U.S. patent application number 09/770386 was filed with the patent office on 2001-08-09 for power transmitting system for four-wheel drive vehicle.
This patent application is currently assigned to HONDA GIKEN KOGYO KABUSHIKI KAISHA. Invention is credited to Arai, Kentaro, Arai, Yasunori, Murakami, Ryuichi, Takahashi, Susumu.
Application Number | 20010011622 09/770386 |
Document ID | / |
Family ID | 18558714 |
Filed Date | 2001-08-09 |
United States Patent
Application |
20010011622 |
Kind Code |
A1 |
Arai, Kentaro ; et
al. |
August 9, 2001 |
Power transmitting system for four-wheel drive vehicle
Abstract
A four-wheel drive vehicle having a first hydraulic pump
operated in operative association with the rotation of front
wheels, and a second hydraulic pump operated in operative
association with the rotation of rear wheels, such that a
difference in rotational speed is produced between the front and
rear wheels, a multi-plate clutch is brought into its engaged state
by a hydraulic pressure generated by the hydraulic pumps, whereby
the mode of the vehicle is shifted into a four-wheel drive mode. A
torque cam mechanism is disposed between a clutch piston and clutch
plates, so that when the difference in rotational speed is produced
between the front and rear wheels, the torque cam mechanism
produces an axial thrust force immediately to promptly bring the
multi-plate clutch into its engaged state. The engagement of the
multi-plate clutch is achieved with a sufficient engagement force
by the hydraulic pressure thereafter produced by the hydraulic
pump.
Inventors: |
Arai, Kentaro; (Wako-shi,
JP) ; Murakami, Ryuichi; (Wako-shi, JP) ;
Takahashi, Susumu; (Wako-shi, JP) ; Arai,
Yasunori; (Wako-shi, JP) |
Correspondence
Address: |
ARMSTRONG,WESTERMAN, HATTORI,
MCLELAND & NAUGHTON, LLP
1725 K STREET, NW, SUITE 1000
WASHINGTON
DC
20006
US
|
Assignee: |
HONDA GIKEN KOGYO KABUSHIKI
KAISHA
Tokyo
JP
|
Family ID: |
18558714 |
Appl. No.: |
09/770386 |
Filed: |
January 29, 2001 |
Current U.S.
Class: |
192/35 ; 180/249;
192/48.7; 192/54.52; 192/70.23 |
Current CPC
Class: |
F16D 47/00 20130101;
B60K 17/3505 20130101 |
Class at
Publication: |
192/35 ;
192/48.7; 192/54.52; 192/70.23; 192/85.0CA; 180/249 |
International
Class: |
F16D 047/00; B60K
017/348 |
Foreign Application Data
Date |
Code |
Application Number |
Feb 7, 2000 |
JP |
2000-34357 |
Claims
What is claimed is:
1. A power transmitting system for a four-wheel drive vehicle
including front wheels to which a driving force from an engine is
transmitted directly, and rear wheels to which a portion of the
driving force from the engine is transmitted indirectly through a
multi-plate clutch, said multi-plate clutch being brought into its
engaged state by a hydraulic pressure generated by hydraulic pumps
in accordance with a difference between rotational speeds of the
front wheels and the rear wheels, said power transmitting system
comprising: a torque cam mechanism including a first cam member
operated in operative association with the rotation of the front
wheels, and a second cam member operated in operative association
with the rotation of the rear wheels, said multi-plate clutch being
brought into the engaged state by an axial thrust force generated
in accordance with a difference between the rotational speeds of
said cam members, said torque cam mechanism being arranged so that
when the rotational speed of the front wheels is greater than that
of the rear wheels during forward traveling of the vehicle, the
thrust force is generated, and when the rotational speed of the
rear wheels is greater than that of the front wheels during forward
traveling of the vehicle, the thrust force is not generated.
2. A power transmitting system for a four-wheel drive vehicle
according to claim 1, wherein one of said first cam member and said
second cam member is connected through a frictional clutch to a
member rotated in operative association with one of the front
wheels and the rear wheels.
3. A power transmitting system for a four-wheel drive vehicle
according to claim 2, wherein the member rotated in operative
association with the front wheels and said first cam member are
connected to each other through said frictional clutch, and said
second cam member is fixed to the member rotated in operative
association with the rear wheels, the hydraulic pressure generated
by the hydraulic pumps urging the entire torque cam mechanism
axially through an end plate to bring said multi-plate clutch into
the engaged state, said power transmitting system further
comprising: a thrust bearing disposed between said end plate and
said first cam member at a location radially inward from an urging
portion of said second cam member for urging said multi-plate
clutch.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] The present invention relates to a four-wheel drive vehicle
including front wheels to which a driving force from an engine is
transmitted directly, and rear wheels to which a portion of the
driving force from the engine is transmitted indirectly through a
multi-plate clutch which is brought into its engaged state by a
hydraulic pressure generated by a difference between rotational
speeds of the front and rear wheels.
[0003] 2. Description of the Related Art
[0004] There is a conventional power transmitting system already
proposed by the present assignees in Japanese Patent No. 2516095,
wherein, whenever the front wheels are slipped to produce a
difference between the rotational speeds of the front and rear
wheels, a driving force is transmitted from the front wheels to the
rear wheels to automatically switch over the mode of the vehicle
from a two-wheel drive mode to a four-wheel drive mode.
[0005] The conventional system is designed such that when a
difference in rotational speed is produced between the front and
rear wheels, a hydraulic pressure is generated by a difference
between the amount of oil discharged from a first hydraulic pump
operated in operative association with the rotation of the front
wheel and the amount of oil discharged from a second hydraulic pump
operated in operative association with the rotation of the rear
wheels, and such hydraulic pressure causes the multi-plate clutch
to be brought into its engaged state, thereby transmitting the
driving force from the front wheels to the rear wheels to switch
over the mode of the vehicle to the four-wheel drive mode. However,
the conventional system suffers from a disadvantage in that there
is a time lag until the multi-plate clutch is brought into the
engaged state by the hydraulic pressure generated based on the
difference between the rotational speeds of the front and rear
wheels. Another disadvantage is that, whenever the forward movement
of the vehicle is initiated, the front wheels are slipped,
resulting in a low responsiveness for bringing the vehicle into the
four-wheel drive mode.
SUMMARY OF THE INVENTION
[0006] The present invention has been derived with the above
circumstance in view, and it is an object of the present invention
to ensure that in a four-wheel drive vehicle designed so that a
multi-plate clutch is brought into its engaged state by a hydraulic
pressure based on a difference between rotational speeds of front
and rear wheels, the responsiveness from the generation of the
difference between the rotational speeds of the front and rear
wheels to the shifting of the vehicle into the four-wheel drive
mode is enhanced.
[0007] To achieve the above object, there is provided a power
transmitting system for a four-wheel drive vehicle including front
wheels to which a driving force from an engine is transmitted
directly, and rear wheels to which a portion of the driving force
from the engine is transmitted indirectly through a multi-plate
clutch, the multi-plate clutch being brought into its engaged state
by a hydraulic pressure generated by hydraulic pumps in accordance
with a difference between rotational speeds of the front wheels and
the rear wheels. The power transmitting system comprises a torque
cam mechanism including a first cam member operated in operative
association with the rotation of the front wheels, and a second cam
member operated in operative association with the rotation of the
rear wheels, the multi-plate clutch being brought into the engaged
state by an axial thrust force generated in accordance with a
difference between the rotational speeds of the cam members, the
torque cam mechanism being arranged so that when the rotational
speed of the front wheels is greater than that of the rear wheels
during forward traveling of the vehicle, the thrust force is
generated, and when the rotational speed of the rear wheels is
greater than that of the front wheels during forward traveling of
the vehicle, the thrust force is not generated.
[0008] With the above arrangement, when the rotational speed of the
front wheels is greater than that of the rear wheels during forward
traveling of the vehicle, the first cam member and the second cam
member of the torque cam mechanism are rotated relative to each
other to generate the thrust force, thereby immediately bringing
the multi-plate clutch into the engaged state. Therefore, as soon
as the front wheels are slipped upon starting of the forward
movement of the vehicle or during sudden acceleration of the
vehicle moved forwards to generate the difference between the
rotational speeds, the driving force is transmitted from the front
wheels to the rear wheels. Thus, it is possible to enhance the
responsiveness for bringing the vehicle into the four-wheel drive
mode to enhance the running performance. Thereafter, the
multi-plate clutch is brought into the engaged state by the
hydraulic pressure generated by the hydraulic pumps with a small
time lag and, hence, a sufficient amount of driving force
transmitted from the front wheels to the rear wheels can be
ensured. On the other hand, when the rotational speed of the rear
wheels is greater than that of the front wheels due to a sudden
braking during forward traveling of the vehicle, the torque cam
mechanism generates no thrust force. Therefore, it is possible to
prevent the driving force from being transmitted from the front
wheels to the rear wheels to avoid interference with an ABS system
or the like.
[0009] One of the first cam member and the second cam member may be
connected through a frictional clutch to a member rotated in
operative association with one of the front wheels and the rear
wheels.
[0010] With the above arrangement, one of the first cam member and
the second cam member is connected through a frictional clutch to a
member rotated in operative association with one of the front
wheels and the rear wheels. Therefore, the moment that the relative
rotations of the front and rear wheels are produced, the first cam
member and the second cam member can be rotated relative to each
other to generate the thrust force. Before the multi-plate clutch
is thereafter brought into the completely engaged state by the
hydraulic pressure, the frictional clutch can be slipped to prevent
an excessive load from being applied to the torque cam
mechanism.
[0011] The member rotated in operative association with the front
wheels and the first cam member are connected to each other through
the frictional clutch, and the second cam member is fixed to the
member rotated in operative association with the rear wheels, so
that a hydraulic pressure generated by the hydraulic pumps urges
the entire torque cam mechanism axially through an end plate to
bring the multi-plate clutch into the engaged state, and a thrust
bearing is disposed between the end plate and the first cam member
at a location radially inward from an urging portion of the second
cam member for urging the multi-plate clutch.
[0012] With the above arrangement, the thrust bearing is disposed
between the end plate and the first cam member. Therefore, the
relative rotations of the end plate rotated in operative
association with the front wheels and the first cam member rotated
in operative association with the rear wheels after the operation
of the torque cam mechanism can be absorbed. Moreover, the thrust
bearing is disposed at a location radially inward from the urging
portion of the second cam member for urging the multi-plate clutch
and hence, the position of the thrust bearing can be displaced
radially inwards as much as possible to alleviate the load and to
enhance the durability.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] FIG. 1 is an illustration showing a power transmitting
system in a four-wheel drive vehicle.
[0014] FIG. 2 is a view showing a multi-plate clutch and a
hydraulic pressure circuit in the four-wheel drive vehicle.
[0015] FIG. 3 is an enlarged sectional view of the multi-plate
clutch.
[0016] FIG. 4 is an enlarged sectional view taken along a line 4-4
in FIG. 3.
[0017] FIG. 5 is a graph for explaining the operation.
[0018] FIG. 6 is a view similar to FIG. 3 but according to a second
embodiment of the present invention.
[0019] FIG. 7 is a view similar to FIG. 3 but according to a third
embodiment of the present invention.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0020] The present invention will now be described by way of
embodiments of the present invention shown in the accompanying
drawings.
[0021] FIGS. 1 to 5 show a first embodiment of the present
invention.
[0022] As shown in FIG. 1, an output from an engine E mounted at a
front portion of a four-wheel drive vehicle is input into a
differential 2 through a transmission 1 for front wheels, and an
output from the differential 2 is transmitted through drive shafts
3, 3 to left and right front wheels Wf, Wf. The output transmitted
from the engine E to the differential 2 is input into a power
transmitting device T which will be described hereinafter through a
bevel gear device 4, and an output from the power transmitting
device T is transmitted through a bevel gear device 5 to a
differential 6 for rear wheels. Further, an output from the
differential 6 is transmitted through drive shafts 7, 7 to left and
right rear wheels Wr, Wr.
[0023] The power transmitting device T is comprised of a first
hydraulic pump Pf driven by an input shaft 8 extending the bevel
gear device 4 for the front wheels, a second hydraulic pump Pr
driven by an output shaft 9 connected to the bevel gear device 5
for the rear wheels, a wet-type multi-plate clutch C which governs
the transmission and the interruption of a driving force between
the input shaft 8 and the output shaft 9, and a hydraulic pressure
circuit (which will be described hereinafter) for controlling the
multi-plate clutch C.
[0024] The arrangements of the multi-plate clutch and the hydraulic
pressure circuit will be described below with reference to FIGS. 2
and 3. The first hydraulic pump Pf comprises a trochoidal pump, and
includes a first port 10 which acts as a discharge port during
forward movement of the vehicle and acts as an intake port during
reverse movement of the vehicle, and a second port 11 which acts as
an intake port during forward movement of the vehicle and acts as a
discharge port during reverse movement of the vehicle. The second
hydraulic pump Pr likewise comprises a trochoidal pump, and
includes a third port 12 which acts as an intake port during
forward movement of the vehicle and acts as a discharge port during
reverse movement of the vehicle, and a fourth port 13 which acts as
a discharge port during forward movement of the vehicle and acts as
an intake port during reverse movement of the vehicle. The amounts
of oil discharged per one rotation by the hydraulic pumps Pf and Pr
are set such that the amount of oil discharged by the second
hydraulic pump Pr is slightly (for example, 2.5%) greater than the
amount of oil discharged by the first hydraulic pump Pf. The first
and second ports 10 and 12 are connected to each other through a
first connecting oil passage 14, and the third and fourth ports 11
and 13 are connected to each other through a second connecting oil
passage 15. The direction of oil discharged by each of the
hydraulic pumps Pf and Pr each comprising the trochoidal pump
depends on the rotational direction of the pump and hence, is
inverted between forward and backward movements of the vehicle.
Arrows in the hydraulic pumps Pf and Pr in FIG. 2 indicate
directions of oil discharged during the forward movement of the
vehicle.
[0025] The multi-plate clutch C includes a clutch housing 41
secured to the input shaft 8, and a clutch hub 42 secured to the
output shaft 9, which is coaxially and relatively rotatably fitted
in a rear end of the input shaft 8 with a roller bearing 29
interposed therebetween. A plurality of clutch plates 43 slidably
carried in a spline 41a defined around an inner periphery of the
clutch housing 41 and a plurality of clutch plates 44 slidably
carried in a spline 42a defined around an outer periphery of the
clutch hub 42 are superposed on one another, so that they can be
put into abutment against one another. A clutch piston 47 is
slidably received in a clutch cylinder 46 defined in a casing 45
with seal members 48, 48 interposed therebetween, and a working
hydraulic pressure chamber 16 is defined on a right side of the
clutch piston 47, so that a hydraulic oil for driving the clutch
piston 47 is supplied into the working hydraulic pressure chamber
16.
[0026] A torque cam mechanism 61 disposed at a rear end of the
multi-plate clutch C comprises a first cam member 62, a second cam
member 63 and a plurality of cam balls 64. The first cam member 62
located on a rear side is a substantially annular member, and is
relatively non-rotatably and axially movably carried at its outer
peripheral portion in a rear portion of the spline 41a in the
clutch housing 41, and relatively non-rotatably and axially movably
carried at its inner peripheral portion at a front end of a sleeve
65 relatively rotatably fitted over an outer periphery of the
output shaft 9. The sleeve 65 is connected to an inner rotor of the
first hydraulic pump Pf and drives the first hydraulic pump Pf in
operative association with the front wheels Wf, Wf.
[0027] The second cam member 63 of the torque cam mechanism 61 is a
substantially annular member superposed on a front surface of the
first cam member 62, and has an urging portion 66 protruding
forwards from an outer periphery at a front end for urging the
clutch plates 43 and 44 of the multi-plate clutch C forwards to
bring them into engagement with one another. The second cam member
63 is supported for slipping movement with a predetermined torque
relative to the clutch hub 42 by a frictional clutch 67 disposed
between a spline 63b defined in an inner periphery of the second
cam member 63 and the spline 42a in the clutch hub 42. A thrust
bearing 68 is disposed between a front surface of the clutch piston
47 and a rear surface of the first cam member 62.
[0028] A cone member 69 extending from the second cam member 63
toward the inside of the clutch hub 42 includes a large number of
small bores, and has a function to disperse a lubricating oil
supplied through an oil passage 9a and oil bores 9b defined in the
output shaft 9 by a centrifugal force to equally lubricate the
clutch plates 43 and 44 of the multi-plate clutch C. In this case,
the output shaft 9 may be a solid shaft, and an oil passage may be
defined between the output shaft 9 and the sleeve 65, so that the
lubricating oil is supplied through the cone member 69 to lubricate
the clutch plates 43 and 44.
[0029] As can be seen from FIGS. 4A and 4B, six recesses 62a, 63a
are defined at distances of 60.degree. in each of opposed surfaces
of the first and second cam members 62 and 63 of the torque cam
mechanism 61, and the cam ball 64 is accommodated between each pair
of the recesses 62a and 63a. The shape of each of the recesses 62a
and 63a is such that a deeper portion accommodating half of the cam
ball 64 and a portion gradually decreased in depth from the deeper
portion are formed continuously to each other. Thus, when the first
and second cam members 62 and 63 are in a phase relationship shown
in FIG. 4A, they are close to each other with their opposed
surfaces being in close contact with each other. When the first and
second cam members 62 and 63 are rotated relative to each other in
a direction of an arrow a from this state, they are moved relative
to each other away from each other, as shown in FIG. 4B, by an
axial thrust force f generated by an effect of the shapes of the
recesses 62a and 63a.
[0030] When the first and second cam members 62 and 63 are in a
phase relationship shown in FIG. 4A, even if they are intended to
be rotated relative to each other in a direction of an arrow b, the
relative rotation of the first and second cam members 62 and 63 in
the direction of the arrow b is limited by the effect of the shapes
of the recesses 62a and 63a and hence, the axial thrust force f is
not generated. Namely, the torque cam mechanism 61 also has a
one-way cam function.
[0031] As can be seen from FIG. 2, the working hydraulic pressure
chamber 16 in the multi-plate clutch C and the first connecting oil
passage 14 are connected to each other through a third connecting
oil passage 17, and the working hydraulic pressure chamber 16 and
the second connecting oil passage 15 are connected to each other
through a fourth connecting oil passage 18. A first one-way valve
19 is provided in the third connecting oil passage 17 for
permitting only an oil flow from the second hydraulic pump Pr to
the working hydraulic pressure chamber 16, and a second one-way
valve 20 is provided in the fourth connecting oil passage 18 for
permitting only an oil flow from the second connecting oil passage
15 to the working hydraulic pressure chamber 16. Provided in a
fifth connecting oil passage 36 connecting an oil tank 21 with the
first connecting oil passage 14 and the second connecting oil
passage 15 are a third one-way valve 22 for permitting only an oil
flow from the oil tank 21 to the first connecting oil passage 14,
and a fourth one-way valve 23 for permitting only an oil flow from
the oil tank 21 to a port 33c.
[0032] A choke-type constriction 24 is provided at a location
upstream of the working hydraulic pressure chamber 16 in the
multi-plate clutch C. An orifice-type constriction 25 and a first
relief valve 26 are provided in series at locations downstream of
the working hydraulic pressure chamber 16 and connected at a
downstream location to the oil passage 9a defined in the output
shaft 9. The oil passage 9a communicates with an area to be
lubricated in the multi-plate clutch C, i.e., an internal space in
the clutch housing 41 having the clutch plates 43 and 44
accommodated therein, through the plurality of oil bores 9b
provided radially through the output shaft 9.
[0033] The first relief valve 26 has a function to prevent air from
flowing backwards from the area to be lubricated in the multi-plate
clutch C to the working hydraulic pressure chamber 16, when the
clutch piston 47 of the multi-plate clutch C is swung by eccentric
rotations of the input shaft 8 and the output shaft 9.
[0034] A second relief valve 28 is provided between a location
upstream of the choke-type constriction 24 and a location
downstream of the orifice-type constriction 25 for limiting an
upper limit value for a hydraulic pressure transmitted to the
working hydraulic pressure chamber 16. The second relief valve 28
is provided with a thermo-switch 51 for forcibly opening the second
relief valve 28 upon an increase in oil temperature.
[0035] A spool valve 32 is provided in the second connecting oil
passage 15 and comprises a spool 31 accommodated in a housing and
biased rightwards by a spring 30. When the spool 31 is in a right
position shown in FIG. 2, the communication between the port 33c
and a port 33d is cut off, and a port 33a and a port 33b are in
communication with each other. When the spool 31 is moved to a left
position against a biasing force of the spring 30, the
communication between the port 33a and the port 33b is cut off by
the spool 31, and the port 33c and the port 33d are put into
communication with each other. A fifth one-way valve 34 is provided
between the port 33a and the port 33d for permitting only an oil
flow from the port 33d to the port 33a, and a sixth one-way valve
35 is provided between the port 33b and the port 33c for permitting
only an oil flow from the port 33b to the port 33c.
[0036] Therefore, during the forward movement of the vehicle, i.e.,
when the second hydraulic pump Pr is discharging the oil through
the fourth port 13, the spool 31 is moved to the left position,
whereby the second connecting oil passage 15 and the first
connecting oil passage 14 are connected to each other by the
communication between the port 33c and the port 33d. On the other
hand, during the backward movement of the vehicle, when the first
hydraulic pump Pf is discharging the oil through the second port
11, the spool 31 is in the right position shown in FIG. 2, whereby
the second connecting oil passage 15 and the first connecting oil
passage 14 are connected to each other by the communication between
the port 33a and the port 33b.
[0037] When the spool 31 of the spool valve 32 is in the right
position, a port 33e put out of communication with the port 33c by
the spool 31 is in communication with the oil passage 9a in the
output shaft 9 through a lubricating oil passage 53 provided in a
seventh one-way valve 52.
[0038] The operation of the first embodiment of the present
invention will be described below.
[0039] Upon the initiation of the forward movement of the vehicle,
the driving force from the engine E is transmitted through the
transmission 1, the differential 2 and the drive shafts 3, 3 to the
front wheels Wf, Wf. The driving force is also transmitted from the
differential 2 through the bevel gear device 4 and the input shaft
8 to the first hydraulic pump Pf to drive the first hydraulic pump
Pf. At this time, the multi-plate clutch C is in a non-engaged
state, and the second hydraulic pump Pr connected to the output
shaft 9 is in a stopped state. Therefore, the oil drawn from the
oil tank 21 through the fifth one-way valve 34 into the second port
11 in the first hydraulic pump Pf is discharged through the first
port 10 into the first connecting oil passage 14. At this time, the
third one-way valve 22 in the fifth connecting oil passage 36 is in
a closed state and hence, the entire amount of oil discharged into
the first connecting oil passage 14 flows into the third connecting
oil passage 17, where the flowing of the oil is obstructed by the
second one-way valve 20. Thus, the oil is supplied through the
first one-way valve 19 and the choke-type constriction 24 into the
working hydraulic pressure chamber 16 in the multi-plate clutch
C.
[0040] When the multi-plate clutch C is brought into an engaged
state in the above manner, the rear wheels Wr, Wr are driven
through the output shaft 9, the bevel gear device 5, the
differential 6 and the drive shafts 7, 7, and the second hydraulic
pump Pr connected to the output shaft 9 is rotated. As a result,
the oil discharged from the first hydraulic pump Pf is drawn into
the second hydraulic pump Pr through the first connecting oil
passage 14 in response to an increase in rotational speed of the
rear wheels Wr, Wr, and the oil discharged from the second
hydraulic pump Pr is drawn into the first hydraulic pimp Pf through
the ports 33c and 33d and the fifth one-way valve 34, while moving
the spool 31 of the spool valve 32 leftwards against the biasing
force of the spring 30. The hydraulic pressure applied to the
working hydraulic pressure chamber 16 in the multi-plate clutch C,
i.e., the engagement force of the multi-plate clutch C, is varied
automatically in accordance with a difference between the amount of
oil discharged from the first hydraulic pump Pf and the amount of
oil drawn into the second hydraulic pump Pr. When the vehicle has
reached, for example, a constant-speed forward-traveling state in
which the difference between the rotational speeds of the front and
rear wheels is substantially equal to 0 (zero), the hydraulic
pressure is not applied to the working hydraulic pressure chamber
16 in the multi-plate clutch C, whereby the distribution of the
torque to the rear wheels Wr, Wr is cut off. In the constant-speed
forward-traveling state, the amount of oil discharged from the
second hydraulic pump Pr is slightly greater than the amount of oil
discharged from the first hydraulic pump Pf, as described above,
but the oil discharged from the second hydraulic pump Pr acts to
move the spool 31 of the spool valve 32 leftwards against the
biasing force of the spring 30, and the surplus amount of oil
discharged from the second hydraulic pump Pr is circulated through
the ports 33c and 33d and the third one-way valve 22 in the fifth
connecting oil passage 36 to the third port 12 in the second
hydraulic pump Pr.
[0041] When the oil discharged from the first hydraulic pump Pf and
the second hydraulic pump Pr is circulated through the first
connecting oil passage 14 and the second connecting oil passage 15
in the above-described constant-speed forward-traveling state, a
hydraulic pressure corresponding to the biasing force of the spring
30 is generated in the second connecting oil passage 15 between the
fourth port 13 and the port 33c by moving the spool 31 of the spool
valve 32 leftwards against the biasing force of the spring 30 by
the oil discharged from the second hydraulic pump Pr. As a result,
air drawn from a side-clearance in each of the hydraulic pumps Pf
and Pr into the circulated oil is compressed by such hydraulic
pressure and discharged sequentially through the side-clearance of
the second hydraulic pump at a location closer to the fourth port
13 and hence, the air cannot be resident in the circulated oil.
Thus, it is possible to reliably prevent a disadvantage produced
when a difference is thereafter generated between rotational speeds
of the front wheels Wf, Wf and the rear wheels Wr, Wr, whereby a
difference is generated between the amounts of oil discharged from
(or drawn into) the first and second hydraulic pumps Pf and Pr, the
rise in hydraulic pressure is retarded due to the resident air and
as a result, the responsiveness of the multi-plate clutch is
reduced.
[0042] Now, when the difference has been generated between the
amounts of oil discharged from (or drawn into) the first and second
hydraulic pumps Pf and Pr, a hydraulic pressure corresponding to a
preset load of the first relief valve 26 is applied immediately to
the working hydraulic pressure chamber 16 in the multi-plate clutch
C. After the first relief valve 26 has been opened, a hydraulic
pressure determined by a difference between amounts of oil
discharged from the first and second hydraulic pumps Pf and Pr,
pressure drop characteristics of the orifice-type constriction 25
and the choke-type constriction 24, the viscosity of the oil or the
like is applied to the working hydraulic pressure chamber 16 in the
multi-plate clutch C. The upper limit value for such hydraulic
pressure is limited by the preset load of the second relief valve
28 and hence, the upper limit value for the torque transmitted from
the multi-plate clutch C can be regulated by properly setting the
preset load of the second relief valve 28.
[0043] The amount of oil passing through the choke-type
constriction 24 is influenced by the viscosity of the oil, so that
when the viscosity of the oil in a low-temperature state is
increased, the amount of flow through the choke-type constriction
24 is decreased and hence, the amount of oil passing through the
working hydraulic pressure chamber 16 in the multi-plate clutch C
and the orifice-type constriction 25 is also decreased. At this
time, the amount of drop in pressure generated across the
orifice-type constriction 25 is proportional to the square of the
amount of oil passed through the orifice-type constriction 25 and,
hence, if the amount of oil passing through the orifice-type
constriction 25 is decreased, the amount of drop in pressure in the
orifice-type constriction 25 is decreased, and the amount of drop
in pressure in the upstream choke-type constriction 25 is
correspondingly increased. Thus, the hydraulic pressure applied to
the working hydraulic pressure chamber 16 in the low-temperature
state, i.e., the pressure resulting from the subtraction of the
amount of drop in pressure produced by the choke-type constriction
24 from the pressure set by the second relief valve 28, is
decreased. Therefore, even if the frictional coefficient is
increased by an increase in viscosity of the oil, the urging force
provided for the clutch plates 43 and 44 by the hydraulic pressure
is correspondingly decreased and hence, an increase in engagement
force of the multi-plate clutch C at a low temperature is prevented
as a whole. On the other hand, in a high-temperature state, the
viscosity of the oil is decreased, whereby the frictional
coefficient is decreased. In this case, the amount of drop in
pressure provided by the choke-type constriction 24 is decreased,
and the hydraulic pressure applied to the working hydraulic
pressure chamber 16 in the multi-plate clutch C is increased.
Therefore, the urging force for the clutch plates 43 and 44 is
correspondingly increased to prevent a decrease in engagement force
of the multi-plate clutch C.
[0044] (1) The oil discharged from the working hydraulic pressure
chamber 16 in the multi-plate clutch C through the orifice-type
constriction 25 and the first relief valve 26 upon the start of the
forward movement of the vehicle or during sudden acceleration of
the vehicle moved forwards and (2) the oil discharged from an
upstream position in the working hydraulic pressure chamber 16
through the second relief valve 28, are supplied through the oil
passage 9a and the oil bores 9b in the output shaft 9 into the
multi-plate clutch C, where the oil is scattered radially outwards
from the oil bore in the cone member 69 rotated relative to the
output shaft 9 by a centrifugal force to equally lubricate the
clutch plates 43 and 44. The oil, which has lubricated the clutch
plates 43 and 44, is circulated through an oil passage (not shown)
to the oil tank 21.
[0045] During forward traveling of the vehicle at a constant speed,
the pressure oil is not supplied to the working hydraulic pressure
chamber 16 in the multi-plate clutch C, as described above, and
hence, the lubricating oil is not supplied via the working
hydraulic pressure chamber 16 to the area to be lubricated in the
multi-plate clutch C. However, when the oil discharged from the
fourth port 13 in the second hydraulic pump Pr moves the spool 31
of the spool valve 32 leftwards against the biasing force of the
spring 30, a predetermined hydraulic pressure is generated and
moreover, an excessive amount of the oil is discharged into the
second connecting oil passage 15 by the difference between the
amounts of oil discharged per rotation from the hydraulic pumps Pf
and Pr. Therefore, a portion of the surplus oil is supplied through
the lubricating oil passage 53 to the oil passage 9a in the output
shaft 9 by a pressure generated upon opening of the spool valve 32,
and is then supplied from the oil passage 9a through the oil bores
9b to the area to be lubricated in the multi-plate clutch C. In
this manner, even during forward traveling of the vehicle at the
constant speed with the multi-plate clutch C being in the
non-engaged state, the lubricating oil can be supplied to the area
to be lubricated in the multi-plate clutch C to effectively
lubricate the clutch plates 43 and 44, thereby preventing the
over-heating of the multi-plate clutch C.
[0046] The seventh one-way valve 52 provided in the lubricating oil
passage 53 exhibits a function, which will be described below. When
a negative pressure or vacuum is generated in the second connecting
oil passage 15 upon the start of the forward movement of the
vehicle or during sudden acceleration of the vehicle moved
forwards, the air drawn from the area to be lubricated in the
multi-plate clutch C can be prevented from being drawn into the
first hydraulic pump Pf through the lubricating oil passage 53, the
port 33e, the port 33d, the fifth one-way valve 34 and the second
connecting oil passage 15.
[0047] When only the front wheels Wf, Wf have treaded on a road
surface of a low frictional coefficient during forward traveling of
the vehicle at a constant speed, or when a driver has attempted to
suddenly accelerate the vehicle, the front wheels Wf, Wf may be
brought into an excessively slipping state in some cases. In such a
state, the amount of oil discharged from the first hydraulic pump
Pf connected to the input shaft 8 is greater than the amount of oil
drawn into the second hydraulic pump Pr connected to the output
shaft 9, and the third one-way valve 22 is closed to cut off the
communication between the first connecting oil passage 14 and the
second connecting oil passage 15 through the fifth connecting oil
passage 36. Therefore, the multi-plate clutch C is likewise brought
into the engaged state to distribute the driving torque to the rear
wheels Wr, Wr.
[0048] When a braking force is applied to the wheels, the front
wheels Wf, Wf are locked earlier than the rear wheels Wr, Wr upon
hard braking, because the distribution of the braking force to the
front and rear wheels is generally set such that the braking force
on the front wheels Wf, Wf is greater than that on the rear wheels
Wr, Wr. An engine brake from the traveling of the vehicle at the
constant speed is applied to only the front wheels Wf, Wf, and even
in this case, the rotational speed of the front wheels Wf, Wf is
transiently lower than that of the rear wheels Wr, Wr. In such a
case, the amount of oil discharged from the second hydraulic pump
Pr is greater than the amount of oil drawn into the first hydraulic
pump Pf, and an excessive amount of the oil is discharged to the
second connecting oil passage 15. Further, when the front wheels
Wf, Wf have been locked completely, the operation of the first
hydraulic pump Pf is stopped, and only the second hydraulic pump Pr
is rotated and, hence, the total amount of oil discharged from the
second hydraulic pump Pr is excessive. However, the excessive
amount of the discharged oil is circulated through the port 33c and
the port 33d in the spool valve 32 and the third one-way valve 22
in the fifth connecting oil passage 36 to the third port 12 in the
second hydraulic pump Pr. Even if the rotational speed of the rear
wheel Wr, Wr is greater than that of the front wheels Wf, Wf in the
above manner, a hydraulic pressure based on a difference between
the amounts of oil discharged from the first and second hydraulic
pumps Pf and Pr is not applied to the working hydraulic pressure
chamber 16 in the multi-plate clutch C. Therefore, the multi-plate
clutch C is maintained in the non-engaged state to inhibit the
transmission of the braking force from the front wheels Wf, Wf to
the rear wheels Wr, Wr, whereby a variation in distribution of the
braking force to the front and rear wheels cannot be produced.
[0049] During the above-described braking of the vehicle moved
forwards, the multi-plate clutch C is not brought into the engaged
state and, hence, the lubrication of the multi-plate clutch C by
the oil passed through the first relief valve 26 is not carried
out. However, as in the above-described forward movement of the
vehicle at the constant speed, a portion of the oil discharged from
the second hydraulic pump Pr is supplied to the multi-plate clutch
C through the spool valve 32 and the lubricating oil passage 53,
whereby the lubrication of the area to be lubricated in the
multi-plate clutch C is carried out without hindrance.
[0050] During reverse movement of the vehicle, both of the
rotational directions of the first and second hydraulic pumps Pf
and Pr are inverted, thereby producing an inverse relationship
between the discharge port and the intake port.
[0051] More specifically, when the rotational speed of the front
wheels Wf, Wf is greater than that of the rear wheels Wr, Wr upon
the start of the reverse movement of the vehicle or during sudden
acceleration of the vehicle moving in reverse, the amount of oil
discharged from the first hydraulic pump Pf is greater than the
amount of oil drawn into the second hydraulic pump Pr and, hence, a
hydraulic pressure is produced in the second connecting oil passage
15. At this time, the spool 31 of the spool valve 32 is retained at
a shown position under the action of the biasing force of the
spring 30, and the oil discharged into the second connecting oil
passage 15 by means of the difference between the amount of oil
discharged from the second port 11 in the first hydraulic pump Pf
and the amount of oil drawn into the second hydraulic pump Pr
through the fourth port 13 is inhibited from flowing into the fifth
connecting oil passage 36 by the fourth one-way valve 23 and the
fifth one-way valve 34, and is permitted to flow into the fourth
connecting oil passage 18, as described above, where the oil passes
through the second one-way valve 20, and is then supplied to the
working hydraulic pressure chamber 16 in the multi-plate clutch C
in such a manner that the flow of the oil is inhibited by the first
one-way valve 19. This causes the multi-plate clutch C to be
brought into the engaged state in order to distribute the driving
torque to the rear wheels Wr, Wr. When the rotational speed of the
rear wheels Wr, Wr and as a result, the vehicle is brought into a
constant-speed reverse-traveling state, the rotational speeds of
the first and second hydraulic pumps Pf and Pr become equal to each
other. However, the amount of oil discharged per rotation from the
second hydraulic pump Pr is greater than the amount of oil
discharged per rotation from the first hydraulic pump Pf and,
hence, an amount of the oil corresponding to a difference between
such amounts of oil discharged is supplied to the first connecting
oil passage 14. As a result, during reverse movement of the
vehicle, the torque is distributed from the front wheels Wf, Wf to
the rear wheels Wr, Wr even in the constant-speed traveling state
of the vehicle.
[0052] In this constant-speed, reverse-traveling state, the load
provided by the spring 30 of the spool valve 32 is not applied to
the oil circulating through a circulation oil passageway comprising
the first and second connecting oil passages 14 and 15. However,
the constant-speed, reverse-traveling state cannot be generally
continued for a long time and hence, the drawing of air from the
side-clearances of the rotors of the hydraulic pumps Pf and Pr and
the stoppage of the supplying of the lubricating oil to the
multi-plate clutch C are substantially not problematic.
[0053] During braking of the vehicle moving in reverse, the
rotational speed of the first hydraulic pump Pf is less than that
of the second hydraulic pump Pr and hence, a hydraulic pressure is
generated by means of a difference between the amount of oil
discharged from the second hydraulic pump Pr through the third port
12 and the amount of oil drawn into the first hydraulic pump Pf
through the first port 10. At this time, the third one-way valve 22
is closed and, hence, the multi-plate clutch C is brought into the
engaged state through the first one-way valve 19 in the third
connecting oil passage 17, whereby the braking force for the front
wheels Wf, Wf is transmitted to the rear wheels Wr, Wr.
[0054] During the above-described reverse movement of the vehicle,
the second port 11 in the first hydraulic pump Pf serves as a
discharge port, and the fourth port 13 in the second hydraulic pump
Pr serves as an intake port. Therefore, the spool 31 of the spool
valve 32 is always retained at a shown right position. Even when
the spool 31 is locked at a left position for any reason at that
time, the pressure of oil discharged from the fist hydraulic pump
Pf through the second port 11 is obstructed by the fifth one-way
valve 34 and applied to the port 33a in the spool valve 32, whereby
the locked spool is pushed back to the right position, which is a
normal position. At this time, even when the locking of the spool
31 is not released, the hydraulic pressure in the second connecting
oil passage 15 escapes from the second relief valve 28 through the
fourth connecting oil passage 18 and, hence, an excessive load
cannot be applied to the first hydraulic pump Pf.
[0055] The above-described engagement of the multi-plate clutch is
performed by advancing the clutch piston 47 by the oil supplied to
the working hydraulic pressure chamber 16 and by urging the clutch
plates 43 and 44 by the urging portion of the second cam member 63
of the torque cam mechanism 61 urged axially trough the needle
bearing 68. At this time, the torque cam mechanism 61 interposed
between the clutch piston 47 and the clutch plates 43 and 44
exhibits a function which will be described below.
[0056] As described for explaining the structure of the hydraulic
pressure circuit, and as also shown in FIG. 5, the multi-plate
clutch C is brought into the engaged state during acceleration of
the vehicle moved forwards, during acceleration of the vehicle
moving in reverse and during deceleration of the vehicle moving in
reverse, whereby the transmission of the torque is carried out
between the front wheels Wf, Wf and the rear wheels Wr, Wr Only
during deceleration of the vehicle moved forwards, the multi-plate
clutch C is not brought into the engaged state, whereby the
transmission of the torque is not carried out between the front
wheels Wf, Wf and the rear wheels Wr, Wr. As described above, the
torque cam mechanism 61 also has the function as the one-way cam
and during acceleration of the vehicle moved forwards, the first
cam member 62 and the second cam member 63 of the torque cam
mechanism 61 are rotated relative to each other from the state
shown in FIG. 4A to the state shown in FIG. 4B to generate a thrust
force f for bringing the multi-plate clutch C into the engaged
state. Therefore, as shown in a right and upper portion of FIG. 5,
the torque cam mechanism 61 can be operated at an initial stage of
the acceleration of the vehicle moved forwards occurring with a
high frequency during traveling of the vehicle to assist in the
engagement of the multi-plate clutch C performed the hydraulic
pressure, thereby enhancing the responsiveness for bringing the
vehicle into a four-wheel drive mode.
[0057] Moreover, the engagement force generated in the multi-plate
clutch C by the torque cam mechanism 61 is gradually decreased in
accordance with an increase in engagement force generated in the
multi-plate clutch C by the hydraulic pressure. Therefore, it is
possible to prevent the multi-plate clutch C from being brought
into the engaged state only by the torque cam mechanism 61 to avoid
the occurrence of the differential locking.
[0058] As shown in a right and lower portion of FIG. 5, during the
deceleration of the vehicle moving forwards, the engagement of the
multi-plate clutch C by the hydraulic pressure is not performed and
moreover, the directions of relative rotation of the first and
second cam members 62 and 63 are inverted, and hence, the thrust
force f for the torque cam mechanism 61 to bring the multi-plate
clutch C into the engaged state by the function of the one-way cam
cannot be generated. Therefore, during the deceleration of the
vehicle moving forwards, the vehicle is maintained in a two-wheel
drive mode to avoid the interference with an ABS system, thereby
ensuring the braking performance of the vehicle.
[0059] As shown in a left and upper portion of FIG. 5, during the
acceleration of the vehicle moving in reverse the engagement of the
multi-plate clutch C by the hydraulic pressure is performed, but
the thrust force f for the torque cam mechanism 61 to bring the
multi-plate clutch C into the engaged state by the function of the
one-way cam cannot be generated.
[0060] As shown in a left and lower portion of FIG. 5, during the
deceleration of the vehicle moving in reverse, the engagement of
the multi-plate clutch C by the hydraulic pressure is performed,
and the thrust force f for the torque cam mechanism 61 to bring the
multi-plate clutch C into the engaged state by the function of the
one-way cam is generated, leading to a state in which the thrust
force of the torque cam mechanism 61 assists in the engagement of
the multi-plate clutch C by the hydraulic pressure.
[0061] Thus, at the initial stage of the acceleration of the
vehicle moved forwards actually occurring with a high frequency,
the multi-plate clutch C can be brought into the engaged state by
both of the torque cam mechanism 61 and the hydraulic pressure,
thereby enhancing the responsiveness for bringing the vehicle into
the four-wheel drive mode and distributing the sufficient driving
force to the rear wheels Wr, Wr. Likewise, during the deceleration
of the vehicle moving forwards actually occurring at a high
frequency, the vehicle can be maintained in the two-wheel drive
mode to avoid interference with the ABS system.
[0062] Even after the torque cam mechanism 61 has been operated to
provide the state shown in FIG. 4B, thereby inhibiting the relative
rotations of the first and second cam members 62 and 63, the
relative rotations of the front wheels Wf, Wf and the rear wheels
Wr, Wr are continued. However, the frictional clutch 67 is slipped
to permit the relative rotations of the clutch hub 42 and the
second cam member 63 and hence, an excessive load cannot be applied
to the torque cam mechanism 61.
[0063] A second embodiment of the present invention will now be
described with reference to FIG. 6.
[0064] In the structure of the first embodiment, after the torque
cam mechanism 61 has been operated to provide the state shown in
FIG. 4B, thereby inhibiting the relative rotations of the first and
second cam members 62 and 63, the rear-end clutch plate 43
continued to be rotated along with the clutch housing 41 and the
urging portion 66 of the second cam member 63 are slid on each
other. For this reason, there is a possibility that a friction
and/or a strange noise may be generated.
[0065] Therefore, in the second embodiment, a thrust bearing 70 is
disposed between the rear-end clutch plate 43 and the urging
portion 66 of the second cam member 63, whereby the generation of a
friction and/or abnormal noise can be prevented.
[0066] A third embodiment of the present invention will be
described below with reference to FIG. 7.
[0067] In a torque cam mechanism 61 in the third embodiment, a
second cam member 63 is axially movably and relatively
non-rotatably carried in the spline 42a of the clutch hub 42, while
a frictional clutch 67 is disposed between a spline 62a of a first
cam member 62 and the spline 41a of the clutch housing 41. The
frictional clutch 67 is designed, so that it is slipped, when a
torque greater than a predetermined value is applied between the
first cam member 62 and the clutch housing 41.
[0068] An end plate 71 is axially movably and relatively
non-rotatably carried in the rear of the torque cam mechanism 61
between the spline 41a of the clutch housing 41 and the sleeve 65
connected to the first hydraulic pump Pf. A thrust bearing 72 is
disposed between a front surface of the end plate 71 and a rear
surface of the first cam member 62, and a thrust bearing 73 is
disposed between a rear surface of the end plate 71 and a front
surface of the clutch piston 47.
[0069] With the above arrangement, when relative rotations occur
between the front wheels Wf, Wf and the rear wheels Wr, Wr to
operate the torque cam mechanism 61, the multi-plate clutch C is
brought into its engaged state by a thrust force f generated by the
relative rotations of the first and second cam members 62 and 63.
Even after the relative rotations of the first and second cam
members 62 and 63 have been limited, the relative rotations of the
front wheels Wf, Wf and the rear wheels Wr, Wr are continued, but
the frictional clutch 67 is slipped to permit the relative
rotations of the clutch housing 41 and the first cam member 62 and
to permit relative rotations of the end plate 71 rotated in unison
with the clutch housing 41 and the first cam member 62 rotated in
unison with the clutch hub 42 by the action of the thrust bearing
72. The thrust bearing 73 disposed between the clutch piston 47 and
the end plate 71 permits the rotation of the end plate 71 relative
to the clutch piston 47 which is not rotated relative to the
housing 45.
[0070] With the third embodiment, upon operation of the torque cam
mechanism 61, only the second cam member 63 is moved axially with
the first cam member 62 being in its axially stopped state, thereby
bringing the multi-plate clutch C into the engaged state.
Therefore, it is possible to prevent the frictional clutch 67
carried on the first cam member 62 from being axially moved to
contribute to the alleviation of the friction. The relative
rotations of the end plate 71 rotated in unison with the cutch
housing 41 and the first cam member 62 rotated in unison with the
clutch hub 42 are permitted by the action of the thrust bearing 72.
However, the rotational speed of the thrust bearing 72 in the third
embodiment can be reduced to enhance the durability, because the
thrust bearing 72 is mounted at a location radially inward from the
thrust bearing 70 (mounted at the tip end of the urging portion 66
of the second cam member 63 in FIG. 6) in the second embodiment,
namely, at a radially intermediate location on the end plate
71.
[0071] Although the embodiments of the present invention have been
described in detail, it will be understood that the present
invention is not limited to the above-described embodiments, and
various modifications may be made without departing from the
subject matter of the invention.
[0072] For example, in the various embodiments, the clutch housing
has been illustrated as a member rotated in operative association
with the front wheels Wf, Wf, and the clutch hub 42 has been
illustrated as a member rotated in operative association with the
rear wheels Wr, Wr. Alternatively, a member other than the clutch
housing 41 and the clutch hub 42 may be employed.
* * * * *