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Патент USA US3095718

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July 2, 1963
|_. 1-. HARRIS
3,095,708
VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY
Original Filed July 9, 1957
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July 2, 1963
L. T. HARRIS
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Original Filed July 9, 1957
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July 2, 1963
1.. T. HARRIS
3,095,708
VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY
Original Filed July 9, 1957
7 Sheets-Sheet 5
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INVENTOR.
LEE T HARRIS
“ya/‘1.417%
July 2, 1963
L. T. HARRIS
3,095,708
VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY
Original Filed July 9, 1957
7 Sheets-Sheet 4
INVENTOR.
LEE T HARRI 8
July 2, 1963
Y
|_. -r. HARRIS
3,095,708
VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY
Original Filed July 9, 1957
'
7 Sheets-Sheet 5
INVENTOR.
LEE T. HARRIS
MW
July 2, 1963
L. T. HARRIS
3,095,703
VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY
Original Filed July 9, 1957
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INVENTOR.
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LEE T HARRIS
BY
July 2, 1963
3,095,708
L. T. HARRIS
VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY
Original Filed July 9, 1957
'
7 Sheets-Sheet 7
TO ENGINE
.
VACUUM
TO IGNITION 5W.
i108 \ 104‘
119
101
T0 BAND BRAKE
SOLENOID SERVO-VALVE
Fig.52
TO BAND BRAKE
;
SOLENOID
SERVO-VALVE
MILL
120
w
BRUSH
BRUSH
ARM
F
COMMUTAT‘ON
ROTOR
.TO
RELAY R
,
TO
RELAY F‘
,
INVENTOR.
BY LEET HARRIS
United States Patent Qv ” Ice‘
1
2
FIGURE 3 is a side elevation of the bearing member
in the same aspect relationship as shown in FIGURE 1;
3,095,708
VARIABLE DISPLACEMENT HYDRAULIC
'
FIGURE 4 is a top plan partially broken away of the
valve and bearing member of FIGURE 1;
ASSEMBLY
Lee T. Harris, 511 William St., Rome, N.Y.
Original application July 9, 1.957, Ser. No. 670,814, now
Patent No. 3,044,409, dated July 17, 1962. Divided
and this application Aug. 24, 1961, Ser. No. 141,555
9 Claims.
3,095,708
Patented July 2, 1963
(Cl. 60-53)
FIGURE 5 is a vertical section of. the valve and hear
ing member along the line V——V of FIGURE 3;
FIGURE 6 is an axial section of the ducting core which
?ts into the bearing member in the same aspect rela
tionship as shown in FIGURE 4 with certain parts shown
‘
This application is a division of my Patent 3,044,409 10 in full lines for clarity;
dated July 17, 1962.
FIGURE 7 is a view similar to FIGURE 6 taken on
line VII-VII of FIGURE 6.
FIGURES is a partial sectional view on line VIII—VIII
(This invention relates to hydraulic pump mechanisms
and more particularly to a variable displacement pump
mechanism that may be used in a driving and ‘driven
of FIGURE 7;
con?guration to comprise a unique hydraulic transmission 15
have generally been satisfactory for various applications
FIGURE 9 is an end elevational view of the disk
support assembly as viewed from the left hand side of
FIGURE 2 and FIGURE 10;
FIGURE 10 is a side elevational view of the disk
support assembly in the same aspect relationship as shown
in FIGURE 2;
FIGURE 11 is a top plan View of the disk support
but have had certain disadvantages such as limited input
assembly of FIGURE 10;
output speed ratios requiring special gear trains, high
frictional and hydraulic losses resulting in low e?iciency
transmissions, costly and expensive machining of highly
FIGURES 12 and 12A are from left to right an end
and side view of either of the two bushings that ?t onto
25 the transverse projections of the disk support assembly;
stressed and critical parts resulting in high cost limited
installations. ‘According to the present invention I have
mission main oil ducting described by 360° counterclock
discovered a variable displacement hydraulic pump and
wlse rotation as viewed from the left hand end of FIG
transmission system that permits in?nite variation in
URE 1 of the ducting core, starting at top center in
system.
For some years now variable displacement pumps have
been used in hydraulic transmission systems for coupling
a prime mover to a load such as in automobiles and other
industrial applications. These transmissions and pumps
FIGURE 13 is a diagrammatic drawing of the trans
input-output speed and torque ratios that is dynamically 30 respect to FIGURE 1;
and statically balanced so as to minimize frictional forces
and that still may be manufactured largely by castings in
a cheap and simple manner.
~
‘
.
_FIGURE 14 is a side View of the pump chamber end
piece in the same aspect relationship as shown in
' FIGURE 1;
Accordingly it is an object of the present invention to
FIGURE 15 is a top plan view of the chamber end
.
provide a variable displacement pump mechanism that is 35 piece of FIGURE 14;
in?nitely variable in displacement. It is another object
FIGURE 16 is a vertical transverse section of the pump
of the present invention to provide a transmission system
chamber end piece along the line XVI—XVI of FIG
that is in?nitely variable in input-output speed ratio
throughout the design limits thereof. It is another object
URE 14;
'
‘FIGURE 17 is a right half end view of the frontal end
of the present invention to provide a variable pump 40 plate of the inner housing as viewed from the right hand
mechanism that is dynamically and statically balanced so ' side of FIGURE 1;
as to minimize frictional forces therein. It is another
1FIGURE 18 is an axial section of the frontal end plate
object of the present invention to provide a transmission
of the inner housing in the same aspect relationship as
system‘ that is extremely ?exible in application and effi
shown in FIGURE 1;
cient in operation. It is another object of the present 45
FIGURE 19 is an axial section of the frontal end plate
invention to provide a transmission system wherein the
of the inner housing in the same aspect relationship as
driving pump may be located remotely from one or more
shown in FIGURE 2;
driven pump mechanisms. It is another object of the
present invention to provide a variable displacement
hydraulic pump mechanism that may be combined with
a corresponding mechanism to provide a transmission that
FIGURE 20 is an end View of the intermeshed rotor
and divider disk of the driving pump as viewed from the
It is a still further object of the present invention to
provide a variable displacement pump and transmission
FIGURE 23 is an enlarged view of a rotor vane and
.
right in FIGURES 1 and 2;
FIGURE 21 is a side view of the driving pump rotor
rotates together at a one-to-one speed ratio without
and divider disk of FIGURE 20 shown with the latter
external gearing or locking. It is another object of the
in section along line XXI—XXI of FIGURE 20;
present invention to provide a transmission system that
_ FIGURE 22tis a partial transverse section of the driv
55
requires no gear train or other speed control mechanism
mg pump ro or and divider dis
a
'
XXII~XXII of FIGURE 21;
k 310mg the hue
to connect it between a prime power source and ‘a load.
a partial circumferential section of the intermeshed divider
disk showing the relationship of the two to each other at
60
operate and economical to construct. These and other
a phase of the rotation cycle where the maximum angle
and further objects will be in part apparent and in part
of incidence occurs in respect to the normal;
pointed out as the speci?cation proceeds.
FIGURE 24 is a partial radial section of the divider
In the drawings:
disk along the line O-XXIV of FIGURE 22 with the
FIGURE 1 is an axial section of the transmission with
the rotatable inner portion shown at a position wherein 65 rotor removed;
FIGURE 25 is‘a partial radial section of the divider
the pitch axis of the driving and driven pump divider disks
disk along the line O—-XXV of FIGURE 22 with the
are perpendicular to the plane of the drawing and wherein
the transmission is at an input-output forward speed _ rotor removed;
FIGURE 26 is an enlarged exploded perspective view
ratio ofone-to-one;
FIGURE 2 is an axial section on line II——II of FIG- 70 of adivider disk insert and the corresponding urging
URE 1 of the transmission except that thedivider disk
of the driving pump is at the zero pitch position;
‘FIGURE 27 is a ‘schematic representation of a method
mechanism that is highly e?icient, extremely simple to
spring;
"
.
3,095,708
3
4
for forcibly maintaining a relatively uniform speed rela
tionship of the divider disk in respect to the rotor;
FIGURES 28, 29, and 30 are enlarged top, side, and
intersecting the geometric center of the pump chamber.
Disk support assembly 16-16’ is subject to control in its
end views respectively of a rotor vane showing a fur
In operation, oil sealing action between divider disk
15-15’ and rotor 14 is provided by contact or proximity
of the spherically contoured inner contact surfaces 23’
ther alternative method of forcibly maintaining a rela
tively uniform speed relationship of the divider disk in
respect to the rotor; and
FIGURES 31 and 32 are schematic drawings of a pos
movements about the axis as will be described herein.
and outer slot contact surfaces 27’ of said divider disk
15-15’ with the mating spherically contoured contact
surfaces 23 and 22’ of said rotor 14 in any supported posi
sible control system for automatically controlling the in
put-output speed ratio of the transmission for automotive 10 tion of said divider disk 15-15’ within the limits of move
ment of disk support assembly 16-16’ about its transverse
application.
axis. Oil sealing contact between the slots 27 and the
Referring now to FIGURES l and 2, the transmission
includes an input shaft 1; an output shaft 2; an outer hous
ing comprising end plate 3 and upper and lower casings
radial surfaces of vanes 22 is maintained by inserts 29
which are slidably mounted in recesses 28, said inserts 29
4 and 5 respectively; an inner housing comprising end 15 being continuously but yieldably urged against vanes 22
by springs 35 or hydraulic pressure channeled into recesses
plate 6, front and rear barrel sections 8 and 9, central sec
28 by shuttle balls 31 or a combination of the two forc
tion 10, and rear section 12; ducting core 13; a ?rst pump
ing means. Shuttle balls 31 actually serve two purposes:
chamber bounded by pump chamber end piece 7, bearing
(1) by being exposed to the pressure appearing on both
member 11, and disk support assembly 16-16’; a second
pump chamber bounded by member 10 and rear section 20 sides of divider disk 15-15’, between any two adjacent
vanes 22, through ori?ces 39 said shuttle balls 31 are auto
12; a driving pump including a rotor 14 and divider disk
matically seated against the ori?ces 39 which are facing
15-15’; a driven pump or motor including rotor 17, and ,
the lowest pressure, thereby leaving the ori?ces 39 which
divider disk 18-18’; over-running clutch 19; and a band
are facing the highest pressure open, and permitting the
brake including band 21 and drum 20.
The driving pump is in?nitely variable in displacement 25 highest available pressure to be transmitted into valve
chambers 30 and then into recesses 28 via oil channels 32
from design maximum forward output through zero out
wherein said high pressure acts against the backsides of
put to design maximum reverse output as will be subse
quently described. The rotor 14 is mechanically coupled
to input shaft 1 and has therein six radial vanes 22 (see
FIGURE 20) circumferentially spaced 60° apart at the
centers with ?at sides and peripheral surfaces which are
inserts 29 thereby counteracting the highest pressure ap
pearing on the exposed face surfaces 40 of said inserts 29
(FIGURES 23 and 26); (2) by leaving one ori?ce 39
open at all times, the oil trapped in recesses 28 by inserts
as surfaces of a common sphere whose center is the geo
metric center point of the pump chamber, and a core sec
29 is provided an escape when said inserts 29 are forced
as to provide an oil sealing contact to the end and periph
eral surfaces of revolution of the vanes 22 of rotor 14.
Thus the six vanes 22 circumferentially divide the pump
ing to the excess clearance between the face contact sur
back into said recesses 28 from extended positions.
tion whose outer surface 23 is spherically contoured with
In operation, it will be understood that when the input
the sphere center coinciding with the same aforementioned 35 shaft 1 is rotated, as for example by a prime power
geometric center point of the pump chamber, said surface
source, that rotor 14 likewise rotates by virtue of its direct
23 being the innermost surface or boundary of the pump ' coupling to input shaft 1 and that since divider disk 15
chamber. Front face surface 24 of chamber end piece 7,
15’ is mechanically intermeshed with rotor 14, it will like
rear face surface 25 of bearing member 11 and peripheral
wise rotate with said rotor 14 but will be permitted some
surface 26 of assembly 16-16’ de?ne the other principal 40 circumferential movement relative to rotor 14, said rela
boundaries of the ?rst pump chamber and are shaped so
tive movement being restricted to the limits correspond
faces 40 of the opposing inserts 29, on each side of a
particular slot 27 when said inserts 29 are in their fully
chamber into six separate and equal volumes of ?xed dis 45 retracted positions, as compared to the thickness of vanes
placement.
22.
Divider disk 15-15’ comprises two main sections (see
FIGURES 20, 21, and 22); six slots 27 for accommoda
the maximum phase deviation of corresponding points of
tion of the vanes 22 of rotor 14, recesses 28 for accom
Said excess clearance should be at least as great as
rotor 14 and divider disk 15-15’ when calculated from
the equation tan B=tan A cos P, in which P is the pitch
modating inserts 29, valve chambers 30 for accommoda 50 angle of divider disk 15-15’, A is any angle of the com
plete 360° cycle of rotation of rotor 14 with either side
tion of shuttle balls 31, oil channels 32, recess grooves 33,
of the pitch axis of divider disk 15-15’ as the zero degree
oil channels 34, and means for securely fastening the two
reference and B is the angle of correspondence of the angle
main parts 15 and 15’ to each other. Said divider disk
A on said divider disk 15-15’ when translated to the
15-15’ intermeshes with rotor 14 as shown in FIGURES
20, 21, and 22 and serves to longitudinally divide each 55 plane of rotation of rotor 14. Therefore, angle B sub
tracted from the angle A represents the phase difference
?xed volume between adjacent vanes 22 into two parts.
of slots 27 in respect to corresponding vanes 22 for various
As visualized in FIGURE 21 divider disk 15-15’ would
angles of rotation and for various pitch angles at which
be assembled on rotor 14 by sliding the part 15 onto said
divider disk 15-15’ may be set between zero and maxi
rotor 14 from the left and sliding the other part 15’ on
from the right, one part, either 15 or 15’ containing the 60 mum design limits. In order to arrive at the minimum
design clearance for vanes 22 in slots 27, angle A should
inserts 29 with corresponding springs 35 and shuttle balls
31, and the two parts 15-15’ being joined by screws 36
as shown in FIGURES 20' and 24 or in some other suit
be taken at angles of 45°, 135°, 225° and 315°, and
angle P at the maximum design pitch angle of divider disk
15-15'. Maximum angle P is 15° in the illustrated em
able manner. Divider disk 15-15’ is supported in axial
position by contact or proximity of the outer face bear 65 bodiment which causes the maximum phase difference be
tween rotor 14 and divider disk 15-15’ to be approxi
ing surfaces 37 of said divider disk 15-15’ with the
mately :1"; therefore the total clearance between face
recessed mating surfaces 37’ of disk support assembly
contact surfaces of opposing inserts 29 when in their
16-16’ (FIGURES 2 and 10) and in radial position by
fully retracted positions in grooves 28 should be at least
rotor 14. Disk support assembly 16-16’ contains two
projections, 33 and 38’, which are journaled in mating 70 equal to the thickness of vanes 22 plus r sin 2°, where r
is the radius from the geometric center of the pump
bearing surfaces contained in the barrel sections 8 and 9
housing.
of the inner housing (FIGURE 2), said disk support as
sembly 16-16’ thereby being pivotally mounted on a
Although inserts 29 will perform most of the yield
transverse axis de?ned by an imaginary line connecting
ing to the phase variation between rotor 14 and divider
the two diametrically opposite bearing centers, said line 75 disk 15-15', the frictional resistance to rotation against
3,095,708
5
divider disk 15-15’ will tend to cause said divider disk
15-15’ to lag behindrotor 14 to the limits imposed by
the excess clearance between opposed inserts 29 which
will tend to cause cyclic variations in the speed of said‘
divider disk 15-15’ in respect to the speed‘ of rotation
of rot-or 14.
6
“reverse” circulation will be achieved when the plane
of rotation of divider disk 15-15’ is aligned with the
letters R-R. Speci?c circulation paths for the ?uid
will be described in detail herein.
Examination of the underlying principles of operation
FIGURE 27 illustrates a means whereby
will disclose that insofar as rotor 14 is concerned, no ap
a more nearly constant speed rotation of divider disk
15-15’ in respect tov rotor 14 may be assured, when
the transmission is operating at an input-output speed
preciable unb-alanced forces in respect to the housing oc
ratio of one to one. A groove 41 is provided in the pe
ripheral surface 22’ of each vane 22, the sides of which
are shaped to provide contact to at least one point of a
mating projection 42, located at the peripheral center of
each slot 27 in divider disk 15-15’, at all phase angles
of rotation between 315° to 45° and 135° to 225°, with
the zero degree reference considered to be in coincidence
with one side or the other of the transverse pitch axis of
divider disk 15-15’.
A further alternative means whereby more nearly con
stant speed rotation of divider disk 15-15’ in respect to
rotor 14 may be assured (FIGURES 28, 29, and 30) is
to shape the sides of vanes 22 so as to keep one or
both of the opposing inserts 29‘ in any one slot 27
fully retracted into their respective recesses 23 at all
cur but that the hydraulic forces on divider disk ‘15-15’
would be severelyunbalanced if not compens-ated for. It
10 is for said purpose of compensation that recess grooves
33 are provided around each peripheral side of divider
disk 15-15’ (FIGURES ‘20' and 21) with oil channels
34 providing hydraulic communication between said recess
grooves 33 and the opposite side of divider disk 15-15’
in the pump chamber. The effective area of recess groove
33 is chosen so that said area multiplied by the radius
of the effective center of ?uid pressure of said recess
groove 33 is equal to the effective area of the part of di
vider disk 15-15’ contained inside the pump chamber
between adjacent vanes 22 multiplied by the radius of the
e?iective center of ?uid pressure of said second area. In
this way any hydraulic forces appearing on the inside
areas ‘of divider disk 15-15’ will be dynamically coun
teracted by proportionate forces created by transmission
phases of cyclic rotation between the angles of 315° 25 of pressure through oil channels 34 to recess grooves 33
to 45 ° and 135 °_to 225” with the zero degree reference
located on the opposite sides. The obvious purpose of
again chosen to coincide with either side of the trans
verse pitch axis of divider disk 15-15’.
‘By inspection of the drawings, it will be apparent that
providing hydrostatic balance for divider disk 15-15’ is
to reduce the ‘frictional resistance to rotation of said
divider-disk 15-15’ in disk support assembly 16-16‘
divider disk 15-15’ operating in a complementary man 30 occurring as a result of hydraulic loading, and thereby
ner with rotor 14 essentially divides the pump chamber
to enhance the e?iciency of power transfer and decrease
mechanical wear. Another method for reducing the fric
into two pumping sections (one of each side of divider
disk 15-15’) which operate in'a diametrically opposite
tion, in lieu of hydrostatic balance as described above, is
. fashion to each other with each section having its own
to support divider disk 15-15’ on ball or tapered roller
input-output ports, represented by the numbers 43 and 35 bearings within disk support assembly 16-16’. A third
43’ for the front pump section and numbers 44 and 44’
method might be ‘a combination of the ‘above two methods.’
for the rear pump section as shown in FIGURE 4 and
The driven pump or motor, hereinafter to be referred
to as the “driven pump,” in the illustrated embodiment is
as shown schematically in FIGURE 13. The ports 43,
43’, 44 and 44’ should each have a circumferential length
essentially identical to the driving pump except that divider
of ‘approximately 120°, leaving an uncut-away circum 40 disk 18-18’ has a ?xed plane (pitch angle of 15°) of
rotation about its axis by virtue of its being supported in
ferential section of approximately 60° in length between
the inner housing and accordingly is a ?xed displacement
each port which will serve to trap the ?uid between ad-'
jacent vanes 22 at maximum of minimum points as
the case ‘may be and thereby to prevent unrestricted cir
driven pump may be of the variable displacement type
culation of oil between input and output ports which would
also and that said driven pump whether of the ?xed or
less than six vanes as shown in the drawings, the ports
should of course be of appropriate circumferential length
corresponding to the number of vanes.
Referring now to FIGURE 1 it is obvious that when
the plane or rotation of divider disk 15-15’ is completely
from the driving pump in basic design. The rotor 17
with six vanes 22A, divider disk 18-18’ and associated
vertical (zero pitch angle) represented by alignment with
the above description of the driving pump.
the letters N-N, the volume of ?uid entrapped between
adjacent vanes 22 and the surfaces of the pump cham
Ducting core 1-3 provides means of ?uid communica
tion between the driving pump and driven pump by pro
viding ?uid channels '45 and 45’ as shown in FIGURES
1 and 2 and as shown schematically in FIGURE 13. Chan
nel 45 connects ports 43' and 44 of the driving pump to
ports 436 and 47 of the driven pump and channel 45’ con
nects ports 43’ and 44’ of the driving pump to ports 46’
pump or motor.
It will be apparent, however, that the
tend to bypass the pump. If the rotor employs more or 45 variable displacement type may :be similar to or different
ber remains constant when rotor 14 and divider disk
15-15’ rotate and that therefore all of the ?uid con
tained within the chamber rotates with the rotor with
no ?uid input or output through ports 43, 43’, 44 and
44’. It should be equally obvious that if the pitch angle
of divider disk 15-15’ is at any valve other than zero
that the individual ?uid volumes between adjacent vanes
22 on each side of divider disk‘15-15’ will vary in a
sine manner as motor 14 progresses from 0° through
pump chamber ‘acting in cooperation with each other
form the essential elements of the driven pump, the op
eration of which will be fully understood by reference to
and 47’ of the driven pump.
7
,
'
'
It will be apparent that when input shaft 1 is. rotated
clockwise when viewing FIGURES 1 and 2 from the left,
and that when divider disk 15-15’ is set at a pitch angle
360° for each rotation cycle, the total changebetween
greater than zero toward alignment with the letters F-F
minima and maxima volumes being a direct function of
‘of FIGURE 1, fluid will be expelled from the chamber of
the pitch angle of divider disk 15-15’. Therefore a
the driving pump through ports 43 and 44 into channel
low compressibility ?uid such as oil when contained there
45 from which said ?uid will be ‘forced into the chamber
in will be forcibly circ-ulated into and out of the pump
of the driven pump through ports 46* and 47 thereby caus
chamber through ports 43, 43’, 44 and 44’ in amounts
ing rotor 17 to rotate in the same clockwise sense as
proportional to the pitch ‘angle of divider disk 15-15’. 70 rotor 14, causing said rotor ‘17 at the same time to expel
In the illustrated embodiment, maximum forward cir~
a like quantity of ?uid from the chamber of the driven
culation is achieved when the plane of ‘rotation of divider
pump through ports 46' and 47’ into channel 45’ from
disk 15-15’ is aligned with the letters F-F (FIGURE
which said like quantity of ?uid will be returned to the
1); i.e. the pitch angle of divider disk 15-15’ isat the
chamber of the driving pump through ports 43’ and 44’.
“forward” maximum of 150°. Likewise, the maximum 75 Fluid ?ow under the above conditions is shown sche
3,095,708
7
matically by the solid arrows in FIGURE 13. When di
vider disk 15-15’ is set at a pitch angle greater than
zero toward alignment with the letters R-R of FIGURE
1 and rotation of input shaft '1 is in the same clockwise
8 .
spect to‘ the outer housing as shown in FIGURES 1
and 2.
In order to obtain a mechanical indication of the trans
mission input-output speed ratio which might be required
schematically by the dashed arrows in FIGURE 13, in
which case output shaft 2 would rotate counter-clockwise
for a follow-up type of control system, a ratio indicating
linkage is provided as shown in FIGURE 6. This link
age starts with a pair of gears 53 and 54 ?xed on shaft
or in the opposite direction to the direction of rotation of
55 and mounted so that gear 53 meshes with a circular
sense, the ?ow of ?uid is reversed and would be as shown
gear segment 56 (FIGURE 11) on disk support assem
input shaft \1.
It is also apparent that regardless of the direction of 10 bly 116-16’ and gear 54 meshes with gear rack 57
mounted in sleeve 58 inside dusting core 13. When disk
rotation of input shaft 1, with divider disk 15-15’ pitched
support assembly 16-16’ rotates about its transverse axis
toward alignment with the letters F-‘F in FIGURE 1,
the output shaft 2 will rotate in the same direction as in
a proportionate angular rotation, corresponding to the
gearing ratio between said gear '53 and said circular gear
put shaft 1, and with divider disk 15-15’ pitched to
ward alignment with the letters R-R in FIGURE 1, 15 segment 56, is induced in gear 54 causing gear rack 57
rotation of output shaft 2 will be opposite to that of in
put shaft 1.
to move longitudinally together with sleeve 58 to which
> It will be understood that the ratio of speed of the in
58 at one end (and therefore moves longitudinally with
it is mechanically ?xed. Shaft 59 is threaded into sleeve
put shaft 1 to the speed of the output shaft 2 will be in
said sleeve 58) and has ?xed in position on the other end
versely proportional to the ratio of the displacement of 20 of shaft 59 by retaining nut 61 a bearing assembly 60
the driving pump to the displacement of the driven pump
which is slidably mounted in output shaft 2 (FIGURE
and that the output-input torque ratio will be directly pro
2). An indicating sleeve 62 having thereon a rim 66 is
slidably mounted on output shaft 2 and mechanically
portional to the input-output speed ratio, neglecting fric
fastened to bearing assembly 60 by screws 63 which are
tional losses. In the illustrated embodiment of the trans
mission, a one to one forward ratio of speed between in 25 free to move longitudinally in slots 64, said indicating
sleeve 62 thereby being caused to move longitudinally
put shaft ‘1 and output shaft 2 occurs when the plane of
rotation of divider disk 15-15’ is in alignment with the
as a unit with bearing assembly 60. A lever assembly
letters F-F in FIGURE 1.
consisting of two arms 65 and 65’, the extremities of
Pitch control of divider disk 15-15’ is effected by
which bear against rim 66 of indicating sleeve 62 ?xed
30 on a shaft 67 to which a third arm 68 is rigidly fastened
control of the movement of disk support assembly 16
16' about its transverse axis by permitting oil under pres
extends the linkage away from the output shaft to the
sure to enter either control chamber A through channel
48 or control chamber B through channel 49 and to exit
control mechanism. Shaft 67 is trunnioned on a trans
verse axis so that its center line is the pivot point for the
from the other (see FIGURE 1). Disk support assembly
lever assembly which is caused to pivot about its axis to
16-16’ has projections or vanes 50 (FIGURES 9, 10, 35 follow the movement ‘of indicating sleeve 62 by spring
85 or other equivalent means. A rod 69 is pivotally
11) which together with cylindrical surfaces 51 (FIG
URES 17, 18, 19) in end plate 6 and the spherical periph
mounted on the end of arm 68 so as to translate the ro
tative movement of said arm 68 into longitudinal motion
eral surfaces 51' of control chambers A and B (‘FIGURE
l1) forms a fluid seal to prevent the unrestricted ?ow of
for utilization by a transmission control means. Rota
oil between control chambers A and B. The axial center 40 tive motion could just as well be imparted to the trans
line of said cylindrical surfaces 511 coincides with the
mission control means by appropriate gearing arrange
transverse axis of rotation of disk support assembly
ment from shaft 67 of the lever assembly if said rotative
16-16’.
Vanes 52 are similar to vanes 50 but are not
motion were more suitable for use by the control means.
necessarily designed to provide oil sealing contact with
When disk support assembly 16-16’ rotates about its
any surfaces of the housing, the essential purpose of said 45 transverse axis a proportionate angular rotation, corre
vanes 52 being to balance the weight of vanes 50 in re
sponding to the gearing ratio between said gear 53‘ and
spect to the longitudinal axis of the transmission when
said circular gear segment 56, is induced in gear 54 mus
the inner housing is rotating, although said vanes 52 also
ing gear rack 57 to move longitudinally together with
serve to strengthen the disk support assembly 16-16’
sleeve 58 to which it is mechanically ?xed. This move
at points where needed.
ment is in turn transmitted through the linkages described
To increase the input-output forward speed ratio or to
until rod 69 moves the indicating mechanism to the cor
decrease the input-‘output reverse speed ratio, oil under
pressure is permitted to enter control chamber B through
channel 49 from the control pressure source and a like
amount of ?uid is permitted to exit from control cham
her A through channel '48 and to return to the sump or
inlet of the control pressure source, which action causes
disk support assembly 16-16’ to move in the direction
responding indication.
Bypass control piston 70 (FIGURES 6 and 7) form
ing the end of shaft 59 which attaches to sleeve 58, con
tains an annular groove 71, the purpose of which is to
bypass a nominal amount of oil from the high pressure
channel in ducting core 13 via ori?ces 72 and 72’ to the
low pressure channel in said ducting core 13 when a load
connected to output shaft 2 is being accelerated from the
toward alignment of the transverse center plane of said
divider disk 16-16’ with the letters \F-F of FIGURE 60 motionless state in either the forward or reverse di
1. To decrease the input-output forward speed ratio or
rection. Bypass control piston 70 acts in cooperation
to increase the input-output reverse speed ratio, oil un
with ori?ces 72 and 72' to form a progressively smaller
der pressure is permitted to enter control chamber A
restriction to the bypass oil path as divider disk 15-15’
through channel 48 and to exit from control chamber B
is increased in pitch in either the forward or reverse
through channel 49 and to return to the sump or inlet
direction from the zero pitch angle up to a predetermined
of the control pressure source, which action causes disk
pitch angle at which point the bypass path will be com
support assembly to move in the direction toward align~
pletely obstructed by the smooth cylindrical surface of
ment of the transverse center plane of said divider disk
the bypass control piston. By this means, torque in vary
16-16’ with the letters R-R of FIGURE 1. Annular
ing degrees may be applied to a motionless load without
grooves 48’ and 49' are provided in end piece 3 of the
stalling the power source, thereby permitting smoother
outer housing as a means of effecting hydraulic com~
starts. For some forms of application of the transmis
munication between a control system located externally
sion a bypass valve obviously would not be required.
to the inner housing and channels 48 and 49‘ respectively
Shuttle valve 73 (FIGURES 6 and 7) may be mounted
under conditions when the inner housing is rotating as
in the ducting core 13 by threaded means as shown and
well as when said inner housing is held motionless in re 75 comprises a valve housing 74 and shuttle spool 75. The
3,095,708
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valve on one end is exposed to fluid channel 45 which
and output shafts due to the uniformity of pressure in‘
directly communicative spaces.
Circulation of oil with the inner housing in synchron
is normally the high pressure channel and on the other
end is exposed to ?uid channel 45' which is normally
the low pressure or return channel. The normally high
pressure and return channels 45 and 45' may alternate
with each other pressure-wise, depending upon the mode
of operation of the transmission, however, the shuttle
spool 75 is always forced against the valve seat on the
high pressure side thus closing off the high pressure from
ous rotation will be near zero; however, this will not .
impair torque transfer inasmuch as the torque transfer is
a function of the net resultant hydraulic forces acting
on vanes 22 and 22A of the rotors 14 and 17 and not
a function of the oil transfer.
When the input-output forward speed‘ ratio is greater
the center of ducting core 13‘; on the oher hand the 10 than one-to-one, and when the power source is supply
low pressure side is always in hydraulic communication
ing torque to input shaft 1, greater torque is applied to
with the center of the ducting core 13 so that except for
the pressure drop due to oil flow through the communica
tion channeling, the low pressure or return prime channel
driven pump rotor .17 then is applied to driving pump
rotor ~14, resulting in a net reverse reaction force acting
on the inner housing'through the respective divider disks
45 [or 45’ will he at the same ?uid pressure as the center
18-13’ and -16—16’ which tends to cause said inner
of ducting core 13. By this device a changing pressure
may be applied to the return ?uid channel 45 or 45’
through channel 76, located at the center of input shaft
1 (FIGURES l and 2), togdiscourage any tendency to
ward cavitation by pumping action of the driving and
driven pumps and also to make up any leakage of oil
from thelinner housing of the transmission. Annular
groove 77 is provided in end piece 3 of the outer housing
housing to rotate in the opposite direction to the rota
tion of the input and output shafts. The function of
over-running clutch 19, shown diagrammatically in FIG
URES 1 and 2, is to prevent the reverse rotation of the
inner housing but to permit free forward rotation.
-When the input-output reverse speed ratio is greater
than ,one-to-one, and when the power source is supplying
torque to input shaft 1, greater torque'is applied to driven
pump rotor 17 than is applied to driving pump rotor 14,
‘as a means of effecting hydraulic communication be
tween an oil pressure source located externally to the 25 and the net reaction force acting on the inner housing in
this case tends to cause said inner housing to rotate in
inner housing and channel 76 under conditions when the
inner housing is rotating as well as when said inner
the forward direction which is undesirable inasmuch as
said forward rotation would cause an effective reduction
housing is held motionless in respect to the outer hous
in the output-input torque ratio. Band 21, when caused
ing as shown in FIGURES, 1 and 2.
to contract around the drum 20 with adequate force by
Referring to FIGURE 2 it will be apparent that spring
loaded piston 78 may be a pressure regulating device
hydraulic servo means or other suitable means, prevents
which will maintain a certain changing pressure depending
forward rotation of the inner housing.
Under conditions when the impelling ‘force is trans-‘
upon the spring characteristics and certain other factors
when oil is forcibly circulated through channel 79 at
mitted via output shaft 2 to rotor 17, thence to rotor
the center of shaft 59. Oil flow would be through oil 35 14 and then via input shaft 1 to the power source which
in this case becomes the load, as ‘for example when a
channel 76 from annular groove '77 into the center of
vehicle is descending a ‘grade, and it is desired to increase
ducting core 13 from where it would be transmitted
the load by forcing the power source to a higher speed
through channel 79 located at the center of shaft 59 to
by increasing the transmission input-output speed ratio,
the face of piston 78, which would cause the piston to
move toward the right in FIGURE 2 until ori?ces 80 40 hydraulic forces tend to rotate the inner housing in,
were exposed to the extent necessary to pass the volume
the forward direction, which rotation if permitted to oc
cur would lower the input~output speed ratio; therefore
of oil being circulated. iOil would flow through ori?ces 89
it is again desirable for the inner housing to be held
into the hollow space in output shaft 2 partially occupied
against forward rotation by constriction of the hand 21
by bearing assembly ‘60 from where said oil would then
pass through slots '64 and eventually along the lower side
around the drum 20.
of the outer housing where it would he returned to the
Shield 84 is provided for the purpose of reducing wind
age losses when the inner housing is rotating and is de
sump through channel ‘81 in end piece 3 (FIGURE 1).
Circulation of oil in this manner may also provide a
signed to cover the circumferentially asymmetrical por
means of transferring heat from the transmission proper
tion of the inner housing.
to the sump or to‘ a heat exchanger; however, under most
‘Possibilities for many different configurations of con-‘
conditions the power loss in the transmission will he suf
trol systems exist for adapting a transmission of this type
' ficiently low as to render the use of a large capacity heat
to automotive, industrial or other applications. FIG
exchanger unnecessary.
URE 3-1 is a simpli?ed schematic drawing of a control
The inner housing comprising parts aforementioned is
system which might be used to ‘automatically control the
journaled at the input shaft end on hearings 82 and at
input-output speed ratio of the transmission in auto
the output shaft end on hearings 83», said inner housing
motive vehicles as a function of output shaft speed and
being disposed to rotate under certain conditions as de
torque demand, or input shaft speed under certain con
scribed herewith. When ‘operating at an input-output
ditions. The control system herein described is jointly
forward speed ratio of one to one, the reaction forces im
electrical, mechanical and hydraulic in operation.
parted to said inner housing through divider disk 15-15’ 60
The hydraulic portion includes an oil pressure source
are equal and opposite to the reaction forces imparted to
100 and a follow up valve assembly 101 containing a
said inner housing through divider disk 18—-18', said
valve spool 102 which is axially slidable in valve sleeve
forces effectively cancelling each other, thus permitting
internal frictional forces to rotate said inner housing
MP3, said valve sleeve 103 being slidably mounted in
cylinder 164. The lands 105 and 106 on the valve spool
without the necessity for “fluid locking.” Assuming some 65 102 are of such width and spacing so as to completely
leakage of fluid which bypasses the normal circulatory
occlude oil inlet channel 107 and oil return channel 108'
paths provided in the transmission, ‘output shaft 2 will
when aligned 5.? shown in FIGURE 31, but to permit
experience a slippage in speed relative to the speed of
circulation of oil from the pressure source 106 through
input shaft 1, in which case it is obvious that the inner
the control passages and back to the sump when said
housing cannot rotate at synchronous speed with both
valve spool 102 is displaced axially in either direction in
input shaft 1 and output shaft 2. Dynamic forces act
respect to valve sleeve 103‘. When valve spool 102 is
ing on divider disk 15~—-15' of the driving pump and di
moved to the left in FIGURE 31, oil under pressure
vider disk 18—18’ of the driven pump when the two are
from inlet channel 187 is permitted to enter control cham
in asynchronous motion cause the inner housing to rotate
ber B through oil channel 109 and oil is permitted to
at a speed intermediate between the speeds of the input 75 return from control chamber A to return channel 108
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through oil passage 110 as indicated by the solid arrows.
By this means disk support assembly 16—‘16' moves in
the direction of increasing forward pitch or decreasing
reverse pitch causing arm 68 of the ratio indicating link
age to move in the direction of the solid arrow.
This
winding of relay F’. Relay F’ then establishes electrical
connection to motor 111 to cause said motor 111 to
rotate in the direction which will cause valve spool 102
to move to the left viewing FIGURE 31, until the gear
119 and shaft 115 are rotated counter-clockwise by an
action continues until inlet channel 107 and return chan—
amount which will cause the brush arm to move the
nel 108 are realigned with lands 105 and 106 at which
brush to the gap position 120 (FIGURE 32) between
the rings F and R thus opening the electrical circuit
time oil circulation into control chamber B and out of
which action causes the motor to stop. When control
control chamber A is blocked. If valve spool 102 is
moved to the right in FIGURE 31 oil is permitted to 10 action causes the ring R of the selected pair of commu
enter control chamber A under pressure and to exit from
tator-brush rotors to come in contact with the brush, a
closed circuit is established between the positive and nega
control chamber B as indicated by the dotted arrows.
tive poles of the vehicle electrical system via the selector
This causes disk support assembly 16-16’ to move in
the direction of decreasing forward pitch or increasing
switch 118, brush arm, ring R and the winding of relay
reverse pitch causing arm 68 in this case to move in 15 R’. Relay R’ then establishes electrical connection to
motor 111 to cause said motor 111 to rotate in the direc
the direction of the dotted arrow. This action continues
tion which will cause valve spool 102 to move to the
until inlet channel 107 and return channel 108 are again
right viewing FIGURE 31 until the gear 119 and shaft
realigned with lands 105 and 106 at which time oil cir
115 are rotated clockwise by an amount which will cause
culation into control chamber A and out of control
chamber B is blocked. It is therefore apparent that 20 the brush arm to move the brush to the gap position 120
between the rings F and R thus opening the electrical
when valve spool 3102 is moved axially in either direction,
circuit which action causes the motor to stop.
that disk support assembly 16-16’ is caused by hydraulic
means to follow said axial movement of valve spool 102
Since the “Drive,” “Neutral,” and “Low” brush arms
are directly coupled to shaft 115, rotation of said brush
‘by a corresponding angular amount in the direction which
will tend to maintain alignment of lands 105 and 106 25 arms is naturally in the same direction as for shaft 115;
however, inasmuch as the “Reverse” brush arm is cou
with inlet channel 107 and return channel 108.
pled to shaft 115 through reverse gearing, rotation of said
An electric motor 111 is provided to effect axial move
“Reverse” brush arm is in opposition ‘to the rotation of
ment of valve spool 102 through cooperation of the
shaft 115. The reverse gearing of the “Reverse” brush
threaded parts of shaft 112 and spool shaft 113. Spool
shaft 113 should be designed for axial movement only. 30 arm together with the oppositely oriented F and R rings
Motor 111 is controlled electrically by commutator-brush
assembly 114 containing four pairs of commutator-brush
permits the “Reverse” commutation rotor to move counter
clockwise in respect to the “Neutral” reference position
the same as for the “Drive” and “Low” commutation
rotors as shown schematically in FIGURE 32; viewing
rotors and to provide the same transmission input-output
from left to right, one pair for “Drive,” one pair for
“Neutral,” one pair .for “Low” and one pair for “Re 35 ratio control characteristics in the “Reverse” speed as in
the forward “Low” speed. The “Neutral” reference posi
verse.” Essentially the purpose of commutator-brush
tion for the commutation rotors is de?ned by vertical
assembly 114 is to compare the actual input-output speed
orientation of the gap position 120 as shown by the com
ratio of the transmission at any instant with what the
mutation rotor for “Neutral” in FIGURE 32. FIGURE
ratio should be as a function of transmission output shaft
speed, torque demand, and selected speed range and to 40 32 shows the selector switch 118 in the “Low” position
in which a forward input-output speed ratio for the trans
control motor 111 in such a way as to cause said motor
mission is indicated by a counter-clockwise displacement
111 to drive valve spool 102 to the position which will
correct any discrepancy.
of the “Drive,” “Neutral,” and “Low” brush arms in re
spect to the vertical “Neutral” reference. With the com
Shaft 115 is geared at one end to spool shaft 113 and
is directly coupled to the “Drive,” “Neutral,” and “Low” 45 mutation rotors in the same position, it is apparent that
if selector switch 118 was put in the “Reverse” position
brush arms or rotors and is coupled through reverse
that the servo action previously described would cause
gearing (not shown) to the “Reverse” brush arm and
valve spool 102 to move to the position representing a
reverse transmission speed ratio equal to the forward
at one end to control rod 117 and is coupled by suitable 50 speed ratio as indicated in FIGURE 32 and the “Reverse”
brush arm would then be at the same counter-clockwise
means to the “Drive,” “Low,” and “Reverse” commuta
displacement angle in respect to the “Neutral” reference
tion rotors so as to provide angular rotation in direct
proportion to axial movement of said control rod 117
as ‘for the “Drive,” “Neutral” and “Low” brush arms
shown in FIGURE 32 and that the latter brush arms
up to certain predetermined limits for each commutation
rotor. For example, the limit of travel for the “Drive” 55 would then be at the same relative position as shown for
the “Reverse” brush arm in FIGURE 32.
commutation rotor might be adjusted to correspond to
From the foregoing description it will be understood
an input-output speed ratio of one-to-one for the trans
thereby causes angular movement of said brush arms
when spool shaft 113 moves axially. Shaft 116 is geared
that input-output speed ratio control of the transmission
mission, whereas the limit of travel for the “Low” and
in either forward or reverse output speeds is effected by
“Reverse” commutation rotors might arbitrarily be ad
justed to correspond to an input-output speed ratio of 60 control rod 117. As a means of causing said control rod
two-to-one for the transmission.
The “Neutral” com
mutation rotor is ?xed in position corresponding to an
input-output speed ratio of in?nity, or zero output for
117 to move as a function of output shaft speed and torque
demand or input shaft speed, an output speed governor
121, a vacuum modulator 122 and an input speed gov
ernor 123 are provided.
the transmission driving pump.
Governor 121 may be a centrifugally operated device
Selector switch 118 may be designed to permit manual 65
selection of any one of the four switch contacts of the
corresponding pairs of commutator-brush rotors and to
provide electrical contact with the brush arm of the one
‘geared to or otherwise coupled to the transmission out
put shaft 2 and coupled to control rod 117 as shown in
FIGURE 31 to cause said control rod 117 to move axially
selected. Only the selected pair of commutator-brush
to the left, viewing FIGURE 31, as a ‘function of output
rotors may effect control of motor 111. In operation, 70 shaft speed. Vacuum modulator 122 may be coupled to
governor 121 in a suitable manner to oppose the action
‘when control action causes the ring F of the selected
of said governor 121 as a function of engine manifold
pair of commutator-brush rotors to come in contact with
vacuum and thus provide higher input-output transmission
the brush, a closed circuit is established between the posi
ratios with decreasing manifold vacuum or increasingly
tive and negative poles of the vehicle electrical system
via the selector switch 118, brush arm, ring F and the 75 open car-bureator throttle positions. With output shaft 2
73,095,708
13
14
motionless such as would occur when a vehicle is at a
readily be implemented with a hydraulic system than with
standstill, governor shaft 124 may be set at such a posi
conventional mechanical power coupling means. The pos
sibility also exists for combining the braking means with
the propelling means.
output or neutral position when spring 125 is holding con
While there is given above ‘a certain speci?c example,
trol rod 117 toward the right, in FIGURE 31, to the limit 01
of this invention and its application in practical use, it
imposed vby shoulder 126 resting against the lip of
should be understood that this is not intended to be ex
sleeve 12'7.
haustive or to be limiting of the invention. On the con
Governor 123 may be a centrifugally operated device
trary, this illustration and explanation herein are given
geared or otherwise coupled to the transmission input
shaft 1 and coupled to control rod 117 as shown in FIG 10 in order to acquaint others skilled in :the art with this in
tion as to cause the transmission pump to be at the zero
URE 31 to cause said control rod 117 to move axially to
vention and the. principles thereof and a suitable manner
of its application in practical use, so that others skilled
in the art may be enabled to modify the invention and
to adapt and apply it in numerous forms each as may be
as for example the idling speed of a gasoline engine, shaft 15 best suited to the requirement of a particular use.
I claim:
128 will remain at the extreme right hand position, in
1. A variable speed transmission of the hydraulic type
FlGURE 31, but that at increasingly higher speeds of in
comprising a housing having therein a driving pump
put shaft 1, shaft 128- will move to increasingly more left
chamber and a driven pump! chamber; input and output
hand positions, viewing FIGURE 31, causing control rod
117 to move leftwardly with said shaft 128 against the 20 shafts respectively mounted in said chambers; driving and
driven pump assemblies respectively mounted in said
tension of spring 125 until a predetermined maximum
chambers, said ‘assemblies including a rotor connected to
left hand position is reached, for instance corresponding
said input and output shafts and a plurality of radial
to a transmission input-output ratio of four-to4one, at
which point shaft 128 will be held against further left~
vanes mounted on said rotors; a divider disk supporting
wardly movement; however, control rod 117 will not be 25 assembly mounted in said housing and having therein an
restrained to further leftwardly movement resulting from
annular groove, a :divider disk having a plurality of radial
slots therein mounted in said groove and adapted to inter
leftwardly movement of sleeve 127. This arrangement
mesh with said vanes, a plurality of peripheral chambers
will permit smooth starts for vehicles at a standstill, when
on each side of said divider disk in the portion thereof
selector switch 118 is in the “Drive,” “Low,” or “Reverse”
disposed in said annular groove, each of said peripheral
positions by permitting governor 122 to exercise control
chambers being connected by a duct to a corresponding
up to a predetermined limit, for instance corresponding
pump chamber formed between adjacent vanes on the
to a vehicle road speed of ?ve miles per hour, and per
mitting governor 121 in conjunction with a torque-demand
opposite sides of said divider disk, said disk support as
sensing element such as vacuum modulator 122 to effect
sembly being pivotally mounted about a diameter thereof
transmission control at ‘higher road speeds. As described
in said housing about said driving pump assembly; a duct
heretofore, bypass control piston 70‘ will prevent positive
ing core extending from said driving pump chamber to
circulation of oil between the driving pump and driven
said driven pump chamber, said ducting core having at
pump of the transmission when the input-output speed
least a pair of channels out therein to sequentially con
the left viewing FIGURE 31 as a function of input shaft
s'pee . Governor \123 and associated coupling linkage
may be designed such that at a nominal input shaft speed,
nect together corresponding pump chambers of said rotor
output speeds such as would be the case when an automo 40 assembly whereby oil may flow from one to the other;
ratio is above a predetermined value in forward or reverse
tve vehicle is being accelerated from a standstill. I
A control means (not shown) of any conventional form
may be provided for constricting‘ band brake 20 around
drum 21 when either the “Low” or “Reverse” positions
are selected on the selector switch .118.
and control means for varying the pitch of said driving
pump divider disk assembly.
2. A device as described in claim 1 wherein the outer
edge of each vane of said driving pump rotor has a slot
An interlock 45 cut therein and said divider disk has at the center of each
switch may be provided to prevent starting of the engine
of said radial slots a cooperating pin adapted to engage
except when disk support assembly 16—16’ is at the zero
in said slot, said slot being contoured to maintain said
pitch or zero output position.
disk in proper phase relationship to said rotor.
The control system herein described is suited for auto
3. A variable speed transmission of'the hydraulic type
matic operation of the transmission for automotive ap 50 comprising a housing having therein a driving pump
plication without any additional ?uid coupling interposed
chamber and a driven pump chamber; input and output
in the power train. The design of the transmission and
shafts respectively mounted in said chambers; driving and
control system for use with a slippage type of fluid cou
driven pump assemblies respectively mounted in said
pling interposed between input shaft 1 and the engine
chambers, said assemblies including a rotor connected
crankshaft could dilfer in several details.
to said input and output shafts, and a plurality of v-anes
An obvious constructional variation in the illustrated
mounted on said rotor; a divider disk supporting assem
embodiment would be the physical separation of the driv
bly mounted in said housing and having therein an annu
ing pump and driven pump with each having its own
lar .groove; a divider disk mounted in said groove; a plu
housing and hydraulically connected by lengths of tubing
rality of radial slots cut in said disk and adapted to
or conduit. Such a variation would be equivalent to cut
ting the transmission at line X-X of’FIGURE 1 into
two parts, each part then being adapted to make appro
priate connections with the interconnecting high pressure
and return tubes or conduits and removing the outer hous
60 intermesh with said vanes mounted ‘on said rotor; said
disk supporting assembly forming with said rotor and disk
a plurality of pump chambers on each side of said disk;
a plurality of peripheral chambers on each side of said
divider disk in the portion thereof disposed in said annu
ing. For instance, by locating the driving pump near the 65 lar groove, each of said peripheral chambers being con
engine and the driven pump near the rear axle of an
automobile, the need for the conventional drive shaft and
associated universal joints would be eliminated thus mak
ing possible a ?at ?oor design for automobiles. By pro
nected by a duct to a corresponding pump chamber; said
disk support assembly in said driving pump assembly
being pivotal-1y mounted in said housing about a diameter
thereof disposed at right ‘angles to the axis of said hous
viding two ‘driven pumps or one for each rear wheel, the 70 ing, a ducting core extending from said driving pump
conventional dilferetnial gearing‘could be also eliminated.
chamber to said driven pump chamber; a plurality of
A further possibility would be to integrally combine the
driven pumps with the wheels.
C-onstructional variations such as four wheel ‘drive and
front wheeldrive in lieu of rear wheel drive could more
channels cut in said ducting core to sequentially connect
together‘ corresponding pump chambers of said rotor as
semblies whereby oil may flow from one to the other;
control means for varying the pitch of said driving pump
3,095,708
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divider-disk assembly; an outer casing surrounding said
assembly on at least one side thereof, duct means for
housing, driving and driven‘ pump assemblies; bearing
introducing and withdrawing oil from said chambers
formed between said housing and said supporting disk
assembly whereby the pitch of said assembly may be con
means within said outer casing to permit rotation of said
driving and driven pump assemblies therein; overrunning
housing to selectively prevent forward rotation of said
trolled, a ducting core positioned within the shafts of
said rotor assemblies and extending from said driving
pump chamber to said driven pump chamber, said duct
ing core having at least a pair of channels cut therein to
housing.
sequentially connect together corresponding pump cham
4. A variable speed transmission of the hydraulic type
comprising a housing having therein a driving pump
chamber and a driven pump chamber; input and output
shafts respectively mounted in said chambers; driving and
driven pump assemblies respectively mounted in said
chambers, said assemblies including a rotor having six
radial vanes, connected to said input and output shafts,
and a ‘divider disk having six radial slots therein adapted
bers of said rotor assemblies whereby oil may ?ow from
one to the other; control means for varying the pitch of
to intermesh with said vanes mounted on said rotor; a
pump assemblies, bearing means mounted in said outer
clutch means mounted between said outer casing and said
housing to prevent reverse rotation thereof and hand
brake means connected between said outer casing and
said driving pump divider disk assembly; angle indicat
ing means geared to said driving pump disk supporting
assembly and extending through said ducting core mem
ber to a movable collar on said output shaft, lever means
interconnecting said collar to said control means; an
outer casing surrounding said housing, driving and driven
casing having said driving and driven pump assemblies
divider disk supporting assembly having therein an annu
lar groove adapted to receive the periphery of said divider 20 journalled therein‘; overrunning clutch means mounted
between said outer casing and said housing to prevent
disk, said disk supporting assembly forming about said
reverse rotation thereof and band brake means connected
rotor and disk an inner chamber Within said housing,
between said housing and inner chamber to selectively
having twelve pump chambers; six peripheral chambers
prevent forward rotation of said housing.
on each side of said divider disk in the portion thereof
6. The device of claim 5 wherein‘ said angle indicating
disposed in said support-ing assembly annular groove,
means includes a sleeve, an annular channel in said
each of said peripheral chambers being connected by a
sleeve, a plurality of ori?ces in the high and low pres
duct to a pump chamber on the opposite side of said di
sure channels of said ducting core, said sleeve being posi
vider disk; said driving pump assembly also having said
tioned about said ori?ces when said driving pump divider
disk support assembly pivotally mounted in said housing
about a diameter thereof, a ducting core positioned with 30 disk assembly is near the zero pitch position to at least
partially interconnect through said channel said ori?ces
in said driving rotor assembly and extending from said
‘whereby a slight bypassing of ?uid is obtained to prevent
driving pump chamber to said driven pump chamber, said
stalling of the prime power source when starting the
‘ducting core having at least a pair of channels out there
driven pump under heavy loads.
in to sequentially connect together corresponding pump
7. A device as described in claim 5 wherein said duct
chambers of said rotor assemblies whereby oil may flow 35
ing core has therein a valve and port assembly intercon
from one to the other; control means for varying the
necting opposite pressure channels thereof to the center
pitch of said driving pump divider disk assembly; an
of said ducting core, said valve mechanism being arranged
outer casing surrounding said housing, driving and driven
to close oif the port leading to the high pressure side of
pump assemblies; bearing means ?xed in said outer cas
ing and having journaled therein said driving and driven 40 said ducting core at any given time whereby oil may be
added to the low pressure side of said ducting core
pump assemblies; overrunning clutch means operatively
through the center thereof.
mounted between said outer casing and said housing to
8. A variable speed transmission of the hydraulic type
prevent rotation thereof in one direction and hand brake
comprising a housing having therein‘ a driving pump
means operatively connected between said casing and
housing to selectively prevent rotation of said housing 45 chamber and a driven pump chamber; input and output
shafts respectively mounted in said chambers; driving
in the other direction.
and driven pump assemblies respectively mounted in said
5. A variable speed transmission of the hydraulic type
chambers, said ‘assemblies including a rotor connected to
comprising a housing having therein a driving pump
said input and output shafts and a pluarlity of radial
chamber and a driven pump chamber; input and output
shafts mounted respectively in said chambers; driving and 50 vanes mounted on said rotors; a divider disk supporting
‘assembly mounted in said housing and having therein
driven pump assemblies respectively mounted in said
chambers, said assemblies comprising a rotor having six
an annular groove, a divider disk having a plurality of
radial vanes mounted about and connected to said input
radial slots therein mounted in said groove and adapted
to intermesh with said vanes, said disk support assembly
being pivotally mounted about a diameter thereof in
said housing about said driving pump assembly; a duct
ing core extending from said driving pump chamber to
said idriven pump chamber, said ducting core having
and output shafts, a divider disk having six radial slots
therein adapted to intermesh with said vanes mounted
on said rotor, a divider disk supporting assembly mounted
in said housing and having therein an annular groove
adapted to receive the periphery of said divider disk,
said disk supporting assembly forming about said rotor
at least a pair of channels cut therein to sequentially con
and disk an inner chamber Within said housing, s-ix pe 60 nect together corresponding pump chambers of said rotor
assembly whereby oil may flow from one to the other;
ripheral chambers formed on each side of said divider
disk in the portion thereof disposed in said supporting
assembly annular groove, each of said peripheral cham
and control means for varying the pitch of said driving
pump divider disk assembly.
9. In a hydraulic variable displacement transmission
bers being connected by a duct and valve mechanism to
a corresponding pump chamber formed between adja 65 of the type having a driving pump connected to a prime
power source and a driven pump connected to the source
cent vanes on the opposite side of said divider disk; said
to be moved, control means comprising a pitch control
disk support assembly in said driving pump assembly
mechanism connected to said driving pump means, said
being pivotally mounted in said housing about a diam
eter thereof disposed at right angles to the axis of said 70 pitch control mechanism having forward, neutral, and
reverse positions; a feed back loop for indicating the
pitch of said driving pump in said pitch control mecha
nism; motor means operatively connected to said pitch
control for moving said pitch control to the desired po
said support assembly and said shaft to form oil sealing
chambers above and below the axis of said support disk 75 sition; a commutator brush assembly having a plurality
housing being contoured on‘ the outer surface to form an
oil sealing contact with the inner surface of said hous
ing; a pair of vanes mounted in the plane of the axis of
an
3,095,708
17
of commutation stators and rotors; a load demand sens
ing element connected ‘to said commutator brush assem
bly; switch means for selectively connecting to the de
sired commutation rotor; battery means connected
through said switch means to said commutation‘ rotors
and load demand sensing elements connected to said
commutator brush assembly whereby variations in load
demand or switch control will cause said pitch control
mechanism to vary the pitch of said driving pump so
that the desired output will be delivered to the output 10
shaft.
References Cited in the ?le of this patent
UNITED STATES PATENTS
951,064
Erickson ____________ __ Mar. 1, 1910
1
18
1,020,271
2,242,058
2,318,386
2,323,926
2,371,228
2,431,122
2,443,074
2,691,349‘
2,808,006
Erickson ____________ __ Mar. 12,
2,828,695
Marshall _____________ __ Apr. 1,
1,123,013
1,009,488
France _______________ __ June 4,
Cuny _______________ __ May 13,
Haines _______________ __ May 4,
McGill ______________ __ July 13,
Dodge ______________ __ Mar. 13,
Jakobsen ____________ __ Nov. 18,
Kraft ________________ __ June 8,
Cuny ________________ __ Oct. 12,
Paulsrn-eier ___________ __ Oct. 12,
1912
A1941
1943
1943
1945
1947
1948
1954
1957
1958
FOREIGN PATENTS
Germany ____________ __ May 29,
1956
1957
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