CROSS-REFERENCE TO RELATED APPLICATIONS This application is a continuation-in-part of application Serial No. 08/066,663 filed May 25, 1993, in the name of Gary D. Lee, entitled "Clutchless Mechanical Gear Transmission", which is commonly assigned with the present application. The entire contents of the '633 Lee application are expressly incorporated herein by reference.

TECHNICAL FIELD OF THE INVENTION The present invention pertains generally to a method and an apparatus for torque transfer between a driving member and a driven member and is more particularly directed to a new and improved method and apparatus for changing a gear ratio in a transmission without interrupting torque transfer during such change. BACKGROUND OF THE INVENTION The advent of the industrial revolution was brought on by the development of reliable and relatively inexpensive rotating power sources of the thermodynamic engine type, first in the form of the steam engine. Industry, transportation and agriculture have all progressed rapidly utilizing these engines. Later the duties of the steam engine were shared or displaced by newer thermodynamic engines of the internal combustion type, such as the diesel and otto cycle engines. Today these rotating power sources are ubiquitous and put to a myriad of uses in modern day society.

One of the general characteristics of the rotating power source is its torque or power curve where an engine will produce different amounts of power at different speeds. To most efficiently utilize the available power, such engines are operated in speed ranges which are based on their mechanical design. These ranges do not generally coincide with the desired speed and torque requirements of the end use or loads

and hence the need and development with these engines of a control mechanism, the transmission, to transfer and regulate the torque from such engines to the load.

A transmission usually comprises multiple sets of gears or speed variance devices to allow different speed ratios between the power source and the load and a means for changing or shifting between those ratios. In this manner the speed of the power source can be varied to deliver more or less torque while the load is changing while maintaining a constant or controlled speed for the load. Conversely, with a transmission a power source which is operated most efficiently at a constant speed can be used over a wide range of output loads and speeds. For many years the changing of gear ratios in a standard transmission has been accomplished by disengaging the engine from the transmission with a clutch. A clutch is a device which allows the power source to be decoupled for an instant so that the gear can be changed and then recoupled. What occurs is that the gear to be shifted into is not loaded and so can be brought up to the speed of the load fairly easily. The clutch is then reengaged by bringing two frictional surfaces together which have some slip to allow the engine and load speeds and torque transfer through the engine to equalize. It is the slip in the clutch that allows the two different rotational speeds of the engine and load to be matched for different power outputs and torque requirements. Several disadvantages pertain to this method where the power source is completely decoupled from the load while the gears are being shifted. The first is that when the engine is under very heavy load, such as a piece of earth moving equipment or the like, the right gear must be chosen to begin a movement or the piece must be stopped in the middle of a push to shift gears which causes a loss of momentum. This is because a

heavily loaded gear can not be "shifted on the fly" because the instant that the clutch is disengaged the load speed begins to slow and the unloaded engine will begin to race causing a speed mismatch too great to be handled by the slip in the clutch. If the engine speed is controlled to match the load speed, then the sudden application of a heavy load to the engine that exceeds its power output at that speed could cause it to stall. A second problem is that the clutch is operated by a person who must coordinate and judge the load/speed requirements between the engine and output and properly time the shift. While this is not a problem for most of the population after some learning, many shifts are not made efficiently and there is always some power wasted in the transition from one gear to another. While this is not significant for any one shift, particularly in passenger automotives, it does become more substantial over the operating life of larger, more heavily loaded engines used constantly such as those in heavy duty trucks where gains in operating efficiencies translate directly into saved fuel and operating costs. The human element of the timing of a gear shift becomes more critical when it has to be made under more extreme conditions. For example, in racing environments during accelerations (upshifting) or decelerations (downshifting) a very short time is available to recognize a change in the load/speed requirement, coordinate actuation of the clutch and actually make the shift. The race driver with the most efficient shifting in these situations will save time, fuel and consequently win races. In one particular type of racing, popularly termed drag racing, the timing of gear shifts to produce maximum acceleration while not losing traction is the most critical element to winning. To a lesser extent, but still importantly, this critical timing and higher coordination is required on the shifts made for heavily loaded trucks on upshifts and

downshifts while going uphill and downhill. Automatic transmissions have been developed in an attempt to overcome some of these problems and usually comprise a hydraulic torque converter where the rotation and power from the engine is used to set a hydraulic fluid in motion which is then coupled to different sets of vanes (or bands) on the output side to transfer the power in the moving fluid. While this type of transmission provides a constant transfer of torque while shifting between bands and inherent slippage to match changing load/speed requirement, it is relatively complex to build and manufacture and thus costly to produce. Further because the principle relied on is one of fluid momentum transfer and there is power loss in the input and output sides, it is much less efficient than the direct geared transmission while not changing gears.

What is needed is a transmission which has the efficiency and simplicity of the direct geared transmission while not shifting, does not interrupt torque transfer while shifting, and does not rely on human intuition for shift timing.

Other non-gear designs have been used for the purpose of changing the ratio while providing continuous torque. One such design is a double pulley arrangement constructed of metal elements that drives a V-belt. This design is presently used in small engine designs, such as those used with snowmobiles, but belt wear resulting in slippage and/or breakage has been a problem. Another design uses rolling elements between an inner input gear and an outer output gear arranged much like an epicyclic gear train. The rolling elements can continuously change the gear ratio but they are torque-limited because they rely on the coefficient of static friction to transmit forces between the rolling elements. Attempts also have been made to change gear ratios by randomly forcing detents between the closely

spaced teeth of internally-toothed gears, but this brute force approach still requires the disengagement of torque from the transmission. In addition, this approach may cause rapid wearing and/or breakage of both the detents and the internal teeth, and binding and/or lockup between adjacent gears.

SUMMARY OF THE INVENTION The invention provides a transmission apparatus for changing the speed of a driven member relative to the speed of a driving member without interrupting the torque transfer between the members. The invention comprises a plurality of driving gears each adapted to move in response to the driving member and a plurality of driven gears each adapted to move the driven member. The driving gears and driven gears are continuously engaged with each other to form gear pairs which are in different ratios and are adapted to vary the speed of the driven member relative to the driving member when there is a shift between gear pairs. At least one of the driving and driven gear sets is journaled for rotation on its respective driving and driven member. An actuating means is used to selectively shift between the different driving and driven gear pairs by simultaneously engaging one of the journaled gears to its respective member while disengaging another one of the journaled gears.

In accordance with one of the objects of the invention, the simultaneous engagement and disengagement of different gear pairs changes the gear ratio, and consequently the relative speeds between the driving and driven members, without interrupting torque transfer between the driving member and the driven member. Because the gear pairs providing the variable speed ratios are in constant engagement with each other there is no need to mesh a gear under load with another gear being driven at different speed, thereby eliminating the need for such devices as synchronizers, clutches and

other speed/load reconciliation apparatus. The apparatus eliminates the torque lost between shifts in a standard clutch type transmission so that the engine works more ef iciently when down shifting and upshifting because of load changes requiring a gear change.

Moreover, the human element in coordinating the timing between shifting gears is eliminated. Shifting can be done much faster than a human can react and more efficiently by essentially an automatic shifting device. Such apparatus retaining the efficient direct gearing of the standard transmission and not requiring the expensive and complex torque converters of the automatic transmission.

In the illustrated implementation, the transmission apparatus includes a plurality of reducing gears driven by the driving member, a plurality of annular cam gears in continuous engagement with the reducing gears forming different gear ratio pairs, and a driven member on which the cam gears are journaled for rotation. The driven member comprises a rotatably mounted hollow shaft section having a plurality of radially extending apertures, each containing a reciprocating detent for engaging and disengaging a corresponding cam gear.

The actuating means includes a shift member reciprocally mounted within the inner bore of the driven member for movement along its rotational axis and having at least one cam surface for selectively moving one detent between an engaged position and a disengaged position on one cam gear while simultaneously moving another detent from a disengaged position to an engaged position on another cam gear.

The simultaneous engagement and disengagement between the two gear pairs associated with the cam gears produces a smooth shift in gear ratios where the torque transfer between the driving member and driven member is never interrupted but is advantageously transferred during the shift by the continuous engagement of at

least one of the detents to one of the cam gears at all times. The angular velocity difference between the two cam gears is reconciled by the timing of the shift during an allowed shift window which is dependent upon the design of the cam gears and the angular velocity difference. The angular velocity difference and reconciliation problem is minimized by having all gear pairs constantly engaged so that the difference in speeds between adjacent gears is dependent primarily on their different gear ratios and not the input speed of the power source.

Preferably, an inner cam surface of each cam gear has peak sections and valley sections and axial movement of the shift member is timed relative to the position of the peak sections and valley sections on adjacent cam gears.

The invention lends itself to several applications and has several variables. For example, the gear ratios as well as the number of peaks and valleys in the cam gears may all be manipulated to accomplish the task for which the particular application of the invention is intended.

Therefore, the essence of the invention lies in the following principles, (a) the cams machined into an inner perimeter of output gears become a common link to other output gears which vary in size and rotational speed. By means of this common link, an engagement device can follow one cam out of engagement with one gear as another engagement device simultaneously follows a cam of an adjacent gear into engagement, (b) the required timing to start the above mentioned exchange can be determined from the position of the cam gears by manipulating the ratio of the gears in relation to the number of cams repeated around the inner perimeter of the cam gears, (c) the movement of the engagement devices can be synchronized by knowing the angular velocity difference between adjacent cam gears by sensing the speed of a driving member and setting the

gear ratios between adjacent cam gears, and (d) differing angular velocities of the gears can be reconciled by a driven member commonly connected to the engagement devices. BRIEF DESCRIPTION OF THE DRAWINGS

The invention may be further understood from consideration of the following detailed description of the preferred embodiment when taken in conjunction with the accompanying drawings in which: Fig. 1 is a partially schematic side elevational view, in cross section, of a transmission apparatus constructed in accordance with the teachings of the invention;

Fig. 2 is an end view of the shift member illustrated in Fig. 1;

Fig. 3 is a side view in cross-section of the shift member illustrated in Fig. 1 taken along section line 3-3 in Fig. 2;

Fig. 4 is an end view of one of the cam gears illustrated in Fig. 1 showing a six valley cam gear;

Figs. 5A to 5F are a pictorial representation of the cam surfaces of two adjacent cam gears, their respective detents and the movement of the driven member during a downshift cycle of the transmission illustrated in Fig. 1; and

Figs. 6A to 6F are a pictorial representation of the cam surfaces of two adjacent cam gears, their respective detents and the movement of the driven member during an upshift cycle of the transmission illustrated in Fig. 1. Fig. 7 is a partially schematic side elevational view, in cross section, of a transmission apparatus with the shift assembly associated with the output shaft;

Fig. 8 is a partially schematic side elevational view, in cross section, of a transmission apparatus with the shift assembly associated with the input shaft;

Fig. 9 is a partially schematic side elevational view, in cross section, of a transmission apparatus with the

shift assemblies associated with both the input and output shafts,-

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT Fig. 1 shows a partially schematic, side elevational view in cross section of a preferred embodiment of a transmission apparatus 10 constructed in accordance with the invention. An input shaft, or driving member, 12 is splined for connection to a source of power, preferably a thermodynamic engine, or the like, at its driven end 13 on the outside of a transmission housing 14. The shaft 12 is journaled for rotation inside the housing 14 by bearings 16. The bearings 16 mount the input shaft 12 for axial rotation without lateral movement, and may be of the thrust bearing type to prevent such movement. The input shaft 12 mounts through the center aperture of, and is locked by a key 18 to, each of a plurality of reducing gears 20 of varying circumferences. An output shaft, or driven member, 22 includes an enlarged hollow shaft section 26 on which a plurality of cam gears (driven gears) 24 are journaled for rotation. The shaft section 26 is rotatably mounted by bearings 28. The bearings 28 also retain the cam gears 24 in their proper lateral positions along shaft section 26.

Each of the input reducing gears 20 is continuously engaged by means of a gear tooth profile on its outer circumference to a corresponding gear tooth profile on one of the output cam gears 24. Each pair of input and output gears, 20 and 24 respectively, that is, a reducing gear 20 on the input shaft 12 and a cam gear 24 on the output shaft 22, has a different gear ratio than an adjacent pair of gears. In many applications, each pair of gears will have a different gear ratio than every other pair of gears in that particular transmission. Preferably, the gear ratios increase in one direction along the shaft 12 and consequently decrease in the other direction.

Each cam gear 24 has a uniform cam surface 58 on its inner circumference, and the peaks 65 of the cam surface 58 ride on the outer surface of the shaft section 26 of the output shaft 22. The cam surface 58 can be shaped to influence the characteristics of a shifting sequence, depending upon the application. For example, in a six cam segment design, where the rotations of adjacent cam gears 24 differ by one third for every rotation of the input shaft 12, the cam segments can be divided into two sets, one set which lines up for a shift on a zero degree (0) mark of the input shaft 12 and the other set which lines up for a shift on a 180 degree mark. One set can have sections with profiles to perform specific functions for upshifting uphill and downhill, and another set can have sections with profiles to perform specific functions for downshifting uphill and downhill. The output shaft 12 includes a hollow cylinder section 26 (referred to herein as the shaft section) whose outer diameter is sized to fit inside the inner diameter of the cam gears 24 with a sliding oil tolerance fit as defined by the peak sections 65 that rest on the surface of the shaft section 26. Guide apertures 30 (generally two or more to correspond with the number of valleys 64 in the cam surface) , extending from the inner surface of the central bore of the shaft section 26 to its outer surface, contain and guide the detents 32 which are shaped to engage the valley sections 64 of the cam surface 58. The detents 32 may have a variety of shapes, such as spherical, rectangular or triangular, and preferably the shape of the engaging portion of the detents 32 conforms to the shape of the cam valley 64 so as to maximize the surface to surface contact area for locking the cam gear 24 to the shaft section 26 during full engagement of that cam gear 24. The detents 32, each of which may be a metal ball, a pin, a bolt, a catch, a rectangular shaft, or the like, have sides

shaped to slide in and out of the guide apertures 30 cut in the shaft section 26, and a bottom end shaped to follow the contoured cam grooves 74 in the cam surface 36 of the shift member 34. In some applications, the inner cam surface 58 of the cam gears 24 and the detent 32 arrangement are so configured as to minimize the forces acting upon the shift member 34 and to maximize the forces acting upon the aperture wall 27 of the shaft section 26. Generally, the steeper the angle or curve of the top of the detent 32 and the corresponding incline sections 62 of the cam surface segment, the more force is directed into the shaft section 26 rotation. The detent 32, cam surface 58 tolerances and an optional oil orifice in the detent 32, control the movement of oil for the purpose of lubricating and dampening the shift. Each set of guide apertures 30 may be located inline or offset with respect to each other in the wall of the shaft section 26 of the output shaft 22. This places each set of the detents 32 in the radial plane of its corresponding cam gear 24.

A generally bullet shaped shift member 34 is mounted within the central bore of the shaft section 26 such that it is capable of reciprocal movement along the rotational axis of the output shaft 22. The shift member 34 contains six cam surfaces 36 in the form of grooves 74 each extending along the rotational axis and having a central portion extending radially upon which the detents 32 ride for radial reciprocating movement. A first follower surface 33 on the inner end of the detents 32 comes into contact with the radially and axially extending cam surfaces 36 of. the shift member 34.

A shift member positioning mechanism is used to control and change the position of the shift member 34. This can be accomplished in a variety of ways, such as by a mechanical lever, a set of gears, an electrical motor, an electrical solenoid, a pneumatic mechanism, or

another device for moving shif member 34. All these positioning mechanisms must be responsive to timing and capable of synchronizing the rate of shift member 34 movement with the detents 32. The shift member positioning mechanism also may be responsive to a torque-sensing device in order to initiate a shift in automatic response to the load on the output shaft 22. The positioning mechanism preferably comprises a positioning cam element that is rotated in timed (synchronized) relation to the cam gear 24 rotations by being driven by the input shaft 12 or output shaft 22. As shown in Figures 1 and 4, the axial movement of the shift member 34 is regulated by a screw shaft 38, one end of which is rotatably attached to the shift member 34 at pivot joint 35, and the other end of which passes through an aperture 37 at the distal end of shaft section 26.

As stated above, all of the gear pairs are continuously engaged, and also the reducing gears 20 are affixed to input shaft 12. However, only one of the cam gears 24 is engaged to the output shaft section 26 via the detents 32 for that cam gear 24 of the pair. Power is transmitted from the input shaft 12 to the output shaft 22 via the one engaged gear pair. The shift actuator gear 50 has teeth on its outer circumference that engage the teeth of the shift collar 56. The center hole of the shift collar 56 is threaded to receive the screw shaft 38. The screw shaft 38 is held from rotating about its own axis by a screw shaft washer guide 57. For this purpose, the screw shaft washer guide 57 is bolted to the housing 14 and has a key (not shown) built into its inner circumference that slides in a slot 60 (shown by broken lines) in the screw shaft 38.

The transmission apparatus 10 includes an actuating means which, in concert with the shift assembly, times the engagement of the one cam gear 24 while simultaneously disengaging another cam gear 24 during

the shift window. The actuating means illustrated in the embodiment of Fig. l comprises a variable speed, reversible electric motor 61 coupled by a shift actuator gear 50 to the shift collar 56. The motor 61 rotates the shift actuator gear 50 in one direction for an upshift and in the other direction for a downshift. This rotation causes the screw shaft 38 to move axially to the output shaft 22 and carry the shift member 34 in one direction of rotation and in the reciprocal direction for the other direction of rotation. The distance which the shift member 34 moves and the speed at which it moves is a function of the speed of the shift actuator gear 50 and the number of its revolutions for each movement. The motor 61 is regulated in its speed and movement with a controller 48 which generates a control voltage to the motor 61 of a magnitude which is proportional to the speed desired and has a polarity which is representative of the direction of the travel of the shift member 34. The controller 63 receives inputs from a position sensor which generates information as to the angular velocity of the input shaft 12 and its instantaneous position and a shift command which indicates that a shift should take place and whether it is to be an upshift or a down shift. Upon the coincidence of a shift command and the start of the next shift window, the controller 63 generates the correct magnitude and polarity of voltage which will cause the motor 61 to move the shift member 34 the requisite amount in the right direction at the correct speed.

With the illustrated implementation, the distance the shift member 34 must move is substantially the same for an upshift or a downshift, and between any gear pair. The number of revolutions of the motor 61 for any shift is substantially the same which greatly simplifies timing and control of the motor 61.

The time during which shift takes place, termed the shift window, is a function of a number of variables but can be designed to be a fixed number of degrees of rotation times the difference in angular velocity between the engaged gear 24 and the to-be-engaged gear 24. Because an engaged gear 24 and a to-be-engaged gear 24 are both referenced to the input shaft through their respective gear ratios of their gear pairs, the difference in angular velocity between them is a constant times the angular velocity of the input shaft 12. This constant, while fixed for either an upshift or a downshift, varies among the different gear ratios but can be designed to linearly decrease for an upshift cycle and to linearly increase for a downshift cycle. The gear shifting assembly is composed of a plurality of cam gears 24, each of which is shaped like a flat donut and has an outer surface with gear teeth 59 in profile. A central aperture through each cam gear 24 provides a continuous inner cam surface 58 having at least one segment, preferably repeating segments, cut into an inner perimeter of the gear 24. Each segment of the cam surface 58 is generally composed of a plurality of indented sections which form a valley 64 into which a correspondingly shaped detent 32 can be inserted. A peak section 65 created between adjacent valleys 64 is also identified as one of the segment sections and includes a top or crest. A side or incline section 62 between the peak section 65 and the valley section 64 and a second side or incline section 63 connecting adjoining segments are each steep enough that when the detent 32 is inserted into the valley section 64 and held there by the underlying shift member 34, the cam gear 24 and the output shaft 22 are locked together. The incline sections 62 and 63 create continuous intermediate channel portions between the valley sections 64 and the peak sections 65 to facilitate alternate movement of the detents 32 between the engaged

and disengaged positions responsive to reciprocating movement of the cam gear 24.

Alternatively, the cam gear 24 assemblies may be associated with the input shaft 12 instead of the output shaft 22, as shown in Figures 7 and 8. Separate cam gear 24 assemblies may also be associated with both the input and output shafts 12 and 22 respectively, as shown in Figure 9. Although these designs may provide different input and output ratios, the rules of interaction between the components of the shifting assembly are the same in each design.

The shift member 34 is preferably a cylindrical piece that fits inside the inner diameter of the shaft section 26 of the output shaft 22 and has contoured grooves cut into its outer surface to receive the bottoms of the detents 32. The bottom of each groove 74 in the shift member 34 provides a cam surface 36 which is designed

(along with the shape of the cam surface 58 on the inner perimeter of the cam gear 24) to hold the detents 32 in a desired radial position and to control the characteristics of the shifting sequence. The shift member 34 is connected to and maneuvered by a bullet positioning mechanism. Generally, the shift member 34 moves in synchronization with the forces created by movement of the detents 32 along the cam surfaces 58 in the active cam gear 24 and the inactive cam gear 24.

The shift member 34 is not limited to having a number of cam grooves 74 equal to the number of detents 32. For example, the shift member 34 can be rotated in the shaft section 26 to change its angular orientation with respect to the shaft section 26 in order to utilize one set of cam surfaces 58 for an upshift cycle and then rotated back to utilize another set of cam surfaces 58 for a downshift cycle. Thus far, this specification has assumed a shift member 34 with longitudinal cam grooves which translates along its longitudinal axis to complete a shift, as

shown in Figures 1 through 3. Other shift member 34 configurations are possible. Any configuration which defines radial detent motion for a corresponding shift member 34 motion may be utilized. For example, the longitudinal cam grooves which define radial detent motion for corresponding longitudinal shift member 34 translation could be cut circumferentially around the outer perimeter of the shift member 34 so as to have the shift member 34 produce a shift upon rotation rather than translation. Or, the cam grooves could follow a corkscrew configuration, for which screw-like rotation translation of the shift member 34 would produce a shift. In addition, the shift member 34 could be replaced by a hydraulic fluid arrangement, which would also actuate proper radial detent movement.

Shifting Sequence The power flow characteristics (speed and torque) through the transmission 10 can be changed by unlocking an engaged cam gear 24 from the output shaft 22 while simultaneously locking an adjacent disengaged cam gear 24 to it. This is a gear "shift" which will vary the relative speeds of the input shaft 12 and output shaft 22 depending upon the change in the gear ratios of the gear pairs shifted between. The shift can be an upshift or a downshift depending upon whether the gear ratio is increased or decreased as a result of the shift.

For this specification, the active gear pair is the pair of gears that are currently transmitting power. An upshift is a change in the active gear pair such that the ratio of the speed of the input shaft 12 to the speed of the output shaft 22 is a numerically smaller number, for example, a shift from the gear ratio (input to output) of 5:1 to a gear ratio of 4:1. A downshift is a change in the active gear pair such that the ratio of the speed of the input shaft 12 to the speed of the output shaft 22 is a numerically larger number, for

example, a shift from the gear ratio (input to output) of 4:1 to a gear ratio of 5:1.

The shifting time and the sequence of detent 32 movements are the result of the configuration of the cam surfaces 58 on the cam gears 24, the shift member 34 and the positioning cam element. The design and arrangement of the components that constitute the output shaft 22 and the associated positioning mechanism for moving the shift member 34 also influence the shifting operation. The following description of the shifting sequence will start with the transmission apparatus 10 rotating with the active gear pair in the mid-range of the transmission 10. This means that the transmission 10 is operating in a gear ratio from which it can be either upshifted or downshifted. The power flow enters through the input shaft 12 and is modified in speed and in torque by the gear ratio of the active gear pair.

The active gear pair is locked to the output shaft 22 because the shift member 34 has been positioned under the detents 32 for that gear pair, thereby forcing the active gear pair's corresponding detents 32 radially outward into the open spaces formed by the cam valleys 64 in the cam surface 58 cut into the inner periphery of the annular cam gear 24. The detents 32 are held in their outermost radial position in abutment with the bottoms of the cam surface valleys 64 by the corresponding cam grooves of the shift member 34. Therefore, the cam gear 24 and the output shaft 22 are locked together with the cam gear 24 engaged and driven by the reducing gear 20.

All of the other detents 32 for the gear pairs, other than the active gear pair, are not being held radially outward in position against the cam surface 58 of their respective cam gears 24 by the shift member 34. Any force acting between the detents 32 and the cam surface 58 of such an "inactive" cam gear 24 has a radial component which will drive the detents 32 radially

inward, and hence, the inactive cam gears 24 are free to rotate relative to the output shaft 22 while being continuously driven by the reducing gears 20. In addition, the shift member 34 grooves can be designed to trap the detents 32 radially inward in order to ensure that they follow the shift member's 34 grooved surface and that the inactive cam gears 24 are therefore free to rotate relative to the output shaft 22.

The inactive gear pairs adjacent to the active gear pair have gear ratios that are different from that of the active gear pair. Therefore, the cam gears 24 on the output shaft 22 that are adjacent to the active cam gear 24 are moving relative to the active cam gear 24. Each adjacent inactive cam gear 24 is also moving relative to the output shaft 22, and the corresponding detents 32 of the inactive cam gears 24 are all retracted into the inactive apertures of the "shaft section 26 of the output shaft 22. The magnitudes of these relative motions are different for each inactive cam gear 24 and depend on the differences between the gear ratios of the active and the inactive cam gears 24.

The relative motion between the active cam gear 24 and an inactive cam gear 24 causes the valleys 64 of the cam surfaces 58 of the inactive cam gear 24 and its retracted detents 32 in the output shaft 22 to periodically coincide or line up. When the valleys 64 of the cam surface and the detents for either of the cam gears adjacent to the active cam gear line up, then a gear shift can occur. In other words, as the cam surface valleys 64 and the detents 32 for either cam gear 24 adjacent to the active cam gear 24 approach coincidence, the shift sequence can start. For every rotation of the engaged cam gear 24, the adjacent higher speed cam gear 24 rotates a fraction more than once. To ensure at least one shifting position per revolution, this fraction must be an integer multiple of the reciprocal of the number of cam segments in the cam gear

24. For example, if the cam gear has six segments (i.e., six valleys 64 and six peaks 65), the additional rotation must be at least one-sixth of a rotation or sixty degrees more than one revolution. This principle applies for each consecutive cam gear 24. For every rotation of the engaged cam gear 24, the adjacent higher speed cam gear 24 must travel at least one-sixth, one-third (two-sixths) , one-half (three-sixths) , two-thirds (four-sixths) , or five-sixths and so on, more than the engaged cam gear 24. When two shifting positions are desired in one rotation, the fractional increase must also be a multiple of two. The different amounts of rotation from one cam gear 24 to the next adjacent cam gear 24 as well as the number of cam gears 24 all affect the positions at which the shift can be initiated. During the shift sequence, the engaged detents 32 of the active cam gear 24 must be allowed to retract in order for the shift to continue to take place. This is accomplished by the timed synchronous movement of the shift member 34.

Downshifting A downshift requires that the cam gear 24 of the adjacent gear pair having a numerically larger gear ratio be engaged by its detents 32 substantially simultaneously while disengaging the detents 32 of the active gear pair. Because the to-be-engaged gear pair has a numerically larger gear ratio, the to-be-engaged cam gear 24 on the output shaft 22 is rotating at a slower rate than the output shaft itself 22. Hence, the detents 32 in the apertures of the output shaft which are to be aligned with the to-be-engaged cam gear 24 are rotating faster than the cam gear 24 itself, and therefore, after a certain angle of rotation, these detents will be movable into the position required to be fully engaged 32 with their cam gear 24.

For a downshift, the output shaft 22 will start to slow immediately when the shift sequence starts. This

is because as the active detents 32 start to withdraw, they slide inwardly along the cam surface 58 of the cam gear 24 causing the output shaft 22 to rotate more slowly than the cam gear 24. Both sets of detents 32 continue to move until the to-be-engaged detents 32 reach their fully inserted position where they lock their cam gear 24 to the output shaft 22, at which time the output shaft 22 rotation rate is determined by the gear ratio of the newly active gear pair. For a downshift, the final output shaft 22 rotation rate will be slower than the rotation rate before the shift. The resulting rotation rate of the output shaft 22 during a downshift can be varied from application to application by changing the shape of the cam gear segments as well as the shape of the shift member 34 grooves. Figs. 5A-5F illustrate a gear shift from a high speed cam gear HG (low ratio) to a lower speed cam gear LG (high ratio) , which is a downshift cycle caused by driving shift member 34 to the right (Fig. 1) . When a downshift cycle commences, the shaft section 26 of the output shaft 22 is locked to and rotating with the faster rotating high speed cam gear HG due to the engaged detents HD (HIGH SPEED CAM GEAR DETENTS) . The disengaged detents LD (LOW SPEED CAM GEAR DETENTS) inside the to-be-engaged slower rotating low speed cam gear LG are rotating faster than the low speed cam gear LG itself. The gear to detent driving force "F" and the shaft section to detent resistance "R" (Fig. 5A) push the detents HD radially inward so that they slide along cam sections HS1 (HIGH SPEED CAM SECTION NO. 1) and are released from the faster rotating cam gear HG. Simultaneously, the detents LD (LOW SPEED CAM GEAR DETENTS) , by means of the shift member 34, are forced radially outward and slide outwardly along cam sections LSI (LOW SPEED CAM SECTION NO. 1) until they reach the valley bottoms 64. The detents HD will slide radially inward just as fast as cam sections LSI will let the

detents LD slide radially outward. These movements are synchronized by the travel speed and cam surfaces of the shift member 34.

As cam sections LSI and the shift member 34 allow the detents LD to move radially outward, the shift member 34 is allowed to move along its longitudinal axis, which allows the detents HD to move radially inward, which slows down the shaft section rotation and simultaneously slows down how fast the detents LD move outward. All of these simultaneous movements affect the output shaft section rotation and the number of degrees that the input shaft must rotate to complete a shift.

As may be seen in Figure 1, all of the gear rotations are fixed relative to the input shaft rotation. The relationship between the force F and the resistance R along cam sections HS1 and at peak tops 65 move the detents HD all the way in. Simultaneous movement of the shift member 34 moves the detents LD all the way out in order to complete a gear shift. Figs 5B-5F include an illustrative example of the relative angular changes in the shaft section angle DA, the low speed cam gear angle LA and the high speed cam gear angle HA during a downshift cycle. Thus, shaft section angles DA of 35 degrees, 72 degrees, 108 degrees, 144 degrees, and 180 degrees may correspond to high speed cam gear angles HA of 41 degrees, 82 degrees, 123 degrees, 165 degrees, and 205 degrees, respectively, and to low speed cam gear angles LA of 31 degrees, 62 degrees, 92 degrees, 123 degrees, and 154 degrees respectively. The gear shift is complete when the peak tops 65 of high speed cam gear HG are in the same position of rotation as the valley bottoms 64 of low speed cam gear LG (Fig. 5F) .

The ability to predetermine the lining up of the valleys 64 of the engaged cam gear with the peaks 65 of an adjacent unengaged cam gear for the purpose of initiating a shift is a product of, and timed with reference to, the angular rotation of the input shaft

12. If the slower speed cam gear travels X degrees for each rotation of the input shaft, then the higher speed cam gear travels X + YX degrees, where Y is a variable parameter determined by the ratio between pairs of adjacent gears. If these gears traveled at the same velocity there would not be any potential for creating a gear shift. The difference between the angular rotations caused by the difference in the gear ratios of adjacent cam gears affects the shift interval (the angular rotation of the input shaft required to complete a shift) . Thus, (X + YX) - X represents the arc difference between output gears. Because the downshift involves cam section HS1 and cam section LSI, the difference in rotation must allow for the full angular arc subtended by cam sections HS1 and LSI. Hence: (X + YX) - X = the arc subtended by HS1 plus the arc subtended by LSI.

Upshifting An upshift requires that the adjacent gear pair which has a numerically smaller gear ratio be engaged while simultaneously disengaging the active gear pair. The upshift is similar to the downshift except that the cam gear on the other side of the active gear pair is the one to be engaged and the cam gear for the to-be-engaged gear pair is rotating faster that the output shaft. Further, because the to-be-engaged gear pair is rotating faster than the output shaft, the corresponding detents 32 approach their respective cam surfaces from the opposite sides of the valleys 64 as compared to a downshift. In other words, each of the detents for the to-be-engaged gear pair will slide outwardly along the opposite side section of the V or U shaped cam surface valley 64 of the corresponding cam gear.

For an upshift, the output shaft will also start to slow down -immediately when the shift sequence starts. This is because as the engaged detents start to withdraw, they slide inwardly along the cam surface of

the cam gear to be disengaged causing the output shaft to rotate more slowly than the cam gear. Both sets of detents continue to move until the to-be-engaged detents reach their fully inserted position where they lock their cam gear to the output shaft, at which time the output shaft rotation rate is determined by the gear ratio of the newly active gear pair. For an upshift, the final output shaft rotation rate will be faster than the rotation rate before the shift. The resulting rotation rate of the output shaft during an upshift can be varied from application to application by changing the shape of the cam gear segments and as well as the shape of the shift member 34 grooves 7 .

Figs. 6A-6F illustrate a gear shift from a low gear (high ratio) to a high gear (low ratio) , which is the upshift cycle caused by driving shift member 34 to the left (Fig. 1) . The detents and cam gear surfaces of Figs. 6A-6F differ from those of Figs. 1 and 4, and as with Figs. 5A-5F, are shown as examples of varying structural modifications.

Initially, the shaft section 26 of the output shaft 22 is rotating with the low gear LG by means of the engaged detents LD. The shaft section 26 of the output shaft 22 and the disengaged detents HD rotate slower than the high gear HG. The force F and the resistance R push the detents LD radially inward, causing them to slide along cam sections LSI and releasing them from the low gear LG.

In the upshift cycle, the resistance R causes most of the movement required for the shift. The active cam surface sections HS2 and LSI of the gears HG and LG, respectively, participate in this negative change, i.e., slowing down, in shaft section rotation. If detents HD were not considered, the force F and resistance R acting on the detents LD of the slower gear LG would alone move the shaft section through an angle equal to the arc angle subtended by cam section LSI. Figs. 6B-6F include

an illustrative example of the relative angular changes in the shaft section angle DA during an upshift cycle. Thus, shaft section angles DA of 2 degrees, 3 degrees, 5 degrees, 7 degrees and 9 degrees may correspond to high speed cam gear angles HA of 7 degrees, 14 degrees, 21 degrees, 27 degrees and 34 degrees respectively, and to low speed cam gear angles LA of 5 degrees, 10 degrees, 15 degrees, 21 degrees, and 26 degrees respectively. If the angle subtended by cam sections HS2 and LSI were equal, no rotational difference would be required between high speed gear HG and low speed gear LG. But if these angles were equal, the shift would theoretically have to finish the instant that it started. Therefore, it is important for the length of cam section HS2 to be longer than the length of cam section LSI to allow time for upshifting. In this case, the detents LD will slide radially inward just as fast as cam sections HS2 will let the detents HD slide radially outward. As indicated above, the resistance force R along cam section LSI is responsible for moving the shift member 34 an axial distance corresponding to the arc angle subtended by cam segment LSI. The difference in length between cam section HS2 and cam section LSI corresponds to the difference in gear ratios between low speed gear LG and high speed gear HG for any given design. There are many manipulations that can be made to change this outcome.

In the upshift cycle, it may be said that the detents HD travel negatively with respect to their cam gear because detents HD are not engaged by .cam gear HG until they are substantially at the valley bottom 64 (Fig. 6F) . It is preferable that the length and shape of all cam surface sections have such a relationship that the cam surfaces of the shift member 34 may be symmetrical. Otherwise a more sophisticated version of the shift member 34 must be implemented.

Similar to the downshifting sequence, (X + YX) - X represents the relative rotation between output gears. Because the upshift involves cam section HS2 and cam section LSI, the difference in rotation must allow for the difference between the angular arcs subtended by cam segments HS2 and LSI. Hence: (X + YX) - X = the arc subtended by HS2 minus the arc subtended by LSI.

Many other modifications of the constructions shown and described above will be apparent to the skilled person, and these modifications are within the scope of the appended claims. For example, the central portion of the cam surfaces of each shift member 34 may be made more elongated and flat. Also, these and the other cam surfaces may have shapes and configurations other than those shown herein by way of example. A similar result may be accomplished by flattening out the shift cams and thereby reducing the ratio at which the shift member 34 moves axially during each gear shift.