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Inventors
Folsom, Lawrence R.
Tucker, Clive
Application #
480772
Filed
Jun-7-1995
Published
Aug-12-1997
Current US Class
060/487
060/491
International Classes
F16D 039/00
Field of Search
60/490 60/491 60/492 60/487
Assignee
Folsom Technologies, Inc. (Pittsfield, MA)
Examiners
Nguyen; Hoang
Attorney, Agent or Firm
Neary; J. Michael
US Patent References
| 4109466 |
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Hydraulic transmis... |
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| 4196586 |
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Multi-mode hydrost... |
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| 4646521 |
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Hydroversion |
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| 5243822 |
|
Hydraulic rotary p... |
|
Referenced by:
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Citation
Cite This Patent
More From Subclass 487
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Abstract
A continuously variable transmission includes a housing in which input and output shafts are journaled for rotation. Both input and output shafts have torque couplings for connection to drive shafts of a prime mover and a driven load such as the drive axle of a vehicle. The input shaft is coupled to a vane rotor of a vane-type pump having radially extending pump vanes driven by the pump rotor around a pump working volume between two axially spaced pump side plates surfaces and within an inside surface of a flexible pump cam ring spaced radially from the pump rotor. The pump vanes each have a radial end edge remote from the pump rotor and in contact with the pump cam surface. A stator hub is fixed to the housing and has radially extending stator vanes projecting radially into a stator working volume between two axially spaced stator side plate surfaces and between a cam surface of a flexible stator cam ring and the radially spaced surface of the stator hub. The stator vanes each have a radial end edge remote from the hub and in contact with the stator cam surface. Torque is transmitted from the stator cam ring and the pump cam ring to the output shaft by a torque coupling. Fluid openings are provided for passing fluid pressurized in the pump directly from the pump working volume into the stator, and for passing fluid displaced from the stator directly from the stator working volume into suction sectors of the pump. A cam ring adjustment mechanism is mounted in the housing for changing the radial spacing of portions of the pump cam ring from the pump rotor, and for changing the radial spacing of portions of the stator cam ring from the stator hub. Rotation of the pump rotor by the input shaft pressurizes fluid in the pump working volume which acts against the pump cam ring to generate torque that is transmitted to the output shaft, and the pressurized fluid in the pump working volume communicates through the fluid openings to pressurize fluid in the stator working volume which acts against the stator cam ring to generate additional torque that is transmitted to the output shaft, and the cam adjustment mechanism is operated to reciprocally change the displacement of the pump and the stator to change the transmission ratio of the transmission.
Claims
We claim:
1. A continuously variable transmission, comprising:
a housing;
an input shaft journaled for rotation in said housing and having a torque coupling for connection to an drive shaft of a prime mover;
an output shaft journaled for rotation in said housing and having a torque coupling for connection to a driven shaft;
a vane-type pump having a pump vane rotor coupled to said input shaft and having radially extending pump vanes driven by said pump rotor around a pump working volume defined between two axially spaced pump side plates surfaces and an inside surface of a flexible pump cam ring spaced radially from said pump rotor, said pump vanes each having a radial end edge remote from said pump rotor and in contact with said pump cam surface;
a torque transmitting connection between said pump cam ring and said output shaft;
a stator hub fixed relative to said housing and having radially extending stator vanes projecting radially into a stator working volume defined between two axially spaced stator side plate surfaces and a cam surface of a flexible stator cam ring radially spaced from said stator hub, said stator vanes each having a radial end edge remote from said hub and in contact with said stator cam surface;
a torque transmitting connection between said stator cam ring and said output shaft;
fluid openings for passing fluid pressurized in said pump directly from said pump working volume into said stator, and for passing fluid displaced from said stator directly from said stator working volume into suction sectors of said pump;
a cam ring adjustment mechanism for changing the radial spacing of portions of said pump cam ring from said pump rotor, and for changing the radial spacing of portions of said stator cam ring from said stator hub, said cam ring adjustment mechanism being independent of said torque transmitting connection between said cam rings and said output shaft and remaining unstressed by output torque during operation of said transmission;
whereby rotation of said pump rotor by said input shaft pressurizes fluid in said pump working volume which acts against said pump cam ring to generate torque that is transmitted to said output shaft, and said pressurized fluid in said pump working volume communicates through said fluid openings to pressurize fluid in said stator working volume which acts against said stator cam ring to generate additional torque that is transmitted to said output shaft, and said cam adjustment mechanism is operated to reciprocally change the displacement of said pump and said stator to change the transmission ratio of said transmission.
2. A continuously variable transmission as defined in claim 1, wherein:
said pump and stator are axially juxtaposed with a flat port plate axially interposed therebetween; and
said fluid openings are in said port plate and are symmetrically arranged therein about a central axis of said port plate.
3. A continuously variable transmission, comprising:
a housing;
an input shaft journaled for rotation in said housing and having a torque coupling for connection to an output drive shaft of a prime mover;
an output shaft journaled for rotation in said housing and having a torque coupling for connection to an output drive shaft;
a vane-type pump having a vane rotor coupled to said input shaft and having radially extending pump vanes driven by said pump rotor around a pump working volume defined between two axially spaced pump side plates surfaces and an inside surface of a cam ring spaced radially from said pump rotor, said pump vanes each having a radial end edge remote from said pump rotor and in contact with said pump cam surface;
a torque transmitting connection between said pump cam ring and said output shaft;
a stator hub fixed relative to said housing and having radially extending stator vanes projecting radially into a stator working volume defined between two axially spaced stator side plate surfaces and a stator cam ring surface radially spaced from said stator hub, said stator vanes each having a radial end edge remote from said hub and in contact with said stator cam surface;
a torque transmitting connection between said stator cam ring and said output shaft;
said pump and stator are axially juxtaposed with a flat port plate axially interposed therebetween;
said torque transmitting connections include radially projecting tenons torsionally coupling said cam rings operatively to said output shaft, and against which said pressurized fluid acts to generate said torque,
fluid openings in said port plate for passing fluid pressurized in said pump directly into said stator, and for passing fluid displaced from said stator directly into suction sectors of said pump;
a cam adjustment mechanism for changing the radial spacing of portions of said pump cam ring from said pump rotor, and for changing the radial spacing of portions of said stator cam ring from said stator hub;
whereby rotation of said pump rotor by said input shaft pressurizes fluid in said pump working volume which acts against said pump cam ring to generate torque that is transmitted to said output shaft, and said pressurized fluid in said pump working volume communicates through said fluid openings to pressurize fluid in said stator working volume which acts against said stator cam ring to generate additional torque that is transmitted to said output shaft, and said cam adjustment mechanism is operated to reciprocally change the displacement of said pump and said stator to change the transmission ratio of said transmission.
4. A continuously variable transmission as defined in claim 3, wherein:
said tenons are received in radially opening channels in an output rotor mechanically connected to said output shaft.
5. A continuously variable transmission as defined in claim 1, wherein:
said pump and stator are radially juxtaposed with a common cam ring radially interposed therebetween.
6. A continuously variable transmission as defined in claim 5, wherein:
said common cam ring has an external surface and an inner surface, said external surface has a cross-sectional profile that is elliptical when a cross-sectional profile of said inner surface is circular, and vise-versa.
7. A continuously variable transmission, comprising:
a housing;
an input shaft having a torque coupling for connection to a drive shaft and journaled for rotation in said housing;
an output shaft having a torque coupling at one end for connection to a driven shaft and journaled for rotation in said housing;
a pump having a pump rotor mounted in said housing for rotation and coupled to the other end of said input shaft for torque transmission therefrom;
said pump having a flexible cam ring having a smooth continuous surface facing said pump rotor and defining between said continuous surface and said pump rotor variable working volumes therein within which fluid therein can be pressurized by rotation of said pump rotor;
structure in said pump against which said pressurized fluid can act to generate torque;
a motor having structure defining a working volume in which pressurized fluid from said pump can act to produce torque;
passages in said pump structure and said motor structure for conveying pressurized fluid from said pump working volume directly to said motor working volumes;
suction passages in said pump structure and said motor structure for conveying unpressurized fluid from said motor working volume directly back to said pump working volumes; and
output torque coupling structure for coupling said pump and said motor to said output shaft for transmission of torque generated in said pump and said motor to said output shaft;
an adjustment mechanism for flexing said cam ring and thereby changing said variable working volumes therein, said adjustment mechanism being independent of said output torque coupling structure.
8. A continuously variable transmission, comprising:
a housing;
an input shaft having a torque coupling for connection to a drive shaft and journaled for rotation in said housing;
an output shaft having a torque coupling at one end for connection to a driven shaft and journaled for rotation in said housing;
a pump having a pump rotor mounted in said housing for rotation and coupled to the other end of said input shaft for torque transmission therefrom;
said pump having a flexible cam ring concentrically arranged with respect to said pump rotor defining variable working volumes therein within which fluid therein can be pressurized by rotation of said pump rotor and against which said pressurized fluid can act to generate torque;
a motor having a flexible cam ring concentrically arranged with respect to a motor stator and defining a working volume in which pressurized fluid from said pump can act to produce torque;
a cam ring adjustment mechanism spanning said cam rings and supported on both axial sides thereof and having bearing surfaces that engage said cam rings and, when actuated, change the shape of said cam rings to change the shape of said variable working volumes to change the transmission ratio of said transmission;
passages in said pump structure and said motor structure for conveying pressurized fluid from said pump working volume directly to said motor working volumes;
suction passages in said pump structure and said motor structure for conveying unpressurized fluid from said motor working volume directly back to said pump working volumes;
structure for coupling said pump and said motor to said output shaft for transmission of torque generated in said pump and said motor to said output shaft.
Description
BACKGROUND OF THE INVENTION
This application relates to a continuously variable hydraulic pump and motor, and to an infinitely variable transmission made by a mechanical and fluid coupling of the pump and motor, and more particularly to a vane type pump and motor of continuously variable displacement and a mechanical and fluid connection of the vane type pump and motor in a small, inexpensive, light weight infinitely variable transmission.
Many variable speed drive mechanisms of various designs are described in the literature or are commercially available. These mechanisms find application in fields as diverse as computers, machine tools, recreational vehicles, construction equipment, trucks and automobiles. They all share the basic function of converting the rotational speed and torque of an input shaft to a selected variable speed and torque at an output shaft.
The motor vehicle is an ideal application for an infinitely variable speed drive mechanism because of the improved economy and low exhaust emissions that can be obtained by operating the vehicle's prime mover, such as an internal combustion engine, in the range of its optimum operating point. Moreover, the potential market is enormous: it has been estimated that the annual world-wide market for automotive transmissions in the 15 years before 2005 will be in the vicinity of 47 million vehicles. Use of an efficient infinitely variable transmission in that number of vehicles would save an inestimable amount of fuel and reduce the worldwide exhaust emissions more than any other known conservation and air purity stratagem now considered feasible.
Although many infinitely variable transmissions and continuously variable transmissions have been proposed and designed for automotive application, none has proven entirely satisfactory. Traction devices have been unable to demonstrate acceptable life at the power levels required and the transient torque conditions occurring in a normal automotive driving cycle. Rubber belt variator devices, similar to snowmobile transmissions, have found limited application in special automotive areas, such as some mini cars, but the durability and efficiency of these devices are marginal at best, even in light vehicles with engine power on the order of only ten to fifteen horsepower. The most common application for this type of transmission is the snowmobile, where component life is not expected to exceed 100 hours.
Hydrostatic transmissions have existed for years and have been developed to a high degree of sophistication. These devices are in use in some military, agriculture and construction equipment, mining and other off-the-road vehicles, and in small garden tractors. A conventional hydrostatic transmission has two principal elements: a hydraulic pump driven by the prime mover, and a hydraulic motor powered by hydraulic fluid pressurized by the pump for driving the load. Either or both of these elements may be variable displacement to achieve the variable gear ratio of the transmission. Regardless of the configuration selected, the overall system efficiency can be no better than the product of the efficiencies of the individual elements. For example, if both the pump and motor are 95% efficient, the hydrostatic unit cannot achieve efficiency greater than (0.95.times.0.95)=90% and in practice it is usually significantly less than this because of flow losses in the hydraulic lines coupling the two elements. This efficiency is inferior to that offered by conventional automatic transmissions which can operate at steady state efficiency levels on the order of 97%-98% with torque converter lock-up, but the advantages of an infinitely variable transmission and the absence of a clutch outweigh the disadvantage of low efficiency in the applications in which conventional hydrostatic transmissions have been used successfully.
In addition to their low operating efficiencies, there are other disadvantages that have militated against the wide use of conventional hydrostatic transmissions. They are usually bulky, heavy and expensive. In addition, conventional hydrostatic transmissions are noisy, especially at the higher gear ratios where most over-the-road driving is done because the flow rate of the hydraulic fluid is greatest at the high gear ratios in these hydrostatic transmissions.
The integrated hydrostatic transmission, in which the motor and pump are combined in one unit to minimize fluid flow losses, is a step in the right direction. However, none of the prior art integrated hydrostatic transmissions overcome the condition which degrades their efficiency and contributes to their noisiness, namely, that the peak power rating of the transmission is attained at maximum pressure and flow. As a consequence, hydraulic losses associated with pressure, such as leakage and hysterisis losses during fluid compression and expansion will be greatest at maximum power throughput. Also, viscous flow losses which are proportional to fluid velocities are greatest at peak power/speed when the flow and pressure are at their highest levels.
The lack of enduring commercial uses of hydrostatic transmissions in production for automotive or other uses that require a high power-to-weight ratio is believed to be due to four main reasons: 1) high cost, 2) high noise levels at normal operating conditions, (3) poor efficiency, and (4) lack of any significant weight and size advantage. However, modern production techniques have been developed that would make it possible to produce a hydrostatic transmission designed specifically for such applications at a cost approximately comparable to that of a prior art adjustable ratio variable transmission. The second and third factors, namely, noise and efficiency, have been the key factors discouraging adoption of a hydrostatic transmission by the automotive and recreational vehicle industries. The size and weight factors could be significant if there were competing designs that satisfied the first three factors.
One effort to overcome some of the disadvantages of the conventional hydrostatic transmission is the power branching transmission. An early example of such a transmission is shown in U.S. Pat. No. 3,175,363 to Hans Molly. The power branching transmission was intended to reduce the fluid flow losses associated with the hydrostatic transmission, particularly as the transmission ratio moves toward unity, by transmitting a portion of the input power mechanically to the output shaft. Since the proportion of mechanically transmitted power increases to 100% at a 1:1 transmission ratio, the hydraulic losses are potentially much less in a power branching transmission.
Unfortunately, attempts to commercialize the power branching hydrostatic transmission have been unsuccessful, probably because the complexity of the system would compromise performance and increase cost to a noncompetitive level versus the conventional transmission. Also, the prior art power branching transmissions have not been able to achieve a dynamic balance of the rotating elements which would be a serious shortcoming since substantial vibration levels at operating speed would not be acceptable. In addition, prior art power branching transmissions have not been readily scalable to make different sizes of a single design usable for different power ranges. Scalability could be an important feature in smaller applications such as snowmobiles and motorcycles where the ability to match the size, weight and cost of the transmission precisely with the power, torque and speed requirements could become competitively important.
If the power available in operation of a vehicle during braking and periods of low power requirements could be stored and made available for use during periods of auxiliary or high power requirements such as hydraulic power take-off, engine starting, and vehicle acceleration, the engine sizing for any given vehicle could be reduced substantially, since engines are normally sized for the maximum anticipated power requirements. The storage of hydraulic energy in an accumulator is a well known and understood technology and should encounter no resistance to use in motor vehicle applications as some new technologies have in the past, and the use of a moderately sized accumulator will add little weight and cost, certainly less than is saved by the use of a small light weight compact vane-type continuously variable transmission that makes possible the elimination of a clutch and a starter motor and makes possible the substantial downsizing of the engine because of the availability of the added hydraulic power source.
Thus, the transmission art has long needed an improved infinitely variable hydrostatic transmission that provides the advantages of the integrated hydrostatic transmission while markedly improving efficiency by reducing the hydraulic fluid losses associated with conventional hydrostatic transmissions, reducing the size, weight, cost, emissions and noise levels of operation, improving the performance near the neutral point, offering scalability of some basic machine designs, and reducing the manufacturing and maintenance costs.
SUMMARY OF INVENTION
Accordingly, it is an object of this invention to provide an improved dynamically balanced power branching transmission in which fluid losses are reduced to near zero at maximum power throughput for steady state operating conditions, resulting in overall operation efficiencies comparable or superior to conventional automatic transmissions, but which enables the engine to maintain an operating level at or near its optimum operating point. Another object of this invention is to provide an improved hydrostatic transmission that is small, light weight, quiet, durable, inexpensive to produce and service, and offers system efficiency in a motor vehicle that is comparable or superior to conventional automatic transmissions. Yet another object of this invention is to provide an improved method of converting power in a rotating input shaft at one speed and torque to nearly the same power in an output shaft at a different speed and torque.
These and other objects of the invention are attained in a power branching transmission which has fluid openings for flow of fluid directly between juxtaposed vane type pump and motor units, and an adjustable cam ring concentrically disposed in relation to each pump and motor unit. The cam rings are coupled to and drive an output shaft. The cam rings are driven by the fluid pressure generated in the pump to provide an infinitely variable transmission ratio for the unit in an efficient and simple manner. The transmission provides the capability of regenerative braking and accelerating, as well as engine starting using energy stored from the transmission.
DESCRIPTION OF THE DRAWINGS
The invention and its many attendant objects and advantages will be better understood on reading the following description of the preferred embodiment in conjunction with the following drawings, wherein:
FIG. 1 is a side sectional elevation of a vane-type hydrostatic transmission in accordance with this invention, shown adjusted to produce a transmission ratio of 1:1;
FIG. 2 is an axial sectional elevation of the pump section of the transmission shown in FIG. 1 along lines 2--2;
FIG. 3 is an axial sectional elevation of the stator section of the transmission shown in FIG. 1 along lines 3--3;
FIG. 4 is a side elevation of an end cap and wear plate in the transmission shown in FIG. 1;
FIG. 5 is a side elevation of the port plate between the pump and stator units in the transmission shown in FIG. 1;
FIG. 6 is a schematic view of the control system for controlling the transmission ratio of the transmission shown in FIG. 1;
FIG. 7 is a side elevation of a second embodiment of a tandem vane-type transmission, shown adjusted to its neutral position;
FIG. 8 is a side elevation of the transmission shown in FIG. 7, shown adjusted to its 1:1 transmission ratio position;
FIG. 9 is an elevation along lines 9--9 in FIG. 11;
FIG. 10 is a sectional elevation of the transmission shown in FIG. 7 along lines 10--10 in FIG. 7;
FIG. 11 is a sectional elevation of the transmission shown in FIG. 7 along lines 11--11 in FIG. 8;
FIG. 12 is a sectional elevation of the transmission shown in FIG. 7 along lines 12--12 in FIG. 7;
FIG. 13 is a sectional elevation of the transmission shown in FIG. 7 along lines 13--13 in FIG. 8;
FIG. 14 is a sectional elevation of the transmission shown in FIG. 7 along lines 14--14 in FIG. 7;
FIG. 15 is a sectional elevation of the transmission shown in FIG. 7 along lines 15--15 in FIG. 7;
FIG. 16 is a sectional elevation of the transmission shown in FIG. 7 along lines 16--16 in FIG. 7;
FIG. 17 is a sectional elevation of a third embodiment of the inventive transmission having vane type pump and stator units arranged concentrically;
FIG. 18 is a sectional elevation of the concentrically designed vane-type transmission shown in FIG. 17 along lines 18--18 in FIG. 17
FIGS. 19 and 20 are sectional elevations of the two halves of the housing of the embodiment shown in FIG. 17;
FIG. 21 is an end elevation of the housing half shown in FIG. 20;
FIG. 22 is a sectional elevation of the output shaft in the embodiment of FIG. 17, along lines 22--22 in FIG. 23;
FIG. 23 is an end elevation of the output shaft shown in FIG. 22;
FIG. 24 is an end elevation of the inside face of the end plate on the output shaft of the embodiment shown in FIG. 17;
FIG. 25 is a sectional side elevation of the output shaft in the embodiment of FIG. 17, along lines 25--25 in FIG. 24;
FIG. 26 is an end elevation of the output shaft of the embodiment shown in FIG. 17 along lines 26--26 in FIG. 25;
FIG. 27 is a side elevation of a speed sensor mounted in the housing adjacent the output shaft end plate shown in FIG. 25;
FIG. 28 is an end elevation of the inside face of the input-side end plate shown in FIG. 17;
FIGS. 29 and 30 are sectional elevations along lines 29--29 and 30--30 in FIG. 28;
FIG. 31 is an end elevation of the input-side end plate shown in FIG. 28, with the wear plate attached;
FIG. 32 is a sectional elevation along lines 32--32 in FIG. 31;
FIG. 33 is a side elevation of the cam ring shown in the embodiment of FIG. 17;
FIGS. 34, 35 and 36 are side views of the cam ring along lines 34--34, 35--35 and 36--36;
FIGS. 37 and 38 are side elevations of the cam ring shown in FIG. 33 in unstrained and deformed conditions, respectively;
FIGS. 39 and 40 are side elevations of the top control rod in the 1:1 and neutral positions, respectively;
FIG. 41 is a side elevation of an array of torque arms used in the embodiment shown in FIG. 17;
FIG. 42 is a view along lines 42--42 in FIG. 41.
FIG. 43 is a partial sectional elevation through the torque arm along the lines 43--43 in FIG. 18;
FIG. 44 is a partial sectional elevation through the check valve along the lines 44--44 in FIG. 18;
FIG. 45 is a partial sectional elevation through the check valve along the lines 45--45 in FIG. 44;
FIG. 46 is a sectional elevation of the check valve seat bushing shown installed in FIG. 44;
FIG. 47 is an end view of the seat bushing shown in FIG. 46
FIG. 48 is an end elevation of the wear plate on the input-side end plate shown in FIG. 17;
FIGS. 49 and 50 are elevations along lines 49--49 and 50--50 in FIG. 48;
FIG. 51 is an enlarged sectional elevation of the socket in the wear plate for one of the check valves;
FIG. 52 is an end view of the socket shown in FIG. 51;
FIGS. 53-56 are schematic diagrams of a hydraulic circuit used in all three embodiments to obtain regenerative braking, regenerative acceleration, engine starting, and auxiliary hydraulic power for take-off applications.
DESCRIPTION OF THE PREFERRED EMBODIMENT
Turning now to the drawings, wherein like reference numerals designate identical or corresponding elements, and more particularly to FIG. 1 thereof, a first embodiment of an infinitely variable vane-type hydrostatic transmission is shown having two tandem hydraulic vane-type units, a pump unit 20 and an axially juxtaposed stator unit 22, mounted on either side of a flat port plate 24 in a housing 30. The pump unit 20 is driven by an input shaft 26 which in turn is driven by a drive shaft 27 of a prime mover 28 mounted in an apparatus such as a machine tool, snowmobile, marine vessel, automobile, motorcycle, etc.
Fluid pressurized by the pump unit 20 acts on a pump unit cam ring 32 keyed to a pump output rotor 34, and also communicates directly through the port plate 24 to act on a stator unit cam ring 36, in turn keyed to a stator output rotor 35. The output rotors 34 and 35 are bolted together to form a rigid back-up sleeve of a rotating assembly 37 which is connected to an output shaft 38 for coupling torque exerted on the output rotors 34 and 35 to the output shaft 38. A control mechanism 40 is coupled to the pump cam ring 32 and the stator cam ring 36 to change the cross-sectional shape of the cam rings 32 and 36 and thereby vary the transmission ratio of input speed and torque to output speed and torque.
The prime mover 28 is typically an internal combustion engine, such as a diesel, two-cycle, four-cycle, rotary or turbine engine, which operates most cleanly and efficiently within a narrow band of engine speed. The continuously variable transmission of this invention enables the transmission of power to driven applications at infinite gradations of speed from zero to 1:1 (the input speed to the transmission) while the prime mover operates continuously at its optimum operating point.
A pump rotor 42 is mounted for rotation on a spline section 44 of the input shaft 26 for driven rotation thereby. Ten radially projecting pump vanes 46 are mounted on the pump rotor 42 for rotation therewith and for limited radial sliding movement to enable the vanes 46 to follow the inner surface 48 of the pump cam ring 32. The number of pump vanes 46 will vary, depending on the power capacity of the pump unit 20. Typically, for a transmission designed to operate at a moderate system pressure of 2000 PSI and a power capacity of 100 HP, a pump rotor with 10 vanes would be used.
The housing 30 includes a cylindrical body 50 which may taper slightly for manufacturing convenience. The body 50 is provided with bosses 52 at an input end which are drilled and tapped for receiving screws 53 for attaching a mounting flange 54 to the housing 30, which closes and seals the flange end opening of the housing 30. The transmission is connected to the apparatus in which it is to be used by attaching the mounting flange 54 to the prime mover 28 or other structure of the apparatus by way of holes (not shown) or other conventional attachment means on the flange 54.
The opposite end, the output end, of the housing 30 is necked down at 56 and terminates in a smaller diameter barrel 58 having an axial cylindrical opening 60 in which is mounted a conical roller element bearing 62 in which the output shaft is journaled for rotation. The housing 30 may be made of cast, welded or superplastically formed aluminum, cast iron, or other low cost material, since it does not carry any significant structural loads.
The housing body 50 is closed at its output end barrel 58 with an annular end cap 64 that has an axial opening for receiving the output shaft 38. A seal 66 in the end cap 64 outboard of the bearing 62 prevents leakage of hydraulic fluid out of the housing 30 around the output shaft 38. A wave spring 68 is provided between the end cap and the bearing 62 to preload the rotating assembly against the rotating seal interface at the input end.
The stator unit includes a stator hub 70 mounted on splines at the distal end of a ground shaft 72, which has an end flange 73 that is attached to the mounting flange 54, so that the stator hub 70 is fixed relative to the housing 30. A plurality of stator vanes 74, shown in FIG. 3 as ten vanes 74, are mounted in radial slots in the stator hub 70 for radial sliding movement to accommodate the varying radial distance to the inside surface 76 of stator cam ring 36.
The rotating assembly 37 includes a, pair of structurally similar end caps 78 and 80, one on each axial end of the tandem pump and stator units 20 and 22, respectively. The end caps 78 and 80 are coupled axially by eight circumferentially spaced tie bolts 82 (only one of which is seen in FIG. 1) which extend through axial holes 83 in the pump side end cap 78, through aligned holes 84 in the pump output rotor 34 and holes 85 in the stator output rotor 35, and also through aligned holes 86 in the port plate 24. The bolts 82 are threaded into tapped holes 88 in the stator side end cap 80 to secure the end caps axially and bear the force exerted by fluid pressure within the rotating assembly 37. The housing is thus not subjected to the stresses incident to containment of high pressure fluid within the pump and stator units, so the housing can be made as a relatively light weight and inexpensive structure
An annular control housing 90 has an inner flange 91 that is fastened to a thick section 92 of the stator side end cap 80 by screws 94. The control housing 90 has an axial lip 96 which overlaps a step 98 formed in the end cap 80 at the radial outer extent of the thick section 92, at its junction with a radially outer stepped section 99, for accurate placement and retention of the control housing on the end cap 80.
The control housing has a cylindrical, axially extending intermediate section 100 having a cylindrical outer surface 102 that is accurately machined to function as the inner surface of an annular control cylinder 103 for the control mechanism 40. The control housing terminates in a radial end flange 104 having a rabbet 106 cut radial outside edge for receiving one end of a cylindrical sleeve 108 which forms the outer shell of the rotating assembly 37. The sleeve 108 is clamped between the rabbet 106 on the control housing end flange 104 and an annular sleeve retainer 110 held against the stepped outer section 112 of the output side end cap 78 by the bolts 82. The port plate 24, disposed between the axially juxtaposed pump unit 20 and the reaction stator unit 22, and both output rotors 34 and 35, as well as the end caps 78 and 80 all extend radially outward to the sleeve 108 and are contained therein, maintaining the the elements in concentric alignment around the axis 113 of the machine. The cylindrical inner surface 114 of the input end of the sleeve 108 forms the outer cylindrical surface of the control cylinder 103 of the control mechanism 40.
An annular piston 120 is positioned in the annular control cylinder 103 for axial movement therein under control of fluid pressure from a conventional gerotor make-up pump 121 driven by the input shaft 26 and selectively exerted on opposite axial faces of the piston 120 through fluid ducts 122 and 124. The gerotor pump 121 has an external cam ring 116 in which an outer gear 117 rotates. An internal gear 118, having one fewer teeth than the outer gear 117, drives the outer gear 117 which revolves around the internal gear while pressurizing fluid in the working space between the inner and outer gears 118 and 117. A control system 125, shown in FIG. 6 and described below, governs the selective delivery of control fluid through the ducts 122 and 124. Two radially spaced seals 126 and 128 in annular grooves in the piston 120 seal the piston 120 in the cylinder for sliding movement therein. Four control rods 130, circumferentially spaced around the rotating assembly 37, are fastened to the annular piston 120 by nuts 132 threaded onto threaded ends of the control rods 120 and holding a shoulder 134 of a stepped end of the control rods 130 against the inner face of the annular piston 120.
The control rods 130 extend axially through aligned holes in the two end caps 78 and 80, in the pump and stator output rotors 34 and 35, and the port plate 24. An "O" ring 136 on each control rod 130 seals the clearance between the rods 130 and the holes in the stator side end cap 80 to prevent leakage of pressurized fluid therethrough. Each control rod 130 has a pair of inwardly opening wedge-shaped notches 138 and 140 positioned in radial alignment with the centerline of the output rotors 34 and 35, respectively. The notch 138 has an upright left side and a sloping right side, hence will be referred to as a "right sloping" notch The notch 140 has an upright right side and a sloping left side, hence will be referred to as a "left sloping" notch. The top and bottom control rods 130 T and 130B shown in FIGS. 2 and 3 and identical to the top control rod 130 shown in FIG. 1. The side control rods 130S have the same kind of notches 138 and 140, but the upright and sloping sides are switched. That is, the notch 138S in the side control rods 130S is a "left sloping" notch as described above, and the notch 140S is a "right sloping" notch.
A complementary wedge-shaped notch 142 is provided in each of four radially projecting tenons 146 on the pump cam ring 32. The tenons 146 extend into radially opening channels 148 on the pump output rotor 34 and provide a radially slidable torque coupling by which the pump cam ring 32 is keyed to the pump output rotor 34. The notches 138 and 142 receive toggles 150 by which axial motion of the control rods 130 is converted to radial translation of the tenons 146. Similarly, a complementary wedge-shaped notch 152 is provided in each of four radially projecting tenons 154 on the stator cam ring 36. The tenons 154 extend into radially opening channels 156 on the stator output rotor 35 and provide a radially slidable torque coupling by which the stator cam ring 36 is keyed to the stator output rotor The notches 140 and 152 receive toggles 160 by which axial motion of the control rods 130 is converted to radial translation of the tenons 154. The control rods 130 all have a square cross-section where they extend through the tenons 146 and 154 and fit closely in the notches in those tenons to ensure that the control rods do not turn and misalign the notches in the control rods 130 with the toggles which they are supposed to operate.
Each of the end caps 78 and 80 has a stepped axially facing surface providing an annular, axially facing recess 164 on end cap 78, and another annular, axially facing recess 166 on the end cap 80 for receiving wear plates 168 and 170, respectively. The wear plates 168 and 170 provide replaceable wear surfaces for the end caps 78 and 80, and also provide lubrication and pressure balancing channels in a manner known in the vane pump art. The fluid pressure on both sides of the cam rings 32 and 36 is balanced as shown in FIGS. 1 and 4, and in other corresponding figures for the other tandem units shown in FIGS. 1-16. The slots in the side plates 168 and 170 communicate through to balancing basins on the outside faces of the side plates to provide axial fluid balancing and wear compensation of the side plates, and also communicate through four circumferentially spaced ports near the outer radial edge of the side plates to allow fluid communication around the edge of the cam rings to the spaces radially outside the cam rings to provide radial fluid balance on both radial surfaces of the cam rings 32 and 36.
Four radial holes 172 are drilled in the port plate 24 which communicate with axial holes 174 in the pump output rotor 34 containing check valves 176. Similar axial holes 178 in the stator output rotor 35 receive check valves 180. Fluid pressurized in the make-up pump 121 flows through an annular space between the ground shaft 72 and the input shaft 26 and then through the radial holes 172 in the port plate 24 where it is available to be drawn into the pump and stator working volumes to make up for fluid leakage out of the pressurized volumes in the pump and stator.
In operation, the housing 30 is filled with hydraulic fluid and the input shaft 26 is driven by a prime mover 28 such as an engine or motor. Assuming that it is desired to drive the output shaft 38 initially at a slow speed with high torque, as would be done in the case of starting up most applications such as vehicles, marine vessels, machine tools or engine powered equipment, the transmission is initially set in the neutral position, wherein the annular control piston 120 is positioned to its right-most position, or about 0.25" to the right from the 1:1 position shown in FIG. 1. The control piston 120 is positioned by delivering pressurized control fluid from a control valve 182 in the control system 125, shown in FIG. 6, through the control duct 122. The duct 122 passes through the mounting flange 54, through the outer cam ring 116 of the make-up pump 121, and through the end flange 73 of the ground shaft 72. An outer annular groove in the input side face of the thick section 92 of the end cap 80 is radially aligned with the portion of the control duct 122 opening in the output-side face of the ground shaft end flange 73, and communicates through a radial portion of the duct 122 in the end cap 80 and a diagonal portion through the control housing with the control cylinder 103 adjacent the end flange 104 of the control housing 90.
Control of the axial motion of the annular piston 120 toward the flange end of the control cylinder 103 is by way of the control duct 124 which also passes through the mounting flange 54, through the outer cam ring 116 of the make-up pump 121, and through the end flange 73 of the ground shaft 72. An inner annular groove in the input side face of the thick section 92 of the end cap 80 is radially aligned with the portion of the control duct 124 opening in the output-side face of the ground shaft end flange 73, and communicates through a radial portion of the duct 124 in the end cap 80 directly with the control cylinder 103 between the annular piston 120 and the end cap 80. The control cylinder 103 is thus in continuous fluid communication through portions of the ducts 122 and 124 in the rotating end cap 80 with the portions of the ducts 122 and 124 through the stationary ground shaft end flange 73.
At the right-most position of the control piston 120 and the attached control rods 130, that is, at the position of the control piston 120 shifted farthest from the end flange 104 of the control housing 90, the top and bottom pump toggles 150 engaged with the top and bottom control rods 130 will be lying at about 45.degree. to the right, and the top and bottom stator toggles 160 engaged with the top and bottom control rods 130 will be upright. At the same time, the side pump toggles 150S will be lying over 45.degree. to the left in a position identical to that shown in FIG. 1 for the toggles 160, and the side stator toggles 160S will be in a position perpendicular to their control rods 130, in a position identical to that of the pump toggles 150 shown in FIG. 1. In this position of the control piston 120, the cross sectional shape of the pump cam ring 32 is circular (like the shape of the stator cam ring 36 shown in FIG. 1) and the shape of the stator cam ring 36 is elliptical (like the shape of the pump cam ring 32 shown in FIG. 1.)
In the neutral position, the input shaft drives the pump rotor 42 and the pump rotor vanes 46 sweep around the pump working volume, which is the volume between the pump cam ring 32 and the pump rotor 42, and between the port plate 24 and the wear plate 168. However, since the pump working volume is cylindrical and co-axial with the pump rotor 42, there is no displacement. Essentially, a cylindrical annulus of fluid merely rotates around the circular pump cam ring 32 with the pump rotor 42. The only output torque exerted on the output shaft 38 in the neutral position of the transmission is produced by the friction force of the pump vanes 46 exerted on the pump cam ring 32, the port plate 24 and wear plate 168 as the vanes 46 travel around the pump working volume.
To begin applying power through the transmission to the wheels of the vehicle or other driven application, fluid pressure is applied by the control valve 182 through the control duct 124 to the portion of the control cylinder between the piston 120 and the end cap 80, and the control valve 182 vents the control duct 122 to sump. The control fluid from the control duct 124 moves the piston 120 toward the end flange 104 and pulls the control rods in the same direction. The pump toggles 150 are rocked by the control rods 130 to the right, camming the pump cam ring 32 inward to an elliptical shape. Simultaneously, movement of the side control rods 130S rocks the toggles 150S to the right to allow the tenons 146S on the sides of the pump cam ring 32 to advance radially outward into the channel 148S to facilitate the change of shape of the cam ring 32 from circular to elliptical. The change of shape of the cam ring from circular to elliptical changes the form of the pump working volume from an annular cylinder to a wider, shorter shape, with a cross section thinner at its top and bottom and thicker at its sides. Rotation of the pump vanes through the pump working volume of this cross-sectional shape pressurizes fluid in two oppositely disposed pressure sectors 184, each having a tapering cross section, and causes suction of fluid into the two oppositely disposed suction sectors, each having a flaring cross section. The "tapering" and "flaring" descriptions are from the perspective of the vanes 46 moving in the sectors "tapering" sector is one in which the vanes 46 are moving into their slots in the rotor 42 and the volume of the pump working volume between adjacent vanes is decreasing as the rotor 42 rotates. A "flaring" sector is one in which the vanes 46 are moving out of their slots in the rotor 42 and the volume of the pump working volume between adjacent vanes is increasing as the rotor 42 rotates. Thus, the volume between adjacent vanes 46 in the tapering pressure sectors 184 decreases as the vanes rotate through those sectors, thereby pressurizing the fluid in those volumes, and increases as the vanes rotate in the flaring sectors, thereby producing a suction of fluid into those volumes.
Fluid pressurized in the pump working volume by the pump vanes 46 contributes toward output torque in two ways: its pressure is exerted against the radial faces of the tenons 146 in the pressure sectors 184. The pressurized fluid thus exerts a hydraulic force on the pump cam ring 32 equal to the product of the fluid pressure and the effective area against which the pressure acts in the circumferential direction. This effective area is the product of the axial length of the cam ring and the difference between the exposed radial surface of the tenons 146 on the two circumferential limits, 90.degree. apart, of the pressure sectors 184.
Fluid pressurized by the pump in its pressure sectors 184 is displaced therefrom through pressure ports 186 in the port plate 24 directly into the stator unit 22 where it acts on the tenons 154 of the stator cam ring 36 in tapering sectors of the working volume to produce torque that is transmitted to the stator output rotor 35 by engagement of stator tenons 154 on the stator can ring 36 in channels 156. The pressurized fluid in the tapering sectors of the stator working volume acts on the stator cam ring 36 in exactly the same way to produce output torque to the output rotor with the exception that the stator hub 70 and its vanes 72 do not rotate; the stator output rotor 35 rotates relative to the stator hub 70 and vanes 72 along with the rest of the rotating assembly 37 in the direction of rotation indicated by the direction arrow in FIG. 3. The function of the stator vanes is to divide the stator working volume into smaller sectors to prevent pressurized fluid in the pressurized sectors, wherein the working volume decreases in the direction of rotation of the stator output rotor 35, from flowing into the suction sectors, wherein the working volume increases.
The fluid displaced by the relative rotation of the stator rotor 35 and the stator hub 70 and vanes 72 is conveyed directly through ports 188 in the port plate 24. Because of the reciprocal relationship of the positions of the pump cam ring 32 and the stator cam ring 36, established the fixed inverse positions of the toggles 150 and 160 on the control rods 130, the displacement of the pump unit 20 is always almost exactly identical to the suction volume of the stator unit 22. This is true even though the pump cam ring 32 rotates in the same direction as the pump rotor and therefore has a lower displacement than such a pump with a non-rotating cam ring would have, because the stator cam ring 36 is also rotating in the same direction at the same speed, so the suction demands of the stator working volume are correspondingly lower. The displacement from the pump as the cross-section approaches the greatest difference between major and minor axis dimensions actually decreases because the stator displacement decreases as the stator cam ring becomes more circular. Since the available volume in the stator unit 22 is limited, and there is no other path in which the fluid driven by the pump vanes 46 can flow, it drives the pump cam ring 32 in the same direction of rotation as the pump rotor 42 instead of flowing out of the pump working volume.
Increases in the output speed of the transmission and corresponding decreases in the transmission ratio toward 1:1 are achieved by moving the control piston 120 further and further to the left until the 1:1 position illustrated in FIG. 1 is reached. The fluid flow rate between the pump and stator units 20 and 22 reaches a maximum in a mid-range of the transmission ratio, it then declines toward zero when the transmission ratio approaches 1:1. This is the ratio where most driving is done, and provides maximal efficiency and minimal noise because fluid flow losses and noise generated in this transmission are minimal.
The friction force exerted by the pump vanes 46 on the pump cam ring 32, the wear plate 78, and the port plate 24 would, in conventional vane pumps, be reacted through the housing to ground, and would be wasted. In this invention, however, this friction force is coupled to the output shaft 38 and thus contributes to the output torque.
THE SECOND EMBODIMENT
Turning now to FIGS. 7-16, an electrically controlled, tandem, vane-type continuously variable transmission is shown having a housing 200 including a cylindrical body 202 having an input end connected to a mounting flange 204 by screws 206, and having an output end connected to an output end closure 208. A ground shaft 210 has an end flange 212 that is rigidly fastened in a recess 214 in a thickened center portion 216 of the mounting flange 204 by screws 218. The ground shaft has a tubular portion 220 that projects axially into the housing 200 and receives an input shaft 222, journaled therein at the flange end of the ground shaft in a journal bearing 224.
The input shaft 222 has an input end with a torque coupling structure such as the spline 226 for connection to an output drive shaft 27 of a prime mover 28. The input shaft 222 has a distal end 228 journaled in a bearing 230 mounted in an axial opening 232 in the center of an end cap 234 at the output end of a rotating assembly 236. The rotating assembly 236 includes a pump unit 238 and a stator unit 240 disposed on opposite sides of a flat port plate 242 and between the output-end end cap 234 and an input-end end cap 244. A suitable number of circumferentially spaced bolts 246 (this embodiment is designed to use twelve such bolts) secure the rotating assembly 236 to an inner end flange 248 of an output shaft 250, mounted for rotation in bearings 252 in the end closure 208, and hold the pump and stator units between the end caps, reacting the outward axial forces exerted by pressurized fluid in the pump and stator units 238 and 240. The output shaft has an end section 253 that is splined for torque transmitting connection to a driven application, such as a vehicle drive shaft for driving the wheels of the vehicle.
The pump unit 238 has an annular pump rotor 254 having an inside spline engaged in a torque transmitting coupling with a spline section 256 on the input shaft 222 adjacent to the distal end 228. The pump rotor 254 has a series of radial slots 258 regularly spaced circumferentially around the rotor, each slot 258 slidably receiving a radially projecting pump vane 260. Ten slots are provided in the disclosed embodiment, but other numbers of vanes can be used. The number of vanes is primarily determined by the maximum number of vane slots 258 that can be fit into the available space on the rotor 254. The pump vanes 260 are driven by the pump rotor 254 around a pump working volume 262 to pressurize hydraulic fluid in the working volume for operation of the transmission. The working volume is confined radially between the outside cylindrical surface 264 of the pump rotor 254 and an inside surface 266 of a pump cam ring 268 radially spaced from and surrounding the pump rotor. The axial confines of the pump working volume 262 are the output-side face 269 of the port plate 242 and an inner face of a wear plate 270 mounted in an annular recess in the inner face of the end cap 234.
Fluid pressurized in the pressure sectors of the pump unit 238 acts directly against the tenons 296 of the pump cam ring 268 in the same manner as the embodiment of FIG. 1 to produce torque, and also passes through pressure ports 272 in the port plate 242 directly into the working volume 274 in the stator unit 240. The stator working volume is the space limited radially between the outside cylindrical surface 276 of a stator hub 278 and an inside surface 280 of a stator cam ring 282. The axial boundaries of the stator working volume 274 are the stator-side face 284 of the port plate 242 and the surface of a wear plate 286 mounted in an annular recess in the input-end end cap 244.
Fluid pressurized in the pump unit 238 and pressurizing pressure sectors in the stator working volume 274 acts against the tenons 298 of the stator cam ring 282 to produce torque in the same manner as that described for the embodiment of FIG. 1, and is reacted by ten stator vanes 288 slidably mounted in radial slots in the stator hub 278. The pressure force reacted by the stator vanes 288 passes through the stator hub 278 to a spline connection 290 at the distal end of the ground shaft tubular extension 220 and thence back to the mounting flange 204.
The pump and stator cam rings 268 and 282 are keyed to pump and stator output rotors 292 and 294, respectively, by four radial tenons 296 on the pump cam ring 268 and four radial tenons 298 on the stator cam ring 282 that slide radially in channels 300 and 302 in the pump and stator output rotors 292 and 294. Torque is transmitted between the cam rings 268/282 and the rotors 292/294 by engagement of the circumferentially facing surfaces of the tenons 296/298 with the circumferentially facing surfaces in the channels 300/302 during transmission of torque through the transmission, and during back-driving from the output shaft to charge a hydraulic accumulator during recovery of braking energy, as will be described below in connection with the third embodiment of FIGS. 17 and 18.
A set of aligned axial holes is provided through both end caps 234 and 244, through both output rotors 292 and 294, and through the port plate 242 on diametrically opposite sides of the rotating assembly 236, for receiving two control rods 304. Each of the control rods 304 has two oppositely sloping wedge-shaped notches 308 and 310 which receive one end of two toggles 312 and 314, the other ends of which are received in complementary notches in abutments 316 and 318 formed on the pump cam ring 268 and the stator cam ring 282 adjacent an opposed pair of tenons 296 and 298 for distributing the radial force exerted by the toggles 312 and 314 on the cam rings. This embodiment differs from the embodiment of FIG. 1 in part by the use of only two control rods 304, whereas the embodiment of FIG. 1 has four control rods 130. This embodiment utilizes the spring-back or restoring force exerted by the material of the cam rings 268 and 282 after it is flexed to the elliptical shape shown in FIGS. 11 and 12 by the toggles 312 and 314. Typical materials that would provide sufficient tensile strength to give an adequate spring-back force are 4140 tool steel and 17-4 PH stainless steel.
Another difference between this embodiment and that of FIG. 1 is that the notches 308 and 310 in the control rods and the toggles 312 and 314 are arranged to shift the transmission from neutral toward 1:1 with a push on the control rods 304 whereas the embodiment of FIG. 1 accomplished that result with a pull on the control rods 130. For that reason, the wedge-shaped notches 308 and 310 on the control rods 304 slope in the opposite direction from the direction that the notches on the control rods 130 in the embodiment of FIG. 1 slope. Specifically, the notches 308 and 310 slope away from each other so that a push on the control rod 304 tends to cock the toggle 312 toward the vertical position and distort the pump cam ring 268 toward and elliptical cross-section, and tends to cock the toggle 314 toward a 45.degree. position (shown in FIG. 1:1) and allow the stator cam ring 282 to spring back toward its unstrained circular cross-section.
Axial control movement of the control rods 304 is by way of a control motor 320 which is coupled to the control rod 304 by a linkage including a control lever 322 and a spider 324. The motor 320 is a servo motor or a stepper motor mounted on pivot pins for swiveling through a small vertical arc between two spaced ears 326 attached to the cylindrical body 202 as by welding or other suitable means of attachment. The motor 320 drives a threaded rod 328 that is threadedly engaged with a internally threaded sleeve 330 pinned for swiveling through a small vertical arc between two upright arms of a fork 332. The fork 332 is attached by a vertical screw 334 to the end of the lever 322.
The control lever 322, shown in FIGS. 7, 8 and 16 has a central slot 336 which receives and is pinned for swiveling movement through a vertical arc to a bracket 338 attached to the inside of the mounting flange 204. The lower end of the control lever 322 is formed in a ring 340 that holds a spherical bearing race 342, in turn engaged with a spherical bearing 344 slidably mounted on the tubular extension 220 of the ground shaft 210. The spherical bearing 344 engages an annular disc 346 in contact with a roller bearing 348 surrounding the ground shaft tubular extension 220 and in rolling contact with the spider 324 which rotates with the rotating assembly 236.
The spider 324 includes a hub 350 surrounding the tubular extension 220 of the ground shaft 210, but spaced slightly therefrom to prevent contact between the rotating spider hub 350 and the stationary ground shaft. The spider has two diametrically opposed radial arms 352 attached at their distal ends to the ends of the two control rods 304 by screws 354 threaded into tapped axial holes in the end of the two rods 304. A conical spring 356 having radial fingers 358 extending inwardly from an outer ring and bearing at their ends against the inside face of the hub biases the spider toward the input end and maintains contact between the roller bearing 346, the disc 342, and the spherical bearing 344.
A shroud 360 encloses the rotating assembly 236 to present a smooth surface to the body of hydraulic fluid within the housing 200 to minimize churning the hydraulic fluid, particularly by the spider arms 324. The shroud has a circumferential ring of closely spaced shallow teeth 362, positioned around the circumference of the shroud 360 at an axial position therealong aligned with a port 364 through the body 202 of the housing 200. A sensor such as a proximity sensor 366 is mounted in the port in close proximity to the ring or teeth and detects the passage of each tooth 362 past the sensor. The sensor output is conducted to the control system 306 via leads 368 where it is converted by a processor 370 into rotational speed. The rotational speed is compared in a comparitor 372 to the desired set speed and, in the event of a deviation, the comparitor generates an error signal which is transmitted to the servo motor as correction signal to readjust the tilt angle of the control lever 322. An interrogation of the sensor 366 by the processor 370 is initiated periodically, for example, 10 times per second, so the transmission is continually adjusted if necessary to maintain the desired speed instruction input to the control system 306.
THE THIRD EMBODIMENT
A third embodiment of the invention, shown in FIGS. 17-18, utilizes a vane-type pump unit 400 and a surrounding concentric vane-type stator unit 402. A unitary cam ring 404 for both pump and stator units has inner and outer cam surfaces 406 and 408 in contact with the free ends of outwardly extending vanes 410 carried by a pump rotor 412 of the pump, and the free ends of inwardly extending vanes 414 carried by a surrounding hub 416 of the stator 402. The stator hub has a radial flange 417 bolted between two facing radial flanges 419L and 419R of two mating halves of an enclosing housing 418, shown in FIGS. 19-21.
The pump rotor 412 is splined to and driven by an inner splined section 420 of an input shaft 422, which in turn is journaled for rotation in the housing 418 in a bearing 424 mounted in a recess on an axial boss 426 on the housing 418. The inner end 428 of the input shaft 422 is journaled in a journal bearing 430 pressed into an axial bore in an inner face 432 of an output shaft 434, as shown in FIGS. 22-27. The output shaft 434 in turn is rotatably supported in a roller bearing 436 mounted in an axial bore 437 through an inwardly projecting nipple 438 on an inner surface of a flat output end wall 440 of the housing 418.
The inner end of the output shaft 434 is radially enlarged to form an end plate 442 which is axially spaced from a similar input-side end plate 444, shown in FIGS. 28-32. The end plates 442 and 444 each have a wear plate 443 and 445, respectively, fastened to the axially inwardly |