WO1996007040A1

WO1996007040A1 - A variable speed planetary transmission

Info

Publication number: WO1996007040A1
Application number: PCT/US1994/009927
Authority: WO
Inventors: Mark J. Egyed
Original assignee: Egyed Mark J
Priority date: 1991-08-16
Filing date: 1994-08-29
Publication date: 1996-03-07

Abstract

In power transmissions having power transmitting structures (141, 142, 147) available for selective coupling to a plurality of power transmitting paths (252, 266, 366), increased kinematic degrees of freedom result by allowing a reaction element (128) grounded in the transmission casing (100), to be inserted topologically inside a power transmitting structure or path (144, 147, 252, 266, 366). Interior placement of the reaction element (128) allows that gearing elements (172) or power transmitting paths (266) normally blocked from access to the transmission case (100) may be selectively coupled to either a power transmitting structure (147) op to a braking structure (122), substantially increasing the number of possible ratio states.

Description

A VARIABLE SPEED PLANETARY TRANSMISSION

1. Field of the Invention

This invention relates to mechanical power transmissions, either vehicular or stationary, where output angular speed or torque is modulated or systematically varied. Emphasis in this disclosure is on automotive transmissions, which provide unique engineering challenges as described below. However, as will be evident, the teaching presented here can be applied to power transmissions generally.

2. Background and Description of the Prior Art

Civilian conservation of motor fuels figures importantly in the health of national economies and air quality worldwide. To date, great efforts have been made worldwide to reduce automotive fuel consumption, including engineering for vehicle weight reduction, creation of aerodynamic body designs, and improved design of engines, tires, braking systems, fuels and lubricants. However, large potential savings remain in better matching the operation of automotive prime movers to their loads. In developing usable output power, the fuel efficiency of internal combustion engines and most other prime movers varies greatly with output speed and torque. There are many times, such as during highway cruising, where engine power requirements are modest (under 12 kW for passenger vehicles) , that engine speeds should be reduced to save fuel. Although recent improved automotive transmissions have better exploited this, drivetrains that allow functioning of the prime mover at optimum or near optimum efficiency for all operating regimes have proved impractical in practice.

To produce the full spread of allowable output power most efficiently given exhaust emissions constraints, one needs to follow an ideal operating schedule for the engine in question. This requires using a wide ratio span transmission that will permit operating the prime mover at low speeds, and with high mean effective pressures, when power requirements are modest.

To provide the engine-drivetrain control needed, engineers have long envisaged storing brake specific fuel consumption and exhaust emission information in electronic PROM (programmable read- only memory) , and using a microprocessor electronic control unit (ECU) , generating command signals that follow a predetermined optimum operating schedule. In this manner, using techniques known in the art, extensive use can be made of what is commonly referred to as overdrive gearing to allow frequent low speed, high torque operation of the engine.

To achieve brake specific fuel consumption reductions, a recent study has recommended that a wide ratio span 8 speed automatic be developed [ref: SAE Paper 810446, "Engine Transmission Matching,'' R. H. Thring, Ricardo Consulting Engineers, 1981].

Although many multiple speed or continuously variable ratio transmissions (CVT's) have been proposed to provide better engine- drivetrain matching, not one has yet satisfied nine essential suitability criteria relating to [1] Transmission Cost; [2] Reliability; [4] Power Density and Torque Capacity; [5] Driveability and Customer Satisfaction; [6] Low Transmission Energy Losses; [7] Ratio Range; [8] Distribution and Size of Ratio Increments; and [9] Ease of Ratio Modulation or Control.

Prior Art Using Epicyclic Drivetrains

Because of the problems encountered with traction and belt drives transmissions, conventional gearing with clutching into discrete ratio states continues to be used for nearly all transmissions for automotive applications. Epicyclic or planetary gearing arrangements are widely used and offer many advantages. They are always in mesh, allowing use of power shifts where the prime mover proceeds uninterrupted during ratio changes. Their design, using multiple radially-spaced pinions, allows several sets of teeth to be in mesh at once, distributing forces to allow for greater strength and torque ratings. Because the shafts and other rotating structures used with planetary gear trains can be arranged on the same centerline, a very compact unit can be realized. With proper ratio control, any wide ratio range transmission possessing numerous discrete ratio steps can reap all or nearly all the fuel economy savings afforded by using a continuously variable ratio transmission (CVT) .

Many multi-speed planetary change-gear transmission designs exist that have a high number of available ratios. However, none satisfy all nine of the above suitability criteria. For example, the binary incrementally variable transmission of Kerr, US Pat. 4,559,848 allows for a large number of equally spaced forward speeds, and could satisfy criteria relating to reliability, driveability, and ease of ratio control, but it uses a large number of complex or twin pinion epicyclic gearsets, and is not suitable because of deficiencies in areas relating to cost, space constraints, power density and efficiency.

Other transmissions presently in use nearly satisfy all criteria except those relating to ratio range and incrementation. At present, to remedy this by adding to these units component gearsets and clutch hardware to give additional well-spaced ratio states and better engine-transmission matching would drive up the unit cost and size considerably. One such presently used transmission is the Automatic Overdrive^tm transmission for rear wheel drive applications made by Ford Motor Company. It uses a compound planetary gearset having

2 sun gears, 3 short pinions, 3 long pinions, and one internal or ring gear, with 4 friction clutches, 2 one-way clutches, and 2 friction bands to produce 4 forward speeds and one reverse. Another similar unit for front wheel drive applications is the Ford AXOD*™ overdrive 4 speed transaxle transmission which uses 3 simple planetary gearsets including a reduction gearset, 4 multiple-disc clutches, two friction bands, and two one-way clutches [ref: US Pat. 4,509,389, Vahratian, et al]. Another such four speed overdrive front wheel drive transaxle transmission is the Hydra- Matic*™ THM 440 T4 transaxle made by General Motors. It also uses

3 simple planetary gearsets including a reduction gearset, 4 multiple-disc clutches, 2 friction bands, and two one-way clutches. Recently General Motors produced the state-of-the-art Hydra- Matic^tm 4T60E, which has an additional sprag clutch and friction band for smoother operation.

To obtain additional ratio states from these and other conventional transmissions for better engine-transmission matching it is necessary to add one or more component gearsets, additional clutches and other hardware. An example is the 5 speed automatic transmission recently developed by Nissan Motor Company, Limited, of Japan. To obtain a fifth ratio state, an auxiliary planetary gearset and additional clutches were added to a conventional 4 speed unit. This brings the total number of planetary gearsets to 3, plus another reduction gearset needed for final reduction gearing. This adds considerably to the cost and axial length of the geartrain. Meshing losses are also increased, lowering the overall transmission efficiency. Similar units have been produced by Toyota of Japan and Mercedes Benz of West Germany.

At present, a transmission unit for passenger vehicles that has acceptable ratio incrementation and control and other characteristics, and has a minimum of friction producing devices to produce 7 to 9 speeds, will require 4 to 5 simple planetary gearsets or the equivalent when reduction gearing is taken into account, and generally will not have tightly spaced ratio increments in the overdrive ratio zone. One example is provided by Hiraiwa, US Pat. 4,653,348 where 3 simple planetary gearsets, 7 friction-producing devices and 3 one-way clutches are use to produce 7 forward speeds. An additional planetary gearset or the equivalent is necessary for final drive reduction. Another example is a multi-speed unit with acceptable ratio control afforded by the transmission gearing arrangement of Klemen, US Pat. 4,683,776 which uses 4 simple planetary gearsets, 6 friction-producing devices and, in practice, an unspecified number of one-way clutches, to produce 9 forward speeds. An additional reduction gearset is still necessary for automotive use, bringing the total to five. Klemen, US Pat. 4,976,670 discloses power transmissions using three planetary gearsets, including one utilized solely as a reversing gear arrangement, to provide 7 forward speeds and one reverse. Because it is topologically impossible to provide independent inputs to all gearing elements used, the two planetary gearsets not used for reversing are doubly linked — the carrier of the first gearset is permanently coupled to the ring of the second gearset, and also the ring of the first gearset is permanently connected to the sun of the second. This limits the number of possible ratio states and the maximum ratio range obtained. The gaps between transmission ratios for the low speeds are large. Additional "very low" forward and reverse speeds can be added, bringing the total number of ratio states to eight forward and two reverse speeds, by adding an optional (fourth) planetary gearset to the drivetrain disclosed. In any case, an additional final reduction gearset must be used for automotive use to give the high overdrive ratios needed for optimum engine operation at cruising speeds. For automotive use, this brings the total number of planetary or equivalent gearsets needed to four or five.

The problem of needing a high number of planetary or other gearsets to produce multi-speed transmissions with a high number of ratios and a high ratio range is not confined to automotive applications, but applies to all power transmission devices. It is also not confined to a particular type of component gearset or compound geartrain — the problem is that the number of kinematic degrees of freedom, or independent shaft speeds, that can be delivered to the constituent gearing elements of any geartrain has always been fewer than that theoretically possible.

OBJECTS OF THE INVENTION

Accordingly it is a broad aim of this invention to: [1] introduce a method to provide additional available kinematic degrees of freedom for every available power transmitting path or unlinked element in any component gearset, as compared to present gearing arrangements;

[2] allow combinational or independent actuation of clutch or torque transfer devices to create useful ratio states; [3] to deliver to each independent (unlinked) element of any geartrain at least one kinematic degree of freedom, making better use of gearing components;

[4] allow direct coupling of the transmission input to any or all independent elements of any geartrain so as to allow the theoretical maximum number of ratio states possible using a minimum of clutch and gearing devices; [5] allow use of any gearing arrangement using conventional coaxial gearing components, with fewer planetary or other gearsets, to form inexpensive high multi-speed transmissions where no final ratio reduction is required for automotive applications, whereby wide range near-CVT operation is obtained in a compact package with time-proven inexpensive components at low technical risk, with no need for countershafts or power transfer cases; whereby the resulting ratio range is large enough to provide both very low ratio states for brisk acceleration and climbing ability, and high overdrive states that allow efficient operation of the prime mover under low power conditions, whereby longer engine life can be obtained through the extensive use of overdrive and high overdrive gearing, with quieter operation; whereby energy conversion efficiency is comparable to present day transmissions; whereby multiple closely-spaced overdrive ratio states eliminate new or alien sensations for vehicle occupants, allowing smooth inconspicuous changes in transmission ratio to suit operating conditions at cruising speeds where most travel occurs;

Other objects of this invention not given above will become clear from further reading of the specification.

SUMMARY OF THE INVENTION

According to the present invention, a input arrangement structure that I call an "input transmitter" is provided to a geartrain whereby a plurality of coaxial shafts or equivalent structures are made available for rotational coupling thereto, such that a torque transmitting structure grounded in a suitable reference frame, usually the transmission casing, is inserted topologically inside a power transmitting structure or power transmitting path. This torque transmitting structure provides a component a reaction stator or reaction element through which reaction forces may be supplied for the purpose of grounding selected gearing elements of the geartrain. Normally the grounding or restraining of selected gearing elements or power transmitting paths is accomplished by using brake clutches that couple the power transmitting path to the transmission casing. This requires that the power transmitting path to be braked is directly accessible to a braking mechanism fixed to the transmission casing. This invention provides for introducing a reaction element into the interior of power transmitting paths or structures so that gearing elements or other power transmitting paths located therein, and blocked from direct access to the transmission case or a mechanical reference frame, may also benefit from being selectively braked or restrained when desired. As described below, this allows that selected power transmitting paths may be selectively coupled to either a driving structure or a braking or restraining structure, giving a substantial increase in the number of possible ratio states available by yielding a greater number of kinematic degrees of freedom for the transmission overall.

One form of an input transmitter that accomplishes this topological insertion of a reaction element provides a set of three or more coaxial or substantially coaxial shafts or equivalent structures that are made available for rotational coupling to a plurality of power transmitting paths, such that no two constituent radially adjacent coaxial shafts or structures normally have the same angular speed. The substantially coaxial or nested arrangement of these structures, from the inner structure to the outer structure must be such that at least one reaction stator or reaction element forms one of the intermediate elements, that is, elements other than the innermost or outermost coaxial structures.

Another equally useful form of input transmitter topologically inserts a reaction element into the interior of a power transmitting path or structure by allowing the path or structure to be split into two separate elements. A power bridge is provided to maintain torque handling continuity from one element to the other, while a torque transmitting structure that serves as a reaction element is inserted therebetween, thereby giving the reaction element access to an interior space that is normally "forbidden." At least one reaction stator or similar member is always present, topologically inserted into the interior of a shaft or rotor. Placement of additional reaction elements may be arbitrary. Any reaction stator may become a reaction rotor when replaced by an equivalent structure in a non-stationary reference frame, such as a structure that is motor-driven at constant or accelerating angular speed.

The gearing arrangements obtained from practicing this invention include combinational incrementally variable transmissions (CIVTs) . These combinational incrementally variable transmissions have a minimum number of gearset linkages. They fully utilize the elements of component geartrains such that the number of possible distinct ratio states Z of any geartrain or compound geartrain can be Z = 2ⁿ - 1

or more, where n is the number of elements not permanently linked or coupled to other elements in said tree. The geartrain(s) employed using this invention may contain any type of gearset, e.g., simple and twin pinion planetary types, compound planetary types with long and short pinions (e.g, Ravigneaux type) , single axis differential gearsets, or may be mixed to make compound gearsets, or have additional transmission mechanisms. As used in this disclosure, a gearset is defined as a set of geared or interconnected machine elements arranged such that their total number of kinematic degrees of freedom is one less than the number of gearing or machine elements. An example is the simple planetary gearset, where the 3 elements commonly known as the sun, ring, and carrier (possessing one or more free-rotating pinions) , have a linear kinematic relationship allowing for 2 kinematic degrees of freedom.

BRIEF DESCRIPTION OF THE DRAWINGS

The following detailed disclosure describes the many aspects of these novel transmissions and refer to the accompanying drawings, in which:

FIG. 1 is an upper half-plane schematic representation of a transmission using two singly coupled simple planetary gearsets.

FIGS. 2(a) - (r) show a set of possible schematic gearing configurations that can be substituted for the right-hand portion of the schematic representation of FIG. 1. FIG. 3 is a schematic representation similar to that of FIG. 1, but with additional clutches that add forward and reverse ratio states and engine braking capability.

FIGS. 4 and 5 show the input transmitters of FIGS. 1 and 3 schematically in the upper and lower half-planes, with dual drive links.

FIG. 6 shows one type of epicyclic input transmitter schematically in the upper and lower half-planes, using a single drive link.

FIGS. 7, 8, 9, 10, 11, and 12 are schematic upper and lower half-plane views of various alternative complex input transmitters. FIG. 13 shows a complex input transmitter in the upper and lower half-planes, having two reaction stators and three driving/driven sprockets fed by a single wide drive link. FIG. 14 shows a complex epicyclic input transmitter in the upper and lower half-planes, having two reaction stators and five driving/driven gearing elements.

FIG. 15 shows another complex epicyclic input transmitter having two centrally-grounded reaction stators and four driving/driven gearing elements.

FIG. 16 shows a schematic representation for a compact geartrain, that expands upon the geartrain shown in FIG. 1, and which may be used as a front drive transaxle having eight forward speeds and one reverse. FIG. 17 is the embodiment of FIG. 16 having no drive link and suitable for rear wheel drive application.

FIG. 18 shows a schematic representation for a transmission that yields three forward speeds and one reverse using one simple planetary gearset. FIG. 19 shows a schematic representation for a transmission that yields four forward speeds and one reverse using one simple planetary gearset.

FIG. 20 shows a partial schematic representation for a transmission that yields five forward speeds and one reverse using one simple planetary gearset. The full schematic can be obtained by substituting the partial schematic as shown for the right-hand side of the schematic of FIG. 19.

FIG. 21 shows an input transmitter using dual radius input transmitter pinions. FIG. 22 shows a schematic embodiment functionally similar to that suggested by FIG. 20, but using instead an output transmitter. FIG. 23 shows a schematic representation of a transmission allowing placement of the input transmitter in the axial middle of the geartrain. FIG. 24 shows a schematic representation of a transmission similar to that shown in FIG. 23, except now the input transmitter is used to insert a reaction element through a power transmitting path.

FIG. 25 shows a schematic representation of a transmission similar to that shown in FIG. 16, allowing placement of the input transmitter in the axial middle of the geartrain.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Reference should be made to US Patent No. 5,342,258, from which this disclosure originates. FIG. 1 shows an upper half- plane schematic representation of one possible gearing arrangement using this invention. The arrangement shown includes an input transmitter having a input sprocket 198 drivingly connected to input shaft 142 and a second input sprocket 199 drivingly connected to input shaft 144. Input shafts 142 and 144 are generally coaxial with the axis of the transmission or geartrain, and they are situated so as to topologically or coaxially surround reaction stator 128, which is grounded to case 100 in a manner not interfering with operation of input sprockets 198 and 199. Particular construction of this and other input transmitters will be given in further detail below. In this disclosure, angular deviations about the transmission axis shall be termed circumferential; movements or indication along the length of the transmission axis or any parallel line will be termed axial; and any relations involving perpendicular distances from the transmission axis will be termed radial. Input shafts 142, 128, and 144 are presented and made drivingly available to clutch and gearing devices shown to the right of the input transmitter. Well known planetary gearsets are used throughout this disclosure, but as mentioned any three-element gearing devices may be substituted in their stead. Gearing devices shown include a first planetary gearset having a first sun gear 162 and a first ring gear 172 each meshingly engaged with a plurality of planet pinions 382, which are rotatably supported by planetary carrier 210. Carrier 210 is drivingly coupled, by means of second sun gear hub 163, to second sun gear 164 of a second planetary gearset, which further comprises a second ring gear 174 and a plurality of planet pinions 384 which are rotatably mounted on planetary carrier 212 and which mesh with second sun gear 164 and second ring gear 174. With carrier 210 drivingly connected to output shaft 410, three gearing elements, namely, second ring gear 174, first ring gear 172 and first sun gear 162, are available for modulation of their angular speeds by the input transmitter via three power transmitting paths. Starting with the third power transmitting path driving first sun gear 162, first sun gear hub 161 drivingly connects first sun gear 162 to C clutch housing 134 which may be selectively coupled to input shaft 142 by means of C clutch 114. Alternatively, when this power transmitting path is not driven by actuation of C clutch 114, reactive force needed to prevent first sun gear 162 from rotating backward during forward loading of the other power transmitting paths may be borne by C one-way clutch (OWC) or drive establishing device 44 whose outer race 295 is drivingly connected to C clutch housing 134 and whose inner race 294 is drivingly connected to reaction stator 128 as shown, thereby preventing backward motion of first sun gear 162.

Similarly, for the second power transmitting path driving first ring gear 172, first ring gear 172 is drivingly connected to drum 266 which may be selectively coupled to input shaft 144 by means of B clutch 112. Alternatively, when this power transmitting path is not driven by actuation of B clutch 112, reactive force needed to prevent backward rotation of first ring gear 172 may be supplied by B one-way clutch 42 whose outer race 293 is drivingly connected to drum 266 and whose inner race 292 is drivingly connected to reaction stator 128, as shown.

For the first power transmitting path driving second ring gear 174, shown symbolically in FIG. 3 as the path fed by clutch A, second ring gear 174 is drivingly connected to drum 366 which may be selectively coupled to input shaft 144 by means of A clutch 110. Alternatively, during non-actuation of A clutch 110, reactive force to prevent backward rotation of second ring gear 174 is provided by A one-way clutch 40 whose outer race 291 is drivingly connected to drum 366 and whose inner race 290 is drivingly connected to case 100 as shown.

Clutches 110, 112, and 114 operate by well known mechanical, electrical, magnetic, pneumatic, hydraulic, or electrorheological means; preferably they are hydraulically actuated multiple-disc wet clutches which are widely used and well known in the art. In this figure and elsewhere in this disclosure, braking or clutch devices may be substituted for all one-way clutches shown.

Using the equation of motion for a simple planetary gearset, ω_c = aω_β + bω_r (Eqn . 1)

such that b > a and a + b = 1, and where

ω_c = angular speed of planetary carrier ω = angular speed of sun gear ω_r = angular speed of ring (internal) gear a = carrier/sun angular ratio, with ring fixed; a < l b = carrier/ring angular ratio, with sun fixed; b < 1

we can derive the equation of motion for this transmission by equating ω„ of the first planetary gearset with ϋ, of the second planetary gearset. We can thus obtain the angular speed of the output shaft, ω_QUt, as a function of the angular speed of the power transmitting paths selectively fed by the A clutch, B clutch, and C clutch, denoted by ω_A, ω_β, and ω_c, respectively:

^ωout ^{= b}2^ωA ^{+ a}2^bι^ωB ^{+ a}ι^a2^ωc ⁽*^Bt?ⁿ- ²⁾

where the new subscripts 1 and 2 refer to the first and second planetary gearsets, respectively. As can be seen from the above equation of motion for this geartrain, the three power transmitting paths A, B, and C — that is, the power transmitting paths fed by selective actuation of A clutch, B clutch, and C clutch, respectively — may be driven independently so long as the paths not driven are grounded. One can therefore execute the clutching of the power transmitting paths to the transmission input in a manner which is combinational and independent — A clutch, B clutch, and C clutch may be actuated in any combination, singly (A, B, C) , in pairs (AB, BC, AC) , or all three at a time (ABC) , with all combinations resulting in useful ratio states. Moreover, the principle of superposition applies, so that one can simply add algebraically the contributions to output angular speed made by each power transmitting path. For convenience, one can name the ratio states according to the power transmitting paths that actively contribute to net output, so that the AC state, for example, is the state obtained by coupling of the prime mover to the two power transmitting paths fed by clutches 110 and 114. Seven forward drive ratio states may be obtained using this transmission by selective actuation of the three clutches 110, 112, and 114 along with the active engagement of one-way clutch (OWC) or drive-establishing devices 40, 42, and 44, when necessary. In neutral, the input shafts 142 and 144 rotate with the engine or prime mover, and may also drive a conventional hydraulic pump, not shown, for use by a control and lubrication system for the transmission, as known in the art. Since all three power transmitting paths A, B and C are operatively restrained from backward motion by one-way clutches 40, 42, and 44, respectively, the output shaft will also be similarly restrained and therefore no neutral rollback is permitted. All power transmitting clutches are released.

From neutral, a first or lowest forward ratio state is achieved in a single transition shift as the clutch control system gradually actuates C clutch 114, which causes C clutch housing 134 and first sun gear 162 to rotate in unison with input shaft 142. With a load on the output shaft 410, the remaining two free gearing elements, first ring gear 172 and second ring gear 174 are urged to rotate backward, or in a sense opposite that of the prime mover. Specifically, first ring gear 172 is urged to rotate backward because the load on the first planetary gearset is borne by first pinion carrier 210, and the forward torque imposed on first sun gear 162 creates a reaction torque in the reverse direction on first ring gear 172. First pinion carrier 210, in turn provides forward torque to the second sun gear 164 of the second planetary gearset since it is drivingly connected thereto via second sun gear hub 163. The forward torque applied to second sun gear 164 thus creates a reaction torque in the reverse direction on second ring gear 174. First ring gear 172, however, is restrained from backward rotation by B one-way clutch 42 which actively engages, causing reaction stator 128 to ground or hold stationary drum 266, which is drivingly connected to first ring gear 172. Similarly, second ring gear 174 is restrained from backward rotation via drum 366 which is arrested by active engagement of A one-way clutch 40 which couples drum 366 to case 100. Thus, when the prime mover is driving the load at output shaft 410, first ring gear 172 and second ring gear 174 are held stationary, and first sun gear 162 driven via C clutch 114 is the sole power transmitting path to contribute to output angular speed at output shaft 410. Power applied to first sun gear 162 causes first pinion carrier 210 to rotate forward at a reduced angular speed. Because first pinion carrier 210 is coupled to second sun gear 164 via second sun gear hub 163, a second reduction in angular speed is achieved in the second planetary gearset, causing output shaft 410 to rotate at a further reduced angular speed. Because the gear reductions provided by the two planetary gearsets are compounded, the overall transmission ratio for this lowest ratio state may be made low enough to eliminate the need for final ratio reduction gearing in automotive applications, if desired. During a coasting condition where the load at output shaft 410 drives the prime mover, the non- engagement of one-way clutches 42 and 44 causes first ring gear 172 and second ring gear 174 to spin freely in the forward direction, resulting in a decoupling of the prime mover from the output shaft 410. Specifically, with output shaft 410 under forward rotation in excess of forward rotation that would be provided by driving the power transmitting path fed by C clutch 114, second ring gear 174 is no longer urged to rotate backward, but instead is driven forward at reduced speed. This forward rotation of second ring gear 174 is now unconstrained because drum 366 to which it is coupled is allowed to turn freely, since A clutch 110 remains in a released state, and A one-way clutch 40 is inactive or disengaged during forward rotation. With no substantial reaction torque applied to second ring gear 174, no driving torque is applied to second sun gear 164. Since second sun gear 164 is the only link to the first planetary gearset, no coupling occurs to drive the prime mover, and thus no engine braking is possible while in the first ratio state. When selected by the transmission control system, a second ratio state is available in another single transition shift through simultaneous release of C clutch 114 and application of B clutch 112. This causes drum 266 and first ring gear 172 to rotate in unison with input shaft 144. In a manner similar to that of the first ratio state, first sun gear 162 and second ring gear 174 are now urged to rotate backward when a load is applied to output shaft 410. First sun gear 162 is restrained from backward rotation by C one-way clutch 44, which causes stationary reaction stator 128 to couple to and ground C clutch housing 134, which is drivingly coupled to first sun gear hub 161 and first sun gear 162. Second ring gear 174 is restrained as above by A one-way clutch 40 via drum 366. Thus when the prime mover is driving the load at output shaft 410, first sun gear 162 and second ring gear 174 are held stationary, and first ring gear 172 driven via B clutch 112 is the only power transmitting path contributing to output angular speed at output shaft 410. Power applied to first ring gear 172 causes first pinion carrier 210 to rotate at moderately reduced angular speed. Again, because first pinion carrier 210 is coupled to second sun gear 164 via second sun gear hub 163, a second reduction in angular speed is achieved in the second planetary gearset, causing output shaft 410 to rotate at further reduced angular speed. During a coasting condition where the load rotates faster than the normal driven speed for this ratio, one-way-clutches 40 and 44 allow second ring gear 174 and first sun gear 162 to spin freely in the forward direction, decoupling the prime mover from the output shaft 410 in a manner similar to the first ratio state above. As before, drum 366 is no longer restrained by A one-way clutch 40, and with A clutch 110 remaining disengaged, it spins freely, along with second ring gear 174. With no substantial reaction torque applied to second ring gear 174, no driving torque is applied to second sun gear 164, and the decoupling occurs as before.

To effect a change to a third ratio state, C clutch 114 is applied while B clutch 112 remains engaged. Power flows from the first and second speeds are thus combined, whereby first sun gear 162 is coupled to input shaft 142 and first ring gear 172 is coupled to input shaft 144. Since in this case input shafts 142 and 144 are both driven at the same transmission input speed, the first planetary gearset locks up as a unit and drives first pinion carrier 210 at the same transmission input angular speed. As before, second ring gear 174 is urged to rotate backward when a load is applied to output shaft 410, but backward rotation is arrested by active engagement of A one-way clutch 40. With first pinion carrier 210 rotating at full transmission input angular speed, only the single gear reduction by means of the coupling of first pinion carrier 210 to second sun gear 164 into the second planetary gearset occurs. As before, A one-way clutch 40 still overruns and decouples the transmission when the output shaft 410 rotates faster than its operative driven speed.

A fourth gear ratio is achieved through a multiple transition shift whereby B clutch 112 and C clutch 114 are released, and A clutch 110 is simultaneously applied. This results in coupling of drum 366 to input shaft 144, driving second ring gear 174 at transmission input angular speed. With a load at output shaft 410, one-way clutches 42 and 44 hold, operatively grounding first ring gear 172 and first sun gear 162, respectively. With first sun gear 162 and first ring gear 172 grounded, the first planetary gearset locks up as a unit, so that first pinion carrier 210 and second sun gear 164, through coupling, are held stationary. With second sun gear 164 making no contribution to output angular speed, A clutch 110 makes the only contribution, a moderately reduced output angular speed. During a coasting condition, one-way clutches 42 and 44 again overrun, and no engine braking occurs.

For successively higher speeds of the output shaft 410, fifth, sixth, and seventh speeds are achieved in a combinational fashion through additional single transition shifts. The four-to-five shift is achieved by additional application of C clutch 114. The five-to-six shift is effected by simultaneous release of C clutch 114 and application of B clutch 112. The final six-to-seventh shift is achieved by additional application of C clutch 114. These higher ratio states mimic the first three ratio states, except that the power transmitting path fed by A clutch 110 adds a large contribution to output angular speed and is no longer modulated by A one-way clutch 40. This leaves one fewer one-way clutch to overrun during coasting conditions, and during the seventh and highest ratio state, all three power transmitting paths are contributing to output angular speed, and all one-way clutches are inactive, so no decoupling of the transmission occurs during coasting. Also, the power flow is simplified in this ratio, since all free gearing elements, first sun gear 162, first ring gear 172, and second ring gear 174 are driven at the same input angular speed via input sprockets 198 and 199. As a result, both planetary gearsets lock up, and the transmission in effect becomes a simple direct drive coupling, with output shaft 410 is driven at input angular speed.

Although this simple embodiment lacks engine-braking or reversing provisions, this transmission can provide superposition gearing with excellent ratio modulation and incrementation characteristics, and most notably can be used as a "building block" to develop other transmissions, some more suitable for automotive use, as will be discussed below. The twin sprocket arrangement shown having input sprockets 198 and 199 is not the only way to achieve the function of an input transmitter. Other structures are shown below, with detailed cross sections.

FIGS. 7(a) - 7(r) show a set of possible schematic gearing configurations, that can be substituted for the right-hand gearing portion of the schematic representation of FIG. 1. Each gearing configuration represents a rearrangement of the single linkage or coupling between one selected element from each three-element planetary gearset, and an output coupling selected on one of the remaining available gearing elements of the second planetary gearset. In the above embodiment shown in FIG. 1, all seven possible ratio states are forward speeds with transmission ratios ascending sequentially to a maximum transmission ratio of 1:1. It is possible, however, to substitute an alternative gearing configuration from FIGS. 2(a) - 2(r) to obtain a transmission with different kinematic properties. Some gearing configurations, such as shown in FIGS. 2(g) and 2(n) offer higher top ratio states, with the ratio of ω_out/ω_in greater than one. They also develop reversing motion internally, whereby one or more reverse ratio states are achieved when the planetary carrier of one planetary gearset is held stationary and power is applied to either its sun or ring gear, as is well known. This causes one or more power transmitting paths to give a negative or reversing contribution to the motion at the output shaft 410, which may result in a net forward or reverse motion, depending on the other power transmitting paths energized.

Using these and other singly coupled planetary gearsets in transmissions of this type allows for maximum kinematic degrees of freedom and produces a maximum number of ratio states, but if desired, one can substitute any of a number of doubly coupled gearsets that are usually employed in planetary geartrain practice. However, this greatly reduces the number of possible ratio states as a function of the number or complexity of the gearsets used. Generally one also may substitute any type of gearset in place of the planetary gearsets shown so long as there are three substantially coaxial free gearing elements or the equivalent available for rotational coupling to the clutch devices and input transmitter shown in FIG. 1. In addition, for this embodiment it is also possible to drive input sprockets 198 and 199 at different speeds, which would alter the mix of gearing ratios provided. Since the ratio patterns generated derive from a combinational summation process, many varied useful ratio patterns may be obtained by varying the number of teeth or the effective radius for each gearing element. FIG. 3 gives a schematic representation similar to that of FIG. 1, but with additional clutches that add forward and reverse ratio states and a capability for engine braking. Power transmitting paths fed by clutches A, B, and C may now be positively grounded when necessary to permit engine braking where the output load drives the prime mover. The transmission thus obtained is then suitable for automotive use, although embodiments discussed below have other added advantages, input sprockets 198 and 199 are again drivingly connected to input shafts 142 and 144, respectively, which coaxially surround reaction stator 128, for presentation to clutch devices to the right. The second sun gear 164 may now be selectively coupled to first pinion carrier 210, by means of forward clutch 350, or to first ring gear 172 by means of reversing (R) clutch 355. To achieve this, second sun gear 164 is drivingly connected, via second sun gear hub 163, to sleeve shaft 221 which drivingly mates with both clutch inner hub 37 and clutch inner hub 38. Clutch inner hub 37 may be selectively coupled using forward clutch 350 to forward clutch housing 250 which is permanently coupled to or forms an integral part of first pinion carrier 210. Similarly, clutch inner hub 38 may be selectively coupled using reversing clutch 355 to a clutch housing (not shown) which is coupled to or forms an integral part of drum 266, to which first ring gear 172 is drivingly coupled. In addition to being drivingly coupled to forward clutch housing 250, first pinion carrier 210 is also drivingly coupled to inner shaft 141. Inner shaft 141 is drivingly engaged with clutch inner hub 30 and may be grounded using carrier brake 356. With this embodiment C clutch housing 134 is also drivingly connected to, or is an integral part of, clutch inner hub 35, which by means of C brake clutch 124, allows C clutch housing 134 to be positively grounded when desired. Actuation of C brake clutch 124 causes clutch inner hub 35 to be coupled to C brake clutch housing 234 which is splined or drivingly connected to reaction stator 128. Also, outer race 293 is now drivingly connected to clutch inner hub 34. This allows B one- way direct clutch 202 to selectively decouple B one-way clutch 42 from drum 266 and first ring gear 172 to allow intentional reverse motion of first ring gear 172 for use by reversing clutch 355 in what can be called "R" and "RA" ratio states. B one-way direct clutch 202 employs clutch inner hub 34 and has an outer housing which is drivingly connected to or forms an integral part of drum 266. Also, drum 266 also comprises or drives a clutch housing for B brake clutch 122 whose inner hub 33 is also drivingly connected to reaction stator 128. Actuating B brake clutch 122 allows positive grounding of first ring gear 172 when desired. Also, drum 366 may be arrested by A brake clutch or band 120, allowing for positive grounding of second ring gear 174 when desired.

A total of nine ratio states may be obtained by selective or active actuation of the clutch devices in this gearing arrangement. In neutral, the input shafts 142 and 144 again are driven by the prime mover, which may also drive a conventional hydraulic pump, not shown, for lubrication and clutch control. All selectively operable clutches are released. Since both the forward and reverse clutches 350 and 355, respectively, are not applied, both second sun gear 164 and second pinion carrier 212 may spin freely. Optionally, however, from a clutch control standpoint any two of the three clutches 202, 350 or 114 may be applied without driving the transmission output.

From neutral, a first forward or C ratio state results as the clutch control system gradually applies C clutch 114, B one-way direct clutch 202 and forward clutch 350. C clutch 114 couples input shaft 142 to first sun gear 162 while forward clutch 350 couples the first pinion carrier 210 to second sun gear 164. With a load driven through first pinion carrier 210, first ring gear 172 is urged to rotate backward, but is restrained by automatic active engagement of B one-way clutch 42. This allows first pinion carrier 210 to drive second sun gear 164 at a reduced rotational speed. A similar reaction force imposed on second ring gear 174 by active engagement of A one-way clutch 40 will allow driving a load forward at second pinion carrier 212 at a further reduced rotational speed. During coasting where the load rotates faster than driven by the geartrain, second pinion carrier 212 drives second ring gear 174 forward and A one-way clutch 40 disengages. With second ring gear 174 freewheeling, no reverse coupling of the load to the engine occurs. The gear reductions provided by this combinational incrementally variable transmission (CIVT) are compounded, eliminating the need for final ratio reduction gearing for automotive applications, if desired.

From the first speed, a second or B ratio state is available through a single transition shift by simultaneous release of C clutch 114 and application of B clutch 112. Both forward clutch 350 and B one-way direct clutch 202 remain applied, although B one¬ way direct clutch 202 does not participate in power transmission, and remains applied solely to simplify clutch control. With first ring gear 172 driven forward, C one-way clutch 44 actively engages to prevent reverse rotation of first sun gear 162, thus allowing first pinion carrier 210 to rotate at reduced angular speed. With forward torque transmitted to second sun gear 164, again by coupling through forward clutch 350, second ring gear 174 is also urged backward, but prevented through the reactive force provided by active engagement of A one-way clutch 40. Second pinion carrier 212 then is driven at a moderate reduced angular speed. During a coasting condition, second ring gear 174 again overruns, decoupling the load from the engine. The third or BC ratio state is achieved by gradual reapplication of C clutch 114, allowing both the "B" and "C" power transmitting paths to contribute to the output motion, so that the sum of the inverse overall ratios or fractional drive contributions for the "C" and "B" states taken individually equal algebraically the inverse overall ratio for the "BC" ratio state. With second sun gear 164 driven at transmission input speed, second ring gear 174 is again restrained by active engagement of A one-way clutch 40. Again, no engine braking occurs during coasting since forward driving of second ring gear 174 goes unchecked. To access the fourth or RA ratio state, a multiple transition shift occurs whereby clutches 112, 202, and 350 are released while reversing clutch 355, carrier brake 356 and A clutch 110 are applied. With C clutch 114 applied, first sun gear 162 is coupled to input shaft 142 while first pinion carrier 210 is held stationary by carrier brake 356, giving rise to reversing motion at first ring gear 172. This reverse motion is no longer prevented by engagement of B one-way clutch 42 because B one-way direct clutch 202 is disengaged. The reverse motion of first ring gear 172 is communicated to second sun gear 164 via drive shell 268 and by application of reversing clutch 355. This "R" or reversing power transmitting path contributes a reverse angular speed component to the net transmission output speed. In addition to this "R" reversing power transmitting path, application of A clutch 110 drives drum 366 and second ring gear 174 at the speed of input shaft 144, contributing a forward speed component at the transmission output. With judicious choice of tooth numbers for the planetary gearsets, the "R" and "A" contributions add up to a net forward speed at the transmission output. During coasting, the load may drive the engine, as there are no free gearing elements to freewheel or decouple the engine from the load.

To shift into the fifth or A ratio state, reversing clutch 355, C clutch 114 and carrier brake 356 are released while B one¬ way direct clutch 202 and forward clutch 350 are again applied. A clutch 110 drives drum 366 and second ring gear 174, and under an output load second sun gear 164 will be urged to rotate backward. By actuation of forward clutch 350, second sun gear 164 is coupled to first pinion carrier 210, which will not rotate backward because first ring gear 172 and first sun gear 162 are restrained by active engagement of one-way clutches 42 and 44, respectively.

From fifth, a sixth AC ratio state is available through a single transition shift whereby C clutch 114 is engaged. The power flow is similar to that for the fifth speed, with an additional contribution made by engagement of C clutch 114, giving an additional forward contribution to the output angular speed. The B power transmitting path is still restrained from backward motion by B one-way clutch 42.

From sixth, a seventh AB ratio state is achieved by further simultaneous application of B clutch 112 and release of C clutch 114. The C power transmitting path is restrained by C one-way clutch 44. Engagement of B one-way direct clutch 202 is again optional, only to simplify clutching, since the B power transmitting path is now driven. From seventh, the eighth or top ratio state occurs when C clutch 114 engages, allowing all forward power transmitting paths to contribute to output motion. The transmission then acts as a direct coupling, with the drive fraction and overall ratio equal to unity. With all power transmitting paths driven, no engagement of one-way clutches occurs, and the engine is always directly connected to the load, permitting engine braking. During the fifth, six and seventh speeds, however, load-engine decoupling occurs, since one-way clutches 42 and/or 44 will disengage, permitting freewheeling of a principal gearing element.

Safety considerations require that manually selectable engine braking states be available. A "manual third" BC ratio state which allows engine braking is achieved through the same clutching given above for the BC state, but with A brake clutch 120 also applied to prevent overrunning of the A clutch housing 130 during coasting. Similarly a "manual second" engine braking B ratio state is achieved through B ratio clutching plus additional application of both A brake clutch 120 and C brake clutch 124. A "manual first" C ratio state giving engine braking is attained with C ratio state clutching plus application of both A brake clutch 120 and B brake clutch 122.

Finally, a pure reverse or R ratio state is available by actuating C clutch 114, carrier brake 356, reversing clutch 355, and A brake clutch 120. As before in the RA ratio state, actuation of C clutch 114 with first pinion carrier 210 held stationary by carrier brake 356 develops reversing motion at first ring gear 172 which is coupled to second sun gear 164 by reversing clutch 355. Now, however, the A power transmitting path is not energized, but is held stationary by A brake clutch 120, which prevents forward spinning of second ring gear 174 when driving an output load. The structure of FIG. 3 is not unique and many variations in layout of clutches, shafts, etc., can be made without departing from the general arrangement given. This embodiment allows a maximum number of significant ratio states (nine, total) using only a minimum number (two) of planetary gearsets or equivalent gearing hardware. Some embodiments presented below, such as that shown in FIG. 16, using more sophisticated input transmitters have added advantages, including having fewer required clutch devices and using simpler transitional shifting. FIGS. 4 and 5 show full plane schematic cross-sectional views of the split/twin input transmitters of the type shown in FIGS. 5 and 9, with dual input drive links 312 and 314 meshing with input sprockets 198 and 199, respectively. These figures show two ways to have the reaction stator emerge from the input transmitter for mechanical connection to the transmission case or other reference frame. In the lower half-plane, input drive links 312 and 314 are shown in cross-section. In FIG. 4, the reaction stator mechanically bonds to transmission case 100 by passing radially between the input sprockets 198 and 199. Coupling to transmission case 100 is shown occurring in a circumferential location where the drive links 312 and 314 engagingly surround input sprockets 198 and 199. If desired, however, the reaction stator may emerge as shown in FIG. 5, bonding to case 100 in a location between the slack and tensioned sides of input drive links 312 and 314. It also allows a single input drive link to replace the twin input drive links 312 and 314 in a manner similar to that shown in FIG. 13, which uses a single wide input drive link 315. Generally, one can also eliminate drive links altogether and use offset or countershaft gearing to feed input shafts 142 and 144 and accomplish the coaxial layering of shafts needed.

FIG. 6 shows a schematic cross-sectional view of an alternate construction, an epicyclic input transmitter, which requires only a single input sprocket 198 and input drive link 312 of normal width. Input sprocket 198 is splined or coupled to input shaft 144. To allow placement of reaction stator 128 and also to drive input shaft 142 without requiring an additional input sprocket, an epicyclic power bridge is used. To form the bridge, input shaft 142 is drivingly connected to or is integral with input transmitter first sun gear 82, while input shaft 144 is drivingly connected to or is integral with input transmitter second sun gear 84. One or more input transmitter pinions 288, which are rotatably mounted on a carrier fixed in case 100 or other suitable reference frame, meshingly engage with both input transmitter first sun gear 82 and input transmitter second sun gear 84. This arrangement allows that both input transmitter sun gears 82 and 84 turn in synchrony. This holds true even if reaction stator 128 is allowed to rotate. The reaction stator 128 exits or passes through the epicyclic input transmitter in a circumferential location between two or more of the input transmitter pinions 288, bonding to case 100 as shown. If only one input transmitter pinion 288 is employed, the reaction stator 128 simply exits in a circumferential location not conflicting with the input transmitter pinion 288. Also, the epicyclic transmitter may be driven from the left side instead of the right side as shown, by having input sprocket 198 or other driving means coupled to input shaft 142.

When using input drive links, this epicyclic input transmitter can give better mechanical efficiency than the twin input transmitters of FIGS. 10 and 11, since fewer meshing and other energy losses are incurred when using a single input drive link of normal width rather than using similarly rated double-width or dual input drive links. Other important advantages include lower cost, and less space required for the additional driven sprocket or sheave. More importantly, this epicyclic construction may be used as a core for more complex input transmitter assemblies disclosed below that can provide additional forward and reverse motions or shaft speeds for use by a drivetrain. For example, a reverse input shaft can be presented and made drivingly available to clutch and gearing devices without requiring availability of an additional braked or grounded planetary carrier inside the powertrain as is customary to develop reverse motion in presently used transmissions. This saves the added cost and hardware needed to brake a planetary carrier under load, which typically requires a torque that is triple that of the prime mover.

For a given gearing arrangement, many input transmitter structures or configurations can be utilized. One can usually devise an input transmitter that will accommodate the requirements of the gearing and clutch devices, such as: the number and type of forward and/or reversing inputs required; the input drive means, whether by use of an end shaft, a sprocket input that uses drive links or belts, offset gearing, or any hybrid combination thereof; and the specific topologic coaxial arrangement of shafts needed to access the gearing elements or structures to be driven. FIGS. 7, 8, 9, 10, 11, and 12 show schematic upper and lower half-plane views of some slightly more complex input transmitters than that shown in FIG. 6. FIG. 7 shows an input transmitter similar to that shown in FIG. 6, except now a main input shaft 140 is integral with the input shaft 142 previously shown. In addition, an input transmitter ring gear 184 now meshes with input transmitter pinions 288, axially to the right of where reaction stator 128. Input transmitter ring gear 184 is shown integral with an input transmitter ring gear hub 183, which in turn is coupled or splined to input shaft 147. With forward driving of main input shaft 140, input transmitter ring gear 184 will generate reverse motion which can be "transmitted" to a subsequent geartrain via input shaft 147. FIG. 8 shows a similar arrangement, but now a input drive link 312 drives a input sprocket 198 which is coupled to input transmitter first sun gear 82 and input shaft 142. Another reaction stator 129 is also added, bonding to case 100 at the axial left end of the input transmitter. This provides an innermost stationary structure for braking or restraining structures in the gearing or clutch device structure. As mentioned before, reaction elements like reaction stator 129 may be driven or allowed to rotate, becoming reaction "rotors." FIG. 9 shows an input transmitter similar to that of FIG. 8, except that the input sprocket 198 and the input transmitter ring gear have "switched sides." Specifically, input transmitter ring gear 184 has been replaced by an input transmitter ring gear 182 which meshes with input transmitter pinions 288 to the axial left of reaction stator 128. Input transmitter first ring gear hub 181 couples ring gear 182 to an input shaft 141. Input drive link 312 driving input sprocket 198 is now coupled to input transmitter second sun gear 84 and input shaft 144.

Even in the description just given, many variations may be made. In FIGS. 7, 8, and 9, the input drive was coupled to either input shaft 142 or 144. Instead, however, the input drive in FIGS. 35 and 36 may be coupled to the input shaft 147, and the input drive of FIG. 9 coupled to inner shaft 141. This interchanges the roles of the "reversing" and "forward" shafts, e.g., input shafts 142 and 144 would then exhibit reverse motion.

FIG. 10 shows an input transmitter having one reaction stator 128 surrounded coaxially by two rotating shafts on both its outside and inside surfaces. A plurality of input transmitter pinions 288, (again rotatably mounted on a carrier fixed in case 100 or other suitable reference frame) gearingly mesh with first sun gear 82 and first ring gear 182 to the axial left of reaction stator 128 and mesh also with input transmitter second sun gear 84 and second ring gear 184 to the axial right of reaction stator 128. In this arrangement, main input shaft 140 is coupled to both input transmitter first ring gear 182 via hub 181 and to inner shaft 141. First sun gear 82 is coupled to input shaft 142 and second sun gear 84 is coupled to input shaft 144. Second ring gear 184 is coupled to input shaft 147 via second ring gear hub 183. The coaxial grouping of shafts, in order of ascending radii, 141, 142, 128, 144, and 147, may be used with gearing and clutch device arrangements where access to reversing inputs (142 and 144) is required in a location radially adjacent to the reaction stator 128. FIG. 11 shows a similar construction where instead a input drive link 312 drives a input sprocket 198 which is coupled to both first ring gear 182 and inner shaft 141. This allows an added second reaction stator 129 to become the innermost shaft, with the left axial end again grounded in case 100. FIG. 12 shows the input sprocket 198 coupled instead to both second ring gear 184 and input shaft 147.

The gearing arrangements made possible by this invention include transmissions having three or more component gearsets, not including the input transmitter structure. This greatly increases the number of available combinational ratio states with a minimum of additional gearing and clutch devices. Many arrangements may be devised to provide an input transmitter appropriate for a particular complex gearing arrangement so that each power transmitting path may be selectively coupled to the desired rotational inputs. Any means may be used to provide the necessary coaxial layering of shafts or rotating structures. For very high multispeed requirements, such as for transmissions having 20 or more speeds, input transmitters such as shown in FIGS. 13 and 14 may be used. FIG. 13 shows an input transmitter, in the upper and lower half-planes, using a single wide input drive link 315 which drivingly meshes with input sprockets 197, 198, and 199. First reaction stator 128 emerges at an axial location between input sprockets 197 and 198 while second reaction stator 129 emerges between input sprockets 198 and 199, both bonding at least mechanically to case 100 at a location between the slack and tensioned sides of input drive link 315. Individual drive links may be used for any of the drive sprockets. With different sizing for input sprockets 197, 198, and 199, different angular speeds will be "transmitted" to the subsequent geartrain along the input shafts 142, 144, and 146 that can enhance the transmission ratio pattern or clutching control strategy.

FIG. 14 shows another input transmitter having a plurality of input transmitter pinions 288 that drivingly mesh with sun gears 82, 84, and 86, which are in turn coupled to input shafts 142, 144, and 146, respectively. Additionally, the input transmitter pinions 288 drivingly mesh with first and second input transmitter ring gears 182 and 184, respectively. Coaxially between input shafts 142 and 144 is a first reaction stator 128 which passes radially outward to bond with case 100 at an axial location between sun gears 82 and 84, and just to the axial right of input transmitter ring gear 182. Coaxially between input shafts 144 and 146 is a second reaction stator 129 which also passes radially outward to bond with case 100 at an axial location between sun gears 84 and 86, and just to the axial left of input transmitter ring gear 184. This would allow an input means to drive either input transmitter ring gears 182 or 184, providing reverse motion at sun gears 82, 84, and 86. However, one may omit one or more of these gearing elements. If, for example, one omits input transmitter ring gear 182 and its associated input shaft 141, a main input shaft 140 may then be fitted to or integral with input transmitter first sun gear 82 or input shaft 142. This would then give reversing motion at the remaining input transmitter ring gear 184. A variation on this input transmitter shown in FIG. 15 shows an a plurality of pinions 288 having support spindles grounded in case 100, gearingly mesh with ring gears 182, 184, and 186 and a single sun gear 86. First and second reaction stators 128 and 129 are now centrally grounded, bonding to case 100 at a location radially inside the input transmitter pinions 288. Sun gear 86 is coupled to input shaft 146; third ring gear 186 is coupled to input shaft 147 via third ring gear hub 185; second ring gear 184 is coupled to an input shaft 149 via second ring gear hub 183; first ring gear 182 is coupled to inner shaft 141 via first ring gear hub 181. One possible driving means is shown where an input sprocket 198 is drivingly coupled to inner shaft 141, but the driving means may be coupled to any of the principal elements 86, 182, 184, or 186 at any axial location, even at the opposite axial end of the geartrain. As an example of the kind of gearing arrangements that can be realized with judicious choice of an input transmitter, FIG. 16 shows a schematic representation for a compact geartrain, that expands upon the geartrain shown in FIG. 1, and which may be used as a front drive transaxle having eight forward speeds and one reverse, and needing no additional final drive reduction gearing for automotive applications. The input transmitter shown in FIG. 16 possesses a plurality of input transmitter pinions 288 drivingly meshing with first and second ring gears 182 and 184 respectively, and with a sun gear 82. Input drive occurs through an input sprocket 198 which is coupled to, or is an integral part of, input transmitter second ring gear 184 and input shaft 147, the radially outermost input shaft. Input transmitter first ring gear 182 is coupled via first ring gear hub 181 to inner shaft 141. Coaxially surrounding input shaft 141 is input shaft 142, which is splined or coupled to input transmitter sun gear 82. Radially outside input shaft 142 is reaction stator 128, which passes radially outward to bond with case 100 at an axial location to the axial right of input transmitter ring gear 182 and sun gear 82, and in a circumferential location not interfering with input transmitter pinions 288. Coaxially surrounding the reaction stator 128 is the outermost input shaft 147. Three free gearing elements, first sun gear 162, first ring gear 172 and second ring gear 174, are coupled to C forward/reverse clutch housing 252, drum 266, and drum 366, respectively, creating three free power transmitting paths. The coaxially nested shafts, that is, forward driven coaxially nested shafts 141 and 147, and the reverse driven input shaft 142, along with the stationary reaction stator 128 give many options for driving the above power transmitting paths, resulting in as many as eleven possible ratio states using just the gearsets shown.

The first power transmitting path driving second ring gear 174 may be driven by A clutch 110 which may selectively couple drum 366 to input shaft 147 using clutch inner hub 31. Braking and reactive forces are provided by A brake clutch 120 and A one-way clutch 40, whose inner race 290 is coupled to case 100 and whose outer race 291 is coupled to drum 366. The second power transmitting path driving first ring gear 172 is driven by B clutch 112 which selectively couples B clutch housing 132, via clutch inner hub 32, to drum 266. B clutch housing 132 is coupled to input shaft 147. Braking force for this power transmitting path is provided by B brake clutch 122 which may selectively couple drum 266 to reaction stator 128 by way of clutch inner hub 33. To prevent backward motion of drum 266, B one-way clutch 42 has an inner race 292 coupled to drum 266, and an outer race 293 splined to clutch inner hub 33, which is in turn coupled to the reaction stator 128. The third power transmitting path driving first sun gear 162 is driven by C clutch 114 which selectively couples C forward/reverse clutch housing 252 with input shaft 147 by way of clutch inner hub 37. Alternatively, this power transmitting path may be reverse driven by actuation of reversing clutch 355, which selectively couples C forward/reverse clutch housing 252 to reversing input shaft 142 via clutch inner hub 36. Reactive force to prevent backward rotation of this power transmitting path is provided by C one-way clutch 44, whose inner race 294 is coupled to C forward/reverse clutch housing 252 and whose outer race 295 is coupled to clutch inner hub 35. This allows C one-way direct clutch 204 to couple or decouple the C one-way clutch 44 from C reaction clutch housing 254, which is itself coupled to the reaction stator 128. C reaction clutch housing 254 is also coupled to clutch inner hub 34, which by means of C brake clutch 124, selectively couples C forward/reverse clutch housing 252 to reaction stator 128. Reaction stator 128 may accommodate internal passages to carry fluid used for lubrication or for actuating clutches in the geartrain, or may house electrical or optical conductors used to convey information (e.g., motion sensing) from the geartrain to a transmission control system, which is not shown and whose construction and operation is well known in the art.

In neutral, with input sprocket 198 driven by the prime mover, input shafts 141 and 147 are driven forward, and input shaft 142 is driven in reverse. The prime mover may also drive a hydraulic pump, not shown, for lubrication and clutch control. All selectively operable clutches are released. Since the C one-way direct clutch 204 is released, first sun gear 162 may spin freely. Through the single linkage that couples first pinion carrier 210 to second sun gear 164, second pinion carrier 212 is also unrestrained. Optionally, C one-way direct clutch 204 may be applied without driving the transmission output.

From neutral, the first forward C ratio state is obtained by gradual application of C clutch 114. C clutch 114 couples inner shaft 141 to first sun gear 162. Under an output load, first ring gear 172 and second ring gear 174 are urged to rotate backward, but are restrained by active engagement of B one-way clutch 42 and A one-way clutch 40, respectively. To simplify clutch control, C one-way direct clutch 204 may also be applied, but does not participate in power transmission.

From first, a second B ratio state is obtained through a single transition shift by release of C clutch 114 and application of B clutch 112. First ring gear 172 is now driven by input shaft 147, while backward motion of second ring gear 174 and first sun gear 162 are prevented by active engagement of one-way clutches 40 and 44, respectively. C one-way direct clutch 204 remains applied and is now necessary for power transmission. The third BC ratio state is attained by gradual application of C clutch 114, so that both the "B" and "C" power transmitting paths contribute to the output motion. This drives first pinion carrier 210 and second sun gear 164 at transmission input speed, and second ring gear 174 is again restrained from backward rotation by active engagement of A one-way clutch 40. C one-way direct clutch 204 may remain applied but simply overruns as in the first ratio state. During the 3-4 shift, it must be released.

Shifting to the fourth or RA ratio state occurs by way of a double transition shift whereby B clutch 112 and C clutch 114 are released and reversing clutch 355 and A clutch 110 are applied. Application of A clutch 110 drives the second ring gear 174 forward, giving a forward contribution to the transmission output, while the "R" or reversing power transmitting path enabled by actuation of reversing clutch 355 gives a reverse output contribution. The "R" and "A" contributions add algebraicly to a net forward output speed and some power recirculates, with second ring gear 174 driving second sun gear 164 backward. Also, when second sun gear 164 is being driven backward by the action of A clutch 110 driving second ring gear 174 with an output load, first ring gear 172 will be urged backward as well, but will be prevented from doing so by active engagement of B one-way clutch 42. From fourth, a fifth or A ratio state is achieved through a single transition shift by simultaneous release of reversing clutch 355 and application of C one-way direct clutch 204. Under an output load with second ring gear 174 driven forward, first sun gear 162 and first ring gear 172 will be urged backward but will be restrained by active engagement of C one-way clutch 44 and B one¬ way clutch 42, respectively. From fifth, a sixth or AC ratio state occurs through additional application of C clutch 114. C one-way direct clutch 204 may remain applied but does not participate in power transmission. Reverse rotation of first ring gear 172 is still prevented by action of B one-way clutch 42. Shifting to a seventh AB ratio state occurs through a single transition shift where C clutch 114 is released simultaneously with application of B clutch 112. Reverse rotation of first sun gear 162 is prevented by action of C one-way clutch 44. C one-way direct clutch 204 remains applied and is now needed for power transmission.

From seventh, the eighth or top forward ratio state is achieved when C clutch 114 engages, so that all forward power transmitting paths "A," "B," and "C" contribute to output motion. The transmission again acts as a direct coupling, providing a deep effective overdrive. C one-way direct clutch 204 again may remain applied but does not participate in power transmission. During coasting conditions, no engine braking occurs for speeds l, 2, 3, 4, 5, 6, and 7 because at least one of the one-way clutches 40, 42, and 44 will overrun, allowing freewheeling of a principal gearing element and decoupling the prime mover from the output load. Engine braking during these ratio states may be obtained by actuation of one or more of brake clutches 120, 122, and 124, as needed. The "manual third" engine braking DC ratio state is thus obtained through BC ratio clutching with A brake clutch 120 additionally applied to prevent overrunning of the A one-way clutch outer race 291 and A clutch housing 130 during coasting. The "manual second" engine braking £ ratio state is achieved through B ratio clutching plus additional application of both A brake clutch 120 and C brake clutch 124, and the "manual first" C ratio state is attained with C ratio state clutching plus application of both A brake clutch 120 and B brake clutch 122.

The reverse or R ratio state is available by applying reversing clutch 355 along with A brake clutch 120 and B brake clutch 122. Applying reversing clutch 355 drives first sun gear 162 in reverse, which under an output load will urge first ring gear 172 and second ring gear 174 to rotate forward. Ring gears 172 and 174 are held stationary, however, by braking clutches 122 and 120, respectively. Introducing the reaction stator 128 inside the power transmitting paths represented by drums 266 and 366 has made it possible to have clutches 124, 204 and 44 brake or restrain the C forward/reverse clutch housing 252, which would not normally have access to the transmission case or housing for that purpose. Overall, close spacing of the effective drive ratios obtained from using this and other similar gearing arrangements yields many advantages as cited above.

FIG. 17 shows an embodiment almost identical to that of FIG. 16, except that now the prime mover drives the input transmitter first ring gear 182, instead of driving second ring gear 184 using an input sprocket and drive link. The input drive is by means of main input shaft 140, making the transmission particularly suitable for automotive rear wheel drive applications. This transmission can be used with an unconventional rear axle. Because final reduction gearing is not necessary, a rear axle reduction ratio of 1:1, or perhaps 1.25:1 may be used. Using such a low final reduction ratio allows a smaller rear axle housing since the large ring or internal gear usually used may be reduced in size. The smaller housing reduces manufacturing cost and unit weight for the axle, and increases ground clearance.

This invention may be practiced with any gearset(s) , including a single simple planetary gearset. FIG. 18 gives a schematic representation that yields three forward speeds and one reverse. Similar to FIG. 1 above, input shafts 142 and 144 are coupled to input sprockets 198 and 199, respectively. The pinion carrier 210 is coupled to forward clutch housing 250 on the axial right side and coupled to drum 266 on the axial left side. Sun gear 162 is coupled to drum 262 via sun gear hub 161 and ring gear 172 is coupled to drum 366. Output shaft 410 is coupled to clutch inner hubs 36 and 37, so that application of forward (F) clutch 350 allows that output shaft 410 is coupled to forward clutch housing 250 and pinion carrier 210 while application of reversing (R) clutch 355 couples output shaft 410 to drum 366 and ring gear 172. The first power transmitting path driving sun gear 162 may be acted upon by B clutch 112, B one-way clutch 42 and B brake clutch 122, while the second power transmitting path driving ring gear 172 may be acted upon by A clutch 110, A brake clutch 120, and A one-way clutch 40, whose outer race 291 may be selectively coupled to drum 366 by A one-way direct clutch 200. Non-application of A one-way direct clutch 200 allows reverse motion of drum 366 during the R ratio state. Axially between A clutch 110 and B brake clutch 122 is carrier brake 356, which couples drum 266 to the reaction stator 128. The first forward B ratio state is achieved by applying B clutch 112, forward clutch 350 and A one-way direct clutch 200. The second forward A ratio state occurs when A clutch 110 and forward clutch 350 are applied. The third forward AB ratio state occurs when both A clutch 110, B clutch 112 and forward clutch 350 are applied. The reverse R ratio state is attained by application of B clutch 112, reversing clutch 355 and carrier brake 356. Engine braking for the A and B ratio states again requires application of B brake clutch 122 and A brake clutch 120, respectively.

We can improve on this transmission with addition of an overdrive ratio state. This is achieved by adding a third power transmitting path to drive the pinion carrier 210 directly. Referring to FIG. 19, another schematic representation, carrier brake 356 has been replaced with C clutch 114 and C brake clutch 124. C brake clutch 124 still performs the braking function of carrier brake 356, but actuation of C clutch 114 drives pinion carrier 210 by coupling the carrier via drum 266 and clutch inner hub 33 to the input shaft 144. By also applying B brake clutch 122 and reversing clutch 355, an overdrive ratio state is achieved by having ring gear 172 coupled to output shaft 410. A second further high overdrive ratio state may be obtained by adding an OD clutch to allow coupling the sun gear 162 to the transmission output shaft 410 while the pinion carrier 210 is driven by actuation of C clutch 114. FIG. 20 shows a partial schematic representation for this transmission. The full schematic can be obtained by substituting the partial schematic of FIG. 20 as shown for the right-hand side of the schematic of FIG. 19. Reversing clutch 355 has been renamed reversing/overdrive clutch 352 and forward clutch 350 is still needed for forward speeds 1- 3. A high overdrive clutch 359 is added now to give a very deep overdrive. This clutch couples output shaft 410 to high overdrive clutch housing 259, which is in turn coupled to sun gear hub 161 using a sleeve shaft 226. By applying C clutch 114, A brake clutch 120 and high overdrive clutch 359, a second overdrive ratio state is obtained, where with ring gear 172 stationary, the pinion carrier 210 is driven and the sun gear 162 is coupled to the transmission output. This gives three underdrive and two overdrive ratios and one reverse ratio for a total of six ratio states, using only a single simple planetary gearset. In addition to changing the number of teeth selected for each gearing element, the ratio patterns obtained may also be altered if desired by introducing a greater variety of input shaft speeds. FIG. 21 shows an input transmitter similar to that used in the transmission of FIG. 16, where the input transmitter pinions 288 have been replaced by dual radius pinions 289 that each are wide at the left axial end, and narrow at the axial right end. First ring gear 182 and sun gear 82 drivingly mesh with the wide left axial side of dual radius pinions 289, while second ring gear 184 drivingly meshes with the narrow right axial end of the dual radius pinions, allowing first ring gear 182 and second ring gear 184 to turn at different angular speeds.

In providing one or more kinematic degrees of freedom to any independent element of a geartrain, often a rearrangement of clutch devices, shafts, etc., will allow replacing the input transmitter structure with an "output transmitter" which provides introduction of the reaction stator 128 into the transmission interior at the output end of the transmission. FIG. 22 shows a schematic embodiment functionally similar to that suggested in FIG. 20, but using instead an output transmitter. At the input end of the transmission is a conventional main input shaft 140, which by means of clutch inner hubs 32, 33, and 34, provides the driving members for clutches 110, 114, and 112, respectively. Drums 366 and 266 are again coupled to ring gear 172 and pinion carrier 210, respectively. A one-way clutch 40, A one-way direct clutch 200, and A brake clutch 120 are located as before in FIG. 20, but drum 262 has been replaced by a B clutch housing 132. Also, the B brake clutch 122, B one-way clutch 42 and C brake clutch 124 have been relocated to the axial right of the planetary gearset. The output transmitter at the axial right resembles the input transmitter given previously in FIG. 6, having a plurality of fixed axis output transmitter pinions 488 which drivingly mesh with output transmitter sun gears 482 and 484. Output transmitter sun gear 482 is coupled to sleeve shaft 222, while output transmitter sun gear 484 is coupled to both sleeve shaft 221 and to output shaft 410. Reaction stator 128 passes coaxially between sleeve shafts 221 and 222 and emerges between the pinions 488 to bond to case 100. Sun gear 162 is now coupled via sun gear hub 161 to sleeve shaft 226, which in turn is coupled to both B clutch housing 132 and to drum 362. Drum 362 is fitted with a clutch housing for high overdrive clutch 359, which via clutch inner hub 35 selectively couples sleeve shaft 221 to drum 362. Drum 362 also accommodates B brake clutch 122, which selectively couples reaction stator 128 to drum 362 via clutch inner hub 36. Backward motion of drum 362 may be arrested by action of B one-way clutch 42, whose outer race 293 is coupled to the drum 362 and whose inner race 292 is coupled to the reaction stator 128. Pinion carrier 210 is coupled on its axial right side to drum 364, which is fitted with a clutch housing for C brake clutch 124. Application of C brake clutch 124 couples the drum 364 to reaction stator 128 via clutch inner hub 37. Drum 364 also houses or is coupled to a clutch housing for forward clutch 350, which selectively couples the drum 364 to sleeve shaft 222 via clutch inner hub 38. Sleeve shaft 222 is also coupled to a clutch inner hub 39 so that application of reversing/overdrive clutch 352 can couple the sleeve shaft to drum 366, as before. The general result is that introducing the reaction stator 128 inside drums 362 and 364 allows placing B brake clutch 122, C brake clutch 124, and B one-way clutch 42 inside power transmitting paths, where access to the transmission case 100 is not normally allowed. It is not possible here to show the nearly infinite number of transmissions that can result by applying the teachings of this disclosure. One may vary the: [1] Number and Type of Gearsets Used This includes the simple planetary gearsets discussed above, along with complex planetary gearsets having long and short dual pinion sets such as the well known Ravigneaux gearset, or single axis differential gearsets, etc. The gearsets may be mixed in any manner, with the permanent linkages or couplings between gearsets in any configuration. Single, double or multiple linkages between gearsets may be used as desired. Traditional gearing or non- geared transmission devices may be used that contain belts, traction rollers, etc. [2] Type of Input Transmitter(s) This includes the various input transmitters shown above, as well as those shown in FIGS. 23 - 25 below. Many arrangements are possible that are not specifically described here. As described above, input drive sprockets or epicyclic gearsets may be used as part of the input transmitter(s) , or countershaft or offset gearing may be used as well, such as using countershaft gearing in place of the input transmitter pinions 288. As demonstrated in FIG. 22, output transmitters may also be used, either alone or with one or more input transmitters. Using both input and output transmitters can be useful in constructing a transmission with a high number of available ratio states while reducing the complexity of the input transmitter. [3] Reaction Element(s)

Any number or placement of reaction stators or elements may be used as part of the input or output transmitter structures, so long as at least one reaction element (128) is placed in the interior of a power transmitting structure or path. As mentioned, any reaction stator may be bonded to a rotating reference frame or allowed to rotate, becoming a reaction rotor. [4] Clutching to Input Transmitter Driven Elements Any arrangement may be used to couple selected input transmitter driven elements or power transmitting structures to available power transmitting paths. [5] Clutching to Internal Gearing Elements or Shafts

As required, selected internal gearing elements or shafts may be coupled to one another, to alter power transmitting paths already driven, or to establish new ones. [6] Clutching to Case or Reference Frame

This includes one-way clutches and brakes used to ground selected power transmitting paths and the various configurations used to obtain various engine braking options. [7] Ratio Development

This includes the number of teeth chosen for each gearing element as well as changes in the overall transmission ratio by using initial or final ratio reduction gearing to suit the application.

As shown above and in FIGS. 23 - 25, the input transmitter, clutch devices, and gearing need not be arranged axially from left to right as shown. One can apply the methods taught here to provide maximum degrees of kinematic freedom to a four element gearset, such as disclosed in US Pat. 4,864,892, Ando et. al., which employs a compound planetary gearset having first and second ring or internal gears, one long sun gear, and a plurality of meshing twin pinions interposed therebetween. The gearset resembles a simple planetary gearset with twin pinions, except that the inner set of the meshing twin pinions extends outward to allow engagement with an additional second ring gear. With one gearing element coupled to the transmission output, three gearing elements remain to be selectively driven. Using an appropriate input transmitter, we can provide each of the three remaining independent or unlinked gearing elements with at least one degree of freedom. This results in a minimum of 2³ - 1 or seven ratio states, instead of the maximum of five as disclosed in the patent. Using an input transmitter having additional driving elements such as an external reversing provision would yield even more ratio states. In a similar way, the compound planetary gearset used in the transmission of US Pat. 4,884,472 to Miura is also not fully utilized. Miura also uses a four element compound planetary gearset, having first and second sun gears and one ring or internal gear. The gearset also resembles a simple planetary gearset with twin pinions, except that the outer set of meshing twin pinions meshingly engages with an additional second sun gear. Again, with one gearing element coupled to an output, three power transmitting paths may be established, giving at least seven ratio states instead of five. Instead, Miura adds a subtransmission to obtain five forward speeds and one reverse.

Another example is provided by US Pat. 4,802,385 (Hiraiwa) which discloses a transmission using a five element compound planetary-type geartrain that uses dual non-meshing sets of pinions mounted on a single pinion carrier. The first set of pinions meshes with first sun and ring gears, while the second set of pinions meshes with second sun and ring gears. It is kinematically equivalent to two simple planetary gearsets having linked carriers. With the first ring gear coupled to the transmission output, this gearset has four kinematically independent or free gearing elements. Using an appropriate input transmitter, the teachings of this disclosure allows a minimum of 2* - 1 or fifteen forward and reverse ratio states, instead of the six as disclosed by Hiraiwa. And of course, a complex input transmitter could present to the gearing more than one degree of freedom for selected gearing elements, allowing for even more possible ratio states.

Similarly, the four planetary gearsets used in the previously mentioned US Pat. 4,683,776 to Klemen could be better utilized using the instant invention. Klemen uses double linkages between the first and second and between the third and fourth planetary gearsets, with a single linkage or permanently coupling between the second and third sets. The transmission output is coupled to the pinion carrier of the fourth planetary gearset. Using only single linkages between gearsets allows that five gearing elements are independent and may be driven as desired. This allows five power transmitting paths, with at least 2⁵ - 1 or thirty-one ratio states, with many more ratio states possible still with use of an appropriate input transmitter that allows two possible driving speeds for selected gearing elements in the geartrain.

FIG. 23 shows a schematic representation of a transmission allowing placement of the input transmitter in the axial middle of the geartrain, and using a different coaxial arrangement of shafts.

Many of the transmissions presented thus far may use instead the input transmitters in the way illustrated by example in FIGS. 23 -

25. Here the input transmitter is used to introduce a reaction element into the interior of a power transmitting path or structure by allowing that path or structure to be split into two separate elements. Also, for illustration, the first planetary gearset has been replaced by a well known simple 3 element twin pinion planetary gearset, having a twin pinion carrier 215 which rotationally supports and includes a plurality of circumferentially spaced twin pinion sets each having an inner twin pinion 381 and a radially offset outer twin pinion 389. The inner twin pinion gearingly meshes with both sun gear 162 and the outer twin pinion 389. The outer twin pinion 389 gearingly meshes with both the inner twin pinion 381 and the ring gear 172. Input sprocket 198 now drives only input shaft 142. The input transmitter is now placed axially between B clutch 112 and B one-way clutch 42. Reaction stator 128 now originates from case 100, passing coaxially rightward through the interior of input shaft 142, and radially outward at an axial location between input transmitter first and second sun gears 82 and 84, which are coupled to input shafts 142 and 144, respectively. If the reaction stator 128 and the input transmitter were absent, the input shaft 142 would continue axially rightward through the geartrain, and would include input shaft 144 as shown. Said differently, a single input shaft "142-144" would pass coaxially rightward through the geartrain, and could allow coupling to the power transmitting paths 366, 266, and 262 by way of clutches 110, 112, and 114. However, without the input transmitter and introduction of the reaction stator 128 into the transmission interior, it would not then be possible to selectively brake drum 262. Instead, this power transmitting structure has been split into two power transmitting elements, namely input shafts 142 and 144. Input transmitter pinions 288 provide a power bridge as before, assuring torque handling continuity of input shafts 142 and 144, while allowing reaction stator 128 to pass between them. Once radially outward of pinions 288, the reaction stator 128 bonds to a fixed input transmitter pinion carrier 218 to the axial left of pinions 288 and also passes axially rightward to B one-way clutch 42, B brake clutch 122 and C brake clutch 124. In this way, a reaction element 128 is placed inside a space that would have been inaccessible, bounded by input shaft "142-144" and power transmitting paths 366, 266, and 262. Operation of this transmission proceeds in a manner similar to that discussed above, with the three power transmitting path selectively fed by clutches 110, 112, and 114; and reactive or brake forces supplied to the power transmitting paths by coupling to the reaction stator 128. FIG. 24 shows a schematic representation of a transmission similar to that shown in FIG. 23, except now the input transmitter is used to insert a reaction element through the third power transmitting path formerly shown as drum 262, now labelled as C forward/reversing clutch housing 252. As shown, the input transmitter is located to the immediate axial left of the first and second planetary gearsets. Input sprocket 198 again drives input shaft 142, which now extends fully axially rightward so as to allow coupling to the power transmitting paths 252, 266, and 366 through clutches 110, 112, and 114. Again, the reaction stator 128 passes coaxially rightward through the interior of input shaft 142. The split power transmitting path now starts with the C forward/reversing clutch housing 252, which now houses C brake clutch 124 and C clutch 114, and in the absence of the input transmitter and reaction stator 128, would continue axially rightward to couple with first sun gear hub 161. Now, however, this third power transmitting path is split into two power transmitting elements, namely C forward/reversing clutch housing 252 and what is now shown as sleeve shaft 222, which are coupled to input transmitter first sun gear 82 and input transmitter second sun gear 84, respectively. As in FIG. 23, reaction stator 128 passes radially outward between first and second sun gears 82 and 84, passes radially outward of pinions 288, and bonds to input transmitter carrier 218 to the axial left, continuing axially leftward to provide selective restraining means to C forward/reversing clutch housing 252 and drum 266 via C brake clutch 124, and B brake clutch 122 and B one-way clutch 42, respectively.

In a similar way, a transmission similar to that shown in FIG. 16 can be realized by moving the input transmitter to the axial middle of the geartrain. The schematic as shown in FIG. 25 shows the third power transmitting path driving first sun gear 162 as a renamed drum 262. A one-way clutch 40 and A brake clutch 120 are located as in FIG. 16. Input sprocket 198 now drives shaft 141, which is drivingly coupled to clutch inner hubs 31 and 32 for use by A clutch 110 and B clutch 112, respectively. Shaft 141 is coupled to input transmitter first ring gear 182 via first ring gear hub 181, with the input transmitter now located to the axial right of B clutch 112. Using a plurality of input transmitter pinions 288, a power bridge is established, allowing continuation of torque transfer from inner shaft 141 to input shaft 147, which is coupled to an input transmitter second ring gear 184. Pinions 288 also drive an input transmitter second sun gear 84, which is coupled to input shaft 144. Reaction stator 128 now passes from case 100 axially rightward through the interior of inner shaft 141, passes radially outward of pinions 288, bonds to input transmitter carrier 218 to the axial left of pinions 288, and then passes radially outward between first and second ring gears 182 and 184 to pass to the axial right for use by B one-way clutch 42, B brake clutch 122, C brake clutch 124 and C one-way clutch 44. Input shaft 147 passes coaxially rightward from input transmitter second ring gear 184 through the interior of reaction stator 128 to couple to clutch inner hub 36 of C clutch 114. Input shaft 144 passes coaxially rightward through the interior of input shaft 147 from input transmitter second sun gear 84 to clutch inner hub 37 which is used by reversing clutch 355. Again, the introduction of reaction stator 128 has allowed clutches 122, 124, 204, and 44 in an interior space not proximate the transmission case. Alternate sources of power can be devised to drive one or more of the power transmitting paths in a CIVT, including other powertrains or assemblies, or even secondary movers such as electric motors. One could, for example, construct a "continuous shift" transmission where an electronically controlled electric motor drive allows ratio shifts with little or no dissipated energy at clutch mechanisms. This could be accomplished by coupling the electric motor armature to one or more of the power transmitting paths, or by using the electric motor to drive one or more reaction rotors. The reaction rotor(s) could be held stationary to provide transmission operation as given above, then driven by the auxiliary electric motor to increase or decrease the transmission output speed without changing the speed of the prime mover. When the transmission output speed reaches a speed synchronous with the targeted or desired speed ratio, a clutch application control system can initiate a ratio shift while de-energizing the auxiliary electric drive motor. This would allow for smooth ratio shifts with little or no energy dissipated at clutch lining surfaces.

Obviously, many modifications and variations of the present invention are possible in light of the above teaching. It is therefore to be understood, that within the scope of the appended claims, the invention may be practiced otherwise than as specifically described or suggested here.

Claims

I claim:

1. An input arrangement to introduce a reaction stator into an interior of a power transmission having a plurality of power transmitting paths, comprising: first and second substantially coaxial power transmitting structures each disposed to permit individual selective coupling to at least one of said power transmitting paths; a mechanical reference frame (100) ; a reaction stator (128) coupled to said mechanical reference frame at a first axial location and positioned radially between said substantially coaxial power transmitting structures so as to permit selective coupling of said reaction stator to at least one said power transmitting path at a second axial location; means for selectively coupling said reaction stator to said power transmitting path at said second axial location, said reaction stator further positioned so as to permit said means for selectively coupling to be substantially located within the interior of at least one of said power transmitting paths, thereby allowing said reaction stator to provide a reaction force to at least one said power transmitting path in said power transmission using said means for selectively coupling at a location other than proximate the mechanical reference frame.

2. The input arrangement of claim 1, wherein said first and second power transmitting structures are driven by first and second input means for delivering power (312, 314) to said power transmission, respectively.

3. The input arrangement of claim 1, wherein the mechanical reference frame to which said reaction stator is coupled may rotate, thereby making said reaction stator a reaction rotor.

4. The input arrangement of claim 1, wherein said first and second power transmitting structures are drivingly coupled to first and second gear means, respectively; and further comprising at least one pinion means (288) gearingly meshing with both said first and second gear means, said pinion means being rotatably mounted on a carrier (218) which is substantially fixed in said mechanical reference frame; said pinion means and said carrier each being positioned circumferentially, radially, and axially so as not to interfere with said reaction stator proximate to said first axial location, thereby establishing a power bridge between said first and second power transmitting structures while still allowing said reaction stator to pass by said pinion means toward said second axial location.

5. The input arrangement of claim 4, wherein said first power transmitting structure is driven by an input means for delivering power (312) to said power transmission.

6. The input arrangement of claim 4, wherein said first and second gear means are sun gears.

7. The input arrangement of claim 4, wherein said first and second gear means are ring gears.

8. The input arrangement of claim 4, wherein said first gear means is a sun gear and said second gear means is a ring gear.

9. The input arrangement of claim 4, wherein said first gear means is a ring gear and said second gear means is a sun gear.

10. An input arrangement to introduce a reaction stator into an interior of a power transmission having a plurality of power transmitting paths, comprising: first and second substantially coaxial power transmitting structures, drivingly coupled to first and second gear means, respectively; a mechanical reference frame (100) ; a reaction stator (128) coupled to said mechanical reference frame at a first axial location and positioned axially between said first and second substantially coaxial power transmitting structures and first and second gear means so as to permit selective coupling of said reaction stator to

43 ^'< at least one said power transmitting path at a second axial location, at least one pinion means (288) gearingly meshing with both said first and second gear means, said pinion means being rotatably mounted on a carrier (218) which is substantially fixed in said mechanical reference frame; said pinion means and said carrier each being positioned circumferentially, radially, and axially so as not to interfere with said reaction stator proximate to said first axial location, thereby establishing a power bridge between said first and second power transmitting structures while still allowing said reaction stator to pass by said pinion means toward said second axial location; means for selectively coupling said reaction stator to said power transmitting path at said second axial location, said reaction stator further positioned so as to permit said means for selectively coupling to be substantially located within the interior of at least one of said power transmitting paths, thereby allowing said reaction stator to provide a reaction force to at least one said power transmitting path in said power transmission using said means for selectively coupling at a location other than proximate the mechanical reference frame.

11. The input arrangement of claim 10, wherein said first and second gear means are sun gears.

12. The input arrangement of claim 10, wherein said first and second gear means are ring gears.

13. The input arrangement of claim 10, wherein said first gear means .is a sun gear and said second gear means is a ring gear.

14. The input arrangement of claim 10, wherein said first gear means is a ring gear and said second gear means is a sun gear.

15. The input arrangement of claim 10, wherein the mechanical reference frame to which said reaction stator is coupled may rotate, thereby making said reaction stator a reaction rotor.

44

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