U.S. patent application number 10/477407 was filed with the patent office on 2004-12-02 for powertrain control.
Invention is credited to Baker, Hannah, Ramsbottom, Mark, Wheals, Jonathan Charles.
Application Number | 20040237681 10/477407 |
Document ID | / |
Family ID | 26246062 |
Filed Date | 2004-12-02 |
United States Patent
Application |
20040237681 |
Kind Code |
A1 |
Wheals, Jonathan Charles ;
et al. |
December 2, 2004 |
Powertrain control
Abstract
A vehicle hydraulic system (10) includes a pump (12), a low
pressure line (16) and a high pressure line (30). A restriction and
bypass system is provided on the low pressure system allowing
charging up of an accumulator (50) on the high pressure system.
Furthermore a torque sensor, D, E, F is provided on a clutch hub
(96) of a multi-plate wet clutch.
Inventors: |
Wheals, Jonathan Charles;
(Leicestershire, GB) ; Ramsbottom, Mark;
(Lancashire, GB) ; Baker, Hannah; (Warwickshire,
GB) |
Correspondence
Address: |
OSHA & MAY L.L.P.
1221 MCKINNEY STREET
HOUSTON
TX
77010
US
|
Family ID: |
26246062 |
Appl. No.: |
10/477407 |
Filed: |
July 7, 2004 |
PCT Filed: |
May 13, 2002 |
PCT NO: |
PCT/GB02/02203 |
Current U.S.
Class: |
74/335 |
Current CPC
Class: |
B60W 30/18 20130101;
B60W 2510/0638 20130101; B60W 2050/0029 20130101; Y02T 10/40
20130101; Y10T 74/19251 20150115; F16H 61/0025 20130101; F04B 17/03
20130101; F16H 61/0437 20130101; F16H 2061/0084 20130101; B60W
2510/0657 20130101; F16H 2061/0034 20130101; F04B 17/05 20130101;
B60W 2050/0031 20130101; B60W 2540/22 20130101; B60W 2540/221
20200201; B60W 10/02 20130101; B60W 2520/105 20130101; F16H
2061/0071 20130101; B60W 10/11 20130101 |
Class at
Publication: |
074/335 |
International
Class: |
F16H 059/00; F16H
061/00 |
Foreign Application Data
Date |
Code |
Application Number |
May 11, 2001 |
GB |
0111582.3 |
Mar 4, 2002 |
GB |
0205006.0 |
Claims
1. A hydraulic fluid pump for a vehicle transmission hydraulic
system, the pump comprising a dual drive pump pressurizsation
source.
2. A pump as claimed in claim 1 in which the dual drive source is
an electromechanical source.
3. A pump as claimed in claim 2 in which the mechanical drive is
from the vehicle engine and the electrical drive is from an
electric motor.
4. A pump as claimed in claim 1 in which the dual drives drives a
common output shaft.
5. A pump as claimed in claim 4 including a clutch arrangement
coupling at least one of the drives to the output shaft.
6. A pump as claimed in claim 4 in which a mechanical drive is
coupled to the output shaft by one of a controlled clutch and a
one-way clutch.
7. A pump as claimed in claim 4 in which the dual drives drives the
common output shaft coupled through an epicyclic coupling.
8. A vehicle transmission system, including a pump and a high
pressure accumulator.
9. A system as claimed in claim 8 including a high pressure and low
pressure circuit drive commonly by the pump.
10. A system as claimed in claim 9 in which the accumulator is
provided on the high pressure circuit.
11. A system as claimed in claim 8 further including a bypass valve
on the low pressure circuit.
12. A system as claimed in claim 8 further comprising a pump
controller and a torque sensor providing feedback thereto.
13. A method of controlling a vehicle transmission hydraulic system
including a pump, a pump drive, a high pressure and low pressure
circuit, a bypass valve on the low pressure circuit and an
accumulator on the high pressure circuit including the steps of
driving the low pressure circuit in a first mode of operation,
overdriving the pump, closing the bypass valve to block the low
pressure circuit and charge the accumulator in a second mode of
operation and overdriving the pump and opening the bypass valve in
a third mode of operation.
14. A method as claimed in claim 13 in which the pump drive is a
dual drive arrangement and in which a first drive only drives the
pump in the first mode and both drives overdrive the pump in the
second and third modes.
15. A method of controlling a vehicle transmission hydraulic system
including a pump and a pump drive in which the pump drive is a dual
drive arrangement and in which a first drive only drives the pump
in a normal mode and both drives drive the pump in an overdrive
mode.
16. A clutch for a transmission comprising a clutch hub and a
torque sensor provided on the clutch hub.
17. A clutch as claimed in claim 16 in which the torque sensor is
provided on a radially extending portion of the clutch hub.
18. A clutch as claimed in claim 16 in which the torque sensor is
provided on an axially extending portion of the clutch hub.
19. A vehicle transmission system comprising a hydraulic pump
system as claimed in claim 1 and a clutch as claimed in claim 16.
Description
[0001] The invention relates to improved powertrain control for
example integrated powertrain control applied to automated manual
transmissions (AMT).
[0002] In Europe there is considerable growth in the application of
automated manual transmissions (AMT) predominantly for reasons of
cost and CO.sub.2 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.
[0003] 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.
[0004] 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.
[0005] 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.
[0006] 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 electromechanical 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.
[0007] 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.
[0008] 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 CO.sub.2 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.
[0009] Embodiments of the invention will now be described by way of
example with reference to the drawings of which:
[0010] FIG. 1 shows a hydraulic system according to the present
invention;
[0011] FIG. 2 shows a first drive system according to the present
invention;
[0012] FIG. 3 shows a second drive system according to the present
invention;
[0013] FIG. 4 shows a third drive system according to the present
invention;
[0014] FIG. 5 shows a drive system and a hydraulic scheme in a
first mode according to the present invention;
[0015] FIG. 6 shows the drive system of FIG. 5 in a second
mode;
[0016] FIG. 7 shows the drive system of FIG. 5 in a third mode;
and
[0017] FIG. 8 shows a torque sensor arrangement according to the
present invention.
[0018] 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 7 kW
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.
[0019] 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.
[0020] 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.
[0021] 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.
1TABLE 1 Estimates of hydraulic requirements 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 = 15 mm
actuators Travel = 10 mm Volume = 2.5 cc Total 2.1 litres/min Total
volume = disengage old gear + engage new gear = 5.0 cc Rapid
Shifting Maximum Clutch Assume maximum prolonged shift frequency
Instantaneous OD piston = 60 mm of 1 Hz 3.8 litres/minute ID piston
= 30 mm 10 Bar Travel = 0.9 mm Flow per shift Volume = 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 * 19 cc = 30 cc Total for Shifting : 2.1
litres/min Leakage Assumptions Valve Leakage : 1.0 litres/min Per
seal : 0.15 litres/ Clutch Leakage : 0.7 litres/min min Per valve :
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 sufficiant heat
rejection capacity to Cooling of Open clutch : 2.0 litres/min
reject 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
[0022] 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.
[0023] In-line with current trends, the pump type could be for
example of either Duocentric.TM. 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.
[0024] 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).
[0025] 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:
[0026] 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.
[0027] 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.
[0028] 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
[0029] 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.
[0030] 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.
[0031] 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.
[0032] 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.
[0033] 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.
[0034] 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.
[0035] 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.
[0036] 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.
[0037] 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.
[0038] 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.
[0039] It will be appreciated that the various aspects and
components described 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.
* * * * *