WO2016193027A1

WO2016193027A1 - Gear ratio shift management

Info

Publication number: WO2016193027A1
Application number: PCT/EP2016/061467
Authority: WO
Inventors: Robert William Thompson
Original assignee: Qinetiq Ltd
Current assignee: Qinetiq Ltd
Priority date: 2015-06-03
Filing date: 2016-05-20
Publication date: 2016-12-08
Anticipated expiration: 2017-12-03
Also published as: GB201719737D0; GB2558077B; HK1257625A1; GB201509619D0; GB2558077A

Abstract

A method comprising the steps of: in response to determining that a gear ratio shift is to be initiated while a vehicle is in motion reducing pressure exerted between a first rotating body and a second rotating body, wherein the first rotating body is rotatably driven by a drive mechanism, and wherein the second rotating body is caused to spin due to friction between it and the first rotating body and comprises part of a linkage for transferring torque to an output portion for propelling the vehicle; implementing a gear ratio shift in a gear arrangement of the linkage; and restoring the pressure exerted between the first rotating body and the second rotating body.

Description

Gear Ratio Shift Management

Field

The present application relates to the management of gear ratio shifts in a vehicle, such as a car, while it is in motion.

Background

WO2014/049317 describes a Q-shift gearbox, which is clutchless in the sense that once a car having such a gearbox is in motion, different gear ratios can be selected without dis-engaging a clutch. This enables gear ratio changes to be carried out without loss of driving power. In other words, during a gear ratio change some torque is always transferred between the car engine and the wheels.

Tests with this type of gearbox have shown that at low engine rpm, gear change shock is more noticeable to car passengers than at higher rpm. Gear change shock refers to an impulse imparted through a car's driveline when a gear ratio change occurs, wherein a driveline collectively refers to components

mechanically coupling the car engine to the wheels for transferring engine torque to the wheels.

Following an upshift in gear ratio, for example a change from third to fourth gear, engine speed is required to be reduced to maintain constant wheel speed. However following a downshift in gear ratio, for example a change from fourth to third gear, engine speed is required to be increased to maintain wheel speed. Since the Q-shift gearbox implements gear ratio changes without using a clutch, a step change in car engine speed occurs when a gear ratio shift happens. This causes the aforementioned impulse (also known as a torque spike or torque pulse) to be generated along the driveline which passengers can feel at certain engine rpms, so reducing the quality of their drive experience. These impulses are also likely to reduce the life and reliability of components in the driveline.

Aspects of the present invention are conceived to address the foregoing. Summary

According to an aspect of the present invention there is provided a method comprising the steps of: in response to determining that a gear ratio shift is to be initiated while a vehicle is in motion reducing pressure exerted between a first rotating body and a second rotating body, wherein the first rotating body is rotatably driven by a drive mechanism, and wherein the second rotating body is caused to spin due to friction between it and the first rotating body and comprises part of a linkage for transferring torque to an output portion for propelling the vehicle; implementing a gear ratio shift in a gear arrangement of the linkage; and restoring the pressure exerted between the first rotating body and the second rotating body. The pressure exerted between the first rotating body and the second rotating body may be reduced until they are determined to be slipping against each other based on output from rotational speed sensors.

The method may further comprise the steps of: after the gear ratio shift adjusting the relative difference between the rotational speeds of the first and second rotating bodies such that it complies with a predetermined condition using rotational speed sensors; and then increasing the pressure exerted between the first and second rotating bodies such that a subsequent change in the rotational speed of one of them results in a corresponding change in the rotational speed of the other.

The step of adjusting the relative difference between the rotational speeds of the first and second rotating bodies may involve changing at least one operational parameter of the drive mechanism.

The drive mechanism may comprise an internal combustion engine and the at least one operational parameter may be selected from the list of: throttle position, fuel/air ratio per ignition cycle, fuel injection rate, ignition timing and variable valve timing.

The drive mechanism may comprise an electric motor and the at least one operational parameter may be selected from the list of: voltage or current supplied to a driven component of the electric motor; and voltage or current supplied to a stator of the electric motor. The step of adjusting the relative difference between the rotational speeds of the first and second rotating bodies may involve varying the pressure exerted between them and thereby the friction between them. The pressure may be varied such that the rate of change of rotational speed of the first rotating body is controlled to substantially comply with another predetermined condition.

The drive mechanism may be operating at substantially maximum torque output.

The or each said predetermined condition may be dependent on a mode of operation of the vehicle.

The mode of operation may be selected by a user, or may be determined on-the- fly based on output received from at least one sensor on board the vehicle indicative of how the vehicle is being operated.

The output received may be indicative of a usage property of part of the drive mechanism, or may be indicative of the magnitude of vehicle acceleration.

The usage property may be the throttle position and/or engine revs.

The first and second rotating bodies may be determined to be slipping against each other if their rotational speeds differ from each other by at least a threshold amount.

The first and second rotating bodies may be determined to be slipping against each other if the rates of change of their rotational speeds differ from each other by at least a threshold amount.

The gear arrangement may be configured such that during the gear ratio shift the amount of torque transferred between an input shaft and an output shaft thereof does not reduce to zero. At least one gear may be provided on either the input shaft or the output shaft and wherein two selector members may be associated with the or each gear respectively, the or each gear comprising a first and second face, and further comprising, on each face, at least one projection, and the or each selector member comprising, on at least one face thereof, at least one complementary projection arranged to selectively engage with a projection of a gear, the arrangement being such that the projection(s) of a selector member and the projection(s) of the first face of a gear may be drivingly engaged so as to transfer torque in a first rotational sense and the projection(s) of a selector member and the projection(s) of the second face of a gear may be drivingly engaged so as to transfer torque in a second rotational sense, opposite to the first rotational sense.

According to another aspect of the present invention there is provided a computer program comprising computer executable instructions which when executed by computing apparatus causes the computing apparatus to perform the method of any of the above clauses.

According to a further aspect of the present invention there is provided a non- transitory computer readable storage medium having stored thereon computer- readable code which when executed by computing apparatus causes the computing apparatus to perform the method of any of the above clauses.

According to an aspect of the present invention there is provided aapparatus for managing a gear ratio shift in a vehicle comprising: means for reducing pressure exerted between a first rotating body and a second rotating body in response to determining that a gear ratio shift is to be initiated while a vehicle is in motion, wherein the first rotating body is configured to be rotatably driven by a drive mechanism, and wherein the second rotating body is configured to be caused to spin due to friction between it and the first rotating body and comprises part of a linkage for transferring torque to an output portion for propelling the vehicle; means for implementing a gear ratio shift in a gear arrangement of the linkage; and means for restoring the pressure exerted between the first rotating body and the second rotating body.

Said means for reducing pressure exerted between the first rotating body and the second rotating body may be configured to reduce the pressure exerted between the first rotating body and the second rotating body until they are determined to be slipping against each other in use based on output from rotational speed sensors.

The apparatus may further comprise: means for, after the gear ratio shift, adjusting the relative difference between the rotational speeds of the first and second rotating bodies such that it complies with a predetermined condition using rotational speed sensors; and means for increasing the pressure exerted between the first and second rotating bodies such that a subsequent change in the rotational speed of one of them results in a corresponding change in the rotational speed of the other.

Brief Description of the Drawings

Embodiments of the invention will now be described by way of non-limiting example with reference to the accompanying drawings, in which:

Fig. l shows the assembly of the main components of a known Q-shift gearbox arrangement in a fully disengaged state;

Fig. 2 is an exploded view of the main components of the arrangement of Fig. l;

Fig. 3 shows a detail view of the dog features on a dog hub 3b and the

corresponding dog features on the gear 2 of Fig. 1;

Fig. 4 shows the assembly of the main components of the arrangement of Fig. 1 in a semi engaged state;

Fig. 5 shows the assembly of the main components of the arrangement of Fig. 1 in a fully engaged state; Fig. 6 is an overall view of a version of a known 4-ratio Q-shift gearbox including a shifting mechanism;

Fig. 7 shows the components of the shifting mechanism of Fig. 6 with one dog hub shown in an exploded view for reference;

Fig. 8 is a section view through the axis of a shift shaft 9, showing a spring connection between the shift shaft and selector drum;

Fig. 9 is a plan view of an embodiment of a known 4-ratio Q-shift gearbox shown with one ratio fully engaged;

Fig. 10 is a plan view of a 4-ratio gearbox of Fig. 9 shown in the process of making an up shift from third to fourth speed;

Fig. 11 is a plan view of a 4-ratio gearbox of Fig. 9 shown with a down shift pre- selected while positive driving torque is present;

Fig. 12 is a schematic view of some components within a car;

Figs. 13a and 13b are schematic cross-sectional views of a clutch mechanism;

Fig. 14 is a method of managing a gear ratio shift according to an embodiment of the present invention; and

Figs. 15 and 16 show respective ways of implementing step S7 in Fig. 14.

Detailed Description

Embodiments of the present invention are concerned with managing a gear ratio shift in a vehicle having a gearbox which can transfer torque between an input shaft and an output shaft thereof even while a gear ratio shift is taking place. An example of a gearbox which can to operate in this manner is the Q-shift type gearbox described inW020i4/0493i7. Other examples of gearboxes which can operate in this manner are described in USi,404,38sA, US2012/0240698A1, WO2004/099654A1, EP2169265A2 and UK patent application no. 1417965.9 for instance; the entire contents of each of the aforementioned documents being incorporated herein by reference.

For completeness of the present description however, an outline of the exemplary gearbox arrangement described in WO2014/049317 will first be provided. After that, an explanation will be given of how gear ratio shifts can be managed in accordance with embodiments of the present invention. Fig. l shows the assembly of the main components of a Q-shift gearbox in a fully disengaged state. The shaft ι passes through a first selector member, herein referred to as a 'dog hub' 3a, a drive member in the form of a gear 2 and a second dog hub 3b. The gear 2 is mounted on the shaft by a low friction bearing (not visible), being, in this embodiment, a combination of plain thrust washers and a needle roller bearing, so that it is axially and radially located on the shaft 1 but free to rotate relative to the shaft 1.

In this arrangement the drive member is a gear 2, although in other

arrangements the drive member may be any part of a drive mechanism which is required to be selectively rotatively engaged to a shaft, for example a roller chain sprocket or a belt drive pulley.

As can be seen with reference to Figs. 1 to 3 in particular, dog hubs 3 are substantially annular having two faces and a means for engaging with the shaft 1, in this example, a toothed inner ring 6a, 6b. They also comprise a plurality (in this example, 3) of projections or engagement 'dog' features 7a, 7b, 7f, 7e, arranged on each face thereof. The gear 2 is also annular, comprising a toothed outer surface and two opposed faces. The gear further comprises a plurality (in this example 3 on each face) of engagement 'dog' features 7c, 7d. The dog features 7 of the hubs 3b are shaped to engage the dog features 7 of the gear 2, as will be described in greater detail below.

In the disengaged state of Fig. 1, the two dog hubs 3a and 3b are axially displaced away from the gear 2 so that the dog features 7a, 7b on the hubs are disengaged from the corresponding dog features 7c, 7d on the gear 2. It will be noted that the dog features 7e, and 7f on the second face (as illustrated, the outer faces) of two dog hubs 3a, 3b do not engage with the gear 2 shown in Figs. 1-5, but are provided to allow for engagement with other gears which may be mounted on the main shaft (see Fig. 6 onwards) .

Fig. 2 is "an exploded view" of the main components of Fig. 1, in which the lower half of Fig. 2 shows the same components as the upper half of Fig. 2 but at a different viewing angle to reveal the dog features 7 on the second side of the gear 2. In the upper half of Fig. 2 the dog features 7a on the dog hub 3a and the corresponding dog features 7d on one side of the gear 2 are visible. In the lower half of Fig. 2 the dog features 7b on the dog hub 3b and the corresponding dog features 7c on the other side of the gear 2 are visible.

In this arrangement as illustrated in the accompanying illustrations, each face of each dog hub 3 and each face of each gear 2 have three such projections provided by the dog features 7 substantially evenly distributed around the axis of the shaft. However, the use of any number of similar dog features 7 per face is possible. For example if higher load capacity was required more dog features 7 could be used, or for simplicity, or to achieve engagement with higher speed difference, fewer dog features 7 could be used. However, providing three dog features 7 per face provides for substantially even load sharing between dog features 7 and gives a self centring action when the dog features 7 are

transmitting torque. If greater than three dog features 7 are used, high accuracy of manufacture is required to ensure that any errors in dog feature position are less than the deflection of the dog features 7 under load if load sharing over more than three dog features 7 is to be achieved. Also the use of three dog features 7 per face provides a large angular gap between features to allow positive engagement with a high relative speed between a dog hub 3 and gear 2. The shaft 1 comprises a plurality of male spline teeth portions 6c, 6d (the number of which corresponds to the number of dog hubs 3 in the assembly), and the inner ring of the dog hubs 3 carry corresponding female spline teeth 6a, 6b. These mating splines 6c, 6d and 6a, 6b are toleranced to engage with a clearance fit. The male spline teeth are wider (i.e. extend further axially along the shaft 1) than the female spline teeth such that, once arranged on the shaft 1, the dog hubs 3a, 3b are radially and rotatively connected to the shaft 1 when aligned with the male spline teeth 6c, 6d but are free to move axially, while maintaining this connection to the shaft 1. This allows the dog features 7a, 7b of the dog hubs 3a, 3b to move in and out of engagement with the dog features 7c, 7d of the gear 2 while still being driven by the shaft 1.

Fig. 3 shows a detail view of the dog features 7b on a dog hub 3b and the corresponding dog features 7c on the gear 2. Each of the dog features 7a-d consist of a ramp-like block projecting above the side face of the gear 2 or dog hub 3. Each dog feature 7a-d rises from a base 4e which is in the same plane as the face of the dog hub 3 and comprises a sloping surface 5b, 5c and a 'mating' or contacting surface 4c, 4b which is angled away from the axis of the shaft 1 to provide positive engagement, or a physical interlock, between a feature on a hub 3 and a feature on the gear 2 when one is rotatively driving the other. Furthermore, taking the example of a single hub, when the dog hub 3b is moved towards the gear 2 by action of a shifting mechanism (described herein below), the leading edges 4b of the dog features 7b of the dog hub 3b will become engaged with the leading edges 4c of the dog features 7c of the gear 2 and provide a means of transmitting torque between the two components. The contact forces between the mating faces 4b, 4c, resulting from the transmission of torque will tend (because of the angle of the faces 4b, 4c) to pull the dog hub 3b toward the gear 2, so ensuring there is no tendency for the mechanism to fall out of engagement when transmitting torque. The sloping surfaces 5b, 5c of the dog features 7b, 7c provide a trailing edge which ramps at a relatively shallow angle. When there is relative rotation between the dog hub 3b and the gear 2 in the other sense, the shallow sloping surface 5b, 5c will easily ride up each other so forcing the dog hub 3b away from the gear 2 and so disengaging any connection between the two components. This 'ramped' form of the dog features 7a-f therefore provides for positive

engagement and transmission of torque in one sense of relative rotation and for disengagement of the dog features 7a-f and no transmission of torque in the other sense of relative rotation. It should be noted that the dog features 7 on one side of the gear 2 and the corresponding dog features 7 on the dog hub 3a, as visible in the upper half of Fig. 2, are adapted to give positive engagement and driving connection between the dog hub 3a and the gear 2 in one sense of relative rotation and the dog features 7 on the other side of the gear 2 and the corresponding dog features 7 on the dog hub 3b are adapted to give positive engagement and driving connection in the other sense of relative rotation. Therefore when both dog hubs 3 are moved toward the gear 2 by action of a shifting mechanism (described later), there exists positive driving engagement in both senses of relative rotation between the gear 2 and the dog hubs 3a, 3b, and since both the dog hubs 3a, 3b are connected to the shaft by splines, to the shaft 1. Fig. 4 shows the assembly of Fig. l in a 'semi engaged' state. Dog hub 3a is in contact with gear 2 and the mating surfaces 4a, 4d of the dog features 7 are in contact with each other. Dog hub 3b is axially displaced away from the gear 2 so that there is a clearance between the tips of the dog feature projections. In this state therefore the gear 2 is rotatively connected to the shaft 1 for relative rotation (e.g. positive torque/forward drive) in one sense but not in the other sense (e.g. negative torque/reverse drive) .

Fig. 5 shows the assembly of Fig. 1 in a fully engaged state. Both dog hubs 3a, 3b are in contact with the gear 2. The leading edges of the dog features 7a, 7b on the dog hubs 3 are in engagement with the corresponding leading edges of dog features 7c, 7d on the gear 2. In this state therefore the gear 2 is rotatively connected to the shaft 1 for relative rotation and drive in both senses. It should be noted that the assembly of the dog hubs 3a, 3b onto the shaft 1 is chosen so that in this state of full engagement only a small tangential clearance exists between the leading edges of the dog features 7. This small tangential clearance ensures that only a small amount of backlash is present between the gear 2 and the shaft 1 in this fully engaged state. For road vehicles, especially passenger vehicles, low backlash is desirable for passenger comfort, to minimise torsional shocks in the driveline each time a torque reversal occurs, for example when the driver lifts off the accelerator pedal. A minimal amount of clearance is provided between the dogs to allow for engagement of the dog features 7a-f having angled leading edges (as shown in this embodiment) to provide positive engagement, and to allow for manufacturing tolerances.

Fig. 6 is an overall view of a 4-ratio, Q-shift gearbox arrangement. The main shaft la carries four gears 2a, 2b, 2c, 2d, of varying diameters each meshing with a lay gear 8, the lay gear 8 having four meshing gears 2 formed along its length. The main shaft la is the input and the lay gear 8 is the output (and so the main shaft la comprises a gearbox input shaft and the lay gear 8 comprises a gearbox output shaft) . When one of the gears 2a-d is selected, by rotatively connecting it to the main shaft using the dog hubs 3, there is a driving connection between the input shaft la and the lay gear 8, the relative speeds of the input shaft la to the lay gear 8 being the ratio of the numbers of teeth on the selected gear 2 and the meshing gear on the lay gear 8. The diameters and tooth numbers on each of the gears are chosen to achieve substantially similar ratio differences between adjacent gear pairs so providing for similar step changes in input speed as shifts are made. The gearbox can therefore undergo a gear ratio shift by changing the selected gear so that the newly selected gear has a different ratio between the number of teeth on it and the corresponding meshing gear than the

corresponding ratio of the previously selected gear. A gear ratio shift thus changes the relative speed of the gearbox input shaft to that of the output shaft.

The shift mechanism comprises a shift shaft 9, connected to a number of shift mechanisms 10. In this four speed example embodiment there are five such shift mechanisms, one for each dog hub 3. The dog hubs 3 between each of the gears 2 on the main shaft la have dog features 7 formed on both sides, as shown in Fig. 1, to engage with either of the adjacent gears 2. The dog hubs 3 on each end of the main shaft la shown in this illustration have dog features 7 on only one side to engage with the adjacent gear 2. However, for convenience, to minimise the number of different components used in a given gearbox, and to guard against incorrect assembly, the same 'two sided' dog hub 3 as used between gears 2 could be used at the end of the main shaft la, with its outer dog feature 7 being redundant.

The main shaft la, the lay gear 8 and the shift shaft 9 are supported in a casing on suitable bearings, one bearing at each end of each shaft. The casing and bearings are not shown in this illustration but suitable bearings would be familiar to the person skilled in the art.

Fig. 7 shows the components of the shifting mechanism coupled to one dog hub 3 shown in an exploded view. A selector drum 11 is mounted on the shift shaft 9. Shift arms 12a, 12b are provided, one mounted above and one mounted below each selector drum 11, and located into the gearbox case (not illustrated) by a pivot pin 13. The pivot pin 13 is located in holes in the gearbox case so as to be stationary with the case. On one end of each pivot arm 12a, 12b there is a pin 15, which engages with a track 16 cut in the outer diameter of the selector drum 11 (see also Fig. 9 for views of the tracks 16) . As can perhaps be better appreciated in the assembled shifting mechanism shown in Fig. 9, the axial position of the track 16 varies around the circumference of the drum 11. As the drum 11, is rotated with the shift shaft 9, the pin 15 therefore is moved axially and the shift arms I2a-b, pivot around the pivot pin 13. On the other end of the shift arms 12a, 12b is mounted the shift thrust ring 14. Pins 17 formed on the outside of the shift thrust ring 14 fit through holes in the ends of the shift arms 12a, 12b so that the shift thrust ring 14 can pivot on the end of the arms 12a, 12b. The shift thrust ring 14 is formed to have a groove 18 around its inside diameter which fits over a ridge 19 formed on the outside diameter of the dog hub 3. When the shift thrust ring 14 is fitted over the dog hub 3 there exists a small axial clearance between the inside of the groove 18 and the profile of the ridge 19 so that the dog hub 3 can freely rotate inside the ring 14, but if the ring 14 is axially displaced the dog hub 3 is also axially displaced. Specifically, when the shift arms 12a, 12b pivot by the action of the selector drum 11 being rotated, the shift thrust ring 14 therefore is axially displaced and with it the dog hub 3.

The gearbox is assumed to be partially filled with oil for cooling and lubrication of the gears and bearings. This oil therefore will provide lubrication between the inside of the groove 18 and the ridge 19 so forming a thrust bearing capable of displacing the dog hub 3 for the purpose of gear selection when the gearbox is rotating at high speed. As will be familiar to the skilled person, some of the inside surface of the groove 18 may be cut back to leave raised thrust pads to reduce the area of contact between the groove 18 and the ridge 19 to reduce friction and aid lubrication.

The thrust bearing between the shift thrust ring 14 and the dog hub 3 may alternatively be any other suitable form of thrust bearing, for example a needle roller thrust bearing, a ball bearing or a spherical roller bearing. As will be familiar to the skilled person, any such bearing is capable of causing an axial displacement of the dog hub 3.

Fig. 8 is a section view through the axis of the shift shaft 9 revealing the spring connection between the shift shaft 9 and the selector drum 11. The shift shaft 9 comprises a tube with a series of slots 22 (see also Fig. 9) which are cut through the tube wall. Pins 21, one each side of each selector drum 11, fit through the slots 22 and engage with cut out tracks 25 on each side of the selector drum 11. Springs 24 are fitted inside the shift shaft 9 between each selector drum 11 with spring cups 23 on each end of each spring 24, the spring cups 23 resting on the pins 21. The springs 24 are selected to be longer than the gap between each spring cup 23 so that they are partially compressed on assembly and so provide a defined preload to the pins 21. The middle portion of each pin 21, on which the spring cups 23 rest, is a smaller diameter so that there is a step each side of the spring cup 23 to retain the pin 21 in position in the shift shaft 9 even if the pin 21 is revealed by axial displacement of the selector drum 11. It should be noted that the gap between the ends of the slots 22 in the shift shaft 9 and the axial thickness of the material between the cut-out tracks 25 in the selector drum 11 are substantially the same, so that preload on the pins 21 from the springs 24 holds the pins 21 in contact with the ends of the slots 22 and the selector drum 11 is positively located between the pins 21 along the axis of the shift shaft 9. The pins 21 passing through the slots 22 and into the cut-outs 25 in the selector drums 11 provide a rotary connection between the shift shaft 9 and the drums 11 but allow axial displacement within the length of the slot 22. If a force acts to push the sector drums 11 along the shift shaft 9 which exceeds the preload in the springs 24 then the pins 21 will move along the slot 22, further compressing the spring 24, allowing axial displacement of the selector drum 11 while maintaining rotary connection. The selector drums 11 therefore are positively located along the length of the shift shaft 9 so can provide positive control over the shift arms and the movement of the dog hubs 3 along the shaft 1, but if a load defined by the preload in the springs 24 is exceeded, then the selector drums 11 may be axially displaced along the shift shaft 9.

It should be noted that, in this example, each selector drum 11 is fitted over the shift shaft 9 with a clearance fit to allow axial displacement and additionally the internal bore 28 of the selector drum 11 is of a double conical form to allow for some misalignment of the drum 11 on the shaft 9. Additionally it should be noted that the two shift arms 12a and 12b in the example embodiment can pivot independently of one another. This provides for an amount of flexibility in the mechanism to allow for variation in the geometry of each component due to manufacturing tolerance and the like. The dog hub 3 is then allowed to contact the gear 2 without being overly constrained by the shift mechanism. It may also be noted that the use of a selector drum 11 with two pins 15 engaging in its track, one directly opposite the other, ensures that the resulting forces from the shift arms 12 on the selector drum 11 act substantially at the centre line of the shaft 9 on the spring 24. This configuration therefore resists any tendency for the drum 11 to bind on the shift shaft 9 if it is axially displaced by the action of the dog features 7 on the dog hub 3 engaging with dog features 7 on the gear 2.

It may be noted, that, in this example, for cost effective manufacture by minimising the number of different components used and preventing incorrect assembly, each of the selector drums 11 fitted to the shift shaft 9 are identical components, and that their angular and axial positions are determined by the location of the various slots 22 cut through the shift shaft 9. Also the selector drums 11 are designed to be symmetrical so that they do not need to be installed in any particular orientation to function correctly. However, this need not be the case in all examples.

Fig. 7 shows a selector drum 11 fitted to the shift shaft 9. Adjacent to the drum 11 is the gate 26. The gate 26 is fixed to the case of the gearbox (not shown), and has a series of slots through which pass the edges of each selector drum 11. There exists a clearance between the drums 11 and the slots in the gates 26 so that in normal operation the drums 11 rotate freely without contacting the gate 26. In the end faces of each drum 11 there are cut-outs 27 adjacent to the slot in the gate 26. At certain angular positions the cut-out 27 of a drum 11 aligns with the gate 26 so that if the selector drum 11 is displaced axially along the shift shaft 9 then the cut-out 27 engages with the gate 26 and the rotation of the drum 11, and the whole shift shaft 9, is limited to the angular extent of the cut-out. This provides a mechanism to block certain combinations of dog hub 3 movement which could otherwise damage the gearbox as now described. In other angular positions, the cut-out 27 does not align with the gate 26 and so the axial movement of the drum 11 is limited by the gate 26.

It should be noted that the cut-outs 27 in the selector drum 11 are aligned with the gate 26 when the pin 15 on the shift arms 12 are in the portion of the track 16 which cause axial displacement of the pin and so moves the shift arms 12 and moves the dog hubs 3 into engagement with the gears 2. The remaining portion of the tracks 16 in the selector drums 11 provide no axial displacement, therefore when the pins 15 are in this portion of the track the dog hubs 3 are held substantially at mid position between gears 2 and so are not in engagement with the gear 2, and cut-outs 27 do not align with the gate 26 so axial displacement of the drum 11 is limited to the clearance between the drum edges and the gate 26. This shift mechanism could also be used in other drive member selection assemblies. Fig. 9 is a plan view of a Q-shift type gearbox having four gear ratios, one of which is fully engaged. Specifically, dog hubs 3a, and 3b are engaged with gear 2a. The shift shaft 9 is in an angular position where the pins 15a, 15b in tracks 16a and 16b in the selector drums 11a and 11b are displaced axially away from gear 2a so that shift arms 12a and 12c are pivoted around the pivot pins 13a, 13b to hold the dog hubs 3 in engagement with the gear 2a.

The gearbox is assembled so that up shifts are performed with substantially positive torque transmission, (i.e. when the sense of torque at the input shaft is the same as the sense of rotation), and down shifts are performed with substantially negative torque (when the sense of torque at the input shaft is opposite to the sense of rotation) . Up shifts are transitions from one gear ratio to another which result in a reduction in input shaft speed for a constant output shaft speed and are normally performed sequentially as the vehicle accelerates. Down shifts are transitions from one gear ratio to another which result in an increase in input shaft speed for a constant output shaft speed and are normally performed sequentially as the vehicle decelerates.

In Fig. 9 it should be noted that the dog features 7 on the right hand faces of each gear 2 and the meshing dog features 7 on the dog hub 3 provide positive driving torque connection between the gear 2 and the shaft and that dog features 7 on the left hand side of each gear 2 provide negative driving torque connection between the gear 2 and the shaft. When the gearbox is transmitting positive torque therefore the dog hub 3b on the right hand side of gear 2a, as illustrated in Fig. 9, is transmitting torque from the input shaft to the gear 2, the gear 2 then meshing with the lay gear drives the output. It should be noted that when transmitting torque, due to the angled faces of the dog features 7 and the friction in the spline connection between the dog hub 3 and the shaft 1, the dog hub 3 may not readily be moved away from the gear 2 to disengage the gear 2 from the shaft. The preload in the springs 24 on the shift shaft 9 is limited to ensure that disengagement can only occur at relatively low torque levels. As shown in the Fig. 10, however, the dog hub 3a, however, when positive torque is being transmitted, is not loaded and so is free to move away from gear 2a if so urged by the shift mechanism. Fig. 10 is a plan view of the gearbox of Fig. 9 including a shifting mechanism, shown in the process of making an up shift from third to fourth speed (i.e. an upshift from the third gear ratio to the fourth gear ratio) . The shift shaft 9 has been rotated fully to the fourth gear position. The dog hub 3b is transmitting torque from the input shaft to the gear 2a and so is held in engagement. The associated selector drum 11b is displaced along the shift shaft 9. Another dog hub 3c is being moved towards the gear 2b by action of the spring 24 in the shift shaft 9. Since the third gear 2a is still engaged, the gear 2b is rotating slower than the shaft la and the dog hubs 3, so the ramped faces of the dog features 7 ride up one another to push the dog hub 3c away from the gear 2b without engagement. The dog hub 3c is allowed to move away from the gear 2b by compressing the spring 24 inside the shift shaft 9. A further dog hub 3a is being moved toward the fourth gear 2b by action of the spring 24 inside the shift shaft 9. The dog features 7 between the fourth gear 2b and the dog hub 3a are positive driving and the main shaft is rotating faster than the fourth gear therefore as the dog hub 3a is moved axially towards the gear 2b the leading faces of the dog features 7 contact and positive drive is achieved between the shaft and the gear 2b. As drive is picked up by dog hub 3a, the load on dog hub 3b is relaxed and the gear 2a is then rotating faster than the main shaft 1. The ramps on the dog features 7 then will cause the dog hub 3b to be pushed away from the gear 2a and also the spring 24 in the shift shaft 9 moves the selector drum 11 back to the neutral position.

Once the dog hub 3a engages with gear 2b the main shaft rotates at the same speed as the gear 2b therefore dog hub 3c is able to move full into engagement with gear 2b. Fourth gear is then full engaged and the shifting process is complete.

It should be noted that the shift is achieved by the axial movement of an intermediate dog hub 3, which is between the two gears 2 which are engaged or disengaged. This dog hub 3 is free to move to initiate the shift since it is the negative driving component for the gear 2 being disengaged, and it makes the engagement because it is the positive driving component for the gear 2 which is to be engaged. Once engagement is made by the intermediate dog hub 3 the positive driving dog hub 3 from the earlier selected gear 2 is released and synchronisation is achieved to allow the negative driving dog hub 3 on the gear 2 being engaged to move fully into engagement.

Expressed more generally, if a positive driving selector member is moved toward a drive member of a higher gear, the leading edges of the projections on each are moving towards each other, because of the different gear ratio. When the leading edges contact, drive is picked up by the higher gear drive member and the selector for the lower gear drive member becomes disengaged. Because drive from the lower gear is only disengaged after the drive is picked up by the higher gear, no interruption in torque transmission occurs and the shift may be substantially instantaneous.

Down shifts are completed by the same mechanism of movement of the intermediate gear 2 while negative torque is being transmitted, in the case of a conventional engine driven vehicle this would be when the diver has lifted off the accelerator pedal.

If a down shift is required when positive driving torque is demanded for example if the vehicle is climbing a slope and vehicle speed is falling, a different shifting process is employed. While the gearbox is transmitting positive torque, a single down shift can be pre-selected by rotating the shift shaft 9 to the next lower gear position. The intermediate dog hub 3 is not free to move since it is the positive driving engagement between the gear 2 and the input shaft. The selector drum 11 for the intermediate gear 2 therefore is axially displaced along the shift shaft 9 and the spring 24 inside the shift shaft 9 is further compressed. The negative driving dog hub 3 for the lower gear to be engaged is brought into contact with the gear 2 but it is rotating slower than the gear 2 so it does not engage. To complete the shift, in this example, the driver momentarily reduces or reverses the driving torque to release the intermediate dog hub 3. This is most easily achieved by lifting the accelerator or alternatively by dipping a clutch pedal, if provided. As soon as the torque transmitted through the intermediate dog hub 3 is relaxed the compressed spring 24 in the shift shaft 9 moves the dog hub 3 into engagement with the lower gear, the negative driving dogs engage, the gear 2 and shaft 1 are synchronised and the forward driving dog falls into engagement completing the down shift.

If there exists positive driving torque which holds a dog hub 3 into engagement when more than one down shift is attempted, without allowing the shift to be completed by reducing the torque, there would be, if no steps were taken to prevent it, the possibility that a negative driving dog hub for a lower gear could be engage while the positive driving dog of the other gear is still engaged. This would result in a lock up condition which is likely to cause significant damage to the assembly, but can be addressed as set out below.

Fig. 11 is a plan view of the gearbox with a down shift pre-selected while positive driving torque is present. Dog hub 3b is held in engagement with gear 2a. Dog hub 3d is moved towards gear 2c but as the gear 2c is rotating faster than the main shaft due to gear 2a being still engaged, the reverse driving dog 3d does not engage. In this condition the shift shaft 9 has been rotated to a position corresponding to the lower gear being fully engaged. As dog hub 3b is held in engagement with gear 2a the selector drum 11b is displaced along the shift shaft 9 and the cut-out 27 is engaged with the gate 26. The shift shaft 9 can therefore not be rotated further to pre select a second down shift and so a lockup condition is prevented. Similar cut outs are present in both sides of each selector drum 11 and the gate 26 is adjacent to each side of each selector drum 11, therefore preselection of more than one up shift while negative torque is transmitted is similarly prevented. The cut-outs 27 therefore prevent the simultaneous engagement by a negative driving dog hub 3 of one gear 2 and a positive driving dog hub 3 of a higher gear 2, and also prevent the simultaneous selection of a positive driving dog hub 3 of one gear 2 and a negative driving dog 3 of a lower gear 2. The shift shaft 9 in this example must be rotated by a predetermined angle on each shift. For manual operation this is most easily achieved using any form of indexing mechanism familiar to one skilled in the art for example a ratchet mechanism used for indexing a selector drum in a motor cycle gearbox.

Alternatively any form of rotary actuator for example an electric servo motor, a hydraulic servo motor or a pneumatic servo motor could be used. Other shifting mechanisms which provide the necessary control of the dog hubs 3 to achieve appropriate selection of gears may be devised by one skilled in the art, for example other mechanical manually operated devices, use of individual actuators for example electrometrical actuators, hydraulic actuator or pneumatic actuator for each of the dog hubs 3. Electronic or other forms of control system may be used to operate the gear mechanism. This could simply take the form of a means to determine when to make a shift and so automating the shifts using a mechanical shift mechanism or could be a system to control the operation of the individual dog hubs 3 in the necessary sequences.

It will be apparent to one skilled in the art that the shifting components and mechanism of the heretofore described Q-shift gearbox could be used in many different configurations of gearbox or transmission for any purpose, for example using any number of different selectable gears or with selectable gears mounted on more than one shaft.

Also, it will be noted that a gear shift as performed by the heretofore described gearbox results in substantially an instantaneous step change in the speed of the driving engine or other rotating power source without requiring use of a clutch (thus reducing complexity of use) . Since this power source will have some inertia there will be an impulse (referred to as a torque pulse, torque spike or gear shift shock) imparted to the connected drive line. Within the drive line of a vehicle there exist various sources of torsional flexibility, for example drive shafts, clutch pressure plate springs and tyres, which will absorb at least some of the torque pulse. However tests have shown that at certain engine rpms, such an impulse is noticeable to a car driver in that it can be felt following a gear shift. In the case of a high performance vehicle, for example a sports car or racing car, generation of impulses along the driveline following gear ratio shifts may be acceptable to the driver, however for a vehicle requiring a higher level of refinement it may not be.

Furthermore, it should be noted that the gearbox or any of the components may be used in any mechanism requiring the selective coupling of components to shafts. In addition it should be noted that a gearbox or any of the components described herein may be used in conjunction with any rotary power source and rotary load for example in a transmission coupling an electric motor to the wheels of a vehicle. Other examples of rotary power sources which could be used include, but are not limited to, hydraulic motors, pneumatic motors, internal combustion engines and gas turbine engines. Management of Gear Ratio Shifts

Embodiments of the present invention are described hereafter in the context of a car, specifically a method for managing a gear ratio shift in a car. However this is not intended to be limiting and embodiments of the present invention could equally be used in other contexts such as managing a gear ratio shift in another type of vehicle whether land based or not (e.g. a truck, lorry, van, motorcycle, a vehicle having one or more tracks for propulsion purposes such as a tank or snow mobile, a boat, jet-ski or aircraft). In water craft and aircraft a gear box could be provided for controlling the rotation of a propeller for example. With the above proviso in mind, Fig. 12 illustrates a highly schematic view of some components within in a car. An internal combustion engine 200 can be coupled to the wheels 202 via a driveline 204. Components forming the driveline 204 cooperate with each other to transfer torque generated by the engine 200 to the wheels 202 for propelling the car. Examples of some components included within the driveline 204 (not all of which are shown) include a clutch, gearbox, shafts, differentials, universal joints, constant velocity joints and wheel axels for instance.

The clutch 206 and gearbox 208 of the driveline 204 are depicted in Fig. 12. The clutch 206 is a friction clutch and thus includes a clutch plate 210 that can be caused to engage and subsequently disengage a flywheel 212 of the engine 200. In other words the clutch 206 has a first configuration in which the clutch plate 210 is caused to be forced against the flywheel 212 and a second configuration in which the clutch plate 210 is not forced against the flywheel 212.

Persons skilled in the art will appreciate that the flywheel 212 is caused to spin by torque generated by the engine 200. When the clutch plate 210 is forced against the flywheel 212, friction between these two components provides that the spinning flywheel 212 drives rotation of the clutch plate 210 and so engine torque is transferred to the driveline 204. The clutch plate 210 is rotationally fixed to a gearbox input shaft 207, so when the clutch plate 210 rotates so does the gearbox input shaft 207. In this manner torque is transferred through the gearbox 208 to an output shaft 209 thereof, and then further down the driveline 204 to the wheels 202 which causes them to rotate. Suitable friction clutch configurations will be apparent to persons skilled in the art, for instance a diaphragm clutch of either the push or pull variety. Diaphragm clutches are well known, an example of which is shown in Figs. 13a and 13b.

Fig. 13a shows a diaphragm clutch in a first configuration in which the clutch plate 308 thereof is forced against an engine flywheel 302. The diaphragm clutch has a clutch housing 300 that is fixed to the flywheel 302 using bolts for example. The clutch housing 300 therefore rotates with the flywheel 302. Inside the clutch housing 300 a diaphragm spring 304, a pressure plate 306 and the aforementioned clutch plate 308 are provided around a shaft 310 (the gearbox input shaft) which extends partially into a recess 307 of the flywheel 302. Of these components, only the clutch plate 308 is rotationally fixed to shaft 310. Thus when the clutch plate 308 rotates, due to friction between it and the spinning flywheel 302, the shaft 310 rotates at the same rotational speed as the clutch plate 308. In the configuration shown, the clutch plate 308 is being forced against the flywheel 302 by the pressure plate 306 and the diaphragm spring 304. Friction between clutch plate 308 and flywheel 302 enables the spinning flywheel 302 to exert a torque on the clutch plate 308, thereby causing it and thus the shaft 310 to rotate. Now looking at Fig. 13b, to disengage the clutch plate 308 from the flywheel 302 a throw-out bearing 311 is caused to exert an axial force on the diaphragm spring 304 along a direction towards the clutch plate 308. As this force is applied, the diaphragm spring 304 flexes and thereby draws the pressure plate 306 away from the clutch plate 308. In particular, forcing the throw-out bearing 311 against the diaphragm spring 304 causes the diaphragm spring 304 to flex about the locations where it is coupled to the pressure plate 306. This will be described in more detail. Mounting elements 312 extend through the diaphragm spring 304. One end of each mounting element 312 is coupled to the clutch housing 300, whereas the other end is wider than the opening in the diaphragm spring 304 through which it extends. Thus when the throw-out bearing 311 presses against the diaphragm spring 304, reaction forces between the widened end of the mounting elements 312 and the diaphragm spring 304 result in the diaphragm spring 304 flexing. This flexing causes the outer edges of the diaphragm spring 304, which are coupled to the pressure plate 306, to move away from the clutch plate 308. The pressure exerted by the pressure plate 306 on the clutch plate 308 is thereby reduced and the torque exerted on the clutch plate 308 by virtue of friction between it and the spinning flywheel 302 is reduced.

The friction clutch arrangement described in connection with Figs. 13a and 13b is only one example of a suitable clutch mechanism that could be used to implement the present invention. On reading this specification however other suitable friction clutch mechanisms will be apparent to persons skilled in the art, which need not necessarily be of the diaphragm variety (e.g. a coil-spring clutch). Some other examples of suitable friction clutch mechanisms include a dry multi plate clutch external to the gear box 208, a wet multiplated clutch internal to the gearbox 208, a cone clutch and a magnetic clutch.

Referring back to Fig. 12 the gearbox 208 is configured such that it can transfer torque between an input shaft 207 and an output shaft 209 thereof even while a gear ratio shift operation is taking place. In another manner of speaking, during a gear ratio shift the amount of torque transferred between the input shaft 207 and the output shaft 209 does not reduce to zero. This provides that the car does not experience a loss of driving power during a gear ratio shift operation.

Suitable gearboxes configured to operate in this manner have already been mentioned towards the beginning of this description, one of which is the Q-shift type gearbox heretofore described.

Before a detailed explanation of how a gear ratio shift in the arrangement depicted in Fig. 12 is managed, a brief summary is first provided.

Essentially, before a gear ratio shift operation takes place while the car is in motion (e.g. driving in first gear) the amount of pressure exerted by the clutch plate 210 on the flywheel 212 of the engine 200 is gradually reduced. When the pressure has been reduced enough, clutch slip will occur, meaning that static friction between the clutch plate 210 and the flywheel 212 will be overcome and they will begin to slide against one another while the torque is still transmitted at substantially the same magnitude as before the shift process was initiated. At this stage the gear ratio shift operation takes place (e.g. a shift from first to second gear), which causes a sudden change in the rotational speed of the clutch plate 210. This is because following a gear ratio shift within the gearbox 208, the gearbox input shaft 207 will experience a sudden change in rotational speed because of a change in the gear that is rotationally fixed relative to the input shaft 207; whereby the selected gear has a different number of teeth than the gear previously rotationally fixed to the shaft. However, since clutch slip is occurring, the flywheel 212 is not caused by friction between it and the clutch plate 210 to undergo a similar sudden change of rotational speed to match that of the clutch plate 210. A step change in engine speed caused by a sudden change in the rotational speed of the flywheel 212 would cause an impulse to be imparted along the driveline 204 to the wheels 202, however instead, due to clutch slip, after a sudden change in the rotational speed of the clutch plate 210, the flywheel 212 slides against it at a different rotational speed. Next, the rotational speeds of the clutch plate 210 and flywheel 212 are substantially matched by causing a variation in the rotational speed of the flywheel 212, for example by controlling the amount of torque generated by the engine 200. After this, when the clutch plate 210 and flywheel 212 are rotating at substantially similar speeds, the pressure exerted by the clutch plate 210 on the flywheel 212 is again increased. In particular, the pressure is increased by an amount such that a subsequent change in the rotational speed of either the clutch plate 210 or the flywheel 212 results in a corresponding change in the rotational speed of the other.

Managing a gear ratio shift in the manner described in the foregoing paragraph provides that a step change in engine speed does not occur following a gear ratio shift inside the gearbox 208. As a result, the magnitude of any impulses imparted through the vehicle driveline 204 following a gear ratio shift operation are reduced, since changes in engine speed occur over an extended controlled period of time while the clutch is slipping, which increases passenger comfort and improves car refinement and also improves the life and reliability of the components of the driveline. In other words, car passengers are less likely to feel that a gear ratio shift has occurred due to the reduction in magnitude of torque spikes along the vehicle driveline 204 after a gear ratio shift operation. How the above is implemented will now be described in more detail.

Referring further to Fig. 12, the car includes a variety of electronic components coupled together by a system bus 214 that is used to transfer signals (e.g. raw data and control signals) between such components, one of which is a controller 216. Non-volatile memory 218 and volatile memory 220 are also coupled to the system bus 214, wherein the controller 216 is configured to execute program code stored in the non-volatile memory 218 and to use the volatile memory 220 to store intermediate results.

It will be appreciated that the controller 216 may take any suitable form, for instance it may be a microcontroller, plural microcontrollers, a processor, or plural processors. Furthermore the non-volatile memory 218 could comprise ROM such as PROM, EPROM or EEPROM and the volatile memory 220 could comprise RAM such as DRAM or SRAM.

A gear ratio shift management application 219 is stored in the non-volatile memory 218. The controller 216 is configured to load the gear ratio shift management application 219 into the volatile memory 220 and subsequently implement the functionality defined by it. Upon doing so the controller 216 interacts with other electronic components coupled to the system bus 214 to manage a gear ratio shift operation. The functionality that the gear ratio shift management application 219 causes the controller 216 to implement when executed is described below, however the nature of the other electronic components coupled to the system bus 214 will be described first.

A first rotational speed sensor 222 (hereafter "the first sensor") connected to the system bus 214 is provided to generate output indicative of the rotational speed of a crankshaft forming part of the engine 200. The flywheel 212 is rotationally fixed relative to the crankshaft and so they rotate at the same speed. The specific configuration of the first sensor 222 is not essential and suitable rotational speed sensors will be apparent to persons skilled in the art. In some embodiments output generated by the first sensor 222 is indicative of a value corresponding to the rotational speed of the crank shaft, whereas in other embodiments the output generated is raw data and the controller 216 determines the rotational speed of the crank shaft from this raw data. A second rotational speed sensor 224 (hereafter "the second sensor") is connected to the system bus 214 and is provided to generate output indicative of the rotational speed of the gearbox input shaft 207. It will be remembered that this is the shaft to which the clutch plate 210 is rotationally fixed, such that when the clutch plate 210 engages the spinning engine flywheel 212 the gearbox input shaft 207 will be caused to rotate. The second sensor 224 can have a similar configuration to the first sensor 222 and as such, suitable second sensors for generating output indicative of the rotational speed of the gearbox input shaft 207 will be apparent to persons skilled in the art.

A clutch actuator 226 is connected to the system bus 214 for controllably releasing and engaging the clutch 206. In particular the clutch actuator 226 is used to control the force between the clutch plate 210 and the flywheel 212, and so controls the frictional forces acting between the clutch plate and the flywheel and correspondingly the torque transmitted by the clutch 206 from the engine 200 to the gearbox input shaft 207.

In the case of the diaphragm clutch mechanism depicted in Figs. 13a and 13b for example, the clutch actuator 226 could be an electric motor (e.g. a servo motor) for controlling the extent of rotation of a clutch fork in engagement with the throw-out bearing 311. Rotating the clutch fork towards the throw-out bearing 311 urges it towards the clutch plate 308 and so causes the pressure exerted by the clutch plate 308 on the flywheel 302 to decrease (see Fig. 13b) . Subsequently rotating the clutch fork the other way enables the diaphragm spring 304 to flex back into is original shape such that the pressure exerted by the clutch plate 308 on the flywheel 302 increases (see Fig. 13a) .

Instead of a servo motor as mentioned in the foregoing example, other types of rotary actuators could be used instead such as a hydraulic servo motor or a pneumatic servo motor. Furthermore, in some embodiments the clutch actuator 226 could instead be a linear actuator having a member that contollably moves forwards and backwards (e.g. an electromechanical, hydraulic or pneumatic linear actuator) . Such forwards motion could be used to reduce clutch plate pressure on the flywheel (by pressing against the throw-out bearing 311 in the example of Figs. 13a and 13b). Similarly backwards motion of the actuator member could increase clutch plate pressure on the flywheel.

In view of the foregoing it will therefore be appreciated that the specific configuration of the clutch actuator 226 is not essential.

A shift actuator 228 is also connected to the system bus 214 and is configured to implement gear ratio shifts of the gearbox 208. In the case of a Q-shift type gearbox of the kind heretofore described, rotation of the shift shaft 9 causes gear ratio shifts to take place. The shift shaft 9 must be rotated by a predetermined angle on each shift. Any form of rotary actuator for example an electric servo motor, a hydraulic servo motor or a pneumatic servo motor could be used for this purpose. Any such actuator could therefore be used as the shift actuator 228.

Furthermore, as already previously mentioned, it is acknowledged that some envisioned Q-shift gearbox arrangements could have a shifting mechanism other than the type heretofore described which includes the drive shaft 9 for providing the necessary control of dog hubs to achieve an appropriate selection of gears. Such gearboxes may instead require the use of individual actuators for example an electromechanical actuator, hydraulic actuator or pneumatic actuator for each of the dog hubs. In the case of such gearbox arrangements the shift actuator 228 comprises a means of automating gear ratio shifts by controlling the operation of the individual dog hubs in the necessary sequences.

In view of the foregoing it will therefore be appreciated that the specific configuration of the shift actuator 228 is also not essential. An instructor 230 connected to the system bus 214 in Fig. 16 is configured to communicate with the controller 216 in order to indicate that a gear ratio shift is to be initiated. More specifically, the instructor 230 is configured to determine whether an upshift or a down shift is required and in response to provide the controller 216 with information indicative of the nature of the gear ratio shift to be initiated. In manual transmission systems, the instructor 230 could receive signals from one or more sensors which monitor the position of a gear stick, or whether one or more gear shift buttons on a steering wheel have been pressed for example. In this manner the instructor 230 can determine when a gear ratio shift is required by monitoring changes in the position of the gear stick, or from the manner that one or more gear shift buttons are pressed, and then provide information to the controller 216 indicative of the nature of the gear ratio shift to be initiated. In other embodiments however, information indicative of the position of a gear stick, or the manner in which one or more gear shift buttons are pressed, could be sent to the controller 216 directly which then itself determines the nature of gear ratio shifts to be initiated, without the use of an instructor 230.

In the example of when a gear stick is manipulated by a user to make a gear shift, a sensor (e.g. a magnetic proximity detector, optical sensor or micro switch) could generate output when the gear stick begins to move out of a so called "in gear" configuration. Such output could be detected by the instructor 230, or the controller 216 directly, and thereby cause a dipping operation of the clutch 206 to be initiated - such that clutch slip between the clutch plate 210 and flywheel 212 occurs before the gear stick travels far enough to complete the shift. For instance in shifting from second to third gear, dipping of the clutch 206 such that clutch slip occurs can be initiated when it is detected that the gear stick has been moved out of the second gear position but before it has been moved fully into the third gear position; whereby a gear shift inside the gear box 208 from the second to third gear ratio only occurs after the gear stick has been fully moved into the third gear position at which time clutch slip will be occurring.

In automatic transmission systems, the instructor 230 could form part of a control system for determining when gear ratio shifts are to be implemented. When such a gear ratio shift is required, as determined by the control system, the instructor 230 sends information to the controller 216 indicative of the gear ratio shift to be implemented. A control system of this type could include a lookup table stored in non-volatile memory which associates respective gear ratios (e.g. second gear, third gear etc.) with both a maximum and minimum threshold rotational speed for the gearbox input shaft 207 (or engine

crankshaft) . When a car is in motion the control system compares the measured rotational speed of the gearbox input shaft 207 (or crankshaft) with the threshold speeds stored in the lookup table for the particular gear ratio that the car is driving in. If the measured rotational speed is determined to exceed the maximum threshold speed then the instructor 230 sends information to the controller 216 indicative that an upshift in gear ratio is to be implemented.

Alternatively if the rotational speed of the gearbox input shaft 207 (or the crankshaft) is determined to move below the minimum threshold speed for the particular gear ratio that the vehicle is driving in, the instructor 230 sends information to the controller 216 indicative that a downshift in gear ratio is required.

In some embodiments the controller 216 itself may be configured to determine whether a gear ratio shift is to be implemented as described in the foregoing paragraph, in which case the instructor 230 is not necessary. A power supply source 232 (for example a battery) is provided for powering the various electrical components of the car.

The functionality that the gear ratio shift management application 219 causes the controller 216 to implement when executed will now be described with reference to Fig. 14. It should be noted that this functionality is only initiated when the car is in motion (e.g. in first, second or third gear etc. and already moving) . In other words, this functionality is not implemented while the car commences moving from a standing start. It is instead used to manage gear ratio shifts when the car is already being driven in a particular gear (e.g. the car is being driven in first, second or third gear etc.) .

In step Si, while the car is moving (in other words while torque is being transferred between the engine 200 and the wheels 202) the controller 216 determines that a gear ratio shift is to be initiated.

This could involve the controller 216 receiving information from an instructor 230, of the kind previously described, indicative that a particular gear ratio shift is to be initiated. Alternatively however step Si could involve the controller 216 itself determining whether a gear ratio shift is to be initiated, in addition to whether it should be an upshift or a down shift. In particular on the basis of information received from one or more sensors the controller 216 could compare the rotational speed of either the engine crank shaft or the gearbox input shaft 207 with maximum and minimum threshold values thereof for the particular gear ratio that the vehicle is currently driving in. If the measured rotational speed is determined to exceed the maximum threshold speed for the particular gear ratio that the car is currently driving in then the controller 216 determines that an upshift is to be initiated. However if the measured rotational speed is determined to be below the minimum threshold speed for the particular gear ratio that the car is currently driving in then the controller 216 determines that a downshift is to be initiated. The threshold values required to implement this functionality can be stored in a lookup table in the non-volatile memory 218, the lookup table associating respective gear ratios with maximum and minimum threshold values of engine crank shaft, or the gearbox input shaft 207, rotational speed.

In step S2 of the method in Fig. 14 the controller 216 communicates with the clutch actuator 226 and causes it to gradually reduce the pressure between the engine flywheel 212 and the clutch plate 210. More specifically the clutch actuator 226 causes the force exerted by the clutch plate 210 on the engine flywheel 212 to be gradually reduced. The exact rate at which force exerted by the clutch plate 210 on the flywheel 212 is reduced is not essential, however, the clutch 206 should be able to be dipped without overshoot, and to give an example of a suitable rate some embodiments could be configured such that steps Si toS4 in Fig. 14 are completed within approximately 40 to 60

milliseconds (e.g. 50 milliseconds).

In the example of the diaphragm clutch described in connection with Figs. 13a and 13b, step S2 involves gradually increasing the force exerted by the throw-out bearing 311 on the diaphragm spring 304, such that the diaphragm spring 304 flexes and draws the pressure plate 306 away from the clutch plate 308.

In step S3, while the force exerted by the clutch plate 210 on the engine flywheel 212 is being gradually reduced the controller 216 determines if static friction between the clutch plate 210 and flywheel 212 is overcome. If so clutch slip will be occurring, meaning that the engine flywheel 212 and the clutch plate 210 will be rubbing against each other while they rotate at different rotational speeds. To implement step S3 the controller 216 monitors the rotational speed of the engine crank shaft (which rotates at the same speed as the flywheel 212) and the rotational speed of the gearbox input shaft 207 (which rotates at the same speed as the clutch plate 210). This is performed on the basis of information received from the first and second sensors 222 and 224, which is indicative of these rotational speeds respectively. As already mentioned, the controller 216 can process raw data received from such sensors 222, 224 to determine the required rotational speeds itself, or alternatively the first and second sensors 222, 224 can include some processing functionality which enables them to process raw data and to work out the rotational speeds such that information output to the controller 216 is indicative of values corresponding to the calculated rotational speeds. More specifically, in implementing step S3, while the force exerted by the clutch plate 210 on the engine flywheel 212 is being gradually reduced the controller 216 determines if the rotational speed of the crank shaft (and thus the engine flywheel 212) differs from that of the gearbox input shaft 207 (and thus the clutch plate 210) by more than a threshold amount. Such a threshold amount can be pre-stored in the non-volatile memory 218 and is ideally substantially zero because this will minimise power loss and also heat generated by friction arising from the clutch plate 210 and fly wheel 212 rubbing against each other. In practice however the threshold value can range from orpm to 100 rpm for instance, and in some embodiments the threshold value is 10 rpm.

Upon determining that the rotational speed of the engine crank shaft differs from that of the gearbox input shaft 207 by more than the pre-stored threshold amount, clutch slip will be occurring. As already mentioned this means that the engine flywheel 212 and the clutch plate 210 will be rubbing against each other while they rotate at different rotational speeds.

Reducing the force exerted by the clutch plate 210 on the engine flywheel 212 lowers the threshold of motion between the engaging surfaces of these two components; the threshold of motion being equivalent to the maximum static frictional force that can be generated by the interfacing surfaces of these two components. In other words, by reducing the force exerted by the clutch plate 210 on the engine flywheel 212 this lowers the maximum force of static friction that can be generated between them.

Therefore in a positive torque condition - while the flywheel 212 is being rotationally driven by torque generated by the engine 200, and while the pressure exerted by the clutch plate 210 on the flywheel 212 is being gradually reduced, at a certain point the torque exerted by the flywheel 212 on the clutch plate 210 due to friction between them will exceed the threshold of motion between the interfacing surfaces of these components. It is at this point that clutch slip occurs and the flywheel 212 will rotate faster than the clutch plate 210 such that they rub against each other.

Alternatively in a negative torque condition - while the clutch plate 210 is being rotationally driven by torque transferred to it from the wheels 202, and while the pressure exerted by the clutch plate 210 on the flywheel 212 is being reduced, at a certain point the torque exerted by the clutch plate 210 on the flywheel 212 due to friction between them will exceed the threshold of motion between the interfacing surfaces of these components. It is at this point that clutch slip occurs and the clutch plate 210 will rotate faster than the flywheel 212 such that they rub against each other.

Looking again at Fig. 14, if clutch slip is determined to be occurring the controller 216 implements step S4. However if clutch slip is not determined to be occurring the clutch actuator 226 continues reducing the force exerted by the clutch plate 210 on the flywheel 212 and step S3 is performed again.

In step S4, after clutch slip is determined to be occurring, the controller 216 communicates with the clutch actuator 226 and instructs it to refrain from further reducing the force exerted by the clutch plate 210 on the engine flywheel 212. As a result, the clutch plate 210 and flywheel 212 are maintained in a position relative to each other in which clutch slip is occurring.

As mentioned earlier, the exact rate at which force exerted by the clutch plate 210 on the flywheel 212 is reduced is not essential, however, some embodiments could be configured such that steps Si toS4 are completed within approximately 40 to 60 milliseconds (e.g. 50 milliseconds). In step S5, the controller 216 communicates with the shift actuator 228 and instructs it to implement the necessary gear ratio shift. If the gearbox 208 is a Q-shift gearbox this could involve causing a servo motor to rotate a shift shaft 9 by a required amount necessary to effect a change in gear ratio. However the shift actuator 228 could have various different configurations depending on the shifting mechanism of the gearbox 208, as already acknowledged.

Since clutch slip is controlled to a low level, the torque transmitted by the clutch 206 from the engine 200 to the rest of the driveline 204, and to the wheels 202, while the gear ratio shift operation takes place is substantially the same as before the shift operation was initiated. Therefore the car does not experience a loss of driving power while the gear ratio shift operation takes place. Thus if a car is accelerating and moving up through the available gear ratios (moving from first to second to third gear etc.), driving torque will be transferred from the engine 200 to the wheels 202 at all times throughout the acceleration process, thereby reducing the time it takes for the car to reach a particular speed. This is achieved while simultaneously improving passenger comfort and vehicle reliability by reducing gear change shock, which has already been mentioned.

As a reminder however this is because by causing clutch slip to occur before initiating a gear ratio shift within the gear box 208, a step change in engine speed does not occur following such a gear ratio shift. As a result, the magnitude of any impulses imparted through the driveline 204 following a gear ratio shift operation are reduced, which increases passenger comfort and improves car refinement. In other words, car passengers are less likely to feel that a gear ratio shift has taken place due to the reduction in magnitude of torque spikes inside the vehicle driveline 204 following a gear ratio shift operation. In step S6, after the required gear ratio shift has occurred inside the gearbox 208 the controller 216 receives information from the shift controller 228 confirming that the gear ratio shift has taken place.

It will be remembered that a gear ratio shift operation causes a change in the rotational speed of the clutch plate 210. Therefore, following a gear ratio shift the clutch plate 210 and flywheel 212 will be rubbing against each other and will be rotating at different speeds.

Subsequently in step 7, the clutch 206 is once again fully engaged (like before step S2 was implemented) and the specific steps carried out to achieve this in a controlled manner are now be described with reference to Fig. 15.

In step S7a, the controller 216 communicates with an engine control unit (not shown) and instructs it to change one or more operational parameters of the engine 200 such that the difference between the rotational speeds of the flywheel 212 and clutch plate 210 is caused to have a magnitude equal to or less than a threshold value T. The threshold value T is pre-stored in the non-volatile memory 218 and can range from orpm to approximately loorpm. The smaller the difference in rotational speeds between the flywheel 212 and clutch plate 210 that is achievable in practice the better (in other words the smaller the value of T is the better), since the magnitudes of impulses generated along the driveline 204 when the clutch 206 is suddenly fully engaged are proportional to the speed difference squared. More specifically the rotational speed of the flywheel 212 is substantially matched with that of the clutch plate 210. Following an upshift (e.g. a change from third to fourth gear) the rotational speed of the flywheel 212 will need to be reduced in order to substantially match it with that of the clutch plate 210.

Following a downshift (e.g. a change from fourth to third gear) the rotational speed of the flywheel 212 will need to be increased in order to substantially match it with that of the clutch plate 210.

Since the engine 200 is an internal combustion engine, examples of operational parameters that could be changed by the engine control unit for the above purpose include the throttle position, fuel/air ratio per ignition cycle, fuel injection rate, ignition timing and variable valve timing. The act of changing the throttle position to implement step S7a is something known as throttle blipping.

It will be appreciated that in some embodiments, rather than communicating with an engine control unit the controller 216 could instead communicate directly with components inside the engine 200 for causing a change to one or more operational parameters thereof.

In step S7D, while the rotational speed of the flywheel 212 is being changed the controller 216 compares it with that of the clutch plate 210 to determine whether the difference between these rotational speeds, denoted Δω, is equal to or less than the aforementioned threshold value T. The controller 216 does this by monitoring the rotational speed of the crank shaft (and thereby the engine flywheel 212) and that of the gearbox input shaft 207 (and thereby the clutch plate 210) and determining whether the difference between them Δω is equal to or less than the threshold value T. If it is then step S7C is implemented. If not then the rotational speed of the engine flywheel 212 is changed further and step S7b is performed again. In step S7C the controller 216 instructs the clutch actuator 226 to fully engage the clutch 206, which involves increasing the force exerted by the clutch plate 210 on the flywheel 212. In particular the pressure exerted by the clutch plate 210 on the flywheel 212 is increased by an amount such that a subsequent change in the rotational speed of the flywheel 212 or the clutch plate 210 results in a corresponding change in the rotational speed of the other. More specifically, the clutch actuator 226 returns the clutch 206 to the configuration it was in before step S2 was implemented and thereby returns the clutch 206 from a partially dipped configuration (in which clutch slip can occur between the clutch plate 210 and flywheel 212 when they rotate at different rotational speeds) to a fully engaged configuration (in which clutch slip does not occur when the rotational speed of the clutch plate 210 or flywheel 212 is changed).

In the example of the diaphragm clutch previously described in connection with Figs. 13a and 13b, restoring the diaphragm clutch from a partially dipped configuration to a fully engaged configuration involves allowing the full force of the diaphragm spring 304 to clamp the clutch plate 308 against the flywheel 302.

After step S7C has been implemented a subsequent increase in the rotational speed of the engine flywheel 212, caused by a driver pressing the vehicle throttle pedal, will result in a corresponding increase in the rotational speed of the clutch plate 210. As a result positive torque transferred by the gearbox 208 will increase in magnitude, causing the wheels 202 to turn faster and the car to accelerate. In another instance, after step Syc has been implemented, if a driver lifts their foot off the throttle pedal while the car is in motion then engine breaking will occur. More specifically the gearbox 208 will enter a negative torque condition in which rotation of the wheels 202 drives rotation of the gearbox input shaft 207 (and thus the clutch plate 210). The engine flywheel 212 will thereby be caused to rotate at the same speed as the clutch plate 210 due to the static friction between it and the clutch plate 210.

It is envisaged that in some embodiments the threshold value T used to implement step S7b could be adjusted, or a different value of T (also stored in the non-volatile memory 218) could be selected depending on the circumstances. Using an increased rvalue in step S7b for instance, essentially increases the acceptable magnitude of gear change shock that will occur when step S7C is implemented. This is because the magnitudes of impulses imparted along the driveline 204 when the clutch 206 is suddenly fully engaged are proportional to the value of Δω squared (as a reminder Δω being the magnitude of the difference between the rotational speeds of the clutch plate 210 and flywheel 212). In contrast, using a lower rvalue in step S7b essentially decreases the acceptable magnitude of gear change shock that will occur when step S7C is implemented. Some cars can be driven in a so-called sport mode, performance mode or enhanced mode etc. (hereafter "sport mode"), whereby if a car is operating in sport mode faster gear ratio shifts are desirable and drivers will more likely be prepared to accept increased gear change shocks to achieve faster gear shifts. By using an increased value of T for implementing step S7b, upon selecting to drive a car in sport mode, the time it will take to implement steps S7a to S7C will be reduced due to the smaller correspondence between the rotational speeds of the clutch plate 210 and flywheel 212 that is required before the clutch 206 is fully engaged. However the magnitudes of impulses imparted along the drive line 204 are increased, due to the bigger value of Δω when the clutch 206 is fully engaged. By operating the car in this manner, greater vehicular acceleration can be realised due to faster gear ratio shifts but passenger comfort will be somewhat reduced due to increased gear change shock. Although it is assumed that this will be acceptable to drivers intending to drive their car in a sport mode like manner.

By using a reduced value of Tfor implementing step S7b, upon not selecting to drive a car in sport mode, or upon selecting to drive the car in a comfort or normal mode, the time it will take to implement steps 7a to 7c will be increased relative to when driving in sport mode due to the greater correspondence between the rotational speeds of the clutch plate 210 and flywheel 212 that is required before the clutch 206 is fully engaged. However the magnitudes of impulses imparted along the drive line 204 will be lower in comparison to when driving in sport mode due to the smaller value of Δω when the clutch 206 is fully engaged. By operating the car in this manner, lower vehicular acceleration is realised due to a reduced speed of gear ratio shifts compared to when driving in sport mode, but passenger comfort is increased due to less gear change shock.

In implementing the previous two paragraphs, when driving in sport mode the controller 216 may select a rvalue from the non-volatile memory 218 intended to be used when driving in sport mode. However when not driving in sport mode the controller 216 may select a different rvalue from the non-volatile memory 218 that is intended to be used when not driving in sport mode. Alternatively, the value of T which is stored in non-volatile memory 218 may be a value intended to be used to implement step S7b when not driving in sport mode, however when sport mode is selected the value of T used to implement step S7b may be increased by a predetermined amount during processing.

In addition it is envisaged that in some embodiments the controller 216 could adjust the threshold value T used to implement step S7b on-the-fly based on how the car is being driven. For example, by monitoring the rate of change of throttle position or average engine revs, if either of these is determined to be high in comparison to normal driving conditions (or if the average throttle position is determined to be wide in comparison to normal driving conditions) the controller 216 could determine that the driver is driving in a sport-like manner. Also if the car is determined to be accelerating by at least a threshold rate, as determined from output from an accelerometer, the controller 216 could determine that the driver is driving in a sport-like manner. Upon determining that the car is being driven in a sport-like manner the car could automatically enter a sport mode, of the kind heretofore described, and the aforementioned threshold value T could be adjusted accordingly to achieve faster gear shifts - and thus faster vehicular acceleration - at the expense of increased gear change shock, which the driver is assumed to deem acceptable when driving in a sportlike manner.

On the contrary, if the controller 216 does not determine that the car is being driven in a sport-like manner, the threshold value T used to implement step S7b is selected with the aim of minimising gear change shock and improving drive quality, passenger comfort and car refinement.

Another way of implementing step S7 in Fig. 14 will now be described with reference to Fig. 16.

In step S7d the controller 216 instructs the clutch actuator 226 to gradually increase the force exerted by the clutch plate 210 on the flywheel 212. In the example of the diaphragm clutch previously described in connection with Figs. 13a and 13b, this involves causing the force of the diaphragm spring 304 to increasingly urge the clutch plate 308 against the flywheel 302.

While step S7d is being implemented, the controller 216 in step S7e monitors the magnitude of the rate of change of rotational speed of the flywheel 212, denoted hereafter as a, caused by friction between the rotating clutch plate 210 and flywheel 212. This is achieved using output from the first sensor 222 because the rate of change of rotational speed of the aforementioned crank shaft will be the same as that of the flywheel 212.

In step S7f the controller 216 causes the force exerted by the clutch plate 210 on the flywheel 212 to be adjusted such that the magnitude of the rate of change of rotational speed of the flywheel a is inside a predetermined range. Expressed mathematically this involves adjusting the force exerted by the clutch plate 210 on the flywheel 212 so that a_min≤ a≤ a_max; whereby this range is pre-stored in the non-volatile memory 218. If a exceeds a_max then the force exerted by the clutch plate 210 on the flywheel 212 is reduced, whereas if a drops below a_min then the force exerted by the clutch plate 210 on the flywheel 212 is increased.

The magnitudes of impulses imparted along the aforementioned driveline 204 are related to the rate of change of engine speed and so are related to a. Thus by controlling the rate of change of rotational speed of the flywheel 212 in step Syi, this enables the magnitude of impulses imparted along the driveline 204 to be controlled such that they remain within acceptable limits. As a result drive quality, passenger comfort and car refinement can be increased.

While the step Syi is being implemented the controller 216 monitors the difference between the rotational speeds of the clutch plate 210 and flywheel 212 Δω using output from the first and second sensors 222, 224. More specifically in step S7g the controller 216 determines whether Δω is equal to or less than a threshold value T in a similar manner to that already described in connection with step S7b. If it is then step S7I1 is implemented, whereas if not then the rotational speed of the flywheel 212 is further adjusted and step S7g is performed again. For completeness it is again mentioned that the threshold value T can range from orpm to approximately loorpm and that the smaller the difference in rotational speeds between the flywheel 212 and clutch plate 210 that is achievable in practice the better (in other words the smaller the value of T is the better, such that the rotational speeds of the clutch plate 210 and flywheel 212 are substantially matched), since the magnitudes of impulses generated along the driveline 204 when the clutch 206 is suddenly fully engaged are proportional to the speed difference Δω squared.

In step S7I1 the controller 216 instructs the clutch actuator 226 to fully engage the clutch 206, which involves causing the clutch actuator 226 to return the clutch 206 to the configuration it was in before step S2 was implemented. In other words the clutch actuator 226 returns the clutch 206 from a partially dipped configuration (in which clutch slip can occur between the clutch plate 210 and flywheel 212 when they rotate at different rotational speeds) to a fully engaged configuration (in which clutch slip does not occur when the rotational speed of the clutch plate 210 or flywheel 212 is changed). In the example of the diaphragm clutch described in connection with Figs. 13a and 13b, restoring the diaphragm clutch from a partially dipped configuration to a fully engaged configuration involves allowing the full force of the diaphragm spring 304 to clamp the clutch plate 308 against the flywheel 302.

By fully engaging the clutch 206 in the controlled manner described in connection with Fig. 16, it is possible to maximise vehicular acceleration. This is because steps S7d to S7I1 can be implemented even while the engine 200 is operating at maximum power output, whereby the throttle is fully open. Drive quality can thus be improved by reducing the magnitude of gear change shocks felt by passengers, while simultaneously maximising vehicular acceleration.

It is envisaged that in some embodiments the values a_min, a_max and/or the threshold value T used in steps Sji and S7g could be adjusted or alternatively different values (stored in non-volatile memory 218) could be selected depending on the circumstances. Using an increased r value in step S7g for instance essentially increases the acceptable magnitude of gear change shock that will occur when step S7I1 is implemented, whereas using a lower r value decreases this magnitude. As already mentioned, this is because the magnitudes of impulses imparted along the driveline 204 when the clutch 206 is suddenly fully engaged are proportional to the value of Δω squared. Furthermore, increasing both a_min and a_max enables the rotational speed of the flywheel 212 to be changed at a greater rate in step S7f. As a result the time in which it takes for the difference between the rotational speeds of the clutch plate 210 and flywheel 212 Δω to reach the threshold value T in step S7g is reduced. Overall this provides that faster gear shifts can be achieved, although at the expense of increased gear change shock magnitude due to the increased rate of change of engine speed following a gear ratio shift. However this is likely to be acceptable to drivers intending to drive their car in a sport mode, as heretofore described.

By using an increased value of T and/or increased values of a_min and a_max for implementing steps S7f and S7g, upon selecting to drive a car in sport mode, the time it will take to implement steps 7d to 7I1 will be reduced. This provides that greater vehicular acceleration can be realised due to faster gear shifts taking place but passenger comfort will be somewhat reduced due to increased gear change shock. Although it is assumed that this will be acceptable to drivers intending to drive their car in sport mode.

By using a reduced value of T and/or reduced values of a_min and a_max for implementing steps S7f and S7g, upon not selecting to drive a car in sport mode, or upon selecting to drive the car in a comfort or normal mode, the time it will take to implement steps S7d to S7I1 will be increased relative to when driving in sport mode. However the magnitudes of impulses imparted along the drive line 204 will be lower in comparison to when driving in sport mode. By operating the car in this manner, lower vehicular acceleration is realised due to a reduced speed of gear ratio shifts compared to when driving in sport mode, but passenger comfort is increased due to less gear change shock.

In implementing the previous two paragraphs, when driving in sport mode the controller 216 may select values of T, a_min and a_max from the non-volatile memory 218 intended to be used when driving in sport mode. However when not driving in sport mode the controller 216 may select different values of T, a_min and a_max from the non-volatile memory 218 that are intended to be used when not driving in sport mode. Alternatively, the values of T, a_min and a_max stored in non-volatile memory 218 may be values intended to be used when not driving in sport mode, however when sport mode is selected such values used to implement steps S7f and S7g may be increased by a predetermined amount.

In addition it is envisaged that in some embodiments the controller 216 could adjust the threshold value T and/or the values a_min and a_max used to implement steps S7f and S7g on-the-fly based on how the car is being driven. For example, by monitoring the rate of change of throttle position or average engine revs, if either of these is determined to be high in comparison to normal driving conditions (or if the average throttle position is determined to be wide in comparison to normal driving conditions) the controller 216 could determine that the driver is driving in a sport-like manner. Also if the car is determined to be accelerating by at least a threshold rate, as determined from output from an accelerometer, the controller 216 could determine that the driver is driving in a sport-like manner. Upon determining that the car is being driven in a sport-like manner the car could automatically enter a sport mode and the aforementioned threshold value T and/or the values of a_min and a_max could be adjusted accordingly to achieve faster gear shifts - and thus faster vehicular acceleration - at the expense of increased gear change shock magnitude, which the driver is assumed to deem acceptable when driving in a sport-like manner. On the contrary, if the controller 216 does not determine that the car is being driven in a sport-like manner, the threshold value T and the values a_min and a_max used to implement steps S7f and S7g are selected with the aim of minimising gear change shock and improving drive quality, passenger comfort and car refinement.

It will be appreciated that whilst various aspects and embodiments of the present invention have heretofore been described, the scope of the present invention is not limited to the embodiments set out herein and instead extends to encompass all other arrangements, and modifications and alterations thereto, falling within the spirit and scope of the appended claims.

For instance, whilst illustrative embodiments have been described as employing software it will be appreciated by persons skilled in the art that the functionality provided by such software may instead be provided by hardware (for example by one or more application specific integrated circuits), or indeed by a mix of hardware and software.

In some embodiments the first sensor 222 does not measure the rotational speed of the engine crankshaft and is instead arranged to measure the rotational speed of another engine component having a rotational speed related to that of the flywheel 212, such as a cam shaft. Alternatively the first sensor 222 could measure the rotational speed of the flywheel 212 itself. In particular it is known to use a speed sensor on a flywheel 212 for control of engine ignition in combustion engines and this sensor could function as the first sensor 222.

In some embodiments the second sensor 224 does not measure the rotational speed of the gearbox input shaft 207 and is instead arranged to measure the rotational speed of another component of the driveline 204 having a rotational speed related to that of the clutch plate 210, such as the lay gear of a gearbox. Furthermore a magnetoresistance sensor could be positioned next to any one of the gears in the gearbox and sense the movement of the gear teeth themselves providing a pulsed output, the frequency of which is proportional to the gearbox input speed (and thereby the rotational speed of the clutch plate 210) .

Alternatively the second sensor 224 could measure the rotational speed of the clutch plate 210 itself.

In the embodiments mentioned in the previous two paragraphs, the heretofore described threshold values stored in the non-volatile memory 218 for use in step S3, S7b and S7g will concern the rotational speeds of the relevant components monitored.

In some embodiments step S3 could instead involve determining if the rate of change of the rotational speed of the crank shaft (or other engine component having a rotational speed related to that of the flywheel 212) differs from that of the gearbox input shaft 207 (or other driveline component having a rotational speed related to that of the clutch plate 210) by more than a threshold amount.

Although the engine 200 is mentioned to be an internal combustion engine, in other embodiments (e.g. in the case of an electrically powered car) the engine 200 could be an electric motor. In such embodiments step S7a could involve changing at least one operational parameter of the electric motor selected from the list of: voltage or current supplied to a driven component of the electric motor; and voltage or current supplied to a stator of the electric motor. In other embodiments the engine 200 could be any other type of driver capable of generating torque (such as a hydraulic motor, pneumatic motor, or heat engine such as a gas turbine) and step S7a involves changing at least one operational parameter thereof.

It will therefore be appreciated that the specific nature of the engine 200, and the operational parameter(s) thereof varied in step S7a are not essential.

In some embodiments a combination of the teachings described in connection with Figs. 15 and 16 could be used, particularly when implementing gear ratio shifts at relatively high but not full throttle. With reference to Fig. 6, embodiments have been described by explaining that an internal combustion engine, electric motor or other driver causes a torque to be exerted on the main shaft la (a gearbox input shaft), which is transferred to a lay gear 8 (a gearbox output shaft) when the gear box is in a positive torque condition. However, in other embodiments the gearbox could be provided the other way around such that an internal combustion engine, electric motor or other driver causes a torque to be exerted on the lay gear 8 (in this case, the gearbox input shaft) which is then transferred to the main shaft la (in this case, the gearbox output shaft) when the gear box is in a positive torque condition.

The method described in connection with Figs. 14 to 16 concerns a way of optimising the drive experience of passengers and optimising the life duration of vehicle drive line components. However a way of improving the drive experience of passengers and the life duration of driveline components to a lesser extent, and thus a method of managing gear ratio changes in a less than optimal manner, is to manage gear ratio shifts by partially dipping the clutch 206 until clutch slip occurs, implementing a gear ratio shift and then re-fully engaging the clutch 206. If the rotational speed of the clutch plate 210 and flywheel 212 are not caused to substantially match before re-fully engaging the clutch 206 (like in steps S7b and S7g) then when the clutch 206 is fully engaged a greater torque spike will be imparted along the driveline 204 than otherwise would have been had such rotational speeds been substantially matched. Nevertheless over the course of the whole gear ratio shift process, from detecting that a gear ratio shift is to be initiated to re-fully engaging the clutch 206, the magnitude of impulses imparted along the driveline 204 will be less than would otherwise have occurred had the clutch 206 not been partially dipped in the first place to cause clutch slip prior to initiating the gear ratio shift. This is because, as already mentioned, implementing a gear ratio shift inside a gearbox 208 causes a step change in the rotational speed of the clutch plate 210. The occurrence of clutch slip due to the clutch 206 being partially dipped prior to implementing a gear ratio shift provides that the step change in clutch plate rotational speed is not transferred to the flywheel 212 while the gear ratio shift takes place, thereby reducing gear change shock.

A further way in which gear ratio shift operations may be managed in another less than optimal manner is by partially dipping the clutch 206 - although by such an amount that clutch slip does not actually occur, then implementing a gear ratio shift and subsequently re-fully engaging the clutch 206. Notably in this method the clutch 206 is dipped by an amount such that the resulting step change in rotational speed of the clutch plate 210 causes it to slip relative to the flywheel 212 only after the required gear ratio shift has occurred inside the gearbox 208, not before. In this method also therefore, the step change in clutch plate rotational speed is not transferred to the flywheel 212 while the gear ratio shift takes place, thereby reducing gear change shock. As a result over the course of the whole gear ratio shift process, from detecting that a gear ratio shift is to be initiated to re-fully engaging the clutch 206, the maximum magnitude of impulses imparted along the driveline 204 will be less than would otherwise have occurred had the clutch 206 not been partially dipped in the first place prior to initiating the gear ratio shift.

Claims

1. A method comprising the steps of:

in response to determining that a gear ratio shift is to be initiated while a vehicle is in motion reducing pressure exerted between a first rotating body and a second rotating body, wherein the first rotating body is rotatably driven by a drive mechanism, and wherein the second rotating body is caused to spin due to friction between it and the first rotating body and comprises part of a linkage for transferring torque to an output portion for propelling the vehicle;

implementing a gear ratio shift in a gear arrangement of the linkage; and restoring the pressure exerted between the first rotating body and the second rotating body.

2. The method of claim l, wherein the pressure exerted between the first rotating body and the second rotating body is reduced until they are determined to be slipping against each other based on output from rotational speed sensors.

3. The method of claim 1 or 2, further comprising the steps of:

after the gear ratio shift adjusting the relative difference between the rotational speeds of the first and second rotating bodies such that it complies with a predetermined condition using rotational speed sensors; and

then increasing the pressure exerted between the first and second rotating bodies such that a subsequent change in the rotational speed of one of them results in a corresponding change in the rotational speed of the other.

4. The method of claim 3, wherein the step of adjusting the relative difference between the rotational speeds of the first and second rotating bodies involves changing at least one operational parameter of the drive mechanism.

5. The method of claim 4, wherein the drive mechanism comprises an internal combustion engine and the at least one operational parameter is selected from the list of: throttle position, fuel/air ratio per ignition cycle, fuel injection rate, ignition timing and variable valve timing.

6. The method of claim 4, wherein the drive mechanism comprises an electric motor and the at least one operational parameter is selected from the list of: voltage or current supplied to a driven component of the electric motor; and voltage or current supplied to a stator of the electric motor.

7. The method of claim 3, wherein the step of adjusting the relative difference between the rotational speeds of the first and second rotating bodies involves varying the pressure exerted between them and thereby the friction between them.

8. The method of claim 7, wherein the pressure is varied such that the rate of change of rotational speed of the first rotating body is controlled to substantially comply with another predetermined condition.

9. The method of claim 7 or 8, wherein the drive mechanism is operating at substantially maximum torque output.

10. The method of any of claims 3 to 9, wherein the or each said

predetermined condition is dependent on a mode of operation of the vehicle.

11. The method of claim 10, wherein the mode of operation is selected by a user, or is determined on-the-fly based on output received from at least one sensor on board the vehicle indicative of how the vehicle is being operated.

12. The method of claim 11, wherein the output received is indicative of a usage property of part of the drive mechanism, or is indicative of the magnitude of vehicle acceleration.

13. The method of claim 12, wherein the usage property is the throttle position and/or engine revs.

14. The method of any of claims 2 to 13, wherein the first and second rotating bodies are determined to be slipping against each other if their rotational speeds differ from each other by at least a threshold amount.

15. The method of any of claims 2 to 13, wherein the first and second rotating bodies are determined to be slipping against each other if the rates of change of their rotational speeds differ from each other by at least a threshold amount.

16. The method of any preceding claim, wherein the gear arrangement is configured such that during the gear ratio shift the amount of torque transferred between an input shaft and an output shaft thereof does not reduce to zero .

17. The method of claim 16, wherein at least one gear is provided on either the input shaft or the output shaft and wherein two selector members are associated with the or each gear respectively, the or each gear comprising a first and second face, and further comprising, on each face, at least one projection, and the or each selector member comprising, on at least one face thereof, at least one complementary projection arranged to selectively engage with a projection of a gear, the arrangement being such that the projection(s) of a selector member and the projection(s) of the first face of a gear may be drivingly engaged so as to transfer torque in a first rotational sense and the projection(s) of a selector member and the projection(s) of the second face of a gear may be drivingly engaged so as to transfer torque in a second rotational sense, opposite to the first rotational sense.

18. A computer program comprising computer executable instructions which when executed by computing apparatus causes the computing apparatus to perform the method of any preceding claim.

19. A non-transitory computer readable storage medium having stored thereon computer-readable code which when executed by computing apparatus causes the computing apparatus to perform the method of any of claims 1 to 17.

20. Apparatus for managing a gear ratio shift in a vehicle comprising:

means for reducing pressure exerted between a first rotating body and a second rotating body in response to determining that a gear ratio shift is to be initiated while a vehicle is in motion, wherein the first rotating body is configured to be rotatably driven by a drive mechanism, and wherein the second rotating body is configured to be caused to spin due to friction between it and the first rotating body and comprises part of a linkage for transferring torque to an output portion for propelling the vehicle;

means for implementing a gear ratio shift in a gear arrangement of the linkage; and means for restoring the pressure exerted between the first rotating body and the second rotating body.

21. The apparatus of claim 20, wherein said means for reducing pressure exerted between the first rotating body and the second rotating body is configured to reduce the pressure exerted between the first rotating body and the second rotating body until they are determined to be slipping against each other in use based on output from rotational speed sensors.

22. The apparatus of claim 20 or 21 further comprising:

means for, after the gear ratio shift, adjusting the relative difference between the rotational speeds of the first and second rotating bodies such that it complies with a predetermined condition using rotational speed sensors; and means for increasing the pressure exerted between the first and second rotating bodies such that a subsequent change in the rotational speed of one of them results in a corresponding change in the rotational speed of the other.

23. Apparatus for managing a gear ratio shift in a vehicle, substantially as hereinbefore described with reference to the accompanying drawings.