HK1142942B - Infinitely variable transmissions, continuously variable transmissions, methods, assemblies, subassemblies, and components therefor - Google Patents

Description

Infinitely variable transmissions, infinitely variable transmissions and methods, assemblies, subassemblies, and components thereof

RELATED APPLICATIONS

This application claims the rights granted by provisional application No. us 60/890,438 filed on 16/2/2007 and incorporated herein by reference.

Technical Field

The field of the invention relates generally to methods, systems, devices, assemblies, subassemblies, and/or components of mechanical power transmissions, and more particularly, infinitely variable transmissions or infinitely variable transmissions.

Background

In some systems, power is expressed as torque and speed. More specifically, power in these systems is generally defined as the product of torque and rotational speed. The transmission is typically connected to a power input that provides an input torque at an input speed. The transmission is also connected to a load that requires an output torque and an output speed, which may differ from the input torque and the input speed. Generally, the prime mover provides the power input to the transmission, while the driven device or load receives the power output from the transmission. The primary function of the transmission is to regulate the power input to deliver a power output to a driven device at an appropriate ratio of input speed to output speed (the "speed ratio").

Some mechanical transmissions include step-ratio, discrete or fixed ratio type transmissions. These transmissions are used to provide discrete or stepped speed ratios within a particular ratio range. For example, such transmissions may provide 1: 2, 1: 1, or 2: 1 speed ratios, but may not provide intermediate speed ratios such as 1: 1.5, 1: 1.75, 1.5: 1, or 1.75: 1. Other drives include a transmission, commonly referred to as a continuously variable transmission (or "CVT"), which includes a continuously variable transmission. Compared to a step transmission, a CVT is designed to provide any fractional ratio within a particular range. For example, in the above ranges, a CVT can generally provide any desired speed ratio between 1: 2 and 2: 1, including ratios of 1: 1.9, 1: 1.1, 1.3: 1, 1.7: 1, and the like. While others employ infinitely variable transmissions (or "IVTs"). Like CVTs, IVTs are also capable of producing any ratio within a specific range of ratios. However, in contrast to a CVT, an IVT can only deliver zero output speed at a steady input speed ("zero power" state). Therefore, if the speed ratio is defined as the ratio of the input speed to the output speed, the IVT is able to (at least theoretically) produce an infinite set of speed ratios, and therefore the IVT is not limited to a particular ratio range. It should be noted that some transmissions provide IVT functionality using continuously variable transmissions connected to other transmissions and/or clutches. However, as used herein, the term IVT is primarily understood to mean a continuously variable transmission that does not have to be connected to additional transmissions and/or clutches to produce IVT functionality.

The field of mechanical power transmissions refers to several types of continuously variable transmissions or infinitely variable transmissions. For example, one well-known continuously variable transmission is a belt and variable radius pulley type transmission. Other existing transmissions include hydraulic, conical disk roller and conical ring transmissions. In some cases, these transmissions are connected to other transmissions to provide IVT functionality. Some hydromechanical transmissions may provide infinite ratio changes without other transmissions. Some continuously variable transmissions and/or infinitely variable transmissions are classified as friction or traction type transmissions because they rely on dry friction or elastohydrodynamic traction, respectively, to transmit torque in the transmission. An example of a traction-type transmission is a ball-type transmission in which the spherical elements are sandwiched between torque-converting elements and a thin layer of elastic power fluid acts as a torque-converting conduit between the spherical elements and the torque-converting elements. Most closely related to the inventive embodiments disclosed in this application is the latter transmission.

There is a continuing pressure in the CVT/IVT industry to improve transmissions and speed change devices, increase their efficiency and packaging flexibility, simplify operation, and reduce cost, size, and complexity. Inventive embodiments of CVT and/or IVT methods, systems, subassemblies, components, etc., disclosed below address some or all aspects of this need.

Disclosure of Invention

The systems and methods described herein have several features, no single one of which is intended to produce the desired attributes. Without limiting the scope of the claims which follow to a greater degree, the more important features of the invention will be discussed briefly below. After considering this discussion, and particularly after reading the section entitled "detailed description of certain embodiments" one will understand how the features of the systems and methods provide several advantages over conventional systems and methods.

One aspect of the invention relates to a carrier input cover for an infinitely variable transmission with a planet-pivot arm assembly. In a certain embodiment, the carrier input cap has an approximately circular body with a central aperture. In one embodiment, the carrier input cap includes a set of carrier pins angularly arranged about the central aperture. The carrier pin carries a first set of fluid passages. The carrier input cover includes a set of surfaces formed on the carrier pin. The surface may be designed to couple to the planet-pivot arm assembly.

Another aspect of the invention relates to a carrier centerblock for an infinitely variable transmission with a planet-pivot arm assembly. The carrier center block comprises an approximately circular body with a central bore, and further comprises a neck extending axially from the circular body and concentric with the central bore. In one embodiment, the carrier center block includes several carrier pins angularly aligned around and radially extending from the central bore. The carrier pin carries a first set of fluid passages. The carrier center block also has a set of surfaces formed on the carrier pin. The surface may be designed to couple to the planet-pivot arm assembly.

Another aspect of the invention relates to a carrier output block for an infinitely variable transmission with a planet-pivot arm assembly. The carrier output block includes a generally circular body with a central bore and a shaft extending from the circular body and coaxial with the central bore. The shaft end is spline type. In a certain embodiment, the carrier output block includes several carrier pins angularly aligned around and radially extending from the central bore. The carrier pin carries a first set of fluid passages. In one embodiment, the carrier output block also has several surfaces formed on the carrier pins. The surface may be designed to couple to the planet-pivot arm assembly.

In addition, another aspect of the invention relates to a pivot arm for an infinitely variable transmission. The pivot arm includes a first arm bar with a first bearing hole and a pivot hole at an end of the first arm bar distal to the first bearing hole. In one embodiment, the pivot arm further includes a second arm bar attached to one end of the pivot hole. The second arm has a second bearing hole at an end of the second arm distal to the pivot hole. At least one of the first and second arms has a plurality of lubricant passages.

One aspect of the invention relates to a pivot arm of a variator of an Infinitely Variable Transmission (IVT) having a planet axle. The pivot arm has a central pivot bore from which the first extension extends. The pivot arm has a second extension extending from the central pivot bore opposite the first extension. In one embodiment, the pivot arm has first and second axle holes in the first and second extensions, respectively. The shaft holes may be used to support the planet shafts.

Another aspect of the invention relates to a planet axle of the transmission. The planet axle has an elongated body with a cylindrical central portion. The cylindrical central portion of the planet shaft also has a plurality of grooves. Wherein at least one groove is used for supporting the elastic ball and at least one groove is used for supporting the retainer clip. In a certain embodiment, the first neck of the planet axle is located at one end of the elongated body. The first neck may be designed to extend radially toward the longitudinal axis of the elongated body and taper. The first neck may also be designed to expand radially toward the cylindrical central portion. The planet axle may position a second neck at the other end of the elongated body, opposite the first neck. The second neck may be designed to extend radially toward the longitudinal axis of the elongated body and taper, or may be designed to expand radially toward the cylindrical center portion.

In addition, another aspect of the invention is directed to a planet axle for an infinitely variable transmission pivot arm assembly. The planet axle has an elongated body with a cylindrical central portion. The cylindrical center section has a set of eccentric grooves. In one embodiment, the first cylindrical portion of the planet axle extends from and is coaxial with the cylindrical center portion. The first cylindrical portion has a smaller diameter than the cylindrical central portion. The second cylindrical portion of the planet axle extends from and is coaxial with the first cylindrical portion. The second cylindrical portion has a smaller diameter than the first cylindrical portion. In one embodiment, the third cylindrical portion of the planet shaft extends from and is coaxial with the second cylindrical portion, and the third cylindrical portion has a smaller diameter than the second cylindrical portion.

In another aspect, the invention relates to an input shaft for an infinitely variable transmission with a hydraulic system. The input shaft includes a cylindrical body with a central bore. The cylindrical body can be designed as a valve housing of the hydraulic system. In one embodiment, the input shaft has a manifold flange extending from a first end of the cylindrical body. The splined portion of the input shaft may extend from the second end of the cylindrical body. The outer surface of the flange of the input shaft may also have several grooves which may be used to mate with the hydraulic system. In one embodiment, the plurality of fluid passages of the input shaft are all aligned on the outer surface of the flange.

Another aspect of the invention relates to a transmission input shaft with an elongated body having an outer surface. The input shaft includes an intermediate cavity formed in the elongated body. The input shaft may have several fluid passages for providing fluid communication between the intermediate cavity and the outer surface. In one embodiment, the input shaft has a spool valve positioned in the intermediate chamber.

One aspect of the invention relates to an input shaft for an infinitely variable transmission with a hydraulic system. The input shaft includes a cylindrical body with a central bore. In one embodiment, the input shaft has a manifold flange extending from the first end of the cylindrical body. The input shaft may include a splined portion extending from the second end of the cylindrical body and several sealing grooves formed in the periphery of the cylindrical body. The seal grooves may be used to provide several fluid pockets arranged between the seal grooves. In one embodiment, the input shaft has several fluid ports arranged around the periphery of the cylindrical body. A fluid interface may be arranged between each seal groove.

Another aspect of the invention is directed to a fluid manifold for an infinitely variable transmission having a hydraulic system. The fluid manifold is a circular body with a first face, a second face, and a central bore. In one embodiment, the fluid manifold has a lubrication fluid port located at a periphery of the first face. The fluid manifold has several lubrication fluid passages for fluid communication with the lubrication fluid port. The lubrication fluid passages may be angularly spaced about the central bore and the lubrication fluid passages may be arranged on the second face. In one embodiment, the fluid manifold has a line pressure port located at a periphery of the first face. The fluid manifold has a line pressure fluid passageway for fluid communication with the line pressure port. The line pressure fluid passages are arranged on the second face. The fluid manifold has a pilot pressure port located at a periphery of the first face. In one embodiment, the fluid manifold has a pilot pressure fluid passage configured to be in fluid communication with a pilot pressure port. The pilot pressure fluid passage is arranged on the second face.

Another aspect of the invention relates to a pivot pin hub for an Infinitely Variable Transmission (IVT). The pivot pin hub includes a cylindrical body with a central bore. The pivot pin hub has a plurality of pin pairs angularly distributed about and concentric with the central bore. The pin pair extends longitudinally from the central bore. The pivot pin hub includes a first face of the cylindrical body with a flat surface. In one embodiment, the pivot pin hub has a second face of the cylindrical body with a plurality of grooves for coupling with a lock washer of the IVT.

In addition, another aspect of the invention relates to a control piston for an Infinitely Variable Transmission (IVT). The control piston includes a cylindrical body with a central bore. In one embodiment, the control piston has a flange at the first end of the cylindrical body. A flange extends longitudinally from the central bore. The grooves of the control piston are arranged on the cylindrical body. The groove may be located at the second end of the cylindrical body and may be used to support a lock washer of the IVT. The control piston has a seal groove located at the periphery of the flange.

One aspect of the invention relates to a traction ring for an infinitely variable transmission. The pull ring includes an eyelet. In one embodiment, the traction ring has a cylindrical surface formed on one side of the annular ring. The traction ring includes a traction surface extending from the cylindrical surface toward the inner periphery of the grommet. The traction surface and the cylindrical surface are intersected to form an angle. The traction ring has a set of splines located on the periphery of the eye.

Another aspect of the invention relates to a drive flange for an infinitely variable transmission. The drive flange includes an annular cylindrical body with a first end and a second end. The first end is located at the end of the second end. The drive flange has a plurality of splines arranged on an inner diameter of the first end. The drive flange has an end cap on the second end. The end cap has a central aperture.

In addition, another aspect of the invention relates to a reaction flange for an Infinitely Variable Transmission (IVT). The reaction flange includes an annular cylindrical body with a first end and a second end. In one embodiment, the reaction flange has a plurality of splines arranged about the inner periphery of the first end. The reaction flange has a flat surface on the second end. The flat surface may be used to react axial forces during IVT operation. The flat surface also has several dowel grooves.

In another aspect, the invention relates to a torque transmitting coupling for an Infinitely Variable Transmission (IVT). The torque transmitting coupling includes an annular cylindrical body with a first end, a middle portion, and a second end. In one embodiment, the torque transmitting coupling has a first set of splines arranged on an inner periphery of the first end and a second set of splines arranged on an inner periphery of the second end. In addition, there is a third set of splines arranged on the periphery of the intermediate portion.

Another aspect of the invention relates to a reaction flange for an infinitely variable transmission with a traction ring. The reaction flange includes a circular body with a first end, a second end, and a central bore, and a first set of splines arranged about an inner periphery of the first end. The first set of splines may be used to connect with a traction ring. The reaction flange also has an end cap at the second end. The end cap has a splined central bore.

One aspect of the invention relates to an input cam flange of an infinitely variable transmission. The input cam flange includes a cylindrical tubular body with a first end and a second end. In one embodiment, the input cam flange has a plurality of splines arranged about an inner periphery of the first end. The flange of the input cam flange extends from the periphery of the cylindrical tubular body. The flange of the input cam flange is provided with a group of cam extension lines. The cam extension line has a group of counterclockwise spiral extension lines and a group of clockwise spiral extension lines. The input cam flange also has a neck extending from the flange.

Another aspect of the invention is directed to a cam base for an infinitely variable transmission. The cam base comprises a hole ring, and a group of cam extension lines are arranged on the surface of the hole ring. The set of cam extensions includes a set of counterclockwise spiral extensions and a set of clockwise spiral extensions. The cam base also has several dowel recesses arranged around the periphery of the annular ring.

Another aspect of the invention relates to a cam-loaded piston for an Infinitely Variable Transmission (IVT). The cam-loaded piston includes an annular flange having a flat surface on one face and a recess on the other face. The recessed portion may be used to connect a compression spring of the IVT. A first sealing ring groove is formed in the inner periphery of the annular flange of the cam load piston. The periphery of the annular flange of the cam load piston is also provided with a second sealing ring groove.

In addition, another aspect of the invention relates to a valve piston for an Infinitely Variable Transmission (IVT) with an unloading cylinder. The valving piston is an orifice ring. In one embodiment, the valve piston has a first ring on the orifice ring surface. The first ring can be used for connecting the unloading oil cylinder. The valve actuating piston has a second ring located on the opposite side of the first ring. A first sealing groove is formed in the periphery of a hole ring of the valve-driven piston. And a second sealing groove is formed on the inner periphery of the hole ring of the valve-driven piston.

One aspect of the invention relates to a center cam base for an infinitely variable transmission. The central cam base includes an annular cylindrical body. In one embodiment, the central cam base includes a set of splines located on the periphery of the annular cylindrical body. The central cam base has a first set of lines of elongation located on a first face of the annular cylindrical body. The central cam base also has a second set of lines of elongation on the second face of the annular cylindrical body.

Another aspect of the invention relates to a cam ring for an infinitely variable transmission. The cam ring includes a circular flange with a central bore. The cam ring has a plurality of splines arranged about the inner periphery of the central bore. In one embodiment, the cam ring has a cam shoulder located at the periphery of the circular flange. The cam shoulder has a neck extending therefrom. The neck has a spring retainer groove on its inner periphery. The cam shoulder is also provided with a group of cam extension lines.

In addition, another aspect of the invention is directed to an output disc of an infinitely variable transmission. The output disc has an annular cylindrical body. In one embodiment, the output disc has a first set of splines arranged about the inner periphery of the first end of the annular cylindrical body. The output disc has a second set of splines arranged about the periphery of the second end of the annular cylindrical body. The output disc also has an elongated flange extending from the second end of the annular cylindrical body.

In another aspect, the invention relates to a transmission housing for an Infinitely Variable Transmission (IVT). The transmission housing includes a cylindrical container with a first end and a second end. The transmission housing has a sleeve extending from the cylindrical container. The sleeve is used for connecting an oil pan of the IVT. In one embodiment, the transmission housing has a plurality of accelerator interfaces arranged on the sleeve. The transmission housing has several instrument ports arranged on a cylindrical container. The transmission housing has a first set of pin holes arranged at a first end of the cylindrical container. The transmission housing further having a second set of pin holes disposed at the second end of the container; in addition, the transmission housing has several lubricant interfaces arranged on the outer peripheral surface of the second end of the cylindrical container.

Another aspect of the invention relates to a housing for an infinitely variable transmission with a hydraulic system. The housing includes a cylindrical body with a central passage. In one embodiment, the housing has a cam-loaded piston interface located at the periphery of the cylindrical body. The housing has a lubricant port located at the periphery of the cylindrical body and a line pressure port located at the periphery of the cylindrical body. The housing includes a pilot pressure interface located at the periphery of the cylindrical body. The cam load piston interface, the lubricant interface, the line pressure interface, and the pilot pressure interface are all for fluid communication with the hydraulic system. The housing also includes a recess formed in the central passage. The recess may be used to connect to a cover plate of the transmission.

One aspect of the invention relates to a powertrain system having a power source, and an infinitely variable transmission coupled to the power source. The infinitely variable transmission includes a first set of traction rollers and a second set of traction rollers. An infinitely variable transmission has a carrier operatively connected to the first and second sets of traction rollers. The carrier is rotatable about a longitudinal axis of the infinitely variable transmission. In one embodiment, the powertrain includes a hydraulic system for controlling a gear ratio of an infinitely variable transmission.

Another aspect of the invention is directed to a powertrain for a tractor having a transmission housing and an infinitely variable transmission operatively connected to the transmission housing. The infinitely variable transmission includes a carrier that rotates about an infinitely variable transmission longitudinal axis. An infinitely variable transmission has a first set of planet-pivot arm assemblies. The carrier is operatively connected to at least one of the planet-pivot arm assemblies. The driveline further comprises a secondary gearbox connected to the infinitely variable transmission.

Another aspect of the invention relates to a transmission with an input shaft aligned along a longitudinal axis of the transmission. The transmission includes a carrier operatively connected to the input shaft. In one embodiment, the transmission includes a row of pivot arm assemblies operatively connected to the bracket. The variator has a set of traction rollers coupled to the pivot arm assembly. The variator also has a set of planet axles connected to the traction rollers. The traction roller is used for rotating coaxially with the planet shaft.

In addition, another aspect of the invention relates to a transmission with an input shaft and a carrier coupled to the input shaft. The variator has a set of first set of planet-pivot arm assemblies operatively connected to the carrier. In one embodiment, the variator has a first fixed traction ring operatively connected to the first set of planet-pivot arm assemblies. The variator has an output traction ring operatively connected to the first set of planet-pivot arm assemblies, and the variator also has a torque conversion device operatively connected to the output traction ring. In a certain embodiment, the transmission has an axial force generating device operatively connected to the first stationary traction ring. The variator also has a second set of planet-pivot arm assemblies coupled to the carrier. The first and second sets of pivot arm assemblies each include a planet and a planet axle operatively connected to the planet. The planet-pivot arm assembly also includes a pivot arm connected to the planet axle. The pivot arm may be operatively connected to a shifting mechanism of the transmission.

One invention of the invention relates to a planetary-pivot arm assembly of an infinitely variable transmission with a variator. The planet-pivot arm assembly includes a planet and a planet axle operatively connected to the planet. The planet-pivot arm assembly also includes a pivot arm connected to the planet axle. The pivot arm is operatively connected to the shifting mechanism.

Another aspect of the invention relates to a planet-pivot arm assembly for a variator of an infinitely variable transmission. The planet-pivot arm assembly has a spherical planet with a central bore. In one embodiment, the planet-pivot arm assembly has a planet axle with a first end and a second end. The planet-pivot arm assembly also has a set of elastomeric spheres mounted in the planet axle for forming a frictional interface between the planet and the central bore of the planet axle.

In addition, one aspect of the invention is directed to a planet-pivot arm assembly for an infinitely variable transmission. The planet-pivot arm assembly includes a spherical planet and a planet axle operatively connected to the planet. The planet-pivot arm assembly has a pivot arm connected to the planet axle. The pivot arm includes a first arm bar with a first bearing hole and a second arm bar with a second bearing hole. The pivot arm further includes a pivot hole connected to the first and second arm levers. The pivot aperture may be located at the ends of the first and second bearing apertures. The planet-pivot arm assembly includes several lubricant passages formed in the first and second arm levers.

In another aspect, the invention relates to a center cam assembly for applying axial load to an infinitely variable transmission component having a traction ring and one or more rows of planets. The center cam assembly includes a first cam ring operatively connected to a first traction ring. The center cam assembly includes a second cam ring operatively connected to a second traction ring. The first and second cam rings are adapted to generate an axial force to urge the first and second traction rings toward one or more of the rows of planets. The center cam assembly includes a plurality of torque transfer rings interposed between the first and second traction rings and the first and second cam rings, respectively. The center cam assembly also includes a center cam base with a set of extension lines. The central cam base may be operatively connected to the first and second cam rings. A central cam base is interposed between the first and second cam rings. The center cam assembly also includes a plurality of cam rollers to cooperate with the first and second cam rings to generate an axial force.

Another aspect of the invention relates to a center cam assembly, a center coupling and a traction ring with a flange. A central coupling is operatively connected to the flange. The center cam assembly includes a drive output element coaxial with the traction ring. The transmission output member may be operatively connected to the central coupling. The center cam assembly has a center output transition member connected to the transmission output member. The center cam assembly also has a plurality of axial force generating elements located between the traction ring and the center output transition element.

One aspect of the invention relates to an input cam assembly for an infinitely variable transmission with a traction ring. The input cam assembly includes a cam flange for coupling to the traction ring. In one embodiment, the input cam assembly includes a cam base positioned coaxially with the cam flange. The input cam assembly has a set of cam rollers supported by roller retainers. A cam roller may be used to interact with the cam base.

Another aspect of the invention is directed to a carrier for an Infinitely Variable Transmission (IVT). The bracket includes a first bracket center block. In a certain embodiment, the bracket has a second bracket center block connected to the first bracket center block. A hydraulic fluid chamber is located at the interface between the first and second carrier center blocks. The bracket includes a bracket input cover connected to the first bracket center block. The carriage also includes a carriage output cover connected to the second carriage center block.

Another aspect of the invention relates to a hydraulic shift control system for a transmission having a plurality of planets operatively connected to a planet axle and a pivot arm. The hydraulic shift control system includes a piston operatively connected to at least one of the shifter pivot arms. The hydraulic shift control system has a regulator for hydraulically driving the piston and thereby the pivot arm. In one embodiment, the hydraulic shift control system includes a control signal device operatively connected to the regulator. The hydraulic shift control system has a synchronizer device operatively connected to the pivot arm. The hydraulic shift control system also has a feedback system connecting the synchronizer device and the regulator.

Further, another aspect of the invention relates to a transmission mechanism of an infinitely variable transmission with an input shaft and a carrier. The gear shifting mechanism includes a hydraulic valve at least partially integral with the input shaft. The variator includes a hydraulic circuit for controlling fluid flow to and from the carrier through a plurality of passages and chambers. At least a portion of the conduit channel and the housing cavity are located within the input shaft.

Another aspect of the invention is directed to a position feedback mechanism for a continuously variable transmission or an infinitely variable transmission. The positioning feedback mechanism includes a hydraulic control valve for cooperating with a shifting device of the transmission. The position feedback mechanism includes a control screw operatively connected to the control valve. In a certain embodiment, the positioning feedback mechanism includes a feedback screw connected to the control screw. The feedback screw is for operative connection to a transmission of the transmission.

Another aspect of the invention relates to a synchronizer arrangement for a transmission of a continuously variable transmission or an infinitely variable transmission. The synchronizer device includes a control screw and a set of pivot pin hubs connected to the control screw. The control screws are used to synchronize multiple planetary gear sets of the transmission to the same tilt angle.

In addition, another aspect of the invention relates to a variator mechanism for a continuously variable transmission or infinitely variable transmission (C/IVT). The variator includes a control valve located within a cavity of the C/IVT input shaft. The variator has a control piston in fluid communication with a control valve. The shift mechanism also includes a pivot pin hub operatively connected to the control piston. The pivot pin hub may be operatively connected to the pivot arm of the C/IVT.

One aspect of the invention relates to a method of shifting an infinitely variable transmission having a set of pivot arms. Operatively, the method may couple the feedback mechanism to the set of pivot arms, and may also couple the adjuster to the feedback mechanism. The method includes the step of transmitting an indication of one or more conditions of the set of pivot arms from the feedback mechanism to the regulator. The method includes receiving a control signal on a regulator and regulating a hydraulic pressure with the regulator. The hydraulic pressure combines, at least to some extent, the control signal and one or more indications of the state of the set of pivot arms. The method further includes the step of facilitating a transmission ratio adjustment by using hydraulic pressure to move the plurality of pivot arms.

Another aspect of the invention relates to a transmission including a first array of planet-pivot arm assemblies. The planet-pivot arm assembly has a first row of traction rollers. The variator includes a second array of planet-pivot arm assemblies. The planet-pivot arm assembly has a second row of traction rollers. The transmission includes a carrier for receiving and supporting a plurality of rows of planet-pivot assemblies. The first and second rows of traction rollers are both angularly aligned about the longitudinal axis of the carriage and are both positioned concentrically with the longitudinal axis of the carriage. The variator includes a first fixed traction ring coupled to the first array of planet-pivot arm assemblies. The variator also includes a second fixed traction ring coupled to the second array of planet-pivot arm assemblies. In one embodiment, the variator has a first output traction ring coupled to the first array of planet-pivot arm assemblies. The variator also has a second output traction ring coupled to the second array of planet-pivot arm assemblies. The transmission includes an axial force generating device operatively connectable to the first and/or second output traction rings. In one embodiment, the variator has a variator coupled to the planet-pivot arm assembly. The variator synchronously actuates the first and second arrays of planet-pivot arm assemblies to tilt the axis of rotation of the traction drum. The transmission also has an input shaft connected to the carrier.

Yet another aspect of the invention is directed to an Infinitely Variable Transmission (IVT) with a carrier for rotation about a longitudinal axis of the IVT. The IVT has a first array of planet-pivot arm assemblies operatively connected to the carrier. The carrier is adapted to receive and support the first array of planet-pivot arm assemblies. Each planet-pivot arm assembly has a set of planet wheels angularly arranged about the longitudinal axis. The IVT has an input shaft connected to the carrier. The input shaft and carrier are adapted to be coaxial with and rotatable about a central axis of the transmission. The IVT includes a first fixed traction ring coupled to the first array of planet-pivot arm assemblies, and the IVT further includes an output traction ring coupled to the first array of planet-pivot arm assemblies. The IVT has an idler assembly coaxial with the longitudinal axis of the transmission. The idler wheel assembly is connected with the planet wheel. The IVT also has a hydraulic control system for fluid communication with the carrier and/or the input shaft. A hydraulic control system may be used to adjust the transmission ratio of the IVT.

In another aspect, the invention relates to a method of operating a transmission of an infinitely variable transmission. The method includes the steps of operatively connecting the input shaft to a carrier of the transmission and receiving power on the input shaft. The method includes the step of transmitting an input torque T1 to the carrier through the input shaft at an input speed W1. The method further includes the step of transmitting a second torque T2 from a drive flange of the transmission to an exterior of the transmission at a second speed W2. Both the second torque T2 and the second speed W2 may be continuously variable. The second torque T2 and the second speed W2 depend at least in part on the angle of inclination of the traction planets of the set of variator planet-pivot arm assemblies. The magnitude of the second speed W2 may be zero. The second speed W2 may rotate in forward and reverse directions.

Another aspect of the invention relates to a method of providing hydraulic axial loading to dynamically react to torque damage in an infinitely variable transmission. The method includes the step of providing a mechanical load cam assembly to generate an axial force to torque damage. The method further includes the step of connecting the hydraulic axial load device to the mechanical load cam assembly. The hydraulic axial load device may provide an axial force based, at least in part, on a steady state of operating transmission torque.

One aspect of the invention relates to a method of changing the transmission ratio of a variator of a continuously variable transmission or infinitely variable transmission. The method includes the steps of providing a hydraulic control valve and operatively connecting the hydraulic control valve to a set of hydraulic loops. The method includes the step of operatively connecting a set of transmission planet shafts to at least one set of hydraulic pistons. In a certain embodiment, the method includes the step of regulating hydraulic pressure with a hydraulic control valve. The method includes the step of providing hydraulic pressure from a hydraulic control valve to at least one set of hydraulic pistons. The method further comprises the step of hydraulically actuating the change in the inclination angle of the planet axles.

Another aspect of the invention is directed to a shifting method of a continuously variable transmission or an infinitely variable transmission. The method comprises the steps of hydraulically connecting the control valve to the control piston and connecting the control piston to the feedback spring. In a certain embodiment, the method comprises the step of providing a pilot pressure to the control valve indicative of the transmission planet axle tilt angle. The pilot pressure range is at least a function of the spring rate of the feedback spring, the total deflection range of the feedback spring, and the control piston area. The method further comprises the step of actuating the planet axles to tilt based on, at least in part, the pilot pressure. Tilting of the planet shafts can shift the transmission.

Drawings

Fig. 1 is a schematic view of a drive system using a continuously variable transmission or an infinitely variable transmission.

Figure 2 is a perspective view of a transmission that may be connected to a housing and a tractor range box that provides a tractor transmission.

FIG. 3A is an isometric view of the transmission connected at one end to an input shaft and at the other end to a gear set and an output shaft.

FIG. 3B is an elevational view of the transmission illustrated in FIG. 3A connected to a housing and a gear unit housing.

Fig. 4A is a schematic diagram of an embodiment of a transmission according to an embodiment of the invention described in the present application.

FIG. 4B is a schematic diagram of a hydraulic shifting mechanism that may be used with the transmission shown in FIG. 4A.

Fig. 5A is a cross-sectional view of certain components of the transmission shown in fig. 3B, taken along line a-a.

Fig. 5B is detail a of the cross-section shown in fig. 5A.

Fig. 6A is detail B of the cross-section shown in fig. 5A.

Fig. 6B is a partially exploded isometric view of some of the transmission components shown in fig. 6A.

FIG. 6C is a cross-sectional view of an alternative center device and center cam assembly that may be used with the transmission shown in FIG. 5A.

Fig. 7A is a partial enlarged view C of the cross-section shown in fig. 5A.

Fig. 7B is a partially exploded isometric view of some of the transmission components shown in fig. 7A.

Fig. 8 is an enlarged partial view D of the cross-section shown in fig. 5A.

FIG. 8A is a perspective view of a bracket assembly that may be used with the transmission shown in FIG. 5A.

FIG. 8B is a right side elevational view of the bracket assembly shown in FIG. 8A.

FIG. 8C is a cross-sectional view of the bracket assembly shown in FIG. 8A taken along line B-B.

FIG. 8D is a cross-sectional view of the bracket assembly shown in FIG. 8A taken along line C-C.

FIG. 8E is a perspective view of the bracket assembly first end member shown in FIG. 8A.

Fig. 8F is a second perspective view of the first end member shown in fig. 8E.

Fig. 8G is a front view of the first end member shown in fig. 8E.

FIG. 8H is a cross-sectional view of the first end member of FIG. 8G taken along line A1-A1.

FIG. 8I is a cross-sectional view of the first end member shown in FIG. 8G taken along line BI-BI.

FIG. 8J is a cross-sectional view of the first end member of FIG. 8G taken along line C1-C1.

Fig. 8K is a perspective view of the bracket intermediate member shown in fig. 8A.

Fig. 8L is a second perspective view of the intermediate member shown in fig. 8K.

Fig. 8M is a right side elevational view of the intermediate member shown in fig. 8L.

FIG. 8N is a cross-sectional view of the intermediate member shown in FIG. 8M taken along line A2-A2.

FIG. 8O is a cross-sectional view of the intermediate member shown in FIG. 8M taken along line B2-B2.

FIG. 8P is a cross-sectional view of the intermediate member shown in FIG. 8M taken along line C2-C2.

Fig. 8Q is a perspective view of the second end member of the bracket shown in fig. 8.

Fig. 8R is a second perspective view of the second end member shown in fig. 8Q.

Fig. 8S is a right side elevational view of the second end member illustrated in fig. 8R.

FIG. 8T is a cross-sectional view of the second end member of FIG. 8S taken along line A3-A3.

FIG. 8U is a cross-sectional view of the second end member of FIG. 8S taken along line B3-B3.

FIG. 9 is a perspective view of a planet-pivot arm subassembly that may be used with the embodiment shown in FIG. 5A.

Fig. 10 is a front view of the planet-pivot arm assembly shown in fig. 9.

FIG. 11 is an elevation view of the planet-pivot arm subassembly shown in FIG. 9.

FIG. 12 is a cross-sectional view of the planet-pivot arm subassembly shown in FIG. 11 taken along line D-D.

Fig. 13 is a partial enlarged cross-sectional view E of fig. 12.

FIG. 14A is a perspective view of a pivot arm that may be used with the planet-pivot arm subassembly shown in FIG. 9.

Fig. 14B is a right side elevational view of the pivot arm shown in fig. 14A.

Fig. 14C is a cross-sectional view of the pivot arm of fig. 14B taken along line E-E.

FIG. 15A is a perspective view of a planet axle that may be used with the planet-pivot arm sub-assembly shown in FIG. 9.

Fig. 15B is a front view of the planet axle shown in fig. 15A.

Fig. 15C is a cross-sectional view of the planet axle shown in fig. 15B taken along line F-F.

FIG. 15D is a perspective view of a guide wheel that may be used with the planet-pivot arm sub-assembly shown in FIG. 9.

Fig. 15E is a second perspective view of the guide wheel shown in fig. 15D.

Fig. 15F is a left side elevational view of the guide wheel shown in fig. 15E.

Fig. 15G is a cross-sectional view of the guide wheel shown in fig. 15F taken along line G-G.

Fig. 16A is a perspective view of a subassembly that can be part of the shifting mechanism of the shifting device shown in fig. 5A.

Fig. 16B is a right side elevational view of the subassembly shown in fig. 16A.

Fig. 16C is a cross-sectional view of the subassembly shown in fig. 16B, taken along line H-H.

Fig. 16D is a cross-sectional view of the subassembly shown in fig. 16B, taken along line I-I.

Fig. 16E is a graph showing the speed of certain subassemblies of the embodiment of the transmission 310.

FIG. 17A is a perspective view of an input shaft that may be used with the shifting mechanism shown in FIG. 16.

FIG. 17B is another perspective view of the input shaft shown in FIG. 17A.

Fig. 17C is an elevation view of the input shaft shown in fig. 17A.

FIG. 17D is a cross-sectional view of the input shaft shown in FIG. 17C taken along line J-J.

FIG. 18A is a perspective view of a manifold that may be used with the shifting mechanism shown in FIG. 16.

FIG. 18B is a second perspective view of the manifold shown in FIG. 18A.

Fig. 18C is a left side elevational view of the manifold shown in fig. 18A.

FIG. 18D is a cross-sectional view of the manifold shown in FIG. 18C taken along line K-K.

FIG. 18E is a cross-sectional view of the manifold shown in FIG. 18C taken along line L-L.

FIG. 18F is a perspective view of a cover plate that may be used with the transmission shown in FIG. 5A.

Fig. 18G is a rear elevational view of the cover plate shown in fig. 18F.

Fig. 18H is a front elevational view of the cover plate shown in fig. 18F.

Fig. 18I is a cross-sectional view of the cover plate shown in fig. 18H taken along line M-M.

Fig. 19 is an exploded view of certain components of the shifting mechanism shown in fig. 16.

FIG. 20 is detail E of the cross section shown in FIG. 16 showing a hydraulic valve system that may be used with the variator shown in FIG. 16.

FIG. 21A is a perspective view of a pivot pin hub that can be used with the shifting mechanism shown in FIG. 16.

Fig. 21B is a front view of the pivot pin hub shown in fig. 21A.

Fig. 21C is a right side elevational view of the pivot pin hub illustrated in fig. 21A.

Fig. 21D is a cross-sectional view of the pivot pin hub of fig. 21C taken along line M-M.

Fig. 22A is a perspective view of a control piston that may be used with the shifting mechanism shown in fig. 16.

Fig. 22B is an elevation view of the control piston shown in fig. 22A.

Fig. 22C is a cross-sectional view of the control piston taken along line N-N of fig. 22B.

FIG. 23A is a perspective view of a gear set and axle assembly that can be connected to the transmission shown in FIG. 5A.

FIG. 23B is a second perspective view of the gear set and shaft assembly shown in FIG. 23A.

FIG. 23C is an exploded view of the gear set and shaft assembly shown in FIG. 23A.

FIG. 23D is a schematic illustration of a gearset or rangebox as may be used with the gearset or rangebox shown in FIG. 5A.

Fig. 23E is a perspective view of an assembly that measures the variator planet axle tilt angle γ shown in fig. 5A.

Fig. 23F is a front view of the assembly shown in fig. 23E.

FIG. 23G is a cross-sectional view of the assembly of FIG. 23F, taken along line A5-A5.

Fig. 24A is a perspective view of a traction ring that may be used with the transmission shown in fig. 5A.

Fig. 24B is a cross-sectional view of the pull ring of fig. 24A.

FIG. 25A is a perspective view of a drive flange that can be used with the transmission shown in FIG. 5A.

FIG. 25B is a second perspective view of the drive flange of FIG. 25A.

FIG. 25C is a left side elevational view of the drive flange illustrated in FIG. 25B.

FIG. 25D is a cross-sectional view of the drive flange of FIG. 25C taken along line P-P.

FIG. 26A is a perspective view of a reaction flange that may be used with the transmission shown in FIG. 5A.

FIG. 26B is a cross-sectional view of the reaction flange shown in FIG. 26A.

FIG. 27A is a perspective view of a torque converter assembly that may be used with the transmission shown in FIG. 5A.

FIG. 27B is a cross-sectional view of the torque conversion assembly shown in FIG. 27A.

FIG. 28A is a perspective view of an alternative reaction flange that may be used with the transmission shown in FIG. 5A.

FIG. 28B is a cross-sectional view of the reaction flange of FIG. 28A.

FIG. 29A is a perspective view of a cam flange that can be used with the transmission shown in FIG. 5A.

Fig. 29B is a cross-sectional view of the cam flange shown in fig. 29B.

FIG. 30A is a perspective view of a cam base that can be used with the transmission shown in FIG. 5A.

Fig. 30B is a front view of the cam base of fig. 30A.

Fig. 30C is a right side elevational view of the cam base shown in fig. 30A.

Fig. 30D is a rear elevational view of the cam base shown in fig. 30A.

FIG. 31A is a perspective view of a cam load piston that may be used with the transmission shown in FIG. 5A.

Fig. 31B is a cross-sectional view of the cam-loaded piston of fig. 31B.

FIG. 32A is a perspective view of a valve operating piston that can be used with the transmission shown in FIG. 5A.

Fig. 32B is a right side elevational view of the piston shown in fig. 32A.

Fig. 32C is a cross-sectional view of the piston shown in fig. 32A.

FIG. 33A is a perspective view of an unloader cylinder that can be used with the shifter shown in FIG. 5A.

Fig. 33B is a cross-sectional view of the unload cylinder of fig. 33A.

FIG. 34A is a perspective view of a center cam base that can be used with the transmission shown in FIG. 5A.

Fig. 34B is a side view of the cam base of fig. 34A.

FIG. 35A is a perspective view of a cam ring that can be used with the shifting apparatus shown in FIG. 5A.

FIG. 35B is a second perspective view of the cam ring shown in FIG. 35A.

FIG. 35C is a cross-sectional view of the cam ring shown in FIG. 35A.

FIG. 36A is a perspective view of an output disc that can be used with the shifter shown in FIG. 5A.

Fig. 36B is a cross-sectional view of the output disk shown in fig. 36B.

FIG. 37A is a perspective view of a carrier guide ring that can be used with the transmission shown in FIG. 5A.

Fig. 37B is a cross-sectional view of the guide ring shown in fig. 37A.

FIG. 38A is a perspective view of a synchronizing ring that can be used with the transmission shown in FIG. 5A.

Fig. 38B is a cross-sectional view of the synchronizer ring shown in fig. 38A.

FIG. 39A is an exploded perspective view of an idler assembly that may be used with the transmission shown in FIG. 5A.

Figure 39B is a cross-sectional view of the idler assembly shown in figure 39A.

FIG. 40A is a perspective view of a transmission housing that can be used with the transmission shown in FIG. 5A.

Fig. 40B is a second perspective view of the transmission housing of fig. 40A.

Fig. 40C is a front view of the transmission housing of fig. 40A.

Fig. 40D is a right side elevational view of the transmission housing illustrated in fig. 40A.

FIG. 40E is a cross-sectional view of the transmission housing taken along line Q-Q of FIG. 40D.

FIG. 41A is a perspective view of a housing that can be attached to the transmission shown in FIG. 5A.

Fig. 41B is another perspective view of the housing shown in fig. 41A.

Fig. 41C is another perspective view of the housing shown in fig. 41A.

Fig. 41D is a front view of the housing shown in fig. 41A.

FIG. 41E is a cross-sectional view of the housing of FIG. 41D taken along line A8-A8.

Detailed Description

Certain inventive embodiments will be described with reference to the accompanying drawings, wherein like numerals refer to like elements throughout. The terminology used in the description of the present application is not intended to be interpreted in a narrow or limiting sense as it is used in conjunction with the detailed description of certain specific embodiments of the invention. Moreover, embodiments of the invention may include several novel features, no single one of which is solely responsible for its desirable attributes and which is essential to practicing the inventions herein described. The CVT/IVT embodiments described herein are generally associated with the transmissions and shifters disclosed in U.S. patent nos. 6,241,636, 6,419,608, 6,689,012, and 7,011,600. The disclosures of all of these patents are hereby incorporated by reference in their entirety.

As used herein, the terms "operatively associated," "operatively connected," "operatively linked," "operatively associable," "operatively connectable," "operatively linkable," and the like refer to a relationship (mechanical, interlocking, connecting, etc.) between elements that results in the operation or actuation of one element relative to the other element, either subsequently or simultaneously. It is noted that in using the above terminology to describe embodiments of the invention, emphasis is placed on the particular structure or mechanism of the linkage or linking elements. However, unless otherwise expressly stated, when the above terms are used, they indicate that the actual linkage or coupling may take a variety of forms that, in some cases, will be readily apparent to those having ordinary skill in the relevant art.

In the description, the term "radial" refers to a direction or position perpendicular to the longitudinal axis of the transmission or variator. The term "axial" as used herein refers to a direction or position along an axis parallel to the main or longitudinal axis of the transmission or variator. For simplicity and clarity, sometimes like components (e.g., control piston 582A and control piston 582B) with like labels will be referred to collectively as (e.g., control piston 582).

As shown in FIG. 1, one embodiment of the drive system 100 employs a Continuously Variable Transmission (CVT) or an Infinitely Variable Transmission (IVT) CVT/IVT 105. Drive system 100 may include a power source 110 connected to a first gear set 120 by a coupling 160. A coupling 170 couples a Continuously Variable (CV) or Infinitely Variable (IV) CV/IV variator 130 to the first gear set 120, and a coupling 180 couples the CV/IV variator 130 to the second gear set 140. The driven device 150 is connected to the second gear set 140 by a coupling 190. It should be appreciated that the description of CVT/IVT 105 as including variator 130, gearsets 120, 140 and couplings 160, 170, 180, 190 is primarily for convenience. The transmission 130 may provide all functions of the transmission alone, depending on the application and the background, and thus the transmission 130 may be regarded as a continuously variable transmission or an infinitely variable transmission.

For example, power source 110 may be an electric motor, an internal combustion engine, or a hybrid prime mover combining an electric motor and an internal combustion engine. The first and second gear sets 120, 140 may be any gear box arrangement, each of which may include one or more planetary gear sets. For example, the driven device 150 may be a coupling, a shaft, a propeller, a differential drive split gearbox, a traction load (e.g., to drive a motorcycle, a car, a truck, or a tractor), an industrial load (e.g., to drive a stationary or semi-stationary device such as a printer), a propulsion load (e.g., to drive a ship such as a ship or boat or to drive an aircraft such as an airplane or a helicopter), a utility load (e.g., to drive a dunpster crane, a trash compactor, or a turboprop), an agricultural load (e.g., to drive a spray attachment on a tractor or combine), and combinations thereof, such as traction and agricultural loads or traction and utility loads, and the like. The driven device 150 may also be a compressor, generator, pump, and accessory drive, including an alternator, water pump, cooling fan, etc.

Couplings 160, 170, 180, and 190 may be any suitable mechanical device that transfers power between connected devices. For example, couplings 160, 170, 180, and 190 may be any type of coupling, from splines, keys, welds, or flanges to individual planetary gear sets and to gearboxes having multiple planetary gear sets and other gears in parallel or in series. The CV/IV transmission 130 may be any embodiment of a continuously variable transmission or infinitely variable transmission as described below.

In some embodiments, the drive system 100 may have one or both of the first and second gear sets 120, 140 absent. Thus, the drive system 100 may be designed such that the power source 110 is connected to the CV/IV variator 130 through a coupling 160, and there is no first gear set 120 connection between the power source 110 and the CV/IV variator 130. In other embodiments, the CV/IV shifting device 130 may be connected to the driven device 150 without a second gear set 140 connection between the CV/IV shifting device 130 and the driven device 150. Also, in some embodiments, additional gear sets may be connected in series or parallel to the first gear set 120, the CV/IV variator 130, or the second gear set 140.

One embodiment of the drive system 100 may be applied to a transmission assembly 300 in a tractor, as shown in FIG. 2. In the illustrated embodiment, the housing 330 is attached to the CV/IV variator 310, which is in turn coupled to a rangebox or gearset 320 having a rangebox housing 325. As discussed below, in some embodiments, both the housing 330 and the gear set 320 are given the features necessary to cooperate with components of the CV/IV variator 310. FIG. 3A shows an embodiment of the CV/IV variator 310 including an input shaft 510 connected to a gear set 320, which in this embodiment is connected to an output shaft 585. Fig. 3B shows that the CV/IV variator 310 may be configured with a housing 531 and a casing 590 as a gear set 320.

Fig. 4A is a schematic diagram of an embodiment of a transmission 200 that may be used in embodiments of the drive system 100. The transmission 200 may be a continuously variable transmission or an infinitely variable transmission. The transmission 200 includes a transmission housing 205 that houses a carrier 215, which in this embodiment is designed to rotate about a longitudinal axis and is operatively connected to a planet-pivot arm assembly 220. Generally, in some embodiments, the planet-pivot arm assembly 220 includes a planet 222 mounted on a planet axle 254, which is operatively connected to a pivot arm 252; the planets 222 can be traction rollers as described below. The pivot arms 252 may rotate or tilt the planet axles 254 to produce shifting of the transmission 200. The planet-pivot arm assemblies 220 are generally angularly spaced evenly about the central longitudinal axis of the transmission 200. In the illustrated embodiment, the input shaft 210 is connected to a carrier 215. As used herein, a "pulling roll" may be a spherical roll or sphere in some embodiments. Thus, as used herein, the terms "traction drum", "spherical drum" or "ball" are interchangeable when referring to rolling elements that transmit power by traction.

The traction rings 225 and 227 are coupled to the planetary gear sets 222A and 222B, respectively, of the planetary pivot arm assembly 220. An idler assembly (not shown in fig. 2A, see the example of fig. 5A) connects and radially supports the planetary gear sets 222A and 222B. In some embodiments, the idler assembly is coaxial with the longitudinal axis of the transmission 200 (see also fig. 5A). Traction rings 230 and 233 connect planetary gear sets 222A and 222B, respectively. The traction rings 230, 233 are operatively connected to a torque output member 232. In some embodiments, the traction rings 225, 277 may be connected to ground through the grounding device 270. In this application, "ground" refers to a stationary element such as transmission housing 205. The grounding device 270 may be a key, spline, clip, or other reinforcement means that substantially prevents axial and/or rotational movement of the traction rings 225, 227. It should be noted that in some embodiments, only one or none of the traction rings 225, 227 are grounded. The transmission 200 may also include several axial force generating ("AFG") devices 235. In some embodiments, the variator 200 includes a variator 240 operatively coupled to the planet-pivot arm assembly 220.

In the embodiment shown in fig. 4A, during operation of the transmission 200, the input shaft 210 transfers torque to the carrier 215, which then transfers torque from the carrier 215 to the planet-pivot arm assembly 220. In certain embodiments, the traction rings 225 and 227 are non-rotatable, providing a rolling surface and torque reaction for the planetary gear sets 222A and 222B, respectively. However, in other embodiments, one or both of the traction rings 225 and 227 are rotatable.

In the embodiment shown in fig. 4A, the planetary gear sets 222A, 222B convert torque to rotatable traction rings 230 and 233, which may be operatively connected to a torque output member 232, thereby transferring torque thereto. In one embodiment, the carrier 215 may be connected to the output shaft 256 or be part of the output shaft 256. Thus, the transmission 200 can receive one power input through the input shaft 210 and form two power outputs: power output through the torque output member 232 and power output through the carrier 215 and the output shaft 256. Of course, the transmission 200 may transmit power along the opposite power path. That is, in some embodiments, transmission 200 may receive power via shaft 256 and/or torque element 232 and may then output power via shaft 210. Since the transmission 200 can provide at least two power paths, the transmission 200 can be a torque splitting device.

In embodiments where the carrier 215 is used to rotate about the shaft of the transmission 200 and the traction rings 225, 227 are not rotatable, the torque output member 232 may be used to achieve zero speed and/or reverse its direction of rotation. Also, the output of the transmission 200 may be zero or negative when the output of the torque output member 232 is combined with the output of the output shaft 256. Since some embodiments of the transmission 200 can produce zero output speed when the input speed is not zero, and the torque ratio is generally inversely proportional to the speed ratio, the transmission 200 can be described as a continuously variable transmission.

In the embodiment shown in FIG. 4A, variator 240 controls the proportion of torque applied to the carrier input that reacts to the torque of planet-pivot arm assembly 220. In one embodiment of the transmission 200, the variator 240 includes a hydraulic control system that varies the angle of rotation of the planetary gear sets 222A and 222B by actuating (e.g., rotating or tilting) elements such as a pivot arm 252 and a planetary shaft 254, as described below. It will be apparent to one of ordinary skill in the relevant art that other devices besides the shifting mechanism 240 may be used to effect shifting of the transmission 200.

Fig. 4B illustrates one embodiment of a hydraulic shift control system 280 (or a shift mechanism 280) that may be used with the transmission 200. The shifter 280 includes a regulator 282 that regulates a hydraulic fluid shift pressure 284 for actuating at least one piston, namely piston 294. A pump 288 may be used to deliver the control fluid to provide the variable speed pressure 284 to the regulator 282. Both the control signal device 290 and the feedback mechanism 292 may be connected to the regulator 282. As previously described, the pivot arm 296 may be connected to the synchronizer 293, thereby connecting the synchronizer 293 to the feedback mechanism 292. However, in other embodiments, a synchronizer is not used and instead the pivot arm 296 may be directly connected to the feedback mechanism 292. In one embodiment, the piston 294 is operatively connected to a pivot arm 296. The regulator 298 is used for fluid communication with a fluid tray or fluid tank 298.

For example, regulator 282 may be a valve. In one embodiment, regulator 282 is a four-way valve with a spool valve that regulates the control fluid pressure and/or flow to piston 294 and fluid tank 298. The control signal device 290 may be a mechanical, electrical, or electromechanical device adapted to send a control signal to the regulator 282. In certain embodiments, the control signal is hydraulic fluid pressure (also referred to as "pilot pressure"). In other embodiments, the control signal device 290 may be used to receive and process electrical or mechanical signals from the feedback mechanism 292 and/or the synchronizer 293.

The control piston 294 may be configured and adapted such that the variable pressure 284 actuates movement of the control piston 294. Pivot arm 296 may be pivot arm 252 shown in fig. 4A. In certain embodiments, the feedback mechanism 292 is mechanical, electrical, and/or electromechanical. In certain embodiments, the feedback mechanism 292 includes a feedback screw and a feedback spring (see fig. 16C and 20). The synchronizer 293 may be a mechanical linkage that synchronizes the sets of planet gears 222A, 222B so that the planet shafts 254 of each planet gear set 222A, 222B have the same inclination angle. In one embodiment, synchronizer 293 includes a device with a turnbuckle function that may be coupled to the planetary gear set 222 via one or more additional linkages. The pump 288 and tank 298 may be typical, well known fluid pumping and collection devices.

During operation, to shift ratios, the control signal device 290 actuates the regulator 282, which allows the shift pressure 284 to actuate the piston 294. In some embodiments, the regulator 282 may regulate the speed at which the variable speed pressure 284 is transmitted. Because the control piston 294 is coupled to the pivot arm 296, the pivot arm 296 can be actuated and responded to by operation of the control piston 294. In some embodiments, the pivot arms 296 are operatively connected to the planet axles 254, the pivot arms 296 thereby causing the angle of the planet axles 254 relative to the transmission longitudinal axis to change or tilt relative to the longitudinal axis of the transmission 200. The change in the angle of the planet axle 254 causes the ratio of the transmission 200 to change.

As previously mentioned, in some embodiments, the pivot arm 296 may be connected to the feedback mechanism 292. In this embodiment, the feedback mechanism 292 can communicate one or more status indicators of the pivot arm 296 to the regulator 282, which indicators can include angular position, axial position, angular velocity, axial or linear velocity, and the like. In some embodiments, the indicators may include hydraulic fluid flow and/or pressure in the piston, electrical signals from speed ratio measurements, position of the carrier 215, angular position of the planets 222, and axial forces on the traction rings 225, 227, 230, and/or 233 caused by centrifugal or gyroscopic forces generated by rotation of the carrier 215. Because the pivot arm 296 is moved by the movement of the control piston 294, the feedback mechanism 292 will transmit any of the indicators described above to the regulator 282. Upon combining the control signal from the control signal device 290 with the indicia communicated by the feedback mechanism 292, the regulator 282 will further adjust the shift pressure 284 to initiate the desired gear ratio adjustment or to maintain the gear ratio in steady state.

In certain embodiments of the transmission 200, the AFG device 235 applies an axial force to the traction rings 225, 227, 230, and 233 to facilitate efficient transfer of torque between the planetary gear sets 222A, 222B and the traction rings 230 and 233. In other embodiments, AFG device 235 can only be connected to a portion, but not all, of traction rings 225, 227, 230, and 233. AFG device 235 may be cam-type, wherein the interaction between the cam surface and the rollers may generate an axial force (which may be proportional to the torque applied to the cam surface); it may also be of the hydraulic actuator type, wherein hydraulic fluid may urge the combination of the piston and cylinder to generate an axial force. In other embodiments, AFG device 235 may incorporate a cam and hydraulic axial force generation method. It should be noted that the hydraulic or cam-type AFG device 235 is not the only option for generating the appropriate axial force in the transmission 200. Preferably, AFG device 235 is used to generate axial forces that can respond quickly to transient torque impairments and at least to some extent to rely on or respond to the highest torque levels that occur in any of traction rings 225, 227, 230, and 233.

As shown in fig. 5A-8, in one embodiment, the transmission 310 shown in fig. 3A has a transmission housing 505 that houses a bracket 515 or the like. The input shaft 510 is connected to a carrier 515. In some embodiments, the axial position of the input shaft 510 and the carrier 515 relative to the housing 510 is not strictly limited. That is, in some embodiments, the input shaft 510 and the carrier 515 can (at least partially) float axially within the transmission housing 505. In this embodiment, both the input shaft 510 and the carrier 515 are rotatable about their central longitudinal axes. In one embodiment, the input shaft 510 and the carrier 515 are rotatably coupled.

In one embodiment, the shifter 310 includes a planet-pivot arm assembly 579 (see detail a and 5B). The planet-pivot arm assembly 579 generally includes a planet axle 554 that provides a motive support and an axis of rotation for the planet 522. In this embodiment, the pivot arms 552 support the planet axles 554, while the idler 562 provides radial support and position for the planet 522. Additional descriptions of the planet-pivot arm assembly 579 are provided below (see also FIGS. 9-15G).

In some embodiments, the cover plate 560 is mounted coaxially with the input shaft 510 and is connected to the manifold 565. The cover plate 560 may provide bearing support for the input shaft 510. In the illustrated embodiment, the housing 531 is used to receive, support and secure the cover plate 560 and the manifold 565.

Traction rings 525A, 530 connect planets 522A, while traction rings 525B, 533 connect planets 522B. In the illustrated embodiment, the traction rings 525A, 525B are designed to be nearly or completely non-rotatable. The drive flange 532 is connected to the traction rings 530, 533 by a center cam assembly 570 (see detail B and fig. 6A-6B). In one embodiment, the drive flange 532 is coupled to the sun gear 2320 of the downstream gear set 320 (see fig. 23A-23C). In the illustrated embodiment, the drive flange 532 may, but need not, be axially fixed.

In the illustrated embodiment, the input cam assemblies 575 are coupled to the traction ring 525A (see detail C and fig. 7A-7B). The center cam assembly 570 and the input cam assembly 575 function to some extent as axial force generators. The center cam assembly 570 and the input cam assembly 575 will be described in further detail below with reference to FIGS. 6A-6B and 7A-7B, respectively.

During operation of transmission 310, in certain embodiments, input shaft 510 applies torque to carrier 515, and carrier 515 transfers torque to traction rings 530 and 533 via planet-pivot arm assembly 579. The traction rings 530, 533 then transfer torque to the center cam assembly 570 and the drive flange 532, causing them to rotate as a rigid body. Thus, certain embodiments of the transmission 310 will force power splitting. That is, transmission 310 may be used to receive power input at input shaft 510 and transmit power through two different paths. For example, assuming that the input shaft 510 transmits power at an input speed w1 and an input torque T1, the transmission 310 can transmit power at a continuously variable speed w2 and an output torque T2 through the drive flange 532, while the transmission 310 can provide power at an output speed w1 and an output torque T3 through the spline shaft 844 connected to the carrier 515. In the illustrated embodiment, spline shaft 844 is part of cradle 515; however, in other embodiments, spline shaft 844 may be connected to the carrier 515 by any suitable means, including keyed, splined, bolted, dowelled, geared, hydraulic or electric motors, and the like.

In some embodiments, either the drive flange 532 or the carrier spline shaft 844 can be used as a drive or power take off. However, in other embodiments, the two torque outputs T2 and T3 may be integrated into the torque output T4 on the output shaft 585 through an auxiliary gear set. As described below with respect to fig. 42, the variator 310 can be coupled to a rangebox 4200 to provide a transmission with multiple modes, including reverse drive. Depending on the angle of inclination of the planet axle 554, the drive flange 532 may reverse its direction of rotation and have substantially zero rotational speed, i.e., w2 is about equal to zero. In some embodiments, the transmission 310 can produce zero output speed when the output of the drive flange 532 is summed into the output of the splined shaft 844. Therefore, the transmission device 310 may function as a continuously variable transmission or an infinitely variable transmission.

As shown in the embodiments illustrated in fig. 5A, 8 and 16A-20, the transmission 310 may have a shifting mechanism 577 in which the input shaft 510 is adapted to receive and interact with various components of a hydraulic valve. The shifting mechanism 577 is partially illustrated in detail D of fig. 5A and 8; additional description of the embodiment of the shifting mechanism 577 is provided in FIGS. 16A-22C and the accompanying text. Embodiments of the input shaft 510 are described further below with reference to fig. 17A-17D. In one embodiment, the variator 577 includes a hydraulic circuit that can control fluid flow to and from the carrier 515 via the passages 814A, 814B and the chambers 580A, 580B (see fig. 8D and 16C-16D).

In certain embodiments, the control piston 582 is coupled to the pivot pin hub 805. Pivot pin boss 810 supported by pivot pin hub 805 supports pivot pin 815 which is connected to pivot arm 552. In a certain embodiment, the control screw 820 comprises a right control screw 820A rigidly connected to a left control screw 820B, wherein the right control screw 820A has a thread lead opposite in direction to the thread lead of the left control screw 820B. When the control screw 820 is turned in a certain direction, the threads of the left control screw 820B and the right control screw 820A, which are opposite to each other, provide a turnbuckle function. In some embodiments, right control screw 820A and left control screw 820B are integral. In one embodiment, the left control nut 825A is rigidly connected to the pivot pin hub 805A and the left control nut 825B is rigidly connected to the pivot pin hub 805B. Thus, when the control screw 820 is axially restrained and turned, the right and left control nuts 825A and 825B move axially in opposite directions to one another (see fig. 19), which causes the pivot arm connected to the planet 522A to rotate in an opposite direction (in the center of the planet 522A) to the pivot arm connected to the planet 522B that rotates in unison.

Further, in some embodiments, the absolute values of the thread leads between the right control screw 820A and the right control nut 825A and the left control screw 820B and the left control nut 825B are equal, which causes the axial movement of the right control nut 825A and the left control nut 825B to be of equal magnitude and opposite direction. In one embodiment, equal but opposite axial movement is translated into equal but opposite rotational movement of the pivot arms 552A, 522B by the pin-and-slider mechanism. Since the pivot arms 552 pivot due to the operative connection of the pivot arms 552 to the planet axles 554, the tilt angle of the axis of rotation of the planets 522 is adjusted, thereby allowing equivalent adjustment of the ratio of the individual variator cavities of the variator 310. In some embodiments, the ratio of the individual transmission chambers can be set to different values by using a differential mechanism by choosing different leads for the control screws 820A, 820B and/or the control nuts 825A, 825B and supplementing other measures. The control screw 820, control nut 825, pivot pin hub 805, and connecting screw end stop 870, shown in fig. 19, ensure that the angle of inclination of the axis of rotation of the planet 522A reflects the planet 522A in contrast to the plane of the center of the variator 310, which bisects the distance between the centers of the planets 522A and 522B.

In certain embodiments, control screw 820A cooperates with control nut 825A to collectively provide mechanical feedback to the hydraulic valve. In the embodiment shown in fig. 5A, the control nut 825B cooperates with the control screw 820B to provide mechanical feedback to the sensor to help determine the angular position of the planet axle 554, which is the angular position of the axis of rotation of the planet 522 (see fig. 19 and 23D-23F). The control nuts 825A, 825B are rigidly connected to the pivot pin hubs 805. The control screw end stop 870 axially secures the control screw 820A. The operation of the hydraulic shift mechanism 577 will be further discussed below with reference to FIGS. 16A-16D and 20.

It should be noted that in the transmission 310 of the present embodiment, the axial thrust bearings are not used to transfer axial loads due to the transfer of tractive forces between the components, at least in part due to the configuration of the center cam assembly 570. More specifically, the traction rings 525A, 525B are rotationally fixed and they can transmit axial forces to the transmission housing 505. Because no axial thrust bearing is used for the traction rings 525A, 525B, the bearing friction losses that are often generated when axial thrust bearings are used to transmit axial loads are avoided.

As shown in fig. 6A and 6B, in certain embodiments, the center cam assembly 570 includes a center cam base 605 splined to the drive flange 532. The central cam base 605 is operatively connected to a right cam ring 610 and a left cam ring 615. In one embodiment, rollers (not shown) supported by the cam roller retainer 650 provide an operative connection between the center cam base 605 and the cam rings 610, 615. The right cam ring 610 is connected to the right torque transfer ring 620, and the left cam ring 615 is connected to the left torque transfer ring 625. The right torque transfer ring 620 is connected to the traction ring 533, while the left torque transfer ring 625 is connected to the traction ring 530. As used herein with reference to cam rings 610, 615 and torque converter rings 620, 625, the terms "left" and "right" refer only to a position relative to the central cam base 605 and have no other meaning.

Bearings 630, 635 support torque converter rings 620, 625, respectively, concentric with the carrier 515. In some embodiments, the bearings 630, 635 are radial bearings, but in other embodiments, the bearings 630, 635 may be ball bearings. Carrier guide ring 640 and carrier center bearing shim 642 are located between bearings 630 and 635. The synchronization ring 645 is mounted on a concentric axis between the cam rings 610, 615 and the torque transfer rings 620, 625. The synchronization ring 645 is coupled to the cam rings 610, 615. The synchronizing ring 645 allows for axial deflection but does not allow the cam rings 610, 615 to rotate relative to each other, thereby centering the center cam base 605 between the two planetary gear sets 522A, 522B. As shown in fig. 6B, the cam roller retainer 650 is positioned between the central cam base 605 and the cam rings 610, 615. The cam roller retainers 650 can support and separate rollers (not shown) that connect the cam rings 610, 615 to the center cam base 605 when torque is applied to the center cam assembly 570. The shape of the rollers may be cylindrical, barrel or spherical.

Because the cam ring 610 is coupled to the cam ring 615 via the synchronizing ring 645, the vertical heights of the respective rollers of the cam rings 610, 615 are substantially equal. This ensures that symmetrical axial displacement of the cam rings 610, 615 relative to the planetary gear sets 522A, 522B is loaded. It is preferable that the distances between the centers of the planetary gear sets 522A, 522B and the center of the transmission carrier 515 during operation be equal. In some embodiments, a skewed axial movement of the carrier 515 in response to an axial force may be preferred. The inner rings of bearings 630, 635 are rigidly mounted on carrier 515. The outer rings of bearings 630, 635 may then be mounted on torque converter rings 620, 625 by, for example, a slip fit. In this embodiment, the outer rings of the bearings 630, 635 may move axially along the center cam assembly 570. To keep the carrier 515 centered between the planetary gear sets 522A and 522B, a wave spring (not shown) is placed between the outer race side 655 of the bearings 630, 035 and the torque converter rings 620, 625. To generate axial forces, axial deflection is best, but it is recommended that the carrier 515 be positioned at a center point between the planetary gear sets 522A and 522B at all times. In some embodiments, the wave spring acts only axially on the outer ring of the bearings 630, 635. However, in other embodiments, the outer rings of bearings 630, 635 are press fit with torque transfer rings 620, 625 and the wave springs act only on the inner rings of bearings 630, 635.

Referring now to FIG. 6C, a center assembly 1000 that can be used with the transmission 310 is illustrated. The hub assembly 1000 does not rely on springs to perform the centering function. In one embodiment, the center assembly 1000 includes a bearing 1002 supported by the carrier 515. A centering coupling 1004 is supported on the bearing 1002 and is connected to a flange of a traction ring 1006. The axial force generating element 1008 is interposed between the traction ring 1006 and the central output transition element 1010, connected to the drive output element 1012. In certain embodiments, the axial force generating element 1008 includes load cams 1014, 1016 connected in operation by a cam roller cylinder 1018. In one embodiment, an axial thrust bearing 1020 is interposed between the centering coupling and the central output transition element 1010. For some applications, it may be recommended to use shims 1022 to ensure minimal clearance and/or accurate positioning of the components. In certain embodiments, splines are used to connect the centering coupling 1004 to the traction ring 1006 and the central output transition element 1010 to the drive output element 1012, respectively. It should be noted that the traction ring 1006, while rotationally fixed to the centering coupling 1004, is not limited to the centering coupling 1004 in the axial direction.

Referring to fig. 7A and 7B, in one embodiment, the input cam assembly 575 includes a cam flange 705 coupled to the traction ring 525 and configured to interact with a cam roller (not shown) supported by a roller retainer 710. The cam rollers may be spherical, barrel or cylindrical in shape. As shown, the cam base 715 can be positioned between the roller retainer 710 and the cam load piston 720. The cam flange 705 is operatively connected to a valve piston 725 that cooperates with the unloader cylinder 730. In this embodiment, the input cam assembly 575 may further include a plurality of compression springs 735 positioned between the sides of the cam load piston 720 and partially within the bore 755 of the housing 531. For purposes of description herein, cam flange 705, roller retainer 710, cam base 715 and associated rollers are referred to as a mechanical load cam assembly 717.

The input cam assembly 575 is an embodiment of an axial force device that combines hydraulic and cam-type axial force generation. Using hydraulic pressure, the cam load piston 720 may apply an axial force to the traction ring 525 via the mechanical load cam assembly 717. In other embodiments, the mechanical load cam assembly 717 can be modified to include a single or multiple components that transfer axial forces from the cam load piston 720 to the traction ring 525 without the use of a cam. However, some embodiments do not use the cam load piston 720, and only the mechanical load cam assembly 717 provides the axial force on the input side of the variator 310. The mechanical load cam assembly 717 is characterized as a passive axial force generator that continuously reacts axial forces in proportion to torque.

In embodiments that utilize hydraulic axial loading, it is recommended to provide a dynamic response to torque damage. This can be achieved by combining a cam-type axial load device (for reacting quickly to torque damage) with a hydraulic axial load device. In embodiments employing active axial force generation, the hydraulic pressure may adjust the axial force generated by the cam to a desired magnitude. For example, the axial force level provided by the mechanical load cam assembly 717 can exceed the maximum specified operating axial force, while the valving piston 725 and the unloader cylinder 730 provide hydraulic control to adjust the axial force generated by the mechanical load input cam assembly 717 to a desired axial force level. Thus, in certain embodiments, the cam load piston 720 is not used, but rather the mechanical load cam assembly 717 is used to control axial force generation such that the cam-type axial force generated is out of range of the transmission. In the embodiment shown in fig. 5A, 7A, and 7B, the valving piston 725 can reduce the force of the planet flange 705 on the traction ring 525, thereby relieving the load on the planet 522 to a selected level. In such embodiments, the axial force is most of the time constant, allowing torque damage to be addressed by cam-type AFG while controlling the steady state axial force required for a particular torque with hydraulic-type AFG. If a hydraulic valving piston is used, it is possible to use only the input load cam assembly 717; in such embodiments, the cam rings 610, 615 and the center cam 605 may be replaced with solid members.

For example, as torque increases, hydraulic pressure decreases to be received by the input cam assembly 717. If a particular application requires 100 pounds of axial force at steady state, it would be desirable to provide a cam that can generate 1000 pounds of axial force. The pressure to the valve actuating piston 725 and the unloader cylinder 730 will reduce the axial force on the traction ring 525 to 100 pounds. This configuration can handle torque damage, minimize friction during normal operation, and help achieve desired axial force requirements.

It is worth mentioning that while fig. 5A describes an embodiment that combines hydraulic and cam-type axial force generating devices on the input side of the variator 310, not all or any of the cam load piston 720, the input cam assembly 717, and the valve actuating piston 725 and the unloader cylinder 730 may be used, depending on the circumstances of the embodiment. In embodiments where no AFG device is used on the input side of the transmission 310, the traction ring 525 is rotationally fixed to a non-rotatable member of the transmission 310, and a thrust member may be interposed between the traction ring 525 and the transmission housing 505.

It should be noted that embodiments of the center cam assembly 570 and the input cam assembly 575 can produce an axial force proportional to the torque on the respective cam assemblies 570, 575. Thus, the one of the cam assemblies 570, 575 that experiences the greatest torque will determine the level of axial force in the transmission 310 because the cam assembly 570, 575 that experiences the greatest torque is able to generate the greatest axial force.

An embodiment of the bracket 515 is shown in fig. 8A-8U. In this embodiment, the carrier 515 includes a carrier input cap 802 connected to a first carrier center block 804. The cradle 515 also includes a cradle output cover 806 connected to a second cradle center block 808. First carrier center block 804 and second carrier center block 808 are coupled together to form a hydraulic fluid chamber 580. As shown in fig. 8C and 8D, the carrier input cap 802 and carrier center blocks 804, 808 incorporate lubricant passages 812 and hydraulic fluid passages 814A, 814B. In this embodiment, the carrier output cover 806 includes only the lubricant channel 812.

As shown in fig. 8E-8H and 9, in this embodiment, the carrier input cap 802 has several fastening pins 816 that help to connect the carrier input cap 802 to the carrier center block 804. The fastening pin 816 has a bolt hole 821 through which a bolt or screw (not shown) can be screwed. In this embodiment, the securing pin 816 is provided as part of the carrier input cap 802, and the lubricant channel 812 and hydraulic fluid channels 814A, 814B are partially formed in the securing pin 816.

The carrier input cap 802 also includes carrier pins 822 whose surfaces 824 can transfer force to bearings 920 (see fig. 9) of a planet-pivot arm assembly 579, which in turn transfers force to shafts 554 that support planets 522. The surface 826 of the bracket pin 822 provides support for the guide wheels 925 as the planet pivot arm assembly 579 shifts. The carrier pin 822 also includes a threaded aperture 828 that can secure the input shaft 510 to the carrier input cap 802.

In the vicinity of the carrier pin 822 and the securing pin 816, the carrier input cap 802 includes a lubricant interface 885 that fills a lubricant turret 887 in this embodiment. In the illustrated embodiment, lubrication fluid is drawn from the input shaft 510 and injected into the channel 812 of the carrier input cap 802 before being delivered to the lubrication port 885.

The first and second carrier center blocks 804, 808 are very similar; the following description of the first carriage center block 804 therefore applies generally to the description of the second carriage center block 808 as well. Fig. 8K-8P show an embodiment of a first carrier center block 804. The securing pin 830 of the center block 804 is similar to the securing pin 816 of the bracket input cover 802. The bolt holes 832 can facilitate the attachment of the bracket input cap 802 to the center block 804. The body of the center block 804 (including the pin 830) incorporates the lubricant passage 812 and the hydraulic fluid passages 814A, 814B.

The center block 804 includes carrier pins 834, the carrier pins 834 being similar in form and function to the carrier pins 822 of the carrier input cover 802. Surface 836 of bracket pin 834 transmits force to bearing 920, while surface 838 provides support for guide wheel 925. As shown in fig. 8L and 8N, bracket pins 834 and neck-like extension rod 843 incorporate bolt holes 840 to facilitate the connection of the fastening center blocks 804 and 808. In certain embodiments, center block 804 also includes pin holes 842 for positioning and assembly. The center block 804 includes an inner cylindrical portion 581 located in the neck-shaped extension rod 843 which, along with a similar inner cylindrical portion of the center block 808, forms a cavity 580.

The carrier output cover 806 is seen in fig. 8Q-8U. The carrier output cover 806 is similar to the carrier input cover 802. In this embodiment, the carrier output cap 806 does not incorporate the hydraulic control fluid passage 814 as does the carrier input cap 802. In this embodiment, the carrier output cover 806 includes a carrier output shaft 844. In this embodiment, splines 846 on the carrier output shaft 844 allow for torque conversion. The carrier output cover 806 includes fastening pins 848 and carrier pins 850 that are identical to the fastening pins 816 and carrier pins 822 of the carrier input cover 802. The support surface 852 and the torque transfer surface 854 of the carrier output cover 806 are similar to the support surface 826 and the torque transfer surface 824 of the carrier input cover 802.

In certain embodiments, the carrier input cap 802, the carrier output cap 806, and/or the first and second carrier blocks 804, 808 may deliver lubricant to the idler 562 (see fig. 5B) or the idler 3905 (see fig. 39A-39C) and the idler bearing 3920 through the lubricant interface.

As shown in fig. 9-15F, a planet-pivot arm assembly 579 includes a planet 522 mounted on a planet axle 554 that is supported by a pivot arm 552. In one embodiment, the planets 522 are spheres having a diameter of about 2.5 inches. The central bore of the planet 522 is about 0.5 inches in diameter to support the planet axle 554. The planet 522 may be manufactured from bearing quality steel, such as 52100 steel. The elastomer balls 907 fit into the grooves 1505 of the planet axle 554 (see fig. 15). In one embodiment, the elastomeric spheres 907 provide a frictional interface between the planet axle 554 and the planet 522 whereby the planet axle 554 and the planet 522 rotate as a unit.

In the illustrated embodiment, the clips 1305 fix the axial position of the planet 522 on the planet axle 554. The ends of the planet axles 554 are connected to branches of the pivot arms 552. In this embodiment, each leg of the pivot arm 552 may support a shell-type needle bearing 910 and an angular contact ball bearing 915. The needle bearing 910 may be press fit into a bore 1410 (see fig. 14A) of the pivot arm 552, which press fit fixes the axial position of the needle bearing 910. The angular contact ball bearing 915 telescopes into the bore 1420 of the pivot arm 552, and the clip 1310 secures the ball bearing 915 within the bore 1420. In some embodiments, when the bearing 915 is used, the guide wheels 925 may not be used. The ends of the planet axles 554 are supported on bearings 910, 915. Ball bearings 920 are attached to each end of the planet axle 554. In the illustrated embodiment, the ball bearings 920 are held in place by the clips 1318 and the shoulders 1540 of the planet axle 554. The guide wheels 925 are mounted on axles 930 that are inserted into bores 1315 (see fig. 13) of the pivot arms 552. The guide wheels 925 react the speed-changing force generated by the piston 582 so that the speed-changing force is not transmitted to the bearings 910, 915, 920. The guide wheels 925 center the pivot arm 552 at all times in the planet 522.

An embodiment of a pivot arm 552 is shown in fig. 14A-14C. The pivot arm 552 includes pivot arm levers 1425 and 1430 positioned about the pivot hole 1435 opposite each other. In the illustrated embodiment, the pivot arm bars 1425, 1430 form a unitary body, improving structural rigidity; however, in other embodiments, the pivot arm bars 1425, 1430 may be separate components operatively connected to a pivot point such as the pivot hole 1435 for flexibility of assembly. As shown in fig. 8, the pivot hole 1435 is connected to a pivot pin 815, which in some embodiments is supported by the pivot pin block 810. In other embodiments, the pivot pin 815 may be integral with the pivot arm extensions 1425, 1430.

In this embodiment, the pivot arm levers 1425, 1430 each include a bore 1410 and a bore 1420 to receive the bearings 910, 915 (see description of FIG. 13 above). In other embodiments, the holes 1410 and 1420 may be replaced with one hole. Also, the bearings 910, 915 may be replaced with a bearing or bearing surface, which may be integral with the pivot arm extensions 1425, 1430. In the illustrated embodiment, the pivot arm bars 1425, 1430 each include a recess 1440 and a hole 1315. The groove 1440 interfaces with the guide wheel 925 (see FIG. 10), while the bore 1315 interfaces with the shaft 930 on which the guide wheel 925 is mounted. In the illustrated embodiment, the pivot arms 552 can transmit primarily radial loads on the planet axles 554 on the bearings 910, in contrast to other methods of tilting the planet axles 554, in which the actuating device transmits bending loads on the planet axles 554. As shown, in some embodiments, the pivot arm 552 includes passages 1445 and 1450 to facilitate the flow of lubricant.

In one embodiment of the pivot arm 552, the distance between the holes 1315 is about 4 inches, and the distance between the center of the pivot hole 1435 and the center of the hole 1410 is about 2.5 to 3.0 inches. In some embodiments, the pivot arm extensions 1425, 1430 extend from pivot arm apertures 1435 having a radius of about 2.5 to 3.0 inches, and more preferably about 2.75 inches. For some applications, the diameter of the holes 1410 may be about 0.5 to 0.8 inches, while the diameter of the holes 1420 may be about 0.8 to 1.2 inches. In one embodiment, pivot hole 1435 may have a diameter of approximately 0.2 to 0.3 inches, groove 1440 may have a diameter of approximately 0.6 to 0.8 inches, and hole 1315 may have a diameter of approximately 0.2 to 0.3 inches. For some applications, the passages 1445, 1450 can have diameters on the order of 0.1 to 0.2 inches. In one embodiment, the pivot arm 552 is fabricated from 4140 heat treated steel. However, due to the centrifugal forces generated by the embodiment in the event that the carrier 515 rotates, it is recommended that the pivot arm 552 be made of a material that is strong but lighter in weight than steel.

An embodiment of the planet axle 554 is shown in fig. 15A-15C. The planet axle 554 includes a cylindrical central portion 1510 with grooves 1505 that can support elastomer spheres 907 (see fig. 13). In some embodiments, at least one groove 1505 is eccentric with respect to the centerline of the planet axle 554, thereby allowing the elastomer spheres 907 to be entangled into a reduced radial spacing of the subassembly as long as there is relative rotation between the planet 522 and the planet axle 554. The central portion 1510 also includes grooves 1507 that support snap rings (not shown) that fix the axial position of the planets 522 on the planet axle 554. Each end of the planet axle 554 has a neck 1515 that extends radially toward the longitudinal axis of the planet axle 554 and tapers and then expands slightly radially as it transitions into the cylindrical portion 1520. The cylindrical portion 1520 may support a bearing 910 (see fig. 13), which in some embodiments is press fit onto the cylindrical portion 1520. The planet axle 554 further includes another cylindrical portion 1525 that is used to support the bearing 915 (see fig. 13). The cylindrical portion 1525 includes a groove 1527 that supports a clip 1310 that holds the bearing 915 in place. The shoulder 1540 of the planet axle 554 also aids in the axial fixation of the bearing 915. The cylindrical portion 1530 of the planet axle 554 serves to receive and support the bearing 920 (see FIG. 13). The groove 1532 on the cylindrical portion 1530 is used to support the clip 1318 that secures the bearing 920.

In one embodiment, the planet axle 544 has an overall length of about 5.5 inches. In some applications, the central portion 1510 has a diameter of about 0.5 inches and is slip fit or press fit into the central bore of the planet 522. In a certain embodiment, the central portion 1510 is about 2.5 inches long. In certain embodiments, the cylindrical portion 1520 is about 0.5 inches long and about 0.45 inches in diameter, the cylindrical portion 1525 is about 0.4 inches long and about 0.40 inches in diameter, and the cylindrical portion 1530 is about 0.3 inches long and about 0.27 inches in diameter.

Figures 15D-15G illustrate one embodiment of a guide wheel 925. Because the guide wheel 925 is typically subject to dead loads or low speed operation, the guide wheel 925 need not be a bearing. However, in some embodiments, the guide wheels 925 may be radial ball bearings. The central bore 1502 of the guide wheel 925 is used to support a pin 930 (see fig. 10) that passes through the pivot arm 522. In some embodiments, the guide wheel may roll on the pin 930. The guide wheel 925 has a chamfered neck 1504 that fits into a recess 1440 (see fig. 14A) in the pivot arm 552. The counter bore 1506 of the guide wheel 925 extends to one end of the pin 930 and supports a spring collar (not shown) that holds the pin 930 in place. In some embodiments, the bore 1502 is designed such that the guide wheel 925 secures the pin 930 by an interference fit; in such embodiments, the circlip may not be used. However, in other embodiments, the pin 930 may roll in the hole 1315 of the pivot arm 552.

In one embodiment, the diameter of the central bore 1502 is about 0.2 to 0.3 inches and the diameter of the counter bore 1506 is about 0.4 to 0.5 inches. The outer diameter of the guide wheel 925 may be about 0.6 to 0.8 inches and the diameter of the neck 1504 may be about 0.4 to 0.6 inches. In one embodiment, the width of the guide wheels 925 is about 2.5 to 3.5 inches. For some applications, the guide wheels 925 may be made of AISI or SAE 52100 steel, tailored to a minimum of about 58 HRC.

One embodiment and operation of the hydraulic shift mechanism 577 (hereinafter simply referred to as "shift mechanism 577") are described next; however, some definitions are listed first. As shown in fig. 16D, the gamma angle 2395 is defined as the angle between the axis 2380 of the planet axle 554 and the longitudinal axis 2390. The axle 2380 is parallel to the longitudinal axis 2385 of the transmission 310 through the centers of the planet wheels 522. Thus, when planet axis 554 is parallel to axis 2380, γ angle 2395 equals zero. It should be noted that the gamma angle 2395 may range from a maximum positive value to a maximum negative value, such as from +35 degrees to-35 degrees. In a preferred embodiment, the gamma angle 2395 ranges from about-30 degrees to about +30 degrees. Various embodiments of the variator 310 can have different ranges of the gamma angle 2395. Also, in certain embodiments, the range of the gamma angle 2395 need not be symmetrical with zero degrees. That is, as in some embodiments, the maximum gamma angle 2395 may be positive +41 degrees and the maximum gamma angle negative 2395 may be negative-20 degrees. It should be noted that in some embodiments, it is advisable to use the value of the γ angle 2395 of the planet 522B to reflect the value of the γ angle 2395 of the planet 522A.

In embodiments where the traction rings 525A, 525B are not rotatable, the speed at which the planets 522 rotate about the axis 2390 is dependent on the gamma angle 2395. In embodiments where the carrier 515 rotates about an axis 2385, the planets 522 have orbital velocities about the axis 2385. In short, the speed at which the planet 522 rotates about the shaft 2390 will be referred to as the rotational speed of the planet 522, and the orbital speed of the planet 522 about the shaft 2385 will be referred to as the orbital speed of the planet 522. The surface speed at any point of the planet 522 depends on the rotational speed and orbital speed of the planet 522. The speed of the output traction 530, 533 is dependent on the surface speed of the planets 522 at the contact points between the planets 522 and the output traction rings 530, 533. The contact point between the planet 522 and the output traction rings 530, 533 will be referred to as the "contact point". The surface velocity at the above-mentioned contact point will be referred to as "surface velocity". The speed of the drive flange 532 is dependent on the speed of the output traction 530, 533.

In short, in one embodiment, to change the ratio of the transmission 310, the control valve 1605 is operatively connected to the planet axle 554. Adjustment of the state of the valve 1605 will result in adjustment of the gamma angle 2395, resulting in a change in the rotational speed of the planet 522. As the rotational speed of the planets 522 changes, the speed of the output traction 530, 533 changes, resulting in a change in the speed of the drive flange 532. Next, various embodiments of the apparatus and methods of causing the gamma angle 2395 to change or remain stable will be described.

Turning now to FIG. 16A, in one embodiment, a carrier 515 supports various components or functions associated with shifting of the transmission 310; thus, the carrier 515 is included as part of the shifter 577 in this application. However, it will be apparent to one of ordinary skill in the relevant art that the functions of the carrier 515 and shifter 577 can be separated and provided by separate components. As shown in fig. 16C and 20, the gearshift 577 includes a hydraulic control valve 1605 housed in a cavity 512 (see fig. 17D) of the input shaft 510. The valve 1605 generally shown in detail E of fig. 16C and fig. 20 will be discussed further below. To deliver control fluid to the valve 1605, in one embodiment, the shifter 577 includes a fluid manifold 565 that is coupled to a cover plate 560. An external pump (not shown) supplies control fluid to the manifold 565. In other embodiments, the pump may be located within the manifold 565, the cover plate 560, or other ranges of the variator 310. In this embodiment, the manifold 565 and the cover plate 560 are mounted coaxially with the input shaft 510. Further illustrations of certain embodiments of the input shaft 510, manifold 565, and cover plate 560 are detailed in fig. 17A-17D, fig. 18A-18E, and fig. 18F-18I, respectively.

As shown in fig. 8D and 16C, the control valve 1605 is in fluid communication with the chamber 580 via passages 814A, 814B in the input shaft 510 and the carrier 515. Control pistons 582A, 582B divide chamber 580 into chambers 580A, 580B. As previously described, the bracket center blocks 804, 808 form a cavity 580 when brought together. As shown in fig. 16D, control pistons 582A, 582B are operatively connected to pivot pin hubs 805A, 805B, respectively. In a certain embodiment, the control pistons 582A, 582B are coupled to the pivot pin hubs 805A, 805B by an interference fit, key, weld, thread, or other fastening method. The pivot arm 552 is operatively connected to the pivot pin hubs 805, 805B. The pivot arms 552 are operatively connected to the planet axles 554. An example of a planet-and-planet-shaft subassembly 579 is detailed in the description of fig. 5B. More details regarding the pivot pin hubs 805 are shown in fig. 21A-21D, more details regarding the control piston 582 are shown in fig. 22A-22C, more details regarding the pivot arm 552 are shown in fig. 14A-14C, or more details regarding the planet axle 554 are shown in fig. 15A-15C.

In the embodiment shown in fig. 20, the control valve 1605 includes a spool bushing 1610 that houses other valve components and receives and distributes hydraulic control fluid. In certain embodiments, the spool bushing 1610 is axially constrained by a clip (not shown) located on a clip groove 2003 (see FIG. 17D) of the input shaft 510. A key (not shown) for connecting the spool bushing 1610 and the input shaft 510 fixes the spool bushing 1610 in a manner to rotate about the input shaft 510. Other methods of axially and rotationally constraining the spool bushing 1610 to the input shaft 510 will be readily apparent to those having ordinary skill in the relevant art.

A compression spring 1615, partially covered by the spool bushing 1610 has one end pressing against a recess of the input shaft 510 and the other end engaging a pilot control piston 1620. The pilot cylinder 1625 abuts against the pilot piston 1620. In this embodiment, the flange of the pilot cylinder 1625 engages the spool bushing 1610. A compression spring 1615 is suggested to ensure that the pilot control piston 1620 continuously compresses the spool 1630. The feedback spring 1635 is connected to the spool valve 1630 at one end and to the feedback screw 1640 at the other end.

In some embodiments, the control valve 1605 is used to balance the pilot pressure acting on the piston area of the pilot control piston 1620 compressing the feedback spring 1635. The control sensitivity or resolution is defined as the operative pilot pressure range divided by the gamma angle 2395 operating range. The pilot pressure range is the difference between the highest and lowest pilot pressures. In some applications, it is suggested that the widest pilot pressure range possible be used for a particular range of gamma angle 2395. For example, in one case, the resolution may be 20psi pilot pressure change per 1 degree change in the gamma angle 2395. The pressure range can be adjusted by selecting the area of the pilot control piston 1620 and the performance of the feedback spring 1635. Thus, the pilot pressure range may be calculated by the pressure range k d/a, where k refers to the spring rate, d refers to the total deflection range of the feedback spring 1635, and a refers to the area of the pilot control piston 1620. In general, a stiffer (i.e., higher k value) feedback spring 1635 will result in a wider pilot pressure range for a particular A. Likewise, if the feedback spring 1635 has a higher d than a certain range of the γ angle 2395, the pilot pressure range will be larger. Finally, if a decreases, the entire pilot pressure range will increase.

Additionally, once the pilot pressure range is determined, the center point of the range can be adjusted by setting an initial preload on the feedback spring 1635. If a higher pilot pressure is desired, a greater preload deflection of the feedback spring may be imparted. If a lower pilot pressure is desired, a lower initial deflection of the feedback spring 1635 may be imparted. For example, if the feedback spring 1635 is given an initial deflection of 0.020 inches and the pilot pressure range is 50 to 250psi over the range of gamma angles 2395-20 to 20 degrees, the pressure range may expand with the initial deflection. Thus, if the feedback spring is given an initial deflection of 0.04 inches, the pilot pressure can range from about 100psi to 300 psi. In some embodiments, the center point of the pilot pressure range may move up or down without affecting the range of pilot pressures as long as the feedback spring 1635 is never deflected outside the linear range of the feedback spring 1635.

In one embodiment, the pilot control piston 1620 is approximately 1.0 inch long, includes a central cavity having a diameter of approximately 0.2 to 0.3 inches, and has an outer diameter of approximately 0.3 to 0.4 inches. In some embodiments, the pilot cylinder 1625 has a through central bore with a diameter of about 0.3 to 0.4 inches to receive the pilot piston 1620. The pilot cylinder 1625 may have an outer diameter of up to about 0.5 to 0.6 inches and may fit within the cavity of the spool bushing 1610, as shown in fig. 20. However, in other embodiments, to increase the area of the pilot piston 1620, the pilot cylinder 1625 is not used, but rather the cavity and outer diameter of the pilot piston 1620 are increased to fit the cavity of the spool bushing 1610. In such embodiments, the flexibility of setting the control sensitivity can be increased by providing a way to effectively vary a (the area of the pilot piston 1620). In some applications, the pilot control piston 1620 and/or the pilot cylinder 1625 are fabricated from 52100 or 440C steel.

In a certain embodiment, the spool bushing 1610 has an overall length of about 4.5 to 5 inches. The middle cavity diameter of the spool bushing 1610 can be up to about 0.4 to 0.6 inches. The outside diameter of the spool sleeve varies from about 0.7 inches at one end to about 1.1 inches at the other end. One end of the spool bushing 1610 may be provided with a set of step 4 acme threads to engage with corresponding threads of the feedback screw 1640. The set of acme threads can have a nominal diameter of up to 0.75 inch, a lead of about 0.25 inch, and a pitch of about 0.125 inch. The spool bushing 1610 may be provided with a plurality of ports (see fig. 20) connecting the outside and the inside of the spool bushing 1610. For some applications, the port typically has a diameter of about 0.125 inches. In some embodiments, spool valve 1630 is about 2.0 inches long, the intermediate chamber diameter is about 0.25 inches, and the outer diameter is about 0.5 to 0.7 inches. The spool valve 1630 is preferably mated to the spool valve bushing 1610. In one embodiment, the spool valve 1630 is made of 52100 or 440C steel. The spool bushing 1610 may be manufactured from 440C steel.

As shown in fig. 19, to provide feedback to the control valve 1605, a control screw 820 is operatively connected to the control valve 1605. More specifically, in certain embodiments, the control screw 820 is connected to a feedback screw 1640. The control screw 820 may be keyed, splined, or connected in other suitable manner to the feedback screw 1640. In certain embodiments, the control screw 820 is rigidly connected by rotation, but is not limited to the feedback screw 1640 in the axial direction. In this embodiment, the control screw 820 includes a left-handed threaded screw 820A and a right-handed threaded screw 820B that are interconnected by a pin (not shown). Thus, the control screw 820 may be used to provide mechanical feedback to the control valve 1605. As previously described, in certain embodiments, the control screw 820 may provide additional feedback for external measurements of the gamma angle 2395 (see fig. 23D-23F and accompanying text).

As shown in fig. 19-20, the attachment screw end stops 870A, 870B are radially and axially constrained within the bracket 515 and, in conjunction with the shoulder of the control screw 820, provide a positive axial stop for the control screw 820. In some embodiments, the thrust bearing 2065 is located between the attachment screw end stop 870A and the input shaft 510. In other embodiments, a thrust washer 2060 may be used between the thrust bearing 2065 and the input shaft 510. The attachment screw end stop 870B is axially restrained by the bracket 515. In certain embodiments, a thrust bearing or thrust washer (not shown) may be located between the bracket 515 and the attachment screw end stop 870B. In the illustrated embodiment, the bearing 2070 is positioned radially beside grooves in the carrier input cap 802 and the carrier output cap 806. The attachment screw end stops 870A, 870B are, in turn, radially positioned next to or within the bearing 2070. The attachment screw end stop provides radial support for the control screw 820. As shown in fig. 20, bearing 2070 is mounted on attachment screw end stop 870A and radially supports bracket 515. In one embodiment, a central bushing 1615 covers a central portion of the control screw 820 and seals against the inner bushing, thereby forming a portion of the cavity 580.

In certain embodiments, the control screw 820 is coupled to the pivot pin hubs 805A, 805B via control nuts 825A, 825B. Thus, since the control screw 820 ensures that the gamma angle 2395 between the planet axle 554 and each of the planetary gear sets 522A and 522B is the same, the control screw 820 may function as a synchronizer or synchronizer.

In one embodiment, the control screw 820 has an overall length of about 19 inches. In one embodiment, screws 820A, 820B each have a lead of about 3 inches, with screw 820A having left-hand threads and screw 820B having right-hand threads. For some applications, the nominal diameter of the control screw 820 may be up to about 1.0 inch. In a certain embodiment, the central bushing 1615 has an overall length of about 6.5 to 7.0 inches. In certain embodiments, the outer diameter of the central bushing 1615 is about 1.5 inches and the inner diameter is about 1.0 inches. In one embodiment, the feedback screw 1640 has an overall length of about 0.8 to 0.9 inches, a nominal trapezoidal thread diameter of about 0.75 inches, and a hexagonal hole diameter of about 0.3 inches.

In addition, as shown in fig. 17A-17D and 20, to change the gamma angle 2395, pilot pressure is communicated through the interface 2005 and the passage 1725 (partially shown in fig. 17D) of the input shaft 510 to the pilot pressure chamber 2010 and applied thereto to the pilot control piston 1620. As the pilot pressure applied to the pilot control piston 1620 increases, the pilot control piston 1620 moves the spool 1630 axially toward the feedback spring 1635 and thereby biases the feedback spring 1635. Line pressure is provided by a port 2015 and enters a port 2020 of the spool bushing 1610 through a channel 1730 (partially shown in fig. 17D) of the input shaft 510. As the spool valve 1630 biases the feedback spring 1635 (and thus moves to the right in the orientation of fig. 20), control fluid may flow from the port 2020 to the port 2025. The control fluid then flows from the fluid port 2025 into the channel 814A (see fig. 16C) that feeds the chamber 580A. As chamber 580A fills, control pistons 582A, 582B slowly move away from the center of chamber 580.

When the port 2030 connected to channel 814B is in fluid communication with the vent port 1650, control fluid in chamber 580B is vented through the vent port 1650 of the spool valve 1630. Further description of controlling fluid discharge is provided below with respect to the description of the input shaft 510 with reference to fig. 17A-17D. When the pilot pressure in pilot pressure chamber 2010 is reduced, control spring 1635 forces spool valve 1630 to move to the left, thereby establishing fluid communication between fluid ports 2020 and 2030. Control fluid then flows into passage 814B and chamber 580B, causing control pistons 582A, 582B to move toward the center of chamber 580. The control fluid in chamber 580A is then exhausted through passage 814A and port 2025.

As previously described, increasing the pilot pressure causes the control fluid to fill the chamber 580A and the exhaust chamber 580B. When chamber 580A is full, the control fluid in chamber 580A may push control piston 582 outward from the center of chamber 580A. Because the pivot pin hubs 805 are connected to the control pistons 582, the pivot pin hubs 805 will move axially away from the center of the chamber 580. That is, the pivot pin hubs 805A move to the left, while the pivot pin hubs 805B move to the right (where "left" and "right" are viewed in plan view from fig. 16C-16D). Axial movement of the pivot pin hub 805 causes the pivot arm 552 to rotate about the pivot pin 815. It should be noted that actuation of the pivot arm 552 by the pivot pin hub 805 also causes the pivot arm 552 to rotate about the center of the planet 522.

Thus, filling or emptying of the chambers 580A, 580B pushes the control piston 582. Axial movement of the control piston 582 causes axial movement of the pivot pin hubs 805, which in turn causes rotational movement of the pivot arms 552. Because the pivot arms 552 are connected to the planet axles 554, as the pivot arms 552 rotate, the planet axles 554 tilt, thereby changing the gamma angle 2395.

As previously mentioned, in one embodiment, the pivot pin hubs 805 are rigidly connected to the control nut 825, which is threaded onto the control screw 820. Because the control screw 820 incorporates two oppositely threaded screws 820A, 820B connected by a pin (not shown), the control screw 820 acts as a turnbuckle. Due to the opposite threads of the control screw 820, the control screw 820 is pushed in a constant rotational direction when the control piston 582 is moved in the opposite direction. The control screw 820 rotates but does not move axially. For example, when the pivot pin hub 805A moves to the left, the control nut 825A causes the left-handed control screw 820A to rotate. Since the control screw 820A is rotatably coupled to the feedback screw 1640, the control screw 820A also causes the feedback screw 1640 to rotate. Rotation of the feedback screw 1640 in the threads of the spool bushing 1610 causes the feedback screw 1640 to move axially, reacting against the control spring 1635 and changing the bias of the control spring 1635. The spool valve 1630 is balanced by the force caused by the pressure on the load piston 1620 and the control spring 1635. When the force of the control spring 1635 is greater than the pilot pressure, the control spring 1635 pushes the spool 1630. The ratio of the variator 310 remains stable when the spool 1630 is near the interfaces 2030 and 2025 (as shown in fig. 20), thereby preventing fluid communication between the interface 2020 (line pressure) and the interface 2030 or 2025 to fill or evacuate the cavities 580A, 580B. The spool valve 1630 relies on the interaction between the valve loading piston 1620 and the control spring 1635 to achieve a steady ratio state through the force balance created by the spool valve 1630. (20) As the control piston 582 moves away from the center of the chamber 580, the bias of the control spring 1635 changes until an equilibrium between the force applied by the control spring 1635 to the spool valve 1630 and the force applied by the pilot piston 1620 to the spool valve 1630 is achieved. In this way, the feedback screw 1640 and the position control of the gamma angle 2393 are also achieved.

Certain behaviors of the variator 310 when the gamma angle 2395 is changed will be described below. For purposes of this description, it will be assumed that the input shaft 510 is rotating in a clockwise direction, with the observer viewing the input shaft 510 on the side of the input shaft flange 1715 that contacts the carrier input cover 802. In the following discussion, all angular direction coordinates are referenced to the angular velocity of input shaft 510.

In the illustrated embodiment, the carrier 515 rotates in the same direction as the input shaft 510 because the carrier 515 is directly connected to the input shaft 510. Thus, even if the input shaft 510 is assumed to rotate in the clockwise direction, the carrier 515 also rotates in the clockwise direction. The carrier 515 pushes on the bearings 920 connected to the planet axles 554. Because the traction rings 525A, 525B are rotationally fixed and the planets 522 roll on the traction rings 525A, 525B and the idler 562, the planets 522 rotate counterclockwise about the shaft 2390. In this embodiment, the rotation of the planet 522 about the axis 2390 is always in the opposite direction to the rotation of the carrier 515. In addition, the carrier 515 causes the planet 522 to orbit clockwise about the axis 2385. In this embodiment, the planet wheels 522 always run in the same angular direction as the rotation of the carrier 515. The rotational speed of the planet 522 about the axis 2390 can vary as the gamma angle 2395 varies, while the speed of operation of the planet 522 is in a "fixed" state determined by the rotational speed of the carrier 515.

With respect to fig. 16D, for ease of description, local coordinate systems 592A, 592B, 592C, and 592D (collectively coordinate systems 592) are defined as coordinate systems having a positive direction in the y-axis pointing outward from the central longitudinal axis of the variator 310 and a positive direction in the x-axis also pointing outward from the central portion of the variator 310, as represented by the cavity 580A. In this application, the "poles" of the planets 522 refer to points on the surface of the planets 522 that are diametrically opposite on an axis defined at the end of the planet axis 554. As used herein, "equator" refers to points on the surface of the planet 522 that lie in a plane passing through the center of the planet 522 and that intersect perpendicularly with an axis defined by the ends of the planet axles 554.

Additionally, the rotational speeds of certain components of the embodiment of the transmission 310 are shown in FIG. 16E. The speeds of the planets 522, carrier 515, and output traction rings 530, 533 are denoted as P, C and T, respectively. When the control piston 582 is in its extreme position at the centre of the chamber 580, the gamma angle 2395 is at its maximum positive value and the speed P of the planet 522 is at its minimum counterclockwise value. When the gamma angle 2395 is at a maximum positive value, the contact point is closest to the pole and the clockwise rotational speed T of the traction rings 530, 533 is at a maximum. Conversely, when the control pistons 582A, 582B are at their extreme positions furthest from the centre of the chamber 580, the gamma angle 2395 is at a maximum negative value and the speed P of the planet 522 is at its maximum counterclockwise value. When the gamma angle 2395 is at a maximum negative value, the contact point is closest to the equator, while the counterclockwise rotational speed T of the traction rings 530, 533 is at a maximum. When the gamma angle 2395 equals zero, the rotation of the planets 522 and the orbital velocity P combine to produce a zero surface velocity at the contact point. Thus, the speed T of the traction rings 530, 533 (and hence the rotational speed of the drive flange 532) is also zero.

In this example, assuming the initial value of the gamma angle 2395 is equal to zero, increasing the pilot pressure on the valve 1605 causes the control piston to move away from the center of the cavity 805A, thereby causing the pivot arms 552A, 552B to rotate about the planet 522 and push the planet axle 554 in a direction that increases the negative gamma angle 2395. Thus, the contact point moves toward the equator, the counterclockwise rotational speed of the planets 522 increases, and the counterclockwise speed of the traction rings 530, 533 also increases. Thus, in this embodiment, since the speed of the drive flange 532 is equal to the speed of the traction rings 530, 533, increasing the pilot pressure created by the gamma angle 2395 equal to zero will cause the counterclockwise speed of the drive flange 532 to also increase.

If the flow is reversed (i.e., pilot pressure is reduced when γ angle 2395 equals zero), chamber 580B will expand and chamber 580A will contract. This causes the control piston 582 and the pivot pin hubs 805 to move toward the center of the chamber 580. Pushed by the pivot pin hubs 805, the pivot arms 552A, 552B will rotate about the planet 522, pushing the planet axle 554 in a direction that increases the positive gamma angle 2395. The point of contact moves towards the pole and the counterclockwise velocity of the planets 522 decreases, causing an increase in the clockwise velocity of the traction rings 530, 533 and ultimately an increase in the clockwise velocity of the drive flange 532. Thus, in this embodiment, reducing the pilot pressure on the valve 1605 will result in an increase in the clockwise speed of the drive flange 532.

The speed of the traction rings 530, 533, and hence the speed of the drive flange 532, is a function of the surface speed of the planets 522 at the contact points. The planet axle 554 is operatively connected to the control piston 582 such that activation of the control piston 582 varies the gamma angle 2395 and results in a change in the rotational speed of the planet 522. That is, the rotational speed of the planet 522 is a function of the γ angle 2395. However, the surface speed is a function of both the rotational speed and orbital speed of the planet 522. Since the effect of the counterclockwise rotational velocity of the planets 522 on the surface velocity overcomes the effect of the clockwise orbital velocity of the planets 522, or vice versa, the direction of the speed of the traction rings 530, 533 is reversed, in effect reversing the direction of the speed of the drive flange 532.

Because the surface speed of the planet 522 can vary smoothly within a particular speed range, the transmission 310 can provide a continuously changing speed ratio. That is, certain embodiments of the transmission 310 may be used to provide a continuously variable transmission. Additionally, since in some embodiments the speed of the planets 522 is counterclockwise and the speed of the carrier 515 is clockwise, the speed of the traction rings 530, 533 (and thus the speed of the drive flange 532) can be reduced by some amount from clockwise to zero or increased by some amount to counterclockwise. Thus, since some embodiments of the transmission 310 may be in a zero power state, the above-described embodiments of the transmission 310 may be a continuously variable transmission unit that is or is applied in an infinitely variable transmission.

One embodiment of an input shaft 510 is shown in fig. 17A-17D. Shaft 510 has drive splines 1705, which may be connected to a drive shaft of a prime mover or power source or other torque transmission device. In one embodiment, drive spline 1705 is coupled to a valve housing 1710 that supports and houses a hydraulic control valve (e.g., valve 1605). Valve sleeve 1710 is coupled to a manifold flange 1715, which, in the embodiment shown, serves to transfer torque to carrier 515 and allow fluid to flow into channels 812, 814A, and 814B of carrier 515. The manifold flange 1715 may be attached to the bracket 515 with bolts (not shown) placed in the bolt holes 1720. The bores 1724, 1726 facilitate formation of a passage 1725 (partially shown) that brings the pilot pressure fluid interface 2005 into fluid communication with the cavity 2010 (see also fig. 20). The bores 817, 1781 facilitate the formation of the passage 814A for fluid communication between the valve 1605 and the chamber 580A, and the bores 818, 819 also facilitate the formation of the passage 814B for fluid communication between the valve 1605 and the chamber 580B.

The groove 1765 of the manifold flange 1715 may allow hydraulic control fluid to vent from the valve 1605 into the cavity of the variator housing 505. As previously described, during operation of the valve 1605, liquid fluid discharged from either chamber 580A or 580B will pass through the vent 1650 of the valve 1605. The exhausted fluid enters a cavity 1655 of the spool valve 1630 and flows toward the control spring 1635. The discharged fluid then passes through passages 1752, 1754 (partially shown) of the input shaft 510 and into the manifold flange 1715, through the grooves 1765, and out of the manifold flange 1715 through the flange vent 1770. The discharged fluid collects in the transmission housing 505. An external pump (not shown) will collect and circulate the control fluid as line pressure to the manifold flange 1715.

In one embodiment, valve sleeve 1710 has three fluid chambers enclosed by seal slots 1756. The first lumen includes a line pressure fluid port 2015 in communication with a line pressure fluid passageway 1730. The second chamber has a pilot pressure port 2005 communicating with a pilot pressure fluid passage 1725. The third chamber includes several lubrication ports 1758 in communication with the lubrication fluid passage 1763. Fig. 20 shows details of the various fluid channels embedded in valve sleeve 1710.

external length in one embodiment, the overall length of input shaft 510 is about 7.5 to 8.0 inches, with the outer length of valve housing 1710 being about 5 to 6 inches and the length of the drive spline being about 2 to 3 inches. In some embodiments, valve sleeve 1710 has an outer diameter of about 2 to 3 inches. In a certain embodiment, the manifold flange 1715 has an outer diameter of about 7 to 8 inches and a width of about 0.5 to 1.0 inches. For some applications, the chamber 512 is comprised of several sections, ranging in diameter from about 0.75 inches to 1.25 inches; likewise, the length may vary from about 0.5 inches to 2.0 inches. The diameter of each port and channel in the input shaft 510 is typically about 0.125 to 0.30 inches. In one embodiment, the input shaft 510 is made of SAE 8620 or SAE 1060 steel.

Fig. 18A-18E depict an embodiment of a manifold 565 that may be used with the hydraulic system depicted in fig. 16A-16D. The manifold 565 is rigidly connected between the housing 531 and the cover plate 560. The manifold 565 functions, in part, as a manifold with the lubricating fluid interface 1802 and the lubricating fluid channels 1804. In one embodiment, the lubricant channel 1804 has four branches that are spaced at 90 degrees along the center of the manifold 656. The bore 1880 facilitates forming a lubricant channel 1804 in the body of the manifold 656 that is properly inserted during operation of the transmission 310. Manifold 565 also includes a line pressure interface 1806 and a pilot pressure interface 1808 to deliver hydraulic fluid to hydraulic valve 1605. The ports 1806 and 1808 have associated hydraulic fluid passages 1810 and 1812, respectively. The bores 1882 and 1884 are suitably inserted during operation of the shifter 310 and can facilitate formation of the hydraulic fluid passages 1810 and 1812, respectively. In one embodiment, the line pressure channel 1810 has at least two branches.

In this embodiment, the manifold 565 has bolt holes 1814 to facilitate the connection of the manifold 565 to the cover plate 560. As shown in fig. 18A-18B, in this embodiment, the manifold 565 has a portion of solid material in which the channels 1804, 1810, and 1812 are formed. The manifold 565 has a central bore 1818 for supporting and mounting to the input shaft 510. To reduce weight, in this embodiment, material has been removed from the manifold 565, leaving grooves 1816.

In one embodiment, the outside diameter of the manifold 565 is up to about 11.0-11.5 inches. For some applications, the diameter of the central bore 1818 is about 2.0-3.0 inches, more suitably 2.25-2.75 inches, and most suitably about 2.5 inches. In the embodiment shown in fig. 18D-18E, the cross-sectional widest point of the manifold 565 is about 2.5 inches wide. In certain embodiments, the lubrication fluid port 1802, the line pressure port 1806, and the pilot pressure port 1808 are about 0.7-0.9 inches in diameter. The lubrication channel 1804, the line pressure channel 1810, and the pilot pressure channel 1812 may have diameters of about 0.2-0.3 inches.

Fig. 18F-18I show an embodiment of a cover plate 560 having a plurality of bolt holes 4205 for connection to an input manifold 565 and a housing 531. The cover plate 560 includes a central bore 4210 that engages a bearing element, such as a needle bearing (not shown). The cover plate 560 also has a counterbore 4215 for engaging the input manifold flange 565. Grooves 4220 are located on the outer diameter of the cover plate 560 for engaging a sealing element, such as an O-ring (not shown). In one embodiment, the overall length of the cover 560 is about 12 inches. The diameter of the counterbore 4215 is about 11.2-11.5 inches. The central bore 4210 is about 3.0 inches in diameter.

The pivot pin hubs 805 are shown in fig. 21A-21D. The pivot pin hub 805 is a generally cylindrical body with a central bore 2105 and several pin pairs 2110 aligned along the outer diameter of the cylindrical body. In the illustrated embodiment, each pin pair 2110 includes two opposing pins that can be used to receive a pivot pin block 810 (see fig. 8). One end of the pivot pin hub 805 has a flat 2115 with a bolt hole 2120 that abuts the pivot pin hub 805 to help secure the pivot pin hub 805 to the control nut 825. The other end of the pivot pin hub 805 has a recess 2125 therein for engaging a lock tab on a lock washer used to rotationally couple the pivot pin hub 805 to the control piston 582. A hole 2015 may be used to engage control nut 825 and control piston 582.

In one embodiment, the central bore 2105 has a diameter of about 1.5-2 inches, wherein the diameter of the central bore 2015 is adapted to fit the control nut 825 and/or the control piston 5832. In some embodiments, the pin pairs 2110 extend radially to a radius of about 1.5-2 inches from the center of the central bore 2105. For some applications, the width of each pin of pin pairs 2110 is approximately 0.5-1.0 inches, and the pin-to-pin spacing of each pin pair 2110 is approximately 0.3-0.5 inches. For example, in certain embodiments, the pivot pin hubs 805 are fabricated from heat treated steel SAE4140 or 4150.

As shown in the embodiment of fig. 22A-22C, control piston 582 includes a generally cylindrical body 2205 coupled to a flange 2210. A central bore 2215 passes through the cylindrical body 2205 and the flange 2210. On its periphery, flange 2210 has grooves 2220 for engaging a seal (not shown). One end of cylindrical body 2205 is reduced in outer diameter and is provided with a groove 2225. The groove 2225 may be used to engage a lock washer tab (not shown) to rotationally secure the control piston 582 to the pivot pin hub 805. In one example implementation, the control piston 582 is approximately 5 inches in overall length. The diameter of the central bore 2215 is about 1.5 inches. The outer diameter of control piston 582 on the outer periphery of flange 2210 is about 4.0-4.5 inches. The outer diameter at the surface of cylindrical body 2205 is about 2.0-2.5 inches, preferably about 2.25 inches.

The disclosure herein relates to a lubrication system for transmission 310. Generally, in one embodiment, the transmission 310 may be equipped with a lubrication system that includes a pump (not shown), a manifold 565, an input shaft 510, and a carrier 515. In some embodiments, the lubrication system utilizes the same type of fluid as the fluid used for the control fluid. In a certain embodiment, the manifold 565 may receive and distribute lubrication fluid, see fig. 18A-18E and accompanying text. The input shaft 510 and the carrier 515 may be provided with interfaces and passageways (such as the interface 1758 and the passageway 812 shown in fig. 17C and 8C) for feeding the lubrication interface 885 and the turret 887 of the carrier 515. In this manner, lubrication fluid may be injected or infused into the planet-pivot arm assembly 579.

It is observed that in certain embodiments of the transmission 310, when the carrier 515 is configured to rotate about the longitudinal axis of the transmission 310, the carrier 515 functions as a centrifugal fluid pump, tending to circulate lubrication fluid without the assistance of a separate lubrication fluid pump. This effect, according to theory, results from increased pressure on turret 887 and other lubrication interfaces, which is a centrifugal force from the lubrication fluid. In one embodiment, the radial diameter of the lubricant outlet is greater than the diameter of the lubricant inlet, and the pressure increase due to centrifugal force can be expressed by the equation P _ outlet + P _ inlet²Is calculated, wherein P _ outThe port is the pressure on the outlet, the P _ inlet is the pressure on the inlet, is the fluid density, r is the inlet to outlet radial distance, and is the rotational speed of the carrier 515. If the outlet has a fixed restriction area and/or fluid restriction, an increase in P _ Outlet will result in an increase in flow through the outlet. The increased flow draws more fluid through the system, and as long as the P _ inlet remains constant, the system flow increases in proportion thereto. This centrifugal pumping would circulate lubricant throughout the system without an external pump.

In one embodiment, the output of the drive flange 532 and the carrier spline shaft 844 are integrated by a gear set 320, which may comprise a compound planetary gear set. Fig. 23A-23C depict an embodiment of a gear set 320 coupled to a continuously variable transmission 310. Generally speaking, the gear set 320 may include a planetary gear set having a planet carrier 2305, a planet shaft 2310, planet gears 2315, a sun gear 2320, and a ring gear 2325. Certain common sub-assemblies (such as bearings, washers, etc.) often found in planetary gear sets are not shown. The sun gear 2320 is operatively connected to the ring gear 2325 through a set of planet gears 2315. In the depicted embodiment, the gear set 320 includes a driven plate 2330 connected to a ring gear 2325. In certain embodiments, driven plate 2330 is connected to and transmits torque to output shaft 585. To support the shaft 585 in the housing 590 (see fig. 5A), a bearing 2335 is mounted in a coaxial manner about the shaft 585 and guided by a bearing sleeve 2340 that contacts the housing 590. To install the housing, a bearing nut 2345, a seal 2347 and a sealing cap 2349 may be provided.

In certain embodiments, such as the embodiment shown in fig. 5A, the traction ring 525B is operatively connected to the grounded planet carrier 2305 through a torque reacting ring 525C. In the above embodiment, the planet carrier 2305 is bolted to the variator housing 505, through the center of the drive flange 532, and then reacts axial forces on the torque reacting ring 525C without a cam or thrust bearing. In a certain embodiment, the drive flange 532 is coupled to the sun gear 2320. In this embodiment there are several stationary members that react axial forces, and the drive flange 532 may be used to transfer output torque through the stationary members. The rotationally fixed planet carrier 2305 may facilitate this task. The sun gear 2320 of the planetary gear set 320 is rigidly mounted to the drive flange 532 by an interference fit. Although the sun gear 2320 and the drive flange 532 are shown here as two separate parts, the sun gear 2320 and the drive flange 532 may be one integral part.

In one embodiment, planet carrier 2305 is coupled to splined extension bar 2307, which has a splined pitch diameter of about 3.0-3.5 inches. As shown in the embodiment of fig. 23C, splined extension bar 2307 may be an integral part of planet carrier 2305. In some embodiments, the planet carrier 2305 for supporting the planet axle 2310 has an axle diameter of about 6.5-7.0 inches, with 6.75 inches being more suitable. The planet carrier 2305 may be fabricated from 4140 heat treated steel.

As shown in fig. 23D, an embodiment of a gearbox 2382 connected to the transmission 310 may include multiple planetary gear sets to provide a continuously variable transmission or infinitely variable transmission with multiple ranges or modes. In the illustrated embodiment, the gearbox 2382 includes a sun gear S1 that may be coupled to the drive flange 532. The sun gear S1 is connected to the coupled-together planet gears P1, P2 supported by the carrier C1, grounded to the housing H1. In a certain embodiment, the elongated rod C1E of the bracket C1 is attached to the reaction ring 525C (see fig. 5A). The planetary gear P2 is connected to the ring gear R1. Thus, the ring gear R1 acts as the power input member from the transmission 310, as the transfer flange 532 drives the sun gear S1, which in turn drives P2 through P1.

The ring gear R1 may be used to connect to the planet gears P3, P4 that are coupled together. The carrier C2 supports the planet gears P3, P4. The planet gear P4 is connected to the sun gear S2, and the planet gear P3 is connected to the sun gear S3. A shaft 844 (see fig. 5A) or an elongate bar thereof may be connected to sun gear S2. The carrier C2 additionally supports the planet gears P5 and P6. The planet gears P6 are connected to a shaft 2387 which can provide a combined output derived from the input of the drive flange 532 and the shaft 844. The carrier C2 also supports planet gears P5 and P6, with P5 and P6 connected to the ring gear R2.

Two clutches, a low range clutch CL and a high range clutch CH, selectively connect elements of the gearbox 2382 to a shaft 2387. The low range clutch CL may be used to connect the carrier C2 to the shaft 2387 to provide a low forward range. The high range clutch CH may be used to connect the sun gear S3 to the shaft 2387 to provide a high forward speed. A reverse clutch CR, which provides reverse mode, may be used to connect carrier C2 to shaft 2387 through sun gear S4, which is connected to ring gear R2 through planet gears P5 and P6. Thus, in certain embodiments, the transmission 310 may be connected to the gearbox 2382 to provide continuous speed change in low and high (two forward modes, two forward clutches) and reverse modes.

Fig. 23D-23F show a gamma angle testing assembly 2350 ("gamma tester 2350") that may be used to indicate the value of the gamma angle 2395 during testing or normal operation of the transmission 310. Gamma tester 2350 measures the relative rotation of the control screw 820B with respect to the bracket 515. Since the rotation of the control screw 820 is kinematically linked to the tilt of the planet axle 554, the indication of the amount of rotation of the control screw 820 directly indicates the value of the gamma angle 2395. In the illustrated embodiment, gamma tester 2350 includes a proximity probe 2352 supported by a mounting plate 2354, which may be coupled to planet carrier 2305. The locking plate 2356 abuts the mounting plate 2354 and is secured to the mounting plate 2354. In certain embodiments, the proximity probe 2352 can be a hall effect sensor, eddy current sensor, non-contact proximity sensor, or contact linear variable displacement sensor (LVDT). The proximity probe 2352 is held a suitable distance from a gamma sensor cap 2358, which covers one end of the gamma screw link 2360. In certain embodiments, the cable 2351 is connected to the proximity probe 2352 so as to convey information emitted by the proximity probe 2352 to an appropriate signal receiver (not shown).

The gamma insert 2364 is a generally cylindrical tube with one end abutting the gamma end cap 2362 and the other end threaded into the gamma screw 2366. In some embodiments, the γ insert 2364 is fixed to the bracket 515. The aperture of the gamma insert 2364 receives the gamma screw link 2360 and the gamma screw 2366, which screws into the gamma insert 2364. A spring 2365 housed in the gamma insert 2364 is coaxial with the gamma screw link 2360 and bears against the gamma end cap 2362. A gamma hex link 2363 is rigidly mounted on the control screw 820B and connected to the gamma screw 2366.

In operation, rotation of control screw 820B causes gamma screw 2366 to rotate, thereby moving axially within the threads of gamma insert 2364. As the gamma screw 2366 moves axially, the gamma screw 2366 drives the gamma screw link 2360 which has a flange 2361 that reacts against the gamma spring 2365. The spring 2365 provides a preload to prevent backlash and hold the gamma screw link 2360 against the gamma screw 2366. In a certain embodiment, the gamma cap 2358 moves axially 0.15 inches to move the planets 522a full range (e.g., +/-30 degrees). The amount of displacement of the endcap 2358 is based on the lead of the gamma screw 2366 and other space factors. Higher resolution can be achieved by providing greater axial movement of the gamma cap 2358 for a given range of gamma angles 2395.

An embodiment of a pull ring 2400 is shown in fig. 24A-24B. The traction ring 2400 can be used as any of the traction rings 525A, 525B, 530, or 533. The traction ring 2400 is generally an annular ring having a traction surface 2405. It is preferable to use the traction surface 2405 to facilitate torque transmission through the layer of elastohydrodynamic fluid created between the traction surface 2405 and the surface points of the planets 522. In the illustrated embodiment, the traction ring 2400 has an additional set of splines 2410 on its outer diameter. However, in other embodiments, the traction ring 2400 can be keyed to the drive flange 532.

In some applications, the pull ring 2400 has an outer diameter of about 12 to 13 inches and an inner diameter of between 9.5 and 10.5 inches. The pull ring 2400 can have a thickness of 1.0 to 1.5 inches. The traction surface 2405 may form an angle of about 10 to 70 degrees, more preferably 20 to 60 degrees, even more preferably 30 to 50 degrees, and most preferably 35 to 45 degrees with the straight surface 2407 of the traction ring 2400. In one embodiment, the traction ring 2400 is fabricated from SAE8630H or SAE8640 steel that has been case carburized and/or heat treated. Traction surface 2405 is more suitable if it is substantially free of contaminants.

Fig. 25A-25D illustrate an embodiment of a drive flange 532 that can be used to transmit torque out of the transmission 310. The drive flange 532 includes an annular cylindrical body 2505 having internal splines 2510 at one end and a cap 2515 at the other end. The cover 2515 is a generally circular plate with a central aperture 2520 that is connected to a shaft, gear, or other torque transmitting element, such as the sun gear 2320 of the gear set 320. As shown, in some embodiments, the cylindrical body 2505 and the cover 2515 may be unitary rather than two separate parts joined together. The drive flange 532 has a large torque capacity due to its large radius. Further, in certain embodiments, the circuit breakers 2525 of the drive flange 532 may reduce their weight. For acceleration, the drive flange 532 additionally includes a bore or opening 2530.

In one embodiment, the drive flange has an outer diameter 532 of about 14 to 14.5 inches. In certain embodiments, the pitch diameter of spline 2510 is about 13 to 13.5 inches. For some applications, the overall length of the drive flange 532 may be up to about 13 inches. The central bore 2520 may be about 4 to 5 inches in diameter, wherein, for some embodiments, a more preferred configuration of the central bore 2520 is to properly connect the drive flange 532 to the sun gear 230 of the planetary gear set 320.

Fig. 26A-26B illustrate a reaction flange 2600 that may be used in certain embodiments of the shifter 310 in place of the cam flange 705, the roller retainer 710, and the cam base 715. In one embodiment, the reaction flange 2600 includes a surface 2602 that reacts the axial force of the cam load piston 720. The internal splines 2604 of the reaction flange 2600 are sized to mate with corresponding splines of the traction ring 525A. A set of notches 2606 facilitate the use of dowel pins (not shown) to prevent rotation of reaction flange 2600. Thus, the reaction flange 2600 is used to react axial forces and prevent rotation of the traction ring 525A

In one embodiment, the reaction flange 2600 has an outer diameter of about 13.5 to 14 inches. The pitch diameter of inner spline 2604 can be up to about 12.5 inches. In certain embodiments, the reaction flange 2600 can have a width of up to about 2.5 to 3.0 inches and a central bore diameter of about 10.5 to 11 inches. In a certain embodiment, the reaction flange 2600 may be made of 4140 heat treated steel.

Fig. 27A-27B illustrate a torque transfer coupling 2700 that can be substituted for the center cam assembly 570 in the transmission 310. Coupling 2700 is generally an annular cylindrical body with internal splines 2705 at its ends. The coupling 2700 additionally has a set of external splines 2710 on its outer diameter near its mid-section. Internal splines 2705 are used to engage traction rings 533 and 530 of transmission 310. External splines 2170 are then used to engage drive flange 532. In the illustrated embodiment, coupling 2700 may have one or more circuit breakers 2715 to reduce weight and facilitate lubricant flow into transmission 310. In one embodiment, the pitch diameter of external spline 2710 is about 13 to 13.5 inches, and the diameter of internal spline 2705 is about 12.5 to 13 inches. For some applications, the torque transmission coupling 2700 may be up to about 5 to 6 inches wide. In certain embodiments, the torque transmission coupling may be made of 4140 heat treated steel.

Reaction flange 2800 is shown in fig. 28A-28B. The flange 2800 in this embodiment functions as an anti-rotational torque reaction element that prevents the traction ring 525B from rotating about the longitudinal axis of the transmission 310. In the illustrated embodiment, flange 2800 has a circular body 2805 with a set of internal splines 2810 at one end and a cap 2815 at the other end. The cap 2815 has a central bore 2820 with a set of splines 2825. The splines 2810 are connected to a corresponding set of gear splines 2410 of the traction ring 525B. The splines 2825 engage a corresponding set of splines 2307 of the planet carrier 2305, which in some embodiments is rigidly affixed to the transmission housing 505. The material breaker 2830 may be used to reduce the weight of the flange 2800 and to facilitate the flow of lubricant throughout the transmission 310. In certain embodiments, reaction flange 2800 is provided with a plurality of lubrication channels (not shown) located in circular body 2805. In a certain embodiment, reaction flange 2800 has an outer diameter of about 13 inches. In some embodiments, the pitch diameter of spline 2810 is about 12.5 and the pitch diameter of spline 2825 is about 3 to 3.5 inches. In a certain embodiment, the overall width of reaction flange 2800 is about 4 inches. For some applications, the reaction flange 2800 may be made of 4140 heat treated steel.

An embodiment of an input cam flange 2900 is shown in fig. 29A-29B. The input cam flange 2900 includes a generally cylindrical and tubular body 2905 having a set of internal splines 2910. In certain embodiments, the input cam flange 2900 has a flange 2915 that includes a set of cam extension lines 2920. The neck 2925 of the input cam flange 2900 is used to engage the rolling element retainer 710 and the cam base 715 (see fig. 7A). Between the flange 2915 and the internal spline 2910, the input cam flange 2900 has a recessed portion 2930 that is used to react against the unloading piston 725 (see fig. 7A) that abuts the flange member 2915 on the opposite side of the cam extension line 2920.

In one embodiment, the diameter of the input cam flange 2900 on the flange 2915 is about 12.5 to 13.5 inches. The diameter of internal spline 2910 is about 11.5 to 12.5 inches. The outer diameter of the neck 2925 may be about 10.8 to 11.2 inches and the inner diameter of the neck 2925 may be about 10.1 to 10.7 inches. The extension lines 2920 may be a set of eight extension lines 2922, 2924 angularly aligned around the center of the input cam flange 2900. In one embodiment, the extension lines 2920 are appropriately arranged and cooperate with the extension lines 3002, 3003 of the cam base 710 (see FIGS. 30A-30D). For some applications, the extension lines 2922, 2924 are approximately 1.25 to 2.0 inches wide. In one embodiment, the lead of the lines 2922, 2924 is about 1.1 to 1.5 inches, more preferably 1.2 to 1.4 inches, and most preferably 1.3 inches. The start line 2922 has a counterclockwise spiral start line surface and the start line 2924 has a clockwise spiral start line surface. In a certain embodiment, the input cam flange 2900 may be made of a metallic material (e.g., 1065 steel). It is more suitable that the propagation surface of the propagation line 2920 be flame or induction quenched to 58-62HRC, with a minimum effective depth of layer of about 0.03 inches.

An embodiment of the cam base 710 is shown in fig. 30A-30D. The cam base 710 is typically an annular ring with a set of cam extension lines 3005 on one side. On the outer edge of the cam base 710 are several recesses 3010 that can be used to engage torque reacting dowels (not shown) that react the torque in the cam base 710 to the transmission housing 505 and prevent the cam base 710 from rotating. The cam base 710 also has a set of bolt holes 3015 that allow the cam base 710 to be attached to the housing 531 during assembly. After assembly of the transmission 310, the bolts can be unscrewed and the cam base 710 removed from the housing 531. This may cause the cam load piston 720 to axially actuate the cam base 710. The bolt holes 3015 are then plugged with plugs (not shown). On the opposite side of the cam extension line 3005, the cam base 710 has a flat surface 3020 that can bear against the housing 531 and engage the cam load piston 720 (see FIG. 7A).

In one embodiment, the cam base 710 has an outer diameter of about 13.7 inches, an inner diameter of about 11.0 inches, and a thickness of about 0.5 to 0.6 inches. As shown in fig. 30B, in an embodiment, the cam base 710 includes eight sets of extension lines 3002, 3003, angularly aligned along the center of the cam base 710. The width of the extended lines 3002, 3003 is about 1.25 to 2.0 inches. In one embodiment, the lead of the extension lines 3002, 3003 is about 1.1 to 1.5 inches, more preferably 1.2 to 1.4 inches, and most preferably 1.3 inches. The start line 3002 has a counterclockwise spiral start line surface, and the start line 3003 has a clockwise spiral start line surface. In a certain embodiment, the cam base 710 may be made of a metallic material (e.g., 1065 steel). It is more suitable that the starter wire face of the starter wire 3005 be flame or induction quenched to 58-62HRC, with a minimum effective depth of layer of about 0.03 inches.

Fig. 31A-31B illustrate an embodiment of a cam load piston 720, which is a generally annular flange with a flat surface 3105 and a recessed portion 3110, the recessed portion 3110 being located on opposite sides of the flat surface 3105. The recessed portion 3110 may engage a set of compression springs 735 (see fig. 7A). In the illustrated embodiment, cam-loaded piston 720 includes seal ring grooves 3115 and 3120. Cam load piston 720 may engage recess 4104 of clutch housing 531 as shown in FIG. 7A.

In certain embodiments, the cam load piston 720 may act as an axial force sensor on the traction ring 525A by way of pressurizing and sealing fluid volume of the cam load piston 720, thereby making the bore 735 of the housing 531 a zero leakage fluid reservoir. As the axial force on the cam-loaded piston 720 increases, the pressure on the piston bore increases proportionally.

An embodiment of the unloader piston 725 is shown in fig. 32A-32C. The unloader piston 725 is typically an annular flange with a ring 3205, the ring 3205 being used to engage the input cam flange 2900. The unloading piston 725 additionally includes a second ring 3210 opposite the first ring 3205 for engaging the unloading cylinder 730 (see fig. 7A). As shown in fig. 32D, the unloader piston 725 can be equipped with seal grooves 3215 and 3220 to engage seal rings (not shown). Fig. 33A-33B depict an embodiment of the unloader cylinder 730, which is a generally annular cylindrical body with a cavity 3305, the cavity 3305 being configured to engage a ring 3210 of the unloader piston 725. Port 3310 of unload cylinder 730 may be used to receive hydraulic fluid into cavity 3305.

Fig. 34A and 34B illustrate one embodiment of a central cam base 605, which is a generally annular cylindrical body having a set of external splines 3405 on its outer race. Splines 3405 are used to engage corresponding splines 2510 of drive flange 532 (see FIG. 6A). Each side of the central cam base 605 has a cam extension 3410 for cooperating with rolling elements (not shown) on the cam rings 610 and 615 and corresponding cam extensions (see fig. 6A and accompanying discussion). In one embodiment, central cam base 605 has an inner diameter of about 11 inches and an outer diameter of about 12.5 to 13 inches at the root of spline 3405. In some embodiments, the pitch diameter of spline 3405 is about 13.0 to 13.5 inches. The width of the central cam base 605 (excluding the height extending along line 3410) may be up to about 1.5 to 1.7 inches.

In an embodiment, the central cam base 605 includes eight sets of lines 3412, 3414 (on each side of the central cam base 605) angularly aligned around the center of the central cam base 605. The width of the extension lines 3412, 3414 is about 1.25 to 2.0 inches. In one embodiment, the leads of the extension lines 3412, 3414 are about 1.1 to 1.5 inches, more preferably 1.2 to 1.4 inches, and even more preferably 1.3 inches. The line 3412 has a counterclockwise helical line surface, and the line 3414 has a clockwise helical line surface. In a certain embodiment, the central cam base 605 may be made of a metallic material (e.g., 1065 steel). More suitably, the surface of the starter wire 3005 is flame or induction quenched to 58-62HRC, with a minimum effective depth of layer of about 0.03 inches.

Embodiments of cam rings 610, 615 are shown in fig. 35A-35C. For convenience, the following discussion will refer only to the cam ring 610; however, the discussion may apply equally to the cam ring 615. The cam ring 610 is generally a flange with a set of internal splines 3505. In some embodiments, cam ring 610 includes a cam neck 3510 and a cam shoulder 3515 having a set of cam extension lines 3520. In this embodiment, the cam ring 610 is provided with a keyway 3525 that can engage a key (not shown) that secures rotation of the cam ring 610 to the synchronizing ring 645 (see fig. 6A). The cam ring 610 may include a spring retainer groove 3527 for engaging a spring retainer (not shown) of the secondary axial restraining synchronizing ring 645.

In one embodiment, the maximum outer diameter of cam ring 610 is about 13 inches. The pitch diameter of the splines 3505 may be up to about 12.5 inches. In certain embodiments, the cam neck 3510 has an outer diameter of about 11 inches and an inner diameter of about 10.5 inches. In some embodiments, cam ring 610 has a cross-sectional width of about 2 to 2.5 inches. It is more appropriate that the cam extension line 3520 be used to cooperate with the extension line 3410 of the central cam base 605. In an embodiment, the cam ring 610 includes eight sets of extension lines 3522, 3524 arranged in an angular shape around the center of the cam ring 610. The width of the extension lines 3522, 3524 is about 1.25 to 2.0 inches. In one embodiment, the wires running from wires 3522, 3524 are about 1.1 to 1.5 inches, more preferably 1.2 to 1.4 inches, and even more preferably 1.3 inches. The extension line 3522 has a counterclockwise spiral extension line surface, and the extension line 3524 is a clockwise spiral extension line surface. In one embodiment, cam ring 610 can be made of a metallic material (e.g., 1065 steel). More suitably, the surface of the starter wire 3005 is flame or induction quenched to 58-62HRC, with a minimum effective depth of layer of about 0.03 inches.

Fig. 36A-36B depict one embodiment of an output tray 620 (which may be identical to output tray 625). Output disc 620 is a generally annular cylindrical body with a set of internal splines 3605 and external splines 3610. In this embodiment, the internal splines 3605 are used to engage a corresponding set of splines 2410 of the traction ring 2400 (see fig. 6A and 24A-24C). The external splines 3610 are used to engage a corresponding set of splines 3505 of the cam ring 610. Output disk 620 also has flange extension 3615 for engaging bearing 630, carrier guide ring 640 and center bearing shim 642 (see fig. 6A). Shoulder 3625 of output disc 620 abuts cam ring 610 to transmit axial force. A circuit breaker 3620 is positioned within the output tray 620 to reduce its weight and facilitate lubricant flow.

In a certain embodiment, the outer diameter of output disc 620 is about 12.5 to 13.5 inches. The pitch diameter of either internal spline 3605 or external spline 3610 can be up to about 12 to 13 inches. In certain embodiments, the overall length of the output disc 620 is about 4.5 to 5.5 inches. For some applications, flange extender 3615 has a length of about 1.5 to 3 inches and an inner diameter of about 7.5 to 8.5 inches. In certain embodiments, the output disc 620 may be made of 4140 heat treated steel.

An embodiment of a carrier guide ring 640 is shown in fig. 37A-37B. Carrier guide ring 640 is a generally annular cylindrical body with internal grooves 3705 for lubrication distribution. Guide ring 640 also includes an aperture 3710 for lubrication distribution. In some embodiments, the carrier guide ring 640 has an outer diameter of about 6.5 inches and an inner diameter of about 6 inches. As shown in fig. 37B, the cross-sectional thickness of the carrier guide ring 640 is about 0.6 to 0.8 inches. In one embodiment, the groove 3705 is about 0.1 inches in diameter and the aperture 3710 is about 0.25 inches in diameter. Fig. 38A-38B illustrate an embodiment of a synchronizing ring 645, which is generally an annular cylindrical body with a keyway 3805. In one embodiment, the synchronizing ring 645 has an inner diameter of about 10 inches and an outer diameter of about 10.5 inches. In some embodiments, the length of the synchronizing ring 645 is approximately 2 to 2.5 inches. For some applications, the keyway 3805 is about 0.25 inches wide and 0.125 inches deep. In certain embodiments, the synchronizer ring may be made of mild steel.

An embodiment of an idler assembly 3900 is shown in figures 39A-39C. Idler assembly 3900 may include an idler 3905, which is generally an annular ring with a flanged extension rod 3910 and a shoulder 3915. The flange extender 3910 is used to provide a rolling surface for the planets 522. Shoulder 3915 is used to support bearing element 3920, shown in the embodiment as a typical radial ball bearing; however, in other embodiments, the bearing element 3920 may be an angular contact bearing. Bearing elements 3920 provide radial and axial support for idler 3905. In one embodiment, the inner diameter of idler 3905 is about 4.0 inches. In certain embodiments, shoulder 3915 has a diameter of about 4.5 inches. For some applications, idler 3905 on flange extender 3910 is approximately 5.25 to 5.75 inches in diameter. The idler wheel 3905 may be made of a metallic material (e.g., 8620 steel). In one embodiment, idler 3905 is case carburized to the HRC60 surface with a minimum layer depth of about 0.60 inches. The bearing elements 3920 may be Kaydon KD045AH6 bearings.

Fig. 40A-40E illustrate an embodiment of a transmission housing 505 that can be used with the transmission 310. The transmission housing 505 is typically a cylindrical receptacle 4005 that is connected to a bushing 4010, which in turn is connected to an oil pan (not shown) by suitable fastening methods such as bolts, welding, adhesives, etc. However, in other embodiments, the container 4005 may be other than cylindrical. For example, the receptacle 4005 may be a rectangular box. As previously described, the sleeve 4010 and sump form an oil reservoir or sump from which a pump (not shown) may be replenished to provide line pressure to the valve 1605. In one embodiment, sleeve 4010 includes an acceleration port 4020 for fluid communication with a pump. In addition, the sleeve 4010 can be provided with a plurality of ancillary interfaces 4025, some of which can be used to receive thermocouples. In certain embodiments, the sleeve 4010 or, in general, the transmission housing 505, can include one or more utility interfaces 4030 to facilitate access to various components or processes of an auxiliary gearbox (such as the gearbox 320 or the gearbox 2382). In one embodiment, the sleeve 4010 is equipped with a port 4035 for fluid level viewing.

The pod 4005 may contain multiple circuit breakers 4015, which may be Plexiglas^TMA window or the like to view the components housed by the container 4005. In a certain embodiment, the reservoir 4005 includes a valve piston interface 4040 for delivering fluid to the valve piston 725 (see fig. 7A and accompanying text). The pod 4005 is also equipped with a plurality of instruments and interfaces 4045. For example, in certain embodiments, certain interface 4045 may be used to support and house a speed sensor. To provide lubrication to certain components of the auxiliary gearboxes 320, 2382, the reservoir 4005 may contain one or more lubricantsInterface 4050. A plurality of auxiliary lubrication interfaces 4055 may be provided in the reservoir 4005 to supply additional lubrication to the planets 522. To assist in handling the transmission housing 505, including any components assembled therein, the receptacle 4005 can include one or more attachment holes 4060 for attachment to a suitable lifting or manipulating tool or machine. It is more appropriate that attachment holes 4060 be properly positioned in receptacle 4005 to provide a pivot point about the center of gravity of shifter housing 505 and/or shifter 310. The receptacle 4005 may receive alignment dowels (not shown) with dowel holes 4065 to facilitate positioning and assembly of the transmission housing 505 with the housing 531 and/or gearboxes 320, 2382. In some embodiments, receptacle 4005 may be fitted with a recess 4091 for engaging, supporting and axially restraining some components of transmission 310; for example, in one embodiment, the recess 4091 is used to axially restrain the unload cylinder 730 (see FIG. 7A).

In some embodiments, the container 4005 may be provided with an end plate 4063 with a hole 4070 to engage and support the planet shafts 2310 of the gearbox 320 (see fig. 23C). Plate 4063 may also be equipped with a plurality of carrier bolt holes 4075 for engaging bolts that secure planet carrier 2305 to container 4005. Plate 4063 has a central aperture 4085 to enable certain components of transmission 310 (such as drive flange 532) and certain components of gearboxes 320, 2382 (such as sun gear 2320) and their passage and engagement with each other. To facilitate positioning and assembly, the plate 4063 may contain one or more pin holes 4080 to receive alignment pins (not shown). In some embodiments, the plate 4063 may be equipped with a seal groove 4095 to accommodate an O-ring (not shown) that can provide a seal between the transmission housing 505 and the gear set 320 housing 590. The transmission housing 505 may be equipped with a lubrication interface 4090 to allow lubricant to flow back from the gear set housing 590 to the transmission housing 505 so that the housing 590 is not flooded with excess lubricant. In certain embodiments, the transmission housing includes one or more lubrication interfaces 4087 that carry lubricant from the transmission housing 505 to the gear set housing 590 to lubricate the clutches and gear sets.

In one embodiment, the transmission housing 505 has an overall length of about 21 to 22 inches and an overall height or outer diameter of about 16.5 to 17.5 inches. The cylindrical container 4005 has an outer diameter of about 15.5 to 16.5 inches and an inner diameter of about 14.5 to 15.5 inches. In one embodiment, the sleeve 4010 generally encloses a volume having dimensions of about 14x13x6 inches. For some applications, the transmission housing 505 may be made of heat treated steel.

An embodiment of the housing 531 attached to the transmission housing 505 is shown in fig. 41A-41E. The housing 531 is a generally cylindrical body 4101 with a central passage 4103. As also shown in fig. 7A, in one embodiment, a housing 531 is used to house and support certain components of the transmission 310. The housing 531 includes a recess 4104 for receiving the cam load piston 720. The housing 531 may additionally include a plurality of apertures 755 that receive and support the compression springs 735. The grooves 4107 of the housing 531 can be used to receive and support the cam base 715, the roller retainer 710, and/or the cam flange 705. The recesses 4108 of the housing 531 may be suitably adapted to receive and support the cover plate 560 and the manifold 565 (see fig. 5A). A plurality of pin holes 4118 may be provided in the housing 531 to receive alignment pins (not shown) that may position the cam base 715 and provide anti-rotation functionality to the cam base 715 (see fig. 7B). To facilitate assembly of the cam base 715 in the housing 531, when the cam load piston 720 and/or the compression spring 720 are used, the housing 531 can include a plurality of bolt holes 4126 to receive bolts (not shown) that secure the cam base 715 in place when certain components of the input cam assembly 575 are assembled.

For some applications, the housing 531 may be equipped with a plurality of ports for fluid communication with the manifold 565. In a certain embodiment, the housing 531 includes a lubrication interface 4112 for delivering lubrication fluid to the manifold 565. The housing 531 may be equipped with a cam load piston pressure interface 4114 to communicate fluid pressure to the cam load piston 720. A pilot pressure interface 4122 may be included in housing 531 to communicate fluid pressure to manifold 565 to urge pilot control piston 1620 (see fig. 20 and accompanying text). In some embodiments, the housing 531 may be equipped with a line pressure interface 4124 that delivers fluid pressure to a manifold 565 for delivery to a hydraulic control valve 1605, which ultimately uses the line pressure fluid to actuate a control piston 582. To facilitate positioning and assembly of the housing 531 with the transmission housing 505 and/or an engine housing (not shown), the housing 531 may be drilled with pin holes 4102, 4110, respectively. In some embodiments, it may be more suitable to provide housing 531 with a recess 4106 to receive and/or support a starter motor (not shown).

In one embodiment, the outer housing 531 has an overall outer diameter of about 17 inches and a length of about 8.5 inches. Grooves 4107 are about 14 inches in diameter and about 2 inches long. Grooves 4104 have an outer diameter of about 12.5 inches, an inner diameter of about 10.5 inches, and a depth of about 0.75 to 1.0 inches. Grooves 4108 have a diameter of about 12 inches and a depth of about 1.5 to 2.0 inches. In certain embodiments, the diameter of the central passage 4103 is about 9 inches. In a certain embodiment, the holes 755 have a diameter of about 0.6 inches and a depth of about 0.6 to 1.0 inches. In one embodiment, the housing 531 may be made of ductile iron 80-55-06.

The embodiments described herein are examples provided to meet the illustrative requirements set forth by law and are examples provided to provide examples. These examples are merely examples that may be employed by any one party and are not intended to be limiting in any way.

Claims (9)

1. A method of shifting an infinitely variable transmission having an input shaft and a plurality of planet-pivot arm assemblies, each planet-pivot arm assembly having a planet mounted on a planet shaft, wherein each planet shaft is connected to a pivot arm configured to tilt the planet shaft, the method comprising the steps of:

operatively connecting a feedback mechanism with the plurality of planet-pivot arm assemblies;

operatively connecting the regulator with a feedback mechanism;

communicating one or more indications of the status of the plurality of planet-pivot arm assemblies from the feedback mechanism to the regulator;

receiving a control signal at a regulator;

adjusting a hydraulic pressure using a regulator, the hydraulic pressure based at least in part on a combination of the control signal and one or more indications of a state of the plurality of planet-pivot arm assemblies;

supplying the hydraulic pressure from the regulator through a plurality of passages and cavities in a carrier of the transmission, the carrier being coupled to each planet-pivot arm assembly, wherein at least some of the passages and cavities are formed in the input shaft; and

ratio adjustments are made by using hydraulic pressure to push a plurality of planet-pivot arm assemblies.

2. The method of claim 1, further comprising operatively connecting a synchronizer to the plurality of planet-pivot arm assemblies, the synchronizer configured to operatively connect to a feedback mechanism.

3. The method of claim 1, wherein receiving a control signal includes providing a pilot pressure indicative of a desired gear ratio.

4. The method of claim 1, wherein operatively connecting the regulator with the feedback mechanism comprises the steps of:

providing a hydraulic control valve;

operatively connecting a hydraulic control valve with a plurality of hydraulic pistons;

operatively connecting a plurality of planet shafts with at least one of the plurality of hydraulic pistons;

the hydraulic pressure is adjusted through a hydraulic control valve;

supplying the hydraulic pressure from a hydraulic pressure control valve to at least one of the plurality of hydraulic pistons; and

the change in the inclination of the planet axles is actuated hydraulically.

5. The method of claim 4, wherein actuating the planet axle tilt angle change comprises axially moving at least one of the plurality of hydraulic pistons.

6. The method of claim 4, further comprising the steps of: a pilot pressure is provided to the hydraulic control valve to indicate the desired inclination of the planet axles.

7. The method of claim 6, wherein providing the pilot pressure comprises setting a resolution of the pilot pressure to about 20psi per 1 degree of inclination.

8. The method of claim 4, further comprising the steps of:

connecting the hydraulic piston with a feedback spring; and

the tilting of the planet shaft is actuated at least partly in dependence on the pilot pressure, whereby the transmission is shifted by tilting the planet shaft.

9. The method of claim 8, wherein coupling the hydraulic piston with the feedback spring comprises providing means for adjusting a preload of the feedback spring, the preload of the feedback spring being used to establish the magnitude of the pilot pressure.