Hydrostatic continuously variable transmission

In a hydrostatic continuously variable transmission, a pump cylinder of a swashplate type hydraulic pump is integrally and coaxially coupled with a motor cylinder of a swashplate type hydraulic motor. A number of distributor valves are radially arranged for reciprocal movement between radially inner and outer positions to put a large number of pump ports in the hydraulic pump into communication with the higher and lower pressure oil passages, and an eccentric ring circumscribed about the distributor valves is connected to an input member of the hydraulic pump for movement between a clutch OFF position and a clutch ON position. A clutch spring is connected to the eccentric ring for biasing the eccentric ring toward the clutch OFF position, and a weight is added to the eccentric ring and adapted to exhibit a centrifugal force for moving the eccentric ring toward the clutch ON position against a spring force of the clutch spring.

BACKGROUND OF THE INVENTION 
1. FIELD OF THE INVENTION 
The present invention relates to a hydrostatic continuously variable 
transmission comprising a swashplate type hydraulic pump having a pump 
cylinder which is integrally and coaxially coupled with a motor cylinder 
of a swashplate type hydraulic motor, an annular higher pressure oil 
passage communicating with motor ports located in an expansion stroke 
region of the hydraulic motor, an annular lower pressure oil passage 
communicating with motor ports located in a shrinkage stroke region of the 
hydraulic motor, both the higher and lower pressure oil passages being 
concentrically formed between the pump and motor cylinders, a large number 
of radially arranged distributor valves reciprocally movable between 
radially inner and outer positions to put a large number of pump ports in 
the hydraulic pump into alternate communication with the higher and lower 
pressure oil passages, and an eccentric ring circumscribed about the 
distributor valves and connected to an input member of the hydraulic pump 
for movement between a clutch OFF position and a clutch ON position, so 
that the distributor valves are controlled to put the pump ports located 
in a discharge stroke region of the hydraulic pump into communication with 
the higher pressure oil passage and the pump ports located in an intake 
stroke region into communication with the lower pressure oil passage in a 
first eccentric position of the eccentric ring and to bring the hydraulic 
pump into a short-circuited state in a second eccentric position of the 
eccentric ring. 
2. DESCRIPTION OF THE PRIOR ART 
The present applicant has already proposed such hydrostatic continuously 
variable transmission, as disclosed in Japanese patent Application 
Laid-open No. 224769/87. 
In the hydrostatic continuously variable transmission already proposed, an 
external operating member is connected to the eccentric ring through a cam 
mechanism, so that the operating member is manually operated to control 
the position of the eccentric ring. Therefore, the structure is 
complicated and the operation is troublesome. 
SUMMARY OF THE INVENTION 
The present invention has been accomplished with such circumstances in 
view, and it is an object of the present invention to provide a 
hydrostatic continuously variable transmission which is simple in 
structure and in which the position of the eccentric ring can be 
automatically controlled in accordance with the speed of rotation of a 
cylindrical input member. 
To attain the above object, a feature of the present invention is that a 
clutch spring is connected to the eccentric ring for biasing the eccentric 
ring toward the clutch OFF position, and a weight is added to the 
eccentric ring and adapted to exhibit a centrifugal force for moving the 
eccentric ring toward the clutch ON position against a spring force of the 
clutch spring. 
With the above construction, when the input member is rotated, the 
eccentric ring is also rotated together with the input member, and a 
centrifugal force is produced in the weight of the eccentric ring in the 
opposite direction from the direction of the spring force applied to the 
eccentric ring by the clutch spring. 
Thus, when the speed of rotation of the input member is relatively low, 
such centrifugal force cannot overcome the spring force of the clutch 
spring and hence, the eccentric ring is maintained at the clutch OFF 
position by the action of the clutch spring, thereby controlling the 
distributor valves to bring the hydraulic pump into a short-circuited 
state. 
When the speed of rotation of the input member is increased to a level 
exceeding a predetermined value, such centrifugal force overcomes the 
spring force of the clutch spring to move the eccentric ring to the clutch 
ON position, thereby controlling the distributor valves to start 
transmitting the hydraulic pressure from the hydraulic pump to the 
hydraulic motor. 
In this way, the control of the position of the eccentric ring can be 
automatically performed in accordance with the speed of rotation of the 
input member by the centrifugal force provided by the weight, so that the 
states of interception and transmission between the hydraulic pump and the 
hydraulic motor can be properly obtained irrespective of whether or not 
the operator is skilled. Moreover, since the weight is added to the 
eccentric ring, there is no need for a special interlocking mechanism for 
transmitting the centrifugal force of the weight to the eccentric ring, 
leading to an extremely simple structure. 
The above and other objects, features and advantages of the invention will 
become apparent from a reading of the following description of the 
preferred embodiment, taken in conjunction with the accompanying drawings.

DESCRIPTION OF THE PREFERRED EMBODIMENTS 
One embodiment, of the present invention will now be described with 
reference to the accompanying drawings. Referring first to FIG. 1, a power 
unit U for a motorcycle comprises an engine E and a hydrostatic 
continuously variable transmission T. A crank shaft 1 of the engine E and 
the continuously variable transmission T are contained in a common casing 
4. The continuously variable transmission T includes a cylindrical input 
shaft 5 and an output shaft 15 disposed in parallel to the crank shaft 1, 
so that the crank shaft 1 drives the cylindrical input shaft 5 through a 
primary speed-reducing device 2, and the output shaft 15 drives a rear 
wheel (not shown) of the motorcycle through a secondary speed reducing 
device 3. 
Referring to FIGS. 1 and 2, the continuously variable transmission T 
comprises a swashplate type constant volume hydraulic pump p and a 
swashplate type variable volume hydraulic motor M. 
The hydraulic pump P is comprised of the cylindrical input shaft 5 as an 
input member, a pump cylinder 7 relatively rotatably fitted in an inner 
peripheral wall of the cylindrical input shaft 5 through a ball bearing 6, 
pump plungers 9, 9--slidable in a large odd number of annularly arranged 
cylinder bores 8, 8 --provided in the pump cylinder 7 to surround an axis 
of the pump cylinder 7, a pump swashplate 10 having a front face abutting 
against outer ends of the pump plungers 9, and a pump swashplate holder 12 
for supporting a back face of the pump swashplate 10 through an angular 
contact bearing 11 to hold the pump swashplate 10 at a state inclined at a 
given angle with respect to the axis of the pump cylinder 7 about a 
phantom trunnion axis O.sub.1 perpendicular to the axis of the pump 
cylinder 7. The pump swashplate holder 12 is fitted in the inner wall of 
the cylindrical input shaft. A driven gear 2b of the primary speed 
reducing device 2 is attached to an outer end of the cylindrical input 
shaft 5 through a torque damper. 
The pump swashplate 10 is adapted to sequentially provide reciprocations of 
the pump plungers 9, 9--upon rotation of the cylindrical input shaft 5 to 
repeat intake and discharge strokes thereof. 
On the other hand, the hydraulic motor M is comprised of a motor cylinder 
17 disposed coaxially with and laterally of the pump cylinder 7, motor 
plungers 19, 19--slidable in the same number of annularly arranged 
cylinder bores 18, 18 --as the cylinder bores 8, 8--, provided in the 
motor cylinder 17 to surround an axis of the motor cylinder 17, a motor 
swashplate 20 having a front face abutting against outer ends of the motor 
plungers 19, 19--, a motor swashplate holder 22 for supporting a back face 
of the motor swashplate 20 through a tapered roller bearing 21, and a 
motor swashplate anchor 23 for supporting a back face of the motor 
swashplate holder 22. The motor swashplate anchor 23 is secured to the 
casing by a bolt 27. 
Each of the cylinder bores 18 is formed larger in diameter than that of the 
cylinder bore 8, so that the maximum volume of the hydraulic motor M may 
be set sufficiently larger than that of the hydraulic pump P, thereby 
providing a large speed-reduction ratio. 
The motor swashplate holder 22 is provided at opposite ends with a pair of 
hemi-cylindrical trunnion shafts 22a and 22b disposed on a trunnion axis 
O.sub.2 perpendicular to the axis of the motor cylinder 17 and rotatably 
fitted respectively in a bearing hole 23a and a bearing recess 23b 
provided in the motor swashplate anchor 23. Respective abutting, opposed 
surfaces f.sub.1 and f.sub.2 of the motor swashplate holder 22 and the 
motor swashplate anchor 23 are each formed into a spherical surface having 
a center provided by an intersection of the axis of the motor cylinder 17 
with the trunnion axis O.sub.2. Therefore, the motor swashplate holder 22 
is rotatable about the trunnion axis O.sub.2 while receiving an aligning 
action from the motor swashplate anchor 23. 
A cylindrical cylinder holder 24 is integrally connected to the motor 
swashplate anchor 23 to extend rightward, and the motor cylinder 17 is 
rotatably supported at its outer peripheral surface by the cylinder holder 
through a ball bearing 25. 
As shown in FIGS. 8 to 10, an operating lever 26 is secured to one of the 
trunnion shafts 22a, and a reversible electric motor 29 is connected to 
the operating lever 26 through a ball/nut mechanism 28 which comprises a 
threaded shaft 30, and a nut 32 threadedly engaged over the threaded shaft 
30 through cyclic balls 31. An output shaft of the electric motor 29 is 
connected to the threaded shaft 32, and a leading end of the operating 
lever 26 is connected through a pin 34 to a pair of forks 33 projectingly 
mounted on an outer surface of the nut 32. 
The electric motor 29 is supported on a bracket 35 projectingly mounted on 
an outer surface of the cylinder holder 24, and the threaded shaft 30 is 
rotatably carried through ball bearings 38 and 39 on a pair of brackets 36 
and 37 projectingly mounted on the outer surfaces of the motor swashplate 
anchor 23 and the cylinder holder 24. 
Thus, if the threaded shaft 30 is normally rotated by means of the electric 
motor 29, the nut 32 can be moved leftward as viewed in FIG. 8, thereby 
rotating the motor swashplate holder 22 about the trunnion axis O.sub.2 
through the operating lever 26 to right the motor swashplate 20. To the 
contrary, if the threaded shaft is reversed, the nut 32 can be moved 
rightward to tilt the motor swashplate 20 down. When the motor cylinder 
17, with the motor swashplate 20 inclined, is rotated, the motor 
swashplate 20 can produce sequential reciprocations of the motor plungers 
19, 19--to repeat the expanding and contracting strokes. 
Referring again to FIGS. 1 and 2, the pump cylinder 7 and the motor 
cylinder 17 are integrally coupled with each other to constitute a 
cylinder block B which is spline-connected with the output shaft 15 passed 
through a central portion of the cylinder block B. 
A drive gear 3a of the secondary speed-reducing device 3 is inteqrally 
formed on the output shaft 15 adjacent a driven gear 2b of the first speed 
reducing device 2. Meshed portions of a drive gear 2a and the driven gear 
2b in the primary speed-reducing device 2 and meshed portions of the drive 
gear 3a and a driven gear 3b in the secondary speed-reducing device 3 are 
disposed on opposite sides with respect to an axis of the output shaft 15 
therebetween. 
A left end of the output shaft 15 is supported on the motor swashplate 
anchor 23 through a tapered roller bearing 41, and a bearing holder 42 
supporting an inner race of the bearing 41 is fixed to the output shaft 15 
by a cotter 43. 
A right end of the output shaft 15 is also supported on the casing 4 
through a ball bearing 44 arranged with the drive gear 3a sandwiched 
therebetween, and supports the cylindrical input shaft 5 through a tapered 
roller bearing 45. 
In order to rotate the pump swashplate 10 synchronously with the pump 
cylinder 7, the pump swashplate 10 is provided with a spherical recess 10a 
in which a spherical end 9a of the corresponding pump plunger 9 is 
engaged. 
In addition, in order to rotate the motor swashplate 20 synchronously with 
the motor cylinder 17, the motor swashplate 20 is also provided with a 
spherical recess 20a in which a spherical end 19a of the corresponding 
motor plunger 19 is engaged. 
Any of the spherical recesses 10a and 20a is formed to have a radius larger 
than the radii of the spherical ends 9a and 19a, so that the engagement 
with the spherical ends 9a and 19a may be assured in any position. 
Each of outer peripheral surfaces of the pump cylinder 7 and the motor 
cylinder 17 is provided with a large number of grooves 46, 47 for reducing 
the weight between the adjacent cylinder bores 8, 8--; 18, 18--, as shown 
in FIG. 11. 
Referring to FIGS. 1, 2 and 5, between the group of the cylinder bores 8, 
8--in the pump cylinder 7 and the group of the cylinder bores 18, 18--in 
the motor cylinder 17, the cylinder block B is provided with an annular 
inner oil passage 52 and an annular outer oil passage 53 which are 
concentrically arranged about the output shaft 15, first and second valve 
bores 54, 54--and 55, 55--of the same number respectively as the cylinder 
bores 8, 8--and 18, 18--radially passed through annular partition wall 
between the both oil passages 52 and 53 as well as an outer peripheral 
wall of the outer oil passage 53, pump ports a, a, --each permitting the 
communication between the adjacent cylinder bore 8 and first valve bore 
54, and a large number of motor ports b, b--each permitting the 
communication between the adjacent cylinder bore 18 and second valve bore 
55. 
The inner oil passage 52 is provided in the form of an annular groove in an 
inner peripheral surface of the cylinder block B, with its opened face 
closed by an outer peripheral surface of the output shaft 15. 
Each of spool type first distributor valves 56, 56--is slidably received in 
the corresponding one of the first valve bores 54, 54--, and each of 
likewise spool type second distributor valves 57, 57--is slidably received 
in the corresponding one of the second valve bores 55, 55--. A first 
eccentric ring 58 is engaged on outer ends of the first distributor valves 
56, 56--through ball bearings 60 to surround them, and a second eccentric 
ring 59 is engaged on outer ends of the second distributor valves 57, 
57--through ball bearings 61 to surround them. To force these engagements, 
the outer ends of the first distributor valves 56, 56--are interconnected 
by a first forcing ring 62 concentric with the first eccentric ring 58, 
and the outer ends of the second distributor valves 57, 57--are 
interconnected by a second forcing ring 63 concentric with the second 
eccentric ring 59. 
The eccentric ring 58, as shown in FIG. 3, is connected to an inner end of 
the cylindrical input shaft 5 through a pivot 64 parallel to the output 
shaft 15, for swinging movement between a clutch ON position n and a 
clutch OFF position f. The first eccentric ring 58 assumes a location 
eccentrically offset a predetermined distance .epsilon..sub.1 from the 
center of the output shaft 15 along the trunnion axis O.sub.1 in the 
clutch ON position n, and also assumes a location eccentrically offset a 
distance .epsilon..sub.2 greater than such eccentric amount 
.epsilon..sub.1 from the center of the output shaft 15 in the clutch OFF 
position f. The clutch OFF position f is defined by an inner peripheral 
edge of the first eccentric ring 58 abutting against a first stopper 65, 
projectingly mounted on an inner end face of the input shaft 5, while the 
clutch ON position n is defined by an inward directed projection 67 of the 
first eccentric ring 58 abutting against a second stopper 66 projectingly 
mounted on the inner end face of the input shaft 5 on the opposite side 
from the first stopper 65. 
The first eccentric ring 58 and the cylindrical input shaft 5 are provided 
respectively with spring-receiving pieces 68 and 69 projecting therefrom 
and circumferentially opposed to each other on the opposite side from the 
pivot 64, and a clutch spring 70 is mounted in compression between the 
spring-receiving pieces, so that the first eccentric ring 58 is biased 
toward the clutch OFF position f by a spring force of the clutch spring 
70. 
Further, the first eccentric ring 58 is integrally formed with a weight 71 
which exhibits a centrifugal force toward the clutch ON position n during 
rotation of the first eccentric ring 58. 
Yet further, the first eccentric ring 58 is formed with a canopy 72 
connected to the spring-receiving piece 68 for covering the spring 70. 
Still further, the first eccentric ring 58 and the cylindrical input shaft 
5 are provided, on the opposite side from the pivot 64, respectively with 
a guide groove 73 and a guide projection 74 which slidably engage each 
other to define a path of swinging movement of the first eccentric ring 58 
about the pivot 64 (see FIGS. 3 and 4). 
Thus, as a relative rotation is produced between the cylindrical input 
shaft 5 and the pump cylinder 7 when the first eccentric ring 58 assumes 
the clutch ON position n as shown in FIG. 5, each first distributor valve 
56 is reciprocated in the first valve bore 54 between radially inner and 
outer positions in the pump cylinder 7 with a stroke two times the 
eccentric amount .epsilon..sub.1 by the first eccentric ring 58. In a 
discharge region D of the hydraulic pump P, the first distributor valve 56 
is moved in the inner position to put the corresponding pump port a into 
communication with the outer oil passage 53 and out of communication with 
the inner oil passage 52. In the intake region S, the first distributor 
valve 56 is moved in the outer position to put the corresponding pump port 
a into communication with the inner oil passage 52 and out of 
communication with the outer oil passage 53. 
As a relative rotation is produced between the cylindrical input shaft 5 
and the pump cylinder 7 when the first eccentric ring 58 assumes the 
clutch OFF position f (see FIG. 6), each first distributor valve 56 is 
reciprocated in the first valve bore 54 between the radially inner and 
outer positions in the pump cylinder 7 with a stroke two times the 
eccentric amount .epsilon..sub.2 by the first eccentric ring 58, and in 
the inner and outer positions, the first distributor valve 56 permits the 
direct communication between the inner and outer oil passages 52 and 53. 
On the other hand, the second eccentric ring 59 is integrally coupled to 
the cylinder holder 24 to assume a location eccentrically offset a 
predetermined distance .epsilon..sub.3 from the center of the output shaft 
15 along the trunnion axis O.sub.2, as shown in FIG. 7. 
Thus, when the motor cylinder 17 is rotated, each second distributor valve 
57 is reciprocated in the second valve bore between radially inner and 
outer positions in the motor cylinder 17 with a stroke two times the 
eccentric amount .epsilon..sub.3 by the second eccentric ring 59. In an 
expansion region Ex of the hydraulic motor M. the second distributor valve 
57 is moved in the inner position to put the corresponding motor port b 
into communication with the outer oil passage 53 and out of communication 
with the inner oil passage 52. In the contraction or shrinkage region Sh, 
the second distributor valve 57 is moved in the outer position to put the 
corresponding motor port b into communication with the inner oil passage 
52 and out of communication with the outer oil passage 53. 
In the above construction, when the cylindrical input shaft 5 of the 
hydraulic pump P is driven from the engine E through the primary speed 
reducing device 2, causing the pump swashplate 10 to alternately provide 
the discharge and intake strokes to the pump plungers 9, 9--, each pump 
plunger 9 allows a working oil to be pumped from the cylinder bore 8 into 
the outer oil passage 53 during passing through the discharge region, and 
allows the working oil to be drawn into the cylinder bore 8 from the inner 
oil passage 52 during passing through the intake region S. 
Now, in an idling condition of the engine E, the speed of rotation of the 
cylindrical input shaft 5 is lower, and the centrifugal force exhibited by 
the weight 71 of the first eccentric ring 58 rotated with the cylindrical 
input shaft 5 is weaker and for this reason, the first eccentric ring 58 
is maintained at the clutch OFF position f by the action of the spring 70. 
Therefore, the first distributor valve 56 permits the direct communication 
between the inner and outer oil passages 52 and 53, as described above, so 
that the cylinder bore 8 containing the pump plunger 9 which is in the 
intake stroke and the cylinder bore 8 containing the pump plunger 9 which 
is in the discharge stroke become short-circuited through the inner and 
outer oil passages 52 and 53, and transmission of any hydraulic pressure 
does not occur. 
As the rotation of the engine E and thus the cylindrical input shaft 5 is 
increased, the centrifugal force provided by the weight 71 is increased, 
and if the centrifugal force exceeds a set load of the clutch spring 70, 
the first eccentric ring 58 moves around the pivot 64 toward the clutch ON 
position n. 
When the first eccentric ring 58 has reached the clutch ON position n, the 
first distributor valve 56 puts the pump port a into communication with 
the inner oil passage 52 and out of communication with the outer oil 
passage 53 in the intake region S, and puts the pump port a into 
communication with the outer oil passage 53 and out of communication with 
the inner oil passage 52 in the discharged regional so that the working 
oil pumped into the outer oil passage by the pump plunger 9 which is in 
the discharge stroke is supplied into the cylinder bore 18 of the motor 
plunger 19 located in the expansion region Ex of the hydraulic motor M, 
and discharged out of that cylinder bore 18 into the inner oil passage 52 
by the motor plunger 19 located in the shrinkage region Sh. 
During this time, the cylinder block B is rotated by the sum of a reaction 
torque received by the pump cylinder 7 from &he pump swashplate 10 through 
the pump plunger 9 which is in the discharge stroke and a reaction torque 
received by &he mo&or cylinder 17 from the motor swashplate 20 through the 
motor plunger 19 which is in the expansion stroke, and such rotating 
torque is transmitted from the output shaft 15 to the secondary 
speed-reducing device 3. 
When the first eccentric ring 58 is passing between the clutch OFF position 
f and the clutch ON position n, a portion of the working oil discharged 
from the hydraulic pump p is short-circuited to the intake side, and the 
remaining thereof is supplied to the hydraulic motor M, resulting in a 
semi clutch condition where the transmission of hydraulic pressure from 
the hydraulic pump P to the hydraulic motor M is moderately suppressed. 
Thus, as the rotation of the engine E is increased, &he transmission of 
hydraulic pressure from the hydraulic pump P to the hydraulic motor M is 
automatically started, and the vehicle smoothly starts to travel. 
The gear shift ratio of the output shaft 31 to the cylindrical input shaft 
5 can be determined by the following equation: 
##EQU1## 
Therefore, if the volume of the hydraulic motor M is varied from zero to a 
certain value, the gear shift ratio can be changed from 1 to a certaIn 
required value. Moreover, because &he volume of the hydraulic motor M is 
determined by the stroke of the motor plunger 19, the gear shift ratio can 
be continuously varied from 1 to a certain value by tilting the motor 
swashplate 20 from its upright position to a certain inclined position. 
During operation of the transmission T, the pump swashplate 10 receives a 
thrust load from the pump plungers 9, 9 --, and the motor swashplate 20 
receives a thrust load in the opposite direction from the motor plungers 
19, 19--. However, the thrust load received by the pump swashplate 10 is 
supported on the output shaft 15 through the angular contact bearing 11, 
the pump swashplate holder 12, &he cylindrical input shaft 5, the tapered 
roller bearing 45 and the drive gear 3a, while the thrust load received by 
the motor swashplate 20 is supported on the output shaft 15 through the 
tapered roller bearing 21, the motor swashplate holder 22, the motor 
swashplate anchor 23, the tapered roller bearing 41, the bearing holder 42 
and cotter 43. Thus, the thrust loads only cause the output shaft 15 to 
produce a tensile stress, and do not act on the casing 4 supporting the 
shaft 15 at all. 
In this case, the motor swashplate holder 22 supports on its front face the 
motor swashplate 20 through the tapered roller bearing 21 and has its back 
face supported on the mo&or swashplate anchor 23 and hence, even if it 
receives a thrust load from the motor plungers 19, 19--through the motor 
swashplate 20, it cannot generate any deflection. Moreover, since the 
motor swashplate holder 22 and the motor swashplate anchor 23 have their 
spherical surfaces f.sub.1 and f.sub.2 opposed to each other and having 
the center provided by the intersection of the axis of &he motor cylinder 
17 with the trunnion axis O.sub.2, the motor sWashplate 22 exhibits an 
aligning function by an interaction of these spherical surfaces. 
Consequently, the motor swashplate holder 22 can be smoothly rotated about 
the trunnion axis O.sub.2 to easily control the inclined angle of the 
motor swashplate 20. In this case, the rotation of the motor swashplate 
holder 22 about any axis other than the trunnion axis O.sub.2 is inhibited 
by engagement of the trunnion shafts 22a and 22b of the mo&or swashplate 
holder 22 with the bearing hole 23a and the bearing recess 23b in the 
motor swashplate anchor 23. 
In addition, the motor swashplate anchor 23 having the concave spherical 
surface f.sub.2 is increased in wall thickness from its central portion 
toward its peripheral edge and has a high rigidity and hence, can 
sufficiently withstand a larger load from the motor swashplate holder 22 
and the tapered roller bearing 21. 
The driven gear 2b and the drive gear 3a provided respectively on the 
concentrically disposed input and output shafts 5 and 15 are disposed 
axially adjacent each other, and moreover, the drive gear 2a meshed with 
the driven gear 2b and the driven gear 3b meshed with the drive gear 3a 
are disposed on the opposite sides from each other with respect to the 
axis of the output shaft 15. Therefore, even if large radial loads are 
applied respectively to the driven gear 2b and the drive gear 3a from the 
mating gears 2a and 3b during transmission, a flexing moment and a 
swinging moment applied to the output shaft 15 due to the both loads are 
extremely small, because the directions of the both loads are opposite and 
moreover, the distance between the points of application of the both loads 
is extremely small. Consequently, it is possible to reduce the loads of 
the bearings 41 and 44 supporting the opposite ends of the output shaft 15 
to prolong the life of the bearings 41 and 44. 
Referring again to FIGS. 1, 2 and 5, the output shaft 15 is projectingly 
provided at its central portion with a central oil passage 80 closed at 
one end, &o which the working oil is supplied from a supplement pump 81 
through primary and secondary oil filters 82 and 83. The supplement pump 
81 is driven from the crank shaft 1 through a transmission device (not 
shown) to draw an oil from an oil reservoir 84 provided in the bottom of 
&he casing 4. The primary oil filter 82 is mounted on a righthand wall of 
the casing 4 in an opposed relation to an inlet of the central oil passage 
80, and the secondary oil filter 83 is mounted in the central oil passage 
80 to extend from the inlet thereof to a central portion thereof. 
A valve case 85 is fitted in the central portion of the output shaft 15 
adjacent an inner end of the secondary oil filter 83 and across the 
central oil passage 80, with its opposite ends faced on the inner oil 
passage 52. As clearly shown in FIG. 5, the valve case 85 includes a cross 
shaped transverse hole 86 opened into the central oil passage 80, and a 
pair of valve chests 87, 87 which are opposed to each other with the 
transverse hole 86 interposed therebetween and permit the transverse hole 
86 to communicate with the inner oil passage 52. A first check valve 88 is 
contained in each of the valve chests 87 for blocking the reverse flow of 
the oil from the inner oil passage 52 into the central oil passage 80. 
The inner peripheral surface of the cylinder block B is opposed to opposite 
ends of the valve case 85 at a very small gap (see FIG. 1), thereby 
preventing the valve case 85 from being withdrawn from the output shaft 
15. 
The output shaft 15 and the cylinder block B are provided with a serial 
supplement oil passage 90 connecting the central oil passage upstream from 
the valve case 85 with the outer oil passage 53, and a second check valve 
91 is interposed on the way to the supplement oil passage 90 for blocking 
the reverse flow of the oil from the outer oil passage 53 into the central 
oil passage 80. 
Thus, if the pressure in the lower pressure side inner oil passage 52 is 
reduced to a lever lower than the pressure in the central oil passage, 
during a normal load opera&ion in which the hydraulic motor M is 
hydraulically driven from the hydraulic pump P, due to leakage of the oil 
through a hydraulic closed circuit between the hydraulic motor and pump, 
the first check valves 88, 88 are opened, permitting the working oil to be 
supplemented from the central oil passage 80 into the inner oil passage 
52. During this time, on the other hand, the working oil in the higher 
pressure-side outer oil passage is blocked from flowing into the central 
oil passage 80 by the second check valve 91. 
During a reverse load operation, i.e., during an engine braking, the 
hydraulic motor M performs a pumping action, while the hydraulic pump p 
performs a motoring action, so that the outer oil passage 53 is changed to 
the lower pressure side, while the inner oil passage 52 is to the higher 
pressure side. Therefore, if the pressure in the outer oil passage 53 is 
reduced to a level lower than the pressure in the central oil passage 80 
due to leakage of the oil, the second check valve 91 is opened, permitting 
the working oil to be supplemented from the central oil passage 80 into 
the outer oil passage 53, and the flow of the working oil out of the inner 
oil passage 52 into the central oil passage 80 is blocked by the first 
check valves 88, 88.