Technical Field The present invention pertains to hydrostatic drive systems which are also known as variators. Background

Modern hydrostatic drive systems typically consist of at least two hydraulic machines or units fluidly coupled to each other. One hydraulic unit may function as the pump and the other as the motor, or vice versa. When at least one of the hydraulic units is a variable unit, a continuously variable transmission or CVT is created. Further, a hydrostatic drive system can also function as the variator in a power split, infinitely variable transmission or IVT such as disclosed in US 7,357,744. Positive displacement hydraulic units can be of vane, plunger, diaphragm or piston designs. Bent axis piston hydraulic machines have become common in many hydrostatic drive systems such as the Fendt (AGCO) Vario Transmission or as disclosed in US 8,240,145. Radial piston hydraulic machines offer certain advantages over bent axis machines including packaging and relative simplicity of parts but traditional designs have had some drawbacks. A hydrostatic radial piston machine has a rotating cylinder block with multiple radial cylinder bores. The cylinder block is connected to a drive shaft. Typically, a fixed central pintle contains fluid passages and radial ports exposed to the cylinder bores. As the cylinder block rotates, the inner ends of the cylinder bores are alternately in fluid communication first with one set of passages, then the other. The outer ends of the pistons ride in an eccentric cam ring resulting in the pistons moving up and down in the cylinder bores as the block rotates. An example of this type of design is disclosed in US 3,750,533. The drawbacks of this design are the limited size of fluid channels in the pintle and high bending loads in the pintle. Other drawbacks include the high tolerances between pintle shaft and rotor for sealing which can create intolerance to thermal differentials between the two elements causing thermal shock seizures and catastrophic failure. Reducing the clearance to compensate however increases the leakage. In effect, there is no way to compensate at this interface both for leakage and potential seizure. A further drawback for this design is an inability to have an additional through shaft in the middle of the machine which is important for packaging of split path IVTs.

A new design of radial piston machine has been disclosed by Juergen Berbuer in EP2510192B1 and US2013/0145929 (issued as US9784252). In this design, the cylinder block has two roughly concave conical sides which match with two similar control plate elements (one on each side of the cylinder block). Fluid ports are positioned axially along these interfaces allowing fluid to move between the cylinder block and the control plates. Fluid pressure creates a pair of hydrodynamic bearings between the cylinder block and the control plates. The main advantages of this design over previous designs is that radial loads from the cylinder block are transferred through the control plates to the housing. Since the proportions of the control plates are much more stout than a pintle or drive shaft, the assembly is much more stable. The drive shaft is only required to transmit torque and the absence of a pintle allows room for an additional through shaft. The new design also compensates for thermal differentials reducing or eliminating the potential for seizure type failure of the pintle design. With the absence of rolling bearings, the new design is also simpler than previous designs creating potential for cost savings.

As disclosed in US9784252, the aforementioned new hydrostatic radial piston machine comprises a housing, a radial cylinder block rotatably supported in the housing about a rotation axis, a plurality of pistons which corresponds to the plurality of bores, a cam ring which is arranged eccentric relative to the radial cylinder block and which circumferentially envelops the radial cylinder block, two control plate elements including a total of at least two control cross-sections, and a plurality of pass through channels in the radial cylinder block corresponding to the plurality of bores in the radial cylinder block. The radial cylinder block includes a plurality of bores extending from an outer enveloping surface of the radial cylinder block into an interior of the radial cylinder block and arranged distributed over a circumference of the radial cylinder block. The plurality of pistons are movably supported in the bores and respectively define an operating cavity for a hydraulic fluid together with the associated bore. In the hydrostatic radial piston machine, ends of the pistons oriented away from the radial cylinder block are movably supported at a continuously cambered inner enveloping surface of the cam ring during a rotation of the radial cylinder block. Further, at least one control cross-section is connected with an inlet channel and at least another control cross-section is connected with an outlet channel. In addition, the two control plate elements extend respectively with a face oriented towards a central plane of the radial cylinder block, in which the central plane is perpendicular to the rotation axis, and the two control plate elements extend with the faces oriented towards the radial cylinder block beyond a plane which is defined by a face of the radial cylinder block, in which the face of the radial cylinder block is oriented towards the respective control plate element at a greatest axial width of the radial cylinder block. In the hydrostatic radial piston machine, the plurality of pass through channels as a function of a rotational position of the cylinder block in the cam ring respectively connect the operating cavity with the control cross-section corresponding with the inlet channel or with the control cross-section corresponding with the outlet channel or are closable by a closing surface arranged at the control plate element, in which each control plate element includes a radial bearing portion in which radial forces from the radial cylinder block are transferrable to a respective opposite radial surface in the housing or to a radial surface of a housing cover supported in the housing through a direct contact of the radial bearing portion with the radial surface. Alternatively as disclosed in EP2510192B1, the aforementioned new hydrostatic radial piston machine has a housing, a cylinder star which is mounted in the housing such that it can rotate about a rotation axis and has a number of bores which extend starting from an outer lateral surface of the cylinder star into the interior thereof and are distributed over the circumference thereof. The hydrostatic radial piston machine also has a number of pistons corresponding to the number of bores which are arranged in a displaceable manner in the bores and each delimit a working space for a hydraulic fluid together with the associated bore. The hydrostatic radial piston machine also has a lifting ring which is arranged eccentrically to the cylinder star and surrounds the cylinder star in a circumferential manner, and on the inner lateral surface of which ends of the pistons which face away from the cylinder star are supported in a movable manner during the rotary movement of the cylinder star. Further, the hydrostatic radial piston machine has two control mirror bodies which have at least two control cross sections in total, of which at least one is connected to the inlet duct and at least one other is connected to the outlet duct, wherein both control mirror bodies each extend with an end face facing the cylinder star towards a central plane of the cylinder star, which plane is perpendicular to the rotation axis, beyond a plane which is defined by an end face of the cylinder star facing the respective control mirror body at the point of the cylinder star with the greatest axial width. Further, the hydrostatic radial piston machine has a number which corresponds to the number of bores in the cylinder star of through-ducts which are arranged in the latter and, depending on the rotation position of the cylinder star in the lifting ring, each connect a working space to a control cross section which corresponds to the inlet duct or to a control cross section which corresponds to the outlet duct or can be closed by a closure face which is situated on the control mirror body, characterised in that each control mirror body has a bearing region in which radially effective forces can be transmitted to a respective counter face in the housing or a housing cover mounted therein. US 5,228,290 discloses a variator consisting of 2 hydraulic radial piston machines fluidly coupled together. Although that device uses axial ports located along the side faces of the cylinder blocks, the faces are purely for control of fluid flow. All radial and axial forces created at the cylinder blocks are contained by several ball bearings. Additionally, the cam ring is also a rather large ball bearing. Since the bearings must be kept to a practical size, the result is a variator with a large outer envelope. Although a through shaft would be possible with that design, again, the bearing configuration limits the diameter of the through shaft.

As discussed earlier, there is a need in both stationary and mobile machinery for a compact variator for various CVT and IVT applications. The present invention addresses these and other needs as described below.

Summary An advantage of the invention presented here to provide a compact hydrostatic variator which is simpler and more cost effective than traditional bent axis, axial, or prior radial variators. The hydrostatic drive systems and variators of the present invention comprise two of the hydrostatic radial piston machines disclosed in the aforementioned EP 2510192 and/or US 2013/0145929 (issued as US9784252). These two radial piston machines are fluidly connected together with a common control plate in between. At least one of the radial piston machines is of a variable displacement design. Each machine has an independent drive shaft which is connected to its cylinder block. In addition, a third shaft may pass through the middle of the hydrostatic drive system or variator for split power IVT applications.

Specifically, a hydrostatic drive system of the invention comprises a housing, two hydrostatic radial piston machines within the housing and fluidly coupled to each other, and an intermediate control plate element common to each of the two hydrostatic radial piston machines. Each of these hydrostatic radial piston machines comprises a radial cylinder block comprising a plurality of bores, a plurality of pistons corresponding to the plurality of bores, a cam ring, two control plate elements in which one of the two control plate elements is the intermediate control plate element, and a plurality of pass through channels in the radial cylinder block. Each of the two control plate elements in each of the two hydrostatic radial piston machines includes a total of at least two control ports. Further, the intermediate control plate element is rotationally fixed in the housing between the radial cylinder blocks in the two hydrostatic radial piston machines.

In embodiments of the hydrostatic drive system, the radial cylinder block in each of the two hydrostatic radial piston machines can be rotatably supported in the housing about a rotation axis and includes a plurality of bores extending from an outer enveloping surface of the radial cylinder block into an interior of the radial cylinder block and arranged distributed over a circumference of the radial cylinder block. The plurality of pistons in each of the two hydrostatic radial piston machines can be movably supported in the bores and respectively define an operating cavity for a hydraulic fluid together with the associated bore. The cam ring in each of the two hydrostatic radial piston machines can be arranged eccentric relative to the radial cylinder block and circumferentially envelops the radial cylinder block and wherein ends of the pistons oriented away from the radial cylinder block are movably supported at a continuously cambered inner enveloping surface of the cam ring during a rotation of the radial cylinder block. At least one control port in at least one control plate element connects with an inlet/outlet channel and at least another control port in at least one control plate element connects with another inlet/outlet channel. In embodiments of the hydrostatic drive system, the two control plate elements in each of the two hydrostatic radial piston machines can extend respectively with a face oriented towards a central plane of the radial cylinder block, wherein the central plane is perpendicular to the rotation axis, and the two control plate elements extend with the faces oriented towards the radial cylinder block beyond a plane which is defined by a face of the radial cylinder block, wherein the face of the radial cylinder block is oriented towards the respective control plate element at a greatest axial width of the radial cylinder block. The plurality of pass through channels in the radial cylinder block correspond to the plurality of bores in the radial cylinder block. Further, the pass through channels in each of the two hydrostatic radial piston machines as a function of a rotational position of the cylinder block in the cam ring respectively can connect the operating cavity with the control port corresponding with the inlet/outlet channel or with the control port corresponding with the another inlet/outlet channel or can be closable by a closing surface arranged at the control plate element. Further still, each control plate element can include a radial bearing portion in which radial forces from the radial cylinder block are transferrable to a respective opposite radial surface in the housing or to a radial surface of a housing cover supported in the housing through a direct contact of the radial bearing portion with the radial surface.

In an embodiment of the hydrostatic drive system, the control plate elements radially and axially support the radial cylinder blocks in the hydrostatic radial piston machines.

In an embodiment of the hydrostatic drive system, the intermediate control plate element is free to move axially in the housing between the radial cylinder blocks in the two hydrostatic radial piston machines, a plurality of compensating pistons axially locates one of the control plate elements other than the intermediate control plate element relative to the housing, and the one of the control plate elements comprises a plurality of compensating piston bores corresponding to the compensating pistons and a plurality of control channels connected to the corresponding compensating piston bores and the corresponding control ports in the bearing faces of the corresponding one of the other control plate elements.

In an alternative embodiment of the hydrostatic drive system, the intermediate control plate element is axially fixed in the housing between the radial cylinder blocks in the two hydrostatic radial piston machines, a plurality of compensating pistons axially locates each control plate element other than the intermediate control plate element relative to the housing, and each control plate element other than the intermediate control plate element comprises a plurality of compensating piston bores corresponding to the compensating pistons and a plurality of control channels connected to the corresponding compensating piston bores and the corresponding control ports in the bearing faces of the corresponding control plate elements. In embodiments of the invention, the interfaces between the control plate elements and the corresponding radial cylinder blocks may for instance be spherically cambered in shape or conical ring shaped. In an exemplary embodiment, at least one of the hydrostatic radial piston machines is a variable displacement hydrostatic radial piston machine. In an alternative embodiment, both of the hydrostatic radial piston machines are variable displacement hydrostatic radial piston machines.

In a hydrostatic drive system of the invention, one of the hydrostatic radial piston machines can be a pump and the other of the hydrostatic radial piston machines can be a motor.

Hydrostatic drive systems of the invention can advantageously be used in variable transmissions, such as a power split, hydro mechanical infinitely variable transmission or a continuously variable transmission.

Brief Description of the Drawings

Figure 1 is a longitudinal cross section of a preferred embodiment of a hydrostatic variator of the invention. Figures la and lb show two detail views of Figure 1.

Figure 2 is a transverse cross section of the embodiment of the hydrostatic variator of Figure 1 according to line B-B.

Figure 3 is a transverse cross section of the embodiment of the hydrostatic variator of Figure 1 according to line C-C.

Figure 4 is a transverse cross section of the embodiment of the hydrostatic variator of Figure 1 according to line D-D. Figure 5 is a transverse cross section of the embodiment of the hydrostatic variator of Figure 1 according to line E-E.

Figure 6 and 6a are isometric views of first control plate element 10 in the embodiment of the hydrostatic variator of Figures 1 and la.

Figure 7 and 7a are isometric views of intermediate control plate element 24 in the embodiment of the hydrostatic variator of Figures 1 and la. Figure 8 is a longitudinal cross section of an alternate embodiment of a hydrostatic variator of the invention. Figures 8a shows a detail view of Figure 8. Figures 9 and 9a are isometric views of second control plate element 18 in the embodiment of the hydrostatic variator of Figures 8 and 8a.

Detailed Description Unless the context requires otherwise, throughout this specification and claims, the words "comprise", "comprising" and the like are to be construed in an open, inclusive sense. The words "a", "an", and the like are to be considered as meaning at least one and are not limited to just one.

Herein, the term "hydrostatic machine" refers to a machine which moves fluid between an inlet and an outlet by trapping fluid in one or more movable chambers. It has one drive shaft to transfer mechanical power to the chambers. It may function as a pump or a motor. It may be of variable (fluid) displacement or fixed (fluid) displacement.

A "hydrostatic radial piston machine" refers to a hydrostatic machine as described above with a radial arrangement of cylinders and pistons.

A "variator" refers to a machine which has an input and an output shaft with variable speed and torque ratios between the two shafts. A "hydrostatic variator" or "hydrostatic drive system" refers to a variator comprising two hydrostatic machines fluidly coupled together. At least one of the hydrostatic machines is of a variable displacement allowing variable speed and torque ratios between the two shafts.

"Displacement" means the volume of fluid moved between inlet and outlet ports of a hydrostatic machine in one revolution of its drive shaft.

The term "inlet/outlet channel" as used herein refers to a channel which may serve either as an inlet channel or as an outlet channel. Figures 1, la, and lb show a preferred embodiment of a hydrostatic variator of the invention. Figures 2-5 show various sections of a preferred embodiment of a hydrostatic variator of the invention. As shown in Figure 1, variator 1 consists of first radial piston machine 2 and second radial piston machine 3 contained in a housing 5. Each radial piston machine 2, 3 is arranged around rotation axis 4. In variator 1, one of the radial piston machines acts as a pump and the other as a motor.

First radial piston machine 2 comprises:

- first radial cylinder block 8;

pistons 11;

piston seals 12;

piston guide rings 13.

first cam ring 9;

- first control plate element 10 and

intermediate control plate element 24.

First radial cylinder block 8 comprises first and second side walls 8c and 8d respectively which are roughly concave in shape. A plurality of radial bores 8b which extend from the outer enveloping surface 8e into an interior of the radial cylinder block and are evenly distributed over a circumference of the radial cylinder block. A plurality of axial pass through channels 8a extend from side wall 8c to side wall 8d and connect with corresponding radial bores 8b. First radial cylinder block 8 is arranged around and is rotatable about rotation axis 4. The greatest axial width of first cylinder block 8 is defined by faces 8g and 8h respectively. A center plane 8f perpendicular to rotation axis 4 defines the center plane of radial cylinder block 8.

A first cam ring 9 circumferentially envelopes first radial cylinder block 8. As shown in Figure 2, the cam ring 9 may be displaced up to a total positive eccentric displacement El as measured from the rotation axis 4. In a preferred embodiment, the cam ring 9 may also be displaced up to a total negative eccentric displacement E2 also as shown in Figure 2. First cam ring 9 comprises an inner enveloping surface 9a which is continuously spherically cambered in shape.

A plurality of pistons 11 corresponding to the plurality of radial bores 8b are movably supported and free to slide and tilt in the radial bores 8b. A plurality of piston seals 12 hydraulically seal the pistons 11 against the walls of the radial bores 8b. Each piston 11 and its corresponding radial bore 8b define an operating cavity 39 for hydraulic fluid. During operation, fluid pressure in the operating cavity 39 forces the ends of the pistons 11 that are oriented away from the radial cylinder block 8 against the inner enveloping surface 9a of first cam ring 9. In absence of adequate fluid pressure in operating cavity 39, the pistons 11 are held against the inner enveloping surface 9a by piston guide rings 13.

A first control plate element 10 is arranged around rotation axis 4 adjacent to first side wall 8c of first radial cylinder block 8. As shown in Figures 1, la, and 6a, first control plate element 10 comprises a roughly convex bearing face lOe the shape of which complements the first wall of radial cylinder block 8. As shown in Figure 6a, A-control port lOf and B-control port lOg are formed in the face of bearing face lOe are located opposite each other and are kidney shaped. Two closing surfaces lOj and 10k are created between A-control port lOf and B-control port lOg. A plurality of small bores and large bores 10c and lOd respectively are located on the face opposite bearing face lOe. As shown in Figures 6 and 6a, the upper small and large bores 10c and lOd are connected to the upper A-portl0f through A- Control Channels 10a while the lower bores 10c and lOd are connected to the lower B-port lOg thru B- Control Channels 10b. First control plate element 10 is radially supported in first end cap 6 by radial bearing portion lOh and opposite radial surface 6a and is free to slide axially but not permitted to rotate. First control plate element 10 is preloaded axially against first radial cylinder block 8 by preload spring 38 installed between first control plate element 10 and first end cap 6.

A plurality of small and large compensation pistons 14 and 15 are moveably supported in small and large bores 10c and lOd. During operation, fluid pressure in large and small bores 10c and lOd forces the small and large compensation pistons 14 and 15 against first end cap 6. As a result, first control plate element 10 is located axially in housing 5 against first radial cylinder block 8.

Alternate embodiments may vary in size and quantities of compensation bores and pistons. In an alternate embodiment, only a single upper kidney shaped compensation bore and piston and a single lower kidney shaped compensation bore and piston are used in each control plate element.

An intermediate control plate element 24 is arranged around rotation axis 4 adjacent to second side wall 8d of first radial cylinder block 8. As shown in Figures 1 and la intermediate control plate element 24 comprises a first and second bearing face 24f and 24g the shape of which is roughly convex and complements the second wall of radial cylinder block 8. As shown in figure 7, an upper first A-control port 24j and a lower first B-control port 24k are formed in the face of first bearing face 24f are located opposite each other and are kidney shaped. As shown in figure 7a an upper second A- control port 24m and a lower second B-Control Port 24n are formed in the face of second bearing face 24g are located opposite each other and are kidney shaped. First A-control port 24j and second A- control port 24m are connected by A-inlet/outlet channel 24a. First B-control port 24k and second B- control port 24n are connected by B-inlet/outlet channel 24b. In a preferred embodiment, intermediate control plate element 24 is fully fixed within the housing 5, that is it cannot rotate nor translate relative to the housing. First drive shaft 25 is coupled to first radial cylinder block 8. First radial cylinder block 8 is supported on first control plate element 10 and intermediate control plate element 24 at interfaces Fl and F2 where interface Fl is defined by first side wall 8c and bearing face lOe; F2 is defined by second side wall 8d and first bearing face 24f. Faces 8g and 8h define the greatest width of first cylinder block 8 and are oriented towards control plate 10 and intermediate control plate 24 respectively. A center plane 8f perpendicular to rotation axis 4 defines the center plane of radial cylinder block 8. Interfaces Fl and F2 are symmetric to each other across center plane 8f. Bearing faces lOe and 24f extend past planes formed by faces 8g and 8h and are oriented towards the center plane 8f . The radial forces generated by first radial cylinder block 8 are transferrable to first control plate element 10 and intermediate control plate element 24 at interfaces Fl and F2. The radial forces in the first control plate element 10 are transferrable to the first endcap 6 through direct contact of radial bearing portion lOh and opposite radial surface 6a. As shown in figure 3, the radial forces in the intermediate control plate element 24 are transferrable to the housing 5 through direct contact of radial bearing portion 24h and opposite radial surface 5a. Thus, the radial forces generated by first radial cylinder block 8 are transferrable to housing 5 and first end cap 6; only torque loads from first radial cylinder block 8 are transferable to first drive shaft 25.

During operation, a gap is created by fluid pressure between first side wall 8c and bearing face lOe and between second side wall 8d and first bearing face 24f resulting in interfaces Fl and F2 acting as 2 hydrodynamic bearings. The areas of each set of compensation pistons 14 and 15 are such that when pressurized, the axial force generated by the compensation pistons is slightly above the hydraulic gap force generated in any of interfaces Fl to F2. In this way, the gap at each interface is controlled. Thus, the first control plate element 10 and intermediate control plate element radially and axially support radial cylinder block 8 at two interfaces Fl and F2.

In a preferred embodiment, first radial piston machine 2 also comprises a means for adjusting fluid displacement shown as first displacement control 28 in Figure 2. Second radial piston machine 3 comprises:

second radial cylinder block 16;

pistons 19;

piston seals 20;

piston guide rings 21.

- second cam ring 17;

second control plate element 118 and

intermediate control plate element 24. The components of second radial piston machine 3 are arranged in an analogous manner to those of first radial piston machine 2 and having second drive shaft 26 coupled to second radial cylinder block 16. Cam ring 17 may be displaced to any portion of a total positive eccentric displacement of second cam ring E3 as shown in Figure 4. In a preferred embodiment, cam ring 17 may also be displaced to any portion of a total negative eccentric displacement of second cam ring E4 also as shown in Figure 4. In a preferred embodiment, second control plate element 118 is constructed similarly to first control plate element 10. In a preferred embodiment, second radial piston machine 3 also comprises a means for adjusting fluid displacement shown as second displacement control 29 in Figure 4.

Figure la shows a detail of interface Fl between first radial cylinder block 8 and first control plate element 10 and interface F2 between first radial cylinder block 8 and intermediate control plate element 24. Likewise, as shown in Figure lb, is interface F3 between second radial cylinder block 16 and second control plate element 118 and interface F4 between second radial cylinder block 16 and intermediate control plate element 24. The shape of these four interfaces Fl, F2, F3, F4 may be any of a conical or conical ring shape and is roughly spherical in a preferred embodiment. In Figure 3, A-inlet/outlet channel 24a and B-inlet/outlet channel 24b are shown axially positioned in intermediate control plate element 24. Depending on the displacement setting of each radial piston machine and the torques applied to each drive shaft, one inlet/outlet channel will be at a relatively high pressure delivering fluid from the radial piston machine operating as a pump to the radial piston machine operating as a motor while the other inlet/outlet channel will be at a relatively low pressure delivering fluid in the opposite direction. A-tap line 24d connects A-inlet/outlet channel 24a with boost inlet port 24c while B-tap line 24e connects B-inlet/outlet channel 24b with boost inlet port 24c. Control valves 27 are installed in A-tap line 24d and B-tap line 24e. The control valves 27 function as relief valves in the event that over pressure occurs and also as check valves allowing makeup fluid to enter the drive circuit from boost inlet port 24c. Not shown is an optional loop flushing valve which can also be incorporated into intermediate control plate element 24 with similar porting as provided for in control valves 27. In a preferred embodiment, a loop flushing valve is not required as heat is removed from the system by conduction through the components as well as leakage from the pistons 11 and 19 and interfaces Fl, F2, F3, and F4.. In operation, one radial piston machine will act as a pump and the other as a motor. For example, referring to Figure 2, if torque is applied in a clockwise direction to first drive shaft 25 of first radial piston machine 2 and first cam ring 9 is eccentrically displaced by some percentage of El, first radial cylinder block 8 is forced to rotate in a clockwise direction as well and first radial piston machine 2 will act as a pump. Pass through channels 8a that are connected control ports lOg and 24k at that moment and the corresponding operating cavities 39 allow pistons 11 to draw fluid into the operating cavities 39. B-inlet/outlet channel 24b functions as the inlet channel for first radial piston machine 2 in this example.

As the first radial cylinder block 8 rotates further clockwise, each pass through channel 8a in turn passes the control ports lOg and 24k and is closed off by closing surfaces 10k and 24q. The operating cavity 39 will be at its maximum volume at that portion of rotation. Upon further clockwise rotation of radial cylinder block 8, each pass through channel 8a in turn are next connected to control ports lOf and 24j. At this portion of rotation, the pistons 11 push fluid out of operating cavities 39 into A- inlet/outlet channel 24a which functions as the outlet channel for first radial piston machine 2 in this example. Continuing this example, the fluid expelled from first radial piston machine 2 is pushed into second A- control port 24m through A-inlet/oulet channel 24a. In this particular example, A-inlet/outlet channel 24a functions as an inlet channel for first radial piston machine 3 in this example. Looking at Figure 4, second cam ring 17 is eccentrically displaced by some percentage of E3. Pass through channels 16a connected to corresponding operating cavities 40 and to control ports 24m and 118f allow fluid to enter operating cavities 40. The fluid pressure acting against the pistons 19 and the radial bores 16b force radial cylinder block 16 and second drive shaft 26 to rotate in a clockwise direction. Thus, second radial piston machine will act as a motor.

As the second radial cylinder block 16 rotates further clockwise, each pass through channel 16a in turn is closed off by closing surfaces 24r and the corresponding closing surface (not shown) on control plate element 118. The operating cavity 40 will be at its maximum volume at that portion of rotation. Upon further clockwise rotation of cylinder block 16, each of the pass through channels 16a are now connected to control ports 24n and 118g. At this portion of rotation, the pistons 19 push fluid out of operating cavities 40 into B-inlet/outlet channel 24b which functions as an outlet channel for second radial piston machine 3 in this particular example. Fluid is returned to first radial piston machine 2.

By varying one or both eccentricities of first and second cam rings 9 and 17 the fluid displacements of first and second radial piston machines 2 and 3. As such, variable speed and torque ratios between first and second drive shafts 25 and 26 respectively may be obtained.

In an alternate embodiment, first radial piston machine 2 and a second radial piston machine 3 are of different sizes and maximum displacements. In another alternate embodiment shown in Figures 8 and 8a, intermediate control plate element 24 is fixed from rotating but allowed to move axially within housing 5. Second radial piston machine 3 is replaced with alternate second piston machine 103; no compensating pistons or preload spring are used.

Second radial piston machine 103 comprises:

second radial cylinder block 16;

pistons 19;

- piston seals 20;

piston guide rings 21.

second cam ring 17;

second control plate element 18 and

intermediate control plate element 24.

Second control plate element 18 is constructed similarly to second control plate element 118 but does not contain bores and passages for any compensation pistons. Second control plate 18 is axially support directly by second end cap 7. A preload spring 39 between is not required between second control plate 18 and second end cap 7. In alternate this embodiment, the gaps at interfaces Fl, F2 F3 and F4 are all controlled simultaneously by the compensating pistons 14 and 15 in first control plate element 10. Additionally, if first radial piston machine 2 and second radial piston machine 103 are not of the same size and displacement, interfaces Fl and F2 need to be of a geometry yielding the same axial force as generated by the geometry at interfaces F3 and F4. In yet another alternate embodiment, intermediate control plate element 24 is fixed from rotating but allowed to move axially within housing 5. First and second control plate elements 10 and 118 are used in first and second radial piston machines respectively with compensating pistons installed in both control plate elements. In this alternate embodiment, a centering mechanism (not shown) is used to bias the intermediate control plate element to a desired neutral position.

In a yet further alternate embodiment, only one of first radial piston machine 2 or second radial piston machine 3 is variable with the other set at a fixed displacement.

In another alternate embodiment, a third radial piston machine is hydraulically connected to the first and second radial piston machines 2 and 3, respectively. In another alternate embodiment, the displacement control may be of pivoting versus sliding design. For instance here, cam ring 9 or 17 pivots about a pin located parallel to the rotational axis. A drive pin attached to the cam ring engages a linear actuator which controls the eccentric displacement of the cam ring.

In another alternate embodiment, variator 1 is integrated into a transmission housing. Housing 5 and end caps 6 and 7 are not present because all the structural requirements of the variator are supplied by the transmission housing. In another alternate embodiment, an accumulator system may be connected at either or both of A-tap line 24d and B-tap line 24e allowing energy to be stored in and released from the accumulator system.

An investigation of integration of the present invention into current and new IVT applications indicates that there are significant advantages of a variator based on radial piston machines versus conventional variators based on bent axis piston machines. These advantages include:

• higher power density (kw per kg)

• reduced number of parts

• more efficient packaging and integration

• cost reduction compared to current bent axis available components due to less complexity · easier retrofit into existing transmission designs replacing traditional power transfer devices such as a torque converter

Reference Numerals and Descriptions used herein

1 Variator

100 Variator

2 First Radial Piston Machine

3 Second Radial Piston Machine

103 Alternate Second Radial Piston Machine

4 Rotation Axis

5 Housing

5a Radial Surface

6 First End Cap

6a Radial Surface

7 Second End Cap

7a Radial Surface First Radial Cylinder Blocka Pass Through Channelb Radial Bore

c First Side Wall

d Second Side Wall

e Outer Enveloping Surfacef Center Plane

g Face

h Face

First Cam Ring

a Inner Enveloping Surface0 First Control Plate Element0a A-Control Channel

0b B-Control Channel

0c Small Compensation BoreOd Large Compensation BoreOe Bearing Face

Of A-Control Port

Og B-Control Port

Oh Radial Bearing Portion

0j Closing Surface

0k Closing Surface

1 Piston

2 Piston Seal

3 Piston Guide Ring

4 Small Compensation Piston5 Large Compensation Piston6 Second Radial Cylinder Block6a Pass Through Channel

6b Radial Bore

6c First Side Wall

6d Second Side Wall

6e Outer Enveloping Surface6f Center Plane

6g Face

6h Face Second Cam Ring

Second Control Plate Element

e Bearing Face

f A-Control Port

g B-Control Port

h Radial Bearing Portion

j Closing Surface

k Closing Surface

8 Second Control Plate Element

8a A-Control Channel

8b B-Control Channel

8c Small Compensation Bore

8d Large Compensation Bore

8e Bearing Face

8f A-Control Port

8g B-Control Port

8h Radial Bearing Portion

Closing Surface Not shown in Figures

Closing Surface Not shown in Figures Piston

Piston Seal

Piston Guide Ring

Small Compensation Piston

Large Compensation Piston Not shown in Figures Intermediate Control Plate Element

a A-Inlet/Outlet Channel

b B-Inlet/Outlet Channel

c Boost Inlet Port

d A-Tap Line

e B-Tap Line

f First Bearing Face

g Second Bearing Face

h Radial Bearing Portion

j First A-Control Port

k First B-Control Port

m Second A-Control Port 24n Second B-Control Port

24p Closing Surface

24q Closing Surface

24r Closing Surface

24s Closing Surface

25 First Drive Shaft

26 Second Drive Shaft

27 Control Valve

28 First Displacement Control

29 Second Displacement Control

30 Actuator Piston

31 Spring

32 Actuator Cap

33 Centering Spring Retainer

34 Shoulder Bolt

35 Actuator Control Valve

36 Fastener - Capscrew

37 Plug

38 Preload Spring

Loop Flushing Valve Not Shown in Figures

39 Operating Cavity

40 Operating Cavity

Fl First Interface

F2 Second Interface

F3 Third Interface

F4 Fourth Interface

El Total +ve Eccentric Displacement of First Cam Ring

E2 Total -ve Eccentric Displacement of First Cam Ring

E3 Total +ve Eccentric Displacement of Second Cam Ring

E4 Total -ve Eccentric Displacement of Second Cam Ring

All of the above U.S. patents. U.S. patent applications, foreign patents, foreign patent applications and non-patent publications referred to in this specification, are incorporated herein by reference in their entirety. While particular elements, embodiments and applications of the present invention have been shown and described, it will be understood, of course, that the invention is not limited thereto since modifications may be made by those skilled in the art without departing from the spirit and scope of the present disclosure, particularly in light of the foregoing teachings. Such modifications are to be considered within the purview and scope of the claims appended hereto.