} the first and second pulleys 15 and 20 are changed. The minimum value F0 is set, and the first and second pulleys 15 and 20 are endless belts 22.
Is wound in a slip-free state, the transfer torque is not transmitted, and the rear left and right wheels 11 and 12 _receive only the split torque 1/2 × T _IN divided by the rear differential 8 with respect to the input torque T _IN. It receives and drives to rotate respectively. On the other hand, if the vehicle is making a left turn, as shown in FIG. 8, the target gear ratio ρo exceeds 1.0, the pressing force F _A of the first pulley 15 is set to increase, and the second
The pressing force F _B of the pulley 20 is kept at the minimum value F0,
The winding diameter of the pulley 15 increases, the winding diameter of the second pulley 20 decreases, and the moving torque ΔT is transmitted from the rear right wheel 12 side to the rear left wheel 11 having a high rotation speed and a high slip ratio. The torque of 11 is T _L = 1/2 × T _IN + Δ
T, the torque of the rear right wheel 12 is T _R = 1/2 × T _IN −ρ ×
It becomes ΔT, and the left turn is made with good responsiveness.

On the other hand, if the vehicle is making a right turn, as shown in FIG. 9, the target gear ratio ρo falls below 1.0, and the first pulley 15
The pressing force F _A of the second pulley 20 is maintained at the minimum value F0.
Pressing force F _B is increased settings, the winding diameter of the first pulley 15 is reduced, the winding diameter of the second pulley 20 is increased,
Rear left wheel 1 to rear right wheel 12 with high rotation speed and high slip ratio
The moving torque ΔT is transmitted from the 1 side, the torque of the rear left wheel 11 becomes T _L = 1/2 × T _IN −ΔT, the torque of the rear right wheel 12 becomes T _R = 1/2 × T _IN + ρ × ΔT, A right turn is made with good responsiveness.

As described above, the vehicular driving force adjusting apparatus of FIG. 1 is provided with the speed changing means A including the speed ratio calculating means A1, the speed ratio determining means A2, and the speed changing operation means A3 (see FIG. 3).

Here, the speed ratio calculating means A1 corresponds to the speed difference ρr corresponding to the speed difference between the rotation speeds V _RR and V _{RL of} the first pulley 15 (same rotation as the right axle 10) and the second pulley 20.
(= V _RR / V _RL ) is calculated, and the gear ratio determination means A2 determines the gear ratio increase / decrease amount Δρ from the gear ratio ρr based on the running state of the vehicle, that is, the actual yaw rate detected by the yaw rate sensor 41. Feedback control is performed in order to bring γr closer to the theoretical yaw rate γo obtained from the steering angle δ _f detected by each steering sensor 40 and the vehicle speed V (= V _fR / V _fL ). Since the movable sheaves 23 and 24 of the first and second pulleys are slid based on the gear ratio increase amount Δρ, the gear ratio increase / decrease amount Δρ
The moving torque ΔT is adjusted between the left and right wheels according to the above, and the vehicle is turned with high responsiveness according to the moving torque, and the turning characteristics are improved.

Furthermore, since the moving torque ΔT is moved from one of the left and right wheels to the other by the continuously variable transmission 13, there is little energy loss, and the current gear ratio ρr and the gear ratio increase / decrease amount for correcting the gear ratio are also present. Since the target gear ratio ρo is set with Δρ, gear ratio control based on the current turning state of the vehicle, that is, turning control can be performed, and this turning control can be performed with high accuracy and responsiveness, resulting in stable control. Turn into. Next, the turning operation characteristics of the vehicle equipped with the vehicle drive force adjusting apparatus of FIG. 1 will be sequentially described.

13 to 24, the characteristic of the conventional vehicle in which the moving torque between the left and right rear wheels is not adjusted is indicated by a broken line b.
The characteristics of the vehicle equipped with the vehicle drive force adjusting apparatus of the present invention are shown by solid lines a. FIG. 13 shows the relationship between the steering ratio δ _f / δ _f o expressed by the equation (7) and the lateral acceleration.
Indicates the yaw moment M acting on the vehicle at that time. Here, the turning radius is 50 m and the longitudinal acceleration is 0.1.
It was set to m / s ² . In this case, the turning amount of the steering angle is kept in the linear characteristic range when turning.
In the conventional vehicle, the lateral acceleration is up to 4 m / s ² , whereas in the vehicle to which the present invention is applied, the linear characteristic range is expanded up to the lateral acceleration of 6 m / s ² , and the turning characteristic due to the addition of the yaw moment M. Is clearly improved.

In FIG. 15, the steering angle is fixed and the turning radius is 30.
m, the turning trajectory when the longitudinal acceleration is set to 2.5 m / s ² , and FIGS. 16 and 17 show the yaw moment M and the yaw rate γr acting on the vehicle at that time. In this case, the vehicle to which the present invention is applied is closer to the reference turning trajectory Br than the conventional vehicle, and it is clear that the turning characteristics are improved by adding the yaw moment M accompanying the fluctuation of the yaw rate γr. In Fig. 18, the steering angle is fixed and the turning radius is 5
FIG. 19 and FIG. 20 show yaw moments −M acting on the vehicle at that time in the case of engine braking running in which the longitudinal acceleration is set to 0 m and the longitudinal acceleration is set to −2.5 m / s ² .
(For right turn), yaw rate γr is shown. in this case,
The vehicle to which the present invention is applied has a reference turning path Br rather than a conventional vehicle.
The turning accuracy is improved by adding a negative yaw moment −M (for right turn) accompanying the fluctuation of the yaw rate γr.

FIG. 21 shows the traveling track when the lane is changed (lane change) at a speed of 30 km / s.
Shows the steering angle δ _f and the yaw moment M in FIG. In this case, the vehicle to which the present invention is applied has a faster convergence of the steering angle δ _{f to} the straight-ahead position (δ _f = 0) due to the addition of the yaw moment than the conventional vehicle, and the steering stability is improved. It has become clear. FIG.
Shows the nonlinear behavior of the steering angle (rad) and the yaw rate (rad / s) at the time of lane change. Here, the ideal line _{BL for the} vehicle to which the present invention is applied is higher than that for the conventional vehicle.
It is clear that the spread width along the line is narrow, and in this respect as well, it quickly converges to straight running and improves steering stability.

In the above description, the continuously variable transmission 13 uses the first cylinder 25 and the second cylinder 26 as hydraulic actuators for driving the movable sheaves 23, 24.
Instead of this, a continuously variable transmission 13a as shown in FIG. 25 may be used. The continuously variable transmission 13a here includes an electromagnetic solenoid 4 as an actuator for the movable sheaves 23 and 24.
4 and 45 are used, and the electromagnetic force of the electromagnetic solenoids 44 and 45 is used as the sheave pressing forces F _A and F _B. In this case as well, the movable sheaves 23 and 24 are provided on the fixed sheaves 16 and 21 side so as to be relatively non-rotatable and capable of contacting and separating at a relative interval. The electromagnetic solenoids 44 and 45 are connected to the CVTE via a drive circuit (not shown).
Is connected to the Cu 32, the sheave pressure F _A, F _B equivalent output current I _A, supplied with I _B, is adjusted to the first pulley 15 the winding diameter of the second pulley 20, the gear ratio ρr is Adjusted. This continuously variable transmission 13a is also the continuously variable transmission 13 of FIG.
Similarly to the above, it is adopted in the vehicle driving force adjusting apparatus of the present invention and exhibits the same operation and effect.

[0055]

As described above, according to the first aspect of the invention, the vehicle driving force adjusting device including the continuously variable transmission includes the speed changing means and the speed changing operating means, and particularly, the speed changing means includes the speed ratio calculating means. A gear ratio determination means and a gear shift operation means are provided.

Here, the gear ratio calculation means and the second rotation speed on the first rotation shaft side detected by the first rotation speed detection means and the second rotation speed.
The gear ratio of the continuously variable transmission corresponding to the rotational speed difference from the rotational speed on the second rotary shaft side detected by the rotational speed detecting means is calculated, and the speed ratio determining means calculates the speed ratio calculated by the speed ratio calculating means. Is determined based on the running state of the vehicle, and the gear shift operating means determines the movable sheave of the first pulley and the movable sheave of the second pulley based on the gear ratio increase / decrease determined by the gear ratio determining means. Since the vehicle is slid, the moving torque according to the increase / decrease in the gear ratio is increased / decreased between the left and right wheels, and a yaw moment is added to the vehicle in accordance with the torque movement to turn with good responsiveness, thereby improving turning characteristics. Furthermore, since the moving torque is moved from one of the left and right wheels to the other by the continuously variable transmission, energy loss is small, and the target gear ratio is determined by the current gear ratio and the gear ratio increase / decrease amount for correcting the gear ratio. Since the setting is performed, turning control based on the current turning state of the vehicle can be performed, and this turning control can be performed with high accuracy and responsiveness, and as a result, turning control is stabilized.

According to a second aspect of the present invention, in the vehicle drive force adjusting apparatus according to the first aspect, in particular, a yaw rate detecting means, a steering angle detecting means, and a vehicle body speed detecting means are provided, and further, a gear ratio determining means. However, feedback control is performed so that the actual yaw rate detected by the yaw rate detecting means approaches a theoretical yaw rate obtained from the steering angle detected by the steering angle detecting means and the vehicle body speed detected by the vehicle body speed detecting means. Therefore, feedback control is performed in order to bring the actual yaw rate closer to the theoretical yaw rate, and the gear ratio increase / decrease amount is determined based on the deviation of the yaw rate.
Since the moving torque is adjusted between the left and right wheels according to the increase / decrease in the gear ratio, the moving torque can be adjusted according to the yaw rate of the vehicle to turn the vehicle with high responsiveness and improve the turning characteristics. .

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram of a vehicle equipped with a vehicle driving force adjusting device as an embodiment of the present invention.

FIG. 2 is a schematic side view of the vehicle of FIG.

3 is a block diagram of constituent means of the vehicle driving force adjusting apparatus of FIG. 1. FIG.

FIG. 4 is a functional block diagram of control performed by CVTECU of the vehicle driving force adjusting apparatus of FIG.

5 is a cross-sectional view of a continuously variable transmission used in the vehicle drive force adjusting apparatus of FIG.

FIG. 6 is a block surface of a hydraulic circuit connected to the continuously variable transmission of FIG.

FIG. 7 is a diagram for explaining torque movement when the vehicle drive force adjusting apparatus of FIG. 1 is traveling straight ahead.

FIG. 8 is a diagram for explaining torque movement when the vehicle drive force adjusting apparatus of FIG. 1 is turned left.

FIG. 9 is a diagram for explaining torque movement when the vehicle drive force adjusting apparatus of FIG. 1 turns right.

FIG. 10 is a characteristic diagram of time-dependent change of steering angle of the vehicle driving force adjusting apparatus of FIG. 1.

FIG. 11 is a characteristic diagram of a change over time of a yaw rate feedback coefficient of the vehicle driving force adjusting apparatus of FIG.

FIG. 12 is a flowchart of a control program executed by the vehicle driving force adjusting apparatus of FIG.

13 is a diagram comparing the steering angle ratio-lateral acceleration characteristics of a vehicle equipped with the vehicle drive force adjusting apparatus of FIG. 1 and a vehicle not equipped with the vehicle driving force adjustment apparatus.

FIG. 14 is a diagram comparing the yaw moment-lateral acceleration characteristics of the equipped vehicle and the unequipped vehicle of FIG.

FIG. 15 is a diagram comparing the characteristics of a turning track with a constant steering angle of a vehicle equipped with the vehicle drive force adjusting apparatus of FIG. 1 and a vehicle not equipped with the vehicle driving force adjustment apparatus.

16 is a diagram comparing the yaw moment-time characteristics of the equipped vehicle and the unequipped vehicle of FIG.

17 is a diagram comparing the yaw rate-time characteristics of the equipped vehicle and the unequipped vehicle of FIG.

18 is a diagram comparing the characteristics of the turning trajectory during engine braking with a constant steering angle of a vehicle equipped with the vehicle drive force adjusting apparatus of FIG. 1 and a vehicle not equipped with the vehicle driving force adjustment apparatus.

FIG. 19 is a diagram comparing the yaw moment-time characteristics of the equipped vehicle and the unequipped vehicle of FIG. 18.

20 is a diagram comparing the yaw rate-time characteristics of the equipped vehicle and the unequipped vehicle of FIG.

FIG. 21 is a diagram showing a traveling track at the time of lane change of a vehicle equipped with the vehicle driving force adjusting device of FIG. 1.

22 is a diagram comparing the time-dependent characteristics of the steering angle when the lane is changed between the equipped vehicle and the non-equipped vehicle of FIG. 21.

23 is a diagram showing a change characteristic of a yaw moment with time when the lane of the equipped vehicle of FIG. 21 is changed.

24 is a characteristic diagram of yaw moment-steering angle when the equipped vehicle of FIG. 21 changes lanes.

FIG. 25 is a cross-sectional view of a continuously variable transmission as another embodiment that can be used in the vehicle drive force adjusting apparatus of FIG. 1.

FIG. 26 is a schematic configuration diagram of a conventional vehicle driving force adjusting device.

FIG. 27 is a schematic configuration diagram of another conventional vehicle driving force adjusting apparatus.

FIG. 28 is a schematic configuration diagram of another conventional vehicle driving force adjusting apparatus.

FIG. 29 is a perspective view schematically showing the behavior of four wheels when the vehicle power transmission system and the vehicle are turning.

FIG. 30 is a diagram illustrating a force applied to two wheels when the vehicle turns.

FIG. 31 is a plan view schematically showing the behavior of the two wheels when the vehicle turns.

FIG. 32 is a diagram showing a relationship between vehicle steering angle ratio-square value of vehicle speed.

FIG. 33 is a diagram showing the relationship between the turning force of the vehicle and the sideslip angle. 11 rear left wheel 12 rear right wheel 15 first pulley 16 fixed sheave 20 second pulley 21 fixed sheave 22 endless belt 23 movable sheave 24 movable sheave 25 first cylinder 26 second cylinder 27 rotation sensor 28 rotation sensor 32 CVT ECU 40 steering angle sensor 42 rotation sensor 43 rotation sensor γr yaw γr actual gear ratio γo theoretical yaw rate Δρ speed ratio decrease amount ρo target gear ratio ρr actual left wheel speed V _fR speed rear right wheel speed V _RL after the speed ratio [delta] _f steering steering angle V _RR V _fL Rotational speed V Vehicle speed A _Gear ratio A1 Gear ratio calculation means A2 Gear ratio determination means A3 Gear shift operation means ΔT Moving torque M Yaw moment L Wheelbase K Stability factor

Claims (2)

[Claims] 1. A first rotary shaft which is interposed between a first rotary shaft and a second rotary shaft and which is connected to the first rotary shaft and has a movable sheave.
At least one of a pulley, a second pulley connected to the second rotation shaft and having a movable sheave, an endless belt connecting the first pulley and the second pulley, and the first pulley and the second pulley. A continuously variable transmission comprising a transmission means for varying the gear ratio between the first rotating shaft and the second rotating shaft by changing the winding radius of the endless belt that slides one of the movable sheaves. In the vehicle driving force adjusting device, the transmission means includes a first rotation speed detecting means for detecting a rotation speed on the first rotation shaft side and a second rotation speed for detecting a rotation speed on the second rotation shaft side. Rotation speed of the speed detection means, the rotation speed of the first rotation shaft side detected by the first rotation speed detection means, and the rotation speed of the second rotation shaft side detected by the second rotation speed detection means. The stepless corresponding to the difference A gear ratio calculating means for calculating a gear ratio of the transmission; and a gear ratio determining means for determining an increase / decrease amount of the gear ratio from the gear ratio calculated by the gear ratio calculating means based on a running state of the vehicle. A drive force adjusting device for a vehicle, comprising: a shift operation means for sliding the movable sheave based on an increase / decrease amount of a gear ratio from the gear ratio determined by the gear ratio determining means. 2. A vehicle driving force adjusting apparatus according to claim 1, wherein a yaw rate detecting means for detecting a yaw rate of the vehicle, a steering angle detecting means for detecting a steering angle of the vehicle, and a vehicle body for detecting a vehicle body speed of the vehicle. Speed change means, the gear ratio determination means, the yaw rate detected by the yaw rate detection means, the steering angle detected by the steering angle detection means and the vehicle body speed detected by the vehicle body speed detection means A drive force adjusting device for a vehicle, wherein the amount of increase / decrease of the gear ratio from the gear ratio calculated by the gear ratio calculating means is determined by performing feedback control so as to approach the theoretical yaw rate obtained from the above.