The present invention relates to a transmission unit for relaying drive from a crankshaft of an internal combustion engine to engine ancillaries

Internal combustion engines typically have a pulley mounted at one end of the driven crankshaft of the engine. The pulley is connected by a belt to front end ancillaries of the engine, e.g. an alternator, a water pump, an oil pump, a power-assisted steering pump, a supercharger. Traditionally, the transmission ratio between the speed of rotation of the crankshaft and the speed of rotation of each ancillary device is fixed for all engine speeds and loads and is a function of the size of the pulley mounted on the front of the crankshaft and the sizes of the pulleys associated with the individual engine ancillaries. This is disadvantageous because each ancillary device typically has to be sized so that it can operate efficiently at low engine speeds and thus each device is effectively oversized for high speed operation. For instance, an alternator has to be rated to give a certain amount of power. The worst case scenario must be chosen, which is at low engine speeds, when the alternator rotates slowly. The alternator must be rated to provide the required power at these low speeds of rotation. Consequently, it is overrated when high engine speeds is considered. Savings can be made both in cost and in weight by reduction in the size of and rating of an alternator. From an alternative perspective the provision in existing engines of a fixed transmission ratio between crankshaft and engine ancillaries is disadvantageous because the engine ancillaries can take up to 30% of an engine's output, whilst full output from the ancillaries is not always needed, e.g. full output is not needed from a supercharger when the engine is idling or from an alternator when the engine is idling and the battery associated with the engine is fully charged. Thus it is desired to .speed down the engine ancillaries in certain engine operating conditions to reduce the power they take from the engine and thus increase engine efficiency and reduce fuel consumption.

A solution to the above problem has been proposed in US 2005/0153813, in which an epicyclic gear arrangement is used to transmit rotation from an input pulley, driven by an internal combustion engine, to an output pulley which drives the front end ancillaries. The epicyclic gear arrangement has an electromagnetic brake which enables a transmission ratio change based upon an engine speed signal. However, the solution proposed is quite complex and requires a large electromagnet, which increases the weight of the device, along with the cost. More importantly, the axial size of the device is very large, hindering the mounting of the device .

According to a first aspect of the present invention, there is provided a transmission unit as claimed in claim 1.

According to a second aspect of the present invention there is provided a transmission unit as claimed in claim 11. In an elegant simple package, embodiments of the present invention provide a well lubricated ' engine ancillary two-speed drive mechanism mountable on an end face of an engine between the engine crankshaft and the pulley from which drive is transmitted to the engine ancillary devices. Thus the device can be used with existing engines with minimal redesign.

A preferred embodiment of the present invention will now be described with reference to the accompanying drawings in which:

Figure 1 shows a first embodiment of a transmission unit according to the invention;

Figure 2 shows a second embodiment of a transmission unit according to the invention;

Figure 3 shows a third embodiment of a transmission unit according to the invention;

Figure 4 shows a fourth embodiment of a transmission unit according to the invention; and

Figures 5a and 5b are respectively a simplified transverse and a simplified axial cross-section through a fifth embodiment of a transmission unit according to the invention.

Figure 1 shows a first embodiment of an engine ancillary drive mechanism 10 which comprises a pulley 11 which is capable of driving a belt (not shown) which relays drive to ancillary devices of an engine, such as an alternator, water pump, oil pump, power-assisted steering pump, an air-conditioning pump and supercharger. The pulley 11 is of a diameter chosen to give a required transmission ratio between an engine crankshaft and a front end accessory device of the engine; e.g. the pulley could be 112mm to 160mm in diameter. Also shown in the figure is an input shaft 12. This shaft will either be an end part of a crankshaft of an internal combustion engine, or will be a separate shaft connectable to a crankshaft of an internal combustion engine to rotate with the engine crankshaft.

A static casing 13 is illustrated in the Figures. It comprises a planar flange portion 13a which extends as an annulus around the transmission unit 10. From the inner edge of the flange 13a there extends a boss 13b having an annular section 13c extending perpendicular to the flange 13a, a front face 13d, extending radially inwardly from the annular section 13c in parallel with the flange 13a and spaced from the flange 13a by the annular section 13c, and a lip 13e extending away from the radially innermost edge of the front face 13d back towards the flange 13a, in parallel with the annular section 13c. The lip 13e provides an annular surface which defines an aperture through the front face 13d. The annular lip 13e and the annular section 13c both have a common central axis which is coincident with the rotational axis of the input shaft 12.

The pulley 11 comprises an annular belt engaging section 11a which takes the axis of rotation of input shaft 12 as its axis of rotation. This annular section 11a overlies the annular section 13c of boss 13b. Extending inwardly from the annular outer section 11a is a web lib, which extends perpendicularly to the annular section 11a. The web lib connects the annular portion 11a with a sleeve portion lie of the output pulley, the sleeve portion lie extending in parallel with the annular outer portion 11a and within the radially outer portion 11a.

The sleeve portion lie is fixed to an annulus gear 14. The annulus gear 14 rotates with the output pulley 11. The output pulley 11 is journalied for rotation in the unit 10 by bearings 15 which act between the annular lip 13e and the sleeve portion lie of the output pulley 11. Seal 17 is provided between the annular 'lip 13e and the sleeve portion lie. The seal 17 prevents escape of lubricant from the unit 10.

Also shown in the Figures are planet gears 19 which are rotatably mounted on a planet carrier 20. The planet carrier 20 is an annular member extending from and fixed to the input shaft 12 for rotation therewith. A sun gear 40 of the epicyclic arrangement of gears of the unit is integrally formed with an annular flange 25. The planets 19 mesh with both radially inward facing teeth of the annulus gear 14 and radially outward facing teeth of the sun gear 40.

The annular flange 25 has a radially outward annular friction disc carrier sectior 25a, with external grooves. Friction discs 21 are mounted in the grooves so that the friction discs 21 rotate with the planet carrier 19 while being slidable axially along the grooves. Each friction disc 21 can engage at least one non-rotatable brake disc 22. The plurality of non-rotating brake discs 22 are mounted in grooves provided on the interior of the annular section 13c of the boss 13b. They are slidable axially along the boss 13b but cannot rotate relative to the boss. An annular recess 42 and a passage 41 are formed within the boss 13b. The annular recess is defined by the radially inner surface of the annular section 13c, the axially inner surface of the front face 13d, and the radially outer surface of the annular lip 13e. The annular recess defines an axis longitudinally through the centre of its circular cross-section. The passage 41 connects the annular recess 42 with a port 48 formed in the casing. Specifically, the passage 41 extends in an axial direction from the engine side of the casing 13 through the annular section 13c.

An external pump (not shown) via a valve (also not shown) for pumping hydraulic fluid may be connected to the port 48, thereby communicating with' the annular recess 42 via the passage 41. In preferred embodiments the valve is operable to connect the port 48 to either the pump, or to a sink, to thereby either connect annular cavity 42 to a pressurised fluid source or to allow fluid to leave the cavity to release the pressure.

A piston 44 in the form of an annulus is slidably located within the annular recess 42. The annular recess 42 and an end face of the piston 44 together define a cavity into which hydraulic fluid may be pumped via the passage 41.

The piston and annular recess are accurately machined for a close fit. 0-rings 45 are formed in a groove on the outer surface of the piston.

The pressure of the fluid within the cavity acts on the piston to thereby apply a force in the axial direction (the axial direction being the axis of rotational symmetry of the annulus-shaped piston, which corresponds in this embodiment with the axis about which the input shaft 12 rotates) .

The assembly of rotatable friction discs 21 and non- rotating brake discs 22 is sandwiched between the piston 44 and a fixed end stop 43 secured to an end wall 55. The application of fluid pressure to the end face of the piston 44 applies an axial force via the pusher to the assembly, pressing the friction discs 21 and brake discs 22 together, to thereby prevent relative rotation therebetween. The axial force is resisted by the end stop 43. Thus, the application of hydraulic pressure can prevent rotation of the sun gear 40 relative to the casing 13.

Preferably, the friction disks 21 and brake discs 22 are arranged alternately in the axial direction, with a non- rotating brake disc 22 at either end of the arrangement.

In preferred embodiments, the annular piston 44 may comprise a cylindrical pusher 46 for transferring the force applied by the hydraulic fluid to the assembly of rotatable friction discs 21 and non-rotating brake discs 22. However, in other embodiments, the piston may simply be extended to directly apply a force to the assembly.

A cylindrical member 12a extends from the planet carrier 20 in the opposite direction to the shaft 12. A sprag clutch 30 acts between the cylindrical member 12a and the sleeve portion lie of the output pulley 11 and functions as a one-way clutch allowing relative rotation between the input shaft 12 and the pulley 11 only in one direction of rotation. The clutch 30 allows the pulley 11 to rotate faster than the input shaft 12, but prevents the pulley 11 from rotating slower than the input shaft 12.

If the teeth of the annulus 14, the planets 19 and the sun 4 are all helical then an axial force arises from the transmission of torque from the planet carrier 20 to the sun gear 40. The applied torque gives rise to an axial force on the sun gear 40 and hence the annular flange 25. The axial force on the annular flange 25 generated by a transmission of torque through the planet gears 19 is resisted by a thrust bearing 31. Thrust bearing 31 is mounted on the end wall 55. Helical gears tend to be less efficient than spur gears, but quieter. If spur gears are used for the annulus 14, the planets 19 and sun 40 then the transmission unit will be more efficient and cheaper and the thrust bearing 31 could be just a reduced friction surface, since no gear- generated axial forces would need to be reacted.

In a first mode of operation of the transmission unit 10, the valve is operated to allow the pump to communicate with the annular cavity 42 so that fluid pressure is applied to the end face of the hydraulic piston 44 via the passage 41. The hydraulic pressure acts on the piston so that the friction discs 21 and brake discs 22 are pressed together, thereby preventing rotation of the sun gear 40.

In this mode, the input .shaft 12 rotates relative to the static sun gear 40, rotating the planet carrier 20. Accordingly, the planet gears 19, which mesh with the sun gear 40, rotate without slip about _t the static sun gear 40. The planet gears 19 mesh with the annulus 14, and therefore rotation is transmitted from the input shaft 12 via the planet carrier 20 and the planet gears 19 to the annulus 14 and thereby to the pulley 11 and onwards to the belt engaged by the pulley. This achieves a transmission ratio whereby the speed of rotation of the pulley 11 is a multiple of the speed of rotation of the input shaft 12 and the pulley 11 rotates faster than the input shaft 12, e.g. 1.1 to 1 or 1.4 to 1.

In a second mode of operation of the transmission unit 10, the valve is operated so that the pump does not communicate with the annular cavity 42 and the pressure within the cavity is removed. Thus, insufficient fluid pressure is applied to the end face of the hydraulic piston 44 via the passage 41, so that the friction discs 21 and brake discs 22 are not pressed together, thereby allowing rotation of the sun gear 40. It may be desired to incorporate a return spring in the unit 10 acting on the piston 44 to bias the piston 44 to disengage the brake.

In this second mode, the free rotation of the sun gear 40 means that torque ceases to be transmitted from the planet gears 19 to the annulus 14 and instead the one-way clutch 30 locks the pulley 11 to rotate with the input shaft 12 at the same rotational speed.

The unit 10 operates such that when the brake is engaged the engine ancillaries are driven at rotational speeds which are a first set of multiples of crankshaft speed and when the brake is disengaged then the front end ancillaries are driven to rotate at rotational speeds which are a second set of lesser multiples of crankshaft speed. This can be done to reduce power consumption of the ancillaries when the engine is idling or so that at lower engine speeds the ancillaries are driven at higher multiples of crankshaft speed to increase their efficiencies.

The absence of hydraulic pressure alone is sufficient to allow free movement of the sun gear 40, although use of a return spring having the piston 44 to disengage the brake could be preferred. The use of helical threads in the epicyclic arrangement of gears is optional and spur gears could equally well be used.

The operation of the unit 10 will be controlled so that the brake is disengaged for selected engine operating ¹ conditions in which it is desired to "speed down" the engine ancillaries. For instance when an engine is idling and the battery associated therewith is fully charged then the alternator need not be rotated at crankshaft speed. Also a supercharger need not be rotated at crankshaft speed when an engine is idling. This "speeding down" reduces the power consumption of the engine ancillaries and thus reduces the fuel consumption of the engine. The operation of the unit 10 will be controlled by an electronic engine management system (not shown) which will control operation of the valve (which will be electrically operated) which controls connection of the chamber 41 to the pump or to a fluid sump.

It is envisaged that the flange 13a will be attached to a front cover of an internal combustion engine using fixing means such as bolts or rivets 3. Alternatively, the flange 13a and the boss 13b extending therefrom could be formed as an integral part of the engine front cover. Typically a sealing gasket would be used between the opposing faces of - li ¬

the unit 10 and the front of the internal combustion engine, to ensure a good fluid seal. ¹ In either event, a plurality of holes are formed in the circular end wall 55 so that the aperture defined by the flange 13a will expose the epicyclic gear arrangement of the invention to lubricating oil of the crank case of the internal combustion engine so that lubricating oil can flow directly into and out of the epicyclic gear arrangement from the crank case. (Alternatively, the end wall 55 may be a frame, supporting the thrust bearings 31 and the end stops 43, whilst allowing the flow of lubricant from the engine to the transmission unit 10.) Thus, there is no need for the engine to be provided with specific oil flow passages, nor is there any need for the epicyclic gear box to be provided with specific lubricant flow passages. This is made possible by having the casing 13 surround and encase the brake assembly of brake discs 22, friction discs 21 and also surround the epicyclic arrangement of planet carrier 20, annulus 14, planets 19 and sun gear 40. ;

In a preferred embodiment, the plurality of holes may be formed through one segment of the circular end wall 55 only. That segment would be aligned with the pistons of the engine to receive lubricant therefrom.

During rotation of the epicyclic arrangement the gears thereof tend to throw lubricating oil radially outwardly away from the input shaft 12. This radially outwardly expelled oil is then caught by the inner surface of the boss 13b of the casing 13 and is kept within the mechanism and/or returned to the engine crank,. It is not lost outside the mechanism. The bearings 15 and the seal 17 prevent escape of lubricant from outside the cover 13. Thus, having the epicyclic gear components encased radially within the boss 13b of the casing 13 enables the creation of an elegant simple design which requires a minimum of parts and does not require special provision for lubrication.

Figure 2 shows a second embodiment of an engine ancillary drive mechanism 10 which includes the same components as the first embodiment but also additionally comprises a pump 50, held within the casing for supplying fluid to the annular cavity 42.

Pump 50 is a pump of known type, such as a vane pump. A valve (not shown) controls the passage of hydraulic fluid from the pump 50 to the annular cavity 42.

As shown in Figure 2, pump 50 is mounted on and driven directly by the input shaft 12. An electrically-operated valve will be incorporated into the pump 50 and will selectively connect the pump 50 to apply hydraulic pressure to the piston 44 to engage the brake.

Figure 3 shows a third embodiment of an engine ancillary drive mechanism 10. A reciprocating internal combustion engine does not apply a consistent torque over time. The actual torque applied to the drive shaft fluctuates over the course of the cycles of the pistons and can even apply a negative torque. This fluctuation creates undesirable vibrations of the shaft and the mechanisms to which it transmits torque. In order to reduce this effect, in the third embodiment, the transmission unit includes a damper 60, which is mounted on the input shaft 12 for rotation therewith.

As shown in Figure 3, this vibration damping mechanism is attached to the input shaft on the engine side of the transmission unit 10, so that the input to the unit 10 is damped. Advantageously, such a damper is retained within the casing 13. The advantage of mounting the damper on the engine side of the unit is that the torque "seen" by the unit 10 is damped, which reduces stress on the components of the unit 10.

Alternatively, in a fourth embodiment shown in figure 4, the damping mechanism may be attached to the output pulley 11, so that the output of the unit 10 is damped. Such an arrangement allows damping of the vibrations caused by the varying torque, without obstructing the flow of lubricant from the engine to the transmission unit 10. A suitable damping mechanism is supplied by Litens Automotive Partnership of 730 Rowntree Dairy Road, Woodbridge, Ontario, L4L 5T7, Canada and Litens Automotive U.K. of Suite F6, 110 The Causeway, Heybridge Business Centre, Maldon, Essex, CM9 4ND, United Kingdom, under the trade mark torqfiltr (TM) . This is a spring isolator mechanism with a spring whose stiffness can be turned to control resonant frequency. The mechanism is packaged to be located within an output pulley and is ideal to be located within the output pulley 11 as shown in Figure 4, acting between the output pulley and the remainder of the components of the transmission unit.

The pump 50 mentioned above may provide a damping effect smoothing the torque "seen" by the engine ancillaries, but only- when it operates to supply hydraulic pressure to the brake.

Although in the above disclosed embodiments the input shaft drives the planet carrier 20, the output pulley 11 is. driven by the annulus 14 and. the sun gear 40 is controlled by the hydraulic braking mechanism, it would be apparent to the skilled person that other arrangement of gears can be used. For example, the input shaft could drive the planet carrier, the annulus could be controlled by the hydraulic braking mechanism, and the output pulley 11 could be driven by the sun gear 40. This gives a different gear ratio.

Although the above disclosed embodiments have been described as having an annular shaped piston in an annular recess, the skilled person v/ould appreciate that a hydraulic brake could be effected by one or more separate cylindrical cavities, each having a piston, spaced around the periphery the front face 13d.

The one-way clutch 30 has been depicted in the Figures as located axially outward of the epicyclic arrangement of gears. However, alternatively, the input pulley 11 may additionally comprise a short shaft protruding axially inwards from the centre of the front face 13d thereof and penetrating the end of the input shaft 12 and the one-way clutch 30- may be mounted between this and the input shaft 12, so as to be located in the same plane as epicyclic gears .

Figures 5a and 5b show a fifth embodiment of transmission unit 50. The majority of the components of the unit 50 and the method of operation of the unit 50 are identical to previous embodiments described. The difference is in the braking mechanism. Instead of using a piston acting on a plurality of plates, the unit 50 has an . arrangement of cylinder 51 and piston 52 which act via the lever arm 53 (pivotally mounted in the unit 50 for pivoting about a pivot pin 54) on a brake band 55 which engages a brake drum 56 connected to the sun gear, the brake band 55 extending around the brake drum 56 between an end 53a of the lever arm 53 and a fixed retaining pin 57. When pressurised hydraulic fluid is supplied to the arrangement of piston 52 and cylinder 51 then a force is applied to the brake band 55 which causes it to engage the brake drum 56 and thereby brake the sun gear. When the force is removed from the brake band, by connecting the piston and cylinder arrangement to a fluid return (a valve, not shown, will be used to control the hydraulic pressure relayed to the piston and cylinder arrangement) , then the band 55 will release the brake drum 56 to allow rotation of the sun gear. The lever arm 53 has two sides separated by the pivot point 54; the piston 52 is connected at an end 53b of a first side of a first length and the brake band 55 is connected at an end 53a of a second side of a second length significantly shorter than the first length. Thus a mechanical advantage is achieved. The lever 53 is provided with an elbow 53c, with the sections of the lever arm 53 on opposite sides of the elbow angled with respect to each other.