JP3424456B2 - Road surface friction coefficient estimation device for vehicles - Google Patents

Abstract

PROBLEM TO BE SOLVED: To enhance the reliability of road surface state estimation by calculating an index unitarily showing a road surface state from parameters showing the irregular state of the road surface and the easiness of slippage of the road surface. SOLUTION: In steady traveling, an irregular state parameter arithmetic part 211 calculates the maximum amplitude of slip difference ratio in a fixed time from the difference in average speed between front and rear wheels or the difference in speed between front and rear wheels. An easiness-of-slippage parameter arithmetic part 212 calculates the average value from the frictional coefficient (hereinafter referred to as μ) judged value of a road surface. On both the parameters, a road surface index calculating means 220 calculate an index corresponding to each state of other parameter for unitarily showing the road state, or high μ degree, middle μdegree and low μdegree by use of fuzzy inference. A road surface frictional efficient calculating means 230 continuously determines the index calculated by the means 220 and accumulatively evaluates it to calculate the road surface state in steady traveling, or road surface μ. Thus, the road surface μ can be precisely grasped to more precisely perform the traveling control.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a device for estimating the friction coefficient of a road surface on which a vehicle is traveling, and more particularly, by estimating the road surface friction coefficient according to the traveling state of the vehicle, the estimation accuracy is improved. The present invention relates to a vehicle road surface friction coefficient estimating device.

[0002]

2. Description of the Related Art In recent years, various traveling controls have been developed in order to improve the traveling performance of automobiles and to assist artificial operations related to traveling. Examples of such traveling control include engine output control called traction control, differential limitation control between the left and right wheels and front and rear wheels, and torque distribution control (power transmission between the left and right wheels and front and rear wheels). Control) etc. have already been developed.

For example, a differential mechanism is provided between the left and right wheels, which are the driving wheels of an automobile, to allow a differential that occurs during turning. In this differential mechanism, one of the left and right wheels is provided. If the wheel of the wheel slips in the sand, for example, only one wheel of the wheel rotates and the other wheel hardly rotates, which may result in a state where the driving torque cannot be transmitted to the road surface.

Therefore, in such a case, a differential limiting mechanism (LSD = Limited Slip Differential) has been developed which can limit the differential. Such left and right wheel differential limiting mechanisms include a type that is proportional to the rotational speed difference between the left and right wheels and a type that is proportional to the input torque. For the right and left wheel rotation speed difference proportional type, VC that utilizes the viscosity of the liquid
(Viscous coupling) type LSD etc.,
There is an advantage that the running stability of the vehicle can be improved. On the other hand, the input torque proportional type includes mechanical type such as friction type such as general LOM (lock automatic) type LSD, which has an advantage that the turning performance of the vehicle can be improved.

However, in the various differential limiting mechanisms described above, the differential control characteristics are determined by the physical properties and the like, and the differential control characteristics can always be adjusted so that the differential control can always be appropriately performed. It's not. There is also a so-called electronically controlled LSD system in which the LSD is electronically controlled, but even in such a system, the torque transfer between the wheels is limited to only from the high speed side to the low speed side. It is thought that it is not possible to sufficiently improve the running performance while the vehicle is turning.

[0006] Therefore, the applicant of the present invention discloses, for example, in Japanese Patent Laid-Open No. 5-131855, so that the torque can be distributed between the left and right wheels in various running states of the vehicle without causing a large torque loss and energy loss. Kaihei 7-108840
No. 7, JP-A-7-108841, 7-108842,
There has been proposed a vehicle-to-right wheel torque control device for left and right wheels as disclosed in JP-A-7-108843 and JP-A-7-156681.

In this torque transfer control device between the left and right wheels, when two rotating bodies arranged coaxially are brought into sliding contact with each other in a state where the rotating speeds are different from each other, the rotating body having the higher rotating speed has a lower rotating speed. This utilizes the characteristic that torque is transmitted to the other rotating body. That is, for example, this device is configured to change the rotational speed input to the differential device or the rotational speed of one of the wheel shafts to a high speed and a low speed, and output the same, and a differential mechanism that receives respective outputs of the speed changing mechanism. And a plurality of speed change interlocking members that rotate at a rotational speed different from that of the moving device or one of the wheel shafts, constant speed interlocking members that rotate at the same speed as the other wheel shaft of the left and right wheels, and these speed change interlocking members, etc. It is provided with a plurality of torque transmission couplings such as a wet multi-plate clutch provided between the speed interlocking member and the like.

In such a device, even if the left and right wheels are rotating at a constant speed, the torque transmission coupling is
Since the rotation speeds of the speed change interlocking member side and the constant speed interlocking member side are different, if the torque transmission coupling is operated by engaging the wet multi-plate clutch, the speed change interlocking member side and the constant speed interlocking member side are The torque is transmitted from the higher speed to the lower speed. If the variable speed by the speed change mechanism is set to be larger than a certain value, torque can be transmitted from the inner wheel side having a low rotation speed to the outer wheel side having a high rotation speed during turning.

Further, for example, in a torque transmission coupling such as a wet multi-plate clutch, transmission torque to one wheel shaft is controlled by performing switching of engagement of each wet multi-plate clutch and control of the degree of engagement. Can be increased or decreased, or the torque transmitted to the other wheel shaft can be increased or decreased. Therefore, since the transmission torque capacity can be variably controlled, the torque can be transmitted in the desired direction in the left and right wheels at the desired transmission torque capacity.

Such a device can be applied regardless of whether the left and right wheels are drive wheels or driven wheels. If the left and right wheels are drive wheels, the distribution of the driving force from the engine to the left and right wheels is adjusted. When the left and right wheels are driven wheels, torque transmission causes the wheels on the torque transmission side to receive the driving force, and the wheels on the torque transmission side to receive the braking force.

In any case, the magnitude of the driving force or the braking force exerted between each of the left and right wheels and the road surface is unbalanced so that the yaw moment is generated in the vehicle and the behavior of the vehicle is changed. Can be controlled. Further, such torque transmission control can be considered not only between the left and right wheels but also between the front and rear wheels.

[0012]

By the way, in the above-described device for performing torque transmission control (power transmission control) between the left and right wheels or between the front and rear wheels of the vehicle, the torque transmission control corresponding to the running condition of the vehicle is performed. Need to do. That is,
Torque transmission control is performed according to whether the vehicle is traveling straight or turning, accelerating or decelerating, and whether or not each wheel has slips, etc., to improve the running performance of the vehicle. .

However, in order to perform more appropriate control, it is necessary to grasp the road surface condition on which the vehicle is traveling. In other words, it is considered that the control effect of the torque transmission control is different depending on whether or not the traveling road surface is slippery. It is necessary to perform transmission control (power transmission control).

Such a road surface condition can be expressed as a so-called road surface friction coefficient (road surface μ), but the road surface condition (road surface friction coefficient) can be grasped only by the torque transmission control (power transmission control) as described above. Instead, it becomes effective in various traveling controls in automobiles. The present invention has been devised in view of the above-mentioned problems, and is capable of accurately grasping a road surface condition (road surface friction coefficient) on which a vehicle is traveling, and thus can more appropriately perform various traveling controls in an automobile, for example. An object of the present invention is to provide a vehicle road surface friction coefficient estimating device.

[0015]

Therefore, the vehicle road surface friction coefficient estimating apparatus according to the present invention according to claim 1 detects the traveling state of the vehicle, and the vehicle state detecting means detects the traveling state of the vehicle. A steady-state parameter calculating means for calculating a parameter indicating the unevenness state of the road surface and a parameter indicating the slipperiness from the running state of the vehicle during steady running, and a unit for expressing the road surface state from the values of the above two parameters and the road index calculating means for calculating an index corresponding to each state of the parameter, respectively, calculate the road friction coefficient by accumulated evaluation seek cumulatively to continue each index calculated by the road surface index calculation means Road surface friction coefficient calculating means for

Furthermore, preferably, the cumulative value of the above-mentioned index
The road friction coefficient is
To calculate.

Preferably, the above vehicle is provided in association with the left and right wheels of the front wheels or the rear wheels, and transmits the driving force from the power unit to the left and right wheels or rotates the left and right wheels. A rotational propulsion force distribution adjusting mechanism that transfers the accompanying rotational force between the left and right wheels to change the rotational propulsion force of the left and right wheels, and a left drive wheel and a right drive on the side provided with the rotational propulsion force distribution adjustment mechanism. Left wheel side rotational speed detecting means and right wheel side rotational speed detecting means for detecting the rotational speed of the wheels, and left wheel for detecting the rotational speeds of the left driving wheel and the right driving wheel on the side where the rotational propulsive force distribution adjusting mechanism cannot be provided. It has a side rotation speed detection means and a right wheel side rotation speed detection means, and a reference rotation speed difference calculation means for calculating the theoretical value of the left and right rotation speed difference between the front wheel and the rear wheel based on the steering angle and the vehicle speed. And the road surface condition The parameters shown are the detection information from the left wheel side rotational speed detecting means and the right wheel side rotational speed detecting means on the side provided with the rotational propulsive force distribution adjusting mechanism, and the information not detected on the side not provided with the rotational propulsive force distribution adjusting mechanism. The vehicle rotational speed difference (parameter indicating the unevenness state) calculated based on the detection information from the left wheel side rotational speed detection means and the right wheel side rotational speed detection means, and the rotational propulsive force distribution adjustment mechanism are provided. Information from the left wheel side rotation speed detection means and the right wheel side rotation speed detection means, and the left wheel side rotation speed detection means and the right wheel side rotation speed detection means that are not provided with the rotation propulsion force distribution adjusting mechanism. Torque movement (power movement) of the drive wheels of the vehicle calculated based on the detection information from the reference rotation speed difference, the calculation information from the reference rotation speed difference calculation means, and the operating state of the rotation propulsion force distribution adjusting mechanism. According road μ determination value is (parameter indicating the slipperiness).

Further, it is preferable that the index corresponding to each state of the other parameters mentioned above indicates the occurrence degree of each state. Furthermore, the road surface friction coefficient estimation condition (road surface μ
(Determination condition) is set, and the road surface friction coefficient (road surface μ) is updated periodically when the road surface μ determination condition is satisfied, and when the road surface μ determination condition is not satisfied, the road surface friction coefficient ( As the road surface μ), it is preferable to adopt the last (nearest) road surface friction coefficient (road surface μ).

A vehicle road surface according to a fourth aspect of the present invention.
The friction coefficient estimation device is a vehicle-like device that detects the running state of the vehicle.
State detection means and the vehicle detected by the vehicle state detection means
Showing the road surface condition from the running condition during steady running of the vehicle
Steady-state parameter calculating means for calculating the parameters of
Other than steady running of the vehicle detected by the vehicle state detection means
The second indicating the road surface condition from the traveling condition during the specific traveling
A specific parameter calculating means for calculating parameters,
The first parameter calculated by the ordinary parameter calculation means
Data continuously and cumulatively
Of the accumulated first parameter and the second parameter of
Road friction coefficient calculation means for calculating road friction coefficient comprehensively
It is characterized by having and. Good Mashiku is the in the specific running an operation and in a non-linear traveling start running of the vehicle. Further, it is preferable that the specific parameter calculating means sets the second parameter so that the following road surface friction coefficient is obtained when the vehicle starts.

That is, the second parameter is set so as to obtain a low road surface friction coefficient even when one wheel slips when the vehicle starts straight ahead. Further, when the wheels do not slip at a certain acceleration or more, the second parameter is set so that the road surface friction coefficient becomes high. When the vibration component of the wheel is large, the second parameter is set so that the road surface friction coefficient is moderate or low.

Further, whether or not the road surface is in the μ-split state is determined from the torque moving direction and the speed difference between the left and right rear wheels. In the μ-split state, the second parameter is set so that the road surface friction coefficient becomes low. To set. Further, it is preferable that the specific parameter calculating means sets the second parameter such that the following road surface friction coefficient is obtained in a specific traveling state such as starting traveling and non-linear traveling.

That is, when the degree of slippage of the tire (wheel) is large and the lateral acceleration generated in the vehicle is equal to or less than the first predetermined value, it is determined that the road surface has a low road surface friction coefficient (low μ road),
The second parameter is set so that the road surface friction coefficient is low. When the lateral acceleration generated in the vehicle is larger than the second predetermined value, the road surface having a high road surface friction coefficient (high μ
Road) and sets the second parameter so that the road surface friction coefficient is high.

[0023]

〔table of contents〕

1. 1. System overview of this device 1.1 Concept of hardware configuration of this device 1.2 Hardware configuration of this device 1.3 Overview of control of this device 1.4 Actions and effects to be obtained by control of this device 2. Control contents of this device 2.1 Input calculation process 2.2 Drift judgment logic 2.3 Vehicle motion control logic 2.3.1 Target ΔN tracking control 2.3.2 Acceleration turning control 2.3.3 Tack-in correspondence control 2 3.3.4 Steering transient response control 2.4 Road surface μ estimation 2.4.1 Road surface μ estimation during steady turning 2.4.2 Road surface μ estimation during non-linear turning 2.4.3 Road surface μ estimation during starting 2 4.4.4 Output value setting 2.5 Actuator drive 2.5.3 Indicator display control 2.5.4 Output value setting 3. Operation of this device and effect of this device 3.1 Operation of this device 3.2 Effect of this device 3.2.2 Effect of vehicle road surface friction coefficient estimation device 3.2.3 Vehicle road surface friction coefficient (road surface μ ) Effect of adaptive control [Embodiment] 1. System Overview of This Device 1.1 Concept of Hardware Configuration of This Device First, the concept of the hardware configuration of this device will be explained.
The power transmission control device for the left and right wheels for a vehicle according to the present invention, when two rotating bodies coaxially arranged are brought into sliding contact with each other in a state where the rotating speeds are different from each other, a rotating body having a higher rotating speed is rotated to a rotating body having a lower rotating speed. The characteristic is that torque is transmitted to the rotating body.

As described above, if the left wheel side is larger than the right wheel side between the left wheel side and the right wheel side, torque can be easily transmitted from the left wheel side to the right wheel side. If the right wheel side is larger than the left wheel side, torque can be easily transmitted from the right wheel side to the left wheel side. Therefore, in order to realize a state in which the left wheel side is larger than the right wheel side even in the region where the left and right wheels originally rotate at a constant speed, for example, a shift that shifts the left wheel side rotation speed VL to the left wheel side at a high speed. If the mechanism is provided, even if the left and right wheels rotate at a constant speed, the left wheel side member is provided between the left wheel side member that receives the output of this transmission and the right wheel side member that rotates at the same speed VR as the right wheel. It is possible to realize a state in which the rotation speed of is higher than the right wheel side. Also,
For example, if a transmission mechanism for changing the rotation speed VR on the right wheel side to a low speed is provided on the right wheel side, even if the left and right wheels rotate at a constant speed,
A state in which the rotational speed on the left wheel side is higher than that on the right wheel side can be realized between the left wheel side member that rotates at the same speed VL as the left wheel and the right wheel side member that receives the output of the transmission mechanism.

Also on the right wheel side, if it is constructed symmetrically with this, it is possible to always realize a state in which the right wheel side is larger than the left wheel side. When the vehicle turns, the inner turning wheel rotates at a lower speed than the outer turning wheel.However, depending on the gear ratio setting of the speed change mechanism, even when the vehicle turns, the rotating member on the inner wheel side has a higher speed than the rotating member on the outer wheel side. You can shift to.

If a torque transmission coupling is provided between the left wheel-side rotating member and the right wheel-side rotating member that have been given such a speed difference, the torque transmitting coupling can be made to operate appropriately, Under constant traveling conditions, torque can be constantly transmitted from the left wheel side to the right wheel side and from the right wheel side to the left wheel side. Of course, even when turning at the maximum steering angle, the driving torque on the inner wheel side is transmitted to the outer wheel side,
If you set the gear ratio by the speed change mechanism, under all driving conditions,
Torque can be constantly transmitted from the left wheel side to the right wheel side and from the right wheel side to the left wheel side.

In the case of a variable torque transmission capacity type coupling such as a wet multi-plate clutch mechanism, the amount of transmission torque can be adjusted according to the engagement pressure (pressing force P).
By the way, the transmission mechanism and the coupling interposed between the right wheel side and the left wheel side are directly provided between the right wheel side and the left wheel side. The transmission mechanism and the coupling are provided between the wheel side (right wheel side or left wheel side), and the power is transmitted between the left wheel side and the right wheel side via the differential input portion ( Torque transfer) may be realized.

The power transmission (torque movement) between the left and right wheels of the vehicle based on such a principle can be applied regardless of whether the left and right wheels are drive wheels or driven wheels.
The distribution of the driving force from the engine to the left and right wheels will be adjusted, and if the left and right wheels are driven wheels, the torque-transmitted wheel will receive the driving force due to the torque transmission. The wheel on the side of doing the braking will receive the braking force. In any case, the magnitude of the driving force or braking force exerted between each of the left and right wheels and the road surface is controlled to be unbalanced in the left-right direction, thereby generating a yaw moment in the vehicle and changing the behavior of the vehicle. Can be controlled.

1.2 Hardware Configuration of the Present Device Next, the hardware configuration of the power transmission control device for the left and right wheels for the present vehicle based on the above theory will be described with reference to FIGS. 1.2.1 Configuration of the power transmission system of the vehicle according to this device
As shown in, it is provided on the rear wheels of a four-wheel drive vehicle.

In FIG. 1, reference numeral 2 is an engine,
The output of the engine 2 is transmitted to a differential gear mechanism (= center differential, hereinafter referred to as center differential) 8 via a transmission 4 and an intermediate gear mechanism 6. On the one hand, the output of the center differential 8 is transmitted through a differential gear mechanism (= front differential, hereinafter referred to as front differential) 10 for the front wheels to the axle 1
2L, 12R is transmitted to the left and right front wheels 14, 16 and, on the other hand, the bevel gear mechanism 18, the propeller shaft 20.
The transmission is transmitted from the axles 26L and 26R to the left and right rear wheels 28 and 30 via a bevel gear mechanism 22 and a rear wheel differential gear device (= rear differential, hereinafter referred to as rear differential) 24. A rotational propulsive force distribution adjusting mechanism (or rotational force adjusting means, hereinafter referred to as a torque moving mechanism) 50 of the power transmission control device between the left and right wheels is provided in the rear differential 24.

The center differential 8 is composed of differential pinions 8A and 8B and side gears 8C and 8D which mesh with these differential pinions 8A and 8B, as in the conventional well-known type, and the rotation input from the differential pinions 8A and 8B. Torque is side gear 8
It is transmitted to C and 8D, and from the side gear 8C to the front wheel side,
The side gear 8D is transmitted to the rear wheel side while allowing each differential.

Here, from the side gear 8C to the front differential 10 on the front wheel side through the front wheel output shaft 32, from the side gear 8D, the rear wheel output shaft 34 and the bevel gear mechanism 1 are provided.
Torque is transmitted from the propeller shaft 20 to the rear wheel side via 8. The center differential 8 distributes the engine output torque (rotational propulsion force) between the front wheel side and the rear wheel side by restricting (or limiting) the differential between the front wheel side output section and the rear wheel side output section. A viscous coupling unit (VCU) 36 is attached as a controllable differential limiting means [namely, limited slip differential (LSD)].

The VCU 36 is interposed between the front wheel output shaft 32 and the rear wheel output shaft 34, and limits the differential between the front wheel side and the rear wheel side with a force corresponding to the differential state. by doing,
It is possible to avoid a situation in which only the light load side of the front and rear wheels spins and the rotational torque is not transmitted to the heavy load side. 1.2.2 Configuration of Rotational Propulsion Force Distribution Adjustment Mechanism of the Present Device By the way, the present left-right wheel power transmission control device includes a rotation propulsion force distribution adjustment mechanism (torque moving mechanism) 50 provided in the differential carrier 51, The control unit (or rotational propulsion force distribution control unit) is composed of a hydraulic unit 38 and an electronic control unit (hereinafter referred to as ECU) 42. Here, the rear differential 24 and the rear differential 24 and the axle 26 are provided.
The configuration of the torque transfer mechanism 50 fitted between the L and 26R will be described with reference to FIG.

As shown in FIG. 2, an input shaft 52 is connected to the rear end of the propeller shaft 20, and a drive pinion gear 54 is connected to the input shaft 52 so as to rotate integrally. A crown gear 56 provided on the outer circumference of a differential case (differential case) 58 meshes with the drive pinion gear 54, and the output of the engine is transmitted from the input shaft 52 to the rear differential 24 via the drive pinion gear 54 and the crown gear 56. It is being communicated.

The rear differential 24 is, similarly to the conventional one, a pair of pinions provided in the differential case 58, that is, differential pinions 60A and 60B, and side gears 62 and 64 which mesh with these differential pinions 60A and 60B. And the rotational torque input from the differential pinion 60A, 60B is
It is transmitted to the side gears 62 and 64, and is transmitted from the side gear 62 to the rotary shaft 66 on the left wheel side and from the side gear 64 to the rotary shaft 68 on the right wheel side while allowing respective differentials. . Also, the left and right rotary shafts 66,
As shown in FIG. 1, 68 is connected to the axles 26L and 26R that are connected to the left and right rear wheels 28 and 30, respectively.

The torque moving mechanism 50 of the present embodiment is provided between the differential case 58 of the rear differential 24, which distributes the driving force between the left and right driving wheels of the rear wheels, and the right wheel side rotating shaft 68,
Transmission mechanism 70 and transmission capacity variable control type torque transmission mechanism 90
It is configured to transmit rotational propulsion force between the left wheel side and the right wheel side, that is, power transmission (torque movement) via the differential case 58.

The speed change mechanism 70 has a speed increasing mechanism 70A for increasing the rotational speed of the input portion of the rear differential 24, that is, the differential case 58, and outputting it to one side of the left and right wheels (here, the right wheel side).
It is also referred to as an acceleration / deceleration mechanism because it is integrally provided with a deceleration mechanism 70B that decelerates and outputs to one side (right wheel side). Further, the transmission capacity variable control type torque transmission mechanism 90 uses a wet hydraulic multi-plate clutch mechanism (hereinafter, also referred to as a clutch) capable of adjusting the transmission capacity according to the control oil pressure.
A clutch (left clutch) 90L provided between the output side of the speed reduction mechanism 70B of the speed change mechanism 70 and the right wheel side to transmit torque to the left wheel side, and an output side of the speed increase mechanism 70A of the speed change mechanism 70 and the right wheel. And a clutch (right clutch) 90R that is provided between the side and the side and transmits torque to the right wheel side is integrally formed. Such a transmission capacity variable control type torque transmission mechanism 90 is also referred to as an integral coupling or simply a coupling.

The acceleration / deceleration mechanism 70 will be described. The acceleration / deceleration mechanism 70 includes a hollow intermediate shaft 72 coupled to the differential case 58 so as to rotate integrally with the differential case 58, and a hollow intermediate shaft 74 connected to the right clutch 90R. It is interposed between the hollow intermediate shaft 76 connected to the left clutch 90L. All of the intermediate shafts 72, 74, 76 are hollow shafts, and the intermediate shafts 72, 74 are provided so as to be able to rotate relative to the outer circumference of the right wheel side rotation shaft 68, and the intermediate shaft 76 is the intermediate shaft. 7
The outer circumference of 4 is also equipped so that it can also rotate relative to it.

Gears 78A, 80A, 82A are provided on the intermediate shafts 72, 74, 76, respectively, and a counter shaft 84 is provided on the outer circumference of the intermediate shafts 72, 74, 76. This counter shaft 84
Has a triple gear 86. Triple gear 86
Is composed of gears 78B, 80B, 82B, and gear 7
8B is the gear 78A of the intermediate shaft 72, gear 80B is the gear 80A of the intermediate shaft 74, and gear 82B is the gear 8 of the intermediate shaft 76.
2A meshes with each other.

The speed increasing / decreasing mechanism 70 has such a gear 78.
Intermediate shafts 72, 74, 76 having A, 80A, 82A
, Counter shaft 84, and gears 78B, 80B, 8
It is composed of a triple gear 86 having 2B. The counter shaft 84 includes the intermediate shafts 72, 74, 76.
A plurality of (three in this case) are provided on the outer periphery of the drive pinion 54 with a phase shifted. As a result, although not provided with a ring gear, the gears 78A, 80A and 82A are sun gears, and the gears 78B, 80B and 82B are planetary pinions, and the arrangement is similar to that of a triple planetary gear mechanism.

Each counter shaft 84 is fixed to a wall portion 51A provided on the differential carrier 51.
Therefore, the gears 78B, 80B, 82B only rotate about the counter shaft 84 as an axis. This allows
The intermediate shafts 72, 74, 76 are supported in the radial direction by the gears 78A, 80A, 82A and the gears 78B, 80B, 82.
Through engagement with B, it is also performed by the plurality of counter shafts 84 fixed to the wall portion 51A as described above.

Then, these gears 78A, 80A, 8
If the numbers of teeth of 2A are Z ₁ , Z ₂ and Z ₃ , respectively, Z ₂
The relationship of <Z ₁ <Z ₃ is set. Also, the gear 78
The numbers of teeth of B, 80B and 82B are Z ₄ , Z ₅ and Z _{6 respectively.}
Then, the relation Z ₆ <Z ₄ <Z ₅ is set.

As a result, the transmission mechanism (acceleration / deceleration mechanism) 70
Then, gear 78A, gear 78B, gear 80A, gear 80
The combination of B forms a speed increasing mechanism 70A that accelerates the rotation input to the rear differential 24 and outputs it to the right wheel side. The combination of the gear 78A, the gear 78B, the gear 82A, and the gear 82B provides the input to the rear differential 24. A deceleration mechanism 70B that decelerates the rotation thus generated and outputs the decelerated rotation to the right wheel side is configured.

By the way, the transmission capacity variable control type torque transmission mechanism 90 to which the outputs of the acceleration / deceleration mechanism 70 are input,
That is, the left clutch 90L and the right clutch 90R are shown in FIG.
As shown in, the space is provided in the diff carrier 51 on the right wheel side of the speed increasing / decreasing mechanism 70. These hydraulic multi-plate clutches 90L and 90R are connected to the clutch plates 90AL and 90AR, which are connected to the clutch case 92 so as to rotate integrally with the right wheel side rotating shaft 68, and to rotate integrally with the intermediate shafts 74 and 76. Clutch plates 90BR, 90B
L and two pistons (not shown) that apply hydraulic pressure (clutch pressure) to the respective clutches 90L and 90R are provided, and the drive hydraulic pressures of the two hydraulic pistons are adjusted through the hydraulic unit 38 by electronic control of the controller 42, The engagement state of the clutches 90L and 90R is adjusted.

The left clutch 90L includes a right wheel side clutch plate 90AL which rotates integrally with the right wheel side rotating shaft 68, and an intermediate shaft 76.
And a clutch plate 90BL on the output side of the speed reduction mechanism 70B, which is coupled so as to rotate integrally with. Since the clutch plate 90BL rotates together with the intermediate shaft 76 together with the gear 82A that has been decelerated by the reduction mechanism 70B, the clutch plate 90B does not become large unless the speed ratio of the left wheel to the right wheel becomes large.
L rotates at a lower speed than the right wheel side clutch plate 90AL that rotates integrally with the right wheel side rotation shaft 68.

Therefore, by engaging the clutch 90L, even when the vehicle is turning to the right and the right wheel rotates at a lower speed than the left wheel, the clutch plate 90 from the right wheel side clutch plate 90AL side.
Torque is transmitted to BL, that is, from the right wheel side to the input side of the rear differential, reducing the distribution amount of torque from the engine to the right wheel side and increasing the distribution amount to the left wheel side. Can be made.

The right clutch 90R is connected to a right wheel side clutch plate 90AR which rotates integrally with the right wheel side rotating shaft 68, and a speed increasing mechanism 70A which is connected so as to rotate integrally with the intermediate shaft 74.
Output side clutch plate 90BR. The clutch plate 90BR and the intermediate shaft 74 together with the speed increasing mechanism 70A.
Since it rotates integrally with the gear 80A that has been increased in speed, the clutch plate 90BR rotates at a higher speed than the right wheel clutch plate 90AR that rotates integrally with the right wheel rotation shaft 68 unless the speed ratio of the right wheel to the left wheel increases. To do.

Therefore, by engaging the clutch 90R, even when the left wheel is rotating at a lower speed than the right wheel during the left turn, the clutch plate 90BR side to the right wheel side clutch plate 90 are engaged.
The torque is transmitted to the AR side, that is, from the input portion side of the rear differential to the right wheel side, and the distribution amount of the torque from the engine to the right wheel side is increased to be distributed to the left wheel side. Can be reduced.

Therefore, the hydraulic pressure adjusting portions for the clutches 90L and 90R in the hydraulic unit 38 are also controlled through the ECU 42 so that the torque distribution to the left and right rear wheels becomes a desired state. In this case, the ECU 42 controls required parts of the hydraulic unit 38 based on engine information, wheel speed information, steering wheel angle information (steering angle information), information on lateral acceleration and longitudinal acceleration of the vehicle body, and the like.

For example, when it is desired to distribute the driving torque from the input shaft 52 more to the left wheel rotation shaft 66, the left clutch 90L should be engaged at an appropriate control pressure according to the degree of distribution (distribution ratio). Of course, when it is desired to distribute more of the drive torque from the input shaft 52 to the right wheel rotation shaft 68, the right clutch 90R may be engaged at an appropriate control pressure according to the degree of distribution (distribution ratio).

The left and right clutches 90L and 90R are set so as not to be completely engaged at the same time.
When one of the left and right clutches 90L and 90R is completely engaged, the other is not engaged. That is, the operation modes of the clutches 90L and 90R are the left clutch 90L.
There are a mode in which only the right clutch 90R is engaged, a mode in which only the right clutch 90R is engaged, and a neutral mode in which neither is engaged.

As described above, in the torque moving mechanism 50, since the distribution of the left and right torque can be adjusted by moving the torque, the torque cross is extremely large as compared with the case where the distribution of the left and right torque is adjusted by simply braking one wheel. The characteristic is that the torque distribution can be adjusted in a wider range and the yaw moment can be generated in the vehicle without any discomfort.

1.2.3 Configuration of Hydraulic Unit According to the Present Device Here, the configuration of the hydraulic unit 38 will be described with reference to FIG. As shown in FIG. 3, the hydraulic unit 38 appropriately adjusts the pressure of the pressure accumulating portion 101 for accumulating the hydraulic oil and the hydraulic oil accumulated in the pressure accumulating portion 101, and the clutches 90L, 90R.
Control pressure output unit 102 for supplying to the oil chamber (not shown).

The pressure accumulator 101 includes an accumulator 103.
And a motor pump 104 that pressurizes the hydraulic oil in the accumulator 103 to a predetermined pressure, and a pressure switch 105 that monitors the differential oil pressure pressurized by the motor pump 104. Further, the control pressure output unit 102 is the motor pump 1
A hydraulic proportional pressure control valve (abbreviated as proportional valve) 106 for adjusting the pressure of the hydraulic oil in the accumulator 103 whose pressure is adjusted through 04, and the hydraulic oil regulated by the proportional valve 106 for either the left or right clutch 90L, 90R. An electromagnetic directional control valve (direction switching valve) 107 for switching whether to supply the oil to the oil chamber (not shown).

Such a hydraulic unit 38 is provided in the ECU 4
Although the operation is controlled by 2, the ECU 42 includes a wheel speed sensor (wheel speed detection means) 48A, a steering wheel angle sensor (that is, steering wheel turning angle detection means for detecting a steering wheel turning angle) 48B, and a longitudinal acceleration sensor (front and rear). G sensor) 4
8C, lateral acceleration sensor (lateral G sensor) 48D, throttle position sensor (TPS) 48E, pressure switch 105, and other sensors (vehicle state for detecting the running state of the vehicle
(State detecting means) is connected.

Based on the information from these sensors, the ECU 42 determines the motor pump 104 of the hydraulic unit 38 and the proportional valve according to the running state of the vehicle, that is, the vehicle speed, the steering state, the motion state of the vehicle body, and the like. The control unit 106 and the direction switching valve 107 are controlled. This proportional valve 1
06 and differential limiting control through control of the direction switching valve 107,
That is, the details of the torque movement control will be described later.

In FIG. 3, reference numeral 108 denotes a battery, 1
A motor relay 09 controls the motor pump 104 by controlling the supply of electric power from the battery 108 via the motor relay 109, and the pressure accumulation management by the pressure accumulation unit 101 is based on the detection information of the pressure switch 105. The operation of the motor pump 104 is controlled through the relay 109. Further, reference numeral 110 is an indicator lamp for displaying whether or not the differential limiting control by the hydraulic unit 38, that is, the torque movement control is being performed.

Further, since the differential limiting control through the hydraulic unit 38 needs to be linked with the engine output control, here, the ECU 42 outputs a control command to the hydraulic unit 38 and controls the engine output control. The output reduction information is also sent to an engine ECU (not shown). Although not shown, the ECU 42 includes a CPU, a ROM, a RAM,
It has an interface etc.

1.3 Outline of Control of the Present Device Here, the outline of control of the present device will be described with reference to the control block diagram showing the functional configuration relating to the control of the present device in FIG. As shown in FIG. 4, the processing by this control includes sensor input processing for receiving sensor inputs, calculation processing for calculating various values based on these sensor input values, and vehicle motion based on the calculation processing results. It can be divided into a control amount calculation process for calculating each control amount of control and a drive process for driving each actuator based on each calculated control amount.

Among these, in the sensor input process, four wheel speed sensors 48A, steering wheel angle sensor 48B, longitudinal acceleration sensor (longitudinal G sensor) 48C, lateral acceleration sensor (lateral G sensor) 48D, throttle position sensor (TP).
S) Receives sensor input from 48E or the like. In the calculation process, the actual measurement value and the theoretical value thereof are calculated for the speed difference between the left and right rear wheels. The actual measurement value (actual rotation speed difference) is based on the wheel speed values from the wheel speed sensors 48A for the four wheels, and
Theoretical value (target value, theoretical speed difference) is the steering wheel angle sensor 4
Based on the steering angle from 8B and the vehicle speed (vehicle speed) obtained from the wheel speed values from the four wheel speed sensors 48A,
Calculated respectively. In addition, front and rear G sensor 48C, lateral G
Based on the detection value from the sensor 48D, G (g
b), calculation side G (gy) is calculated. Further, in the arithmetic processing, drift determination and road surface μ estimation are further performed.

In the control amount calculation process, each control amount of the vehicle motion control is calculated based on each calculation result as described above.
In this control, the control amount of the target rotational speed difference follow-up control (target ΔN follow-up control) related to the control during the normal turn (target ΔN follow-up control amount) and the control amount of the acceleration turn control during the acceleration turn (acceleration turn control amount) And the control amount of the tuck-in control when the vehicle is tuck-in (tuck-in control amount)
And a control amount of steering transient response control regarding steering transient (steering transient response control amount) are respectively provided, and these added control amounts are added to output the added control amount (output control amount). ing.

The function of performing the control amount calculation processing is called control amount calculation means. Among these functions (control amount calculation means), the function of setting the control amount related to the target ΔN tracking control,
Alternatively, this setting function and the function of outputting a control signal based on the control amount obtained by this setting may be the ΔN tracking control means or the target rotation speed difference tracking control means, the function of setting the control amount of the acceleration turning control, or The setting function and the function of outputting a control signal based on the control amount obtained by this setting are the functions for setting the control amount for the acceleration turning control means and the tack-in control, or this setting function and the control amount obtained by this setting. The control signal output function based on the tuck-in control function, the function for setting the control amount (transient control amount) of the steering transient response control during the steering transient, or this setting function and the control amount obtained by this setting The function of outputting a control signal based on the above is referred to as steering transient response control means. Particularly, a means for calculating a control amount (vehicle behavior corresponding control amount) corresponding to the vehicle behavior, such as a target rotational speed difference tracking control means, an acceleration turning control means, a tuck-in response control means, is used as a vehicle behavior response control amount calculation means, a steering wheel operation. A means for calculating a control amount (transient control amount) based on a driving operation state such as a throttle operation or the like is also referred to as a transient control amount means.

Regarding the target ΔN follow-up control, it has a function (target value calculation means) for calculating or storing a target value of the yaw angle or the left / right wheel rotation speed difference corresponding to the turning state of the vehicle, and further, a steady turn. It has a function (steady turning control means) of calculating a control amount according to a target value at the time. Further, in the drive processing, a proportional valve output that outputs a command signal to the proportional valve 106 to adjust the torque movement amount and a command signal to the directional valve (direction switching valve) 107 to set the torque movement direction. A directional valve output and an indicator display output for outputting a display command signal to the indicator lamp 110 are provided.

2. Control Content of the Present Device Here, the content of the torque control as described above will be described more specifically in the order of input calculation processing, drift determination logic, vehicle motion control logic, road surface μ estimation, and actuator driving. 2.1 Input Calculation Process In the input calculation process, as shown in FIG. 6, the rear left wheel speed v
rl, rear right wheel speed vrr, steering wheel angle θh, vehicle body speed vb, steering wheel angular speed dθh, front left wheel speed vfl,
While receiving a detection signal related to the front right wheel speed vfr from each sensor, the previous calculated value (torque movement amount ta, road surface μ determination coefficient γ) and pressure switch, idle switch,
Upon receiving detection signals from the lateral G sensor, TPS (throttle position sensor), etc., the following numerical value calculation processing is performed.

2.1.1 Rear-wheel left-right speed difference (dvrd) The difference between the rear-left wheel speed vrl and the rear-right wheel speed vrr is calculated, and the actual left-right wheel speed generated during turning or torque movement control is calculated. The speed difference dvrd (= vrl-vrr) is obtained. 2.1.2 Digital filter value (d
vrf) Since the actual speed difference dvrd is used to determine the operating state of the torque transfer control, the actual speed difference dvrd is filtered by a digital filter to remove the influence of noise. Here, filter processing is performed as shown in equation (2.1.2.2), which performs smoothing processing as in equation (2.1.2.1).

Dvrf1 = (dvrd + odvrd) / 2 (2.1.2.1) dvrf = LPF [dvrd] = LPF [dvrf1, dvrf] (2.1.2.2) However, odvrd: keeps the previous dvrd Value dvrfl: smoothed value 2.1.3 Rear wheel average speed (vr) By averaging the rear left wheel speed vrl and the rear right wheel speed vrr, the rear wheel average speed vr [= (vrl + vrr)
/ 2] is obtained and used to determine the operating state of the torque transfer control.

2.1.4 Estimated vehicle speed (vb), turning radius (RR) This device has a function of estimating the vehicle speed by calculation (vehicle speed calculating device or vehicle speed detecting means). In the speed calculation device (vehicle speed detecting means), the estimated vehicle speed vb
Is basically calculated based on the third fastest wheel speed v3 of the left and right front wheels and the left and right rear wheels. this is,
Since this car is a four-wheel drive vehicle, each wheel becomes a drive wheel,
Such a drive wheel causes slippage with the road surface when the driving force is transmitted to the road surface. Therefore, if the vehicle speed is calculated based on the drive wheel, even if the vehicle speed is small, it will be faster than the actual vehicle speed. Becomes Therefore, the slowest wheel speed of the four driving wheels corresponds most to the actual vehicle speed. However, the detected value of the wheel speed may not be an appropriate value due to noise or the like. Therefore, considering the reliability of the detected value, the second lowest wheel speed of the four drive wheels (that is, the third wheel speed). Fast wheel speed)
The estimated vehicle speed vb is obtained by using v3.

By the way, since the wheel speed and the vehicle body speed correspond to each other at a constant ratio during straight traveling, for example, the vehicle body speed (simple calculation vehicle body speed) vbd obtained by multiplying the rotational speed of the wheel by the wheel outer peripheral length is used as the vehicle body speed. It can be vb. Therefore, the function of calculating the estimated vehicle body speed vb from the third wheel speed (that is, the third highest wheel speed) v3 is referred to herein as straight-ahead vehicle body speed estimation means.

However, during turning, the wheel speed changes between the turning inner wheel and the turning outer wheel, and such a change in wheel speed between the inner wheel and the outer wheel also differs depending on the turning radius and the vehicle speed. Therefore, when turning, it is necessary to make a correction according to the turning radius and the like. That is, at the time of turning, it is considered that the third highest wheel speed becomes the inner wheel of the rear wheel, and this inner wheel side becomes the simply calculated vehicle body speed vbd.
Calculated from geometrical relationships.

Therefore, the third wheel speed vbd (= v3) as the vehicle speed (simple calculation vehicle speed) estimated (calculated) by the straight-ahead vehicle speed estimating means, and the steering wheel angle sensor (steering wheel turning angle detecting means). The vehicle width at the time of turning of the vehicle from the steering wheel angle (steering wheel turning angle) θh detected in 48B and constants specific to the vehicle body of the vehicle, that is, the vehicle wheel base, tread width, stability factor, steering gear ratio, etc. Calculate and estimate the vehicle speed centered on the direction. The function of calculating the vehicle body speed centered in the vehicle width direction during turning is referred to as turning vehicle body speed calculating means.

That is, the turning radius RRi on the inner wheel side can be calculated by the following equation (2.1.3.1) based on the vehicle speed vbd on the inner wheel side.

RRi = (1 + A * vbd ² ) * Lw / δ = (1 + A * vbd ² ) * Lw * GR / θh (2.1.3.1) where δ: actual steering angle (= θh / GR) [ Where θh: steering wheel angle (steering wheel turning angle)] A: stability factor Lw: wheel base Lt: tread width GR: steering wheel gear ratio Further, the ratio of the vehicle body speed vbd to the vehicle body speed vb is the turning radius RRi on the inner wheel side. Is equal to the turning radius RR at the center of the vehicle body, and the turning radius RR uses the turning radius RRi to calculate the following formula (2.1.
Since the vehicle speed vb can be expressed as shown in 4.1), the vehicle speed vb can be divided into the following equation (2.1.
As in 4.2) to (2.1.4.4), it can be obtained from the vehicle body speed vbd and the steering wheel angle θh.

RR = RRi + Lt / 2 (2.1.4.1) Right turn vb = (RRi + Lt / 2) / RRi * vbd (2.1.2.1) Straight ahead vb = vbd (2.1. 4.3) When turning left vb = (RRi-Lt / 2) / RRi * vbd (2.1.4.4) The turning radius RR at the center of the vehicle is such a vehicle speed vb.
Based on, it can be expressed as the following equation (2.1.4.5).

RR = (1 + A * vb ² ) * Lw / δ = (1 + A * vb ² ) * Lw * GR / θh (2.1.4.5) Further, the wheel applied to the third wheel speed v3 (= vbd) If the vehicle slips significantly, the third wheel speed v3 will deviate significantly from the actual vehicle speed. In this case,
Since the wheel acceleration of the wheel applied to the third wheel speed v3 has a large difference from the actual longitudinal acceleration (front-rear G) of the vehicle, the wheel acceleration is increased by comparison with the actual acceleration of the wheel. The vehicle slips and the third wheel speed v3 (= vb
It can be determined that d) cannot be adopted as the vehicle speed.

Therefore, in this vehicle body speed computing device, the longitudinal G by the longitudinal acceleration sensor 48C provided on the vehicle body and the third
The wheel acceleration d (v3) / dt of the wheel at the wheel speed is used to determine whether or not the wheel is largely slipping,
When determining such a slip, the longitudinal acceleration sensor 48
The vehicle body speed is estimated by using the front and rear G by C, and the vehicle body speed based on the front and rear G is adopted instead of the vehicle body speed based on the wheel speed.

When the tire slips, the wheel speed v3
Rapidly increases, and the wheel acceleration d (v3) / of the wheel at the third wheel speed
When the difference between dt and the front-rear G measured by the front-rear acceleration sensor 48C becomes larger than a predetermined amount, it is considered to be a non-linear region where the tire slips. At this time, the front-rear G estimated vehicle body speed vbs is calculated, and the front-rear G estimated vehicle body speed vbs is adopted instead of the estimated vehicle body speed (wheel body speed corresponding vehicle body speed) vbd (= v3) based on the wheel speed v3.

This vehicle body speed will be described later (item,
2.2.2) is calculated by the following equation. vbs == gx _SL · t + vb _SL However, vb _SL : vehicle speed when tire slip occurs vb _SL gx _SL : front and rear G detected when tire slip occurs Also, immediately after this slip occurrence, wheel speed v3 With the increase, the difference between the wheel speed v3 and the front-rear G estimated vehicle body speed vbs, that is, the vertical slip coefficient dvvbs of the tire described later increases, but as the slip converges, the wheel speed v3 decreases and the front-rear G estimation. As it approaches the vehicle speed vbs,
The vertical slip coefficient dvvbs of the tire is reduced.

Therefore, the tire vertical slip coefficient dvvb
Based on s, it is possible to estimate the slip state of the tire, that is, the linear region where the tire does not slip or the nonlinear region where the tire slips. Here, when the vertical slip coefficient dvvbs converges to a certain value or less, the tire returns to the linear region where it does not slip.
Since the estimated vehicle body speed vbs is adopted, the estimated vehicle body speed is restored to the estimated vehicle body speed vbd (= v3) based on the wheel speed v3.

2.1.5 Longitudinal acceleration (gx) First, the following formula (2.1.5.1) is used to calculate from the change in the simple calculation vehicle speed vbd calculated in a predetermined cycle, Since the acceleration gxd varies greatly, it is processed by a low pass filter [see (2.1.5.2)] to obtain the longitudinal acceleration gx.

Gxd = vbd−ovbd (2.1.5.1) However, ovbd: simple calculation vehicle speed vbd gx = LPF [gxd] (2.1.5.2) one cycle or a predetermined cycle before 2.1.5.2. 6 Reference Lateral Acceleration (gy) The reference lateral acceleration (gy) can be calculated from the radius RRi and the estimated vehicle speed vb considering the centrifugal force acting on the vehicle at the time of turning, and the radius RRi is calculated from the steering wheel angle θh as described above. Since it is required, the reference lateral acceleration (gy) is calculated by the following equation (2.1.6.1)
As described above, it is calculated from the steering wheel angle θh and the estimated vehicle speed vb. This reference lateral acceleration (gy) is also called a calculated lateral G.

Gy = vb ² / RR = vb ² * θh / [(1 + A * vb ² ) * Lw * GR] ... (2.1.6.1) 2.1.7 Rear wheel reference rotational speed difference (dvhf) Rear wheel The reference rotational speed difference dvhf is the turning radius RR during turning.
Is the rotational speed difference of the rear wheels that can be calculated geometrically from the relationship as shown in Fig. 7, and using the relationship of equation (2.1.4.5), firstly the following equation (2.1.7.1) Estimated vehicle speed v
b, the rotation speed difference dvhr by the function of the steering wheel angle θh
Ask for. The above-mentioned rear left wheel speed vrl and rear right wheel speed vrr are low-pass filtered, and the rotational speed difference dvhr is processed by a low-pass filter in order to match the phases with these [see (2.1.7.2)]. , The rear wheel reference rotational speed difference dvrf is obtained. The function of calculating such a rear wheel reference rotational speed difference dvrf is referred to as target value calculating means.

Dvhr = Lt * vb / RR = Lt * vb * θh / [(1 + A * vb ² ) * Lw * GR] (2.1.7.1) dvhf = LPF [dvhr] (2.1.7.2) ) 2.1.8 Front wheel reference rotational speed difference (dvhff) The front wheel reference rotational speed difference dvrff is the turning radius R when turning.
It is the rotational speed difference of the front wheels that can be geometrically calculated from the relationship as shown in Fig. 7 according to R and the steering angle δ, and the formula (2.1.4.5)
First, the rotational speed difference dvh is obtained from the function of the estimated vehicle speed vb and the steering wheel angle θh as shown in the following equation (2.1.8.1), and this is processed by the low-pass filter. (Refer to 2.1.8.2)], front wheel reference rotational speed difference dvrf
get f.

Dvrf = Lt * vb * cos (θh / GR) / RR = Lt * vb * cos (θh / GR) * [θh / [(1 + A * vb ² ) * Lw * GR] ... (2.1. 8.1) dvrff = LPF [dvrf] (2.1.8.2) 2.1.9 Speed difference between left and right front wheels (dvfd) When calculating the difference between the front left wheel speed vfl and the front right wheel speed vfr, turning And the actual speed difference between the left and right rear wheels
d (= vfl-vfr) is obtained.

2.1.10 Torque Movement Amount (taf: Temporary Delay Value) Since the torque movement has a time delay from the output of the command value to the appearance of the actual vehicle behavior, the torque movement command value. A low-pass filter is applied to ta to match the phases [see (2.1.10.1)] to obtain the torque transfer amount taf. taf = LPF [ta] (2.1.10.1) 2.2 Drift judgment logic In this control, it is judged whether or not the vehicle is about to drift, and the judgment result is determined by the torque transfer control of the left and right wheels of the vehicle. Used to control the movement of. Therefore, in this control,
The drift determination is performed by each process as shown in FIG. The function of determining whether the vehicle is in a drift state or a non-drift state is referred to as drift determining means (turning state determining means).

That is, in this control, it is determined that the drift occurs when the tire has slipped or slipped vertically. The tire skid can be determined when the relationship between the calculated lateral G and the actual lateral G becomes non-linear, and the tire longitudinal slip has a non-linear relationship between the estimated vehicle body speed vb and the front-rear G estimated vehicle body speed vbs described later. You can judge when it becomes. Normally, when the vehicle drifts, side slips and vertical slips are involved, so both are considered in this control.

Therefore, in this control, as shown in FIG.
The drift determination coefficient srp as the degree of slip of the wheels is set based on the coefficient (tire slip coefficient) dgy according to the tire skid state and the coefficient (tire slip coefficient) dvvbs corresponding to the tire longitudinal slip state ( (Calculation) and outputs this, and further, based on this drift determination coefficient srp, drift correction coefficients srp1 to srp5
To set. The function of setting the drift determination coefficient srp as the degree of slip of the wheel is referred to as slip degree detecting means. In addition, such a drift determination coefficient sr
The control in which p is reflected in the torque movement control is also referred to as slipping control.

2.2.1 Tire Side Slip Coefficient (dgy) In this control, as described above, the calculated lateral G, that is, the reference lateral acceleration gy is calculated from the steering wheel angle θh and the estimated vehicle speed vb. On the other hand, the lateral G sensor detects the actual lateral acceleration (actual lateral G) rgy. When the vehicle is traveling without skidding, the relationship between the calculated lateral G and the actual lateral G is linear. Therefore, in order to make a drift decision,
The calculated lateral G and the actual lateral G are compared.

However, the calculated lateral G (gy) calculates the lateral G from the input information such as the steering wheel angle θh, and there is a phase delay before the lateral G occurs in the vehicle depending on the steering wheel. In control, the calculated lateral G is filtered by a low-pass filter to perform phase matching [see (2.2.1.1)]. gyf = LPF [gy] (2.2.1.1) Also, due to the influence of tires and the difference in gear ratio, there is an error between the calculated lateral G (gy) and the actual lateral G (rgy) even in the linear region. Since it occurs, the actual lateral G (rgy) is corrected by the coefficient k as shown in the following equation (2.2.1.2), and coefficient matching is performed.

Rghyh = k * rgy (2.2.1.2) As a result, it is possible to compare the calculated lateral G (gyf) in which the phases are matched with the actual lateral G (rgyh) in which the coefficients are matched. Then, the calculation lateral G calculated by the following formula (2.2.1.3)
(Gyf) and the actual lateral G (rgyh) based on the dimensionless value (tire slip coefficient) dgy, that is, the non-linearity generated between the calculated lateral G and the actual lateral G, that is, the lateral direction of the tire. Determine non-linearity.

FIG. 9 is a diagram showing an example of the correspondence between the actual lateral G (rgy) and the calculated lateral G. If there is no tire skid or the like, the actual lateral G (rgy) and the calculated lateral G are as shown by a straight line A. And have a linear relationship, but actually, when the grip limit of the tire is exceeded, skid or the like occurs, and the actual lateral G does not increase unlike the calculated lateral G. On the high μ road, the linearity can be maintained up to the region where the lateral G is high as shown by the curve B, but on the low μ road, the linearity cannot be maintained in the region where the lateral G is low like the curve C.

The sideslip coefficient dgy of the tire is calculated by the following equation (2.2.
Define as in 1.3). dgy = | (gyf-rgyh) / rgyh | (2.2.1.3) However, in the calculation of the sideslip coefficient dgy of such a tire, calculation start conditions such as the following formula (2.2.1.4), and Clear conditions such as the following formula (2.2.1.5) are provided. This is the size of the actual lateral G (rghyh), the calculated lateral G and the actual lateral G.
If the magnitude of the difference (gyf-rgyh) from the value does not become larger than a certain level, there is no risk of the vehicle drifting. In such a case, the sideslip coefficient dgy is not calculated and the calculation frequency is It is decreasing.

| Rgyh | <a [m / s ² ] and | gy
f-rgyh | <b [m / s ² ] However, when a and b are constants, dgy = 0 ..................... (2.2.1.4) Generally, the actual lateral G and the calculated lateral After passing the linear region with G, the actual lateral G does not increase like the calculated lateral G, so the above equation (2.2.
1.3) can be transformed as follows.

Gyf = (1 + dgy) rgyh (2.2.1.3.a) When leaving the linear region, dgy gradually increases from 0, and the relationship of the above equation (2.2.1.3.a) is For example, it can be shown as a straight line D in FIG. Therefore, theoretically, it is possible to determine that the linearity has collapsed when the sideslip coefficient dgy is other than 0. However, in reality, even if phase matching and coefficient matching are performed on the actual lateral G and the calculated lateral G, perfect matching is always achieved. Since it is difficult to do so, the skid coefficient dgy often occurs (becomes other than 0) even in the actual linear region. Therefore, in this control, as shown in FIG. 10, if the sideslip coefficient dgy is equal to or less than the first predetermined value (dgy1), the linear region is set, and if the sideslip coefficient dgy is equal to or more than the second predetermined value (dgy2), the non-linear region is set. When the skid coefficient dgy is between the first predetermined value and the second predetermined value, it is assumed that the degree of non-linearity increases as it approaches the second predetermined value.

2.2.2 Tire vertical slip coefficient (dvvb
s) In the present control, as described above, the estimated vehicle body speed vb is calculated based on the wheel speed v3 that is the third fastest among the four wheels. However, if the tire slips significantly, the vehicle body speed based on such wheel speed v3 is calculated. vb will be higher than the actual vehicle speed. Therefore, if the occurrence of tire slip is estimated, the front-rear G is calculated based on the vehicle speed and the front-rear G at this time instead of the wheel speed.
The estimated vehicle speed vbs is calculated.

The front-rear G estimated vehicle speed vbs is equal to the front-rear G.
Based on the front and rear G of the vehicle body detected by the sensor, it is calculated by the following formula (2.2.2.1) from the vehicle body speed vb _SL when the tire slip occurs and the front and rear G (gx) _SL detected values. Note that t
Is the elapsed time after the occurrence of slip, and the wheel speed (for example,
It can be estimated that the slip has occurred when the third highest wheel speed v3) has increased rapidly.

Vbs == gx _SL · t + vb _SL (2.2.2.1) The vertical slip coefficient dvvbs of the tire is the estimated front-rear G vehicle speed vbs calculated as described above and the third detected simultaneously. Based on the extremely fast wheel speed v3, the following equation (2.2.2.
The value is calculated by 2). In consideration of the noise influence on the calculated value dvvbsd, the value is further filtered by a low-pass filter [see (2.2.2.3)] to obtain the tire vertical slip coefficient dvvbs.

Dvvbsd = v3-vbs (2.2.2.2) dvvbs = LPF [dvvbsd] (2.2.2.3) For the estimated front-rear G vehicle speed vbs, for example, the actual vehicle speed VR is almost constant. If the tire slips after entering the extremely low μ road and then the slip converges, if the tire slips, the wheel speed v
3 rapidly increases, and the front-rear G estimated vehicle body speed vbs comes to be calculated.

Immediately after the occurrence of the slip, the wheel speed v3
Therefore, the difference between the wheel speed v3 and the front-rear G estimated vehicle body speed vbs, that is, the vertical slip coefficient dvvbs of the tire increases. When the slip converges, the wheel speed v
Since 3 decreases and approaches the front-rear G estimated vehicle body speed vbs, the vertical slip coefficient dvvbs of the tire decreases. Therefore, based on the tire vertical slip coefficient dvvbs, it is possible to estimate the slip state of the tire, that is, the linear region where the tire does not slip or the nonlinear region where the tire slips.

Therefore, theoretically, the vertical slip coefficient dvvbs
It can be determined that the value becomes non-linear when is not 0, but
Actually, the slip occurrence is estimated and the front-rear G estimated vehicle speed v
Since an error also occurs in the estimation of bs, in this control, as shown in FIG.
As shown in, the vertical slip coefficient dvvbs is equal to the first predetermined value (d
vvbs1) or less, linear region, vertical slip coefficient dvvb
If s is greater than or equal to the second predetermined value (dvvbs2), the vertical slip coefficient dvvbs is set to the first predetermined value and the second
If it is between the predetermined value and the second predetermined value, the degree of nonlinearity is assumed to increase.

2.2.3 Drift Judgment Coefficient (srp) In this device, the drift judgment is performed in consideration of both the side slip coefficient dgy and the vertical slip coefficient dvvbs as described above.
Therefore, the following formula (2.2.3.1) is used to calculate a value (this is referred to as a drift determination coefficient) srp (= srpd ² ) obtained by combining the sideslip coefficient dgy and the vertical slip coefficient dvvbs,
Used for drift judgment.

Srp = (a · dgy) ² + (b · dvvbs) ² (2.2.3.1) However, coefficient adjustment for making a and b circles The drift determination coefficient srp is shown in FIG. It can be evaluated by such a drift judgment circle (friction circle). FIG. 12 shows a value (a
.Dgy) and vertical slip coefficient dvvbs (coordinate adjusted values) (b.dvvbs) are plotted on the abscissa and the ordinate, respectively, and the orthogonal coordinates are shown. The drift determination coefficient srp corresponds to the square of the distance from the origin at this coordinate.

The drift judgment circle is a circle centered on the origin of such coordinates, and has a first radius r ₁ and a second radius r ₂ (r ₁ <r ₂ ). And the radius r ₁
Within the circle (the area where the tire is not slipping), outside the circle with a radius r ₁ is a non-linear area (the tire is slipping), and within the nonlinear area is a drift area outside the circle with a radius r _2. Is set.

That is, the square root (srp ^1/2 ) of the drift determination coefficient srp is within the radius r ₁ (that is, srp ^1/2 ≤r).
₁ ) is a linear region, srp ^1/2 is larger than the radius r ₁ (that is, srp ^1/2 > r ₁ ), and srp ^1/2 is larger than the radius r ₂ ( That is, srp ^1/2
If> r ₂ ) then it is said to be in the drift region. In the non-linear region, the region where r ₁ <srp ^1/2 ≦ r ₂ is not a complete drift, but the drift determination coefficient sr
It is assumed that there is a degree of drift tendency corresponding to p.

For example, FIG. 13 shows the drift determination coefficient srp.
Shows the correspondence of the drift determination to the linear region when srp is radius r ₁ ² or less (that is, srp ≦ r ₁ ² )
If srp is larger than the radius r ₂ ² (that is, srp> r ₂ ² ), the drift region and the region where srp is r ₁ ² <srp ≦ r ₂ ² are not complete drifts, but the drift determination coefficient sr
It is said that the degree of drift corresponds to p.

(Drift Corresponding Control Starting Condition) If the drift judgment coefficient srp is equal to or more than a predetermined value and the counter steer is turned off and the steering wheel angular velocity of this counter steer is equal to or more than a predetermined speed, it is determined that the vehicle is drifting (drift determination). Means or turning state determining means). It is determined that the counter steer is turned off when the steering angle exceeds the neutral position, that is, when the calculated lateral G direction and the actual lateral G direction are opposite. That is, when all the conditions of the following three equations are satisfied at the same time, it is determined that the vehicle is drifting and the drift coping control (slip coping control) is started. The function of determining the start of such drift control (slip control) is called start determination means.・ Drift judgment coefficient srp is greater than or equal to a predetermined value srp> srp0 (2.2.6.1) ・ Calculation lateral G (gy) direction and actual lateral G (rgyh) direction are opposite gy ・ rgyh <0 ・ ・ ・ (2.2.6.2) ・ The steering wheel angular velocity Δθh is greater than or equal to the predetermined velocity Δθ ₁ Δθh ≧ Δθ ₁ (deg / s) ・ ・ ・ (2.2.6.3) In addition, the conditions of the above three equations are the same. Even if it is not established, it may be determined that the vehicle is drifting when the drift determination coefficient srp is equal to or greater than a predetermined value. The steering wheel angular velocities Δθh and Δθ ₁ are also described as dθh and dθ ₁ , respectively.

(Drift Corresponding Control End Condition) When the steering angle returns to the neutral position again, that is, when the calculated lateral G direction and the actual lateral G direction are equal, it is judged that the drift traveling has ended, Stop the drift control. Note that the function for determining the end of such drift coping control (slip coping control) is referred to as end judgment means.・ The calculated lateral G (gy) direction and the actual lateral G (rgyh) direction are the same. Gy ・ rgyh> 0 (2.2.6.4) Also, it is determined that the vehicle is not drifting according to the above-mentioned conditional expression. Then, the drift determination coefficient srp is set to zero (sr
p = 0).

2.2.6 Turning lateral G (drift compatible, gg
y) By the way, in this control, there is a torque movement control based on the lateral acceleration (turning lateral G) applied to the vehicle at the time of turning, and for example, tuck-in response control or accelerated turning control corresponds to this.
The calculated lateral G and the actual lateral G correspond to this turning lateral G,
Since there is no difference between the calculated lateral G and the actual lateral G when the tires are gripping the road surface firmly (when gripping), the calculated lateral G also corresponds to the behavior of the vehicle along with the actual lateral G. Therefore, as the turning lateral G, a calculated lateral G having a processing speed faster than the actual lateral G can be used. However, since a large difference occurs between the calculated lateral G and the actual lateral G during drift running, the calculated lateral G cannot be used. In this case, the actual lateral G corresponding to the behavior of the vehicle is regarded as the turning lateral G. Need to be used.

Therefore, in this apparatus, the calculated lateral G is normally used, and when the calculated lateral G cannot cope with the actual situation, the actual lateral G is used. For this reason, when it is determined that the vehicle is drifting under the drift-corresponding control start condition, the lateral turning G
Is switched from the adoption of the calculated lateral G to the adoption of the actual lateral G, and when it is determined that the drift running is ended under the drift-corresponding control termination condition, the adoption of the actual lateral G is set to return to the adoption of the calculated lateral G. ing.

The lateral G is selected by the lateral G selection coefficient dor.
It is represented by i, and when the calculation horizontal G is selected, dori = 0
When the actual lateral G is selected, dori = dori1 (constant). The turning lateral G: ggy corresponding to the drift can be represented by the following equation by the lateral G selection coefficient dori. ggy = [dori.rgyh + (dori1-dor
i) gy] / dori1 where gy: calculated lateral G, rgyh: actual lateral G Furthermore, an example of selecting the turning lateral G for dealing with such drift will be described with reference to FIG. In FIG. 18, the solid line indicates the calculated lateral G (gy), the broken line indicates the actual lateral G (rgyh),
As shown in the figure, lateral G occurs in the vehicle when the vehicle turns, and there is no difference between the calculated lateral G and the actual lateral G when the vehicle is gripping. G changes suddenly. The reason why the calculated lateral G suddenly changes is that the driver applies the steering wheel operation in the drift state, and when the steering wheel operation is applied, it is calculated based on the steering wheel angle θh as shown in the equation (2.1.6.1). That is, the calculated lateral G greatly changes. In particular, when the counter steer is turned off during drift, the calculated lateral G becomes the actual lateral G
And changes in the opposite direction. The calculation is performed only until the calculated lateral G changes in the opposite direction to the actual lateral G and the calculated lateral G becomes in the same direction as the actual lateral G, that is, only during the period indicated as "during drift control" in FIG. The actual lateral G input is adopted instead of the lateral G.

2.3 Vehicle Motion Control Logic As described above, in the present power transmission control device, the control modes are the target rotational speed difference tracking control (target ΔN tracking control),
Acceleration turning control, tuck-in response control, and steering transient response control are provided. Here, each of these controls will be described in detail.

2.3.1 Target ΔN Tracking Control The target ΔN tracking control is a control aiming at both the action as the yaw rate feedback control (yaw rate FBC action) and the action as the LSD (LSD action). Control is performed so as to eliminate the difference between the rear wheel reference rotational speed difference (theoretical value, dvhf) and the left and right rear wheel speed difference (actual speed difference, dvrd) obtained in 2.1.7.2). . Therefore, the broken line block B3 in FIGS.
As shown in 1, a plurality of control amounts corresponding to μ (high μ road control amount tbh, low μ road control amount tbl) are set.

These control amounts are set, for example, in JP-A-7-10.
Although it is obtained by the method disclosed in Japanese Patent No. 8840, the gain characteristic with respect to the rotational speed difference between the left and right wheels is set differently for the control amount for the high μ road and the control amount for the low μ road.
Further, when drift traveling is determined by the drift determination coefficient srp and the like, the drift determination coefficient srp1 is adjusted so that the theoretical value dvhf becomes zero, and control is performed to eliminate the rotational speed difference between the left and right wheels.

Further, the high μ road control amount and the low μ road control amount are subjected to gain adjustment by the drift determination coefficients srp2, srp3 to calculate the control amount suitable for drift running.

2.3.2 Accelerating Turning Control The acceleration turning control is a control for suppressing an increase in the understeer tendency during a sharp turning as described above, and this control is required because the vehicle stability is non-linear. Is the case. That is, as shown in FIG. 19, the spherical center acceleration (that is,
When the relationship between the turning G) and the steering ratio deviates from the linear region (see the broken line portion), the turning radius of the vehicle increases. This is because the steering characteristics of the vehicle are strengthened to the understeer side during a sharp turn.

As described above, during a sharp turn, the target ΔN
In the follow-up control, the torque is moved to the outer wheel turning side to generate a moment in the turning direction to increase the cornering force of the front wheels. However, the target ΔN follow-up control causes a slight reaction delay due to the feedback control. Therefore,
During such a sharp turn, the acceleration turning control is performed to move the torque to the turning outer wheel side to generate or increase the yaw moment in the turning direction, thereby increasing the cornering force of the front wheels in the region where the longitudinal acceleration is large. It is trying to suppress understeering.

(1) Acceleration turning control amount (teh, tel) In this control, as shown in the block B32 of FIGS. 14 and 15, when the turning lateral G (ggy) is equal to or larger than a predetermined value, control during acceleration turning is performed. The basic control amounts tehd and teld of are set. Further, this control amount is output on condition that the tuck-in correspondence control is not being performed. Note that the basic control amount tehd, t
The eld is determined based on the turning lateral G (ggy) when the turning lateral G (ggy) is equal to or larger than a predetermined value according to the maps shown in FIGS. 16 and 17, and is set based on the turning lateral G (ggy). Will get better. Note that FIG. 16 is a map for high μ roads (that is, a map for high road surface friction resistance), and FIG. 17 is a map for low μ roads (that is, a map for low road surface frictional resistance), and turning is performed based on these maps, respectively. Basic control amount (control amount) corresponding to lateral G (ggy), that is, basic control amount for high μ road (control amount for high road surface friction resistance) tehd, basic control amount for low μ road (control for low road surface friction resistance) Amount) teld is set.

As shown in FIGS. 16 and 17, the lateral G (gg
In a small region of y), the same turning lateral G (ggy)
On the other hand, the low road surface friction resistance map has a larger control amount than the high road surface friction resistance map, and the high road surface friction resistance for a similar turning lateral G (ggy) in a large lateral G (ggy) area. The control map gives a larger control amount than the low road friction resistance map. In addition, as shown in FIGS. 16 and 17, each map is provided with a dead zone region in which the control amount is 0 in a region where the turning lateral G (ggy) is small, thereby stabilizing the control. The chain line in FIG. 16 shows the characteristics of the low road surface friction resistance map (see FIG. 17).
The chain line in the middle shows the characteristics of the high road surface friction resistance map (see FIG. 16).

In this embodiment, the turning lateral G
When (ggy) is equal to or greater than a predetermined value, it is determined that the vehicle is turning suddenly. However, if the turning lateral G (ggy) is detected even if it is small, the control amount is calculated so as to increase the rotational propulsive force of the outer turning wheel. Good. When it is determined that the vehicle is drifting, the vehicle turns laterally G
As the actual lateral G, the control amount is calculated.

When it is determined that the vehicle is drifting, the acceleration turning control amounts teh and tel are gain adjusted by the drift correction coefficient srp5 to calculate the control amount suitable for the drifting. For example, during drift driving, the acceleration turning control amount te
It may be configured to set h and tel to zero.

2.3.3 Control for Tack-in As described above, during deceleration turning, the steering characteristics of the vehicle tend to oversteer as the cornering force of the front wheels increases, as opposed to acceleration turning, and the vehicle tends to tuck-in. Become. As described above, during the deceleration turn, the target ΔN
In the follow-up control, the torque is moved to the inner wheel side of the turning to generate the yaw moment in the turning restraining direction.
Although the oversteering is suppressed, the target ΔN follow-up control causes a slight reaction delay because it is a feedback control.

Therefore, at the time of decelerating turning, the torque is moved to the turning inner wheel side to perform the tack-in response control for generating or increasing the yaw moment in the turning restraining direction, and the cornering force of the front wheels is reduced to suppress the oversteering. To do. As a result, the turning behavior of the vehicle is controlled to avoid tuck-in of the vehicle and spin due to tuck-in.

Regarding these control amounts, for example, according to the method disclosed in JP-A-7-108840,
The control amount tdh for the high μ road and the control amount tdl for the low μ road are obtained. Further, similar to the acceleration turning control, when it is determined that the vehicle is drifting, the actual lateral G is adopted as the turning lateral G to calculate the control amount.

Further, when it is determined that the vehicle is drifting, the tuck-in corresponding control amounts tdh and tdl are set to the drift correction coefficient sr.
The control amount suitable for drift running is calculated by adjusting the gain with p5. For example, when drifting, the tuck-in control amounts tdh and tdl may be set to zero. Further, in the tuck-in response control, as in the acceleration turning control, in a small region of the turning lateral G (ggy), the low road surface friction resistance map (low μ road map) is higher than the high road surface friction resistance map (high road map). A control amount larger than the μ road map) may be given. Alternatively, a dead zone is provided in a small region of the turning lateral G (ggy), and the high road surface friction resistance map (high μ road map) has a larger control amount than the low road surface friction resistance map (low μ road map). May be given.

2.3.4 Steering transient response control Steering transient response control is control performed during steering transient.
As shown in the block B33 of FIGS. 14 and 15, the control is performed so as to be proportional to the change of the steering angle, that is, the steering angular velocity. That is, the ECU 42 has a function of setting the steering transient response control amount (transient control amount), that is, means steering transient response control amount setting means (transient control amount calculation means), and steering transient response Control amount (steering angular velocity proportional control amount) t
c can be set. Therefore, first, the basic control amounts tchd, tcld corresponding to the steering angular velocity dθh
Is set, and the correction according to the vehicle speed, the correction according to the turning or turning back of the steering wheel, and the drift correction coefficient 4 are set.
Drift correction is performed by (srp4), and control is performed by the control amounts tch and tcl thus obtained.

Regarding these control amounts, for example, according to the method disclosed in JP-A-7-108840,
The control amount tch for the high μ road and the control amount tcl for the low μ road are obtained. If it is determined that the vehicle is drifting, the tuck-in control amounts tch and tcl are set to the drift correction coefficient sr.
The control amount suitable for drift running is calculated by adjusting the gain with p4. For example, during drift traveling, the tuck-in control amounts tch and tcl may be set to zero.

2.4 Road Surface μ Estimation (Explanation of the Road Surface Friction Coefficient Estimating Device for This Vehicle) In the torque movement control, the control effect differs depending on whether or not the road on which the vehicle runs is slippery, that is, the state of the road surface friction resistance. Therefore, this device (the power transmission control device for the left and right wheels for the vehicle) is provided with a vehicle road surface friction coefficient estimation device, and this vehicle road surface friction coefficient estimation device expresses the road surface friction coefficient ( Hereinafter, also referred to as road surface μ) is configured to estimate μ.

In the present road surface friction coefficient estimating apparatus, the road surface μ is estimated in three steps: μ estimation during steady turning, μ estimation during starting, and μ estimation during non-linearity. That is, each of the steady turn, start, and non-linear stages exists in the region as shown in FIG. 20 regarding the turning lateral G and the vehicle speed. Note that the μ estimate when
The initial value of is set. Further, the non-linear time is when the vehicle is non-linear with respect to the steering of the steering wheel. Here, in each of these cases, the road surface μ
A determination coefficient (road surface friction coefficient, that is, a coefficient representing the degree of the road surface μ) γ is obtained, and the output gain value (output value) of each control amount is determined from the road surface μ determination coefficient γ value. The vehicle road surface friction coefficient estimating device is also referred to as a road surface μ estimating device, a road surface μ determining device, or a road surface friction coefficient detecting device or a road surface μ detecting device.

2.4.1 Road Surface μ Estimation During Steady Turning The optimum value of the torque movement amount in the vehicle motion control differs depending on the level of the road surface μ (road surface friction coefficient). Therefore,
In this vehicle road surface friction coefficient estimation device, during steady running (particularly during steady turning), a parameter corresponding to the vibration component of the wheel speed due to the unevenness of the road surface (a parameter indicating the uneven condition of the road surface), that is, the wheel speed difference. β [However, here
As will be described later, a vibration component βpp of the wheel speed difference β is used] and a parameter according to the slipperiness of the road surface (a parameter indicating the slipperiness of the road surface), that is, the driving force in the grip region of the wheel (tire). The road surface condition is estimated by fuzzy inference from the road surface μ judgment value α (however, the average value αh of the road surface μ judgment value α is used as will be described later), which is the rate of increase of the slip ratio with respect to Is determined by a counter (learning function) to determine the road surface μ.

Therefore, the vehicle road surface friction coefficient estimating device is functionally configured as shown in FIG. FIG. 27
In 201, the average speed vr of the rear wheels, the average speed difference vf of the front wheels from the average speed vf of the front wheels β2 [= (vr-vf)
/ K, K: constant]. Reference numeral 202 denotes a wheel speed difference calculation unit that obtains a wheel speed difference β between the front and rear wheels. The calculation unit 202 calculates a rotational speed difference of a driving force difference calculated from a rear wheel left / right speed difference dvrd and a front wheel left / right speed difference dvfd. From ddv (= dvrd-dvfd), the wheel speed difference β (d
dv / K, K: constant) is calculated.

Reference numeral 204 denotes a road surface μ judgment value calculation unit for obtaining a road surface μ judgment value α. In this calculation unit 204, the speed difference ddvf due to the torque movement amount taf is obtained by dividing by the estimated vehicle speed vb. taf is obtained as a temporary delay value of the torque movement amount in which the phases are adjusted as described above, and the speed difference ddvf is calculated by the reference rotation speed difference calculation unit 205 in the turning correction rotation speed difference calculation unit 206. Reference speed difference ddvh (= dvhf-d calculated from speed difference dvhf and front wheel reference rotation speed difference dvhff
vhff) and the rotational speed difference ddv of the driving force difference calculated by the calculation unit 203, ddvfd (= ddv-ddvh).
Is processed by a low pass filter to obtain.

Reference numeral 207 denotes a road surface μ determination unit which determines whether or not a road surface μ determination determination condition, which will be described later, is satisfied. In the road surface μ determination unit 207, various vehicle state detection means detect. The traveling state of the vehicle, that is, the torque movement amount taf, the road surface μ determination value α, the brake switch information bks
w, steering wheel angular velocity dθh, drift determination coefficient srp,
The determination is made based on information such as the reference lateral acceleration gy. The road surface μ determination determination condition also includes a condition whether the vehicle is in steady running (here, steady turning running).

211 is the maximum amplitude βpp of the slip difference ratio β.
Is a calculation unit (concavo-convex state parameter calculation unit) that calculates the maximum amplitude βpp of the slip difference ratio β at a fixed time (sampling time). By performing this calculation every predetermined cycle, The maximum amplitude βpp is updated. Reference numeral 212 denotes a calculation unit (slipping parameter calculation unit) for calculating an average value αh of the road surface μ determination value α, which calculates an average value αh of the road surface μ determination value α for a fixed time (sampling time). The average value αh is updated in this predetermined cycle by also performing every predetermined cycle.

Such a concavo-convex state parameter calculator 21
1. Since the slipperiness parameter calculation unit 212 calculates the parameters when the vehicle is in a steady state (here, in a steady turn), these are referred to as steady parameter calculating means. Also,
The unevenness parameter calculation unit 211 uses the slip difference ratio β calculated by the wheel speed difference calculation unit 202 when the road μ determination unit determines that the road μ determination determination condition is satisfied.
When it is determined that the road surface μ determination determination condition is not satisfied, the maximum amplitude βpp is calculated using the average speed difference β2 between the front and rear wheels calculated by the calculation unit 201.

Also, the slipperiness parameter calculation unit 212
Only when the road surface μ determination unit determines that the road surface μ determination determination condition is satisfied, the road surface μ determination value α from the road surface μ determination value calculation unit 204 is taken in to calculate the average value αh. Note that the road surface μ determination value α from the road surface μ determination value calculation unit 204
Is limited by the limiter 208 to a value less than or equal to a predetermined value αmax, and is then input to the calculation unit 212. By not taking in an excessive road surface μ determination value (including a signal error, etc.), The accuracy of calculating αh is secured.

Further, 220 is a road surface index calculating means,
In the road surface index calculating means 220, other parameters, that is, a high value, which unifies the road surface condition, from the maximum amplitude βpp of the slip difference ratio β and the average value αh of the road surface μ judgment value α, Index corresponding to each state of μ road degree (high μ road), medium μ road degree (medium μ road), low μ road degree (low μ road) (here, because fuzzy reasoning is used, this index is used as a goodness of fit) Is calculated).

Reference numeral 230 denotes a road surface friction coefficient calculating means. The road surface friction coefficient calculating means 230 continuously and cumulatively obtains the index (degree of conformity) calculated by the road surface index calculating means 220, and cumulatively evaluates it. A value indicating the road surface friction coefficient (road surface μ determination coefficient) γ1 is calculated by a so-called learning function. The road surface μ determination coefficient γ1 estimated during such steady running of the vehicle is called a first parameter.
Of calculating the road surface μ determination coefficient γ1 as a parameter of the steady state parameter calculation means 210 (211 and 211).
12), the road surface index calculating means 220, the road surface friction coefficient calculating means 230, and the like are referred to as the steady-state parameter calculating means 240.

As shown in FIG. 35, the road surface μ determination coefficient γ1 as the first parameter calculated by the steady parameter calculating means 240 and the specific parameter calculating means (specific running other than the steady running of the vehicle) to be described later. The value indicating the road surface friction coefficient (the road surface friction coefficient is calculated by totalizing (selecting) the road surface μ determination coefficient γ2 as the second parameter calculated by the means 250 for calculating a parameter from the running state in the inside). μ determination coefficient) γ is calculated.

That is, during steady running, the road surface μ determination coefficient γ1 calculated by the steady parameter calculating means 240 is set to the road surface μ.
The determination coefficient γ is selected, and the road surface μ determination coefficient γ2 calculated by the specific parameter calculating means 250 during the specific operation is selected as the road surface μ determination coefficient γ. Note that FIG.
The block B81 shown in FIG.
The function as 0 and a part of the function as the road surface friction coefficient calculation means 260 are combined, and the road surface friction coefficient calculation means 260 is provided.
Is composed of a part of the function of the block B81 and the selector 264.

In the following, the processing performed by each of the above components in the vehicle road surface friction coefficient estimating device will be described. (1) Wheel speed difference (coefficient of road surface unevenness) β When the driving wheel transmits the driving force to the road surface, the road speed of the driving wheel vibrates due to the road surface unevenness. Since the vibration component of the wheel speed can be captured by the difference in rotational speed between the left and right wheels, the present device estimates the value of the road surface μ corresponding to the unevenness of the road surface based on the difference in rotational speed between the left and right wheels.

Then, here, the wheel speed difference β is obtained by the following equation. β = ddv / K = (dvrd-dvfd) / K (2.4.1.1) where dvrd: rear wheel left / right wheel rotational speed difference dvfd: front wheel left / right wheel rotational speed difference The rotation speed difference ddv obtained by subtracting the left and right wheel rotation speed difference dvfd from the left and right wheel rotation speed difference dvrd is used because the rotation speed difference between the left and right wheels is affected by the load movement of the vehicle. This is to remove the influence of load movement. Also, the constant K
Is for LSB matching.

(2) Road Surface μ Judgment Value α By the way, the rotational speed difference due to the torque movement of the left and right wheels during turning is from the actual rotational speed difference (ie, ddv) to the reference rotational speed difference (ie, theoretical rotational speed difference). ) Can be obtained by subtracting. Similar to the actual rotation speed difference, the reference rotation speed difference ddvh is obtained by subtracting the front wheel reference rotation speed difference dvhff from the rear wheel reference rotation speed difference dvhf as in the following equation (2.4.1.2).

Ddvh = dvhf-dvhff (2.4.1.2) Then, the rotational speed difference at the time of turning (this value is referred to as the turning correction rotational speed difference because the load movement is corrected. ) Ddvf is obtained by subtracting the reference rotational speed difference ddvh from the actual rotational speed difference ddv by further processing the rotational speed difference ddvfd with a low-pass filter to perform phase matching, as shown in the following equation (2.4.1.3). I am trying to obtain [(2.4.
1.4)].

Ddvfd = ddv−ddvh (2.4.1.3) ddvf = LPF [ddvfd] (2.4.1.4) FIG. 28 shows the magnitude of the driving force with respect to the slip ratio difference β ′. FIG. 9 is a diagram showing an example of the change in the height, where slip ratio difference β ′
Is the number of revolutions difference ddvf divided by the vehicle speed vb, and is calculated by the following equation. The driving force is the driving force transmitted by the wheels to the road surface.

Slip rate difference β ′ = rotational speed difference / vehicle body speed = ddvf / vb (2.4.1.5) In FIG. 28, a curve Hμ indicates the characteristics of a high μ road (a road surface having a high friction coefficient), Lμ is low μ road (road surface with low friction coefficient)
Shows the characteristics of. As shown in FIG. 28, on each road surface, the wheel (tire) grips the road surface in a region where the slip ratio is small (grip region), so that the driving force increases linearly with respect to the slip ratio. When the slip ratio increases, the wheels (tires) no longer grip the road surface (slip region), and therefore the driving force rather decreases even if the slip ratio increases.

Focusing on the increasing rate of the driving force with respect to the slip ratio in the grip area, that is, the slopes α _H and α _L of the characteristic lines Hμ and Lμ, the high μ road (Hμ) and the low μ road (Lμ) are obtained.
and μ) have different slopes α _H and α _{L in} this grip region. That is, as the road has a higher μ, the slope α _x becomes smaller, so that the road surface μ can be estimated based on the slope α _x . Such an inclination α _x is expressed as the following equation as a value of the ratio between the increase amount of the slip ratio difference β ′ and the drive force increase amount (torque movement amount) when the drive force increases from T1 to T2. be able to. Note that Tm is the amount of torque movement (T
m = T1-T2).

Α _x = β ′ / Tm (2.4.1.6) The road surface μ judgment value α is determined by the increase ratio of the driving force with respect to the slip ratio difference in the grip region, that is, the slope α _x . It corresponds to the value after various corrections, but this road surface μ judgment value α is
As shown in the following equation, the speed difference ddv due to the torque movement amount (temporary delay value of the torque movement amount in which the phases are adjusted) taf
It can be obtained by dividing f by the estimated vehicle speed vb.

Α = ddvf / taf / vb (2.4.1.7) (3) Road Surface μ Judgment Condition Then, based on the wheel speed difference β and the road surface μ judgment value α calculated as described above. The road surface μ determination is performed based on the road surface μ determination, and this road surface μ determination is performed when the following road surface μ determination conditions are satisfied.

The torque movement amount taf is a predetermined value tax [N
m] or more [taf ≧ tax (Nm)] This means that if the torque movement amount taf is less than or equal to a predetermined value tax,
Since the pressing force of the clutch portion varies and the torque movement amount is not stable, an attempt is made to make a reliable determination in a region where the torque movement amount is stable.

The road surface μ judgment value α is not negative [α ≧
0] This is because when the response delay at the time of turning back by steering the steering wheel or the tire vertical direction becomes non-linear, the road surface μ determination value α
This is because such a case is excluded because is negative. This is because the brake switch bksw is off, and during braking (that is, when the brake switch bksw is on), an influence of a speed difference other than the torque movement amount appears due to the braking force, and this is excluded.

The front and rear G (gb) is a predetermined deceleration g1 (g
1 is a negative small value) or more [gb ≧ g1] This is to limit the road surface μ determination to a range where the tack-in response control is not included. The drift determination coefficient srp is greater than or equal to a predetermined value (for example, r ₁ ² ) [srp ≧ r ₁ ² ] This is for limiting the case where all the directions of the tire are in the grip region.

The turning lateral G (gy) is smaller than a predetermined value gy1 [gy <gy1] When the turning lateral G (gy) becomes large, that is, when the turning lateral G (gy) becomes a predetermined value gy1 or more, This is because the stability factor of the vehicle becomes non-linear with respect to the lateral turning G (gy), and the theory for determining the road surface μ based on the linear region cannot be established.

(4) Vibration component of wheel speed difference β (βpp) In determining road surface μ, the vibration component βpp is calculated from the wheel speed difference β calculated as described above, and the vibration component of this wheel speed difference is calculated. βpp is used. As a result, more stable determination data can be obtained than when the wheel speed difference β itself is used, and the accuracy of the road surface μ determination can be improved. This vibration component βpp has a predetermined period T ₁ second (T ₁
Is updated every ten and several msec). Here, by sampling the wheel speed difference β every T ₁ seconds,
For example, the latest N (here, N = 12) samples of the wheel speed difference β are stored, and the maximum value βmax [see (2.4.1.8)] and the minimum value βmin of the wheel speed difference β are stored.
From [(2.4.1.9)] and the following equation (2.4.1.10), the vibration component βpp of the wheel speed difference is calculated.

Β max = max [β (n-11), ... β (n)] (2.4.1.8) β min = min [β (n-11), ... β (n)] .. (2.4.1.9) where β (n): wheel speed difference obtained in the current cycle β β (n−k): wheel speed difference obtained β cycles before the current cycle β βpp = [βmax -Βmin] (2.4.1.10) (5) Average value of road surface μ judgment value α (αh) The value of road surface μ judgment value α is calculated by calculating the average value αh.
The h is used for the road surface μ determination so as to increase the reliability of the determination. Here, similarly to the vibration component βpp of the wheel speed difference, the average value of the latest N (here, N = 12) road surface μ determination values α sampled every predetermined period T ₁ second is calculated [ (See 2.4.1.11)].

Αh = [α (n-11) + α (n-10) + ... + α (n)] / 12 ... (2.4.1.11) where α (n): Road surface μ determination value α α (n−k) obtained in the current cycle: Road surface μ obtained k cycles before the current cycle
Judgment value α Average value αh of road surface μ judgment value α calculated in this way
FIG. 30 shows the μ determination region that finally converges with respect to the vibration component βpp of the wheel speed difference. Road surface μ
In a region where both the determination average value αh and the vibration component βpp of the wheel speed difference are small, the high μ road, the vibration component βp of the wheel speed difference are detected.
In the region where p is large, the medium μ road, the vibration component βp of the wheel speed difference
In a region where p is small but the road surface μ determination average value αh is large, it is determined to be a low μ road, and the vibration component βpp of the wheel speed difference is determined.
In the medium range, if the road surface μ judgment average value αh is small, it is a high μ road or a medium μ road, and if the road surface μ judgment average value αh is medium, it is a high μ road or a medium μ road or a low μ road, a road surface μ judgment. Average value α
If h is large, it is determined to be a medium μ road or a low μ road, and in a region where the vibration component βpp of the wheel speed difference is small and the road surface μ determination average value αh is medium, it is determined to be a high μ road or a low μ road.

(6) Fuzzy inference (membership function,
hig3, mid3, low3) Average value αh of road surface μ judgment value and vibration component β of wheel speed difference
A road surface μ determination index is created from pp. The road surface μ determination index is the road surface μ determination value αh and the vibration component βp of the wheel speed difference.
Obtained by fuzzy reasoning with p as input condition. Here, each road surface based on the average value αh of the road surface μ determination values [that is,
High μ road, medium μ road, low μ road]
1, low1 and each road surface based on the vibration component βpp of the wheel speed difference [ie, high μ road, medium μ road, low μ road]
Road] to find the goodness of fit hig2, mid2, low2,
By the minimum method (minimum method), each road surface [ie high μ road,
For each [medium μ road, low μ road], select the smaller one of these suitability values, and select that road surface [that is, high μ road, medium μ road, low μ road]
Road] fitness (fuzzy number) hig3, mid
3, low3.

That is, with respect to the average value αh of the road surface μ judgment values, membership functions as shown in (A), (B) and (C) of FIG. 29 are set, and the membership of FIG. The fitness degree hig1 to the high μ road corresponding to the road surface μ determination value αh is obtained from the function, and the fitness degree mi to the medium μ road corresponding to the road surface μ determination value αh from the membership function of FIG.
Then, d1 is obtained, and the conformity low1 to the low μ road corresponding to the road surface μ determination value αh is obtained from the membership function of FIG.

Further, with respect to the vibration component βpp of the wheel speed difference, membership functions as shown in (D), (E) and (F) of FIG. 29 are set, and the membership function of FIG. 29 (D) is set. From the membership function of FIG. 29 (E) to the high μ road corresponding to the vibration component βpp of the wheel speed difference, the fitness mid2 to the medium μ road corresponding to the vibration component βpp of the wheel speed difference. Then, from the membership function of FIG. 29 (F), the fitness low2 for the low μ road corresponding to the vibration component βpp of the wheel speed difference is obtained.

Then, the goodness of fit hig1 and hi for the high μ road
g2 is compared, and the one with the smaller value is the high μ road conformance degree hig
Select 3. Also, the suitability mid1 and mi for the medium μ road
d2 is compared, and the one with the smaller value is the medium μ road fitness mid.
Select 3. Furthermore, the low-μ road compatibility low1 and l
ow2 is compared, and the one with the smaller value is the low μ road compatibility lo
Select to w3.

(7) Counter function (high, mid, l
ow) As described above, the high μ road conformity high3, the medium μ road conformance m
When the id3 and the low μ road conformity low3 are obtained, the weights (high, mid, low) of each road surface [that is, the high μ road, the medium μ road, and the low μ road] are obtained based on these. Here, high μ road compatibility high3, medium μ road compatibility mid
3, the low μ road compatibility low3 is continuously calculated and cumulatively evaluated, that is, by using a so-called learning function, the weight (h
ig, mid, low) respectively. That is, the counter values (empirical values) Nh, Nm, Nl of each road surface μ (these are collectively referred to as Ni. (I = h, m, l)] is set, and the high μ road compatibility high obtained as described above is set.
3, the intermediate μ road suitability mid3, the low μ road suitability low, the counter values (experience values) Nh, Nm, Nl are respectively increased or decreased to increase the high μ road weight hig, the medium μ road weight mid, The weights low of the low μ road are obtained respectively.

That is, the high μ road compatibility degree hig3, the medium μ road suitability
The judgment value regarding the goodness of fit mid3 and the low μ road compatibility low3 and
And then h₁, H₂, H₃, H_Four(H₁<H₂
<H ₃<H_Four), M₁, M₂, M₃, M_Four(M₁<M₂
<M₃<M_Four), L₁, L₂, L₃, L_Four(L₁<L₂
<L₃<L_Four) Is set, and each goodness of fit hig3, m
By comparing id3 and low3 with these judgment values, FIG.
And update the counter amount as shown below. Note that n
n is a natural number.

[0161] When the high-μ road hig3> h _4, when _{Nh = Nh + nn h 3 <} hig3 ≦ h 4, when _{Nh = Nh + 1 h 2 <} hig3 ≦ h 3, Nh = Nh h 1 < a hig3 ≦ h ₂ When Nh = Nh-1 hig3 ≦ h ₁ , Nh = Nh-nn However, when the counter range is 0 ≦ Nh ≦ Nhmax medium μ road mid3> m ₄ , Nm = Nm + nn m ₃ <mid ₃ ≦ m ₄ When, Nm = Nm + 1 m ₂ <mid3 ≦ m ₃ , Nm = Nm m ₁ <mid3 ≦ m ₂ , Nm = Nm−1 mid3 ≦ m ₁ , Nm = Nm−nn However, the counter range when the 0 ≦ Nm ≦ Nmmax low μ road _{low3> l 4, Nl = Nl} + when _{nn l 3 <low3 ≦ l 4} , when _{Nl = Nl + 1 l 2 <} low3 ≦ l 3, Nl = Nl l 1 <low3 when ≦ l _2, Nl When _{Nl-1 low3 ≦ l 1,} Nl = Nl-nn However, the counter range, 0 ≦ Nl ≦ Nlmax · high μ road weight hig, middle μ road weight mid, this calculation of the weight low low μ road , The counter value (experience value) N corresponding to each road surface μ
When h, Nm and Nl are calculated, these counter values N
Weights hig, mi of each road surface μ according to h, Nm, Nl
d and low are obtained by a map.

That is, the high μ road weight hig is obtained from the high μ road counter value Nh by the map shown in FIG. 32A, and the medium μ road counter value is obtained by the map shown in FIG. 32B. The weight μ of the medium μ road is calculated from Nm, and the weight low of the low μ road is calculated from the counter value Nl of the low μ road according to the map shown in FIG. (8) Calculation of Road Surface μ Judgment Coefficient γ Thus, the high μ road weight hig and the medium μ road weight mi
d, when the low weight of the low μ road is obtained, the weight average value γ of these weights hig, mid, low is obtained from the following equation, and this weight average value γ is used as the road surface μ determination coefficient γ.

Γ = (w1 * hig + w2 * mid + low) / (hig + mid + low + α) (2.4.1.12) Note that w1 and w2 in the above equation are weights for calculating the weight average value γ. Where w1 is the weight value of the weight hig, w2 is the weight value of the weight mid, and the weight value of the weight low is 1, and the weight value is such that w1 is the largest and w2 is the largest (w1>w2> 1). Further, α is an adjustment value, and for example, α = 1.

Weight of each road surface, that is, weight hi of high μ road
When g, the weight of the medium μ road, mid, and the weight of the low μ road are high ₁ , mid ₁ , and low ₁ , respectively, the value (area) obtained by multiplying the weight of each road surface by each weight value is 33
As shown in (A), they are Sh, Sm, and Sl, respectively. Then, γ is obtained by summing the sum Sh + Sm + Sl of the area values Sh, Sm, and Sl into the value (high + mid + low +
Since it is divided by α), the weight average value (road surface μ determination coefficient) γ is obtained as a value on the horizontal axis of FIG. 33 (B).

2.4.2 Road Surface μ Estimation During Non-linear Turning Next, the road surface μ estimation during non-linear turning will be described. The determination as to whether or not the vehicle is nonlinear turning is based on the side slip coefficient dgy of the tire. The road surface μ during the nonlinear turning is based on the actual lateral G (rgy) when the side slip coefficient dgy becomes a non-linear value. Estimate as follows. Here, FIG.
As shown in (1), during the non-linear turning, the low μ road determination is performed when the forced low μ determination condition is satisfied, and the high μ road determination is performed when the forced high μ determination condition is satisfied.

(1) Forced Low μ Judgment Condition For the forced low μ judgment condition, each of the following conditions is to be satisfied. The side slip coefficient dgy has a non-linear magnitude [dgy> dgy1]. The magnitude of the actual lateral G (rgy) is less than the set value (rgy1) [| rgy | <rgy1]. The steering wheel angular velocity dθh is less than the set value (dθh1) [dθh <dθh1].・ The vibration component βpp of the slip ratio difference is the set value (βpp1)
It is less than [βpp <βpp1]. -The vehicle speed vb is less than the set value (vb1) [vb
<1vb]. The state in which all of the above conditions are satisfied must continue for a predetermined duration ct1 (ct1 is, for example, 100 msce).

When the above condition (AND condition) is satisfied,
The road surface μ judgment coefficient γ, each road surface μ
The counter values Nh, Nm and Nl corresponding to the above are respectively set as follows. γ = 0, Nh = 0, Nm = 0, and Nl
= Nlmax (2) Forced high μ determination condition The forced high μ determination condition is based only on the value of the actual lateral G (rgy) when the sideslip coefficient dgy becomes non-linear. That is, the actual lateral G (rgy) is a preset value rgy2
Is larger than [| rgy |> rgy2]. The actual lateral G (rgy) becomes equal to or larger than the predetermined value rgy2 because it cannot occur unless the road is a high μ road.

Further, if the high μ determination condition is met, it is not preferable that the control is suddenly changed to the high μ road control, so that the following condition is satisfied when the high μ determination condition is met. In addition, the road surface μ determination coefficient γ (γ2) and the counter values Nh, Nm and Nl corresponding to the respective road surface μ are set as follows. Note that mm is a natural number larger than nn described above.

Γ2 = γ1 + 10, and Nh = Nh + m
m and Nm = Nm-mm and Nl = Nl-mm where γ2 ≦ γmax, Nh ≦ Nhmax, Nm ≧ 0, N
Let l ≧ 0. In this way, the high μ road is gradually approached. 2.4.3 Road Surface μ Estimation at Start (1) μ Split Road Judgment Condition Here, as shown in FIG. 34, whether the road surface friction coefficient (road surface μ) of the left and right wheels at start is different (μ split road). Whether or not the road surface μ is estimated is estimated based on this determination. The μ-split road is determined as follows mainly based on the torque moving direction taf and the left-right wheel speed difference dvrd of the rear wheels. However, the right turn and the right moment are positive.

Condition 1 taf> taf1 and dvrd <-vd1 or taf <-taf1 and dvrd> vd1 (taf1 is a positive set value, vd1 is a positive set value) That is, the torque movement direction taf is leftward (taf> taf).
f1) and the left-right wheel speed difference dvrd is negative (dvrd <-
vd1) [That is, the right wheel rotates faster than the left wheel]
Or, the torque movement direction taf is rightward (taf <−
taf1) and the left-right wheel speed difference dvrd is positive (dvrd
> Vd1) [that is, the left wheel is rotating at a higher speed than the right wheel]. In other words, this means that the wheel to which the torque has moved is slipping.

The vehicle body speed vb is less than a predetermined value vb2 (vb <
The vehicle speed is vb2). This is a condition for starting. The steering angle θh is a low steering angle less than a predetermined value θh1 (θh <θh1). This is a condition that the vehicle is in a straight traveling state. The throttle opening tps is larger than a predetermined value tps1 (tps> tps1) (that is, there is a start operation).

The state where all of the above conditions are satisfied is a predetermined duration ct2 (ct2 is 100 mse, for example).
c) Continue above. If the above condition (AND condition) is satisfied, it is determined that the road is a μ-split road and is a completely low μ road, and a coefficient myu, a road surface μ determination coefficient γ (γ2), and a counter value N corresponding to each road surface μ
Set h, Nm, and Nl as follows, respectively.

Myu = 1, and γ2 = 0, and Nh
= 0, and Nm = 0, and Nl = Nlmax, where myu is a coefficient for preventing hunting between the forced high μ determination condition and the forced medium μ determination condition, and the vehicle body speed vb is equal to or less than the predetermined value vb3 (vb ≦ vb3), the torque movement amount ta
The magnitude of f is less than or equal to the set value taf2 (| taf | ≦ taf
In the case of 2), myu = 0.

Further, even when it is not judged as a μ-split road, the following forced low μ judgment condition, forced medium μ judgment condition and forced high μ judgment condition are set at the time of starting, and if each condition is satisfied: , Low μ road, medium μ, respectively
Road, high μ road. (2) Forced low μ determination condition In this case, even if one wheel slips when the vehicle starts straight ahead, it is forcibly determined to be a low μ road (ie, μ split road).

Therefore, the forced low μ determination condition is as follows. The steering angle θh is less than a predetermined value θh1 (θh <
The steering angle is θh1) low (that is, the vehicle is traveling straight ahead). The vibration component βpp of the wheel speed difference is a predetermined value ( βpp
It is less than 2 ) (βpp < βpp2 ) (that is, the vibration component βpp is not large).

The throttle opening tps is a predetermined value (tps
It is larger than 1) (tps> tps1) (that is, there is a start operation). The vehicle speed vb is a low vehicle speed that is less than a predetermined value vb2 (vb <vb2) (that is, the vehicle is starting). Even one wheel has slipped. That is, each wheel speed v
Any of fl, vfr, vrl, and vrr is the vehicle speed v
Being higher than b by a predetermined speed (v1) or more (v
fl> v1, or vfr> v1, or vrl> v1, or vrr> v1).

The state where all of the above conditions are satisfied is a predetermined duration ct3 (ct3 is, for example, 100 msc).
e) Continue above. When the above condition (AND condition) is satisfied, it is determined that the road is a completely low μ road, and the coefficient myu, the road surface μ judgment coefficient γ (γ2),
Counter values Nh, Nm, Nl corresponding to each road surface μ are set as follows.

Myu = 1, and γ2 = 0, and Nh
= 0, Nm = 0, and Nl = Nlmax (3) Forced medium μ determination condition When the vibration component of the wheel is large at the time of starting, forcibly take an intermediate value between medium μ and low μ To do. However, when the forced low μ judgment and the forced high μ judgment are made, my
From u = 1 to myu = 0 (that is, vb ≦ vb
3, and until | taf | ≦ taf2), this determination is not performed.

Therefore, the forced medium μ determination condition is as follows. The vibration component βpp of the wheel speed difference is a predetermined value ( β
pp2 ) (βpp> βpp2 ) (that is, the vibration component βpp is large). The throttle opening tps is larger than a predetermined value (tps1) (tps> t)
ps1) (that is, there is a start operation).

The vehicle body speed vb is less than a predetermined value vb2 (vb <
vb2) Low vehicle speed (that is, when starting). myu = 0. The state in which all of the above conditions are satisfied should continue for a predetermined duration ct4 (ct4 is, for example, 200 msce) or more.

When the above condition (AND condition) is satisfied,
The road surface μ determination coefficient γ (γ2) and the counter values Nh, Nm, and Nl corresponding to each road surface μ are set as follows. γ2 = γ ₁ and Nh = 0 and Nm = Nmmax,
Moreover, Nl = Nlmax, and γ ₁ is a value of about ¼ of γmax.

(4) Forced High μ Judgment Condition When the wheel does not slip at a certain acceleration or more at the time of starting, the high μ judgment is forcibly made. However, when the forced low μ determination is made, from myu = 1 to myu = 0 (that is, vb ≦ vb3, and | taf | ≦ ta
Until f2), this determination is not performed.

Therefore, the forced high μ determination condition is as follows. The vibration component βpp of the wheel speed difference is a predetermined value ( β
pp2 ) (βpp> βpp2 ) (ie,
Vibration component βpp is not large). The throttle opening tps is larger than a predetermined value (tps2) (tps> t
ps2) (that is, there is an acceleration operation above a certain level).

The vehicle speed vb is less than a predetermined value vb2 (vb <
vb2) Low vehicle speed (that is, when starting). myu = 0. G (gb) before and after calculation is a predetermined value gb1 or more (gb ≧ gb
1) (that is, when starting).

The average speed vf of the front wheels is sufficiently close to the vehicle speed vb (| vf | <vb) and the average speed vr of the rear wheels is
Is sufficiently close to the vehicle speed vb (| vr | <vb). These indicate that the wheels do not slip. Each of the above conditions
The state where both are established continues for a predetermined duration ct5 or more.
Do (see FIG. 34). When the above condition (AND condition) is satisfied, it is determined to be high μ, and the road surface μ determination coefficient γ and the counter values Nh, Nm, and Nl corresponding to each road surface μ are set as follows.

Myu = 1, and γ2 = 0, and Nh
= Nhmax and Nm = 0 and Nl = 0 2.4.4 Output value setting (1) Output value setting of each control amount (γtb, γtc, γt
d, γte, tb, tc, td, te) As described above, as the control amounts, the target ΔN tracking control amounts tbh, tbl, the acceleration turning control amounts teh, tel, the tuck-in corresponding control amounts tdh, tdl, the steering transient Response control amounts tch and tcl, and a high μ road control amount (high road surface friction resistance corresponding control amount) and a low μ road control amount (low road surface friction resistance corresponding control amount) are set, respectively. The output control amount tad is calculated while interpolating both control amounts in an interpolation manner according to the road surface μ determination coefficient γ as the road surface friction coefficient calculated by the road surface friction coefficient calculating means.

That is, as shown in FIG. 35, for each control amount, there is no step between the high μ road and the low μ road according to the value of the road surface μ determination coefficient γ. The value after gain adjustment is used as the output value (output gain).

For example, if the control amount for the high μ road (control gain for the high μ road) is txh and the control amount for the low μ road (control gain for the low μ road) is txl, the output value (output gain) tx is
It is calculated from the road surface μ determination coefficient γ by the following formula. The road surface μ determination coefficient γ is a value of 0 to γmax. Note that, here, when the road surface μ determination coefficient γ is 0, it is a low μ road, and the road surface μ determination coefficient γ is
Is the high μ road when γmax is, and between the low μ road and the high μ road,
That is, when the road surface μ determination coefficient γ is an intermediate value between 0 and γmax, it is called a medium μ road.

Tx = {γ · txh + (γmax−γ) · txl} / γmax = {(txh−txl) · γ + γmax · txl} / γmax = {(txl−txh) · (γmax−γ) + γmax · txh} / Γmax ··· (2.4.1.1) Further, here, the control gain (control amount) tx is set to be shifted to a high μ road side or a low μ road side. Then, the output value is finely adjusted.

The control gain tx is set to the high μ side (target Δ
N follow-up control: tb) The output value fine adjustment formula to the high μ side is tx for the corrected output value.
If a and the output value fine adjustment coefficient are a (a> 1), the following equation is obtained. txa = {a (txh−txl) · γ + γmax · txl} / γmax = {a · γ · txh + (γmax−a · γ) · txl} / γmax where 0 ≦ a · γ ≦ γmax (2.4.1.2) Note that txa is clipped to the upper limit at txh due to 0 ≦ a · γ ≦ γmax.

As described above, when the high μ and low μ control amounts are interpolated, the reflection degree of the high μ road control amount is set to be larger than that of the low μ road control amount. Such setting to the high μ side is performed with respect to the control gain tb of the target ΔN tracking control. The control gain tx is set between the high μ side and the low μ side [Steering angular velocity proportional control (transient response control): tc, tuck-in response control: td] In this case, the output value fine adjustment is not substantially performed. , The control gain tx is calculated using the above equation (2.4.1.1). Such calculation is performed for each of the control gain (control amount) tc of the steering angular velocity proportional control (transient response control) and the control gain (control amount) td of the tack-in control.

The control gain tx is set to the low μ side (acceleration turning control: te). The output value fine-adjustment formula for the low μ side is the corrected output value tx
When b and the output value fine adjustment coefficient are b (b> 1), the following equation is obtained. txb = {b (txl-txh) * ([gamma] max- [gamma]) + [gamma] max * txh} / [gamma] max = [b * ([gamma] max- [gamma]) * txl + {[gamma] max-b * ([gamma] max- [gamma])) * txh] / [gamma] max. (2.4.1.3) As 0 ≦ b · γ ≦ γmax, txb is clipped by txl at the lower limit.

As described above, when the high μ and low μ control amounts are interpolated, the degree of reflection of the low μ road control amount is set to be larger than that of the high μ road control amount. The setting to such a low μ side is for setting the control gain te of the acceleration turning control.
Do about. Output values (output gains) tx, txa, txb obtained by appropriately performing such output value fine adjustment.
Can be illustrated as shown in FIG. 24 in terms of the road surface μ. In FIG. 24, the alternate long and short dash line indicates the output value txa obtained by finely adjusting the control gain tx to the high μ side.
(That is, the characteristic of the target ΔN tracking control amount tb) is shown, and the solid line of the output value tx (that is, the tuck-in response control amount td, the steering transient response control amount tc) when the control gain tx is not finely adjusted. Shows the characteristic, and the broken line of is the control gain t
The characteristic of the output value txb (that is, the acceleration turning control amount te) obtained by finely adjusting x to the low μ side is shown.

As shown in FIG. 24, the lower the road surface μ (the smaller the road surface μ determination coefficient γ), the smaller the control amount (output value) tx. However, the lower the road surface μ, the higher the control effect. Therefore, in order to obtain the same control effect, the control amount (output value) tx needs to be smaller as the road surface μ is lower. Further, the reason why the target ΔN tracking control amount tb is increased on the medium μ road is that the target ΔN tracking control is a control that can maintain the behavior stability of the vehicle even if the road surface μ is relatively low. This is because it is desired to emphasize the target ΔN tracking control and positively stabilize the behavior of the vehicle. The reason why the acceleration turning control amount te is decreased on the medium μ road is that the acceleration turning control amount te has a property that it is difficult to secure the behavior stability of the vehicle when the road surface μ is low.

Further, in the characteristic of the output value tx in FIG. 24, the inclination can be changed by a parameter such as a constant peculiar to the vehicle. As a result, control matching, that is, fine adjustment of the output value, can be performed according to the vehicle, and more stable control can be performed.
There is also the advantage that the same basic logic can be used depending on the vehicle.

(2) High-pass processing & final output value tad Here, in order to solve the response delay, FIG. 35 and FIG.
As shown in 6, the high-pass processing is performed for the target ΔN tracking control amount tb, the tuck-in corresponding control amount td, and the acceleration turn control amount te. This processing is performed, for example, to correct the control delay with respect to the high frequency input due to the fast steering by the high-pass processing and to advance the phase of each of these control terms.

That is, when the actuator (rotational propulsive force distribution adjusting mechanism) is driven, it is inevitable that the actuator delays in response to the output of the control signal.
Therefore, it is necessary to perform processing so that the response delay of the actuator does not deteriorate the control performance. Further, the control signal includes, for example, a control amount (transient control amount) calculated based on a driving operation state such as a steering wheel angle or a steering angle (including a steering angular velocity) θh and a throttle opening tps.
For example, steering transient response control amount (steering angular velocity proportional control amount) t
c, a control amount calculated based on the vehicle behavior such as a left / right wheel rotational speed difference or lateral acceleration occurring in the vehicle (vehicle behavior corresponding control amount), for example, a target ΔN tracking control amount tb, a tuck-in corresponding control amount td, acceleration There is a turning control amount te. Driving operation is originally the main element of control command, and command delay is not a problem for the control amount according to driving operation, but the behavior of the vehicle occurs as a result of the control command. The control amount set based on this is already delayed when the control signal is issued, which may cause a problem.

For example, when the vehicle behavior changes abruptly, such a delay in the output of the control amount significantly reduces the control performance. Therefore, in the present device, for example, in order to correct the delay of each control response to the steering input, the control amount according to the vehicle behavior, that is, the target ΔN tracking control amount tb, the tuck-in corresponding control amount td, and the acceleration turning control amount te are set. The high pass processing is performed to speed up the output of the control signal. As described above, since the steering transient response control amount (steering angular velocity proportional control amount) tc is a control for advancing the phase, it does not need to be corrected and high-pass processing is not performed.

Further, in the present control, each control amount tb, td,
The final output value tad is determined by adding te and tc. That is, the ECU 42 functions as an output control amount calculation unit that individually calculates the control amounts tb, te, td, and tc based on various parameters and then integrates them to obtain the output value tad. . Therefore, here, the control amounts tb and t required for high-pass processing are set.
For d and te, these are added in advance, and then the high-pass processing is performed on the added value tfd (= tb + td + te).

High-pass processing The high-pass processing is processing for extracting only the high-frequency component of each control output by a high-pass filter. Here, the high-pass processing is performed for the added value tfd of the control amounts tb, td, te for performing the high-pass processing. To obtain a high-pass processed value tff.

Tfd = tb + td + te (2.4.1.4) tff = HPF [tfd] (2.4.1.5) This high-pass processing causes the control output as shown in FIG. 37 (A). A high-pass processed signal tff as shown in FIG. 37B is output from the signal tfd. That is, in the high-pass processing, only the large-sized portion of the differential value of the control output signal tfd is output as a signal (high-pass processed value calculation means).

Further, the processed value tff thus high-pass processed is added to the control output signal tfd (= tb + td + te) which is the target of the high-pass processing [FIG. 37].
(See (C)], the output control amount (total value) tf is obtained (output control amount calculation means). tf = tfd + tff (2.4.1.6) As shown in FIG. 37, the processed value tff is corrected by the gain (high-pass coefficient) kf (that is, tff * kf).
As a), the balance with another control amount may be adjusted.

The final output value (tad) output control amount calculation means adds the output control amount tf and the steering transient response control amount (steering angular velocity proportional control amount) tc which does not perform high pass processing, as shown in the following equation. Thus, the final output control amount (final output value) tad is calculated. tad = tf + tc ··· (2.4.1.7) ・ In torque transfer control between the left and right wheels of the limiter, if the torque transfer amount is too large, the behavior stability of the vehicle may be deteriorated, so in this control In accordance with the friction coefficient state of the road surface (road surface μ state), the magnitude of the torque movement amount between the left and right wheels is limited to the maximum value (this is referred to as limit).

This limit value or maximum value limit is shown in FIG.
As shown in the block B83 in FIG. 6, the relationship between the road surface μ determination coefficient γ and the straight line LIM is set. That is, the limit value limit is calculated by the following equation. limit = mg · γ + tal1 (2.4.1.8) However, mg is the slope of the straight line LIM, and tal1 is l.
This is the minimum value of imit. As shown in block B83 of FIG. 36, this minimum value tal1 is a limit value limit corresponding to the road surface μ determination coefficient 1 of the low μ road.
tal2 is a limit value limit corresponding to the road surface μ determination coefficient γmid for medium μ roads, and tal3 is a limit value limit corresponding to the road surface μ determination coefficient γmax for high μ roads. In addition,
The road surface μ determination coefficient γmid for medium μ roads is set to 1/2 of the road surface μ determination coefficient γmax for high μ roads (γmid = γmax /
2).

With such a limit value limit, the final output value tad is limited as follows. The following equation also takes into consideration the case where the final output value tad becomes negative depending on the torque moving direction. −Limit ≦ tad ≦ limit ··· (2.4.1.9)

2.5 In the actuator drive drive processing (actuator drive processing or proportional valve / directional valve switching control processing), as shown in FIG. 21, the output value (torque movement amount) tad is received, and this output value is received. tad
Is converted into an actuator drive signal for performing torque movement in a direction and amount according to the output value tad, a proportional valve control signal is output to the proportional valve 106 according to the torque movement amount, and the direction is changed according to the torque movement direction. Valve (direction switching valve) 10
7 to output a directional valve control signal to these proportional valves 10
6. Drive the directional valve 107. At the same time, a display command signal is output to the indicator lamp 110 (reference numeral 1).
(Refer to FIG. 3 for 06, 107 and 110).

The control of the proportional valve 106 and the directional valve 107 is performed by the method disclosed in Japanese Patent Laid-Open No. 7-156681, for example. For example, proportional valve 106
Regarding, regarding the final output value ta, the torque transfer-current map (see FIG. 22) and the current correction map (see FIG. 23).
Is used to convert to a target current baseh for control.

2.5.3 Indicator Display Control Here, the indicator display control related to the display device of the vehicle power transmission control device of this embodiment will be described. The power transmission control device for the left and right wheels for the vehicle is as shown in FIG. 25 so that the driver can understand the control state when the torque transfer control (that is, the power transmission control) is performed between the left and right wheels. A display device (hereinafter referred to as an indicator) 201 having a display unit 202 is provided.

In this example, the display unit 202 is configured as a three-dot display including three lighting units 202A, 202B, and 202C that are LEDs. Then, when the control state is weak (that is, the torque movement amount is small), only the first lighting unit (LED1) 202A is lit, and when the control state is medium (that is, the torque movement amount is medium), the first lighting is performed. Unit (LED1) 202A and second lighting unit (LED2) 20
When 2B and 2B are lit and the control state is strong (that is, the torque movement amount is large), all of the first lighting unit (LED1) 202A to the third lighting unit (LED3) 202C are lit. . As a result, the control state can be grasped as the number of lit dots (the number of LEDs).

Of course, when the control state is weak, only the first lighting unit (LED1) 202A, only the second lighting unit (LED2) 202B when the control state is medium, and only the third lighting unit (LED3) 202C when the control state is strong. Although it is possible to display each of them, it is preferable to increase the number of display dots according to the strength of the control state in order to grasp the state more quickly.

As such a display device, display means other than LED such as liquid crystal display or electric light display may be used, and a graph display (for example, bar graph) or numerical display may be used instead of dot display. Good. In addition, display devices for the left wheel and right wheel may be provided respectively.
In this case, there is an advantage that the torque movement state can be clearly understood.

By the way, the strength of the above-mentioned control state is a control degree such as the degree to which the torque movement control (power transmission control) is performed. What is useful for the driver in the driving operation is the torque movement control. It is a control effect such as how much (power transmission control) acts on the behavior of the vehicle. Therefore, we would like to display the control status so that the control effect can be understood.

On the other hand, it is suitable and easy to use the output value tad which is the final control amount as an index of the normal control degree. However, as described above, such a control amount tad does not always correspond to the control effect, as described in the section “Problems to be solved by the invention”. That is, the control amount tad is set according to the road surface μ (road surface friction coefficient), which is a type of traveling environment of the vehicle, but the control effect is higher on the low μ road than on the high μ road. The lower the road surface μ, the smaller the control amount tad is set. Therefore, when the degree of control is displayed so as to be proportional to the control amount tad in this way, on the low μ road,
Even if the degree of control displayed is low, the control effect that actually appears is large. Conversely, high μ
On the road, the degree of control displayed is high, but the control effect that actually appears is small.

Therefore, the present display device does not directly correspond to the control amount tad itself, but to the ratio (= tad / tadmax) of the set control amount tad to the maximum control amount tadmax in the road surface μ (road surface friction coefficient). The display amount is displayed based on the display amount. This maximum control amount tadmax corresponds to the road surface μ, and since the present device uses the road surface μ determination coefficient γ as the amount indicating the road surface μ (road surface friction coefficient), the present display device uses the control amount tad as The display amount is set based on the road surface μ determination coefficient γ. The function of setting the display amount (converting into the display amount) from the control amount tad or the like is referred to as a conversion unit.

In this converting means, as shown in FIG.
As the display amount determination reference value of the control amount tad, six types are provided according to the road surface μ determination coefficient γ. In FIG. 26, the horizontal axis represents a value (γ / γmax) obtained by dividing the road surface μ determination coefficient γ by the maximum coefficient value γmax, and the vertical axis represents the control amount tad. And
The straight lines t _{1 to} t ₆ in the figure are the values (γ /
The display amount determination reference value for γmax) is shown. In particular, the display amount judgment reference value is t _L1 to t _L6 on a low μ road (a road surface with a road surface μ judgment coefficient γ of 0), and is displayed on a high μ road (a road surface with a road surface μ judgment coefficient γ of γ max). Reference value is t _H1
~ T _H6 . The straight lines t _{1 to} t ₆ indicating the display amount determination reference values are points t _{L1 to} t _L6 and points t _{H1 to} t _H6 , respectively.
It is a straight line connecting the and.

Of these display amount judgment reference values, t
₂(T_L2, T_H2Is a base for lighting LED1.
Quasi value (display value), t₁(T_L1, T_H1Is included) is L
It is a reference value (extinguishing value) for turning off ED1. Well
T_Four(T_L4, T_H4(Including) lights LED2
Reference value (display value) for₃(T_L3, T_H3Including
Is a reference value (extinguishing value) for turning off the LED 2.
It And t₆(T_L6, T_H6Includes LED)
It is the reference value (display value) for lighting, and t_Five(T_L5, T
_H5Is a reference value for turning off the LED3 (off)
Value).

As shown in the figure, the smaller the road surface μ determination coefficient γ, the larger the display amount even if the control amount tad is small. This is because the smaller the road surface μ determination coefficient γ, the greater the control effect. Therefore, in order to obtain the similar control effect, the smaller the road surface μ determination coefficient γ, the more the control amount tad.
Is set to a small value (of course, the maximum control amount tadmax is also small). Therefore, when attention is paid to the control effect, the smaller the road surface μ determination coefficient γ, the larger the display amount needs to be even if the control amount tad is small.

Therefore, although the maximum control amount tadmax is set smaller as the road surface μ determination coefficient γ is smaller, each reference value is contradictory to this when the conversion gain for converting the control amount tad into the display amount is considered. The smaller the road surface μ determination coefficient γ, the larger the setting. Further, the reference value for turning on (display value) is set to a value larger than the reference value for turning off (off value) because a so-called hysteresis for stabilizing the display is provided, As a result, even if the control amount tad slightly changes in the vicinity of the reference value, the display state does not change and stable display can be realized.

3. Operation of this device and effects of this device 3.1 Operation of this device Since this device is configured as described above, for example, FIG.
Control is performed as shown in FIG. That is, first, control is started based on various initial setting inputs, and first, in step S10, the input calculation process as shown in FIG. 6 is executed (see item 2.1 Input calculation process). Then, in step S20, a drift determination logic as shown in FIG. 8 is executed based on the result of this input calculation processing (item 2.2).
See Drift Judgment Logic). Further, step S3
In step 0, the vehicle motion control logic is executed based on the result of the input calculation process and the drift judgment (see item 2.3 Vehicle motion control logic).

In this vehicle motion control logic, the target ΔN
Follow-up control (see item 2.3.1 Target ΔN follow-up control),
Acceleration turning control (Refer to item 2.3.2 Acceleration turning control),
Control amounts tb, td, te, tc for tack-in control (see item 2.3.3 Tack-in control) and steering transient response control (see item 2.3.4 Steering transient response control)
For each of these controlled variables tb, td, te, t
c is a high μ road control logic as shown in FIG.
With the low μ road control logic as shown in FIG. 5, the control amounts tbh, tdh, teh, tch on the high μ road and the control amounts tbl, tdl, tel, tcl on the low μ road.
Calculate as

Then, the processing proceeds to step S40, and the μ judgment logic is executed (see item 2.4 Road surface μ estimation). In this μ determination logic, the road surface μ determination coefficient γ is set (step S50), and road surface μ determination is performed (step S50).
60) and various output values are set (step S7).
0). Next, in step S80, the actuator driving logic is executed as shown in FIG. 21 (item 2.
5 Actuator drive). That is, in response to the output value (torque movement amount) tad, a proportional valve control signal is output to the proportional valve 106 according to the torque movement amount according to the output value tad, and according to the torque movement direction according to the output value tad. To output a directional valve control signal to the directional valve (direction switching valve) 107 to drive the proportional valve 106 and the directional valve 107. At the same time, a display command signal is output to the indicator lamp 110.

Such processing is performed for each required period T ₁ through the determination step S90. 3.2 Effects of this device 3.2.2 Effects of vehicle road surface friction coefficient estimation device In the vehicle road surface friction coefficient estimation device, a parameter indicating a road surface unevenness, which is a wheel speed difference (road surface unevenness coefficient) β, and a road surface are used. Since the road surface μ determination is performed based on the μ determination value α, which is a parameter indicating the slipperiness of the road surface, the road surface μ determination can be accurately performed. Further, the vibration component βpp of the wheel speed difference β is used without directly using the wheel speed difference β, or the average value αh of the road surface μ determination value α is used without directly using the road surface μ determination value α.
Since the road surface μ judgment is performed using, the accuracy of the road surface μ judgment,
The reliability can be increased.

Then, from the vibration component βpp of the wheel speed difference β and the average value αh of the road surface μ judgment value α, the high μ road,
Road friction by calculating the index (fitness) corresponding to each condition (road surface condition) of medium μ road and low μ road, and also by continuously calculating and accumulating this index (fitness) Since the coefficient (a road surface μ determination coefficient corresponding to the road surface friction coefficient) γ is calculated, an appropriate road surface μ determination coefficient γ can be obtained easily with less estimation error, and the vehicle left / right wheel power transmission control device There is an advantage that it greatly contributes to the improvement of control performance.

In particular, in this embodiment, fuzzy inference is used, and the membership function for determining the index (fitness) corresponding to each state (road surface state) of the high μ road, the medium μ road, and the low μ road is used as a vehicle. It is possible to easily obtain an extremely accurate road surface μ determination coefficient γ by setting the value in accordance with the type and the like. Also,
The calculation of the road surface μ determination coefficient γ (that is, γ1) based on the vibration component βpp of the wheel speed difference β and the average value αh of the road surface μ determination value α as described above is performed during steady running of the vehicle (particularly during steady turn). Therefore, the reliable road surface μ determination coefficient γ can be obtained.

Further, with respect to the road surface μ determination coefficient γ1 obtained during such a steady running, during a specific running other than the steady running (specifically, during starting, non-linear running, or μ split running) The road surface μ determination coefficient γ2 is estimated by a method different from that during steady running, and the road surface μ determination coefficient (first parameter) γ1 obtained during steady running and the road surface μ determination coefficient (second parameter) obtained during specific running ) Γ2 and (2) are integrated (selected) to obtain the final road surface μ determination coefficient γ, so that the road surface friction coefficient can be estimated in a wider traveling state including other than steady traveling, There is an advantage that it greatly contributes to the improvement of the control performance of the power transmission control device between the left and right wheels for use.

In the μ-split state, the torque is moved from the low μ wheel side to the high μ road wheel side by the torque movement control according to the road surface μ. Therefore, as shown in FIG. The driving force transmitted from the vehicle to the road surface is increased, so that the vehicle can be started and accelerated more quickly and efficiently. 3.2.3 Effect of control for vehicle road surface friction coefficient (road surface μ) In the present embodiment, during normal control (other than drift control), the control amount is reduced when the road surface friction coefficient is low. The control amount is set to be large when the road surface friction coefficient is high, and at the time of drift control, the control amount becomes an intermediate value so that the change in the control amount according to the road surface friction coefficient is small. By performing the correction (gain adjustment), the correction according to the road surface friction coefficient is suppressed, but this device is not limited to this, and during drift control, the control amount correction according to the road surface friction coefficient is performed. An intermediate control amount may be provided, which is prohibited so that the control amount does not change depending on the road surface friction coefficient.

For example, the basic control amount when the road surface friction coefficient is high is set, and during the normal control other than the drift control, the basic control amount is corrected by the correction coefficient α such that the basic control amount becomes small when the road surface friction coefficient is low. During drift control, a value unrelated to the road friction coefficient (correction coefficient β)
Then, the basic control amount is corrected to an intermediate value. In this case, the correction coefficient α continuously changes from 1 to the minimum value αmax (αmax <1) according to the road surface friction coefficient, and the correction coefficient β is a fixed value smaller than 1 and larger than the minimum value αmax. .

Further, in this device, the control amount for the high μ road (control amount for high road surface friction resistance) and the control amount for low μ road (control amount for low road surface friction resistance) are set, and these control amounts are set. The output control amount tad is calculated while being reflected in an interpolation manner in accordance with the road surface μ determination coefficient γ as the road surface friction coefficient calculated by the road surface friction coefficient calculating means. In particular, for the target ΔN tracking control amount tb, For a similar difference ddvr, the high road surface friction resistance map gives a larger control amount than the low road surface friction resistance map, and the degree of reflection of the high μ road control amount is set in proportion to the road surface friction coefficient. Since the control amount for the high μ road and the control amount for the low μ road are set so as to be higher than the proportional distribution value for the road surface μ determination coefficient γ, the target ΔN The control gain tb of the follow-up control becomes relatively high.

Since the target ΔN follow-up control is a control in which the rotational speed difference between the wheel speeds is in-between, even if the road surface friction coefficient is low, the control influence thereof does not easily deviate from the assumed range. Therefore, as described above, in the target ΔN follow-up control, a relatively large torque movement is performed while the control amount for the road surface friction coefficient is set to be relatively large, so that the vehicle behavior can be quickly set to the target. You get the advantage of being able to do it.

Regarding the acceleration turning control amount te, the same turning lateral G is obtained in the region where the lateral G (ggy) is small.
The map for low road friction resistance gives a larger control amount than the map for high road friction resistance for (ggy), and
The control amount for the high μ road and the control amount for the low μ road are set to the road μ determination coefficient γ so that the reflection degree of the low μ road control amount becomes larger than the reflection degree set in proportion to the road surface friction coefficient. Since it is set to the μ side which is lower than the proportional distribution value, the control gain te of the acceleration turn control amount becomes relatively low.

In the case of using the lateral acceleration as a parameter such as acceleration turning control, especially when the road surface friction coefficient is low, as shown in FIG. 9, the relationship between the calculated lateral G and the actual lateral G immediately becomes a non-linear region. Therefore, the control effect is likely to deviate from the expected range. Therefore, as described above, in the acceleration turning control, the low road surface friction resistance map gives a larger control amount than the high road surface friction resistance map in the region where the lateral G is small, and
The reflection degree of the low μ road control amount (control amount corresponding to low road surface friction resistance) is set to be larger than the reflection degree set in proportion to the road surface friction coefficient, whereby the road surface friction coefficient is low. Even in this case, stable control can be performed quickly.

If the road surface friction coefficient is high, high μ
By making the degree of reflection of the road control amount (control amount corresponding to high road surface friction resistance) larger than the reflection degree set in proportion to the road surface friction coefficient, the control in the region where the lateral G is small can be reduced as much as possible. Therefore, the energy loss can be suppressed accordingly. Also, in the tuck-in control, it is possible to set the same as the acceleration turn control,
Also in this case, the same effect as in the case of the acceleration turning control can be obtained.

Note that such fine adjustment of the control amount with respect to the road surface friction coefficient is not limited to the target ΔN follow-up control and the acceleration turning control, but may be applied to those that are less likely to be affected by the road surface friction coefficient and those that are likely to be affected. it can. Further, since the dead zone region is provided in the control amount setting map, the control becomes stable.

Further, the fine adjustment of the control amount with respect to the road surface friction coefficient need not be limited to the above-mentioned embodiment. For example, as a control map for setting the control amount based on parameters such as rotational speed difference and lateral acceleration between the left and right wheels, a high road friction resistance map that gives a control amount corresponding to high road friction resistance and a low road friction resistance map. A map for low road surface friction resistance giving a corresponding control amount is provided, and the output control amount is calculated while reflecting the control amount corresponding to the high road surface friction resistance and the low road surface friction resistance corresponding control amount in an interpolation manner according to the road surface friction coefficient. At the same time, when the two control amounts are reflected in an interpolated manner, the reflection degree of the control amount corresponding to the high road surface friction resistance and the control amount corresponding to the low road surface friction resistance in the medium road surface friction resistance is changed in accordance with the parameter such as the vehicle-specific constant You may adjust in this way.
In this case, a more appropriate control amount according to the vehicle can be given.

In the above embodiment, the target ΔN follow-up control amount tb and the acceleration turn control amount te are added to ensure the turning performance of the acceleration sharp turn. Immediately after the start, control may be performed temporarily by the acceleration turning control amount, and thereafter, control may be performed so as to switch to the target ΔN tracking control amount for steady control. In short, it is important to control so that the turning force of the turning outer wheel is increased immediately after the start of the sharp turning.

Further, although the present embodiment has been described for a four-wheel drive vehicle, the power transmission control device for the left and right wheels for this vehicle is
It can be provided between the left and right drive wheels and between the left and right driven wheels of a two-wheel drive vehicle such as a front-wheel drive vehicle and a rear-wheel drive vehicle.
It may be applied between the front and rear wheels of a wheel drive vehicle, and in this case, it is configured as a vehicle power transmission control device. In the present embodiment, fuzzy inference is used in the vehicle road surface friction coefficient estimation device, but the invention is not limited to this and the parameter (βpp
Correspond to each other) and a parameter indicating the slipperiness of the road surface (αh
(Corresponding to), and the index (high3, mid3, low3) corresponding to each state (high μ road, medium μ road, low μ road corresponds) of other parameters that unitarily express the road surface condition from ) May be calculated, and another method may be used.

Further, in the present embodiment, the vehicle road surface friction coefficient estimation device is used as the vehicle left and right wheel power transmission control device, but needless to say, the application of the vehicle road surface friction coefficient estimation device is not limited to this. Yes.

[0238]

As described in detail above, according to the vehicle road surface friction coefficient estimating apparatus of the present invention, the parameter indicating the unevenness of the road surface and the parameter indicating the slipperiness of the road surface are used. In order to estimate the friction coefficient of the road surface on which the vehicle travels during steady running, the road surface friction coefficient can be estimated with high accuracy, and the calculated index can be calculated continuously and cumulatively to evaluate the road surface friction. Since the coefficient is calculated,
The reliability of the estimation of the road surface friction coefficient can be improved. As a result, there is an advantage that various traveling controls in, for example, an automobile can be more appropriately performed.

[0239] Further, according to the vehicle road surface friction coefficient estimating device of the present invention according to claim 4, calculates a first parameter during steady traveling, and calculates a second parameter when a specific running other than during steady running Te, continue to first parameter accumulated evaluation asking cumulatively, to estimate the friction coefficient of the road surface from this accumulated first parameter and the second parameter, the estimated road friction coefficient appropriate moreover can accurately rows that Ukoto in various driving conditions. As a result, there is an advantage that various traveling controls in, for example, an automobile can be more appropriately performed.

[Brief description of drawings]

FIG. 1 is a schematic overall configuration diagram of a drive system of a vehicle including a power transmission control device for a left and right wheels for a vehicle according to an embodiment of the present invention.

FIG. 2 is a schematic configuration diagram showing a rotational propulsive force distribution adjusting mechanism (torque moving mechanism) of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 3 is a schematic diagram showing a configuration of a hydraulic unit and a control system of a rotational propulsive force distribution adjusting mechanism (torque moving mechanism) of the vehicle left / right wheel power transmission control device according to the embodiment of the present invention.

FIG. 4 is a control block diagram of a power transmission control device for a left and right wheels for a vehicle according to an embodiment of the present invention.

FIG. 5 is a diagram illustrating target control contents of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 6 is a control block diagram related to input calculation processing of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 7 is a diagram illustrating input calculation processing of the vehicle-side left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 8 is a control block diagram relating to drift determination processing of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 9 is a diagram illustrating a drift determination process of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 10 is a diagram illustrating a drift determination process of the vehicle-side left and right wheel power transmission control device according to the embodiment of the present invention.

FIG. 11 is a diagram illustrating a drift determination process of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 12 is a diagram illustrating a drift determination process of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 13 is a diagram illustrating a drift determination process of the vehicle-side left and right wheel power transmission control device according to the embodiment of the present invention.

FIG. 14 is a control block diagram relating to control amount calculation processing (processing for a high μ road) of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 15 is a control block diagram relating to control amount calculation processing (processing for low μ roads) of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 16 is a map (high μ) relating to the acceleration turning control of the vehicle power transmission control device for the left and right wheels according to the embodiment of the present invention.
It is a figure which shows the map for roads.

FIG. 17 is a map (low μ) relating to the acceleration turning control of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.
It is a figure which shows the map for roads.

FIG. 18 is a diagram illustrating drift response control of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 19 is a diagram illustrating acceleration turning control of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 20 is a diagram for explaining road surface μ determination of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 21 is a control block diagram relating to a driving process of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 22 is a diagram showing a map for explaining a drive process of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 23 is a diagram illustrating a map for describing drive processing of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 24 is a diagram for explaining fine adjustment of the output value of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 25 is a schematic diagram showing a display device of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 26 is a diagram for explaining display control of the vehicle-side left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 27 is a control block diagram illustrating a vehicle road surface friction coefficient estimating device as an embodiment of the present invention.

FIG. 28 is a diagram for explaining the vehicle road surface friction coefficient estimating device as one embodiment of the present invention, and is a diagram showing an example of changes in the magnitude of the driving force with respect to the slip ratio of the wheels.

FIG. 29 is a diagram showing a membership function for explaining fuzzy inference by the vehicle road surface friction coefficient estimating apparatus as one embodiment of the present invention, and (A), (B), and (C) are road surface μ judgment values. The suitability for high μ road, medium μ road, and turning to αh is shown in (D), (E), and (F) for high μ road, medium μ road, and turning for vibration component βpp of slip ratio difference. The goodness of fit is shown respectively.

FIG. 30 is a diagram illustrating a surface frictional resistance estimation region (road surface μ determination region) in estimation by the vehicle road surface friction coefficient estimation device as one embodiment of the present invention.

FIG. 31 is a diagram illustrating fuzzy inference performed by a vehicle road surface friction coefficient estimation device according to an embodiment of the present invention.

FIG. 32 is a diagram for explaining fuzzy inference by the vehicle road surface friction coefficient estimating device as one embodiment of the present invention;
(A) relates to a high μ road weight hig, (B) relates to a medium μ road weight mid, and (C) relates to a low μ road weight low.

FIG. 33 is a diagram for explaining road surface frictional resistance estimation (calculation of road surface μ determination coefficient γ) by a vehicle road surface frictional coefficient estimation device as one embodiment of the present invention, and (A) and (B) respectively show the estimation. Explain the process of.

FIG. 34 is a control block diagram illustrating road surface frictional resistance estimation (road surface μ determination) by the vehicle road surface friction coefficient estimation device as one embodiment of the present invention.

FIG. 35 is a control block diagram illustrating output value setting based on road surface frictional resistance estimation (road surface μ determination) by the vehicle road surface frictional coefficient estimation device as one embodiment of the present invention.

FIG. 36 is a main-part control block diagram relating to an output value by road surface μ determination (road surface friction coefficient determination) of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 37 is a diagram for explaining high-pass processing of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

FIG. 38 is a flowchart showing the outline of the operation of the vehicle left-right wheel power transmission control device according to the embodiment of the present invention.

[Explanation of symbols]

2 engine 4 transmission 6 intermediate gear mechanism 8 differential gear mechanism [center differential (center differential)] 8A, 8B differential pinion 8C, 8D side gear 10 front wheel differential gear mechanism [front differential (front differential)] 12L, 12R axle 14, 16 Front Wheel 18 Bevel Gear Mechanism 20 Propeller Shaft 22 Bevel Gear Mechanism 24 Rear Wheel Differential Gear Device [Rear Differential (Rear Differential)] 26L, 26R Axle 28, 30 Rear Wheel 32 Front Wheel Output Shaft 34 Rear Wheel Output Shaft 36 Differential Viscous coupling unit (VCU) 42 as limiting means 42 Electronic control unit (ECU or controller) as controlling means (rotational propulsion force distribution controlling means) 48A Wheel speed sensor 48B Steering wheel angle sensor (handling angle detection) Step) 48C Front-rear acceleration sensor (front-rear G sensor) 48D Lateral acceleration sensor (lateral G sensor) 48E Throttle position sensor (TPS) 50 Rotational propulsion force distribution control mechanism (rotational force adjusting means, torque moving mechanism) 51 Differential carrier 51A Wall part 52 input shaft 54 drive pinion gear 56 crown gear 58 differential case (differential case) 60A, 60B differential pinion 62, 64 side gear 66 left wheel side rotating shaft 68 right wheel side rotating shaft 70 speed change mechanism 70A speed increasing mechanism 70B speed reducing mechanism 72, 74, 76 Intermediate shaft 78A, 80A, 82A Gear (sun gear) 78B, 80B, 82B Gear (planetary pinion) 84 Counter shaft 86 Triple gear 90 Transmission capacity variable control type torque transmission mechanism 90L clutch (left clutch) 90R clutch Right clutch) 90AL, 90AR, 90BL, 90BR Clutch plate 92 Clutch case 38 Hydraulic unit 101 Accumulation unit 102 Control pressure output unit 103 Accumulator 104 Motor pump 105 Pressure switch 106 Electromagnetic proportional pressure control valve (proportional valve) 107 Electromagnetic directional control valve ( Direction switching valve) 108 Battery 109 Motor relay 110 Indicator lamp 111 Control stop means 112 Control amount change means 201, 203 Calculation unit 202 Slip ratio difference calculation unit 204 Road surface μ judgment value calculation unit 205 Reference rotation speed difference calculation unit 206 Turning correction rotation Number difference calculation unit 207 Road surface μ determination unit 210 Steady parameter calculation unit 211 Concavo-convex state parameter calculation unit 212 Slipperiness parameter calculation unit 220 Road surface index calculation unit 230 Road surface friction coefficient calculation unit 240 Steady parameter calculation unit 2 0 specific parameter calculating unit 260 road friction coefficient calculating means 264 selecting unit

Front page continuation (72) Inventor Hiroyuki Suzuki 5-3-8, Shiba, Minato-ku, Tokyo Within Mitsubishi Motors Corporation (72) Inventor Kazuki Ishiguro 5-33-8, Shiba, Minato-ku, Tokyo Mitsubishi Motors Kogyo Co., Ltd. (72) Inventor Satoshi Manabe 5-3-8 Shiba, Minato-ku, Tokyo Mitsubishi Motors Co., Ltd. (56) Reference JP-A-7-101258 (JP, A) JP-A-8- 40241 (JP, A) JP-A-3-258650 (JP, A) JP-A-6-281539 (JP, A) JP-A-6-255510 (JP, A) JP-A-6-286630 (JP, A) JP-A-8-201235 (JP, A) JP-A-5-131855 (JP, A) JP-A-7-108840 (JP, A) JP-A-7-108841 (JP, A) JP-A-7-108842 (JP, A) JP 7-108843 (JP, A) JP 7-156681 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) B60K 17/28-17 / 36 B60K 23/00-23/08 F16H 48/00-48/30 B62D 6/00 B60T 8/00-8/96 G0 1N 19/02

Claims (6)

(57) [Claims] 1. A vehicle state detecting means for detecting a traveling state of a vehicle, and a parameter indicating the unevenness state of a road surface from the traveling state of the vehicle detected by the vehicle state detecting means during steady traveling and the slipperiness. A steady-state parameter calculating means for calculating a parameter, a road surface index calculating means for respectively calculating an index corresponding to each state of other parameters representing the road surface state from the values of the both parameters, and the road surface index calculating means. A road surface friction coefficient estimating device for a vehicle, comprising: road surface friction coefficient calculating means for calculating a road surface friction coefficient by continuously and cumulatively obtaining and evaluating the respective indexes calculated by the above. 2. The road surface coefficient of friction calculating means comprises:
The cumulative value of the target is weighted according to each state, and
A surface friction coefficient is calculated, The claim 1 characterized by the above-mentioned.
Vehicle road surface friction coefficient estimation device. 3. The vehicle according to claim 3 , wherein the left and right wheels are front wheels or rear wheels.
It is provided in relation to the space between the right and left driving power from the power unit.
Rotation that is transmitted to the wheels or is accompanied by rotation of the left and right wheels
Force is transmitted and received between the left and right wheels to rotate and propel the left and right wheels.
Rotational propulsion force distribution adjusting mechanism for changing force and the rotational propulsion
Left drive wheel and right drive on the side equipped with force distribution adjustment mechanism
Left wheel side rotation speed detecting means for detecting wheel rotation speed and right
The wheel side rotation speed detecting means and the rotation propulsion force distribution adjusting mechanism are
The rotation speed of the left drive wheel and the right drive wheel that cannot be provided
Left wheel side rotation speed detection means for detecting and right wheel side rotation speed detection
Based on the output means, the steering angle and the vehicle speed,
Reference rotation speed that calculates the theoretical value of the left-right rotation speed difference between the rear wheels
And a parameter for indicating the road surface condition, the rotational propulsion force distribution
The left wheel side rotation speed detection means and the adjustment mechanism equipped side
And the detection information from the right wheel side rotation speed detecting means,
Left wheel side rotation speed on the side that cannot be equipped with a power distribution adjustment mechanism
Detection information from the detection means and the right wheel side rotation speed detection means
And the vehicle rotation speed difference calculated based on
Left wheel side rotation speed detection on the side equipped with the power distribution adjustment mechanism
Detection information from the output means and the right wheel side rotation speed detection means,
The left wheel side on which the rotation propulsion force distribution adjusting mechanism cannot be provided
Detection from the rotation speed detection means and the right wheel side rotation speed detection means
And Dejo paper, calculation information from the reference rotational speed difference calculating means
And the operating state of the rotation propulsion force distribution adjusting mechanism,
By the torque movement of the drive wheels of the vehicle calculated based on
The road surface μ judgment value is used as a judgment value.
The vehicle road surface friction coefficient estimation device described. 4. Vehicle state detection for detecting a traveling state of a vehicle
Means and during the steady running of the vehicle detected by the vehicle state detecting means.
The first parameter from the running condition to the road condition
Steady parameter calculating means for calculating and other than steady running of the vehicle detected by the vehicle state detecting means
Showing the road surface condition from the traveling condition during the specific traveling of the vehicle
Specific parameter calculation means for calculating the parameters of the first parameter and the first parameter calculated by the stationary parameter calculation means.
The parameter is continuously calculated and evaluated cumulatively.
The accumulated first parameter and the second parameter
Calculate the road surface friction coefficient by calculating and
And a road surface friction coefficient for a vehicle, comprising:
Estimator. 5. The specific parameter calculation means is for a vehicle.
At the time of starting, even if it is a straight start and one wheel slips
The second parameter to ensure a low road surface friction coefficient.
Set the meter, and the wheel does not slip at a certain acceleration or more
The second parameter to obtain a high road surface friction coefficient.
Set the meter, and if the vibration component of the wheel is large,
The second parameter to obtain a low or low road friction coefficient
Is set, and the road is calculated from the torque movement direction and the speed difference between the left and right rear wheels.
Low if the surface is determined to be in the μ-split state.
Set the second parameter so that the road surface friction coefficient becomes
The road surface friction member for a vehicle according to claim 4, wherein
Number estimator. 6. The specific parameter calculation means is for starting
In specific driving conditions such as driving and non-linear driving,
Lateral acceleration caused by the above-mentioned vehicle with a large degree of wheel slippage
Is less than the first predetermined value, the road surface has a low road friction coefficient.
Therefore, the second parameter is set so that the road friction coefficient is low.
The lateral acceleration generated in the vehicle is set to a second predetermined value.
If it is larger than the above, it is judged that the road surface has a high road friction coefficient.
The second parameter to obtain a high road surface friction coefficient.
The vehicle according to claim 4 or 5, characterized by being set.
Road surface friction coefficient estimation device.