BRPI0611908A2 - gear-shift orbital transmission - Google Patents

Abstract

ORBITAL TRANSMISSION WITH GEAR WITH GEAR. The present invention relates to a transmission including an orbital gear complex in combination with a variable hydraulic pump and motor. The input to the transmission is increased by orbital transmission so that when the pump and engine are not running, the orbiter is stationary and the orbital transmission produces an overdrive condition. A downshift is accomplished by rotating the web with the web rotation device, providing a high downshift. The pump and motor are preferably long piston hydraulic machines with infinitely variable oscillating plates. Hydraulic machines preferably have full gimbal stabilized agitators and elongated bore clamping plates for long pistons to eliminate possible impacts between the clamping plates and long piston head ends when the swinging plates are at or near their maximum angle of inclination.

Description

Patent Descriptive Report for "ORBITAL TRANSMISSION WITH GEAR WITH GEAR".

Related Order Reference

This application claims priority for co-pending U.S. application serial number 11 / 153.112, filed June 15, 2005, entitled "ORBITAL TRANSMISSION WITH GEARED OVERDRIVE". This request is incorporated herein by reference.

This application also claims priority for co-pending US Serial No. 11 / 153,111 filed June 15, 2005 entitled "DUAL HYDRAULIC MACHINE TRANSMISSION" which is a patent application in part of US Patent No. 6,983 .680, issued January 10, 2006 to Gleasman et al., Entitled "LONG-PISTON HYDRAULIC MACHINES", which is a continuation in part of Order Serial No. 10 / 647,557, filed August 25, 2003, entitled "LONG - PISTON HYDRAULIC MACHINES ", now abandoned, which is a continuation in part of the patent of origin serial number 10 / 229,407, filed August 28, 2002, entitled" LONG-PISTON HYDRAULIC MACHINES ", now abandoned. The patent and the above mentioned applications are hereby incorporated herein by reference.

Background of the Invention

Field of the Invention

The invention relates to the field of automotive transmissions. More particularly, the invention relates to an automotive transmission with orbital gear and a variable web rotation device.

Description of Related Art

Hydraulic pumps and motors with adjustable or swinging plates to prevent adjustable liquid free movement have been discussed for use in automotive transmissions for decades, but there have been difficulties constructing a lightweight hydraulic pump and motor that is powerful enough at speeds and pressures. required for use in a car. Traditionally, hydraulic pumps / motors with adjustable swing plates have a fixed oscillator and a rotating cylinder block. This arrangement works well for pumps / motors for applications such as golf carts and machinery, but for high pressure and high speed use in automobiles, a rotary cylinder block is too large, too heavy and too inefficient. Although prior art relating to hydraulic machines has shown stationary cylinder blocks and oscillating plates divided for nearly a century, no design has been commercially successful for use with the combination of high speeds and required pressures for automotive drives. The problem was primarily the difficulty of providing a sufficiently stable piston / oscillator interface.

In U.S. Patent No. 5,440,878, "VARIABLE HYDRAULIC MA-CHINE", issued August 15, 1995 to Gleasman et al., The piston / oscillator interface problem is considered. Long dog bones interconnect pistons and the oscillator, and to prevent collapse of dog bones under rotating stresses, the oscillator shaker is supported by a gimbal structure. A full gimbal frame can be used with fixed angle swing plates used in motors, but a car-dan medium is used on variable angle pumps. The prototypes of these machines exhibited undesirable vibrations and pulsations, indicating that the hydraulic machine could be improved.

In U.S. Patent No. 5,513,553, "HYDRAULIC MACHINE WITHGEAR-MOUNTED SWASH-PLATE", issued May 7, 1996 to Gleasman et al., An alternative to the cardan is shown. A spherical gear with spherical gear teeth mesh with gear teeth on the agitator to stabilize the dog and agitator bones by providing additional contact points as compared to the gimbal. This design, however, proved to be complicated to manufacture.

In U.S. Published Application No. 2004/0168567, "LONG-PISTONHYDRAULIC MACHINES", published September 2, 2004 to Gle-asman et al., Dog bones are replaced by long and ocardan pistons and spherical transmission are eliminated. A spring-loaded "hold" is used to maintain long piston shims in contact with the agitator. No restriction is required to prevent collapse as long pistons do not collapse under rotating stresses or in the absence of hydraulic pressure. However, there is a rotating stress imposed on the agitator by the high rotor speed of rotation, and the tensile effects cause undesirable inertial rotation of the agitator.

There is a need in the art for powerful, efficient, lightweight, and small enough variable hydraulic motors to be suitable for use in an automotive transmission.

In US Patent No. 6,748,817, "TRANSMISSION WITH MINI-MAL ORBITER", issued June 15, 2004, to Gleasman et al, a variable pump and motor are combined with a gear orbiter to form a infinitely variable transmission. In this transmission, as the hydraulic motor speed increases, the output shaft speed increases and the vehicle speed increases.

Although an internal combustion engine is the industry standard for automobiles in the United States, several leading car manufacturers are researching a homogeneous car-compression-ignition (HCCI) engine. In a conventional gasoline engine, the fuel - fuel mixture is ignited by a spark plug for power generation. In an HCCI engine, similar to a diesel engine, a piston compresses the air - fuel mixture to raise its temperature until it ignites. An HCCI engine is estimated to be capable of a 30% increase in fuel economy over a standard gasoline internal combustion engine. A major obstacle to implementing HCCI technology in automobiles is a difficulty in controlling combustion at low and high engine speeds.

There is a need in the art of a transmission which provides the power required for a car to operate while allowing its engine speed to remain in a low to moderate range where combustion in HCCI engines is controlled. -controlled more easily. Such a transmission allows for the implementation of more fuel efficient HCCI engines in gasoline powered vehicles.

Summary of the Invention

The transmission includes an orbital gear complex in combination with a variable pump and hydraulic motor that operates differently from known automotive transmissions. Unlike conventional transmissions: when the rotary core device rotates soul at its highest speed in the same direction as the engine, the transmission produces reverse output at its highest speed; then, when the soul is moved at a slightly slower speed, the transmission output produces a neutral (no output); thereafter, as the engine's thrusting of the web is slower, the transmission produces a downshift. When the soul is at rest, transmission provides an overdrive condition. When the core is rotated in a direction opposite to the engine, the transmission provides continuously higher gear ratios. The pump and motor are preferably long piston hydraulic machines with infinitely variable oscillating plates. Hydraulic machines preferably have stabilized agitators with full cardan shafts and long-bore clamping plates for the long pistons.

Brief Description of the Drawings

Figure 1 shows a partially schematic and cross-sectional view of a hydraulic machine with a variable oscillating plate angle. This hydraulic machine is used as the preferred pump and preferred motor for this invention.

Figure 2 shows a partially schematic and cross-sectional view of the hydraulic machine of figure 1 taken along plane 2-2 with parts being omitted for clarity.

Figure 3A shows a partially schematic view of a clamping plate where the oscillating plate is inclined at + 25 ° as seen from plane 3A-3A of Figure 1.

Figure 3B shows a partially cross-sectional view of a clamping plate and the piston clamping assembly, the view taken in plane 3B-3B of figure 1.

Figure 4 shows a cross-sectional view of a single cylinder with a long spring.

Figure 5 shows a partially schematic and cross-sectional view of a hydraulic machine with a split oscillating plate.

Figure 6 shows a view of a "closed loop" arrangement of two hydraulic machines as known in the prior art.

Figure 7 shows a hydraulic machine of the present invention with a full gimbal.

Figure 8 shows a clamping plate of the present invention with elongated holes, the view being taken along plane 8-8 of Figure 7.

Figure 9 shows an orbital transmission of one embodiment of the present invention.

Figure 10A is a roughly schematic diagram of the modular hydromechanical transmission of the invention in place behind a standard automotive engine.

Figure 10B is an end view of the modular transmission shown in Figure OA.

Detailed Description of the Invention

A transmission according to the invention includes a gear complex with an orbital web in combination with a web rotation device for variable gear ratios. Preferably, the web rotation device is a hydraulic variable pump and motor. The input to the transmission is increased speed by orbital transmission, so that when the pump and engine are not operating and the orbiter is stationary, the orbital transmission produces an overdrive condition. A downshift is accomplished by rotating the web with the web rotation device, providing high gear reduction. The transmission is suitable for automotive use. Although an orbital gear complex of the present invention may appear to be similar to the complex shown in U.S. Patent No. 6,748,817, the differences provide substantially different results. The relative size of the input gear and its combination grouping gear are reversed. Instead of a conventional input speed reduction, the input speed is increased by orbital transmission. Input speed reduction is controlled by hydraulics, and overdrive is achieved purely with orbital transmission. This shift eliminates the need for additional downshifting and simplifies the overdrive structure. In other words, when the hydraulic motor is stopped because the hydraulic pump is at a "zero" oscillator angle, the transmission output shaft is rotating faster than the input shaft.

The change in continuous gear ratio and infinite progression of the transmission of the invention occurs without any significant change in vehicle engine speed, and this continuous and infinite progression extends over a remarkably wide range from a predetermined low transmission ratio ( for example, 22: 1) even through an extended overdrive (for example, as high as 0.62: 1). The engine can be kept at a relatively low and efficient operating level (eg 500 rpm) for the entire throttle from one stop to overdrive. This feature not only results in fuel economy but, more importantly, a significant reduction in pollutants. This is particularly true for diesel demotor vehicles, as a selected demotor operating speed can be predetermined to an "ideal point" which optimizes performance. As is well known, when a diesel engine operates at a constant speed, it discharges few, if any, pollutants.

This same feature can provide significant fuel savings despite the fact that the engine runs well below the "ide-al" point for which the transmissions of the present invention are designed. The ideal point of an engine is the conventional optimal efficiency engine speed, which is the region of the efficiency map where the engine is most efficient at converting fuel to mechanical power. For most automotive engines, the optimum point is found in the region of 1500 rpm (maximum torque converter efficiency region for most conventional automatic transmissions). A transmission of the present invention improves fuel economy despite the fact that it can keep the engine at a speed where the engine is not running at maximum efficiency. This loss of efficiency is more than compensated for by reduced fuel requirements at lower operating speeds (eg at 500 rpm instead of 1500 rpm). An additional advantage of operating at these low engine speeds is reduced pressure on the hydraulic pump and motor, thereby reducing the duty cycle on the hydraulics and improving durability.

In terms of fuel economy, an orbital transmission of the present invention leads to city driving efficiency for motorway driving efficiency. Since driving speeds total around 60% of all driving, incorporation of a transmission of the present invention into automobiles should provide significant fuel savings. Obviously, since today's engines are designed to operate more efficiently at the 1500 rpm range, even greater fuel savings can be achieved by combining a transmission of the present invention with a motor engine designed to run more efficiently at about 500 rpm.

A transmission of the present invention is capable of varying drive shaft speed with minimal changes in engine speed. Thus, the present invention allows the engine speed to remain in a relatively narrow low to moderate range where the combustion of the newly proposed HCCI engine is most easily controlled. A transmission of the present invention is highly compatible with the implementation of more fuel efficient HCCI engines in gasoline powered vehicles.

The pump and motor of the present invention are preferably long piston hydraulic machines with infinitely variable oscillating plates. Both hydraulic machines preferably have split oscillating plates which include a rotating and nut rotor that is driven by the input shaft and a nut only stirrer that overlaps the rotor surface in bearings. In one embodiment of the invention, the bearings are needle bearings. The sliding wedges on the long plunger heads move in a "number eight" path over the nutation stirrer surface. However, during the relative movement of the sliding shims on the agitator, this "number eight" is actually a three dimensional dimension on the surface of an imaginary sphere having a diameter that decreases as the angle of the oscillator increases.

In long piston hydraulic machines of the present invention, the agitator is stabilized by a full gimbal. Depending on the bare agitator, the distribution of piston shim pressure in the agitator varies during each cycle as individual pistons change direction. This variable pressure tends to introduce unwanted vibrations into the stirrer stirring motion. The full gimbal helps maintain the stirrer-only movement and reduces unwanted vibrations. Shim sliding speed is restricted by the gimbal mounted agitator.

In another embodiment of the present invention, a lint plate is used to help maintain the piston shims against the agitated cardan mounted swing plate. The holes in the clamping plate are long rather than circular. A computer modeling shows that an increase in the elongation of the piston holes in the clamping plate, especially in the holes which are farthest from the two gimbal anchor points, eliminates an impact between the piston shims and the edges of the platform holes. clamping drive.

Although the hydraulic pump and motor of the invention may be used in combination as an independent transmission, the addition of the orbital gear complex allows a significant reduction in pump and motor size. In an orbital transmission of the present invention, the hydraulic pump and motor are never doing 100% of the work. Reducing load on the hydraulics as a result of the transmission also helps with pump and engine durability.

An orbital gear complex is combined with a variable web rotation device to form a minimal orbiter. The variable web rotation device can be any device capable of producing a variable output. An electric generator in combination with an electric motor may be used as the rotary device of the same. A variable brake can also be used as the web rotation device. In a preferred embodiment, a common variable hydraulic pump variable hydraulic motor serves as the web rotation device. Orbital transmission produces an overdrive condition when the soul is stationary. A gear reduction is accomplished by rotating the soul with the web rotation device, which allows high gear reduction. In a preferred embodiment, the pump and motor are preferably long piston hydraulic machines with infinitely variable oscillating plates with full cardan shafts and elongated holes in the sludge plate. In a preferred embodiment, a pair of hydraulic machines is used in combination with an orbital transmission system to form a transmission with a minimum orbiter. Hydraulics affect gear reduction and overdrive can be achieved with the transmission.

Long Piston Hydraulic Machine

Referring to Figure 1, a variable hydraulic machine 110 includes a modular fixed cylinder block 112. The cylinder block 112 has a plurality of cylinders 114 (only one being shown), in which a respective plurality of combination pistons 116 alternate between piston retracted position 116 and variable extended positions (the maximum length being shown at piston position 116 '). Each piston has a spherical head 118 which is mounted in a nip 120 at one end of the elongated axial cylindrical body portion 122 which is substantially as long as the length of each respective cylinder 114. Each spherical piston head 118 fits into a respective shim 124 which slides over a flat face 126 formed on the surface of a rotor 128 which is pivotally affixed to a drive element, specifically a shaft 130 which is supported on bearings with a hole in the center of the block. cylinder 112.

In one embodiment, the hydraulic machine 110 is preferably provided with a modular valve assembly 133 which is bolted as a plug at the left end of the modular cylinder block 112e and includes a plurality of reel valves 134 (only one shown) that regulate the fluid to and from cylinders114. In another embodiment, a plurality of check valves is alternatively used.

Machine 110 can be operated as a pump or as an engine. For operation as a motor, during the first half of each drive shaft revolution 130, a high pressure fluid from an inlet 136 enters the valve end of each respective cylinder 114 through a window 137 for each respective piston from from its retracted position to its fully extended position. During the second half of each revolution, a fluid at a lower pressure is withdrawn from each cylinder through window 137 and fluid outlet 139 as each piston returns to its fully retracted position.

For operation as a pump, during a drive shaft half-length 130, a lower pressure fluid is drawn into each respective cylinder 114 entering a window 137 from a closed loop of circulating hydraulic fluid through the inlet. 136 as each piston 116 is moved to an extended position. During the next half of each revolution, the actuation of each respective step 116 back to its fully retracted position directs a fluid at high pressure from window 137 to the closed hydraulic Iop through outlet 139. High pressure fluid then it is sent through an appropriate closed loop pipe (not shown) to a combining hydraulic machine, for example, the hydraulic machine 110 discussed above, causing the combining machine pistons to move at a varying speed. with the volume (gallons per minute - 1 gal / min = 3.785 l / min) of high pressure fluid being shipped in a well-known natechnic manner.

The cylindrical wall of each cylinder 114 in the modular cylinder block 112 is radially cut transversely by a respective lubrication channel 140 formed circumferentially therein. A plurality of passages 142 interconnect all lubrication channels 140 for forming a continuous lubrication passage in the cylinder block 112.

Each respective lubrication channel 140 is substantially closed by the axial cylindrical body 122 of each respective piston 116 throughout the entire stroke of each piston. That is, the outer circumference of cylindrical body 122 acts as a wall that surrounds each respective lubrication channel 140 at all times. Thus, even when the pistons 116 are alternating through maximum strokes, the continuous lubrication passage interconnecting all lubrication channels 140 remains substantially closed. The continuous lubrication port 140, 142 is simply and economically formed within the cylinder block 112.

During an operation of the hydraulic machine 110, all interconnected lubrication channels 140 are almost instantly filled by a minimum flow of high pressure fluid from the inlet 136 entering each cylinder 114 through the window 137 and being forced between the walls. cylinders and outer circumference of each piston 116. A loss of lubrication fluid from each lubrication channel 140 is restricted by a surrounding seal 144 located near the open end of each cylinder 114. However, the lubrication fluid in this Closed continuous lubrication passage of lubrication channels 140 flows moderately but continuously as a result of a continuous minimum flow of fluid between each of its respective cylindrical cylinder walls and the axial cylindrical body of each respective piston. response to a piston movement and pressures changing at each half shaft rotation engine 130 as the pistons alternate. As the pressure in each cylinder 114 is reduced to a low pressure, in the stroke around each piston 116, the fluid at a higher pressure in the otherwise closed lubrication passageway 140, 142 is again directed between the walls of each cylinder 114 and the cylindrical body outer circumference 122 of each piston 116 to the valve end of each cylinder114 experiencing such pressure reduction.

Lubricating fluid flow in the closed continuous lubrication passage 140 142 is moderate but continuous as a result of secondary minor fluid flow in response to piston movement and changing pressures at each half-drive shaft 130 cycle as the pistons alternate.

Pump rotor 128 110 is pivotably mounted to axle 130 about a geometry axis 129 which is perpendicular to axle geometry 132. Therefore, while rotor 128 rotates with driving axis 130, its inclination angle with respect to geometry axis 130 is preferably variable from 0- (i.e. perpendicular) to + 25 °. In Figure 1, Orotor 128 is inclined to + 25 °. This variable inclination is controlled as follows: the pivoting of the rotor 128 about the geometric axis129 is determined by the position of a sliding collar 180 which surrounds the drive shaft 130, and is axially movable relative thereto. A control link 182 connects collar 180 to rotor 128 so that a docolar motion 180 axially over the surface of the drive shaft 130 causes orotor 128 to pivot about geometry 129. For example, as the ocular 180 is moved to the right in figure 1, the rotor incline 128 varies an entire continuum from the + 25 ° inclination shown back to 0 ° (i.e. perpendicular) and then to -25 °.

The axial movement of collar 180 is controlled by the jaws 184 of a fork 186 as fork 186 is rotated about the geometric axis of a fork axle 190 by the articulation of a fork control arm 180. Fork 186 is actuated by a conventional linear servo mechanism (not shown) connected to the bottom of the fork arm 188. While the remaining fork e-elements 186 are all wrapped in a modular swing plate housing 192, and the fork axle 190 is supported on bearings fixed to housing 192, fork control arm 188 is positioned external to housing 192. Oscillating plate rotor 128 is balanced by a parallel connection 194 which is substantially identical to control connection 182 and is similarly connected to collar 180, but in a location exactly opposite the collar 180.

Referring to both Fig. 1 and Fig. 2, the cylindrical wall of each cylinder 114 is radially cut transversely by a respective lubricating channel 140 formed circumferentially therein. A plurality of passages 142 interconnects all lubrication channels 140 to form a continuous lubrication passageway in cylinder block 112. Each respective lubrication channel 140 is substantially closed by the axial cylindrical body 122 of each respective piston 116 during the entire stroke of each piston. That is, the outer circumference of cylindrical body 122 acts as a wall that surrounds each respective lubrication channel 140 at all times. Thus, even when the pistons 116 are alternating through maximum strokes, the continuous lubrication passage interconnecting all lubrication channels 140 remains substantially closed. The continuous lubrication passageway 140, 142 is simply and economically formed within the cylinder block 112, as may best be appreciated from the schematic illustration in Figure 2, in which the relative size of the fluid channels and the connection passages were exaggerated, for clarification.

During an operation of the hydraulic machine 110, all interconnected lubrication channels 140 are almost instantly filled by a minimum flow of high pressure fluid from the inlet 136 entering each cylinder 114 through the window 137 and being forced between the walls. cylinders and outer circumference of each piston 116. A loss of lubrication fluid from each lubrication channel 140 is restricted by a surrounding seal 144 located near the open end of each cylinder 114. However, the lubrication fluid in this Closed continuous lubrication passage of lubrication channels 140 flows moderately but continuously as a result of a continuous minimum flow of fluid between each of its respective cylindrical cylinder walls and the axial cylindrical body of each respective piston. response to a piston movement and pressures changing at each half shaft rotation engine 130 as the pistons alternate. As the pressure in each cylinder 114 is reduced to a low pressure, in the stroke around each piston 116, the fluid at a higher pressure in the otherwise closed lubrication passageway 140, 142 is again directed between the walls of each cylinder 114 and the cylindrical body outer circumference 122 of each piston 116 to the valve end of each cylinder114 experiencing such pressure reduction.

Referring to FIG. 3A and FIG. 3B, a clamping assembly for a hydraulic machine includes a clamping member 154 with a plurality of circular apertures 160, each of which surrounds the nip 120 of a respective piston 116. A oscillating plate are at an angle of +259 in FIG. 3A and FIG. 3B. Figure 3A shows the securing element 154 from the perspective of looking downward through the drive shaft 128, or from a plane 3A-3A of figure 1. A plurality of special washers 156 are positioned respectively between the deflection member 154 and each piston pad 124. Each washer 156 has an extension 158 which contacts the outer circumference of a respective pad 124 for maintaining the pad in contact with the rotor flat face 128 at all times. Each respective shim cavity is connected via a shim channel 162 and a piston channel 164 to ensure that the fluid pressure present at the shim-rotor interface is at all times equivalent to a fluid pressure at the head of each piston. 116

A fluid pressure constantly guides the pistons 116 toward the rotor 128, and the illustrated backplate assembly is provided to carry that load. However, at the required operating speeds for automotive use (eg 4000 rpm), additional orientation loading is required to ensure constant contact between piston shims 124 and flat face 126 of rotor 128. Hydraulic machines - Variables provide this additional orientation by using three single spring-oriented clamping sets.

The first clamping assembly for the hydraulic machine 110 includes a coil spring 150 which is positioned about the axis 130 and received in a suitable slot 152 formed in the cylinder block 112 circumferentially about the geometric axis 132. The coil spring 150 orientates a clamping member 154 which is also positioned circumferentially about the axis 130 and the geometrical axis 132. The clamping element 154 is provided with a plurality of circular apertures 160, each of which surrounds the nip 120 of a respective piston. 116. A plurality of special washers 156 are positioned, respectively, between the locking member 154 and each piston pad 124. Each of the washer 156 has an extension 158 which contacts the outer circumference of a respective pad 124 for maintaining the shim in contact with flat rotor 126 aface 128 at all times.

The positions of the oscillating plate and the piston shim clamping assembly change relative to each other as the incline of the motor 128 changes during a machine operation. With fluid flow at the relative position of these parts at the 0 ° inclination, each piston channel 164 has the same radial position with respect to each respective circular opening 160 at the fastener 154. At all other inclinations other than 0 °, the relative radial position of each piston channel 164 is different for each opening 160, and the relative positions of each special washer 156 is also different. The different relative positions in each of the new openings 160 themselves are constantly changing as the rotor 128 rotates and nurtures through a complete revolution on each incline. For example, at the 25 ° inclination shown in Figure 3A, if during each rotor revolution 128 someone were to look at the movement occurring through only the opening 160 at the top (i.e. a 12-hour clock position) of the fastener 154, the relative position of the parts seen at the top opening 160 would change serially to match the relative positions shown at each of the other eight openings 160.

At slopes other than 0 °, during each revolution of the rotor 128, each special washer 156 slides over the clamping element surface 154 as each shim 124 slides over the flat rotor face 126. Each of these parts changes relative to each other. its own opening 160 through each of the various positions that can be seen in each of the other eight openings 160. Each follows a cyclic path (which appears to trace a lemniscata, that is, a "hi-to number"), which varies in size. with the angular inclination of oscillating plate rotor 128 and the horizontal position of each piston 116 in the fixed cylinder block112. To ensure proper contact between each respective shim 124 and flat face 126 of rotor 128, a size is preferably selected to the boundaries of each opening 160 so that opening edges 160 remain in contact with more than half of the surface of each special washer 156. at all times during each revolution for all rotor inclines 128.

A second clamping assembly is shown schematically in Figure 4 in an enlarged partial and cross-sectional view of a single piston of a hydraulic machine 210. Each piston 216 is positioned in the modular fixed cylinder block 212 in a cylinder 214. , the last direction is radially cut transversely by a respective grease channel 240 formed circumferentially therein. In the same manner as described with respect to the other hydraulic machines already detailed above, each lubrication channel 240 is interconnected with similar channels in the other machine cylinders for forming a continuous lubrication passage in cylinder block 212. A seal Optional surrounding 244 may be located near the open end of each cylinder 214 for further fluid loss minimization of each lubrication channel 240.

The fixed cylinder block 212 does not include an axially circumferential coil spring or an axially circumferential slot for maintenance thereof. Hydraulic Modular Fixed Cylinder Block 212 210 can be connected to a modular fixed angle swivel plate assembly or a modular variable angle swing plate assembly, but in any case, the hydraulic machine 210 provides a much simpler subject matter. Specifically, the mode-fastening assembly includes only one respective conventional piston pad 224 for each piston 216 in combination with only one respective spirally spring 250, the latter also being associated with each respective piston 216.

Each piston shim 224 is similar to the conventional shims shown in the first clamping assembly and is mounted on piston head 218 to slide over the flat face 226 formed on the surface of the machine oscillating plate rotor 228. Each spring Spiral 250 is respectively circumferentially seated around the hydraulic valve window 237 at the valve end of each respective cylinder 214 and positioned within the body portion of each respective piston 216.

Each shim 224 slides over the flat face 226 of the rotor 228 by common movement of Iemniscata which varies in size with the horizontal position of each piston 216 and the inclination of the rotor 228 with respect to the geometrical knitting axis 232. During normal operation 210, the shims 224 are kept in contact with the flat face 226 of the oscillating plate by hydraulic pressure. Therefore, the spring orientation provided by spiral springs 250 is minimal but sufficient to maintain effective sliding contact between each shim 224 and flat face 226 in the absence of hydraulic pressure at the valve end of each respective cylinder214. The minimum spring orientation 250 not only facilitates mounting, but also prevents entrapment of minimal dirt and metal debris encountered during assembly and caused by wear.

Referring to Figure 5, a third clamping assembly for a hydraulic machine 310 includes an improved conventional split plate arrangement. A plurality of pistons 316, each including respective sliding block 324, alternate in respective cylinders 314 formed in cylinder block 312 which is identical to cylinder block 112. Each foot 324 slides on a flat face 326 formed on an agitator 327 which It is mounted on a combination rotor 328 by appropriate bearings 372, 374 which allow the agitator 327 to be rotated without a rotation, while the rotor 328 is rotated and rotated in a manner well known in the art. The inclination of the agitator 327 and rotor 328 about a geometric axis 329 is controlled by the position of a slide collar 180, a control link 382 and a parallel balance link 394.

Shims 324 are subjected to a fastening assembly substantially identical to the first fastening assembly, although the single large spiral spring 150 is replaced by a plurality of smaller individual spiral springs.

A clamping plate 354 is attached to the agitator 327. Each shim 324 receives the circumferential extension of a respective special washer 356, and the narrowing of each piston 316 is positioned in a corresponding plurality of respective openings 360 formed through the clamping plate. 354. While agitator 327 does not rotate with rotor328, agitator feed movement 327 is identical to rotor denutation motion 328 and, therefore, the relative movements between shims324 and agitator flat surface 327 are also identical to those in the first one. clamping assembly.

A plurality of individual coil springs 350 provide minimal spring orientation for maintaining effective sliding contact between each shim 324 and agitator flat face 327, in the absence of hydraulic pressure at the valve end of each cylinder 314. Spiral 350 is positioned circumferentially around each shim 324, being caught between each special washer 356 and a collar formed just above the bottom of each shim 324.

Referring to Figure 6, each hydraulic machine, whether a motor or a pump, is preferably paired with another hydraulic machine, a pump, or a combination motor, in a well-known closed loop arrangement. For example, the high pressure fluid exiting from the hydraulic machine outlet 139 is directly sent to the inlet 136 'of a 110' combination hydraulic machine, while the low pressure fluid exiting from the hydraulic machine outlet 139 '110 'is directly sent to the 110-combination hydraulic machine input 136. The 110 hydraulic machine and 110' hydraulic machine can be identical in structure except that the 110 hydraulic machine is used as a pump and the 110 'hydraulic machine be used as an engine. A portion of the fluid in this closed Ioop system is continuously lost to a "combustion exhaust" and is collected in a reservoir, and the fluid is automatically sent from the reservoir back to the closed Ioop to maintain a predetermined volume of fluid in the system. Ioop closed at all times.Full Cardan Hydraulic Machine

During recent coating work on a long piston hydraulic machine, vibration at higher speeds and pressures was noted. Some interference between the long piston heads and clamping plates also caused the bronze shims on the pistons to loosen. Repetitive impacts between the bronze shims and the debris plate increase the operating noise of these hydraulic machines. Although these hydraulic machines show a remarkably low combustion exhaust of lubricating hydraulic fluid compared to prior art hydraulic pumps and motors, a significant amount of this combustion exhaust results from loosening of the fourteen shims over time. Eliminating repetitive impacts significantly improves the performance of these machines. This undesirable interference occurs during the movement of the relative flush by the piston pads as they slide over the agitator portion of the split oscillating plate.

To improve the performance of the hydraulic machines of U.S. Patent Application No. 2004/0168567, the agitator is further stabilized, and this is accomplished with a full gimbal. The force applied by the slide shoe of each piston on the agitator has an axial component and a radial component. As the oscillating plate angle increases, the radial force component increases, and the full gimbal provides structural support to counteract this force and maintain the stirring movement of the agitator.

A split oscillating plate of a long-step hydraulic machine includes a rotation and nut rotor that is driven by the input shaft and a nut only "stirrer" that rises on the bearing surface on the bearings. The sliding shims on the long piston heads rotate in a "number eight" path over the surface of the denutation shaker. However, during the relative movement of the agitator sliding wedges, this "number eight" is actually a three dimensional limb on the surface of an imaginary sphere having a diameter that decreases as the angle of the oscillator increases.

In one embodiment of the present invention, the agitator is stabilized by a full gimbal. Depending on the bare agitator, the distribution of piston shim pressure in the agitator varies during each cycle as individual pistons change direction. This variable pressure tends to introduce unwanted vibrations into the nut stirrer movement. Cardanplen helps maintain the stirrer's only nourishing movement and reduces unwanted vibration. The speed of shim sliding is driven by the restricted gimbal shaker.

Referring to Figure 7, a long piston hydraulic machine 330 having a gimbal is shown. A plurality of pistons 316, each including a respective sliding shim, alternate in respective cylinders 314 formed in a cylinder block 312. Each shim 324 slips over the flat face formed on agitator 327 which is mounted on the combination rotor 328 by appropriate bearings. which enable agitator 327 to rotate while rotor 328 is rotated and rotated in a manner known in the art. The inclination of the agitator 327 and rotor 328 around axle 329 is controlled by the position of slide collar 380, control link 382 and balance parallel link 394. Shims are subjected to a fastening assembly substantially identical to the first assembly. although the single large coil spring is replaced by a plurality of smaller individual coil springs.

The cardan includes a fork 332, a first pair of caran pins 334 connecting fork 332 to agitator 327. Fork 332 forms a complete annular space around the agitator. Gimbal pins 334 are located 180 degrees from each other. The cardan pins 336 are located 180 degrees relative to each other and 90 degrees relative to the cardan pins 334. The cardan structure allows the agitator 327 to nute, but inhibits a rotary motion of agitator 327.

A clamping plate 338 is attached to the agitator 327. Each shim receives the circumferential extension of a respective special washer, and the tightening of each piston 316 is positioned in one of a corresponding plurality of respective openings formed through the clamping plate 338. Although the agitator 327 does not rotate with rotor 328, the agitation movement of agitator 327 is identical to rotational movement of rotor 328e, so the relative movements between shims and the flat surface of agitator 327 are also identical to those in the first gripping assembly.

A plurality of individual coil springs provide minimal spring orientation for maintaining an effective sliding contact between shim and agitator flat face 327 in the absence of hydraulic pressure at the valve end of each cylinder 314. Each spring Spiral is positioned circumferentially around each shim, being captured between each special washer and a collar formed just above the bottom of each shim.

In another embodiment of the present invention, the holes in the clamping plate are elongated rather than circular. The full gimbal is anchored at two points around the agitator. Computer modeling allows an increase in the elongation of the piston holes in the clamping plate, especially the holes which are the furthest from the two anchor points, to eliminate an impact between the piston shims and the edges of the clamping plate holes. .

With reference to fig. 8, a clamping plate 338 preferably has an opening 340 directly aligned with the two decardan pins 336 connecting fork 332 to agitator 327. Relative locations of gimbal pins 334, 336 with respect to clamping plate 338 are shown. schematically in figure 8. Aperture 340 is clearly circular. The shapes of the openings 340, 342, 344, 346, 348 are elongated the further the openings are from the cardan pins 336.

Computer models of this new design indicate that it is possible to use only a combination of two hydraulic machines, one as a pump and one as a motor, in a closed hydraulic Ioop, as an infinitely variable transmission for many models of the deliver them today. However, the inventors believe that the addition of the described orbital gear complex provides a more efficient transmission than these hydraulic machines acting alone. This hydraulic / orbital gear transmission provides a significantly improved automotive transmission that can be scaled up or down to suit a wide range of weight and size requirements.

Orbital Transmission

Although the hydraulic pump and engine can be used in combination as an independent transmission, the addition of the orbital gear complex allows a significant decrease in pump and engine size. In an orbital transmission of the present invention, the hydraulic pump and motor are never doing 100% of the work. Reduced load on the hydraulics as a result of the transmission also helps pump and engine durability.

A transmission 400 is shown in Figure 9 in one embodiment of the present invention. A motor 402 is shown connected to the transmission of the invention, which includes only a minimum orbiter 404 and a variable web rotation 403. In one embodiment, a variable web rotation 403 includes an electric motor 406 in combination with an electric generator 408 with an electrical connection 440 between them. In another embodiment, the variable speed web 403 includes a 406 variable hydraulic motor in combination with a variable hydraulic pump 408 with a hydraulic connection therebetween. The transmission is described below with a hydraulic pump and a hydraulic motor as a preferred web rotation device.

Orbiter 404 includes only one input gear 410 and one output gear 412, both mounted around a first geometry shaft 414 and a rotating-mounted mounting gear 416 about a second geometry axis 418 parallel to the first geometry shaft 414. input 410 is fixed for one revolution with motor drive shaft 420, while outboard gear 412 is fixed for one rotation with an output shaft 422. Gang gear 416 is fixed to an orbit shaft 424 supported for rotation on a web 426, and the web 426 itself is mounted to rotate around the first geometry axis 414, thereby allowing orbiting shaft 424 and grouping gear 416 to orbit, respectively, around the first geometry axis. 414 as well as around input gear 410 and output gear 412. Cluster gear 416 has two sets of gear teeth 428, 430 which engage anchor, respectively, with the teeth of the input gear 410 and output gear 412.

The gear tooth ratios between inlet gear 410 and grouping gear 428 and between grouping gear 430 and output gear 412 are selected such that when a web rotation 426 is avoided, output gear 412 rotates to a gear. predetermined gear rotation overdrive 410. This is in contrast to the teachings of the prior art, where a reduction in gear is taught, as in the aforementioned orbital transmission of US Pat. No. 6,748,817. For example, in a preferred embodiment of the present invention, gear tooth ratios are selected as shown in Table 1. Table 1

<table> table see original document page 25 </column> </row> <table>

With this example transmission, when a speed rotation 426 is avoided, the output gear 412 rotates to approximately 0.6: 1 overdrive of the input gear rotation 410.

A gear 432 is secured to the outside of the web 426 and engages with a motor gear 434 which is connected to the drive shaft 436. The drive shaft 436 may be made detachable from the motor gear 434 by an optional first clutch 438 which is preferably It is, of course, a simple clutch clutch, but it can be any kind of clutch. The clutch 438 provides a true Kom safety feature by ensuring that the vehicle is in neutral, especially at start, identity of the output of the 403 variable speed rotary device. For illustration purposes, the drive shaft 436 is driven by 406 hydraulic control motor and rotate motor gear 434 and web gear 432 in a 1: 1 ratio. The relationship between engine gear 434 and web gear 432 may fall into a range without departing from the spirit of the invention. This ratio, together with the gear teeth ratio, may be varied for the production of input-to-output transmission gear ratios in particular for particular settings of variable web rotation 403.

The control motor 406 is in turn operated by a hydraulic fluid delivered from the hydraulic pump 408 via a "closed loop" hydraulic circuit 440. An auxiliary drive gear 442 that is attached to the motor drive shaft 420 causes a rotation of a first combination gear 444 and a pump shaft 446 in a 1: 1 ratio.

Although the operation of hydraulic pump / motor combinations is well known in the art, an especially suitable pump / motor operation for this transmission is discussed in detail in this disclosure. The auxiliary rotation of pump shaft 446 by drive shaft 420 allows hydraulic pump 408 to create hydraulic fluid flow to control motor 406 according to the adjusted angle of pump 408 oscillating plate (not shown).

IVT Forward Operation

As an illustration of the operation of a transmission of the present invention, the orbital gear tooth ratios set forth above are used herein for the calculation of the values given in Table 2. Table 2 shows the conditions of the swing and pump plates. engine, resulting core rotation rate, output shaft rotation rate, and transmission arelation in discrete stages from reverse to overdrive. However, it must be understood that the infinitely variable transmission passes through a continuum of transmission relations over its entire range. An idle speed of 500rpm is used for sample calculations.

Table 2

<table> table see original document page 26 </column> </row> <table>

Referring to Table 2, it should be noted initially that as the angle of the motor swing plate 406 is continuously increased in a positive direction, the rotation rate of the motor shaft 436 gradually decelerates, thereby slowing down the speed. of the carrier belt 326 and causing the output shaft 422 to slowly increase its rotation rate.

With control motor 406 connected to core 426 and pump swing plate 408 set to its maximum inclination (ie 25 °), adjusting engine swing plate angle 406 to 10.9 ° causes shaft 426 to rotate. a speed that causes the output shaft 422 to stop, that is, for what is, in effect, a "neutral" condition.

For changing from "Neutral" to "Driving" condition, pump oscillator angle 408 is maintained at 25 °, while motor oscillator angle 406 is continuously increased to 25 ° where the transmission ratio equals 1: 1. . From "Driving" to a predetermined "Overdrive2", motor oscillator oscillator angle 406 is maintained at 25 °, while pump oscillator angle 408 is continually decreased to 0 °. When the pump oscillating plate 408 reaches 0 °, the core 426 is stopped, and the output gear 412 and output shaft 422 rotational speed is greater than the motor drive shaft 420e speed of the input gear 410 by the predetermined overdrive. For the basic gear complex referred to above.

As a further important feature of the present invention, the pump oscillator angle 408 may be reduced to a slightly negative angle, thereby reversing the motor shaft direction 436 to extend the "infinite overdrive" of the invention through a band at in addition to "Overdrive 2".

Therefore, as engine speed 406 and alma426 continuously decrease, output shaft forward rotation 422 continuously increases speed through an infinite range of gear ratios from a high range (significantly greater than 3: 1) to 1: 1 and then through continuous extended overdrive, without any gear shifting, clutch change, or any significant engine speed fluctuation. During any required speed reduction, such as braking, plate angles The oscillating knobs are adjusted in the direction opposite to the geared neutral position for proper gear reduction. Similarly, if additional horsepower is required when the vehicle is in a full overdrive, such as climbing a slope or passage, the swinging plate angle may be increased to provide a reduction in speed. proper gear or, obviously, engine speed can be increased.

Particular attention is drawn to the fact that the newly written continuous gear ratio and infinite progress (from start to overdrive) change occurs without any significant change in engine speed 402. The engine can be maintained at a relatively operational level. low and efficient throughout the entire throttle from a permanent stop to overdrive. This remarkable feature not only results in fuel savings but, more importantly, a significant reduction in pollutants. This is particularly true for diesel engine vehicles, since a selected engine operating speed can be predetermined at a "sweet spot", which optimizes performance. As is well known, when a diesel engine operates at a constant speed, it discharges few, if any, pollutants.

An additional feature of the invention provides a tow / winch mode for higher fuel efficiency and reduced duty cycle on hydraulics when the vehicle is hoisting heavy loads, towing a trailer, or traveling over steep terrain at speeds. from the highway. This feature deviates from the hydraulic system and locks driving in a 1: 1 transmission with the vehicle engine. This is accomplished by the addition of a second clutch 439 which is used to engage an additional gear 435 on the shaft to engage an additional gear 435 on the pump shaft 446. The gear 435 is interconnected in a gear train of and of the same size. that the auxiliary drive gear 442, first combination gear 444 and a web gear 432. Therefore, when first clutch 438 disengages motor 406 from motor gear 434 and second clutch 439 engages gear 435 to the drive shaft. At pump 446, the direct gear train of motor 402 through gears 442, 444, 435 and 432 rotates web 426 at the same speed as motor 402 is rotating input gear 410. This results in direct geared driving of both output gear 412. and the output axle motor 422 in a 1: 1 ratio with motor 402.

"Stop" and Operation Behind IVT

When the 406 control motor is connected to the 426 core, and the 408 pump oscillating plate is set to its maximum inclination (ie 25 °), the 406 motor oscillating plate is adjusted to obtain the "Dead Spot" (ie , at 10.9 ° for previously given gear ratios).

Under these conditions, the output shaft 422 is stationary. As indicated above, this in effect provides a "geared neutral" in which the soul 426 is held at constant torque in a stopping position for starting and when starting in traffic. However, it should be noted that at any time under these conditions, clutch 438 can be disengaged and a "true neutral" can be obtained to completely disconnect the wheel drive of the vehicle.

If the angle of the 406 engine swing plate is moved in a slightly negative direction from its idle gear setting (for example, from 1 to 3 °), the control motor 406 will continue to arm core 426 in the same direction. that gets a "geared neutral", but the 426 will rotate at a slightly faster speed. The net effect is that the output shaft 422 now rotates at a relatively high gear reduction in the backward direction, that is, in "Reverse".

When the setting of the motor swing plate 406 is continuously increased in a negative direction (ie, in addition to the setting used to bring the output gear 412 to a stop), the web revs 426, output gear 412 and output shaft 422 will all continuously increase in the forward direction. If the control motor 406 is "neutralized" (for example, by disengaging clutch 438), an idle speed rotation of input gear 410 will automatically cause the clutch gear 416 to rotate line 426 in one direction. reverse direction at the exact speed that causes the 412 output gear to see a complete stop. That is, when the web rotation control is neutralized, the minimum orbiter of this invention automatically seeks the minimum torque position.

Therefore, it may not be necessary to precisely program the tuning of the motor swing plate 406, and in order to create the required predetermined speed reversal of the web to bring the transmission to zero speed upon stopping the vehicle. A preferred mode of hydraulic pump / motor has been developed for the invention which, without a first clutch 438, still allows the control motor406 to be properly adjusted to allow the vehicle to come to a complete stop whenever the speed of the input gear410 is reduced to idle engine speed.

As is well known in the art, supplementary power axles are often provided on tractors and trucks to allow auxiliary equipment to be operated from the vehicle engine. Therefore, another feature of the transmission is a supplemental power package 450 that includes a supplemental power shaft 452 and a supplemental power gear 454 connected by a clutch 456.

Supplementary power gear 454 is driven by auxiliary drive gear 442. Supplementary power gear 454 is generally "freewheeling" being disconnected from the supplemental power shaft 452 by the normally disengaged clutch 456.However, when clutch 456 is engaged, the supplementary power shaft 452 also rotates to operate auxiliary equipment.

Hydraulic Bypass Circuit

A valve-regulated "bypass" assembly is preferably incorporated in the closed Ioop hydraulic circuit 440 shared by the hydraulic pump 408 and motor 406. Such a bypass arrangement is shown in US Patent No. 6,748,817 referenced. above. A pair of "bypass" passages connect to opposite sides of the closed Iop and pass through a cylinder, being blocked by the piston portions of a reel valve. A pair of rods are located in the reel valve, so that when the reel valve is moved in one direction, the rods allow a hydraulic fluid to flow through the bypass passages. A sensor responds to upper and lower levels in the selected vehicle operating parameters (eg vehicle speed and / or hydraulic pressure at closed Ioop). Detecting a first level of these selected parameters causes the reel valve to move one direction to open passages (for example, whenever vehicle speed is reduced and approaching a stopped condition), while detecting a second level restores valve to the opposite position, returning the closed Ioop hydraulic circuit 440 to its normal condition.

An activation of the shift circuit reel valve allows the 436 axis of the control motor 406 to be moved independently, although the pump swing plate 408 is being energized or is being stopped at 0 °. Therefore, the shift assembly may be used to reduce the transmission load during a motor part, thereby replacing a vehicle flywheel clutch. In this sense, since the sensor can be used to detect a significant change in fluid pressure in the closed loop 440 hydraulic circuit, the bypass assembly can also serve as a safety device, preventing any exceptional overload of the hydraulic system.

As an additional feature of an orbital transmission of the present invention, the efficiency of the pump and hydraulic motor in combination with the orbital transmission allows sufficient power to be transmitted to the wheels while the engine speed is maintained at around 500rpm. This feature provides significant fuel savings despite the fact that the engine runs well below the "sweet spot" for which the transmissions of the present invention are designed. The ideal point of an engine is the conventional optimal efficiency engine speed, which is the efficiency range where the engine is most efficient at converting fuel to mechanical power. For most automotive engines, the ideal spot is found in the 1500 rpm region. A transmission of the present invention improves fuel economy despite the fact that it can keep the engine at a reduced efficiency speed. This loss in efficiency is more than offset by reduced fuel requirements at 500 rpm. An additional advantage of running at these low engine speeds is a reduced pressure demand on the hydraulic pump and motor, thereby reducing the hydraulic duty cycle and improving its durability.

In terms of fuel economy, an orbital transmission of the present invention leads to city driving efficiency for motorway driving efficiency. Since driving speeds total around 60% of all driving, incorporation of a transmission of the present invention into automobiles should provide significant fuel savings. Engines are currently designed to run more efficiently in the 1500 rpm range. Of course, even greater fuel savings can be achieved by combining a transmission of the present invention with an automotive engine designed to run more efficiently at about 500 rpm.

A transmission of the present invention is capable of varying drive shaft speed with minimal changes in engine speed. Thus, the present invention allows the engine speed to remain in a relatively narrow low to moderate range where combustion of the newly proposed HCCI engine is most easily controlled. In a preferred embodiment, engine 402 is an HCCI engine. A transmission of the present invention is highly compatible with implementing more fuel efficient HCCI motors in gasoline powered vehicles.

A final feature of the invention is now described with reference to figure 10A and figure 10B which schematically illustrate the modular hydromechanical transmission of the invention in place behind a standard automotive engine 402. In the same general arrangement as described above , the output shaft rotation 414 of the unique orbiter of invention 404 is regulated by motor hydraulics 406 and pump 408, which in turn are driven by motor 402. As shown, orbiter 404 is modularly mounted on the plate. 404a which is bolted to the flywheel housing 502 at the rear of the engine 402. Similarly, the engine 406 and the pump 408 are modularly mounted on the plate 408a.

For some vehicles, the 406 and 408 modular engine combination can meet the full-drive requirements of the vehicle.

Under these circumstances, modular orbiter 404 may be omitted, and minimal commodities, engine 406, pump 408, and plate 408a may be bolted modularly directly to flywheel housing 502 in the rear of engine 402.

Accordingly, it is to be understood that the inventive embodiments described herein are merely illustrative of the application of the principles of the invention. A reference herein to the details of the illustrated embodiments is not intended to limit the scope of the claims, which embody those features considered essential to the invention.

Claims (22)

1. Transmission to a prime mover, the transmission comprising: an orbiter comprising: an input gear mounted on a first geometry axis and responsive to an input drive provided by the prime mover; first eixogeometric; and at least one meshed assembly gear with said inlet and outlet gears and mounted for rotation on an orbit axis positioned parallel to said first eixogeometric; an orbiter core supporting said recessed orbit axis for rotation about said first geometry axis, to allow the orbit axis and clustering gear to orbit said first geometry axis and said gears respectively. wherein the gear tooth ratios between the assembly gear and said inlet and outlet gears are selected such that when a rotation of the web is prevented, a rotation of said input gear produces a rotation of said output gear to a predetermined overdrive of the input drive. Transmission according to claim 1, which further comprises a variable web rotation device operably connectable to the orbiter web for controlling web rotation and wherein when said web is rotated in a first direction as shown in FIG. as the speed of rotation of the web increases, said predetermined overdrive input drive is decreased to a ratio of 1: 1 and thereafter to a downshift of the input drive which increases relative to said web rotation speed. Transmission according to claim 2, wherein rotating the web in said first direction at a predetermined speed for rotation of said output gear, and rotating the web in said first direction at a speed greater than said speed. predetermined reverses the direction of rotation of said output gear. Transmission according to claim 2, wherein the rotation of the soul in a second direction opposite the first direction provides for an infinitely variable overdrive greater than said predetermined overdrive. Transmission according to claim 2, further comprising a first clutch for the selective connection of said variable web rotation device and said orbiter web. Transmission according to claim 5, further comprising: a plurality of consecutively interconnected gears for forming a separate gear train from said orbiter; a second clutch for the selective connection of said inlet relationship provided by said primary motor to said orbiting shaft, wherein when said first clutch disconnects said variable web rotation device and said orbitantee web when said second clutch connects the said inlet drive and said orbiter web, said output gear and said orbiter web both rotate 1: 1 with said inlet drive. Transmission according to claim 2, wherein said variable web rotation device comprises an electric motor in combination with an electric generator which is driven by said motor. Transmission according to claim 2, wherein said variable web rotation device comprises a variable hydraulic motor driven by a variable hydraulic pump which is driven by said motor. Transmission according to claim 8, wherein said variable hydraulic motor and variable hydraulic pump are each enclosed in a respective housing and comprise: a cylinder block having a plurality of cylinders formed in the cylinder block and positioned circumferentially at a first radial distance about the geometric axis of rotation of a drive element, a plurality of respective pistons alternatively mounted on said cylinders, each piston comprising a piston body and a a ball head connected to the piston body by a tapered channel and each respective cylinder having an open head portion beyond which the piston head extends at all times, a split oscillating plate driven by said drive member and comprising: a rotor inclined at an angle rotating and bare variable; an agitator having a flat face which is only bare, wherein the stroke of each piston varies according to the inclination of said upward oscillating plate to a predetermined maximum; and a sliding chock affixed pivotally and directly to each piston head without any intermediate dog bone, each sliding chock being held in direct sliding contact with said flat face during all relative rotational movements between the piston and the flat face. Transmission according to claim 9, said hydraulic pump and hydraulic motor each further comprising a cardan which comprises an annular fork connected to said housing by a first pair of 180 ° spaced cardan pins and connected to the agitator of said plate. The oscillating beam is divided by a second pair of also 180 ° spaced cardan pins, wherein each cardan pin of said first pair is spaced 90 ° from each cardan pin of said second pair. Transmission according to claim 10, in which the hydraulic pump and hydraulic motor are added, each further comprising a clamping assembly for guiding each sliding block towards the flat face. Transmission according to claim 11, wherein each piston body has an elongate axial cylindrical length sufficient to be supported on the respective cylinder to ensure minimal lateral displacement of the piston head when the shim is on a relative sliding contactor. face flat at all times during piston stroke. Transmission according to claim 10, wherein said split oscillating plate further comprises bearings for supporting said agitator and said rotor. Transmission according to claim 10, wherein the hydraulic pump and hydraulic motor are further comprising: a respective lubrication channel formed in the cylindrical wall of each cylinder in the cylinder block for retaining a pressurized fluid; lubrication channels being interconnected to form a continuous lubrication passage in the cylinder block, wherein the pressurized fluid is retained in said continuous lubrication passage by substantially closing each respective lubrication channel through the outer surface of the axially cylindrical body of each piston. During the entire stroke of each piston, the only pressurized fluid source received by the continuous lubrication passage is a minimum flow between each respective cylindrical wall and the cylindrical body of each respective piston; and wherein said continuous lubrication passage is formed entirely within the cylinder block, transversely cuts each cylinder and is substantially circumferentially centered at the same distance as the cylinders are centered about the geometric axis of rotation of said drive member. . Transmission according to claim 11, wherein said clamping assembly comprises: a clamping element having a plurality of respective openings, the boundary of each respective opening being located near a narrow narrowing of each respective piston. ; a respective washer adapted around the narrow narrowing of each piston between the clamping plate and its sliding shim, each respective washer having a cylindrically aligned extension to circumferentially contact each said sliding shim; wherein said washers are in sliding contact with said locking plate for movement relative thereto, in response to the change of relative positions of the sliding wedges, when the flat face is inclined with respect to the geometric axis of rotation of said drive element. . Transmission according to claim 15, wherein the edge of each respective opening in the clamping plate is designed to contact more than half of the outer circumference of each washer at all times during relative movements. Transmission according to claim 15, wherein said apertures in the clamping plate are elongated, said elongation being relatively greater the further each hole is positioned from said first pair of card pins. dan supporting said gimbal. Transmission according to claim 15, wherein the hydraulic pump and the hydraulic motor are added, each further comprising a minimum spring orientation sufficient to maintain the effective sliding contact between each respective shim and the flat face in the absence of hydraulic pressure. at one valve end of each respective cylinder. Transmission according to claim 18, wherein said minimum spring orientation is provided by a plurality of springs, each spring being positioned respectively between the clamping plate and one of the respective washers. Transmission according to claim 18, wherein the minimum spring is provided by a coil spring positioned circumferentially about the geometric axis of rotation of the drive element at less than the first radial distance for orientation of the bearing plate. - rejection against the washers. 21. A hydromechanical transmission for a primary engine, the transmission comprising: a variable hydraulic component driven by a drive axle of said engine and producing a hydraulic output; and a mechanical component driven by said drive axle and producing a mechanical output, wherein said hydraulic output and said mechanical output are combined to drive an output shaft to drive a vehicle such that an increase in the speed of the hydraulic component decreases. the ratio of the speed of said output shaft with respect to the speed of said drive shaft. Transmission according to claim 21, wherein said transmission provides a power to said engine output shaft rotating at a speed well below a conventional optimal efficiency engine speed.