Drive force control device and method of controlling vehicle

A drive force control device, which controls a drive force distribution device that distributes a drive force to right and left rear wheels at variable distribution ratios, computes a steering angle-based turning radius determined in accordance with a steering angle, computes a limit turning radius, which is a minimum value of the turning radius with which the vehicle is turnable while keeping a stable travel state, in accordance with a vehicle speed, sets the larger one of the steering angle-based turning radius and the limit turning radius as a target turning radius, computes target rotational speeds for the right and left rear wheels on the basis of the target turning radius and the vehicle speed, and adjusts the ratios of distribution of the drive force to the right and left rear wheels such that actual rotational speeds of the right and left rear wheels approximate the target rotational speeds.

INCORPORATION BY REFERENCE

The disclosure of Japanese Patent Application No. 2016-220614 filed on Nov. 11, 2016 including the specification, drawings and abstract, is incorporated herein by reference in its entirety.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a drive force control device that can distribute a drive force of a drive source to right and left wheels of a vehicle at variable distribution ratios, and to a method of controlling a vehicle.

2. Description of the Related Art

There has hitherto been known a control device described in Japanese Patent Application Publication No. 2014-40852 (JP 2014-40852 A) as a control device that controls a drive force transfer device that transfers a drive force to right and left wheels of a vehicle.

The drive force transfer device described in JP 2014-40852 A includes a turning radius estimation unit, a target slip angle computation unit, a target rotational speed computation unit, and a drive force control unit. The turning radius estimation unit estimates a turning radius of the vehicle on the basis of the steering angle of a steering wheel etc. The target slip angle computation unit computes a target slip angle in accordance with the estimated turning radius. The target rotational speed computation unit computes respective target rotational speeds for the right and left wheels on the basis of the target slip angle and the vehicle speed. The drive force control unit controls drive forces to be transferred to the right and left wheels such that the actual rotational speeds of the right and left wheels approximate the target rotational speeds.

With the drive force control device described in JP 2014-40852 A, the behavior of the vehicle during a turn can be stabilized. However, if the steering angle becomes larger than an angle at which the vehicle is able to turn while keeping a stable travel state, such as in the case where the steering wheel is operated greatly during a turn at a high speed, for example, the effect in stabilizing the behavior of the vehicle may not always be fully demonstrated.

SUMMARY OF THE INVENTION

It is an object of the present invention to provide a drive force control device that is capable of stabilizing the behavior of a vehicle compared to the case where drive forces to be transferred to right and left wheels are controlled such that the actual rotational speeds approximate target rotational speeds that match a steering angle, and a method of controlling a vehicle.

An aspect of the present invention provides a drive force control device that controls a drive force distribution device that distributes a drive force of a drive source to right and left wheels of a vehicle at variable distribution ratios.

The drive force control device includes:

a first turning radius computation unit that computes a steering angle-based turning radius that is a turning radius of the vehicle determined in accordance with a steering angle;

a second turning radius computation unit that computes a limit turning radius, which is a minimum value of the turning radius with which the vehicle is turnable while keeping a stable travel state, in accordance with a vehicle speed;

a target turning radius setting unit that sets the steering angle-based turning radius as a target turning radius in the case where the steering angle-based turning radius is larger than the limit turning radius, and that sets the limit turning radius as the target turning radius in the case where the steering angle-based turning radius is smaller than the limit turning radius;

a target rotational speed computation unit that computes respective target rotational speeds for the right and left wheels on the basis of the target turning radius, which is set by the target turning radius setting unit, and the vehicle speed; and

a drive force distribution ratio adjustment unit that adjusts the ratios of distribution of the drive force to the right and left wheels such that actual rotational speeds of the right and left wheels approximate the target rotational speeds.

Another aspect of the present invention provides a method of controlling a vehicle that includes a drive force distribution device that distributes a drive force of a drive source to right and left wheels at variable distribution ratios.

The control method includes:

computing a steering angle-based turning radius that is a turning radius of the vehicle determined in accordance with a steering angle;

computing a limit turning radius, which is a minimum value of the turning radius with which the vehicle is turnable while keeping a stable travel state, in accordance with a vehicle speed;

setting the larger one of the steering angle-based turning radius and the limit turning radius as a target turning radius;

computing respective target rotational speeds for the right and left wheels on the basis of the target turning radius and the vehicle speed; and

adjusting the ratios of distribution of the drive force to the right and left wheels such that actual rotational speeds of the right and left wheels approximate the target rotational speeds.

With the drive force control device and the method of controlling a vehicle according to the aspects described above, it is possible to stabilize the behavior of the vehicle compared to the case where drive forces to be transferred to right and left wheels are controlled such that the actual rotational speeds approximate target rotational speeds that match a steering angle.

DETAILED DESCRIPTION OF EMBODIMENTS

FIG. 1is a schematic diagram illustrating a schematic example of the configuration of a four-wheel-drive vehicle according to a first embodiment of the present invention. As illustrated inFIG. 1, a four-wheel-drive vehicle100includes a vehicle body101, an engine102, a transmission103, right and left front wheels104band104a(right and left wheels on the front side), and right and left rear wheels105band105a(right and left wheels on the rear side). The engine102serves as a drive source that generates torque for travel. The right and left front wheels104band104aserve as a pair of right and left main drive wheels to which a drive force of the engine102is always transferred. The right and left rear wheels105band105aserve as a pair of right and left auxiliary drive wheels to which a drive force of the engine102is transferred intermittently in accordance with the travel state.

The four-wheel-drive vehicle100also includes, as a drive force transfer system, a front differential106, a propeller shaft107, and a drive force distribution device1. The drive force of the engine102, the speed of which has been varied by the transmission103, is always transferred to the right and left front wheels104band104avia the front differential106and a pair of drive shafts106band106a. The right and left front wheels104band104aare steered by an operation of a steering wheel109by a driver.

The drive force of the engine102, the speed of which has been varied by the transmission103, is transferred to the right and left rear wheels105band105avia the propeller shaft107, the drive force distribution device1, and a pair of drive shafts108band108a. The drive force distribution device1can distribute the drive force of the engine102to the left rear wheel105aand the right rear wheel105bat variable distribution ratios. The configuration of the drive force distribution device1will be discussed in detail later.

The four-wheel-drive vehicle100includes a drive force control device10that controls the drive force distribution device1. The drive force control device10can adjust the drive forces to be transferred to the left rear wheel105aand the right rear wheel105bindependently of each other. The drive force control device10has a first turning radius computation unit11, a second turning radius computation unit12, a target turning radius setting unit13, a target rotational speed computation unit14, a drive force distribution ratio adjustment unit15, and a road surface friction coefficient estimation unit16. The drive force control device10will be discussed in detail later.

Rotational speed sensors111to114are connected to the drive force control device10. The rotational speed sensors111to114are configured to detect the rotational speeds of the right and left front wheels104band104aand the right and left rear wheels105band105a. The rotational speed sensors111to114are each composed of a Hall IC disposed to face a magnetic ring that has a plurality of magnetic poles that rotate together with the right and left front wheels104band104aand the right and left rear wheels105band105a, for example, and output a pulse signal in cycles that match the rotational speed. This enables the drive force control device10to detect the rotational speeds of the right and left front wheels104band104aand the right and left rear wheels105band105a.

A steering angle sensor115is also connected to the drive force control device10. The steering angle sensor115detects the rotational angle (steering angle) of the steering wheel109from the neutral position. This enables the drive force control device10to detect the steering angle of the steering wheel109.

An accelerator operation amount sensor116is further connected to the drive force control device10. The accelerator operation amount sensor116detects the amount by which an accelerator pedal110is depressed. This enables the drive force control device10to detect the accelerator operation amount which matches the amount by which the accelerator pedal110is depressed.

FIG. 2is a sectional view illustrating an example of the configuration of the drive force distribution device1.

The drive force distribution device1includes a case member20, an input rotary member3, and a pair of torque couplings4. The case member20has first to third housing spaces20ato20cinside. The input rotary member3is housed in the first housing space20aof the case member20. The pair of torque couplings4are housed in the second and third housing spaces20band20cwhich interpose the first housing space20a.

The torque coupling4which is housed in the second housing space20band the torque coupling4which is housed in the third housing space20chave a common configuration. In the case where it is necessary to distinguish the torque couplings4from each other in the following description, the torque coupling4which is housed in the second housing space20bwill be referred to as a first torque coupling4A, and the torque coupling4which is housed in the third housing space20cwill be referred to as a second torque coupling4B.

The case member20is provided with a pair of partition walls21that separate the first housing space20aand the second housing space20bfrom each other and that separate the first housing space20aand the third housing space20cfrom each other. The pair of partition walls21are each formed with a through hole21athat communicates the first housing space20awith the second and third housing spaces20band20c.

The input rotary member3has a first member31and a second member32. The first member31is rotatably supported by the case member20. The second member32is an annular ring gear. The first member31and the second member32are coupled to each other by a plurality of bolts33. The first member31has a cylindrical portion311and a flange portion312that are integral with each other. A through hole31ais formed in the center portion of the cylindrical portion311which is in a cylindrical shape. The flange portion312is formed to project outward from the outer peripheral surface of the cylindrical portion311. The second member32is fixed to the distal end portion of the flange portion312, and meshed with a gear portion107aformed at one end of the propeller shaft107which is inserted through a first opening200aof the case member20. The first member31is rotatably supported by a pair of bearings22disposed between the inner surfaces of the through holes21aand the first member31.

The torque couplings4each have a multi-plate clutch41, an electromagnetic clutch42, a cam mechanism43, an inner shaft44, and a housing40that houses such components.

The housing40is composed of a first housing member401and a second housing member402coupled so as not to be rotatable relative to each other. The first housing member401has a bottomed cylindrical shape. The second housing member402is disposed so as to block an end portion of the first housing member401on the opening side.

The multi-plate clutch41is disposed between the first housing member401of the housing40and the inner shaft44which is in a cylindrical shape. The multi-plate clutch41is composed of inner clutch plates411and outer clutch plates412. The inner clutch plates411are spline-engaged with the outer peripheral surface of the inner shaft44so as not to be relatively rotatable. The outer clutch plates412are spline-engaged with the inner peripheral surface of the first housing member401so as not to be relatively rotatable.

The electromagnetic clutch42has an annular coil421and an armature cam422, and is disposed on the rotational axis of the housing40. The electromagnetic clutch42is configured to move the armature cam422toward the coil421through generation of an electromagnetic force by the coil421, and to frictionally slide the armature cam422against the second housing member402.

The cam mechanism43has a main cam431and cam followers432. The main cam431is parallel to the armature cam422along the rotational axis of the housing40. The cam followers432are in a spherical shape, and are interposed between the main cam431and the armature cam422. The cam followers432can roll in respective cam grooves formed in the armature cam422and the main cam431so as to extend in the circumferential direction. The cam grooves are gradually varied in depth in the axial direction in accordance with the position in the circumferential direction. The cam mechanism43is configured such that the armature cam422receives a rotational force from the housing40through energization of the coil421and converts the rotational force into a pressing force that serves as a clutch force of the multi-plate clutch41.

When the amount of energization of the coil421becomes larger, the friction force between the armature cam422and the second housing member402is increased, and the main cam431presses the multi-plate clutch41more strongly. That is, the torque coupling4can variably control the pressing force of the multi-plate clutch41in accordance with the amount of energization of the coil421, and hence can adjust the amount of torque transferred between the housing40and the inner shaft44.

One end of the drive shaft108afor the left rear wheel, which is inserted through a second opening200bof the case member20, is coupled to the inner shaft44in the first torque coupling4A through spline fitting such that the drive shaft108ais not rotatable relative to the inner shaft44. Meanwhile, one end of the drive shaft108bfor the right rear wheel, which is inserted through a third opening200cof the case member20, is coupled to the inner shaft44in the second torque coupling4B through spline fitting such that the drive shaft108bis not rotatable relative to the inner shaft44. The multi-plate clutch41in the first torque coupling4A is an embodiment of the “left clutch” of the present invention which transfers a drive force to the left rear wheel105a. The multi-plate clutch41in the second torque coupling4B is an embodiment of the “right clutch” of the present invention which transfers a drive force to the right rear wheel105b.

The housings40of the first torque coupling4A and the second torque coupling4B and the cylindrical portion311of the first member31of the input rotary member3are coupled by a pair of coupling members23such that the housings40are not rotatable relative to the cylindrical portion311. The coupling members23each have a columnar boss portion231and a disk-shaped flange portion232that are integral with each other. The boss portions231are spline-fitted with the inner surface of the through hole31aof the first member31so as not to be relatively rotatable. The flange portions232are spline-fitted with the housings40so as not to be relatively rotatable. The boss portions231are inserted through the through holes21aof the partition walls21.

An excitation current is supplied from the drive force control device10to the coils421of the torque couplings4. The drive force control device10can control a drive force transferred from the input rotary member3to the left rear wheel105aby increasing and decreasing a current supplied to the coil421of the first torque coupling4A. Meanwhile, the drive force control device10can control a drive force transferred from the input rotary member3to the right rear wheel105bby increasing and decreasing a current supplied to the coil421of the second torque coupling4B.

The drive force control device10includes a CPU and a storage element, for example. When the CPU executes a process on the basis of a program stored in the storage element, the drive force control device10functions as the first turning radius computation unit11, the second turning radius computation unit12, the target turning radius setting unit13, the target rotational speed computation unit14, the drive force distribution ratio adjustment unit15, and the road surface friction coefficient estimation unit16.

FIG. 3is a diagram illustrating an example of the control configuration of the drive force control device10in the form of a block diagram. The first turning radius computation unit11computes a steering angle-based turning radius which is a turning radius of the vehicle determined in accordance with the steering angle which is detected by the steering angle sensor115. In the embodiment, the steering angle-based turning radius is computed on the basis of the vehicle speed which is obtained from the steering angle and signals output from the rotational speed sensors111to114. The steering angle-based turning radius is a turning radius for a case where the four-wheel-drive vehicle100travels at a steering angle detected by the steering angle sensor115when the wheels (right and left front wheels104band104aand the right and left rear wheels105band105a) are not idling (slipping).

The second turning radius computation unit12computes a limit turning radius, which is the minimum value of a turning radius with which the vehicle is able to turn while keeping a stable travel state, in accordance with the vehicle speed. The phrase “turn while keeping a stable travel state” refers to turning while maintaining a state in which the wheels grip on the road surface without causing oversteer that requires countersteer, for example. In the embodiment, the limit turning radius is computed in consideration of the friction coefficient of the road surface which is estimated by the road surface friction coefficient estimation unit16to be discussed later. The limit turning radius becomes larger as the estimated value of the road surface friction coefficient which is computed by the road surface friction coefficient estimation unit16is smaller.

The target turning radius setting unit13sets the steering angle-based turning radius, which is computed by the first turning radius computation unit11, as a target turning radius in the case where the steering angle-based turning radius is larger than the limit turning radius, which is computed by the second turning radius computation unit12, and sets the limit turning radius as the target turning radius in the case where the steering angle-based turning radius is smaller than the limit turning radius. In other words, the target turning radius setting unit13sets the larger one of the steering angle-based turning radius and the limit turning radius as the target turning radius.

The target rotational speed computation unit14computes respective target rotational speeds for the right and left rear wheels105band105aon the basis of the target turning radius, which is set by the target turning radius setting unit13, and the vehicle speed. The target rotational speeds are the rotational speeds of the left rear wheel105aand the right rear wheel105bfor a case where the four-wheel-drive vehicle100travels while keeping a stable travel state through a turning road with the target turning radius at the vehicle speed which is obtained from signals output from the rotational speed sensors111to114. In the embodiment, a target slip angle is obtained from the accelerator operation amount and the vehicle speed, and the target rotational speeds are computed on the basis of the target slip angle.

FIG. 4is a diagram illustrating a target slip angle α at the time when the four-wheel-drive vehicle100is traveling through a curve with a turning radius R. The target slip angle α is an angle formed by the direction which is perpendicular to a line segment L that connects between a turning center point C and a center-of-gravity point G of the four-wheel-drive vehicle100(the direction of movement of the center of gravity indicated by the arrow A inFIG. 4) and a vehicle center line CL that extends in the front-rear direction of the four-wheel-drive vehicle100through the center-of-gravity point G. The target rotational speed computation unit14obtains the target slip angle α with reference to a map stored in the storage element, for example, and calculates the target rotational speeds through computation using the obtained target slip angle α. In this case, the map which is referenced by the target rotational speed computation unit14defines the relationship among the accelerator operation amount, the vehicle speed, and the target slip angle α.

The drive force distribution ratio adjustment unit15adjusts the ratios of distribution of the drive force to the right and left rear wheels105band105asuch that the actual rotational speeds of the right and left rear wheels105band105awhich are obtained from signals output from the rotational speed sensors113and114approximate the target rotational speeds. More specifically, with reference to a drive force that matches the amount by which the accelerator pedal110is depressed which is detected by the accelerator operation amount sensor116, the drive force to be transferred to the left rear wheel105avia the multi-plate clutch41of the first torque coupling4A is increased if the actual rotational speed of the left rear wheel105awhich is obtained from a signal output from the rotational speed sensor113is lower than the target rotational speed which is set by the target rotational speed computation unit14, and the drive force to be transferred to the left rear wheel105ais reduced if the actual rotational speed of the left rear wheel105ais higher than the target rotational speed. The drive force to be transferred to the right rear wheel105bvia the multi-plate clutch41of the second torque coupling4B is also adjusted in the same manner.

The road surface friction coefficient estimation unit16estimates the friction coefficient of the road surface on the basis of the outside temperature, the tire reaction force during steering, the frequency of operation of a wiper, the state of the road surface which is grasped from an image that captures the road surface, etc., for example.

The drive force control device10executes the processes of the first turning radius computation unit11, the second turning radius computation unit12, the target turning radius setting unit13, the target rotational speed computation unit14, the drive force distribution ratio adjustment unit15, and the road surface friction coefficient estimation unit16repeatedly in control cycles of 5 ms, for example. That is, the drive force control device10computes a steering angle-based turning radius and a limit turning radius, sets the larger one of the steering angle-based turning radius and the limit turning radius as a target turning radius, computes respective target rotational speeds for the right and left rear wheels105band105aon the basis of the target turning radius and the vehicle speed, and adjusts the ratios of distribution of the drive force to the right and left rear wheels105band105asuch that the actual rotational speeds of the right and left rear wheels105band105aapproximate the target rotational speeds, in predetermined control cycles.

With the first embodiment described above, in the case where the steering wheel109is operated to a greater degree than a steering angle corresponding to the limit turning radius, respective target rotational speeds for the right and left rear wheels105band105aare computed on the basis of the target turning radius, which is set to the limit turning radius, and the vehicle speed, and the drive forces for the right and left rear wheels105band105aare adjusted such that the actual rotational speeds of the right and left rear wheels105band105aapproximate the computed target rotational speeds. Consequently, it is possible to suppress an excessive oversteer state, and to stabilize the behavior of the four-wheel-drive vehicle100.

In the embodiment, the limit turning radius is computed in consideration of the friction coefficient of the road surface which is estimated by the road surface friction coefficient estimation unit16. Thus, the limit turning radius can be computed with higher precision, and the limit turning radius can be prevented from becoming larger than necessary, for example.

In the embodiment, the target rotational speeds for the right and left rear wheels105band105aare computed on the basis of the target slip angle. Thus, the target rotational speeds can be computed accurately.

In the embodiment, further, the drive forces are transferred to the left rear wheel105aand the right rear wheel105bvia the respective multi-plate clutches41of the first and second torque couplings4A and4B. Thus, the ratios of distribution of the drive force to the right and left rear wheels105band105acan be adjusted easily and reliably.

A second embodiment of the present invention will be described next with reference toFIGS. 5 to 7. The configuration of a drive force distribution device1A according to the second embodiment, which distributes a drive force to the right and left rear wheels105band105b, is different from that of the drive force distribution device1according to the first embodiment. Differences of the second embodiment from the first embodiment will be mainly described below.

FIG. 5is a sectional view illustrating the configuration of a drive force distribution device1A according to a second embodiment.FIG. 6is a diagram illustrating a schematic configuration of the drive force distribution device1A.FIG. 7is a perspective view illustrating a planetary carrier72that holds a plurality of planetary gears71.

The drive force distribution device1A is configured to have a case member5, a motor50, a differential gear mechanism6, a planetary gear mechanism7, and a speed change mechanism8. The motor50is integrated in the case member5. The differential gear mechanism6distributes a drive force input from the propeller shaft107to the pair of drive shafts108aand108bwhile allowing differential motion. The motor50is controlled by a drive force control device10A.

The case member5is formed by coupling first to third members51to53to each other. The differential gear mechanism6is housed in the first member51. The planetary gear mechanism7and the speed change mechanism8are housed in the third member53. The second member52is disposed between the first member51and the third member53.

The differential gear mechanism6has a ring gear61and a differential case62. The ring gear61is meshed with the gear portion107aof the propeller shaft107. The differential case62rotates together with the ring gear61. The differential case62is rotatably supported by bearings54and55coaxially with the pair of drive shafts108aand108b, and constituted by coupling a body portion621and a lid portion622to each other. An internal gear621ais formed on the inner peripheral surface of the body portion621which is in a bottomed cylindrical shape. The lid portion622is disposed on the opening side of the body portion621.

A sun gear63disposed in the differential case62is coupled so as to rotate together with the drive shaft108a. A plurality of planetary gear pairs64are disposed between the sun gear63and the inner peripheral surface of the body portion621of the differential case62. The planetary gear pairs64are each composed of a first planetary gear641and a second planetary gear642. The first planetary gear641is meshed with the internal gear621a. The second planetary gear642is meshed with the sun gear63. The first planetary gear641and the second planetary gear642are rotatably and revolvably held by a planetary carrier65while being meshed with each other. The planetary carrier65is coupled so as not to be rotatable relative to the drive shaft108bin the differential case62.

A drive force input from the propeller shaft107is transferred from the ring gear61to the differential case62. The input drive force is transferred to the drive shafts108aand108bwhen the sun gear63and the planetary carrier65, which are coupled to the differential case62via the plurality of planetary gear pairs64, are rotated. In the case where there occurs a difference in rotation between the right and left rear wheels105band105a, such as during a turn of the vehicle, the first planetary gear641and the second planetary gear642revolve around the sun gear63while rotating. Consequently, the differential gear mechanism6distributes the drive force which is input from the propeller shaft107to the pair of drive shafts108aand108bwhile allowing differential motion.

As illustrated inFIG. 6, the planetary gear mechanism7and the speed change mechanism8are provided between the drive shaft108aand the drive shaft108b. The planetary gear mechanism7can cause a difference in rotation between the drive shafts108aand108b. The speed change mechanism8is disposed adjacent to the planetary gear mechanism7. The planetary gear mechanism7and the speed change mechanism8are an embodiment of the gear mechanism according to the present invention which varies the rotational speed difference between the left rear wheel105aand the right rear wheel105b.

The planetary gear mechanism7is drivably coupled to the motor50which is formed in a hollow shape, and generates a difference in rotation between the drive shafts108aand108bon the basis of motor torque output from the motor50. The planetary gear mechanism7includes the plurality of (four) planetary gears71and the planetary carrier72. The planetary gears71are each constituted by coupling a first pinion711and a second pinion712, which have different pitch circle diameters, so as not to be relatively rotatable. The planetary carrier72revolvably and rotatably supports the planetary gears71. The planetary gears71are formed such that the pitch circle diameter of the second pinions712is slightly larger than the pitch circle diameter of the first pinions711.

As illustrated inFIG. 7, the planetary carrier72has a cylindrical outer peripheral wall721and a pair of lid portions722that face each other and that partially block both end portions of the outer peripheral wall721. A plurality of opening portions720are formed in the outer peripheral surface of the outer peripheral wall721. The number of the opening portions720corresponds to the number of the planetary gears71. The planetary gears71are rotatably housed in the planetary carrier72with respective tooth portions711aand712aof the first pinions711and the second pinions712projecting outward from the opening portions720. The gear support structure for the planetary gears71is similar to that for planetary gears81of the speed change mechanism8to be discussed later. Therefore, symbols corresponding to the speed change mechanism8are given in parentheses to omit detailed description of the configuration of the speed change mechanism8.

Support holes722aare formed in the lid portions722of the planetary carrier72so as to face each other at positions corresponding to the opening portions720. The planetary gears71are supported so as to be rotatable with respect to the planetary carrier72with shaft portions of the planetary gears71, which extend along the axial direction, inserted into the support holes722a.

Insertion holes722bare formed in the center portions of the lid portions722of the planetary carrier72. The insertion holes722benable insertion of the drive shaft108aalong the axial direction. The planetary carrier72is rotatably supported by the drive shaft108awith the drive shaft108ainserted through the insertion holes722b.

A first ring gear91and a second ring gear92are meshed with the first pinions711and the second pinions712, respectively, which project outward via the opening portions720of the planetary carrier72. The second ring gear92which is meshed with the second pinions712is coupled so as not to be rotatable relative to the planetary carrier65which constitutes the differential gear mechanism6.

The second ring gear92has a tubular portion921through which the drive shaft108ais inserted. The tubular portion921is rotatably supported by a ball bearing56and a needle bearing57. The second ring gear92is coupled to the drive shaft108bvia the planetary carrier65of the differential gear mechanism6with the distal end of the tubular portion921spline-fitted with the planetary carrier65.

The outer peripheral wall721of the planetary carrier72is provided with a flange portion723that extends radially outward. External teeth723aare formed on the outer periphery of the flange portion723. The flange portion723is coupled to the motor50.

The motor50is composed of a brushless motor, for example. The motor50is disposed coaxially on the radially outer side of the planetary gear mechanism7, and rotated by a motor current supplied from the drive force control device10A. The drive force control device10A uses a battery (not illustrated) as a power supply source, and adjusts a motor current through switching based on PWM control. The planetary carrier72is coupled to the motor50with the flange portion723, which is provided on the outer peripheral wall721, spline-fitted with the inner periphery of a rotor of the motor50.

The drive force distribution device1A includes the speed change mechanism8which is configured to correct the speed change ratio which is set for the planetary gear mechanism7. The first ring gear91which is meshed with the first pinions711of the planetary gear71is coupled to the drive shaft108avia the speed change mechanism8. That is, the planetary gear mechanism7has a predetermined gear ratio based on the difference in pitch circle diameter between the first pinions711and the second pinions712of the planetary gear71. Thus, in the case where the speed change mechanism8were not provided, the planetary carrier72would be rotated even if no differential motion were caused between the drive shafts108aand108bduring travel, and a load would be imposed on the motor50etc.

Therefore, in the embodiment, the speed change mechanism8which cancels out the speed change ratio which is set for the planetary gear mechanism7is interposed between the first ring gear91of the planetary gear mechanism7and the drive shaft108a. Consequently, the motor50is configured not to be rotated, even during travel, in the case where no differential motion is caused between the drive shafts108aand108b. That is, the motor50is rotated at a speed corresponding to the rotational speed difference between the drive shafts108aand108b. Motor torque output from the motor50acts to accelerate one of the right and left rear wheels105band105a, and to decelerate the other. Consequently, the ratios of distribution of the drive force to the right and left rear wheels105band105aare adjusted.

More specifically, the speed change mechanism8has a plurality of (four) planetary gears81. The planetary gears81are each constituted by coupling a third pinion811and a fourth pinion812so as not to be relatively rotatable. The third pinion811has the same pitch circle diameter as that of the first pinion711which constitutes the planetary gear71. The fourth pinion812has the same pitch circle diameter as that of the second pinion712. The planetary gears81are revolvably and rotatably supported by a planetary carrier82. The planetary carrier82is rotatably supported by the drive shaft108awhich is inserted through an insertion hole822bformed in the axial center.

A third ring gear93and a fourth ring gear94are meshed with the third pinions811and the fourth pinions812which project outward from opening portions820of the planetary carrier82. The third ring gear93has the same configuration as that of the first ring gear91. The fourth ring gear94has the same configuration as that of the second ring gear92. The third ring gear93, which is meshed with the third pinions811, is coupled so as not to be rotatable relative to the first ring gear91on the planetary gear mechanism7side. The fourth ring gear94is coupled so as not to be rotatable relative to the drive shaft108a. The planetary carrier82which supports the planetary gears81is coupled so as not to be rotatable relative to the third member53of the case member5with a flange portion823(external teeth823a), which is provided on an outer peripheral wall821, serving as a coupling portion.

The first ring gear91and the third ring gear93are integrally formed with each other with inner teeth of the same shape disposed in parallel at both ends of the inner periphery of a sleeve9which is in a tubular shape. The fourth ring gear94has a tubular portion941through which the drive shaft108ais inserted. The tubular portion941is rotatably supported by a ball bearing58. The fourth ring gear94is coupled so as not to be rotatable relative to the drive shaft108awith the tubular portion941spline-fitted with the drive shaft108a.

In the drive force distribution device1A configured as described above, the planetary carrier72of the planetary gear mechanism7which is coupled to the motor50is not rotated in the case where no differential motion is caused between the drive shafts108aand108b. Meanwhile, difference in rotation can be caused between the drive shafts108aand108bby rotationally driving the planetary carrier72via the flange portion723using motor torque. The drive force of the engine102can be distributed to the drive shafts108aand108bat variable distribution ratios by controlling the motor torque to be input to the planetary gear mechanism7as control torque.

As with the drive force control device10according to the first embodiment, the drive force control device10A which controls the drive force distribution device1A has the first turning radius computation unit11, the second turning radius computation unit12, the target turning radius setting unit13, the target rotational speed computation unit14, the drive force distribution ratio adjustment unit15, and the road surface friction coefficient estimation unit16. However, the content of control by the drive force distribution ratio adjustment unit15of the drive force control device10A is different from the content of control by the drive force distribution ratio adjustment unit15of the drive force control device10.

Specifically, the drive force distribution ratio adjustment unit15of the drive force control device10according to the first embodiment adjusts the drive force which is transferred to the right and left rear wheels105band105avia the multi-plate clutches41of the first and second torque couplings4A and4B. However, the drive force distribution ratio adjustment unit15of the drive force control device10A according to the present embodiment adjusts the magnitude and the direction of a motor current to be supplied to the motor50of the drive force distribution device1A to adjust the rotational speed difference between the right and left rear wheels105band105aby the planetary gear mechanism7and the speed change mechanism8such that the actual rotational speeds of the right and left rear wheels105band105aapproximate the target rotational speeds which are computed by the target rotational speed computation unit14.

Also with the second embodiment described above, as with the first embodiment, it is possible to suppress an excessive oversteer state, and to stabilize the behavior of the vehicle, even in the case where the steering wheel109is operated to a greater degree than a steering angle corresponding to the limit turning radius.

The present invention can be modified as appropriate without departing from the scope and spirit of the present invention. For example, the drive force control device10,10A has the road surface friction coefficient estimation unit16which estimates the friction coefficient of the road surface in the first and second embodiments. However, the present invention is not limited thereto, and the drive force control device10,10A may not have the road surface friction coefficient estimation unit16. In this case, the second turning radius computation unit12of the drive force control device10,10A computes a limit turning radius using a predetermined constant set in advance as the friction coefficient of the road surface. It is desirable that the predetermined constant should be a value (e.g. 0.4 to 0.6) corresponding to the friction coefficient of a wet road, for example, which is smaller than the friction coefficient of a dry paved road (dry road), in consideration of the safety.