Automatic clutch device for hydrostatic continuously variable transmission

A hydrostatic continuously variable transmission is provided in which a hydraulic pump and a hydraulic motor are communicated to each other through a low pressure oil passage and a high pressure oil passage and a clutch valve capable of releasing an oil pressure of the high pressure oil passage is connected to the high pressure oil passage. The clutch valve can be automatically operated by making use of an oil pressure of the low pressure oil passage. A clutch valve is applied with an oil pressure of a low pressure oil passage for biasing the clutch valve in a valve opening direction, and an oil pressure governor is connected to a supply oil passage for supplying an oil pressure from an oil supply pump to the low pressure oil passage through a check valve for increasing an oil pressure of the oil passage in accordance with an increase in an input rotational speed of a hydraulic pump.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates to an automatic clutch device for a 
hydrostatic continuously variable transmission in which a hydraulic pump 
and a hydraulic motor are in communication with each other through a low 
pressure oil passage and a high pressure oil passage. An oil supply 
passage is connected to an oil supply pump and is connected to the low 
pressure oil passage and the high pressure oil passage through a first 
check valve and a second check valve, respectively. A path between the 
hydraulic pump and the hydraulic motor is automatically controlled into an 
oil pressure transmission state or an oil pressure transmission cut-off 
state. 
2. Description of Background Art 
As set forth in Japanese Patent Publication No. Sho 61-23414, an automatic 
clutch device is known, which includes a clutch valve and an actuator 
connected to the clutch valve for automatically operating the clutch 
valve. More specifically, the clutch valve is provided in a 
short-circuiting oil passage connecting a low pressure oil passage and a 
high pressure oil passage to each other, for opening/closing the 
short-circuiting oil passage. The actuator is operated by an oil pressure 
generated by an oil pressure governor interlocked with rotation of a motor 
of a hydraulic pump. 
The above automatic clutch device has a disadvantage in complicating and 
enlarging the structure and increasing the cost because it requires, in 
addition to a clutch valve, an actuator for operating the clutch valve. 
SUMMARY AND OBJECTS OF THE INVENTION 
An object of the present invention is to provide an automatic clutch device 
of the above-described hydrostatic continuously variable transmission, 
which is capable of solving the above-described disadvantage. 
To achieve the above object, according to a first feature of the present 
invention, there is provided an automatic clutch device for a hydrostatic 
continuously variable transmission, which includes a clutch valve, 
provided between the high pressure oil passage and the low pressure oil 
passage or a discharge port connected to an oil reservoir, for switching a 
valve position between a valve opening position in which the high pressure 
oil passage and the low pressure oil passage or the discharge port 
connected to the oil reservoir are communicated to each other and a valve 
closing position in which they are cut-off from each other. The clutch 
valve is applied with oil pressure from the low pressure oil passage for 
biasing the clutch valve in the valve closing direction. An oil pressure 
governor is provided for increasing the oil pressure of the low pressure 
oil passage applied to the clutch valve linearly with an input rotational 
speed of the hydraulic pump, the oil pressure governor is connected to the 
clutch valve. 
According to a second feature of the present invention, in addition to the 
first feature, the clutch valve is applied with oil pressure of the high 
pressure oil passage in the valve opening direction when the clutch valve 
is closed. In addition, the clutch valve is opened when the oil pressure 
of the high pressure oil passage is increased to exceed the oil pressure 
of the low pressure oil passage by a specified value or more. 
Further scope of applicability of the present invention will become 
apparent from the detailed description given hereinafter. However, it 
should be understood that the detailed description and specific examples, 
while indicating preferred embodiments of the invention, are given by way 
of illustration only, since various changes and modifications within the 
spirit and scope of the invention will become apparent to those skilled in 
the art from this detailed description.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
Hereinafter, a preferred embodiment of the present invention will be 
described with reference to the drawings. 
Referring to FIGS. 1 to 3, a power unit of a motorcycle includes a 
transmission case 1 composed of right and left case halves 1B, 1A 
connected to each other by means of a bolt 2. The transmission case 1 
contains a continuously variable transmission T including a constant 
displacement type swash plate hydraulic pump P and a variable displacement 
type swash plate hydraulic motor M. The hydraulic pump P and the hydraulic 
motor M constitute an oil pressure closing circuit. 
The hydraulic pump P includes an input cylindrical shaft 4 rotatably and 
axially movably supported on the outer end portion of the right case half 
1B through a radial bearing 3. A pump cylinder 6 is relatively rotatably 
supported on the input cylindrical shaft 4 through a first angular contact 
bearing 5. A plurality of pump plungers 8 are each slidably inserted in a 
plurality of the odd number of cylindrical holes 7 annularly arranged in 
the pump cylinder 6 so as to surround the rotational axis of the pump 
cylinder 6. A pump swash plate 9 is provided with a front surface brought 
into contact with the outer ends of the pump plungers 8 and a pump swash 
plate holder 12 for supporting the swash plate 9 through a thrust bearing 
10 and a radial bearing 11 for holding the pump swash plate 9 in a posture 
in which the pump swash plate 9 is tilted around a virtual trunnion axial 
line 0.sub.1 perpendicular to the axial line of the pump cylinder 6 at a 
specified angle with respect to the axial line of the pump cylinder 6. The 
pump swash holder 12 is formed integrally with the input cylindrical shaft 
4. 
The right end portion of the input cylindrical shaft 4 projects outwardly 
from the right case half 1B, and is fixed with an input gear 13 to which 
power from an engine (not shown) is to be inputted. 
When the input cylindrical shaft 4 is rotated, each pump plunger 8 is 
reciprocated with the aid of the pump swash plate 9 to repeat suction and 
discharge strokes. 
On the other hand, the hydraulic motor M is disposed coaxially with and 
leftward from the pump cylinder 6. The hydraulic motor M includes a motor 
cylinder 16 rotatably supported on the left case half 1A through a second 
angular contact bearing 15. A plurality of motor plungers 18 are each 
slidably inserted in a plurality of, the odd number of, cylinder holes 17 
annularly arranged in the motor cylinder 16 so as to surround the 
rotational axis of the motor cylinder 16. A motor swash plate 19 is 
provided with its front surface brought in contact with the outer ends of 
the motor plungers 18. A motor swash holder 22 supports the motor swash 
plate 19 through a thrust bearing 20 and a radial bearing 21. A motor 
swash plate anchor 23 supports the back surface of the motor swash plate 
holder 22. The motor swash plate anchor 23 is fixed on the outer end 
portion of the left case half 1A by means of a plurality of bolts 24. 
Each of opposite contact surfaces f.sub.1, f.sub.2 of the motor swash plate 
holder 22 and the motor swash plate anchor 23 is formed into a 
semi-cylindrical surface centered at a trunnion axial line 0.sub.2 
perpendicular to the axial line of the motor cylinder 16 for allowing the 
tilting of the motor swash plate holder 22 around the trunnion axial line 
0.sub.2. 
To restrict relative sliding in the direction of the trunnion axial line 
0.sub.2 between the motor swash plate holder 22 and the motor swash plate 
anchor 23, as shown in FIG. 3, a flange 22a which is brought into contact 
with one end surface of the motor swash plate anchor 23 is formed at one 
end of the motor swash plate holder 22. In addition, a speed change lever 
25 is brought into contact with the other surface of the motor swash plate 
anchor 23 and is fixed on the other end surface of the motor swash plate 
holder 22 by means of a bolt 26. Accordingly, the turning of the speed 
change lever 25 allows the motor swash plate holder 22 to be tilted around 
the trunnion axial line 0.sub.2 so that the tilting angle of the motor 
swash plate 19 is adjusted between an upright position where it is upright 
to the axial line of the motor cylinder 16 and the maximum tilting 
position where it is tiled at a specified angle. 
When the motor cylinder 16 is rotated into a state in which the motor swash 
plate 19 is tilted, each motor plunger 18 is reciprocated by the motor 
swash plate 19 to repeat expansion and contraction strokes. 
The pump cylinder 6 and the motor cylinder 16 are integrated with each 
other, to form a cylinder block B. An output shaft 27 is connected to the 
central portion of the cylinder block B in the rotational direction 
through a spline 28 in such a manner so as to pass through the cylinder 
block B. The output shaft 27 is also axially connected to the cylinder 
block B through a pair of right and left snap rings 30, 29. 
The output shaft 27 has the left end portion terminated in front of the 
motor swash plate 19 and the right end portion passing through the input 
cylindrical shaft 4 and projecting outwardly therefrom. The projecting end 
of the output shaft 27 is additionally provided with an output gear 31 for 
outputting power to a rear wheel of a motorcycle. At this time, the output 
shaft 27 is rotatably and axially movably supported on the input 
cylindrical shaft 4 through a radial bearing 32. 
The input cylindrical shaft 4 is divided into a funnel-shaped portion 4a 
integrated with the pump swash plate holder 12 and supported by the right 
case half 1B through the radial bearing 3, and a cylindrical portion 4b 
supporting the pump cylinder 6 through the first angular contact bearing 
5. Both the portions 4a, 4b are connected to each other by means of a bolt 
33. At this time, an outer race 5o of the first angular contact bearing 5, 
which is held between both the portions 4a, 4b, and an inner race 5i of 
the bearing 5, is fixed on the outer peripheral surface of the pump 
cylinder 6 by means of a sleeve 34 (described later) and a snap ring 35. 
The input cylindrical shaft 4 and the pump cylinder 6 are thus axially 
fixedly connected to the first angular contact bearing 5. 
The second angular contact bearing 15 is disposed radially inwardly from 
and in the vicinity of a plurality of bolts 24 for connecting the motor 
swash plate anchor 23 to the outer end portion of the left case half 1A. 
An outer race 15o of the bearing 15 is fixed on the left case half 1A by 
means of a pressing plate 36 held between the left case half 1A and the 
motor swash plate anchor 23, and an inner race 15i of the bearing 15 is 
fixed on the outer peripheral surface of the motor cylinder 16 by means of 
the sleeve 34 (described later) and a snap ring 47. The motor cylinder 16 
and the motor swash plate anchor 23 are thus axially fixedly connected to 
the second angular contact bearing 15. 
To rotate the pump swash plate 9 in synchronization with the pump cylinder 
6, spherical recesses 9a each engaged with spherical end portions 8a of 
the pump plungers 8 are formed in the front surface of the pump swash 
plate 9. 
To rotate the motor swash plate 19 in synchronization with the motor 
cylinder 16, spherical recesses 19a each engaged with spherical end 
portions 18a of the motor plungers 18 are formed in the front surface of 
the motor swash plate 19. 
The spherical recesses 9a, 19a are larger in radius than the corresponding 
spherical end portions 8a, 18a so as to be kept in engagement with the 
corresponding spherical end portions 8a, 18a even in any operating 
position. 
As shown in FIGS. 2, 3 and 6, an annular inner oil passage 50 (low pressure 
oil passage) and an annular outer oil passage 51 (high pressure oil 
passage) are formed between the pump cylinder 6 and the motor cylinder 16. 
More specifically, the oil passage 50 is positioned between the inner 
periphery of the cylinder block B and the outer periphery of the output 
shaft 27, and the oil passage 51 is positioned between the outer periphery 
of the cylinder block B and the sleeve 34 fitted and brazed on the outer 
peripheral surface of the block B. The annular bulkhead and the sleeve 34 
between both the oil passages 50, 51 have a group of first valve holes 52 
and a group of second valve holes 53 which are formed in such a manner so 
as to radially pass through the bulkhead and the sleeve 34 and to be 
adjacent to a group of the cylinder holes 7 and a group of the cylinder 
holes 17, respectively. The adjacent first valve hole 52 and the cylinder 
hole 7 are communicated to each other through a pump port "a" and the 
adjacent second valve hole 53 and the cylinder hole 17 are communicated to 
each other through a motor port "b." A spool type first distribution valve 
55 is inserted in each first valve hole 52 and a spool type second 
distribution valve 56 is inserted in each second valve hole 53. 
As shown in FIG. 4, a first eccentric ring 57 is disposed around the outer 
peripheries of a group of the first distribution valves 55 in such a 
manner so as to be engaged with the outer ends thereof. A first forcing 
ring 58, coaxial with the first eccentric ring 57, is connected to the 
first distribution valves 55 by means of clips 59 for usually ensuring 
engagement between the first distribution valves 55 and the first 
eccentric ring 57. 
The first eccentric ring 57 is formed of a ball bearing, and is mounted in 
the input cylindrical shaft 4 in such a manner so as to be held at a 
position being eccentric along the virtual trunnion axial line 0.sub.1 
from the center of the output shaft 27 by a specified distance 
.epsilon..sub.1. 
When a relative rotation is generated between the input cylindrical shaft 4 
and the pump cylinder 16, each first distribution valve 55 is reciprocated 
in each first valve hole 52 by means of the first eccentric ring 57 by a 
stroke equivalent to a distance twice the eccentric amount 
.epsilon..sub.1. In a discharge region D of the hydraulic pump P, the 
first distribution valve 55 is moved to the inner end side of the first 
valve hole 52 to allow all of the corresponding pump ports to be 
communicated to the outer oil passage 51 and not to be communicated to the 
inner oil passage 50, with a result that the pump plunger 8 in a discharge 
stroke acts to press-feed a working oil from the cylinder hole 7 into the 
outer oil passage 51. Also, in a suction region S of the hydraulic pump P, 
the first distribution valve 55 is moved to the outer end side of the 
first valve hole 53 to allow the corresponding pump port "a" to be 
communicated to the inner oil passage 50 and not to be communicated to the 
outer oil passage 51, with a result that the pump plunger 8 in the suction 
stroke acts to suck a working oil from the inner oil passage 50 into the 
cylinder hole 7. 
As shown in FIG. 5, a second eccentric ring 60 is disposed around the outer 
peripheries of a group of the second distribution valves 56 in such a 
manner so as to be engaged with the outer ends thereof. A second forcing 
ring 61, coaxial with the second eccentric ring 60, is connected to the 
second distribution valves 56 by means of clips 62 for usually ensuring 
engagement between the second distribution valves 56 and the second 
eccentric ring 60. 
The second eccentric ring 60 is formed of a ball bearing, and is mounted in 
the left case half 1A so as to be held at a position being eccentric along 
the trunnion axial line O.sub.2 from the center of the output shaft 27 by 
a specified distance .epsilon..sub.2. 
Accordingly, when the motor cylinder 16 is rotated, each second 
distribution valve 56 is reciprocated in each second valve hole 53 by 
means of the second eccentric ring 60 by a stroke equivalent to a distance 
twice the eccentric amount .epsilon..sub.2. In an expansion region E of 
the hydraulic motor M, the second distribution valve 56 is moved to the 
inner end side of the second valve hole 53 to allow the corresponding 
motor port "b" to be communicated to the outer oil passage 51 and not to 
be communicated to the inner oil passage 50, thereby supplying a high 
pressure working oil from the outer oil passage 51 into the cylinder hole 
17 in which the motor plunger 18 is in the expansion stroke. Also, in a 
contraction region R of the hydraulic motor M, the second distribution 
valve 56 is moved to the outer end side of the second valve hole 53 to 
allow the corresponding motor port "b" to be communicated to the inner oil 
passage 50 and not to be communicated to the outer oil passage 51, thereby 
returning a working oil into the inner oil passage 50 from the cylinder 
hole 17 in which the motor plunger 18 is in the contraction stroke. 
The cylinder block B is thus rotated by the total of a reaction torque 
received by the pump swash plate 9 through the pump plungers 8 in the 
discharge stroke and a reaction torque received by the motor swash plate 
19 through the motor plungers 18 in the expansion stroke, and the 
rotational torque thereof is transmitted to the output shaft 27. 
In this case, a speed change ratio of the output shaft 27 to the input 
cylindrical shaft 4 is given by the following equation: 
EQU speed change ratio=1+(capacity of hydraulic motor M)/(capacity of hydraulic 
pump P) (1) 
Accordingly, the speed change ratio can be changed from 1 to a necessary 
value by changing the capacity of the hydraulic motor M from zero to a 
specified value. Also, the speed change ratio can be continuously 
controlled by tilting the motor swash plate 19 from the upright state to a 
specified tilting position because the capacity of the hydraulic motor P 
is determined on the basis of the stroke of the motor plungers 18. 
Incidentally, a thrust load acting to axially separate the input 
cylindrical shaft 4 from the pump cylinder 6 is generated therebetween by 
a pressing action of the pump plungers 8 to the pump swash plate 9 during 
operation of the hydraulic pump P. However, since the input cylindrical 
shaft 4 and the pump cylinder 6 are axially connected to each other 
through the first angular contact bearing 5, such a thrust load is 
absorbed by the bearing 5, to thereby prevent the load from being applied 
to the transmission case 1 and the output shaft 27. 
A thrust load acting to axially separate the motor cylinder 16 from the 
motor swash plate 23 is generated therebetween by a pressing action of the 
motor swash plate 19 to the motor plungers 18 during operation of the 
hydraulic motor M. However, since the motor cylinder 16 and the motor 
swash plate 23 are axially connected to each other through the second 
angular contact bearing 15 and the outer end portion of the left case half 
1A, the load is absorbed by the bearing 15 and the outer end portion of 
the left case half 1A to thereby prevent the load from being applied to 
the output shaft 27. 
In this case, since the second angular contact bearing 15 is disposed 
radially inwardly of and in the vicinity of a plurality of the bolts 24 
connecting the motor swash plate anchor 23 to the outer end portion of the 
left case half 1A, it becomes possible to make narrower a thrust load 
acting region of the left case half 1A to the utmost while suppressing an 
increase in the axial dimension of the transmission T. As a result, the 
transmission case 1 can be ensured in durability only by increasing the 
thickness of a portion applied with a thrust load of the left case half 
1A. 
Moreover, since the pump cylinder 6 and the motor cylinder 16 are 
integrated with each other to form the cylinder block B and the input 
cylindrical shaft 4 is axially movably supported on an outer end portion 
of the right case half 1B through the radial bearing 3, it becomes 
possible to positively prevent a thrust load generated between the input 
cylindrical shaft 4 and the motor swash plate anchor 23 from being applied 
between both the case halves 1A, 1B and to reduce the thickness and the 
weight of the transmission case 1. 
Additionally, since the output shaft 27 has the right end as the output end 
projecting outwardly from the input cylindrical shaft 4 and the left end 
terminated in front of the motor swash plate 19, the transmission T is 
reduced in its axial dimension and also the tilting range, that is, the 
speed change range of the motor swash plate 19 can be extended without 
interference with the output shaft 27. In addition, an oil introducing 
pipe 66 (described later) passes through the central portions of the motor 
swash plate 19 and the motor swash plate holder 22. However, the oil 
introducing pipe 66 does not obstruct the tilting of the motor swash plate 
19 and the like because the size thereof is very small relative to the 
size of the output shaft 27. 
Referring to FIGS. 1, 2 and 8, the output shaft 27 has at its central 
portion an oil supply hole 65 extending from the hydraulic motor M side. 
The oil introducing pipe 66 extending from the motor swash plate anchor 23 
and passing through the central portions of the motor swash plate holder 
22 and the motor swash plate 19 is relatively rotatably inserted in the 
inlet of the oil supply hole 65 through a bushing 67. The oil introducing 
pipe 66 is communicated to a discharge port of an oil supply pump 69 
driven by the input cylindrical shaft 4 through a gear row 68, by way of a 
series of oil passages 70 formed in the transmission case 1 and the motor 
swash plate anchor 23. The oil supply pump 69 operates to pump a working 
oil from an oil passage 71 on the bottom of the transmission case 1 and to 
supply it to the oil supply hole 65 through an oil passage 70 and the oil 
introducing pipe 66. These oil supply hole 65, oil introducing pipe 66 and 
the oil passage 70 constitute an oil supply passage L. 
The oil supply hole 65 is connected to the inner and outer oil passages 50, 
51 through first and second branched passages 72, 73 formed in the output 
shaft 27 and the cylinder block B, respectively. First and second check 
valves 74, 75 are provided in the first and second branched passages 72, 
73, respectively. Accordingly, upon normal operation, when the inner oil 
passage 50 is reduced in pressure due to oil leakage from the hydraulic 
pump P and the hydraulic motor M, the first check valve 74 is opened to 
supply a working oil from the supply oil passage L into the inner oil 
passage 50. Upon braking, when the outer oil passage 51 is reduced in 
pressure, the second check valve 75 is opened to supply a working oil from 
the supply oil passage L into the outer oil passage 51. 
The peripheral wall of the oil introducing pipe 66 has a plurality of blow 
holes 76 which allow a working oil to be blown for lubricating 
surroundings of the motor swash plate 19. The output shaft 27 has a blow 
hole 77 in communication with the oil supply hole 65 which allows a 
working oil to be blown for lubricating the surroundings of the pump swash 
plate 9. 
Referring to FIGS. 1, 7 and 8, the continuously variable transmission T 
includes a clutch valve 78 for cutting-off/re-starting the oil pressure 
transmission of the hydraulic pump P and the hydraulic motor M. A 
hydraulic servomotor 99 for operating the speed change lever 25 and an oil 
pressure governor 80 and a throttle valve 81 for automatically controlling 
the clutch valve 78 and the hydraulic servomotor 99. 
The hydraulic governor 80 has a rocking cylinder 84 supported through a 
pivot 83 by a rotational shaft 82 of the oil supply pump 69 driven by the 
input cylindrical shaft 4 through a gear row 68 in such a manner as to 
surround the rotational shaft 82. The rocking cylinder 84 has a weight 
portion 84a at one rocking end so as to be rocked on the weight portion 
84a side when the centrifugal force of the weight portion 84a is increased 
linearly with the rotational speed of the rotational shaft 82. The 
rotational shaft 82 has a governor oil passage 86 communicated to the oil 
passage 70 connected to the discharge port of the oil supply port 69 
through an orifice 85 and a conical valve seat 87 to open the oil passage 
86 on the outer side surface of the rotational shaft 82 opposite to the 
weight portion 84a. A ball-like valve body 88 for opening/closing the oil 
passage 86 in co-operation with the valve seat 87 is contained in the 
rocking cylinder 84. The oil passage 86 is connected to a governor oil 
pressure chamber 89 of the throttle valve 81. 
Accordingly, when a rotational speed of the rotational shaft 82, that is, 
the rotational speed of the engine driving the input cylindrical shaft 4 
is relatively low and thereby the centrifugal force of the weight portion 
84a is relatively small, the seating force of the valve body 88 to the 
valve seat 87 due to the centrifugal force is relatively weak so that oil 
pressure is released from the valve seat 87. However, when the engine 
speed becomes higher, the seating force of the valve body 88 to the valve 
seat 87 due to the centrifugal force of the weight portion 84a is 
increased to suppress the release of the oil pressure. As a result, an oil 
pressure PG in the governor oil passage 86 and the governor oil pressure 
89 downstream from the orifice 85 is increased linearly with the engine 
speed, as shown in FIG. 10. 
The throttle valve 81 has a valve hole 91 of a valve body 90 mounted on the 
transmission case 1. A discharge port 97 is opened in an intermediate 
portion of the valve hole 91 for opening the oil passage 70 to an oil 
reservoir 71. A cylindrical valve body 92 is slidably inserted in the 
valve hole 91 for opening/closing the discharge port 97, a return spring 
93 is provided for biasing the valve body 92 in the direction of closing 
the discharge port 97. An operating lever 95 supports the base portion of 
the operating spring 94. A throttle lever 96 is capable of pushing the 
operating lever 95. The throttle lever 96 is interlocked with the 
opening/closing of a throttle valve (not shown) of the engine. More 
specifically, when the throttle valve of the engine is opened by a 
specified opening degree or more (2/8, in the figure), the throttle lever 
96 pushes a throttle bar 95 on the basis of increasing the opening degree 
to thereby increase a load applied to the operating spring 94. 
The set load of the return spring 93 is set to be smaller than that of the 
operating spring 94 so that the return spring 93 holds the valve body 92 
at the position where the discharge port 97 is closed when the operating 
lever 95 is retracted. The end surface, on the oil passage 70 side, of the 
valve body 92 usually faces to the oil passage 70 to receive an oil 
pressure of the oil passage 70. The above governor oil pressure chamber 89 
to which the end surface of the valve body 92 faces on the side opposite 
to the operating spring 94, is defined in the valve hole 91. 
Accordingly, the valve body 92 is operated by a balancing action between a 
load of the operating spring 93 applied to press the valve body 92 in the 
direction of closing the discharge port 97 and an oil pressure of the 
governor oil pressure chamber 89. A load of the operating spring 93 is 
applied to press the valve body 92 in the direction of opening the 
discharge port 97 and an oil pressure of the oil passage 70. Accordingly, 
when the oil supply pump 69 is first driven by the input cylindrical shaft 
4 upon the starting of the engine to generate an oil pressure in the oil 
passage 70, the valve body 93 is pushed in the direction of opening the 
discharge port 97 by the oil pressure of the oil passage 70 and is 
simultaneously pushed in the direction of closing the discharge port 97 by 
an oil pressure P.sub.G of the governor oil pressure chamber 89, to thus 
start the adjustment of the oil pressure P.sub.F of the oil supply passage 
L. 
Incidentally, since the operating lever 95 does not push the operating 
spring 94 until the throttle valve of the engine is opened by the opening 
degree of 2/8, the load of the operating spring 94 is held at a minimum 
and thereby the opening of the discharge port 97 by the valve body 92 is 
the smallest, so that the oil pressure P.sub.F of the supply oil passage L 
is controlled to be relatively higher as shown by a dotted line positioned 
uppermost in FIG. 10. On the other hand, when the throttle valve of the 
engine is opened in the opening degree (2/8) or more, a throttle lever 96 
pushes the operating lever 95 to increase a load applied to the operating 
spring 94, thereby increasing the opening of the discharge port 97 by the 
valve body 92. As a result, the oil pressure P.sub.F of the supply oil 
passage L is controlled to be reduced, that is, it is shifted from the 
upper one to the lower one in a plurality of the dotted lines in FIG. 10. 
On the other hand, in each opening degree of the throttle valve of the 
engine, when the engine speed is increased, the oil pressure P.sub.G of 
the governor oil pressure chamber 89 is increased by the action of the oil 
pressure governor 80, thereby increasing the closing of the discharge port 
97 by the valve body 92. As a result, an oil pressure of the supply oil 
passage L is controlled to be increased. 
The oil pressure P.sub.F of the supply oil passage L thus controlled is 
supplied into the inner oil passage 50 through the first check valve 74, 
so that the oil pressure of the inner oil passage 50 becomes substantially 
the same as that of the supply oil passage L. 
Referring to FIG. 7, the clutch valve 78 has a valve piston 99 and a plug 
body 100. The valve piston 99 is slidably inserted in a cylindrical 
mounting hole 98 formed in the cylinder block B in a range from the inner 
and outer oil passages 50, 51 and faces at one end to the inner oil 
passage 50, and the plug body 100 is fixed in the mounting hole 98 
opposite to the other end of the valve piston 99. An oil chamber 102 
opened in the oil reservoir 71 through the discharge port 101 is defined 
between the valve piston 99 and the plug body 100. A return spring 103 for 
biasing the valve piston 99 on the inner oil passage 50 side is contained 
in the oil chamber 102. 
The plug body 100 includes a through-hole 104 communicated to the outer oil 
passage 51, and a conical valve seat 105 for opening the through-hole 104 
to the oil chamber 102. A valve body 106 which is smaller in diameter 
relative to the valve piston 99 is additionally provided on the valve 
piston 99 through a retainer 107. 
Accordingly, upon idling of the engine, the inner oil passage 50 is 
controlled to be relative low in pressure by a pressure reducing action of 
the oil pressure governor 80, so that the valve piston 99 is pushed onto 
the inner oil passage 50 side by the return spring 103 to release the 
valve body 106 from the valve seat 105, thereby releasing the oil pressure 
of the outer oil passage 51 to the discharge port 101 through the 
through-hole 104 and the oil chamber 102. The hydraulic pump P and the 
hydraulic motor M are thus positioned in an oil pressure transmission 
cutting-off state, that is, in a clutch-off state. 
When the engine speed is increased, the oil pressure of the inner oil 
passage 50 is increased by a pressure increasing action of the oil 
pressure governor 80, and the valve piston 99 allows the valve body 106 to 
be gradually moved toward the valve seat 105 against a force of the return 
spring 103 and to be finally seated on the valve seat 105, thereby 
restricting and stopping the release of the oil pressure from the outer 
oil passage 51. As a result, the hydraulic pump P and the hydraulic motor 
M are shifted to an oil pressure transmission state, that is, clutch-on 
state through a semi-clutch state, thus allowing the vehicle to be 
smoothly started. 
In such a clutch-on state, a high oil pressure of the outer oil passage 51 
is usually applied to part of the spherical valve body 106 which is very 
small in diameter relative to the valve piston 99, and accordingly if an 
excessive load is applied to the hydraulic motor M and an excessive oil 
pressure is generated in the outer oil passage 51, a pressing force of the 
excessive oil pressure of the outer oil passage 51 against the valve body 
106 exceeds a pressing force of the oil pressure of the inner oil passage 
50 against the valve piston 99, to thereby open the valve body 106. As a 
result, the excessive oil pressure of the outer oil passage 51 is 
discharged in the discharge port 101, to avoid the excessive load. In this 
way, the clutch valve 78 has a function of an oil pressure limiter for 
protecting the hydraulic motor M from excessive oil pressure. 
The hydraulic servomotor 79 has a hydraulic cylinder 110 and a control 
valve 111, as shown in FIGS. 1 and 8. The hydraulic cylinder 110 includes 
a cylinder body 102 formed integrally with the transmission case 1, an 
operating piston 115 for defining the interior of the cylinder body 102 
into right and left oil chambers 113, 114, and a return spring 116 for 
biasing the operating piston 115 to the right oil chamber 114 side. The 
operating piston 115 has a rod 115a passing through the left oil chamber 
113 and projecting outwardly from the cylinder body 112, and is connected 
at the leading end with the speed change lever 25. 
The speed change lever 25 is held at a position, low position, where the 
motor swash plate 19 is tiled at a maximum at the rightward movement limit 
of the operating piston 115 where the left oil chamber is extended at a 
maximum. When the operating piston 115 is moved leftward from the low 
position, the speed change lever 25 can be operated in the direction, top 
direction, where the motor swash plate 19 is raised. 
The control valve 111 includes a valve cylinder 117 formed integrally with 
the valve body 90, and a spool type valve body 118 contained in the valve 
cylinder 117. The valve cylinder 117 contains an operating oil chamber 119 
to which the left end surface of the valve body 118 faces, and an 
atmospheric air chamber 120 to which the right end surface of the valve 
body 118 faces. The operating oil chamber 119 receives oil pressure from 
the oil passage 70 through an orifice 121, and the atmospheric air chamber 
120 contains a return spring 122 for biasing the valve body 108 to the 
operating oil chamber side 119. 
A first discharge port 123.sub.1, an input port 124 and a second discharge 
port 123.sub.2 are provided in one side of the valve cylinder 117 from the 
left side as illustrated in FIG. 8, in this order, and a first output port 
125.sub.1, and a second output port 125.sub.2 are provided in the other 
side of the valve cylinder 117. 
The input port 124 is opened to the oil passage 70, and the first and 
second discharge ports 123.sub.1, 123.sub.2 are opened to the oil 
reservoir 71. 
On the other hand, the first output port 125.sub.1, is connected to the 
right oil chamber 114 of the hydraulic cylinder 120, and the second output 
port 125.sub.2 is connected to the left oil chamber 113 of the hydraulic 
cylinder 120. 
Accordingly, when the oil pressure of the operating oil chamber 119, that 
is, the oil pressure of the oil passage 70 is relatively low, the valve 
body 118 is held at the leftward movement limit so that the input port 124 
is communicated to the second output port 125.sub.2 and the first output 
port 125.sub.1, is communicated to the first discharge port 123.sub.1. 
Thus, an oil pressure of the oil passage 70 is supplied to the left oil 
chamber 113 of the hydraulic cylinder 110 to move the operating piston 115 
rightwardly, thereby holding the speed change holder 25 at the low 
position. When the valve body 118 is moved rightwardly with an increase in 
pressure of the operating oil chamber 119, the input port 124 is 
communicated to the first output port 125.sub.1, and the second output 
port 125.sub.2 is communicated to the second discharge port 123.sub.2. 
Thus, an oil pressure of the oil passage 70 is supplied to the right oil 
chamber 114 of the hydraulic cylinder 110 to move the operating piston 115 
leftwardly, thus operating the speed change lever 25 on the top side. 
Incidentally, the oil pressure introduced from the oil passage 70 to the 
operating oil chamber 119 of the control valve 118 is controlled by the 
oil pressure governor 79 and the throttle valve 81 as described above, and 
consequently the control valve 101 is automatically controlled on the 
basis of input signals such as an engine speed and an opening degree of 
the throttle valve. More specifically, when the engine speed is increased, 
the operating piston 115 tends to be operated onto the top side, and when 
the opening degree of the throttle valve is increased, the operating 
piston 115 tends to be operated onto the low side. 
FIG. 9 shows a modification of the clutch valve 78, which includes a valve 
cylinder 130 inserted in a mounting hole 98 of the cylinder block B, a 
spool valve body 121 is slidably inserted in the valve cylinder 130, and a 
cup-like plug body 132 is fixedly fitted in the valve cylinder 130. The 
valve cylinder 130 has a plurality of horizontal holes connected to the 
outer oil passage 51, and the spool valve body 131 has on the outer 
periphery an annular groove 134 connected to the inner oil passage 50. The 
spool valve body 131 is biased on the inner oil passage 50 side by means 
of a return spring 135 contained between the valve body 131 and the plug 
body 132. 
Accordingly, when the oil pressure of the inner oil passage 50 is 
relatively low, the spool valve body 131 is pushed to the inner oil 
passage 50 side by means of the return spring 125 to allow the annular 
groove 134 and the horizontal holes 123 to be communicated to each other 
for short-circuiting both the inner and outer oil passages 50, 51 to each 
other, with a result that the hydraulic pump P and the hydraulic motor M 
are positioned in a clutch-off state. On the other hand, when the oil 
pressure of the inner oil passage 50 is increased, the spool valve body 
131 is pushed to the plug body 132 side by the oil pressure of the inner 
oil passage 50 to allow the annular groove 134 and the horizontal holes 
133 not to be communicated to each other for preventing short-circuiting 
between both the oil passages 50, 51, with a result that the hydraulic 
pump P and the hydraulic motor M are positioned in a clutch-on state. 
The present invention is not limited to the above-described embodiment and 
many changes in design can be made without departing from the scope of the 
present invention. 
For example, in the clutch valve 78 shown in FIG. 7, a through-hole 
communicating the oil chamber 102 to the inner oil passage 50 may be 
provided in the piston valve 99 in place of the discharge port 101 for 
short-circuiting the inner oil passage 50 and the outer oil passage 51 to 
each other upon opening the valve body 106. 
According to the first feature of the present invention, there is provided 
an automatic clutch device for a hydrostatic continuously variable 
transmission, which includes a clutch valve, provided between the high 
pressure oil passage and the low pressure oil passage or a discharge port 
connected to an oil reservoir, for switching a valve position between a 
valve opening position in which the high pressure oil passage and the low 
pressure oil passage or the discharge port connected to the oil reservoir 
are communicated to each other and a valve closing position in which they 
are cut-off from each other. The clutch valve is applied with an oil 
pressure of the low pressure oil passage for biasing the clutch valve in 
the valve closing direction. An oil pressure governor, for increasing the 
oil pressure of the low pressure oil passage, is applied to the clutch 
valve linearly with an input rotational speed of the hydraulic pump and 
the oil pressure governor being connected to the clutch valve. 
Accordingly, the clutch valve can be automatically operated in accordance 
with an input rotational speed of the hydraulic pump by making direct use 
of an oil pressure of the inner oil passage. This is advantageous in 
eliminating any special actuator and in simplifying and miniaturizing the 
structure, resulting in a significantly reduced cost. 
According to the second feature of the present invention, the clutch valve 
is applied with an oil pressure of the high pressure oil passage in the 
valve opening direction when the clutch valve is closed. The clutch valve 
is opened when the oil pressure of the high pressure oil passage is 
increased to exceed the oil pressure of the low pressure oil passage by a 
specified value or more. As a result, it becomes possible to release an 
excessive oil pressure from the high pressure oil passage by the clutch 
valve upon an excessive load of the hydraulic pump. This is advantageous 
in relaxing a torque shock and in protecting the transmission system and 
improving riding comfortability. 
The invention being thus described, it will be obvious that the same may be 
varied in many ways. Such variations are not to be regarded as a departure 
from the spirit and scope of the invention, and all such modifications as 
would be obvious to one skilled in the art are intended to be included 
within the scope of the following claims.