JP2014015141A - Driving force distribution system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a driving force transmission system which enables a shift to two-wheel drive even though a failure occurs in an actuator for controlling a pressing force in a radial direction during a run in four-wheel drive.SOLUTION: An electromagnetic brake 59 is provided which applies a braking-torque according to an energization amount for a coil 59a to a pinion shaft 56 meshed with crank shafts 51L and 51R.

Description

  The present invention relates to a driving force distribution device.

  Patent Document 1 discloses a first roller that rotates together with a main drive wheel transmission system, a second roller that rotates together with a driven wheel transmission system, and a second roller that is rotatable about an eccentric axis that is offset from the rotation axis. A driving force distribution device is disclosed that includes a crankshaft to be supported and an electric motor (radial pressing force control actuator) that rotationally drives the crankshaft.

  In the above prior art, the crankshaft is rotationally driven by the operation of the electric motor and the second roller is turned around the eccentric axis, thereby causing the second roller to be displaced relative to the first roller in the radial direction, whereby both rollers It controls the radial pressing force between them, that is, the distribution of the driving force between the main driving wheel and the sub driving wheel.

  In this prior art, in addition to the above, the rotation of the electric motor is configured to be transmitted to the crankshaft via the torque diode, and when the radial pressing force between the two rollers is maintained, the electric motor is deactivated. Thus, the crankshaft rotation angle is maintained by utilizing the irreversible transmission operation of the torque diode.

JP 2012-11794 A

  However, in the above prior art, when the electric motor breaks down during four-wheel drive traveling, the crankshaft rotation angle is fixed due to the irreversible transmission operation of the torque diode, and it becomes impossible to shift to two-wheel drive traveling. In this case, since the driving force is always distributed to the driven wheels, overheating of the oil temperature of the differential gear of the driven wheels becomes a problem.

  The present invention has been made paying attention to the above-mentioned problems, and the object of the present invention is to move to two-wheel drive traveling even when a failure occurs in the radial direction pressing force control actuator during four-wheel drive traveling. It is in providing the driving force transmission device which can implement | achieve transfer of this.

  In the present invention, an electromagnetic brake that applies a braking force corresponding to the amount of current supplied to the coil is provided to the second roller eccentric support member that supports the second roller so as to be rotatable about the eccentric axis.

Therefore, in order to maintain the radial pressing force between the two rollers, the braking force is generated by energizing the coil of the electromagnetic brake, and the eccentricity of the second roller with the radial pressing force control actuator deactivated. The rotation around the axis can be stopped.
In addition, since the braking force becomes zero by stopping energization to the coil, the second roller eccentric support member has zero driving force distribution to the driven wheels due to the radial pressing reaction force acting between the first roller and the second roller. It returns to the position that becomes.
Therefore, even when a failure occurs in the radial pressing force control actuator during four-wheel drive traveling, the shift to two-wheel drive traveling can be realized.

1 is a schematic plan view showing a power train of a four-wheel drive vehicle equipped with a driving force distribution device of Embodiment 1 as a transfer as viewed from above the vehicle. It is a vertical side view of the driving force distribution device according to the first embodiment. It is a vertical front view which shows the crankshaft used with the driving force distribution apparatus of Example 1. FIG. FIG. 4 is an operation explanatory diagram of the driving force distribution device of the first embodiment, where (a) is an operation explanatory diagram showing a separated state of the first roller and the second roller at a position where the crankshaft rotation angle is 0 ° of the reference point; ) Is an operation explanatory view showing the contact state of the first roller and the second roller when the crankshaft rotation angle is 90 °, and (c) is the first roller and the second roller when the crankshaft rotation angle is 180 °. It is operation | movement explanatory drawing which shows the contact state of a roller. 6 is a flowchart illustrating a flow of a current command value calculation control process for an electromagnetic brake in the transfer controller according to the first embodiment. 3 is a calculation map of a current command value according to the crankshaft rotation angle of the first embodiment. It is a schematic diagram of the driving force distribution apparatus of Example 2. FIG. 6 is a schematic diagram of a driving force distribution device according to a third embodiment.

Hereinafter, embodiments of the present invention will be described in detail based on the illustrated examples.
[Example 1]
FIG. 1 is a schematic plan view showing a power train of a four-wheel drive vehicle equipped with the driving force distribution device 1 of the first embodiment as a transfer as viewed from above the vehicle.

  The four-wheel drive vehicle shown in FIG. 1 is a rear-wheel drive vehicle in which rotation from the engine 2 is transmitted to the left and right rear wheels 6L and 6R through the rear propeller shaft 4 and the rear final drive unit 5 after being shifted by the transmission 3. Is the base vehicle. Then, a part of the torque to the left and right rear wheels (main drive wheels) 6L, 6R is sequentially passed through the front propeller shaft 7 and the front final drive unit 8 by the driving force distribution device 1, and left and right front wheels (secondary drive wheels) 9L, 9R This is a vehicle that enables four-wheel drive traveling by transmitting to the vehicle.

  As described above, the driving force distribution device 1 distributes and outputs a part of the torque to the left and right rear wheels 6L and 6R to the left and right front wheels 9L and 9R, thereby outputting the left and right rear wheels 6L and 6R and the left and right front wheels 9L and 9R. In this embodiment, the driving force distribution device 1 is configured as shown in FIG.

In FIG. 2, reference numeral 11 denotes a housing of the driving force distribution device 1, and the input shaft 12 and the output shaft 13 are inclined in the housing 11 so that the respective rotation axes O ₁ and O ₂ intersect with each other. To do. The input shaft 12 is rotatably supported with respect to the housing 11 by needle roller bearings 14 and 15 at both ends thereof. Here, a ball bearing may be used instead of the needle roller bearing. Both ends of the input shaft 12 are protruded from the housing 11 under liquid-tight sealing by seal rings 25 and 26, respectively. 2, the left end of the input shaft 12 is drivingly coupled to the output shaft of the transmission 3 (see FIG. 1), and the right end is drivingly coupled to the rear final drive unit 5 via the rear propeller shaft 4 (see FIG. 1).

  Arranged near the both ends of the input shaft 12 and the output shaft 13, respectively, a pair of bearing supports 16, 17 are installed between the input / output shafts 12, 13, and the bearing supports 16, 17 are arranged in the middle of each bolt. (Not shown) is attached to the axially opposed inner wall of the housing 11. Here, the bearing supports 16 and 17 may not be fixed to the housing 11. The bearing supports 16 and 17 include input shaft through holes 16a and 17a through which the input shaft 12 passes, and output shaft through holes 16c and 17c through which the output shaft 13 and the crankshaft (second roller eccentric support member) 51L and 51R pass. And vertical walls 16b and 17b connecting between the input shaft through holes 16a and 17a and the output shaft through holes 16c and 17c, and has a substantially glasses shape in the axial front view. Roller bearings 21 and 22 are interposed between the bearing supports 16 and 17 and the input shaft 12 so that the input shaft 12 can be rotated with respect to the bearing supports 16 and 17. However, the input shaft 12 is rotatably supported in the housing 11.

A first roller 31 is coaxially formed integrally in the axial direction of the input shaft 12 between the bearing supports 16 and 17 (between the roller bearings 21 and 22), and hydraulic power can be transmitted to the first roller 31 through hydraulic oil. The second roller 32 is coaxially and integrally formed at an intermediate position in the axial direction of the output shaft 13 so as to be in frictional contact.
The outer peripheral surfaces 31a and 32a of the first roller 31 and the second roller 32 are conical tapered surfaces that can come into line contact with each other even when the input shaft 12 and the output shaft 13 are inclined as described above. Here, instead of the conical tapered surface, a crowning shape may be used. For example, the first roller 31 may have a convex crowning shape and the second roller 32 may have a concave crowning shape.
Thrust bearings 31cL, 31cR and 32cL, 32cR abut on both sides of the radially extending portions of the first roller 31 and the second roller 32, and holding grooves for holding the thrust bearings 31cL, 31cR and 32cL, 32cR in the radial direction. 31b and 32b are formed. The thrust bearings 31cL and 31cR contact the side walls 16a1 and 17a1 of the bearing supports 16 and 17, thereby positioning the first roller 31 in the axial direction. On the other hand, the thrust bearings 32cL and 32cR position the second roller 32 in the axial direction by abutting against roller side abutting portions 51Ld and 51Rd of crankshafts 51L and 51R, which will be described later.

The output shaft 13 is pivotally supported on the bearing supports 16 and 17 in the vicinity of both ends 13L and 13R, so that the output shaft 13 is pivotally supported in the housing 11 via the bearing supports 16 and 17.
Thus, when the output shaft 13 (13L, 13R) is pivotally supported on the bearing supports 16, 17, the following eccentric support structure is used.

A hollow outer shaft type crankshaft 51L, 51R is loosely fitted between the output shaft 13 (13L, 13R) and the bearing supports 16, 17 through which the output shaft 13 (13L, 13R) passes.
The crankshaft 51L and the output shaft 13 (13L) are protruded from the housing 11 at the left end in FIG. 2, respectively, and the seal ring 27 is interposed between the housing 11 and the crankshaft 51L at the protruding portion. By interposing a seal ring 28 between (13L), the crankshaft 51L protruding from the housing 11 and the protruding portion of the output shaft 13 (13L) are liquid-tightly sealed.

  In FIG. 2, the left end 13L of the output shaft 13 discharged from the housing 11 is drivingly coupled to the left and right front wheels 9L and 9R via the front propeller shaft 7 (see FIG. 1) and the front final drive unit 8.

The roller shafts 52L and 52R are interposed between the hollow holes 51La and 51Ra (radius Ri) of the crankshafts 51L and 51R and the corresponding ends 13L and 13R of the output shaft 13, respectively, so that the output shaft 13 (13L and 13R) crankshafts 51L, hollow hole 51La of 51R, in 51Ra, supports that can freely rotate around these central axis O _2.

As clearly shown in FIG. 3, the hollow holes 51La, 51Ra (center axis O ₂ ) of the crankshafts 51L, 51R are eccentric hollow holes eccentric to the outer peripheral portions 51Lb, 51Rb (center axis O ₃ , radius Ro). The central axis O ₂ of these eccentric hollow holes 51La and 51Ra is offset from the central axis O ₃ of the outer peripheral portions 51Lb and 51Rb by the eccentricity ε between them.
The outer peripheries 51Lb and 51Rb of the crankshafts 51L and 51R are rotatable to the inner periphery of the output shaft through holes 16c and 17c of the bearing supports 16 and 17 on the corresponding side via roller bearings 53L and 53R, which are radial bearings To support. Further, the roller side contact portions 51Ld and 51Rd of the crankshafts 51L and 51R are rotatably supported by the thrust bearings 32cL and 32cR. Further, the thrust bearings 32cL, 32cR and axially disposed thrust bearings 54L, 54R are provided.The thrust bearings 54L, 54R are in contact with the spacers 60L, 60R in a freely rotating manner, and ring gears 51Lc, 51Rc described later. The crankshafts 51L and 51R are rotatably supported by contacting with each other.

The spacers 60L and 60R are in contact with the wall surfaces 16b1 and 17b1 of the vertical walls 16b and 17b facing the second roller 32 and are on the inner diameter side of the inner peripheral surfaces of the output shaft through holes 16c and 17c. First spacer portions 61L and 61R extending to a position not in contact with 51R, and second spacer portions 62L and 62R (extending portions) extended so as to be inserted into the output shaft through holes 16c and 17c. .
Then, the spacers 60L and 60R are positioned in the radial direction by abutting between the outer periphery of the second spacer portions 62L and 62R and the inner peripheral surfaces of the output shaft through holes 16c and 17c, and the roller bearings 53L and 53R And mutual interference between the thrust bearings 54R and 54L.

  Ring gears 51Lc and 51Rc of the same specification are integrally provided at adjacent ends of the crankshafts 51L and 51R that face each other, and a common crankshaft drive pinion 55 is engaged with the ring gears 51Lc and 51Rc, respectively. 55 is coupled to the pinion shaft 56.

  As described above, when the crankshaft drive pinion 55 is engaged with the ring gears 51Lc and 51Rc, the crankshafts 51L and 51R are rotated in such a manner that their outer peripheral portions 51Lb and 51Rb are aligned with each other in the circumferential direction. In this state, the crankshaft drive pinion 55 is engaged with the ring gears 51Lc and 51Rc.

  Both ends of the pinion shaft 56 are rotatably supported with respect to the housing 11 by bearings 56a and 56b. A large-diameter output gear (reduction gear) 57b is fixed to the right end side of the pinion shaft 56 on the right side of FIG. On the outer diameter side of the large-diameter output gear 57b, as shown by an arrow A, a crank for detecting the rotation angle of the large-diameter output gear 57b by detecting irregularities 57b1 and 57b2 on the tooth surfaces of the large-diameter output gear 57b. A shaft rotation angle sensor 115 is provided. The crankshaft rotation angle sensor 115 is a magnetic sensor that detects a change in magnetic flux density due to a change in the unevenness of the tooth surface of the large-diameter output gear 57b, and determines the rotation angle of the pinion shaft 56 and thus the rotation angles of the crankshafts 51L and 51R. Detect.

  As described above, in the case of the rotation angle sensor that detects the unevenness of the tooth surface of the large-diameter output gear 57b, components such as a rotary encoder that detects the motor rotation angle require parts on both the rotating body side and the stator side. Compared to an expensive configuration, the rotation angle can be detected at low cost while achieving compactness in terms of space. In addition, since it can be attached from the outer periphery side of the housing 11 at a part of the outer peripheral space of the large-diameter output gear 57b, there can be an advantageous arrangement in terms of space.

  A small-diameter output gear (reduction gear) 57a that meshes with the large-diameter output gear 57b is provided on the outer periphery of the large-diameter output gear 57b. The small-diameter output gear 57a is integrally formed with the small-diameter output gear shaft 57a1, and is further assembled to the motor drive shaft 58a of the inter-roller radial direction pressing force control motor (electric motor) 35 on the left end side in FIG. And rotate together.

On the right end side of the small-diameter output gear shaft 57a1, an electromagnetic brake 59 that can fix the rotation of the small-diameter output gear shaft 57a1 by applying a braking force to the small-diameter output gear shaft 57a1 is provided. The electromagnetic brake 59 includes a coil 59a that generates electromagnetic force, and a clutch plate 59b that is spline-fitted at the right end of the small-diameter output gear shaft 57a1 so as to be capable of stroke in the axial direction.
A seal ring 63 and a seal ring 64 for sealing the electric motor 35 and the electromagnetic brake 59 against the inside of the housing 11 are provided on the left and right in the axial direction of the small diameter output gear 57a.

  The clutch plate 59b is provided with an armature. When the coil 59a is energized, the clutch plate 59b is moved in the axial direction by an electromagnetic attractive force and is attracted and fixed to the yoke on the outer periphery of the coil 59a. When the electromagnetic brake 59 is on (engaged state) by supplying current to the coil 59a, a braking torque corresponding to the current value is applied to the pinion shaft 56, and this braking torque acts on the pinion shaft 56 from the second roller 32 side. By setting the torque to be greater than or equal to the torque to be applied, the pinion shaft 56 can be fixed and the desired distance between the roller axes can be maintained. On the other hand, when the current supply to the coil 59a is stopped and the electromagnetic brake 59 is off (in a released state), the rotation operation of the electric motor 35 can be transmitted to the pinion shaft 56, so that a desired roller shaft distance can be achieved.

When the crankshafts 51L and 51R are rotationally controlled by the electric motor 35 via the pinion 55 and the ring gears 51Lc and 51Rc, the rotation axis O ₂ of the output shaft 13 and the second roller 32 is a locus circle indicated by a broken line in FIG. pivots about the central axis O ₃ along the alpha.

As will be described later in detail, the second roller 32 is rotated by the rotation axis O ₂ (second roller 32) along the locus circle α in FIG. 3 as shown in FIGS. 4 (a) to 4 (c). As the rotation angle θ of the crankshafts 51L and 51R increases with the distance L1 between the roller axes of the first roller 31 and the second roller 32 in the radial direction, the radius of the first roller 31 and the second roller 32 are increased. It can be made smaller than the sum value with the radius. Due to such a decrease in the distance L1 between the roller shafts, the radial pressing force of the second roller 32 against the first roller 31 (the transmission torque capacity between the rollers: the traction transmission capacity) increases. The inter-roller radial pressing force (inter-roller transmission torque capacity: traction transmission capacity), that is, the driving force distribution ratio can be arbitrarily controlled.

As shown in FIG. 4A, in the first embodiment, the second roller rotation axis O ₂ is located immediately below the crankshaft rotation axis O ₃ , and the inter-axis distance L1 between the first roller 31 and the second roller 32 is The distance L1 between the roller axes at the maximum bottom dead center is made larger than the sum of the radius of the first roller 31 and the radius of the second roller 32. Thereby, at the bottom dead center of the crankshaft rotation angle θ = 0 °, the first roller 31 and the second roller 32 are not pressed against each other in the radial direction, and traction is transmitted between the rollers 31 and 32. No traction transmission capacity = 0, and the traction transmission capacity is between 0 at the bottom dead center and the maximum value obtained at the top dead center (θ = 180 °) shown in Fig. 4 (c). Can be controlled arbitrarily. In the first embodiment, the description will be given assuming that the rotation angle reference point of the crankshafts 51L and 51R is the bottom dead center of the crankshaft rotation angle θ = 0 °.

[Driving force distribution]
The drive force distribution action of the transfer 1 described above with reference to FIGS. On the other hand, the torque that has reached the input shaft 12 of the transfer 1 from the transmission 3 (see FIG. 1) passes directly from the input shaft 12 through the rear propeller shaft 4 and the rear final drive unit 5 (both see FIG. 1) to the left and right rear wheels 6L. , Transmitted to 6R.

  On the other hand, the transfer 1 controls the rotational position of the crankshafts 51L and 51R via the pinion 55 and the ring gears 51Lc and 51Rc by the electric motor 35, and sets the distance L1 between the roller shafts (see FIG. 4) to the first roller 31 and the second roller When the roller is smaller than the sum of the radii of 32, the rollers 31 and 32 have a torque transfer capacity between the rollers according to the radial mutual pressing force. Therefore, depending on the torque capacity, the left and right rear wheels 6L and 6R A part of the torque is directed from the first roller 31 to the output shaft 13 via the second roller 32, and the left and right front wheels 9L and 9R can also be driven. Thus, the vehicle is capable of four-wheel drive running by driving all of the left and right rear wheels 6L and 6R and the left and right front wheels 9L and 9R.

  Note that the radial pressing reaction force between the first roller 31 and the second roller 32 during the transmission is received by the bearing supports 16 and 17 which are rotation support plates common to these rollers, and does not reach the housing 11. The radial pressing reaction force is 0 when the crankshaft rotation angle θ is 0 ° to 90 °, and increases as θ increases while the crankshaft rotation angle θ is 90 ° to 180 °. The maximum value is obtained when the crankshaft rotation angle θ is 180 °. These angle assignments can be freely set.

  During such four-wheel drive traveling, the rotation angle θ of the crankshafts 51L and 51R is 90 ° of the reference position as shown in FIG. 4B, and the first roller 31 and the second roller 32 are mutually In this case, when it is pressed by a radial pressing force corresponding to the offset amount OS and is in frictional contact, power is transmitted to the left and right front wheels 9L, 9R with a traction transmission capacity corresponding to the offset amount OS between these rollers. .

  Then, the crankshafts 51L and 51R are rotated from the reference position in FIG. 4B toward the top dead center of the crankshaft rotation angle θ = 180 ° shown in FIG. 4C to increase the crankshaft rotation angle θ. As a result, the distance L1 between the roller shafts further decreases and the mutual overlap amount OL of the first roller 31 and the second roller 32 increases. As a result, the first roller 31 and the second roller 32 increase the radial mutual pressing force. Thus, the traction transmission capacity between these rollers can be increased.

When the crankshafts 51L and 51R reach the top dead center position in FIG. 4 (c), the first roller 31 and the second roller 32 are in the radial direction with the maximum radial pressing force corresponding to the maximum overlap amount OL. The traction transmission capacity between them can be maximized. The maximum overlap amount OL is the sum of the eccentric amount ε between the second roller rotation axis O ₂ and the crankshaft rotation axis O ₃ and the offset amount OS described above with reference to FIG.

  As is apparent from the above description, the crankshaft rotation angle is controlled by rotating the crankshafts 51L and 51R from the rotation position of the crankshaft rotation angle θ = 0 ° to the rotation position of the crankshaft rotation angle θ = 180 °. As θ increases, the traction transmission capacity between rollers can be continuously changed from 0 to the maximum value. Conversely, by rotating the crankshaft 51L, 51R from the rotation position of the crankshaft rotation angle θ = 180 ° to the rotation position of θ = 0 °, the traction transmission between the rollers is reduced as the crankshaft rotation angle θ decreases. The capacity can be continuously changed from the maximum value to 0, and the traction transmission capacity between the rollers can be freely controlled by rotating the crankshafts 51L and 51R.

[Traction transmission capacity control]
Since the transfer 1 distributes and outputs a part of the torque to the left and right rear wheels 6L and 6R to the left and right front wheels 9L and 9R as described above during the four-wheel drive driving, the first roller 31 and the second roller It is necessary to make the traction transmission capacity between 32 correspond to the target front wheel driving force to be distributed to the left and right front wheels 9L, 9R, which can be obtained from the driving force of the left and right rear wheels 6L, 6R and the front and rear wheel target driving force distribution ratio.

In order to control the traction transmission capacity that meets this requirement, in the first embodiment, as shown in FIG. 1, a transfer controller 111 is provided to control the rotational position of the electric motor 35 (control of the crankshaft rotational angle θ). And
Therefore, the transfer controller 111 detects the signal from the accelerator opening sensor 112 that detects the accelerator pedal depression amount (accelerator opening) APO that adjusts the output of the engine 2, and the rotational peripheral speed Vwr of the left and right rear wheels 6L, 6R. A signal from the rear wheel speed sensor 113, a signal from the yaw rate sensor 114 that detects the yaw rate φ around the vertical axis passing through the center of gravity of the vehicle, and a crankshaft rotation angle sensor that detects the rotation angle θ of the crankshafts 51L and 51R The signal from 115 and the signal from the oil temperature sensor 116 for detecting the temperature TEMP of the hydraulic oil in the transfer 1 (housing 11) are input.

The transfer controller 111 performs the traction transmission capacity control of the transfer 1 (front and rear wheel driving force distribution control of a four-wheel drive vehicle) based on the detection information of the sensors 112 to 116 as described below.
That is, first, the transfer controller 111 first obtains the driving force and the front / rear wheel target driving force distribution ratio of the left and right rear wheels 6L, 6R based on the accelerator opening APO, the rear wheel speed Vwr, and the yaw rate φ in a known manner.
Next, the transfer controller 111 obtains a target front wheel driving force to be distributed to the left and right front wheels 9L and 9R from the driving force of the left and right rear wheels 6L and 6R and the front and rear wheel target driving force distribution ratio.

Further, the transfer controller 111 determines the radial pressing force between the rollers required for the first roller 31 and the second roller 32 to transmit the target front wheel driving force (the traction transmission capacity between the first roller 31 and the second roller 32). Of the crankshafts 51L and 51R (see FIGS. 2 and 3) necessary for realizing the radial pressing force between the rollers (the traction transmission capacity between the first roller 31 and the second roller 32). The rotation angle target value tθ, that is, the target turning position of the second roller axis O ₂ is calculated.

  Then, the transfer controller 111 changes the crankshaft rotation angle θ to the crankshaft rotation angle target value tθ in accordance with the crankshaft rotation angle deviation between the crankshaft rotation angle θ detected by the sensor 115 and the crankshaft rotation angle target value tθ. The electric motor 35 is driven and controlled so as to match. When the rotation angle θ of the crankshafts 51L and 51R coincides with the target value tθ by the drive control of the electric motor 35, the first roller 31 and the second roller 32 can mutually transmit the target front wheel driving force. The traction transmission capacity between the first roller 31 and the second roller 32 can be controlled so as to be the front and rear wheel target driving force distribution ratio.

When the rotation angle θ of the crankshafts 51L and 51R matches the target value tθ, the transfer controller 111 turns on the electromagnetic brake 59 to maintain the rotation angle θ and deactivates the electric motor 35.
At this time, the current value supplied to the electromagnetic brake 59 is set to the minimum current value necessary for maintaining the rotation angle of the crankshafts 51L and 51R, thereby suppressing power consumption.

FIG. 5 is a flowchart showing a flow of current command value calculation control processing of the electromagnetic brake 59 in the transfer controller 111 of the first embodiment, and each step will be described below.
In step S1, the crankshaft rotation angle θ is input.
In step S2, it is determined whether or not the crankshaft rotation angle θ matches the crankshaft rotation angle target value tθ. If YES, the process proceeds to step S3, and if NO, the process proceeds to return.
In step S3, a current command value corresponding to the crankshaft rotation angle θ is calculated with reference to the map shown in FIG.
In step S4, the supply current to the electromagnetic brake 59 is increased or decreased so that the supply current to the electromagnetic brake 59 becomes the current command value calculated in step S3.

  FIG. 6 is a calculation map of a current command value according to the crankshaft rotation angle θ of the first embodiment. The current command value is a minimum value (> 0 while the crankshaft rotation angle θ = 0 ° to 90 °). ), Θ is set so as to increase as θ increases in the range of 90 ° to 135 °, and decreases as θ increases in the range of 135 ° to 180 °. The characteristic of the current command value with respect to the crankshaft rotation angle θ shown in FIG. 6 is to maintain the crankshaft rotation angle θ against the radial pressing reaction force of the first roller 31 and the second roller 32, that is, the crankshaft 51L, This is the minimum current value for obtaining the braking torque of the electromagnetic brake 59 necessary for stopping the rotation of 51R.

According to the driving force distribution device 1 of the first embodiment described above, the following operational effects are obtained.
In the first embodiment, an electromagnetic brake (radial pressing force control actuator) 59 is provided that applies a braking torque corresponding to the amount of current supplied to the coil 59a to the crankshafts (second roller eccentric support members) 51L and 51R.
When the radial pressing force between the rollers 31 and 32 is maintained by adopting the above configuration, a braking force is generated by energizing the coil 59a of the electromagnetic brake 59, and the electric motor 35 is deactivated. In this state, the rotation of the second roller 32 around the eccentric axis can be stopped.
Also, since the braking force becomes zero by stopping energization of the coil 59a, the crankshafts 51L and 51R are driven to the left and right front wheels 9L and 9R by the radial pressing reaction force acting between the first roller 31 and the second roller 32. It returns to the position where the power distribution becomes zero.
Therefore, even when a failure occurs in the electric motor 35 during the four-wheel drive traveling, the shift to the two-wheel drive traveling can be realized.

In the first embodiment, the second roller eccentric support members that rotatably support the second roller 32 around the eccentric axis offset from the rotation axis of the second roller 32 are the crankshafts 51L and 51R, The electric motor 35 is the radial pressing force control actuator that rotationally drives the 51R.
By controlling the crankshaft rotation angle θ by the electric motor 35, high responsiveness and high reliability of the front and rear wheel driving force distribution control can be realized.

Further, in the first embodiment, a small-diameter output gear 57a is provided on the small-diameter output gear shaft 57a1 assembled to the motor drive shaft 58a of the electric motor 35, and the electromagnetic brake 59 is disposed on the opposite side of the electric motor 35 with the small-diameter output gear 57a interposed therebetween. Arranged.
By adopting the above configuration, compared with the case where the electromagnetic brake 59 is arranged closer to the electric motor side than the small-diameter output gear 57a, the small-diameter output gear shaft 57a1 has a small diameter from the assembly position with the motor drive shaft 58a. Since the shaft length to the output gear 57a can be shortened, the torsional rigidity of the small-diameter output gear shaft 57a1 can be increased even with the same shaft diameter, and the shaft strength can be improved. In other words, the shaft diameter of the small-diameter output gear shaft 57a1 can be made smaller, which is advantageous in terms of weight reduction.

[Example 2]
FIG. 7 is a schematic diagram of a driving force distribution device (transfer) 70 according to the second embodiment. In the second embodiment, the electromagnetic brake 59 is located closer to the electric motor 35 than the small-diameter output gear 57a, and the small-diameter output gear 57a. And the electric motor 35 is different from the first embodiment.
In FIG. 7, a seal ring 71 for sealing the electromagnetic brake 59 against the inside of the housing 11 is provided on the left side in the axial direction of the small diameter output gear 57a.
Since other configurations are the same as those of the first embodiment, illustration and description thereof are omitted.

According to the driving force distribution device 70 of the second embodiment, the following operational effects are obtained.
In the second embodiment, the small diameter output gear shaft 57a1 assembled to the motor drive shaft 58a of the electric motor 35 is provided with the small diameter output gear 57a, and the electromagnetic brake 59 is disposed closer to the electric motor 35 than the small diameter output gear 57a.
In the driving force distribution device 1 according to the first embodiment, the electric motor 35 and the electromagnetic brake 59 are separately attached to the housing 11, and a structure for sealing each of the electric motor 35 and the electromagnetic brake 59 with respect to the inside of the housing 11 is necessary. On the other hand, in the driving force distribution device 70 of the second embodiment, only the structure for sealing the electromagnetic brake 59 is sufficient, and the structure for sealing the electric motor 35 is not necessary. Therefore, the seal structure can be simplified and the structure can be simplified.
The driving force distribution device 1 according to the first embodiment requires two seal rings 63 and 64 (see FIG. 2), whereas the driving force distribution device 70 according to the second embodiment requires only one seal ring 71. Therefore, the number of parts can be reduced.
In addition, the effect by the structure other than the above is the same as that of Example 1.

Example 3
FIG. 8 is a schematic diagram of a driving force distribution device (transfer) 80 according to the third embodiment. In the third embodiment, the electromagnetic brake 59 is disposed closer to the electric motor 35 than the small-diameter output gear 57a. The second embodiment is different from the first embodiment in that it is disposed on the opposite side to the small-diameter output gear 57a.
In FIG. 8, a seal ring 81 that seals the electric motor 35 against the inside of the housing 11 is provided on the left side in the axial direction of the small diameter output gear 57a.
Since other configurations are the same as those of the first embodiment, illustration and description thereof are omitted.

According to the driving force distribution device 80 of the third embodiment, the following operational effects are obtained.
In the driving force distribution device 80 according to the third embodiment, the electromagnetic brake 59 is disposed on the opposite side of the small diameter output gear 57a with the electric motor 35 interposed therebetween.
By adopting the above configuration, the shaft length from the assembly position of the small-diameter output gear shaft 57a1 to the motor drive shaft 58a to the small-diameter output gear 57a can be made the same as in the first embodiment, so that the shaft strength is improved and the seal structure is simplified. And reduction of the number of parts can be realized.

1,70,80 Driving force distribution device (transfer)
6L, 6R left and right rear wheels
9L, 9R left and right front wheels
12 Input shaft
13 Output shaft
31 1st roller
32 Second roller
35 Electric motor (actuator for radial pressing force control)
51L, 51R Crankshaft (second roller eccentric support member)
57a Small-diameter output gear (reduction gear)
59 Electromagnetic brake
59a coil

Claims (5)

An input shaft that rotates with the main drive wheel transmission system;
A first roller provided on the input shaft;
An output shaft that rotates with the driven wheel transmission system;
A second roller provided on the output shaft;
A second roller eccentric support member that rotatably supports the second roller around an eccentric axis offset from the rotation axis of the second roller;
A radial pressing force control actuator for rotationally driving the second roller eccentric support member;
An electromagnetic brake for applying a braking force corresponding to the amount of current supplied to the coil to the second roller eccentric support member;
A driving force distribution device comprising: The driving force distribution device according to claim 1,
The second roller eccentric support member is a crankshaft that supports the second roller so as to be rotatable around an eccentric axis offset from the rotation axis of the second roller,
The driving force distribution device according to claim 1, wherein the radial pressing force control actuator is an electric motor that rotationally drives the crankshaft. The driving force distribution device according to claim 2,
A reduction gear is provided on the drive shaft of the electric motor,
The driving force distribution device, wherein the electromagnetic brake is disposed on the opposite side of the electric motor with the reduction gear interposed therebetween. In the driving force distribution device according to claim 1 or 2,
A reduction gear is provided on the drive shaft of the electric motor,
The driving force distribution device, wherein the electromagnetic brake is disposed closer to the electric motor than the reduction gear. The driving force distribution device according to claim 4,
The driving force distribution device according to claim 1, wherein the electromagnetic brake is disposed on the opposite side of the reduction gear with the electric motor interposed therebetween.

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