GB2345521A - A stepless variable-ratio transmission using a variable-throw crank - Google Patents

Abstract

A variable ratio gearing has a crank pin 19 variably positioned between a pair of driving discs and connected by a plurality of connecting rods 20 to sliders with rack teeth 21, the movement of the sliders as the disc rotate varying with the eccentricity of the crank pin. The sliders slide in radial guides of a fixed member. Each slider has two sets of teeth 21 cooperating with a respective pair of gears 23a, 23b under the control of a cam mechanism 22 such that one set of teeth engage a gear 23a on the outward stroke and the other set engages a gear 23b, coaxial with gear 23a, on the inward stroke, the two gears being separately geared to the output such that both strokes rotate it in the same direction (Fig. 1). Eight connecting rods will thus provide sixteen driving impulses in succession to the output. A helical spline arrangement on the output gearing allows the teeth positions of the racks 21 relative to the gears 23 to be adjusted to so that the teeth engage properly.

Description

Infinitely Variable Gearbox This invention relates to an infinitely variable speed gearbox.

Gearboxes are normally used between component to enable the driven component to fiction at the required speed irrespective of the input speed of the driving component.

In matching these requirements a typical gearbox will provide a series of gear ratios to give set steps in the relationship of input to output speeds. With the use of gears ont it has not been possible to provide a continuously variable ratio.

According to the present invention there is provided an infinitely variable gear ratio from an output speed of zero to full speed using gears to transmit the drive.

A specific embodiment of the invention will now be described by way of example with referez to the accompanying semi diagramatic drawings in which: Figure 1 illustates an input shaft to a gearbo2c and a pair of discs comected bs a crank pin. The crank pin can be moved down the discs thmugh the disc centrelines to a position 180 firm the original position. A possible crank pin position control arrangement is shown.

Figure 2 illustrates the side view of the inlet discs and shows a connecting rod connecting the crank pin to a slide. For clarity only one connecting rod is shown, the two operating positions for that rod are both indicated. A total of eight slide positions are shown, each position being served by a separate connecting rod. All connecting rods connect onto the one common crank pin on the discs.

Also shown are a series of fixed gear positions which engage with the sliding gear racks and the position of the common outlet gears which are permanently engaged with their respective fixed gears. The inlet drive to the discs is shown.

Figure 3 illustrates the slide arrangement and the attached sliding gear racks. Also shown is the fixed cam shaft and cams which engage with the two racks of gears.

Figure 4 shows x detailed view of the connecting rod, gear rack and cam rocker arrangement.

Figure 5 illustrates a helical spline used to adjust outlet gear angular position.

Figure 6 illustrates the drive to the camshafts from the discs.

Figure 7 shows a block diagram of the transverse gearbox anangement.

Figure 8 illustrates a possible alternative crankpin positioning arrangement.

Figure 9 illustrates a possible alternative slide and gear arrangement.

Referring to the drawings, the input speed passes through shaft 17 into discs 18a and 18b causing crank pin 19 to rotate about the disc centreline. This gives a balance drive to crank pin 19. Eight connecting rods are connecte to the one crank pin and each connecting rod terminates at a separate slider nos. 1 to 8. On each slider is positioned sliding gear racks 21 which move forward and backward with the slider. The gear racks pass over cams 22 on the forward and backward strokes in such a manner that the gear racks are liftez for a short space of time to engage in the fixed gears 23a and 23b respectively (for slider 1) on the forward then backward stokes. This is obtained by arranging for the cams to rotate at the same speed as the discs 18, cam lift occurring on the gear racks only during one sixteenth of the revolution forward and one sixteenth of the revolution backwards (plus a slight overlap). The cam drive is taken from the disc drive to cam shafts on each slide, figure 6. Thus during one revolution of the disc 18 the sliding gear racks 21 drive the gear wheels 23 a to q sequentially through sixteen separated, but synchronised, movements.

The choice of eight connecting rods operating with eight sliders using increments of one sixteenth of the revolution is optional and the system would work with other increments using more or fewer connecting rods and sliders. This example uses one sixteenth increments in order to reduce the speed pulsation's which occur in transferring the drive from rotary to linear movement. Advantage is taken of this pulsation to obtain the transfer of drive between increments of rotation. This is explained as follows :- at the start of the one sixteenth increment the gear rack is moving at exactly the same speed as the preceding gear rack (which is at the end of its one sixteenth of travel). During the one sixteenth of travel the gear rack increases speed to its mid point of travel then decreases speed to the end of its one sixteenth of travel. Whilst this is happening its associated cam is holding the gear rack in engagement with its respective gear wheel 23 as described above. Thus when the cam engages the rack gear at the start of the stroke the acceleration ensures it takes the drive from the preceding rack gear which is decelerating. The preceding rack gear has to be disengaged immediately to allow the accelerating rack uninterrupted movement. The run down of the cam contour should allow the necessary clearance to occur at the teeth during the run down part of the disengagement cycle.

The speed of output gear wheels (24a, b & c) is related to the diametrical position of the crank pin 19, as the crank pin approaches the centreline of the discs the linear velocity of the crank pin reduces and this reduced velocity results in the output gear angular velocity reducing until, when the crank pin is on the disc centreline the output wheels are stationary as there is no reciprocating motion at the respective slides.

Full reverse rotation occurs when the crank pin 19 moves to the 180'position at the other extreme of travel across the discs. This is because the cam operation continues on its original cycle and is 180 reversed to the disc cycle.

Positioning the crank pin 19 can be completed by various methods using normal engineering techniques such as those shown on figures 1 and 8. Similarly the drives to the cams can be achieved by varous methods using normal engineering techniques.

The whole scheme relies on the synchronised operation of the rack gear teeth engaging with the relevant gear wheels cleanly at the correct velocity with no clashing. The system as described can be set up initially completely synchronously to givc smooth nmning. Even at minimum slide stroke when the gear racks are adjacent to the mating gear teeth for the complete stroke the synchronised cam action will still allow the gears to be driven successively in the required manner. To allow the synchronised operation to occur using both the forward then reverse stroke of the slider it is necessary to have three output gears (24a b & c) driven from gears 23a-q. The outward stroke gears driving gear wheel 24a and the inward stroke gears driving gearwheel 24b. Gear 24b will turn in the opposite direction of rotation to output gear 24a. A third output gearwheel 24c is provided and this is driven from gear 24b through an interconnecting gear 27. By this means gears 24c and 24a rotate at similar angular speeds and in the same direction of rotation-On the shaft connecting gearwheel 24c to gearwheel 24a an axial moving sleeve is fitted. This sleeve contains a straight spline on its inner diameter mating with a spline on the shaft, on the sleeve outer diameter a helical spline is formed and this engages with a helical spline on gear 24c.

Thus axial movement of the sleeve can cause gear 24c to rotate a few degrees in either direction. This is necessary to enable the inward stroke rack gear 21 to engage with gears 23 in a synchronous manner during and following gear ratio changes caused by movement ofcrankpin 19 across the discs and compensates for the', xl movement indicated in figure 4.

The sleeve movement control is directly related to crankpin movement control and it is possible with appropriate design input for the one controller to actuate the two movements.

The final drive outlet is taken fi-om gear 24a. this arrangement of gears enables both rack gears to remain n in automatic synchronous engagement with gears 23 at all gear ratios.

Thus when the crank pin position is varied and bo gear rack operating positions move along the slide by a few millimetres both engagement positions automatically adiust to suit. Other methods of adjusting the back stroke synchronisation are possible but the sliding helical sleeve is put forward in this scheme. An alternative would be to use controlled rotation of gears 27 around the gearbox centreline.

The sliding action against the cam in the system described can be largely reduced if the cam is arranged to drive a rocker (similar to an overhead valve engine) with the opposite end of the rocker from the cam containing a bearing, or bearings, which would have rack contact area free to rotate at a speed to match the linear velocity of the rack, thus the iift action on the gear rack would be mainly a rolling action. There are many alternatives to the arrangement shown for operating the gear rack incln the use of hydraulic tappets or other techniques. All would be aimed at reducing the potential wear at this point caused by the sliding action and maintaining proper tooth contact when operated.

The gear racks are made as light as possible subject to design requirements to reduce the energy required to operate them and to enable higher output speeds to be coped with. The gear racks would be designed to limit their movement to ensure correct gear tooih engagement. Gear design would require to allow for the slight pulsation in speed during the gear rack operating cycle. A cushion drive design in the fixed gears 23 (or elsewhereeven in the rack) might reduce the pulsation but should not be allowed to interfere with synchronisation between the gear systems. An alternative rack gear operating arrangement whereby the two racks are arrange to pivot on spindles set onto the moving sliders could be designed and would give a positive drive arrangement.

Each slider unit could be made as a separate unit complete with slide hume, camus rockers and gears. All units would be identical or mirrored and would fit into a suitable gearbox fiame at the required spacing. The sliders would be ananged across the gearbox to enable the sliding mechanism to be positioned in line with its connecting rod.

There are many variations to the gear iayouts that the principe of synchronised gears lends itself to. For instance it is possible to arrange for the inward stroke rack gears 21 to drive their respective gears 23 in the same direction of rotation as the outward stroke rack gears drive their fixed gears 23. This can be achieved by inverting the rack gears on one side of each slide and driving the other side of fixed gears 23. In this event only two output gears 24 would be required thus eliminating gears 24b and 27 in figure 1 and therefore avoiding this additional backlash in the gear train. In figure 2 this would be achieved by modifying the rack and cam arrangement shown at slides 1/9 and 8/16 to accomodate the gear 23 sizes in the reduced space and also realigning the gear 23 at the other slides. This is better illustrated by reference to figure 9 which shows a more compact gear train with the slides spaced around a 157.5 are of rotation. For this arrangement the double shaft design would still be required for gears 23 to enable gears for slides 9 to 16 to be adjusted by the helical gear to suit the changing ratio synchronisation requirements. An additional advantage of the common rotation arrangement is that the relative rotation of the gear 23 double shaft is only a few degrees in either direction.

Due to the interaction of adjacent slides with gears 23 in figure 9 the gear racks would also have to vary from figure 3 arrangement to enable adjacent siides to accommodate their connecting rods at the optimum spacings. The rod sequence across the crankpin would vary from figure 7 and would number from 1 to 8 staight across the crankpin. This would also apphJr if figure 2 adopted the modified outlet gear arrangement.

The shafts carrying gears 23 a to q terminate in output gears meshing into the final output sears 24. The size of these gears can be optimised to suit the particular arrangements chosen and it is possible to reduce the output gears 24 diameter by this means. In optimising output gear 24 size the 22. 5 repeat of slide spacing should be accommodated with exact pitching of gear teeth for ease of rack gear initial setting up.

The following should be noted. All the drawings making up figures 1 to 9 are semi diawnatic axld show the principe only. Exact sizing will be dependent on many design factors. As the crankpin approaches the centrieline of the discs and output rotation faUs to zero the gear ratio rises huards s infinity and the torque required falls towards zero. Soft starts are possibly best suited to the system and under some conditions an automatic clutch miaht be required on start up. Deveiopment will indicate the requirements.

Claims (7)

1. CLAIMS I A method of using a single crank pin and multiple connecting rods, on the radial engine principe, to provide a synchronised gear drive for an infinitely variable ratio gearbox by varying the rotation diametrical position of the crank pin between the driving discs.
2. 2 An arrangement of sliding gear racks and multiple connecting rods, as claimed in claim 1, which engage synchronously and successively with fixed position gears.
3. 3 An arrangement of synchronised gearing, as claimed in claims 1 and 2, with controlled helicai, or other, adjustment which automaticaiiy sets synchronous gear teeth positions to suit the requirements of infinitely adjusting the gear ratio output.
4. 4 An infinitely variable speed gearbox, as claimed in claims 1, 2 and 3, capable of giving an output speed of 0 to full speed in both forward and reverse directions.
5. 5 An infinitely variable speed gearbox, as claimed in claims 1, 2 and 3, capable capable operating between set ratio requirements if needed.
6. 6 A gearbox, as claimed in claims 1, 2 and 3, capable of being used with a constant speed prime mover leading to a more environmentally acceptable economic operation.
7. 7 A gearbox, as claimed in claims 1, 2 and 3, capable of giving any required gear ratios to suit varying load and driving conditions.

   Amendments to the claims have been filed as follows CLAIMS 1 An arrangement of variable position single throw crankpin combined with multiple connecting rods driving rack gears arrange to engage successively and synchronously with fixed gears, driving outlet gears, to provide an infinitely variable ratio gear drive dependent on crankpin position- 2 An arrangement of synchronised gearing and multiple connecting rods, as claimed in claim 1, using both the forward and backward strokes of the rack gearing to provide the drive to outlet gears which have had their relative angular positions adjusted by a helically splined sleeve, or other means, to compensate for relative changes of gear position resulting from crankpin throw movement.