JPH04342615A - Torque distribution controller for four-wheel drive vehicle - Google Patents

Abstract

PURPOSE:To prevent a driver from being applied with the discomfortable steering power variation by installing a yaw moment generating direction judging means having the car speed, longitudinal G, and the longitudinal torque distribution ratio as parameters, and a correcting means for correcting the rear differential limit torque to zero when the yaw moment is generated in the understeering direction under the gentle acceleration condition. CONSTITUTION:Though a rear clutch is feed/forward-controlled by directly claculating the rear differential limit torque Td from the longitudinal G, lateral G, supposed road surface (u), and a stability factor A, a yaw moment is generated on a vehicle according to the driving power state and rear differential limit torque Td, and the yaw moment and the generation direction are actualy calculated from the car speed V, longitudinal G, and the longitudinal torque distribution ratio alpha, and judged. Through an oversteered state tends to be generated in the case where the car speed or the longitudinal G is large, sometimes an understeered state is generated through the torque shift from outer wheels to inner wheels under the gentle accelerating traveling condition. In this case, the rear differential limit torque is corrected to zero, and the generation of torque shift is prevented, and the aimed steering characteristic is obtained.

Description

[Detailed description of the invention]

0001

[Industrial Application Field] The present invention is a full-time four-wheel drive vehicle equipped with a center differential. The present invention relates to a torque distribution control device that variably controls torque distribution, and more specifically, to correction measures when the steering angle and driving force are small in this torque distribution control system.

0002

2. Description of the Related Art Generally, it is known that vehicles have different unique driving performance depending on the drive system. In a full-time 4-wheel drive vehicle equipped with a center differential, all four wheels are driven at all times to avoid slips and skids that occur in FR and FF vehicles. Marginal performance is improved. In addition, since the effects of throttle on and off are distributed to the front and rear wheels at the same time, the tendency for both understeer and oversteer is weakened, resulting in characteristics that are intermediate between the two. This type of four-wheel drive vehicle is also becoming widespread among ordinary vehicles. In addition, in 4-wheel drive vehicles equipped with this center differential,
Torque distribution between the front and rear wheels and the left and right rear wheels further influences turning performance and changes in vehicle behavior, and by optimizing these torque distributions, it is possible to further improve driving performance and dynamic stability. . Therefore, by controlling the torque distribution between the front and rear wheels, etc.
Research and development is underway to provide optimal variable control according to driving conditions.

Conventionally, regarding torque distribution control between the front and rear wheels of a four-wheel drive vehicle equipped with the above-mentioned center differential, there is a prior art, for example, disclosed in Japanese Patent Laid-Open No. 63-13824. Here, the hydraulic multi-plate clutch is configured to be able to move torque with its differential limiting torque to vary the torque distribution between the front and rear wheels with respect to the center differential. In addition, the turning state of the vehicle can be detected by the lateral G, and as the value of this lateral G increases, the grip force of the tires gradually decreases and approaches the limit state, causing the vehicle to spin or drift. . Therefore, it has been shown that the differential limiting torque of the multi-disc clutch is set according to the value of the lateral G, and the torque distribution between the front and rear wheels is variably controlled so as not to cause spin or drift.

0004

[Problem to be Solved by the Invention] However, in the prior art described above, since the turning state is judged only by the value of lateral G, the lateral force is proportional to the side slip angle of the tire. Limited to varying linear grip areas. In other words, in the limit state where the tire grip force reaches its limit on a low μ road and the vehicle begins to spin, the lateral force changes nonlinearly and the actual lateral G value can be arbitrarily determined based on the vehicle's spinning behavior. This is because the turning state may change, making it impossible to accurately judge the turning state. On the other hand, in order to prevent spins in limit states, it is necessary to accurately judge the behavior of the vehicle in the nonlinear spin region and control the torque distribution between the front and rear wheels, and in this respect, the prior art is lacking. That's enough.

[0005] Torque distribution control for a four-wheel drive vehicle is as follows:
It is desirable to control the vehicle to improve stability, maneuverability, turning performance, etc. on all road surfaces such as low-μ roads and in driving conditions such as acceleration and deceleration during turns. To this end, it is possible to control the vehicle so that its steering characteristics are always constant.As a means of achieving this, a control principle is established in which the equation of motion of the vehicle, tire characteristics, etc. are extended to the nonlinear region, and longitudinal and lateral G. G.
Center differential, using the target stability factor parameters of road surface μ and steering characteristics.
Calculate the differential limiting torque of the rear differential,
A system for optimally controlling torque distribution has been proposed by the applicant. By the way, since this control principle is calculated using equations of motion, etc., there are some undesirable points if it is directly applied to an actual vehicle, and these need to be corrected as appropriate.

In other words, the differential limiting torque of the rear differential is controlled to reduce understeer during acceleration such as turning, but under conditions such as when the driving force is small, yaw moment in the understeer direction may occur. be. In addition, the differential limiting torque of the center differential is controlled closer to the front wheels in areas with small longitudinal and lateral G to reduce understeer during acceleration and prevent tuck-in during deceleration. For this reason, there is no problem on a low μ road, but on a high μ road with a small steering angle, the front wheel torque fluctuates when the accelerator is turned on and off, and an unpleasant change in steering force is transmitted to the driver. Furthermore, in this case, the differential limiting torque becomes large, which increases running resistance due to internal circulation torque, leading to a decrease in fuel efficiency. Therefore, it is desirable to make corrections to avoid such problems.

The present invention has been made in view of this point, and provides a system for variably controlling the torque distribution between the front and rear wheels and the left and right rear wheels so as to ensure good steering characteristics over the nonlinear region of low μ roads. The purpose of this invention is to perform appropriate correction in areas where the steering angle and driving force are small.

[0008]

[Means for Solving the Problems] In order to achieve the above object, the present invention provides front and rear wheels that extend into a non-linear region to obtain constant steering characteristics based on longitudinal G, lateral G, road surface μ, and target steering characteristics. In a control system that variably controls torque distribution between left and right rear wheels, a yaw moment generation direction determining means determines the direction of yaw moment generation using vehicle speed, longitudinal G, and longitudinal torque distribution ratio as parameters, and understeer direction under slow acceleration conditions. In the case where a yaw moment is generated, the rear differential limiting torque is corrected to zero.

[0009]

[Function] Based on the above configuration, when driving in four-wheel drive, target steering such as longitudinal G, lateral G, road surface μ, and stability factor, etc., which is set to correct steering characteristics during acceleration and deceleration during turning, etc. Using the characteristics as parameters and an equation of motion extended to the nonlinear range, torque distribution is controlled between the front and rear wheels and the left and right rear wheels so as to always obtain constant steering characteristics. At this time, if a yaw moment occurs in the understeer direction under slow acceleration conditions, the rear differential limiting torque is controlled to zero and torque movement between the left and right rear wheels is prevented, thereby achieving the desired steering characteristics. Obtainable.

0010

Embodiments Hereinafter, embodiments of the present invention will be explained based on the drawings. In FIG. 2, to explain the outline of the drive system of a full-time four-wheel drive vehicle equipped with a center differential, reference numeral 1 is an engine, 2 is a clutch, and 3 is a transmission, and the transmission output shaft 4 is connected to the center differential 20. is being entered. Center differential 2
0, the front drive shaft 5 is in front, and the rear drive shaft 6 is in the rear.
The front drive shaft 5 is connected to the left and right front wheels 9L and 9R via the front differential 7 and the axle 8, and the rear drive shaft 6 is connected to the propeller shaft 10 and the rear differential 1.
1. It is connected to the left and right rear wheels 13L and 13R via the axle 12, respectively, for transmission.

The rear differential 11 is of a bevel gear type, and a hydraulic multi-disc rear clutch 28 is provided as a differential limiting device between, for example, a differential case 11a and one side gear 11b. When the rear differential limiting torque Td of the rear clutch 28 is zero, the left and right rear wheels 13L, 1
When the torque is distributed equally to 3R and a predetermined rear differential limiting torque Td is generated, the torque is transferred from the high-speed wheels to the low-speed wheels by this torque Td, and the largest rear differential limiting torque Td is generated.
When the differential is locked, the torque is distributed according to the product W·μ of the wheel load W applied to the left and right rear wheels 13L, 13R and the road surface friction coefficient μ.

The center differential 20 is of a compound planetary gear type, and includes a first sun gear 21 integrated with the transmission output shaft 4, a second sun gear 22 integrated with the rear drive shaft 6, and a second sun gear 22 integrated with the rear drive shaft 6.
A plurality of pinions 23 are arranged around these sun gears 21 and 22, and the first pinion gear 23a of the pinion 23 is connected to the first sun gear 21, and the second pinion gear 2 is connected to the first sun gear 21.
3b mesh with the second sun gear 22, respectively. In addition, the transmission output shaft 4 is equipped with a reduction drive gear 25.
is rotatably provided, a pinion 23 is pivotally supported by a carrier 24 that is integrated with the drive gear 25, and the drive gear 25 is configured to mesh with a driven gear 26 that is integrated with the front drive shaft 5. As a result, the shift power input to the first sun gear 21 is transmitted to the carrier 24 and the second sun gear 22 in a predetermined standard torque distribution, and the difference in rotation between the front and rear wheels during turning is compensated for by the planetary rotation of the pinion 23. It begins to absorb. Here, the standard torque distribution is based on the four sun gears 21 and 22 and the two pinion gears 23a and 23b.
The gear meshing pitch circle radius can be set freely. Therefore, the standard torque distribution et of front wheel torque TF and rear wheel torque TR is, for example,

[Math 1] This makes it possible to set the weight to be sufficiently biased towards the rear wheels.

[0013] Also, the center differential 2
Immediately behind 0, a hydraulic multi-plate center clutch 27 is disposed coaxially with a drum 27a coupled to a carrier 24 and a hub 27b coupled to a rear drive shaft 6 integral with the second sun gear 22. The center differential limiting torque Tc of the center clutch 27 limits the differential of the center differential 20, and also makes it possible to transfer torque from the rear wheel side to the front wheel side. Here, in the case of a front engine installation, the static weight distribution ew of the front wheel weight WF and rear wheel weight WR of the vehicle is, for example,

In the case of direct connection using the center clutch 27, assuming that the road surface friction coefficient μ of the front and rear wheels is equal, torque is distributed to the front wheels in accordance with this weight distribution ew. Therefore, the torque distribution between the front and rear wheels is determined by the center differential limiting torque Tc of the center clutch 27, and the reference torque distribution e of the rear wheels is biased.
This makes it possible to control the weight distribution ew of the front wheel bias over a wide range.

Next, the hydraulic control system for the center clutch 27 and rear clutch 28 will be explained. First, if the transmission is an automatic transmission, check the oil pump 3 of its hydraulic control system.
It is constructed by using line pressure that is regulated by the regulator valve 31 from the zero oil pressure. Therefore, the center clutch hydraulic pressure control means 32 has a clutch control valve 34 communicating with a line pressure oil passage 33, and this clutch control valve 34 communicates with the center clutch 27 via an oil passage 35. The line pressure oil passage 33 also communicates with a solenoid valve 40 through an oil passage 38 having a pilot valve 36 and an orifice 37, and duty pressure from the solenoid valve 40 acts on the control side of the clutch control valve 34 via an oil passage 39. . When the solenoid valve 40 receives a duty signal corresponding to each running condition from the control unit 50, it drains the hydraulic pressure to generate a duty pressure Pc, and operates the clutch control valve 34 according to this duty pressure Pc. Then, the center differential limiting torque Tc of the center clutch 27 is variably controlled. Similarly, the rear clutch hydraulic pressure control means 32'
, 39', and a clutch control valve 34' communicating with the oil passage 35'.
, a solenoid valve 40', and is configured to variably control the rear differential limiting torque Td of the rear clutch 28 based on the duty pressure Pd of the solenoid valve 40'.

Referring to FIG. 1, an electronic control system for front and rear wheel torque distribution and left and right rear wheel torque distribution will be described. First, to explain the basic control principle, we will explain the handling stability on high μ roads,
In order to prevent a vehicle from spinning on a low μ road and obtain stability, it is necessary to control torque distribution so that good steering characteristics, that is, a stability factor, are always constant on various road surfaces and driving conditions. Here, when accelerating while turning in four-wheel drive, the front tire's cornering power and lateral force decrease due to the front-rear load shift, and conversely, the rear tire's cornering power and lateral force increase. As a result, the vehicle exhibits strong understeer behavior. On the other hand, the relationship between driving force and lateral force in tires is set by a friction circle determined by road surface μ and load, and as driving force increases, lateral force decreases, which affects steering characteristics. Therefore, in the above case, control may be performed to increase the driving force of the rear wheels and reduce their lateral force to cancel the increase in the cornering power of the rear tires due to the front-rear load shift.

On the other hand, the driving force state of the left and right rear wheels generates a yaw moment on the vehicle, which directly affects the steering characteristics. In other words, in this case, the state of the left and right rear wheel driving force differs depending on the magnitude of the input driving force, and when the driving force is small or when engine braking is applied, the braking force acts on the outer wheel and the driving force applies to the inner wheel. In this state, a moment is applied in the direction of understeer. Furthermore, when the driving force is large, the driving force is generated in accordance with the ground loads of the inner and outer wheels, so a moment in the oversteer direction is added. Therefore, the torque distribution between the left and right rear wheels may be controlled to maintain better steering characteristics depending on the driving condition.

To specifically implement the center differential control described above, first, when accelerating, the target steering characteristic is set in the direction of weak understeer with reference to the neutral point, and conversely, when decelerating, the target steer characteristic is set in the direction of strong understeer. . In addition, the equation of motion of the vehicle is extended to the nonlinear region of low-μ roads by taking into account the influence of driving force, and the relationship between the stability limit and stability factor in steady circular turning is determined, and equivalent cornering is Set the power. On the other hand, as an introduction to tire nonlinearity, the relationship between cornering force and sideslip angle is approximated nonlinearly, and the concept of a friction circle is adopted for the longitudinal and lateral forces generated in the tire. Find the equivalent cornering power of the front and rear wheels using lateral G. In this case, the road surface μ is set as a virtual road surface μ calculated from the square root of the sum of the squares of longitudinal G and lateral G as a substitute for the actual road surface μ. Since this virtual road surface μ is always smaller than the actual value, it is convenient to be controlled to the stable side, and in the limit range of a low μ road, it approaches the actual value.

[0018] As described above, when the turning acceleration/deceleration motion that causes large longitudinal G and lateral G is expanded to the nonlinear range and modeled, the cornering power of the front and rear wheels is replaced by the equivalent cornering power expressed by a predetermined equation based on the equation of motion. It was found that it can be analyzed using As a result, when the front and rear torque distribution ratio is changed, the equivalent cornering power of the front and rear wheels changes, resulting in a change in the steering characteristics of the vehicle.The equivalent cornering power that takes this nonlinearity into consideration is It turns out that it is a function of In this way, the front and rear wheel torque distribution ratio α controlled by the center differential is calculated by the following formula.

[0019]

[Math 3]

[Math 4]

[Math 5]

[Math 6]

[Formula 7] In the above equation, Gx: longitudinal G, Gy: lateral G, μ: virtual road surface, A: target stability factor, W: vehicle weight, h: center of gravity height, L: wheel base, Lf, Lr :
Distance between center of gravity and axle, Kfo, Kro: equivalent cornering power, Kfc, Krc: load dependence of cornering power obtained by partially differentiating Kf, Kr with respect to W.

Next, to explain the case in which the upper rear differential control is specifically implemented, when the driving force is small, the moment M1 in the understeer direction is proportional to the yaw rate as follows, assuming that the understeer direction is negative. It is calculated in inverse proportion to vehicle speed.

[0021]

[Formula 8] In the above equation, Kx: driving (braking) stiffness of the tire, d: tread, V: vehicle speed, and γ: vehicle yaw rate.

Further, the moment M2 in the oversteer direction when the driving force is large is calculated in proportion to the driving force to the rear wheels and the longitudinal G as follows.

[Math. 9] In the above formula, Kφf, Kφr: roll steel properties.

Therefore, the overall yaw moment M3 acting on the vehicle is the sum of the above two equations, and is as follows.

[Math. 10]

Therefore, this yaw moment M3 is substituted into the equation of motion for steady circular turning expanded to a nonlinear region, expressed using a stability factor, and further rewritten as rear differential limiting torque. Thereby, the rear differential limiting torque Td is calculated using the following formula.

[Formula 11] In the above formula, Rt is the tire effective diameter. The sign is positive when Gx≧0 and negative when Gx<0.

Therefore, based on the above basic control principle, a longitudinal G sensor 43 for detecting the longitudinal G of the longitudinal acceleration of the vehicle and a lateral G sensor 44 for detecting the lateral G of the lateral acceleration are provided as input information. In order to estimate the input torque of the center differential 20, an engine rotation speed sensor 45,
It has an accelerator opening sensor 46 and a gear position sensor 47,
Furthermore, a signal from the ABS control unit 48 is input.

The control unit 50 has a virtual road surface μ setting section 51 into which the longitudinal G and lateral G are input, and the virtual road surface μ as a substitute for the road surface μ is set by the square root of the sum of the squares of the longitudinal G and the lateral G. do. It also has a target steering characteristic setting section 52 into which the longitudinal G is input, and according to the map in FIG. Set a large stability factor A2. Furthermore, it has an input torque estimating section 53 which receives the engine rotation speed N, accelerator opening degree φ, and gear position P, and estimates the engine output Te from the engine rotation speed N and the accelerator opening degree φ with reference to the engine output characteristics. , the center differential input torque Ti is calculated by multiplying this engine output Te by the gear ratio g of the gear position P.

These longitudinal G, lateral G, stability factor A during acceleration/deceleration, and virtual road surface μ are input to the target longitudinal torque distribution ratio calculating section 54, and the above-mentioned equations (3) to (7) are calculated. is used to calculate the front-rear torque distribution ratio α. This front-rear torque distribution ratio α and center differential input torque Ti are input to the center differential limit torque calculation unit 55,
The center differential limiting torque Tc is calculated as follows. In other words, if the front/rear torque distribution ratio α is set between 0 for RWD and 1 for FWD, and the reference torque distribution ratio Di is set to bias the rear wheels as in the embodiment, the center differential limit torque Tc is calculated by Tc=(α-Di)Ti. Here, if the calculated value is negative, the value of the center differential limiting torque Tc is set to zero. Incidentally, when the reference torque distribution ratio Di is set to bias the weight on the front wheels, subtraction may be performed in the opposite manner to the above. Then, this torque signal is input to the duty ratio converter 56 and converted to a predetermined duty ratio D, and this duty signal is output to the solenoid valve 40. In addition, the longitudinal G, lateral G, stability factor A during acceleration/deceleration, and virtual road surface μ are determined by the rear differential limit torque calculation unit 57.
The rear differential limiting torque Td is calculated using the above-mentioned formula (11). This torque signal is also input to the duty ratio converter 58 and converted, and a predetermined duty signal is applied to the solenoid valve. 40'. Furthermore, A
The signal from the BS control unit 48 is input to each differential limit torque calculating section 55, 57, and when the ABS control signal is input, each differential limit torque Tc, Td is forcibly set to 0.

To further explain the correction measures, the above control system includes a vehicle speed sensor 60 that detects the vehicle speed V, and a steering angle sensor 61 that detects the steering angle when the steering wheel is operated. In addition, in the control unit 50, the vehicle speed V, the longitudinal G
, a yaw moment generation direction determining section 62 to which the front-rear torque distribution ratio α is input. Here, the yaw moment M3 can be directly calculated using the above-mentioned equation (10) using the above-mentioned parameters and vehicle specifications, with the understeer direction set as a negative value. For this reason, the yaw moment generation direction determination unit 62
calculates the value of the yaw moment M3, and if the sign is positive, determines an oversteer direction in which the torque is transferred from the inner wheel to the outer wheel, and if it is negative, determines an understeer direction in which the torque is transferred from the opposite outer wheel to the inner wheel. Then, this determination signal is input to the correction unit 63 on the output side of the rear differential limiting torque calculating means 57, and when the longitudinal G is positive, that is, in an acceleration state, and in the understeer direction, the rear differential limiting torque Td is set to zero. stipulated in The control unit 50 also includes a small steering determination section 64 into which the steering angle is input, and determines a small steering when the steering angle is in a straight-ahead region of about ±20 degrees, for example. This determination signal is input to the correction section 65 on the output side of the center differential limit torque calculation section 55,
At the time of small steering, the center differential limiting torque Tc is corrected to decrease at a predetermined rate.

Next, the operation of this embodiment will be explained. First, when the vehicle is running, power from the engine 1 is input to the transmission 3 via the clutch 2, and shifting power is input to the first sun gear 21 of the center differential 20. Here, since the reference torque distribution et is set to be biased toward the rear wheels according to the specifications of each gear of the center differential 20, the power is distributed to the carrier 24 and the second sun gear 22 according to this torque distribution, and the power is output. At this time, the center clutch 27
If it is released, the power is further transmitted to the front and rear wheels using the above standard torque distribution et, resulting in FR-like power performance even though it is a 4-wheel drive, and the center differential 20 becomes free and the rotation of the front and rear wheels is increased. It becomes possible to turn freely while absorbing the difference. Further, the hydraulic control means 32 applies the center differential limiting torque Tc to the center clutch 27.
occurs, the second sun gear 22 changes according to this torque Tc.
The torque is further transferred between the carrier 24 and the carrier 24,
Torque distribution is variably controlled from rear-wheel biased weight to front-wheel biased torque distribution according to the vehicle weight distribution when directly coupled, preventing slipping of the front wheels or rear wheels, and effectively transmitting power by limiting the differential of the center differential 20. Escape, running ability, stability, etc. will improve.

The power that is distributed to the rear wheels through torque distribution by the center differential 20 and center clutch 27 is input to the rear differential 11, and the torque is further distributed to the left and right rear wheels 13L and 13R by the rear differential 11 and rear clutch 28. Controlled and transmitted. That is, when the rear clutch 28 is released, the rear differential 11 becomes free, and the torque is equally distributed according to the gear specifications. Additionally, the hydraulic control means 32
'When a rear differential limiting torque Td is generated in the rear clutch 28, power is effectively transmitted to the grip wheels by limiting the differential of the rear differential 11, and torque is transferred from high-speed wheels to low-speed wheels according to the rear differential limiting torque Td. When the differential lock is directly connected, the torque is distributed unequally according to the vehicle weight distribution between the left and right rear wheels 13L and 13R.

On the other hand, when driving in the four-wheel drive mode, each sensor signal is input to the control unit 50 of the electronic control system, and the virtual road surface μ is determined according to the longitudinal G and lateral G, and the stability factor A is determined according to the longitudinal G. The front and rear torque distribution ratio α is calculated from these, and the center differential limiting torque Tc is calculated from this front and rear torque distribution ratio α and the input torque Ti. And center differential limiting torque T
A duty signal corresponding to c is output to the hydraulic control means 32, and feedforward control is performed so that the same center differential limiting torque Tc is generated in the center clutch 27. Therefore, the control characteristics of the front-rear torque distribution in this case are shown in FIG. 4.

In other words, when driving straight ahead with small lateral G in an acceleration range where longitudinal G is positive, torque is distributed toward the front wheels and the FF
At this time, when accelerating on a high μ road with large front and rear G, the torque distribution is approximately equal, and the four wheels effectively exhibit driving performance and stability. Additionally, as the lateral G gradually increases during corner acceleration, the torque is distributed closer to the rear wheels and becomes FR-like, thereby canceling out the strong understeer and ensuring that the steering characteristics remain in a good constant state. Ru. Furthermore, when driving on a low μ road where longitudinal G is small, the torque is distributed approximately equally or toward the front wheels, thereby preventing slippage. During deceleration driving where the longitudinal G is negative, the engine is always in the FF or direct connection state, and the engine brake is effectively applied, reducing tuck-in when the accelerator is off.

On the other hand, in the center differential control described above, the center differential limiting torque Tc is further corrected in accordance with the actual steering state based on the steering angle. That is, FIG.
The characteristics of the parameter are virtual road surface μ using longitudinal G and lateral G as a substitute for road surface μ. Therefore, under conditions of small turning on a high μ road with small longitudinal G and lateral G, the virtual road surface μ
is relatively significantly lower than the actual value, and as a result, the front and rear G
When is near zero, the center differential limiting torque Tc is controlled to a larger value on the safe side than necessary, and torque fluctuation becomes large. Therefore, in the case of small steering, the center differential limiting torque Tc for acceleration and deceleration is corrected to decrease as shown by the broken line in FIG. 4, and the above-mentioned problem caused by the virtual road surface μ is corrected. For this reason,
In straight-line slow acceleration/deceleration running with small steering and small longitudinal and lateral G, the center differential limiting torque Tc is kept low and no internal circulation torque or the like occurs. Furthermore, when the accelerator is turned on or off, the amount of torque transferred to the front wheels is reduced, and fluctuations in the front wheel torque are also reduced.

The control unit 50 also has longitudinal G, lateral G.
, the virtual road surface μ, and the stability factor A, the rear differential limiting torque Td is directly calculated, and the rear clutch 28 is similarly feedforward controlled. The control characteristics of left and right rear wheel torque distribution in this case are shown below.
The result will be as shown in Figure 5. That is, by controlling the rear differential limiting torque Td to a large extent in accordance with the increase in the driving force of the front and rear G,
Understeer is reduced by generating a moment in the oversteer direction. At this time, idling of the rear inner wheel is prevented and traction performance is also improved. On a low μ road with small longitudinal G, the rear differential limiting torque Td is the minimum, the rear differential 11 becomes almost free, and the left and right rear wheels 13L
, 13R are prevented from spinning due to simultaneous slip. Furthermore, during deceleration, the rear differential limiting torque Td is controlled to increase in accordance with the increase in deceleration and lateral G, so that a moment in the understeer direction is generated in accordance with this torque, thereby preventing tuck-in.

On the other hand, in the rear differential control described above, a yaw moment M3 is generated in the vehicle depending on the state of the driving force and the rear differential limiting torque Td, and this yaw moment M3 and the direction of its generation are determined by the vehicle speed V, longitudinal G, and longitudinal torque. The determination is made by actually calculating the distribution ratio α. Therefore, according to equation (10), if the vehicle speed V or longitudinal G is large, the direction will be oversteer, but under driving conditions of slow acceleration with small longitudinal G and low speed, the influence of equation (8) becomes large, and at this time, the direction of oversteer will shift from the outer wheel to the inner wheel. Torque may shift and result in understeer. Then, in this case, the rear differential limiting torque Td is corrected to zero as shown by the broken line in Fig. 5, and the above-mentioned torque shift no longer occurs.In this way, even during slow acceleration, the desired steering characteristics can be obtained. It will be done. In addition, when decelerating with a negative longitudinal G, the yaw moment M3 is calculated using equation (10).
is always in the negative understeer direction, promoting the above-mentioned tuck-in prevention effect.

Although the embodiments of the present invention have been described above, the drive system for torque distribution control can be similarly applied to other systems, and the control system is not limited to this.

[0037]

[Effects of the Invention] As explained above, according to the present invention, the target steering characteristics can always be achieved by using the vehicle motion equation extended to the non-linear region with longitudinal G, lateral G, road surface μ, and target steering characteristics as parameters. In the variable torque distribution control for a four-wheel drive vehicle that is controlled in this way, correction is made so that yaw moment in the understeer direction does not occur during slow acceleration, making it easier to achieve the target steering characteristics under these driving conditions. In addition, in the case of acceleration/deceleration with small steering changes, unnecessary center differential limiting torque is suppressed, reducing fluctuations in front wheel torque when the accelerator is on and off, eliminating unpleasant steering force fluctuations for the driver. . Furthermore, in this case, internal circulation torque is reduced, resulting in improved fuel efficiency.

[Brief explanation of the drawing]

FIG. 1 is a block diagram showing an electronic control system of an embodiment of a torque distribution control device for a four-wheel drive vehicle according to the present invention.

FIG. 2 is a configuration diagram showing the configuration of a drive system and a hydraulic control system of a four-wheel drive vehicle to which the present invention is applied.

FIG. 3 is a diagram showing target steering characteristics.

FIG. 4 is a diagram showing control characteristics of front and rear wheel torque distribution according to the present invention.

FIG. 5 is a diagram showing control characteristics of left and right rear wheel torque distribution according to the present invention.

[Explanation of symbols]

11 Rear differential 20 Center differential 27 Center clutch 28 Rear clutch 32, 32' Hydraulic control means 50 Control unit 60 Vehicle speed sensor 61 Rudder angle sensor 62 Yaw moment generation direction determining section 63, 65
Correction section 64 Small steering determination section

Claims (2)

[Claims] Claim 1. A control system that variably controls torque distribution between front and rear wheels and left and right rear wheels so as to extend to a nonlinear range and obtain constant steering characteristics based on longitudinal G, lateral G, road surface μ, and target steering characteristics, Yaw moment generation direction determination means for determining the direction of yaw moment generation using vehicle speed, longitudinal G, and longitudinal torque distribution ratio as parameters, and a rear differential limiting torque that is set to zero when yaw moment is generated in the understeer direction under slow acceleration conditions. A torque distribution control device for a four-wheel drive vehicle, comprising a correction means for correcting the torque distribution. 2. A control system that variably controls the torque distribution between the front and rear wheels and the left and right rear wheels so as to extend to a non-linear range and obtain constant steering characteristics based on longitudinal G, lateral G, road surface μ, and target steering characteristics, A four-wheel vehicle characterized by comprising: a small steering determination means for determining a small steering operation based on a steering angle; and a correction means for reducing and correcting center differential limit torque during acceleration/deceleration in the case of a small steering operation. Drive vehicle torque distribution control device.