GB2464479A - A gyroscopic transmission for variably transferring torque - Google Patents

Abstract

A gyroscopic transmission comprises a flywheel 18 rotatably mounted to an inner gimbal 20 and coupled to an electric motor that is controllable in order to control a speed of rotation of the flywheel. The inner gimbal 20 is rotatably mounted to an outer gimbal 22 which is driven by an input shaft 12. Bevel gear 24 is rigidly attached the inner gimbal 20 and meshes with bevel gear 26 fixed to an output shaft 14 that has an axis which is coincident with an axis of the input shaft 12. In operation sinusoidal torque is applied to the output shaft 14 and as a consequence a torque rectifying unit 40 is provided. The torque rectifying unit 40 comprises a pair of one-way clutches 44, 45 each having gears 43, 41 which mesh with gears 46, 48 and 50 so as to form a gear train. A multi-flywheel arrangement having multiple bevel gears may be provided (fig 5) and the transmission may be used in an AC rail traction drive with regenerative braking (fig 6). In another embodiment a rocking to and fro type transmission (fig 7) for use in, for example, earth moving vehicles or fork lifts is provided.

Description

A Transmission The present invention relates to a transmission for coupling an input shaft to an output shaft and in particular, although not exclusively, to a transmission for use in an automobile.

Transmission systems are widely used in engineering and other applications. The automotive industry is perhaps the biggest market for transmission systems.

Accordingly, whilst the invention as herein described is relative to automotive transmission systems, it will be appreciated that it is equally applicable to any torque convertor.

The need for a transmission in an automobile is a consequence of the characteristics of the internal combustion engine. Combustion engines typically operate over a range of around 600 to 7000 rpm, while the car's wheels rotate between 0 and 1800 rpm. Thus a transmission is required to match the speed of the engine shaft to the speed of the wheel's axle.

Fundamentally, an automotive transmission is required to adjust to a range of output to input velocity ratios.

The most basic transmission system comprises a gearbox having a set of gears. As sets of gears have fixed ratios, the transmission is subdivided into a sequence of set ratios requiring disruption of the transmission at each change of ratio. This disruption involves disconnecting the transmission from the engine by means of a manually operated clutch. A clutch is used to select different gear ratios from the set of gears. Each ratio selection requires two clutch operations, a gear shift, and engine power relaxation during the change. Moreover, because combustion engines provide their highest torque outputs approximately in the middle of the range of rpm and because gears have fixed ratios, to sustain crawl speeds there is no option but to slip the clutch to avert engine stall. The manual operations required of such a basic system are tolerable provided the frequency of operation is not too high. However, the density of modern traffic often culminates in crawl driving conditions with frequent stop/starts placing an unacceptably high work load on the driver. This situation is exacerbated whilst ascending an incline by the additional frequent application of the brakes to avoid rolling backwards.

To avert the tedium of manual operation of a basic gear box, the automatic transmission has been introduced.

The problems of providing both high torque and preventing engine stall at low road speeds is solved by placing a hydrodynamic torque converter between the engine and gear box. The change of gear ratios is automated by planetary gears engaged by band clutches. All of this adds considerable complexity and cost. Power is lost by fluid churning in the torque convertor and supplementary cooling is necessary to avoid overheating. In addition to the disruption of transmission, the fixed ratio gear box, irrespective of manual or automatic operation, may not allow the engine to operate at its most favourable conditions.

Consequently, in an endeavour to solve these issues and to provide vehicles with greater fuel economy and reduced exhaust emissions, there has been considerable effort devoted to developing Continuously Variable Transmissions (CVT5) that enable the combustion engine to operate closer to its most economical conditions irrespective of road speeds. Known CVT's are based on friction drives. These take two basic forms: one based on pulleys and the other on rollers.

Pulley based CVT's connect the input shaft to the output shaft via a high-strength metal belt. The belt connects a pair of variable-diameter pulleys, which are connected to each of the input and output shafts respectfully. Generally the variable-diameter pulleys comprise two 20° cones facing each other. When the cones are moved apart, the belt rides lower towards the central axis. When the cones are moved together, the belt rides higher, increasing the diameter. When one pulley increases its radius, the other decreases its radius in order to keep the belt tight. As the two pulleys change their radius relative to one another, they create a continuous variation of velocity ratio between defined limits. It will be appreciated that the transmission thereby connects the input shaft to the output via friction between the belt and pulleys.

An alternative to the pulley based CVI is a toroidal CVT. Here, the input and output shafts include a pair of opposed, conical discs. One conical disc is arranged on each shaft. The conical discs are arranged point-point and are connected by rotating rollers. The rollers are also tiltable so that they contact the conical discs in different positions. For example, the rollers can be tilted so that they contact the conical disc on the input shaft towards the smallest diameter and the conical disc on the output shaft towards the largest diameter. Thus a "low gear" is selected.

As the toroidal CVI drive uses curved rollers, the contact area between the rollers is virtually reduced to a point resulting in high bearing pressures. This means the parts are prone to wear and. slippage caused by friction.

The friction drives, like gear transmissions, derive velocity ratio from the ratios of the radii of two rolling circles. This imposes a practical limitation on the minimum velocity ratio attainable.

To extend the velocity ration down to zero output, friction drives have been incorporated with fluid torque convertors which generate torque by change of fluid momentum.

To extend the velocity ratio into the negative domain to achieve reverse, the output drive is taken from a different gear train placed. across the friction drive.

As an alternative to friction drives [-lydrostatic transmissions transfer power by positive displacement of virtually incompressible fluid by using a variable displacement pump to create a continuous variation in velocity ratio from positive to negtive. Such hydrornechanical' transmissions are prohibitively expensive and inefficient for pedestrian automotive application but are suited to heavy-duty track laying vehicles for differential steering and where constant changes from forward to reverse are necessary as in earth-moving equipment.

It is an object of the present invention to attempt to overcome at least one of the above or other disadvantages.

It is a further aim of the present invention to provide a continuously variable transmission.

According to the present invention there is provided a transmission, a method and a vehicle as set forth in the appended claims. Other features of the invention will be apparent from the dependent claims, and the description which follows.

In one aspect of the present invention there is provided a transmission for variably transferring torque from an input shaft to an output shaft wherein the torque on the output shaft is induced by a change in angular momentum.

Preferably the transmission varies the torque in order to couple the output shaft over a range of output speeds.

The range may be from zero percent of the input speed.

The transmission may be continuously variable between a range of output speeds. The range maximum maybe commensurate with the input speed.

The transmission may comprise at least one rotating mass. The mass rotates about a first axis. The first axis may be rotatably mounted about a second axis. The second axis may be rotatably mounted about a third axis.

The first, second and third axes may be perpendicular.

The first, second and third axes intersect at a common point. Suitably the third axis is the axis of the input shaft. The rotating mass may induce an internal torque about the second axis. A drive means may connect rotation of the second axis to rotation of the output shaft.

The drive means may comprise a first gear and a second gear which may be bevel gears. The first gear may rotate about the second axis. The second gear may rotate about the third axis. The first and second gears may be arranged at right angles. The second gear may be fixed fast to the output shaft.

The transmission may include a torque rectifying unit.

The second bevel gear may be fixed to an alternator shaft of the torque rectifying unit. The alternator shaft may be coupled to the output shaft by at least a first one-way clutch. Preferably the alternator shaft is coupled to the output shaft by a first one-way clutch and a second one-way clutch. The first and second one-way clutches may be opposed to each other. One of the one-way clutches may couple to the output shaft via an intermediate gear.

Suitably, the or each one-way clutch may comprise a sprag clutch.

The transmission may comprise a plurality of rotating masses. Each rotating mass may be arranged to rotate about a first axis. Each first axis may form a tangent of a circumference having a centre coincident with a third axis. Each first axis may be rotatable about a second axis. Each second axis may be perpendicular to each first axis. Accordingly, each second axis may form a radius of the circumference. Each second axis may be rotatable about the third axis. A drive means may couple rotation of each second axis to rotation of the output shaft.

Suitably, the drive means comprises at least a first drive train. The first drive train may comprise a plurality of first gears and a second gear which may be bevel gears. Each first gear may be arranged at right angles to the second gear. Each first gear may rotate about each second axis. The second gear may rotate about the third axis and may be connected to the output shaft.

Preferably, the drive means comprises a second drive train. Suitably the first drive train and second drive train include a first and second one-way clutch respectfully. The first and second one-way clutches may be arranged in opposite directions. Suitably the one-way clutches are sprag clutches. The second drive train may comprise a plurality of third gear and a fourth gear which may be bevel gears. Each third gear may rotate about each second axis. Each third gear may be arranged at right angles to the fourth gear. The fourth gear may rotate about the third axis. The fourth gear may drive the output shaft via intermediate gearing.

Preferably the or each rotating mass comprises a rotating flywheel. The speed of rotation of the or each rotating mass may be controllable. Suitably an electric motor rotates the or each rotating mass. Controlling the speed of rotation of the or each mass may control the torque applied to the output shaft. The torque applied may be applied irrespective of output speed. The torque applied may be applied within prescribed limits. The prescribed minimum may be zero.

In a further aspect of the present invention there is provided a method of driving an output shaft, the method comprising causing an input shaft to rotate, and inducing a change in angular momentum in order to variably transfer torque from the input shaft to the output shaft.

Change in angular momentum may be induced by controlling the speed of rotation of at least one rotating mass. The method may comprise controlling the rotating speed of a plurality of rotating masses. Preferably, the speed of the rotating masses are controllable individually.

The method may comprise the change of angular momentum for a given speed of rotation of the mass imparting a varying degree of torque from the input shaft to the output shaft. That varying degree may comprise a cylclical variation which may be from a positive to a negative degree, for instance in a sine wave. The method may comprise converting the negative phase of the cycle to a positive degree. When a plurality of masses are rotating, the method may comprise coordinating the varying degrees of torque from each mass at a given rate of rotation whereby, for instance, when the torque from one mass is rising the torque from another mass is falling.

Preferably, the method comprises using a transmission according to the previous aspects.

In a further aspect of the present invention, there is provided a vehicle comprising at least one transmission.

The transmission is in accordance with previous aspects.

In one exemplary embodiment, the vehicle is a road vehicle.

In another exemplary embodiment the vehicle comprises at least two transmissions according to previous aspects, both transmissions driving a common output in order to provide a push-pull functionality.

In another exemplary embodiment, the vehicle operates using an AC power supply, to drive the input shaft.

Various embodiments will now be described, by way of example, and with reference to the following drawings, in which: Figure 1 is a side view of a transmission.

Figure 2 is a top view of the transmission.

Figure 3 is a pictorial representing showing the forces acting on a flywheel.

Figure 4 is a perspective view of a rectified transmission.

Figure 5 is a side, partial cross-sectional view of a multiple flywheel transmission.

Figure 6 is a schematic drawing showing an AC rail tractions drive.

Figure 7 is a perspective view of a push-pull transmission.

As shown in Figure 1, a transmission 10 comprises an input shaft 12 and an alternator shaft 14 that are arranged along a common axis X-X. Power is transmitted between the two shafts by generating an internal torque.

Output torque on the alternator shaft 14 is a consequence of this internal torque reacting against the input torque.

The internal torque is generated using gyroscopic principles. Accordingly a gyroscopic arrangement 16 couples the two shafts.

The gyroscopic arrangement 16 comprises a flywheel 18 rotatably mounted about its central axis W-W to an inner gimbal 20. The flywheel 18 is coupled to an electric motor (not shown) that is controllable in order to control the speed of the flywheel rotation Q. The inner gimbal 20 is rotatably mounted to an outer gimbal 22 along axis Y-Y, which is perpendicular to the axis W-W of the flywheel.

The outer gimbal 22 is formed from a forked yoke at the end of the input shaft 12. A bevel gear 24 is rigidly attached to the inner gimbal 20. The bevel gear 24 meshes with a second bevel gear 26 centered and rigidly attached to the alternator shaft 14. The first and second bevel gears are arranged at right angles to each other. As such they translate rotation of the first bevel gear 24 about axis Y-Y and/or movement of the first bevel gear about a circular path around axis X-X into rotation of the alternator shaft 14.

In use, the input shaft from the engine rotates thereby rotating outer gimbal 22 about the X-X axis. The rotation of the outer gimbal spins the inner gimbal 20 about axis Y-Y. Thus the first bevel gear 24 is caused to move along a circular path about axis X-X. If an internal torque is applied to the inner gimbal 20 such that the inner gimbal 20 is effectively locked fast (i.e. it does not rotate about axis Y-Y), the first bevel gear will drive the second bevel gear and the alternator shaft speed will equal the input shaft speed. If no internal torque is generated, and assuming a minimal torque resisting rotation is applied to the alternator shaft sufficient to overcome the friction of the inner gimbal resisting rotation about the Y-Y axis, as the outer gimbal is spun, the inner gimbal rotates freely about axis Y-Y.

Accordingly, no torque is generated on the second bevel gear 26. Thus there is no output torque. Accordingly, it will be appreciated that adjusting the internal torque between these two limits allows the output speed to be adjusted between zero and 100% of the input shaft speed.

As discussed in more detail below, rotation of the flywheel generates the internal torque. With the flywheel at rest, no internal torque is generated. By accelerating the flywheel from rest, the internal torque is induced that acts to rotate inner gimbal 20 about the Y-Y axis.

Output torque on the alternator shaft is a consequence of this internal torque reacting against the input torque to turn the second bevel gear 26.

Referring to Figures 2 and 3, the generation of the internal torque is described in more detail. The torque induced on the Y-Y axis is the product of the flywheel's moment of inertia I and spin rate Q about the W-W axis and rotation Ct) about an axis transverse to W-W. How this is achieved and especially how the sense of direction of the induced torque is determined can be inferred from Figure 3. As shown by pictorial arrows 30, 31, on each of two identical mass particles situated on the X-X axis at the same radius either side of the W-W axis, the flywheel anticlockwise angular velocity Q results in a tangential velocity V. On one side of the W-W axis V acts downwardly, whereas on the other side V acts upwardly.

When a clockwise rotation Wx is applied to the flywheel about the X-X axis, each mass particle experiences a lateral acceleration (i.e. at right angles to the X-X axis) equal to the product of V and Wx which induces a force F on each of the particles parallel to the W-W axis.

Since the Wx rotation is common to both particles but the velocities V of the two particles are opposite, the induced forces F will be opposed. In having equal magnitude and equidistance from the Y-Y axis but by acting in opposite directions it can be seen that the two forces F combine to induce a clockwise torque about the Y-Y axis.

The flywheel is shown in Figure 3 with axis W-W inclined to axis X-X by angle e of 9Q0* Here the flywheel lateral rotation rate becomes that of the clockwise input shaft rotation rate thereby inducing peak clockwise torque on the Y-Y axis. When axis W-W is inclined to axis X-X by an angle e of Q0 or 180°, the flywheel is isolated from the input shaft rotation. As the flywheel lateral rotation rate is now zero, the induced torque on the Y-Y axis is also zero. When the angle of rotation e is 270°, the flywheel will be rotating in the opposite direction to that of 90 degrees i.e. Q is now clockwise, thereby reversing the sense of the torque induced on the Y-Y axis to that at 90 degrees. It will therefore be appreciated that the induced internal torque follows a sinusoidal pattern. Moreover, it follows that the induced torque about axis Y-Y generated by the change of angular momentum by rotation Wx is a sinusoidal product of the flywheels inertia I, spin rate Q and rotation Wx.

Due to the sinusoidal torque applied to the alternator shaft 14, it is necessary (in most applications) to rectify the torque in order to achieve infinite flexibility between the input and output speeds.

Accordingly, and with reference to Figure 4, a torque rectifying unit 40 couples the alternator shaft 14 to an output shaft 42. Here, the torque rectifying unit 40 is shown suitably as a pair of one way sprag clutches 44, 45 each connected to a gear train. One of the one way sprag clutches 44 controls a gear 43 that is connected directly to a gear 46 formed fast to the output shaft 42.

Consequently, when the alternator shaft 14 is turned in one direction, the first sprag clutch 44 engages the drive train and turns the output shaft and when the alternator shaft turns in the opposite direction, the sprag clutch 44 slips and therefore does not apply a torque through the drive train to the output shaft 42. The other sprag clutch 45 is arranged on the alternator shaft to transmit and slip in the opposite direction to the first sprag clutch 44. Accordingly, sprag clutch 45 drives output shaft 42 via a drive train comprising gear 41, an intermediate gear 48 and gear 50 that it is fixed to the output shaft 42. Thus the torque rectifying unit rectifies the torque transferred to the output shaft as a uni-directional, rectified sine wave.

For instance, whilst the Y-Y axis is rotating through 0 to 180°, the induced internal torque on the Y-Y axis will be negative in being opposed to the Y-Y axis rotation. This torque will appear as a positive torque on the alternator shaft 14 and thence via one-way clutch 45, intermediate gear 48 and gear 50, as positive torque on output shaft 42. After the Y-Y axis rotates through 1800, the induced torque acting on bevel gear 24 will become positive to appear as a negative torque on alternator shaft 14. The reversal of torque on alternative shaft 14 will cause one-way clutch 45 to disengage and one-way clutch 44 to engage thereby enabling the negative torque on alternator shaft 14 to be transmitted via gear 46 to also appear as positive torque on the output shaft 42.

As positive torque on the output shaft 42 induces positive output rotation, the alternator shaft 14, in response to the torque it is transmitting, will alternate in rotation rate between the positive and negative values of the rate of rotation of output shaft 42. The effect of this alternation in alternator shaft 14 is to cause the rotation rate of inner gimbal 20 about the Y-Y axis to fluctuate between two positive levels. When the torque on the Y-Y axis is negative, inner gimbal rotation rate is the difference in the rates of input shaft 12 and output shaft 42. When the torque on Y-Y Axis is positive, inner gimbal 20 rotation rate is the sum of the input and output rates. Thus the rotation rate about the Y-Y axis is greater when the alternating torque is positive than when the torque is negative, a pre-requisite for power transmission through the coupling.

As the output speed rises with fixed input speed, the Y-Y axis rotation rate progressively reduces. The Y-Y axis rotation rate cannot become negative as this would initiate an immediate switch into positive torque mode.

Consequently, the maximum attainable output speed corresponds to zero Y-Y axis rotation. For a system where all the gear ratio are unity, the maximum output speed will be equal to the input speed. This represents a steady state condition for the Y-Y axis where the angle e remains constant at a fixed angle commensurate with the torque load on the output shaft.

At the beginning of each half cycle, the rotation rate of the Y-Y Axis has to change. At the start of the cycle the rate must drop from the sum of the coupling input and output rotation rates to the difference between input and output rates. At the beginning of the second half, the rotation rate must be restored to the greater value.

These changes in speed proceed naturally as they are driven by the prevailing induced torque on the Y-Y Axis.

During these transitional speed changes, the speed of the alternator shaft 14 lies between plus and minus the output shaft speed and is thus temporarily disconnected from the output shaft 42. This disconnection reduces the mean transmitted power which can be compensated for by raising the speed of the flywheel 18 at the expense of raising the peak transmitted torque.

As described above, the torque output from a single flywheel is a fully rectified sine wave alternating between zero and a maximum. For applications demanding less torque fluctuation, the outputs form an odd number of flywheel assemblies can be combined to reduce torque fluctuation by phase interlacing. Figure 5 shows a second embodiment wherein the end of the input shaft 12 comprises a housing plate 60 for supporting multiple, evenly spaced inner gimbals 20 about a circumference. Each inner gimbal is rotatably mounted to the housing plate 60 about an axis Y-Y that forms a radius of the circumference.

Flywheels 18 are rotatably mounted to each inner gimbal about an axis w-w that is perpendicular to the Y axis.

As phase separation isolates the operating cycle of each inner gimbal 20, the multiple inner gimbals are not able to share a common rectification stage. Accordingly, the one-way sprag clutches have been transferred to the axle of the inner gimbal 20 thereby eliminating the alternator shaft. Thus sprag clutch 44 is connected to the output shaft 42 by the first bevel gear 24 and second bevel gear 26. Bevel gear 26 being fixed fast to the output shaft 42. Sprag clutch 45 is connected to the output shaft 42 by a third bevel gear 62, fourth bevel gear 64, intermediate gears 66, 68 and 70, and gear 72 that is fixed to the output shaft 42. Bevel gear 64 is rotatably mounted about the output shaft, which is coincident with the circumference's centre. Thus bevel gears 26 and 64 serve to connect all the inner gimbals 20 to the output shaft 42.

During operation of the multi-flywheel transmission, it is essential to maintain a uniform phase separation between each flywheel cycle in order to sustain minimum torque fluctuation. The duration of sprag clutch disengagement at the start of each half cycle of the Y axis is largely dependent upon the flywheel rotation rate and the inertia of the inner gimbal. Any disparity in these factors between member flywheels is a potential source of phase drift. It is therefore necessary to monitor the phase relationship with respect to an elected reference member and apply any correction by adjusting the flywheel speed.

The transmission as herein described provides the versatility to regulate the torque transmission characteristic through adjustment of the flywheel rotation rate. As an illustration of how this works, consider an acceleration of a vehicle from rest with the engine driving the input shaft such that it is governed to run at a constant speed and a limit placed on its maximum power output.

The output torque from the coupling is proportional to the product of the input speed and the flywheel speed and similarly the reaction torque on the input shaft is proportional to the product of output speed and flywheel rotation rate. As the output speed increases, the torque loading on the input shaft will increase proportionately and with it the power demand on the input shaft until the power output limit is attained. During this phase, as both the input shaft and flywheel speed are constant, the output torque will remain constant.

To prevent exceeding the set power limit, the flywheel speed must be reduced inversely with the output speed.

Consequently, as the output speed rises, the output torque will fall with reduction in flywheel speed thereby sustaining a constant power output.

It will be appreciated that the transmissions described herein are restricted to one way rotation. For automobiles and other situations where reverse is only required for relatively infrequent selections, a conventional gear box selection for reverse is sufficient.

For example, in Figure 4 a reverse drive can be taken from gear 48 and in Figure 5 from gear 70 via shaft 82.

A further application of the transmission is in AC rail traction. Here alternating current has the advantage over direct current in having a more amenable power distribution system and motors that are less costly to build and maintain. Existing transmissions do not adequately adapt the AC motor to a wide range of speeds.

However, the transmissions herein described enable the AC motor to run at a constant speed in synchronism with the AC power supply frequency. Thus the electrical supply to the motor is relived of a control function and the tractive effort is simply determined by the gyration rate of the coupling flywheel. Similarly, regenerative breaking can be facilitated by applying a second transmission acting in the reverse sense.

Figure 6 shows a schematic layout for an AC rail traction drive with regenerative braking. The forward drive 80 transmission is essentially similar to the transmission herein described with the shaft 82 fixed to intermediate gear 70 extended to provide a reverse axle and shaft 42 providing a forward axle shaft 42. Both axles terminate in a conventional forward/reverse gear box 83. The forward/reverse gearbox 83 allows forward or reverse to be selected by coupling the respective shaft to an output drive shaft 84.

For regenerative braking, a second transmission 85 has been added driven from an extension of output shaft 42 of the forward drive transmission 80. A limitation of the transmissions 80, 85 is that output speed is limited by the input speed. Thus, to sustain regenerative braking as the train reduces speed, the output from the second transmission 85 must be increased by a step up gear 86.

One-way clutches within the second transmission 85 prevent drive back through the step-up gearing. An additional one-way clutch 87 at the step-up connection to a motor/generator shaft 88 allows that shaft to overrun the step-up gears in forward drive to avert their drag.

With the above arrangement, the electrical supply to the motor/generator 90 is relived of all complication, requiring only an on/off switch 89 to the supply. Because the motor 90 runs at constant speed, the tractive effort is proportional to rotation rate of the forward drive driver's flywheel (or wheels) . The torque on the synchronous motor is the product the flywheel speed and train speed so that the flywheel speed may have to be reduced as train speed increases to keep the power demand on the electrical supply within specified limits. A similar situation applies during regenerative braking.

A control box 91 controls the flywheels and is selectable between brake, drive and neutral modes. In selecting neutral mode, the flywheels of both couplings would be at rest enabling the traction motor 90 to spin freely at synchronous speed.

A further application of the transmission is in vehicles that are not expected to travel at high speed nor descend long inclines but sustain frequent "to and fro" movements such as earth moving vehicles and forklift trucks.

Figure 7 shows a push-pull transmission for such applications. Here two side by side transmissions lOa and lOb provide continuously available change in direction of drive. The two couplings are conjoined by sharing a common output shaft 42. This necessitates reversing the sense of the one-way clutches on one coupling from those on the other. Thus gear train 43b, 46 and 43a connects the alternator shaft of the upper coupling by one-way clutch 44a and the alternator shaft of the lower coupling by one-way clutch 44b.

The consequence of the transmissions herein described is that the output torque is proportional to the rotation rate of the flywheel. This endows the transmission with an additional degree of freedom of operation conferring advantages not found in other forms of transmission.

For instance, irrespective of the input shaft speed, there can be no output torque when the flywheel is at rest. This enables the engine to be decoupled from the transmission without need of a separate clutch. By accelerating the flywheel from rest with the engine speed held constant and output shaft stationary, the transmission is thus able to produce a corresponding progressive increase in output torque from zero up to a maximum commensurate with the engine speed. Thus the engine idling and stall prevention requirements for automotive application can be handled by simply regulating the flywheel speed without recourse to supplementary clutches or differential gear trains, etc. Furthermore, the torque output from hydrodynamic torque convertors is roughly proportional to the difference in speed between input and output. This necessitates a difference in speed in order to transmit torque, which results in power loss. Consequently, single stage hydrodynamic torque convertors with a fixed stator is limited to a maximum torque gain of the order of 2.5.

Advantageously, because in the transmission herein described, the output torque is proportional to the product of the speeds of the input and flywheel, the transmission has no fundamental limitation on the maximum gain attainable.

Torque transmission by angular momentum affords a further advantage over variable gear ratio transmission by being simpler in operation. For instance, the transmission can sustain acceleration under constant output torque by simply maintaining both engine and flywheel speeds constant. In this situation, the gear ratio of an infinitely variable gear ratio transmission must be continuously readjusted as output speed increases.

The transmission, in contrast to gearbox transmissions, can be said to be a "soft coupling" in the following sense. If the output were suddenly jammed then, in the case of a gear transmission, all parts of the transmission and prime mover would come to a shock stand still. With the present transmission, since the torque reaction on the prime mover is proportional to output speed, the effect would be to take the load off the prime mover, thereby averting any shock loading.

There are practical limits to maximum attainable velocity Ratio in contemporary CVT friction drives which prevent zero output with the engine running. This limitation has been circumvented by coupling a differential gear across the CVI so that there is no output from the differential when the CVI velocity ratio is unity. Moreover, the attainment of an overall velocity ratio of zero implies a mechanical advantage of infinity capable of generating excessive torques. This has been known to cause problems in 4-wheel drive vehicles. For instance, an axle developing slightly greater tractive effort will drive the other axle resulting in a torque feed back windup in the transmission system culminating in failure of either a differential or half-shaft. Solutions include linking one axle via a one-way clutch to an under driven shaft so that it is only engaged if the normally driven wheels turn slightly faster as a consequence of slippage on soft terrain.

Advantageously, in the present coupling, angular momentum is vested in the rotation of the flywheel. Thus with the flywheel at rest there can be no torque output irrespective of the input rotational speed thus obviating the need for either a decoupling input clutch or output differential gear appendage.

Attention is directed to all papers and documents which are filed concurrently with or previous to this specification in connection with this application and which are open to public inspection with this specification, and the contents of all such papers and documents are incorporated herein by reference.

All of the features disclosed in this specification (including any accompanying claims, abstract and drawings), and/or all of the steps of any method or process so disclosed, may be combined in any combination, except combinations where at least some of such features and/or steps are mutually exclusive.

Each feature disclosed in this specification

(including any accompanying claims, abstract and drawings) may be replaced by alternative features serving the same, equivalent or similar purpose, unless expressly stated otherwise. Thus, unless expressly stated otherwise, each feature disclosed is one example only of a generic series of equivalent or similar features.

The invention is not restricted to the details of the foregoing embodiment(s). The invention extends to any novel one, or any novel combination, of the features

disclosed in this specification (including any

accompanying claims, abstract and drawings), or to any novel one, or any novel combination, of the steps of any method or process so disclosed.

Claims (13)

1. Claims 1. A transmission for variably transferring torque from an input shaft to an output shaft wherein the torque on the output shaft is induced by a change in angular momentum.
2. 2. The transmission as claimed in claim 1, wherein the change in angular momentum is induced by a first gyroscopic arrangement.
3. 3. The transmission as claimed in claim 1, wherein the change in angular momentum is induced by a plurality of gyroscopic arrangements.
4. 4. The transmission as claimed in claim 2 or claim 3, wherein each gyroscopic arrangement comprises at least one mass that is rotatable about a first axis wherein the first axis is rotatable about a second axis and the second axis is rotatable about a third axis that is coincident with an axis of rotation of the input shaft.
5. 5. The transmission as claimed in claim 4, wherein a drive means connects rotation of each second axis to rotation of the output shaft.
6. 6. The transmission as claimed in claim 5, wherein the drive means comprises a first gear and a second gear, the first gear being connected to the rotation of the second axis and the second gear being connected to rotation of the output shaft.
7. 7. The transmission as claimed in any preceding claim, wherein each gyroscopic arrangement is coupled to the output shaft via a rectifying unit.
8. 8. The transmission as claimed in claim 7, wherein each rectifying unit comprises a first one-way clutch and a second one-way clutch, the first and second one-way clutches being opposed to each other and the second one-way clutch being arranged to drive an intermediate gear in order to rectify the rotation of the drive output shaft.
9. 9. The transmission as claimed in claim 4, wherein the speed of rotation of each rotating mass is controllable in order to control the torque on the output shaft irrespective of output speed.
10. 10. A method of driving an output shaft that is coupled to an input shaft via a transmission, the method comprising causing the input shaft to rotate, and inducing a change in angular momentum within the transmission in order to variably transfer torque from the input shaft to the output shaft.
11. 11. The method as claimed in claim 10, wherein the method comprises using a transmission as claimed in any of claims 1 to 9.
12. 12. A vehicle comprising at least one transmission as claimed in any of claims 1 to 9.
13. 13. A transmission, method or vehicle substantially as herein described with reference to the drawings.