In Europe there is considerable growth in the application of automated manual transmissions (AMT) predominantly for reasons of cost and CO2 efficiency.

There is a range of such transmissions available, and they can generally be grouped into single clutch systems (AMT-1) and twin path systems (AMT-2).

Yet more sophisticated versions are under development. However, the poor shift quality of this"first wave"of manual-based transmissions, that is to say, the performance of the vehicle during gear changes, particularly as perceived by the driver is a significant problem.

In a related aspect, known designs of hydraulic supply for AMT powertains have concentrated upon known pump design for conventional automatic transmissions and the simple application of a known engine oil lubrication pumps. However this does not recognise that flow requirements for engines and transmissions are fundamentally different, AMTs having requirements for intermittent high pressure flow to control clutches and shift rails, whereas engines have lower pressure requirements and the flow rates are steadier.

According to the known proposals overall efficiency of electrical generation, storage to battery devices, and motoring efficiency is inevitably poor.

In a further aspect, to a great extent, conventional powerpack design is defined by the following characteristics: pump type (fixed displacement or variable displacement), pump characteristic (defined by volumetric efficiency, mechanical efficiency and flowrate), and torque source (fixed mechanical drive, clutched mechanical drive, electric motor and so forth). Within

automated transmissions of any type a high proportion of system losses are due to the energy consumption of the powerpack, whether mechanically or electrically driven. Within the majority of current transmissions the actuator system is hydraulic and the pump supply is mechanically driven. The most common arrangement comprises a fixed displacement pump driven by a fixed mechanical shaft. Whilst simple and inexpensive this design has poor overall efficiency arising from the requirement to satisfy the hydraulic requirements at both low and high engine speeds. In such a configuration, if a pump is sized to provide enough flow at low speeds then it will generate excess flow at higher vehicle speeds. Variable displacement pumps (vane, piston) could be applied to overcome this problem and have the potential to provide improved efficiency, but they are considerably more costly. Also, some claims of high efficiencies must be treated with caution since bypass flow to maintain pressure stability must be considered.

Problems also arise in conventional torque sensing arrangements in clutch packs.

Torque sensing requires the transducer to be in the torque bearing path and the majority of technologies rely upon sensing small changes in the magnetic properties of ferrous materials when subject to stresses. Thus the transducer becomes a treatment applied to a component and an associated pick-up device. To date, publicised devices have focused upon sensing of torque in shafts. However, this usually incurs a penalty of 30-40 mm in axial length, or more complicated packaging arrangements, and the system has a poor resolution at low torque.

The invention provides a hydraulic fluid pump for a vehicle transmission hydraulic system, the pump comprising a dual drive pump pressurisation source. The invention further provides vehicle transmission hydraulic system including a pump and a high pressure accumulator. In preferred embodiments the pump further comprises a dual drive electro-mechanical pump

pressurisation source in which the dual-drive pressurisation source is dual clutched for respective drives and a pump controller and a torque sensor providing feedback thereto.

The invention further provides a clutch for a transmission comprising a clutch hub and a torque sensor provided on the clutch hub. In a preferred embodiment the torque sensor is provided on a radially extending portion of the clutch hub, and/or the torque sensor is provided on an axially extending portion of the clutch hub.

As a result the invention achieves an overall objective of designs comprising both mechanical and electrical drives to allow improved matching between the hydraulic requirements of the transmission and the generation of pump output to meet both low and high pressure flowrate requirements, as provided by the hydraulic scheme and combined electrical/mechanical pump drive for the hydraulic pump. The efficiency of the transmission and sustained-slip performance may be improved through the use of clutched, dual-drive electro- mechanical pump systems with a high pressure accumulator, and electrical generation of cooling flow in a wet clutch design has been shown to allow sustained hill-hold on a 20% gradient, equivalent or improved shift quality may be achieved between a current automatic and a twin clutch AMT. The invention is thus directed to C02 efficiency through improved powerpack design and control, sustained clutch slip capability through improved powerpack design, and shift quality and characterisation through improved algorithms and sensor technology.

Embodiments of the invention will now be described by way of example with reference to the drawings of which:

Fig. 1 shows a hydraulic system according to the present invention; Fig. 2 shows a first drive system according to the present invention; Fig. 3 shows a second drive system according to the present invention; Fig. 4 shows a third drive system according to the present invention; Fig. 5 shows a drive system and a hydraulic scheme in a first mode according to the present invention; Fig. 6 shows the drive system of fig. 5 in a second mode; Fig. 7 shows the drive system of fig. 5 in a third mode; and Fig. 8 shows a torque sensor arrangement according to the present invention.

Throughout the description, common reference numerals relate to common elements. The basic transmission design discussed herein follows a simple twin clutch design with two lay shafts, two input shafts and pre-selection of gears using simple single-cone synchronizers controlled by hydraulic shift actuators. Both clutches are of a wet, multiplate type with static pistons, controlled hydraulically and for example incorporating a 7kW electrical motor/generator device for application within a mild parallel hybrid driveline. The design will be well known to the skilled person and is not discussed in detail here.

Within the AMT transmission two different types of flow are required: high pressure and low pressure for which estimated values are shown in Table 1. It should be noted that these are ideal values and would be subject to the inefficiencies of pressure control valves and general flow losses within the valve block.

An appropriate hydraulic system design is shown in Fig. 1. The system which is designated generally 10 includes a pump 12 driven by a drive and clutch arrangement 14 discussed in more detail below. The pump drives through a low pressure route 16 through a normally open valve 18 and a flow restriction

valve 20 forming a restriction and bypass system 22. The low pressure route runs to lubrication, gear meshes, bearings and so forth generally designated 24 and lubrication and cooling 26 for clutches 28a, 28b.

A high pressure line 30 runs through a one-way valve 32 to a pressure sensor 34 and pressure reducing valve 36a, 36b through to a clutch control arrangement designated generally 38. The high pressure line 30 further runs to a second one-way valve 40 and a second pressure reducing valve 42 to a shift rail actuation system generally designated 44. The high pressure line runs yet further through a pressure regulator 46 to a return line including a filter/tank/cooler system generally designated 48. Flow Characteristics Purpose Approximate Values High pressure, low flowrate Normal Running Average over MVEG (circa 10 Bar) # Valve Leakage : 1. 0 litres/min test # Clutch Seal Leakage : 0. 8 litres/min 2. 1 Litres/min 10 Bar Shift Rail Actuator # Control Pressures for : 0.3 litres/min DIA. Piston =35 mm clutches and shift rail # DIA. Rail =15mm actuators Travel = 10mm Volume = 2. 5 cc Total 2.1 litreslmin Total volume = disengage old gear + engage new gear = 5. 0 cc Rapid Shifting Maximum Clutch Assume maximum prolonged shift frequency of Instantaneous # OD piston = 60 mm 1Hz 3.8 litres/minute # ID piston = 30 mm 10 Bar Travel 0.9 mm Flow per shift # voume = 19 cc #Flow for shift rails 5. 0 cc Assume total volume = disengage. Flow for clutches : 30 cc old clutch + engage new clutch Total for Shifting : 35 cc/second = 1.5* 19cc = 30 cc Total for Shifting :2.1 litres/min Leakage Assumptions a Valve Leakage : 1. 0 litres/min # Per seal :0.15 litres/min # Clutch Leakage : 0. 7 litres/min Pervalve : 0. 10 litres/min Total :3.8 litres/min Low Pressure, high flowrate Normal Running Average over MVEG (circa 1.5 Bar) Lubrication and cooling: 2.5 litres/min test of gear-meshes and. 4.5 litres/min Assume that oil-to-air cooler has bearings. # 1.5 Bar sufficient heat rejection capacity to reject. Cooling of Open clutch : 2. 0 litres/min requisite energy at the assumed flowrates. Total :4.5 litres/min Sustained Slip Maximum Cooling of clutch during : 10 litres/min Instantaneous launch # 15 litres/min 1.5 Bar Specific cooling of clutch : 15 litres/min sustained slip, such as creep and hill-hold. Worst case : 15 litres/min

Table 1-Estimates of hydraulic requirements Also provided in conjunction with the high pressure line is an accumulator 50.

The accumulator is intermittently charged up to provide additional capability to meet high pressure demand, hence making use of additional pump capability from the pump 12. Yet further the bypass valve 22 can be closed if there is further demand allowing additional charging of the accumulator for example if additional shift rail demand is encountered. Alternatively the bypass valve 22

allows electrical generation of high flow-rate, for example, for clutch cooling during conditions of sustained slip as discussed in more detail below. The provision of the bypass system 22 and the accumulator 50 allow an improved yet simplified arrangement in conjunction with appropriate drives.

In-line with current trends, the pump type could be for example of either DuocentricTM or hypocycloidal type, although the powerpack design would be applicable to either type. In the present discussion a hypocycloidal type pump is adopted. As discussed in more detail below, the design of the hydraulic circuit in conjunction with the design of the powerpack and drive system extracts the maximum synergies.

The generation and supply of varying requirements for high pressure (HP) and low pressure (LP) flowrates causes a fundamentally inefficient compromise for mechanically driven systems which is overcome by the present invention. In particular a drive system capable of speed control that is at least partially independent of engine speed will improve matching between pump work and the hydraulic requirement, to enable the powerpack system to satisfy the hydraulic requirements of the transmision in the most fuel efficient way. To achieve this de-coupling of the following fundamental relationships is possible: Engine speed and pump speed (ie dispensing with direct mechanical drive, suggesting variable ratio mechnaical drive or an electrical drive), flow rate and energy usage by the pump (ie use of an hydraulic accumulator), drive torque to the pump and the load on the engine (ie use of a battery).

To implement the desired de-coupling between the operation of the pump and the speed of the engine whilst sourcing a variable drive torque from the engine, a new design of mechanical coupling is required. Various types and their merits are discussed below:

A preferred approach comprises use of clutches between multiple drives such as a pump shaft with an integrated electrical drive and a clutched connection to a shaft driven by the engine. This requires a simple clutch device able to withstand a maximum torque of 10 N. m.

A first preferred embodiment shown in Fig. 4 comprises a single pump (not shown) with mechanical drive 60 from the engine for all LP flow and electrical drive from 42V motor 62 for all HP flow. This features a"clutched" mechanical drive using simple friction clutch designated generally 64 normally closed to drive pump shaft 66 from the engine. The direct electrical drive 62 to allows the speed to be varied independently of the speed of the mechanical drive, by simple disengagment of the mechanical drive 60 by the clutch 64.

A variant of the arrangement in Fig. 4 is shown in Fig. 5. This uses a simple mechanical one-way clutch 68 configured so that the pump speed could be increased or decreased by the electrical drive 62 relative to the speed of the mechanical drive 60, but not both, by overfeeding/slipping the mechanical drive 60. The one way clutch allows effective disengagement by rotating one shaft faster with respect to another-the device is inexpensive and does not require an actuation system A further alternative is shown in Fig. 6. A single pump element 12 is shown with drive via a torque summing device such as an epicyclic designated generally 70 with inputs from the engine 60 and the electric motor 62.

As a result the following design criteria are taken into account: Functional flexibility, including the ability to control the speed of the pump rotor independent from the speed of the engine, to source simultaneously torque

from the engine crank and from an electrical motor, and to increase intermittently the flow to cool clutches under conditions of sustained slip.

Complexity Considerations-number of pump devices, requirement for an HP accumulator to satisfy transient requirements, number and complexity of any clutching systems, number and complexity of any gearing, complexity of speed control required for any electrical motors within system is reduced.

Alternative possibilities for the clutch arrangement comprise: electromagnetic clutch-as found in many automotive air conditioning drives; controlled ball ramp with pilot activation-applied in automotive driveline clutching systems (a small clamp load applied to the pilot clutch (electromagnetic) causes a drag torque which activates a ball ramp device to generate the clamp load of the main clutch); electro-rheological coupling-used in engine fan drives; conventional dog clutch with cone synchroniser-as found in current manual transmissions; or controlled roller clutch-by controlling the loading of elements within roller clutch devices it is possible to establish a torque path equivalent to dog clutch devices, but with substantial cost and packaging benefits although some devices of this type require a reversal of torque to effect disengagement. It will be appreciated that any other type of appropriate clutch may be used.

The operation of the hydraulic circuit in conjunction with the clutch arrangement will now be discussed. Three modes are discussed with reference to Figs. 5,, 6 and 7. In each case the arrangements are used in conjunction with the drive of the type shown in Fig. 3, the one way clutch arrangement, although any of the clutch arrangements of Figs 2 to 4 can of course be adopted.

Referring first of all to Fig. 5 the drive designated generally 80 is operated at zero slip such that the input speed of rotation at drive 60 equals the output

speed of rotation at pump shaft 66. Valve 18 is open in the low pressure circuit and one-way valve 32 is closed in the high pressure circuit. Accumulator 50 is partially uncharged. Accordingly the pump output is at low pressure and high flow-rate.

In the mode shown in Fig. 6 clutch slip is provided such that the rotational speed at the pump shaft 66 is greater than the rotational speed at the drive input 60. The valve 18 in the restriction and bypass element is closed such that flow passes through the high pressure circuit. Mechanical drive and electrical drive are combined in this case for the initial charging of accumulator 50. The pump output is thus low pressure and high flow-rate.

In the third mode shown in Fig. 7, once again clutch slippage is provided such that the electrical drive 62 increases the pump shaft rotation speed relative to the input rotational speed. However the valve 18 is open providing medium pressure pump output at a maximum flow-rate. As a result the electrical drive assists in driving the pump to cool the clutch for example during hill-hold while the bypass is open.

A further aspect addressed by the invention relates to torque sensing. Fig. 8 shows a simple wet clutch pack having an input shaft 90, output shaft 92 and static piston 94 and the potential sites for torque sensing. The torque sensor itself can be of any appropriate type, for example as available from ABB, Sweden. Site A and Site B use conventional shaft sensing. Sites C to F are 'located on the clutch hub 96. Site F and Site E require a technology suitable for "thin wall"tubular sensing. Site D and C require a technology suitable for "disc"sensing.

With regard to minimising any increase to axial length, sensing at sites C and D is preferred. Sensing on the face of the"disc", it would be plausible to apply two sensors to make use of the different stress levels at these two sites: sensor C at the inner radius is subject to higher stress levels and would therefore be suited to resolution of low torque levels associated with creep control, sensor D at the outer radius is subject to lower stress levels and would therefore be suited to resolution of full range torque.

As a result of torque sensing, accurate control of clutch capacity offers significant benefits for damping of torsional vibrations and offer overload protection of driveline components. The hydraulic system of Fig. 1 can further include a pump controller and a torque sensor of the type shown in Fig 8 providing feedback thereto, allowing improved control of the system.

It will be appreciated that the various aspects and components desribed herein can be combined or juxtaposed where appropriate. Although discussion is made specifically in relation to AMT transmissions, it applies equally to other transmissions where similar considerations apply.