U.S. patent application number 09/761214 was filed with the patent office on 2001-07-05 for method for controlling internal combustion engine valve operating mechanism.
Invention is credited to Buehrle, Harry W. II, Clark, Raymond C., Gross, Jarrid, Long, Ron, Nist, Lance E..
Application Number | 20010006049 09/761214 |
Document ID | / |
Family ID | 22233251 |
Filed Date | 2001-07-05 |
United States Patent
Application |
20010006049 |
Kind Code |
A1 |
Buehrle, Harry W. II ; et
al. |
July 5, 2001 |
Method for controlling internal combustion engine valve operating
mechanism
Abstract
The reciprocating valve actuation and control system includes a
poppet valve moveable between a first and second position; a source
of pressurized hydraulic fluid; a hydraulic actuator including an
actuator piston coupled to the poppet valve and reciprocating
between a first and second position responsive to flow of the
pressurized hydraulic fluid to the hydraulic actuator; an
electrically operated valve controlling flow of the pressurized
hydraulic fluid to the actuator; and an engine computer that
generates electrical pulses to control the electrically operated
valve. The electrically operated valve preferably comprises a three
path rotary latched magnetic motor actuating a rotary valve portion
having a housing, a rotor, and a stator receiving and supplying
hydraulic fluid pressure to the rotor, which alternately directs
the hydraulic fluid pressure to the valve cylinder for opening of
the valve, or to return to the engine oil sump, for closing the
valve. In a presently preferred embodiment, the hydraulic actuator
comprises a self-contained cartridge assembly including an actuator
piston with dampers for damping motion of the actuator piston,
limiting the actuator stroke to assure soft seating of the
actuator, and to avoid overshoot during the engine valve opening
stroke and the engine valve return stroke. The electro-hydraulic
valves are electrically controlled by the engine computer, which
generates electrical signals carried to the electro-hydraulic
valves. The engine computer typically senses conventional engine
variables, and optimizes performance of the valve actuation and
control system according to preestablished guidelines, with
information being supplied to the engine computer by sensors. The
engine computer controls all aspects of engine performance,
interfaces with all of the peripheral sensors, and calculates fuel
parameters, ignition timing and engine valve timing based upon
prior mapping of the engine. In this manner the engine can be
controlled so as to provide maximum fuel economy, minimum
emissions, maximum engine torque, or a compromise between these
parameters.
Inventors: |
Buehrle, Harry W. II;
(Irvine, CA) ; Clark, Raymond C.; (Huntington
Beach, CA) ; Gross, Jarrid; (Fullerton, CA) ;
Long, Ron; (Garden Grove, CA) ; Nist, Lance E.;
(Santa Ana, CA) |
Correspondence
Address: |
FULWIDER PATTON LEE & UTECHT, LLP
HOWARD HUGHES CENTER
6060 CENTER DRIVE
TENTH FLOOR
LOS ANGELES
CA
90045
US
|
Family ID: |
22233251 |
Appl. No.: |
09/761214 |
Filed: |
January 16, 2001 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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09761214 |
Jan 16, 2001 |
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09480098 |
Jan 10, 2000 |
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6173684 |
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09480098 |
Jan 10, 2000 |
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09092445 |
Jun 5, 1998 |
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6024060 |
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Current U.S.
Class: |
123/90.12 ;
123/90.15 |
Current CPC
Class: |
F01L 2001/34426
20130101; F01L 13/06 20130101; F01L 9/10 20210101; F01L 2001/34446
20130101; F01L 2001/34423 20130101 |
Class at
Publication: |
123/90.12 ;
123/90.15 |
International
Class: |
F01L 009/02; F01L
001/34 |
Claims
What is claimed is:
1. A reciprocating valve actuation and control system for the
cylinders of an internal combustion engine, comprising: a poppet
valve moveable between a first and second position; a source of
pressurized hydraulic fluid; a hydraulic actuator including an
actuator piston coupled to the poppet valve and reciprocating
between a first and second position responsive to flow of the
pressurized hydraulic fluid to the hydraulic actuator; an
electrically operated valve controlling flow of the pressurized
hydraulic fluid to the actuator; and control means generating
electrical pulses to control the electrically operated valve.
2. The reciprocating valve actuation and control system of claim 1,
wherein the electrically operated valve controlling flow of the
pressurized hydraulic fluid to the actuator supplies pressurized
hydraulic fluid to the hydraulic actuator when electrically pulsed
to a first position, and dumps pressurized hydraulic fluid to a
system return when electrically pulsed to a second position.
3. The reciprocating valve actuation and control system of claim 2,
wherein the electrically operated valve comprises a three path
rotary latched magnetic motor.
4. The reciprocating valve actuation and control system of claim 1,
wherein the electrically operated valve comprises: a rotary valve
having a housing; a stator having an inlet pressure port receiving
pressurized hydraulic fluid, an inner bore in fluid communication
with the inlet pressure port through a plurality of radially
oriented apertures; a cylinder port groove in fluid communication
with the hydraulic actuator; a plurality of axial slots formed in
the stator allowing fluid communication between the cylinder port
groove and the inner bore of the stator; a generally cylindrically
shaped rotor disposed within the stator, the rotor having a
pressure supply groove at one end for receiving pressurized
hydraulic fluid from the inlet pressure port of the stator; a
plurality of axial pressure grooves in fluid communication with the
pressure supply groove of the rotor for supplying pressurized
hydraulic fluid to the actuator; and a plurality of return groove
formed in the rotor in fluid communication with a pressurized
hydraulic fluid return, for receiving hydraulic fluid from the
hydraulic actuator.
5. The reciprocating valve actuation and control system of claim 3,
wherein the three path rotary latched magnetic motor comprises: a
first pole piece connected to a first electromagnetic coil
energized by electrical pulses from said control means; a second
pole piece connected to a second electromagnetic coil energized by
electrical pulses from said control means, said first and second
pole pieces being connected at a magnetic junction; a magnetic
rotor disposed for rotation between a first position and a second
position contacting said first and second pole pieces,
respectively; a third pole piece disposed adjacent to the magnetic
rotor so as to define an air gap between the magnetic rotor and the
third pole piece; a permanent magnet connected to third pole piece;
a fourth pole piece connected between the permanent magnet and the
magnetic junction; and an output shaft mounted on the magnetic
rotor operatively connected to rotary valve means for controlling
flow of the pressurized hydraulic fluid to the hydraulic
actuator.
6. The reciprocating valve actuation and control system of claim 1,
wherein said hydraulic actuator comprises a self-contained
cartridge assembly including an actuator piston having means for
damping a stroke of the actuator piston to assure soft seating of
the actuator, and to avoid overshoot of the actuator piston.
7. The reciprocating valve actuation and control system of claim 6,
wherein said means for damping comprises first damping means to
avoid overshoot during an opening stroke of the engine valve.
8. The reciprocating valve actuation and control system of claim 7,
wherein said means for damping comprises second damping means to
decelerate the actuator piston to avoid high impact of the engine
valve into the valve seat.
9. The reciprocating valve actuation and control system of claim 6,
wherein said means for damping comprises a stepped land on the
actuator piston.
10. The reciprocating valve actuation and control system of claim
6, wherein said self-contained cartridge assembly further comprises
a main generally tubular sleeve having a bore, said bore having a
surface defining a damper cavity, said actuator piston having a
damper land member, and said damper cavity receiving said damper
land member during an actuating stroke of said actuator piston,
whereby hydraulic fluid is trapped in the damper cavity to damp
motion of the actuator piston during a stroke of the actuator
piston.
11. The reciprocating valve actuation and control system of claim
10, further comprising a secondary generally tubular sleeve having
a bore, said secondary sleeve bore having a surface defining a
secondary damper cavity, and said actuator piston having a surface
defining a damper orifice for fluid communication of said hydraulic
fluid from one of said main sleeve damping cavity and said
secondary sleeve damping cavity to the hydraulic fluid return.
12. The reciprocating valve actuation and control system of claim
10, when said self-contained cartridge assembly further comprises
an alignment tube within which said main sleeve is disposed, a
generally tubular damping spacer disposed within said alignment
tube adjacent to the main sleeve, a damping ring disposed within
said alignment tube adjacent to said damping spacer, and said
actuating piston having a surface defining a damping orifice for
fluid communication of hydraulic fluid from said damper cavity to
the hydraulic fluid return.
13. The reciprocating valve actuation and control system of claim
12, wherein said damper land member comprises a split ring, said
split ring having a surface defining a damper orifice through said
split ring for communicating hydraulic fluid to the hydraulic fluid
return.
14. The reciprocating valve actuation and control system of claim
12, wherein said damper land member comprises a laminar sealing
ring, said sealing ring having a surface defining an orifice in the
sealing ring for communication of hydraulic fluid to the hydraulic
fluid return.
15. The reciprocating valve actuation and control system of claim
1, wherein said source of pressurized hydraulic fluid comprises an
engine driven hydraulic positive displacement pump for supplying
said hydraulic fluid pressure, said hydraulic fluid is engine oil,
and an engine oil sump connected in fluid communication with said
pump for supplying engine oil to the pump, and said engine oil sump
being connected in fluid communication for receiving return engine
oil from the valve actuation and control system.
16. The reciprocating valve actuation and control system of claim
15, further comprising an unloader valve connected in fluid
communication with the pump for limiting output pressure of the
pump.
17. The reciprocating valve actuation and control system of claim
16, further comprising a check valve to prevent backflow from the
accumulator.
18. The reciprocating valve actuation and control system of claim
16, further comprising an accumulator connected in fluid
communication with the pump and the unloader valve for storing a
volume of the hydraulic fluid.
19. The reciprocating valve actuation and control system of claim
16, wherein said unloader valve comprises a pressure sensing valve
for sensing pump output pressure, said unloader valve opening when
the pump output pressure reaches a preset threshold value, said
unloader valve returning flow of said hydraulic fluid to
return.
20. The reciprocating valve actuation and control system of claim
1, wherein said control means comprises a computer and a plurality
of sensors disposed in the engine for sensing engine variables, and
optimizing performance of the reciprocating valve actuation and
control system.
21. The reciprocating valve actuation and control system of claim
1, the internal combustion engine having a cylinder head and a
combustion chamber, and wherein the engine cylinder head has a
bridge dividing the combustion chamber.
22. The reciprocating valve actuation and control system of claim
1, wherein said control means comprises a digital signal processor
to take advantage of its high speed real time signal processing
capability, whereby crankshaft dynamic related problems are
diagnosed, and dealt with in real time.
23. A method for controlling reciprocating valve actuation for the
cylinders of an internal combustion engine in a reciprocating valve
actuation and control system, the system including a poppet valve
moveable between a first and second position; a source of
pressurized hydraulic fluid; a hydraulic actuator including an
actuator piston coupled to the poppet valve and reciprocating
between a first and second position responsive to flow of the
pressurized hydraulic fluid to the hydraulic actuator; an
electrically operated valve controlling flow of the pressurized
hydraulic fluid to the actuator; and an engine control unit gene
rating electrical pulses to control the electrically operated
valve, wherein the source of pressurized hydraulic fluid comprises
an engine driven hydraulic positive displacement pump for supplying
the hydraulic fluid pressure, an unloader valve connected in fluid
communication with the pump for limiting output pressure of the
pump, and an accumulator connected in fluid communication with the
pump and the unloader valve for storing a volume of the hydraulic
fluid, the method comprising the step of: storing hydraulic energy
in the accumulator.
24. The method of claim 23, further comprising the step of:
controlling the accumulator in a way that commands the engine
driven pump to "run free" or be disconnected during brief power
bursts.
25. The method of claim 23, further comprising the step of:
controlling the accumulator in a way that forces the accumulator to
be charged during braking.
26. The method of claim 23, further comprising the step of:
controlling the accumulator in a way that forces the accumulator to
be charged during the time the vehicle needs to decelerate.
27. A method for controlling reciprocating valve actuation for the
cylinders of an internal combustion engine in a reciprocating valve
actuation and control system, the system including a poppet valve
moveable between a first and second position; a source of
pressurized hydraulic fluid; a hydraulic actuator including an
actuator piston coupled to the poppet valve and reciprocating
between a first and second position responsive to flow of the
pressurized hydraulic fluid to the hydraulic actuator; and an
electrically operated valve controlling flow of the pressurized
hydraulic fluid to the actuator; the method comprising the step of:
controlling the electrically operated valve with an engine control
unit generating electrical pulses.
28. The method of claim 27, wherein said step of controlling
comprises: controlling the engine control unit in a way that
commands a delay to take place in the opening of multiple intake or
exhaust valves in the cylinder.
29. The method of claim 27, further comprising the step of: the
engine control unit controlling the valve timing to create a swirl
effect in the combustion chamber.
30. The method of claim 27, further comprising the step of: mapping
the engine control unit in a manner that optimizes the swirl
effect.
31. The method of claim 27, wherein said step of controlling
comprises: the engine control unit controlling the valve timing of
the intake and exhaust valves of an engine having at least three
valves per cylinder, such that the intake and exhaust valves will
not open at the same time, and controlling the valve timing of the
intake and exhaust valves of the engine to provide a delay to off
load driver electronics and reduce peak current load, allowing
smaller current traces and preventing ringing of power
transistors.
32. The method of claim 27, the engine having a multi-inlet valve
cylinder having shaped and directed inlet ports, wherein said step
of controlling comprises: the engine control unit controlling the
valve timing to provide a delay of the opening of intake valves, to
cause a swirl effect to take place that is augmented by the shaped
and directed inlet ports.
33. The method of claim 27, the engine having a multi valve
cylinder having first and second exhaust valves, and first and
second hydraulic actuators, the second exhaust valve being larger
than the first exhaust valve, the first exhaust valve to open being
smaller in head diameter, resulting in lower actuation pressure,
wherein said step of controlling comprises: the engine control unit
controlling the timing of the valves to create a delay between the
opening point of exhaust valves in the multi valve cylinder to
reduce the demand placed on the second actuator, to lower
horsepower required to drive the larger exhaust second valve.
34. The method of claim 27, the engine having four intake and
exhaust valves, wherein said step of controlling comprises: the
engine control unit controlling the timing of the valves in the
following sequence: a. number 1 Intake valve opens (large valve) b.
number 4 Exhaust valve closes (after start up) c. number 2 Intake
valve opens (smaller valve) d. number 2 Intake valve closes e.
number 1 Intake valve closes f. compression and power stroke take
place g. number 4 Exhaust valve opens (smaller valve w/less surface
area) h. number 3 Exhaust valve opens (larger valve w/more volume)
i. number 3 Exhaust valve closes j. number 1 Intake valve opens
(overlap begins) k. number 4 Exhaust valve closes (overlap
ends).
35. The method of claim 27, the engine control unit commanding a
first set of exhaust valve opening and closing events, wherein said
step of controlling comprises: the engine control unit controlling
the timing of the valves by commanding a second set of exhaust
valve opening and closing events to take place.
36. The method of claim 27, the engine having four intake and
exhaust valves, wherein said step of controlling comprises: the
engine control unit controlling the timing of the valves in the
following sequence: a. number 1 Intake valve opens (largest valve)
b. number 2 Intake valve opens (smaller valve) c. number 2 Intake
valve closes d. number 1 Intake valve closes e. compression and
power stroke take place f. number 4 Exhaust valve opens (smaller
valve w/less surface area) g. number 3 Exhaust valve opens (larger
valve w/more volume) h. number 3 Exhaust valve closes i. number 1
Intake valve opens (overlap begins) j. number 4 Exhaust valve
closes (overlap ends).
37. The method of claim 27, wherein said step of controlling
comprises: the engine control unit controlling the valve timing by
opening and closing the valves several times during the same
stroke.
38. The method of claim 27, wherein said step of controlling
comprises: the engine control unit controlling the valve timing by
opening and closing the valves several times to control throttling
and braking.
39. The method of claim 27, wherein said step of controlling
comprises: the engine control unit controlling the valve
timing.
40. A method for controlling reciprocating valve actuation for the
cylinders of an internal combustion engine in a reciprocating valve
actuation and control system, the system including a poppet valve
moveable between a first and second position; a source of
pressurized hydraulic fluid; a hydraulic actuator including an
actuator piston coupled to the poppet valve and reciprocating
between a first and second position responsive to flow of the
pressurized hydraulic fluid to the hydraulic actuator; an
electrically operated valve controlling flow of the pressurized
hydraulic fluid to the actuator; a crankshaft; and a control means
controlling the electrically operated valve, the method comprising
the step of: determining the position and direction of rotation of
the crank shaft from the electrical outputs of a sin/cosine
crankshaft position sensor.
41. The method of claim 40, the method further comprising the step
of: operating valve openings and closings that are correct for
forward/reverse crankshaft rotation, based upon the crankshaft
position and direction information, to eliminating possible
mechanical interference for crankshaft reverse rotation.
42. The method of claim 40, further comprising the step of:
reversing the direction indication by electronically inverting one
signal to cause the engine to run backwards.
43. The method of claim 40, the engine being a four cycle engine
having a distributor and camshaft, the method further comprising
the step of: dividing the electrical position output of the
electrical crankshaft position sensor by two electronically to
eliminate costly mechanical components that drive the distributor
and camshaft at half speed, and determining the initial timing
during startup sequencing for valves, fuel and ignition.
44. The method of claim 40, the method further comprising the step
of: inputting preset default valve, fuel and ignition operating
values into registers of the control means upon application of
power, to be utilized if the control means fails to operate,
allowing operation in emergencies.
45. The method of claim 44, further comprising the step of: the
control means closing any open valves upon application of power to
eliminate mechanical interference until correct crankshaft location
is determined by a startup sequence.
46. The method of claim 40, further comprising the step of: the
control means inhibiting fuel injection but not ignition during a
shut-off command by the operator.
47. The method of claim 46, further comprising the step of:
sequentially commanding all intake and exhaust valves of the engine
to close before power termination to the control system, to
eliminate any possible mechanical interference after power removal,
in order to provide smooth termination, a low pollution
termination, or a rapid deceleration termination, depending on the
actual valve closure and ignition sequencing.
48. The method of claim 40, further comprising the step of:
comparing whether the electrical crankshaft position and direction
outputs of the position sensor are greater than but not equal to
desired values that open valves, to allow for commanded valve event
value changes asynchronously to crankshaft position without missed
events, such as a missed valve open event causing a misfire and
greater vibrations, noise and emitted pollution problem.
49. The method of claim 40, further comprising the step of:
comparing whether the electrical crankshaft position and direction
outputs of the position sensor are greater than but not equal to
desired values that open valves, to allow for commanded valve event
value changes asynchronously to crankshaft position without missed
events, such as a missed valve closure event causing a mechanical
interference problem.
50. The method of claim 40, further comprising the step of:
comparing whether the electrical crankshaft position and direction
outputs of the position sensor are greater than but not equal to
desired values that open valves, to allow for commanded valve event
value changes asynchronously to crankshaft position without missed
events, such as a missed fuel injection event causing a misfire and
greater mechanical vibrations and noise.
51. The method of claim 40, further comprising the step of:
comparing whether the electrical crankshaft position and direction
outputs of the position sensor are greater than but not equal to
desired values that open valves, to allow for commanded valve event
value changes asynchronously to crankshaft position without missed
events, such as a missed ignition event causing a misfire and
greater emitted pollution.
52. A method for controlling reciprocating valve actuation for the
cylinders of an internal combustion engine in a reciprocating valve
actuation and control system, the system including a poppet valve
moveable between a first and second position; a source of
pressurized hydraulic fluid; a hydraulic actuator including an
actuator piston coupled to the poppet valve and reciprocating
between a first and second position responsive to flow of the
pressurized hydraulic fluid to the hydraulic actuator; an
electrically operated valve controlling flow of the pressurized
hydraulic fluid to the actuator; and a control means controlling
the electrically operated valve, the method comprising the step of:
controlling the dynamic performance of the system by newly added
dimensions of mapping strategy, consisting of existing mapping of
sensory inputs such as gas pedal position, inlet and exhaust
manifold and barometric pressures, exhaust gas composition, coolant
and ambient air temperatures with ignition timing and fuel
injection, along with new dimensions of inlet/exhaust valve
timing.
53. The method of claim 52, the engine having individual fuel
injectors, the method further comprising the steps of: controlling
the valve actuation solenoids; and controlling the individual fuel
injectors, as well as individual spark events on each cylinder used
on the system, based upon said sensory input, and based upon
multidimensional mapping.
54. The method of claim 52, further comprising the step of:
determining if any cylinders are operating unsatisfactorily, based
upon use rotational rate measurements.
55. The method of claim 54, further comprising the step of: the
control means disabling defective cylinders entirely, reducing
pollution and potential further engine damage, while offering a
limited "limp home" operation.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] This invention relates generally to a valve actuating
apparatus for engines, and more particularly concerns a system for
actuating and controlling reciprocating valves for the cylinders of
an internal combustion engine.
[0003] 2. Description of Related Art
[0004] Conventional piston type internal combustion engines
typically utilize mechanically driven camshafts for operation of
intake and exhaust valves, with fixed valve lift and return timing
and duration. Electrically or hydraulically controlled valves for
improved control of valve operation have also been used in order to
improve fuel economy and reduce exhaust emissions.
[0005] For example, a variable engine valve control system is known
in which each of the reciprocating intake or exhaust valves is
hydraulically controlled, and includes a piston receiving fluid
pressure acting on surfaces at both ends of the piston. One end of
the piston is connected to a source of high pressure hydraulic
fluid, while the other end of the piston can be connected to a
source of high pressure hydraulic fluid or a source of low pressure
hydraulic fluid, under the control of a rotary hydraulic
distributor coupled with solenoid valves.
[0006] Another engine valve actuating system is known in which each
cylinder is provided with a coaxial venturi shaped duct having
inwardly facing vanes that hold an electro-mechanical valve
actuator. When the electro-mechanical valve actuator receives a
pulsed electrical signal, the actuator operates to reciprocate the
valve.
[0007] While a camshaft driven intake or exhaust valve will
typically open and close with a constant period as measured in
crankshaft degrees, for any given engine load or rpm, there is a
need for an indirect valve actuation system for internal combustion
engines that can operate more rapidly, and that will open the valve
at the same rate regardless of engine operating conditions.
Ideally, a valve actuation system should match the optimum, maximum
valve rate of operation at maximum speed of operation of an engine
to provide a rapid, optimum valve operation rate. It would also be
desirable to provide a valve actuation system for internal
combustion engines offering a speed of operation that will allow
greater flexibility in programming valve events, resulting in
improved low speed torque, lower emissions, and better fuel
economy. Conventional approaches to providing higher rates of valve
opening and closing have used non-latching control valves commonly
involving systems using either spool valves or poppet valves,
neither of which provide for a high flow open area in a small, low
inertia system or energy efficient latching mechanisms. It would be
desirable to provide a valve actuation and control system with an
electro-hydraulic valve system, having a high flow open area, low
inertia of operation, a small size, and ease of manufacture. The
present invention meets these needs.
SUMMARY OF THE INVENTION
[0008] Briefly, and in general terms, the present invention
provides for an intake/exhaust (I/E) reciprocating valve actuation
and control system for the cylinders of an internal combustion
engine, comprising I/E poppet valves moveable between a first and
second position; a source of pressurized hydraulic fluid; a
hydraulic actuator including an actuator piston coupled to the
poppet valve and reciprocating between a first and second position
responsive to flow of the pressurized hydraulic fluid to the
hydraulic actuator; an electrically operated hydraulic valve
controlling flow of the pressurized hydraulic fluid to the
hydraulic actuator; and electronic control means generating
electrical pulses to control the electrically operated valve.
[0009] In one presently preferred embodiment, the invention
provides for a three way electrically operated valve controlling
flow of the pressurized hydraulic fluid to the actuator, supplying
pressure when electrically pulsed to open, magnetically latching,
and dumping actuator oil to an engine oil sump when the valve is
electrically pulsed to close. The electrically operated valve
preferably comprises a three path rotary latched magnetic motor
actuating a rotary valve portion having a housing, a rotor, and a
stator receiving and supplying hydraulic fluid pressure to the
rotor, which alternately directs the hydraulic fluid pressure to
the valve cylinder for opening of the valve, or to return to the
engine oil sump, for closing the valve.
[0010] In a presently preferred embodiment, the hydraulic actuator
comprises a self-contained cartridge assembly including an actuator
piston with means for damping motion of the actuator piston,
limiting the actuator stroke to assure soft seating of the I/E
valve, and to avoid overshoot during the engine valve opening
stroke and the engine valve return stroke. In a currently preferred
embodiment, the source of pressurized hydraulic fluid comprises an
engine-driven pump supplying engine oil under pressure as the
hydraulic fluid, an accumulator is used to provide a reservoir of
high pressure fluid, and an engine oil sump for receiving return
hydraulic fluid. An unloader valve limiting pump output pressure is
also provided, along with a check valve preventing backflow from
the engine oil sump. An accumulator is also preferably provided for
storing a sufficient volume of pressurized hydraulic fluid to
moderate the pump and unloader valve duty cycle. The unloader valve
preferably comprises a pressure sensing valve that senses pump
output pressure and opens when the pressure reaches a preset value,
so that when the unloader valve is open, flow from the pump returns
to the engine oil sump. The accumulator is also used to store
energy primarily dissipated under deceleration by the brakes or as
a compression brake by filling the accumulator during that time.
The engine would use the torque from the wheels in reverse driving
the hydraulic pump and filling the accumulator, thus recycling
velocity energy that would normally be lost to wheel braking.
[0011] Thus, the hydraulic pump could be temporarily disconnected
so that under high load, the valve train would run off stored
accumulator energy. This would use more of the power lost during
braking. In a presently preferred embodiment, the control means
comprises a computer, and sensors are operatively connected to the
computer, for monitoring engine variables, and for optimizing
performance of the system.
[0012] These and other aspects and advantages of the invention will
become apparent from the following detailed description and the
accompanying drawings, which illustrate by way of example the
features of the invention.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] FIG. 1 is a schematic diagram of the internal combustion
engine reciprocating valve actuation and control system of the
invention;
[0014] FIG. 2 is a sectional view of a first embodiment of a
hydraulic actuator of the reciprocating valve actuation and control
system of FIG. 1;
[0015] FIG. 3 is a sectional view of a second embodiment of a
hydraulic actuator of the reciprocating valve actuation and control
system of FIG. 1;
[0016] FIG. 4 is a sectional view of a damping spacer of the
hydraulic actuator of FIG. 3;
[0017] FIG. 5A is a sectional view of a third embodiment of a
hydraulic actuator of the reciprocating valve actuation and control
system of FIG. 1;
[0018] FIG. 5B is a plan view of the split ring of the hydraulic
actuator of FIG. 5A;
[0019] FIG. 6 is a sectional view of a fourth embodiment of a
hydraulic actuator of the reciprocating valve actuation and control
system of FIG. 1;
[0020] FIG. 7A is a sectional view of a fifth embodiment of a
hydraulic actuator of the reciprocating valve actuation and control
system of FIG. 1;
[0021] FIG. 7B is a plan view of the laminar sealing ring of the
hydraulic actuator of FIG. 7A;
[0022] FIG. 7C is a side elevational view of the laminar sealing
ring of FIG. 7B;
[0023] FIG. 8 is a sectional view of the electrically operated
valve controlling flow of the pressurized hydraulic fluid to the
actuator of the reciprocating valve actuation and control system of
FIG. 1;
[0024] FIG. 9 is a cross-sectional view of the electrically
operated valve motor taken along line 9-9 of FIG. 8;
[0025] FIG. 10 is a plan view of the rotor of the rotary valve of
the electrically operated valve of FIG. 8;
[0026] FIG. 11 is a sectional view of the rotor taken along line
11-11 of FIG. 10;
[0027] FIG. 12 is a sectional view of the stator of the rotary
valve of the electrically operated valve of FIG. 8;
[0028] FIG. 13 is a cross-section of the stator taken along line
13-13 of FIG. 12;
[0029] FIG. 14 is a sectional view of the rotary valve assembly of
the electrically operated valve of FIG. 8;
[0030] FIG. 15 is a cross-sectional view of the rotary valve
assembly taken along line 15-15 of FIG. 14;
[0031] FIG. 16 is a perspective view of the rotary latched magnetic
motor of the electrically operated valve of FIG. 8;
[0032] FIG. 17 is a schematic front view of the rotary latched
magnetic motor of FIG. 16, illustrating operation of the motor;
[0033] FIG. 18 is a graph comparing operating speeds of valves
driven by a mechanical camshaft and valves driven by the
reciprocating valve actuation and control system of the invention;
and
[0034] FIG. 19 is a schematic diagram of paired intake and exhaust
valves of unequal sizes.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0035] While mechanical camshafts for the intake and exhaust valves
of internal combustion engines typically have a period of opening
and closing that remains constant in terms of crankshaft degrees
for any engine load or rpm; this has limited the ability of the
automotive industry to improve fuel economy, reduce harmful exhaust
emissions, and to improve low end torque. Typical approaches to
providing variable valve opening and closing positions have
involved either variable mechanical linkages or phasing by motors
connecting the camshaft to the cam drive. These methods do not
provide a high flow open area in a small low inertia system.
[0036] The present invention accordingly provides for an improved
reciprocating valve actuation and control system for the cylinders
of an internal combustion engine. As is illustrated in the
drawings, and as is generally shown in FIG. 1, the reciprocating
valve actuation and control system of the invention is a camless
valve control system 20 for an engine poppet valve 22 moveable
between a first, open position, and a second, closed position in
which the engine poppet valves are reseated by common valve springs
24. The engine poppet valves are driven by hydraulic actuators 26,
which are controlled by electrically operated electro-hydraulic
valves 28 supplying hydraulic fluid to the actuators via conduit
29. The hydraulic fluid is preferably engine oil, supplied to the
electro-hydraulic valves by the pressure rail 30. An engine driven
hydraulic pump 32 supplies the oil pressure that is used to open
the engine, poppet valves, receiving the oil from an engine oil
sump 34. In a presently preferred embodiment, the electro-hydraulic
valves are three way type hydraulic valves, supplying pressure when
electrically pulsed to open, magnetically latching, and dumping the
actuator oil to the sump when pulsed to close. Each engine I/E
valve is preferably provided with an actuator and an
electro-hydraulic valve.
[0037] In a presently preferred embodiment, the engine driven pump
32 is a hydraulic pump driven directly by the engine, so that the
output of the pump will increase in direct proportion to the engine
speed. The positive displacement pump is preferably sized to
provide about 110% of the oil flow required by the engine system of
valves. The engine oil return from the electro-hydraulic valve and
piston actuator assembly is to the engine oil sump, typically by
gravity through the normal engine drainage passage (not shown). The
positive displacement pump output pressure is also preferably
limited by an unloader valve 36, as moderated by an accumulator 38
connected to the oil pressure rail. The nature of the actuator and
the valve utilizing the normal engine oil supply allows the engine
oil supply to be used as a hydraulic fluid even if the engine oil
supply contains some entrained air, drastically simplifying the
system and accessories that would otherwise be required to
condition the hydraulic fluid, and obviating the need for a
separate hydraulic fluid supply.
[0038] The unloader valve 36 preferably comprises a pressure
sensing valve that senses pump output pressure and opens when the
pump output pressure reaches a preset threshold value. When the
unloader valve is opened, all of the flow from the positive
displacement pump is to return to the engine oil sump, so that the
output from the pump is then "unloaded". A check valve 40 is also
preferably provided in the fluid line between the accumulator and
the unloader valve to prevent backflow from the accumulator.
[0039] The accumulator in the system is provided to receive oil
from the pump, accepting a volume of engine oil from the pump as an
accumulator piston 42 moves within in the accumulator to create the
interior accumulator volume. A means for biasing the piston to
maintain pressure on the piston is also provided, preferably in the
form of a coil spring 44, although other means of biasing the
piston to provide system oil pressure could also be used, such as a
pneumatic pressure chamber, for example. When the unloader valve
senses that pump output pressure has reached the preset threshold
value, opening to allow flow from the pump to return to the engine
oil sump, the hydraulic fluid flow and pressure are supplied to the
system from the accumulator. When this supply is exhausted, the
system pressure drops, the unloader valve senses the system
pressure drop below a lower, preset minimum oil pressure threshold,
and closes, allowing the pump to reload the accumulator volume. The
cycling rate of this action depends on the settings of the minimum
and maximum oil pressure thresholds of the unloader valve. The
unloader valve settings can be relatively close together, so that
the system cycles rapidly, or can be set relatively far apart, so
that the cycle rate is slower, and resulting in a greater variation
of hydraulic fluid supply pressure, as desired. Unloader valve
settings can be controlled by the engine control unit (ECU), or
engine computer 50.
[0040] The electro-hydraulic valves are preferably electrically
controlled by the engine computer 50 (ECU), which generates
electrical signals carried to the electro-hydraulic valves via
electrical connectors 52a-d. The engine computer typically senses
conventional engine variables, and optimizes performance of the
valve actuation and control system according to preestablished
guidelines, with information being supplied to the engine computer
by sensors 54a-c. The valve actuation and control system typically
includes a manifold pressure sensor, a manifold temperature sensor,
a mass flow sensor, a coolant temperature sensor, a throttle
position sensor, an exhaust gas sensor, a high resolution engine
position encoder, a valve/ignition timing decoder controller,
injection driver electronics, valve coil driver electronics,
ignition coil driver electronics, air idle speed control driver
electronics, power down control electronics, and a standard
communications port. In addition to controlling the engine valves
through the hydraulic actuation system, the engine computer also
typically sequences engine ignition, fuel injection and OBD
(onboard diagnostics).
[0041] The engine computer preferably utilizes a high performance
digital signal processor (DSP), so that control of all aspects of
the engines performance can be attained. The DSP interfaces with
all of the peripheral sensors, and calculates fuel parameters,
ignition timing and engine valve timing based upon prior mapping of
the engine. Mapping is performed multi-dimensionally using engine
speed, manifold pressure, induction mass flow and temperatures. In
this manner the engine can be controlled so as to provide maximum
fuel economy, minimum emissions, maximum engine torque, or a
compromise between these parameters.
[0042] An alternate mapping method to simplify system complexity
and reduce parts count would be induction mass flow, temperatures,
barometric pressure, engine speed and pedal position sensors.
[0043] The engine computer will determine if the current operating
conditions are within or not within the normal driving cycle of the
engine, and will adjust the operation as is required. Configuration
software is utilized that allows the reciprocating valve actuation
and control system to be tailored for an exact engine system.
Engines can be mapped on any engine dynamometer, and evaluated
across engine speed and load, so that independent maps can be
developed for fuel economy, emissions or torque. Maps are stored
for ignition, fuel control and valve control and can be used
separately or in combination.
[0044] The crankshaft position sensor is used to provide the engine
control unit with a method of controlling engine valve/fuel
injection/ignition events. The engine crank position sensor must be
reliable, accurate, low cost and have a long life. The accuracy and
repeatability should ideally be better than or equal to that of a
conventional mechanical camshaft, and with a simple electrical
interface to the engine control unit. Analog and digital rotational
position sensors can meet these requirements.
[0045] Most analog position sensors can be eliminated if they have
any contacting parts that wear out. Resolvers and sin/cosine (hall
effect) potentiometers have output signals that must be phase
decoded, digitized, and then require a table lookup to generate a
digital angle output. These analog sensors usually suffer from long
term drift or linearity/drive problems. A digital sensor eliminates
these problems, and is available at low cost. Two types of position
encoders are in wide use today: magnetic (hall effect), and optic
(photoelectric).
[0046] Both of these position encoder types are generally available
as absolute position encoders. In addition, an automotive sensor
should also be inexpensive and readily mounted to an engine
crankshaft. A typical engine crankshaft has up to .+-.0.003 inch of
axial end play, but good axial rotational concentricity. Absolute
position encoders need to have precision end play and axial
alignment and need to be mounted in a vibration and shock free
environment to give accurate readouts.
[0047] A 360 count, sin/cosine optical encoder can meet all of the
above requirements, because recent optical encoder array sensor
developments allow the encoder to be mounted on the crankshaft and
function well in an automotive environment. A magnetic encoder can
also be used, but this presently requires a larger space, and
presents somewhat greater difficulty to initially index the sensor
on the crankshaft for proper synchronization of the engine in an
automotive environment.
[0048] For either magnetic or optic encoders, the sin/cosine &
index pulses must be converted into a shaft angle output to control
valves, fuel injection, and ignition. It is also desirable for the
position sensor to be able to operate in 2, 3, 4, 5, 6, 8, 10, 12
or 16 cylinder engines; therefore the sensor output counts must be
divisible by 2, 3, or 5 to give the same timing to all cylinders
(without odd offsets which cause vibration and uneven operation).
This requirement eliminates a 256 or 512 count/rev encoder and
their simple base 2 arithmetic. With a 360 count encoder, a
resolution of 1/4degree and accuracy of about 1/3degree is obtained
from the quadrature output decoding of the sin/cosine signals (and
the count is divisible by 2, 3, or 5).
[0049] The engine computer must make valve timing/fuel injection
and ignition timing computations (or lookup tables) that ensure
engine horsepower/RPM/torque requirements and clean combustion for
the engine. Since the engine computer is busy checking many other
sensors that ensure clean combustion and efficient operation, it is
desirable to "unload" the engine computer by controlling valve
timing, fuel injection, and ignition timing with fixed hardware
circuits. This unloading also will allow a smaller and lower cost
microprocessor to be used in the engine control unit.
[0050] It is desirable to allow the engine computer to give valve
timing and ignition or fuel injection updates to the valve control
circuits at any time during the engine rotation without risk of
damage to valve or piston position. This becomes more apparent in 8
to 12 cylinder engines, since more events occur during the same
engine revolution and at different times than in 4 or 6 cylinder
engines. An update to any engine parameter is effective during the
current and subsequent control events until the next update occurs.
Thus, the engine computer will not delay updates until a "safe"
point in the cycle is reached to update timing events. Especially
if a cylinder misfires, it is necessary to change something
immediately if gross pollution is to be avoided, and the engine
computer may shut that cylinder off if necessary.
[0051] Engine starting and stopping are a problem using a
sin/cosine encoder. During start (power application), the engine
sensor does not determine its absolute position until the first
index pulse is received. Further, at engine shutoff, power will be
removed that prevents further valve control, so all valves must be
quickly closed (for further uncontrolled engine rotations). These
shutdowns can be easily handled by the sensor and/or the engine
control unit. During a controlled shutdown (ignition switch turned
off), valves and engine ignition can be fully controlled until zero
rotation by the engine computer, sequentially shutting off fuel,
then closing intake valves, then closing exhaust valves, then
turning off power to itself and engine position sensor. This can be
handled with minimum pollution, if desired, or any other
requirement.
[0052] In case of other, sudden, unexpected power failures, the
engine computer will shut valves (uncontrolled) with a power fault
detect circuit and local power hold up capacitor. This will prevent
engine damage, and contain most pollutants within the engine.
[0053] During power application (and engine cranking), the engine
position sensor immediately loads default starting values for all
valve/ignition/fuel injection settings. When engine cranking
begins, the engine position sensor will command all valves to close
(in case any are open). The engine position sensor will not command
and output events until the first sine/cosine index pulse is
received (so absolute crank position is known). The vehicle driver
may have to crank the engine up to one full revolution before this
occurs (with all valves closed), but this will assure adequate
hydraulic pressure for a good clean start. The engine computer may
update default engine starting values at any time after power
application.
[0054] The engine position sensor must also be able to handle
reverse engine rotation (safely) if the engine accidently rotates
backwards, (if parked on a hill or during a misfire at startup).
These conditions occur only occasionally, but in all cases, valves
must be closed when the piston is at or near top dead center (TDC)
to prevent engine damage. This is performed as a result of standard
quadrature decoding.
[0055] The valve actuation and fuel control system software is a
fully interrupt driven control system that is centered around a DSP
processor as a real time engine controller. The valve actuation and
interrupt system software is written in the DSP processor's native
instruction set for speed and efficiency. The other engine sensors
operate independently from the processor, and their routines can be
written in a higher language such as BASIC or C.sup.++, for
example.
[0056] The valve actuation and fuel control system can operate both
synchronously as well as asynchronously with respect to engine
rotation intervals. The major operating tasks such as data
acquisition and digital filtration are performed asynchronously in
constant time intervals, but the calculation of some engine
parameters, particularly fuel injection and valve angles, are
calculated during degree based intervals.
[0057] The valve actuation and fuel control system contains a
real-time monitor that allows another software package to query the
valve actuation and control system for "while running" information.
This feature allows dynamic data updates to be done by another host
computer system for emissions, diagnostic and custom tuning
work.
[0058] The valve actuation and fuel control system interfaces to
the engine position decoder via an 8 or 16 bit word. This interface
sets individual registers within the decoder, that define starting
and stopping points for events in degrees. The degree based events
controlled by the valve actuation and engine control system is
ignition dwell, engine valve open position and engine valve closed
position of all intake and exhaust valves as well as the start of
the fuel injection event. In addition, the start of the fuel
injection event is timed such that the end of injection event will
occur approximately simultaneous with the spark instant. Because
the engine ignition is degree based, the degrees that the ignition
coil are held powered is its dwell, and can be held either at a
constant dwell or at a constant coil energy. The latter is the most
desirable for lower power consumption and cooler ignition coil
operation.
[0059] The propagation delay of the engine valves must be taken
into account for top performance. This can be accomplished as part
of valve/ignition/fuel injection mapping, but as the system ages,
and some valve velocity may be lost, the valve actuation and
control system can measure its own average valve velocity and
recommend a tuneup.
[0060] The valve actuation and fuel control system controls the
fuel by setting the individual injector time periods proportional
to the amount of fuel calculated by the engine computer. The start
of each injector pulse can be set at any crank angle and can run
for times up to 720 crank degrees. The valve actuation and fuel
control system can run the injectors in true sequential or phased
sequential patterns for better atomization. This system could also
operate a direct injected gasoline engine.
[0061] With reference to FIGS. 2-7C, the hydraulic valve actuators
of the reciprocating valve actuation and control system are
preferably provided as self-contained cartridge assemblies. The
hydraulic actuators preferably include an actuator piston 60
coupled to the poppet valve, and reciprocating between a first,
open position and a second, closed position, in response to flow of
the pressurized hydraulic fluid to the hydraulic actuator. The
actuator pistons are preferably sized to efficiently move the
engine valves against their return spring forces. This sizing is
typically determined by a computer design program that takes into
account all of the necessary mechanical and hydraulic variables. An
ideal piston size is generally one that distributes half of the
pressure drop to the electro-hydraulic valve, and the other half of
the pressure drop to the piston area for actuation. As will be
explained further below, the actuator strokes are preferably
terminated with hydraulic dampers to assure soft seating of the
engine valves.
[0062] As is illustrated in FIG. 2, in one preferred embodiment of
the hydraulic actuator of the reciprocating valve actuation and
control system of the invention, the actuator piston 60 is mounted
to the engine 62 by bolts 64. The hydraulic actuator assemblies
include a main sleeve 66 and a secondary sleeve 68, and the
actuator piston is disposed within the bore 70 of the main sleeve
and the bore 72 of the secondary sleeve. Each of the main and
secondary sleeves have precision lapped bores that mate with the
outside diameter 74 of the actuating piston. In addition, each
sleeve contains secondary bores 76 that fit closely with a damper
land 78 of the actuator piston. The bores and the piston diameters
are all concentric, typically with very close tolerances on the
order of plus or minus 0.00005 inch (0.00125 mm). The hydraulic
actuator piston preferably includes a hydraulic damper system for
limiting the actuator piston stroke to assure soft seating of the
actuator piston, and to avoid overshoot during the engine valve
opening stroke and the return stroke. The secondary bore 76 of the
main sleeve therefore defines a damping cavity 80, and the actuator
piston includes a damping orifice 82 to decelerate the moving parts
to avoid overshoot during the engine valve opening stroke. The
secondary bore also preferably defines a damping cavity 84, and the
actuator piston includes a damping orifice 86 to decelerate the
system to avoid high impact of the engine valve into the valve seat
on the return stroke. The stepped land 78 enters these secondary
diameters in the damping cavities at the ends of the opening and
closing strokes, and the oil trapped in the respective cavities
exits through the respective orifices, thus creating a controlled
high back pressure, slowing down the motion of the piston and
bringing the moving parts of the valve to a soft landing.
Conventional engine valve return springs are used as a return
device, so that energy stored in the spring drives the closing
stroke, and so that energy for the closing stroke does not need to
be supplied by the pumping system.
[0063] As is illustrated in FIGS. 3 and 4, in a second embodiment,
the actuator piston 90 is mounted in the engine 92 within an
alignment tube 94, sealed within the engine by the o-ring 95. The
actuator piston cartridge assembly includes a main sleeve 96
disposed within the alignment tube and having a bore 100 mated to
the outside diameter 104 of the actuator piston. The secondary
sleeve of the piston assembly of FIG. 2 is replaced in this
embodiment by the damping ring 106 disposed within the alignment
tube, and a damping spacer 108. The damping spacer is preferably
drilled to provide a gap 110, and is disposed within the alignment
tube between the main sleeve and the damping ring. The actuator
piston assembly is preferably contained either as a shrink fit or a
pressed fit in the alignment tube. The inside diameter of the main
sleeve can easily be formed to be matched to the outer diameter of
the actuating piston, while the outside diameter of the actuating
piston can be sized while on a mandrel that is concentric to the
inner bore of the sleeve. These considerations allow the
manufacturing cost of the actuator piston and the main sleeve to be
relatively inexpensive. Similarly, the damping ring 106 is
preferably configured as a bushing, and can easily be manufactured
to close tolerances and perfect concentricity. The damping spacer
is also preferably manufactured as a bushing, and the gap provided
by 110 provides limits for the undamped portion of the stroke of
the actuating piston. The orifices 120 provide the damping. The
inside diameter of the damping spacer must fit closely to the
damping land 112 on the actuator piston, and the outside diameter
is preferably concentric and sized as an interference fit with the
alignment tube. However, concentricity and sizing for these close
tolerance fits are easily obtained at low manufacturing costs with
modern machining. The alignment tube is preferably manufactured
from precision tubing, and is preferably made from a seamless tube
that is either honed or roller swaged to size to fit the
surrounding bushing parts. The main sleeve, the damping spacer, the
damping rings and the actuating piston are preferably preassembled,
and are preferably either press fit or shrink fit into the
alignment tube. Once in place and checked for free action, the ends
of the alignment tube are typically roller swaged or electron beam
spot welded to permanently lock the parts in place. The resulting
assembly can then be handled as a cartridge, and mounted in the
engine with a sealing plug 115, o-ring 114, and a snap ring 116. A
damping cavity 118 is provided between the outside diameter of the
actuator piston and the inside diameter of the damping spacer 108,
and damping orifices 120 are provided on either side of the damping
land 112 of the actuator piston.
[0064] Referring to FIGS. 5A, 5B, and 6, in another embodiment, the
actuator piston 90' has been modified to replace the stepped
actuating piston land shown in FIG. 3, in order to reduce
manufacturing costs of the actuating piston, by allowing the
actuator piston to be manufactured as a cylindrical ground or
lapped part. The actuator piston 90' is mounted in the engine 92'
within an alignment tube 94', sealed within the engine by the
o-ring 95'. The actuator piston cartridge assembly includes a main
sleeve 96' disposed within the alignment tube and having a bore
100' mated to the outside diameter 104' of the actuator piston. The
damping ring 106' is disposed within the alignment tube, and a
damping spacer 108' that is preferably drilled to provide a gap
110' is disposed within the alignment tube between the main sleeve
and the damping ring. The actuator piston assembly is preferably
contained either as a shrink fit or a pressed fit in the alignment
tube. The inside diameter of the damping spacer must fit closely to
the damping land 112' on the actuator piston, and the outside
diameter is preferably concentric and sized as an interference fit
with the alignment tube. The alignment tube is preferably
manufactured from precision tubing, and is preferably made from a
seamless tube that is either honed or roller swaged to size to fit
the surrounding bushing parts. The main sleeve, the damping spacer,
the damping rings and the actuating piston are preferably
preassembled, and are preferably either press fit or shrink fit
into the alignment tube. Once in place and checked for free action,
the ends of the alignment tube are typically roller swaged or
electron beam spot welded to permanently lock the parts in place.
The resulting assembly can then be handled as a cartridge, and
mounted in the engine with a sealing plug 115', o-ring 114', and a
snap ring 116'. A damping cavity 118' is provided between the
outside diameter of the actuator piston and the inside diameter of
the damping spacer 108', and a damping orifice 120' is provided
through the side of the damping land 122' of the actuator
piston.
[0065] As is shown in FIGS. 5A and 6, the stepped land of the
actuator piston can be replaced by a hardened split ring 122', and
the actuating piston can be machined with a groove to accept this
ring. Since the outside diameter of the actuating piston is a
straight cylinder, the actuator piston can be centerless ground,
roller lapped, or otherwise machined as a straight rod. The
hardened split ring is a low cost part that has a closely sized
outside diameter to fit closely to the damping spacer 108'. The
inside diameter of the ring is not critical, and can be fit with a
high clearance to the actuating piston groove. The hardened ring is
typically machined, notched, heat treated, finished to size, and
then is slipped onto a tapered mandrel and split at the notches.
The two parts are kept as a pair and assembled to the actuating
piston during assembly with the alignment tube. One or more damping
orifices 120', such as a multiplicity of holes, slots, flats, and
the like, are preferably formed in the ring, although only a single
orifice is shown in FIG. 5B.
[0066] As is illustrated in FIGS. 7A, 7B, and 7C, in another
embodiment, the actuator piston 90" is assembled in the actuator
piston cartridge assembly with an alternative type of replacement
of the damping land of the actuator piston of FIGS. 2 and 3. The
actuator piston 90" is mounted in the engine 92" within an
alignment tube 94", sealed within the engine by the o-ring 95". The
actuator piston cartridge assembly includes a main sleeve 96"
disposed within the alignment tube and having a bore 100" mated to
the outside diameter 104" of the actuator piston. The damping ring
106" is disposed within the alignment tube, and a damping spacer
108" that is preferably drilled to provide an orifice 110" is
disposed within the alignment tube between the main sleeve and the
damping ring. The actuator piston assembly is preferably contained
either as a shrink fit or a press fit in the alignment tube. The
inside diameter of the damping spacer must fit closely to the
damping land 112" on the actuator piston, and the outside diameter
is preferably concentric and sized as an interference fit with the
alignment tube. The alignment tube is preferably manufactured from
precision tubing, and is preferably made from a seamless tube that
is either honed or roller swaged to size to fit the surrounding
bushing parts. The main sleeve, the damping spacer, the damping
rings and the actuating piston are preferably preassembled, and are
preferably either press fit or shrink fit into the alignment tube.
Once in place and checked for free action, the ends of the
alignment tube are typically roller swaged or electron beam spot
welded to permanently lock the parts in place. The resulting
assembly can then be handled as a cartridge, and mounted in the
engine with a sealing plug 115", o-ring 114", and a snap ring 116".
A damping cavity 118' is provided between the outside diameter of
the actuator piston and the inside diameter of the damping spacer
108", and damping orifices 120" are provided on either side of the
damping land 112" of the actuator piston.
[0067] In this embodiment, the actuator piston damping land is
replaced by a sealing ring, such as a two turn laminar sealing
ring, such as a Smalley laminar sealing ring. Such a ring is
generally available from manufacturers of spiral snap rings at a
relatively low cost. Either one, two or three of these rings
typically can be assembled into the actuating piston groove. The
radial spring action of the ring keeps the rings in contact with
the damping spacer 108", thus assuring low hydraulic fluid leakage.
Small holes can also be drilled through these rings to act as one
or more damping orifices 120", one of which is shown in FIG. 7B.
Alternatively, the damping orifices in the actuator piston of FIG.
2 can be used. An advantage of using the laminar sealing rings is
that the bore in the damping spacer can have a much relaxed
tolerance, and all that is necessary is that a reasonably smooth
surface be provided.
[0068] With reference to FIGS. 8-15, the electrically operated
electro-hydraulic valves are generally of a rotary design. The
electro-hydraulic valves 28 provide multiple paths for flow of the
hydraulic fluid, such that the sum of the open areas in the valve
is large, and relatively small rotational angles switch the
cylinder ports from a pressure supply configuration to a return
path configuration. Referring to FIGS. 8-11, the electrically
operated electro-hydraulic valves preferably include a rotor or
rotary valve element 130, assembled in combination with a three
path latched magnetic motor 132.
[0069] The rotor is provided with a pressure supply groove 134 that
communicates with a plurality of axial pressure grooves 136 that
branch from the pressure supply groove 134 and dead-end. A second
set of axial return grooves 138 is also provided in the rotor,
communicating at the opposing end of the rotor with the return to
the system via the engine oil sump, and are dead-ended at their
ends adjacent to the pressure supply groove. The rotor is
preferably manufactured of high strength, hardened steel or an
equivalent durable material. The outside diameter of the rotor is
typically machined to a high finish and is precision sized to fit
within the stator, or fixed valve element 140.
[0070] With reference to FIGS. 8 and 12-15, the stator is
preferably provided with an inlet pressure port and an inner bore
144, with which the inlet pressure port is in fluid communication
through a plurality of radially oriented holes 146. The stator also
includes a cylinder port groove 148 in fluid communication with the
inner bore and the axial grooves 136 and 138 of the rotor through a
plurality of axial stator slots 150. The stator is also preferably
fabricated of high strength, hardened steel or an equivalent
durable material, and the inside diameter is also typically
machined to a high finish and precision sized to mate with the
rotor. The stator is installed in a housing 152 that provides the
necessary fluid connections with the pressure supply and pressure
return lines of the hydraulic fluid system, and the rotary valve
housing 152 is assembled together with the housing 154 of the
magnetic motor assembly.
[0071] FIGS. 14 and 15 show the rotor and stator mated for
operation, with FIG. 15 illustrating how the pressure will be
distributed, in the valve cross-section. As can be readily
appreciated, alternate grooves of the rotor will be either
pressurized with the supply of pressurized hydraulic fluid, or will
be at return pressure, depending upon the orientation of the rotor
within the stator. The cylinder ports 150 are vented to the return
grooves 138, and when the rotor is turned, preferably 9% clockwise,
the cylinder ports will be connected to the pressure grooves 136. A
hydraulic actuator connected to the cylinder port will then receive
flow from six pressure grooves.
[0072] The open flow area of the valve depends upon the axial
length of the cylinder port slots, and the diameter of the
rotor-stator interface. The electrically driven magnetic motor
assembly, connected to the rotor, can thus on command rotate the
rotor first clockwise, and then counterclockwise, 9%. Other angles
of rotation may, of course, also be suitable. It should thus be
apparent that the rotary valve can open a very high flow area when
rotated through relatively small angles. If additional area is
required, the rotor and stator can be designed with increased
length and the stator provided with longer cylinder port slots, as
desired. In this manner, the valve design can be adapted to a
variety of applications. The rotor design also inherently provides
a very small rotational mass moment of inertia, since the numerous
grooves on the outside diameter of the rotor have removed a
substantial amount of material mass that would otherwise contribute
to rotational inertia of the rotor. The small rotational angle
required for operation of the rotary valve, and the low mass moment
of inertia of the rotary valve both optimize the operation of the
reciprocating valve actuation and control system of the invention
for operation at very high cyclic rates, with a low power
consumption by the electrical actuator.
[0073] The rotation of the cylindrical rotor element also entails
very low friction, since the radial loading on the rotor is
pressure balanced at all times, so that wear on the rotor and
stator of the rotary valve will be minimized. It should be readily
appreciated that the rotary valve design could easily be modified
to provide a return passage similar to that used for the inlet
pressure port, and an elongated version could also include a
secondary group of cylinder ports to create a four way valve. It
should also be readily appreciated that the rotor and stator are
ideally configured for manufacture by investment casting or metal
injection molding methods, which will permit greater economy in the
manufacturing process.
[0074] Referring to FIGS. 8, 9, 16 and 17, the electrically
operated electro-hydraulic valves preferably are provided with a
rotating motor driver capable of fast response to electrical
pulses, with magnetic latching at two positions. Briefly, the
magnetic motor consists of a three path magnetic circuit, with each
of the three paths meeting at a central point. Two of the magnetic
paths pass through individual magnetizing coils, while the third
path includes a rotor and a stationary permanent magnet that holds
or latches the rotary element in the position last commanded by the
engine computer.
[0075] As is best seen in FIG. 16, the first path of the magnetic
motor is comprised of a first pole piece 160, connected to a first
electromagnetic coil 162 energized by the electrical signals from
the engine computer, and the magnetic junction 164 connected to the
first pole piece and first coil. The second path of the magnetic
motor similarly is comprised of a second pole piece 166 connected
to a second electromagnetic coil 168 energized by electrical
signals from the engine computer, and the magnetic junction 164 to
which the second pole piece and second electromagnetic coil are
connected. The third path of the magnetic motor is comprised of the
magnetic rotor 170 mounted for rotation between a first position
and a second position contacting the first pole piece and second
pole piece, respectively, an air gap 172 between the magnetic rotor
and a third pole piece 174, a permanent magnet 176 connected to the
third pole piece, and a fourth pole piece 178 connected between the
permanent magnet and the magnetic junction 164. A rotary output
shaft 180 is provided on the rotor of the magnetic motor for
transferring the rotary motion of the rotor of the magnetic motor
to the rotor of the rotary valve 130. Referring to FIG. 17, the
first and second pole pieces are preferably arranged to form 30%
gaps at the end of the rotor of the magnetic motor, to provide
maximum leverage and maximum torque. When the rotor is attracted to
either the first pole piece or the second pole piece, the gap
between the rotor and one of the pole pieces closes, creating a
minimum reluctance path, and the permanent magnetic flux in the
third path of the magnetic motor latches the rotor of the magnetic
motor in place, as indicated by reference number 182.
[0076] The operation of the magnetic motor will be further
described with reference to FIGS. 16 and 17. If the permanent
magnet is oriented to produce a north pole at the rotor of the
magnetic motor, at rest, both the first and second pole pieces
would be at south polarity. The latched position then completes the
permanent magnet flux path, such that the north polarity end of the
rotor is magnetically latched to the south polarity of the pole
piece which the magnetic rotor contacts. In FIG. 17, the magnetic
rotor is shown latched to the first pole piece 160, so that in
order to move the magnetic rotor from the position shown to latch
with the second pole piece, the second electromagnetic coil 168 is
pulsed with direct current. The current flow in the second
electromagnetic coil is preferably phased to produce a strong south
pole at the second pole piece 166. When this occurs the second pole
piece attracts the magnetic rotor, and since the first pole piece
is on the opposite end of the magnetic path of the second
electromagnetic coil 168, the first pole piece assumes a north
polarity. Since the magnetic rotor is permanently provided with a
north polarity by the permanent magnet, the magnetic rotor is
repelled from the first pole piece, and is attracted to the second
pole piece. At the same time, the north polarity flux from the
second electromagnetic coil 168 enters the third path through the
junction 164, reinforcing the permanent magnet, and strengthening
the north polarity of the rotor. The magnetic rotor is then very
strongly urged to close the gap with the second pole piece, and
once this gap is closed, and the coil electrical pulse has ended,
the permanent magnetic flux from the third magnetic path latches
the magnetic rotor in contact with the second pole piece. If the
electromagnetic coil 162 is then pulsed, the opposite action
occurs, with the first pole piece acquiring a strong south
polarity, and the second pole piece acquiring a north polarity, and
the permanent magnet and magnet rotor receiving reinforcement of
the north polarity. The second pole piece then repels, and the
first pole piece attracts the magnetic rotor, and the permanent
magnet again latches the magnetic rotor to the new position at the
first pole piece. As should be readily apparent, the permanent
magnet may also be installed to produce a south polarity at the
rotor, at which both of the electromagnetic coils require current
flow phased to produce north polarity at the first and second pole
pieces. The resulting functions will then be the same as described
above, with all of the magnetic polarities described reversed.
[0077] Testing of the three path rotary latched magnetic motor has
shown that the motor is capable of very high speed operation. With
a rotation cyclic angle of 9%, cyclic rates of 260 Hertz can be
achieved with 12 volts, 5 ampere electrical pulses of 1.0 ms
duration (0.06 watt-seconds). At the 260 Hertz rate, the magnetic
motor drew a steady operational current of 1.172 RMS amperes.
[0078] The improvement in the speed of operation of the
reciprocating valve actuation and control system of the invention
can be readily appreciated with reference to FIG. 18, comparing
valve speeds of a mechanical camshaft driven engine and the camless
engine valve control system of the invention. The graph shows the
length of the valve stroke in inches vs. degrees of rotation of a
mechanical camshaft. When graphed, the cycle of opening and closing
of a valve driven by a mechanical camshaft will display a shape
similar to a sine curve. The period (as measured in crankshaft
degrees) remains constant for any engine load or rpm. However, the
cycle of opening and closing of valves driven by the reciprocating
valve actuating and control system of the invention operates much
faster. Designed to match valve opening rates at the maximum engine
rpm, the valve actuation and control system of the invention opens
the valve at this same rate regardless of engine operating
conditions. Thus, the valve actuation and control system of the
invention will match the valve rate at a maximum rpm of an engine,
but will be faster at all lesser engine speeds. Because of this
improved speed, the reciprocating valve actuation and control
system of the invention allows greater flexibility in programming
valve events, allowing for improved low end torque, lower emissions
and improved fuel economy.
[0079] The reciprocating valve actuation and control system has the
ability to alter the valve cyclical stroke number (i.e., 2 cycle)
to a desired valve cycle combination. It is therefore conceivable
to start and run an engine in standard 4 cycle mode, then change
over at some time to 2 cycle mode and effectively double the
potential available torque.
[0080] The reciprocating valve actuation and control system also
has the ability to control the effective engine speed without the
use of a throttle valve. This is accomplished by controlling the
valve duration from its idle duration to its maximum torque
duration as a function of the desired throttle position. This
allows simplification of the induction system and allows for a more
compact engine design. The throttle control abilities also provide
the ability to control an engine's volumetric efficiency under
certain conditions, and allow the engine to have a softer RPM
limiting function.
[0081] Upon sensing ignition switch shutoff of system power
failure, the reciprocating valve actuation and control system and
valve spring puts the valve in the most desirable "generally
closed" state, so that the valve positions are not ambiguous and
will thus protect engines from valve/valve or piston valve contact.
After the valve positions are guaranteed, the reciprocating valve
actuation and control system will turn off the power to itself, and
operations will cease.
[0082] The stored energy in the accumulator can be used for engine
power bursts. During these brief power bursts, the hydraulic pump
can be disengaged, allowing the valves to be powered solely from
stored energy from the accumulator with additional energy savings
derived by not operating the hydraulic pump. Also, during braking,
some energy that would normally be absorbed by the vehicle friction
braking system can be stored in the accumulator. This is possible
because the crankshaft (ultimately) is connected to the vehicle
wheels and can drive the hydraulic pump to fill the accumulator for
future hydraulic valve actuation.
[0083] A controller chip can eliminate the need for a half
crankshaft speed cam position sensor along with all of its
mechanical and electrical interfaces. (Typically the distributor or
cam position sensor.) The chip can calculate and determine overlap
and firing sequencing of a 2, 4, 5, 6, etc cycle engine during the
start-up sequencing.
[0084] While the preferred embodiment describes the use of engine
oil from the engine lubrication circuit, an alternative would be a
secondary fluid (e.g. diesel fuel, ATF, steering fluid, etc.). The
hydraulic fluid may be also be a separate system with another fluid
type on a separate fluid circuit. Also, the fluid return reservoir
may be the engine crankcase, or a separate and different
location.
[0085] By use of the invention, multiple intake or exhaust valves
of a cylinder need not open at the same time. A delay of even a
small amount can off-load the driver electronics and reduce peak
current load. This will allow smaller current traces on the circuit
board and prevent ringing of the power transistors. The delay of
the intake valves opening in a multi inlet valve cylinder can
enhance the swirl effect. Both opening and closing events of the
set of valves can be mapped to enhance various operating
characteristics. This effect can also be combined with the use of
shaped and directed inlet ports. The invention can also enhance
mechanical simplicity of the intake system. Installing a Pedal
Position Sensor at the velocity/accelerator pedal will allow
simplification of the induction system by eliminating throttle
plates and effectively throttling the engine using only the
conventional intake and exhaust valves that open into the
cylinder.
[0086] Since the invention allows broad control of a variety of
combination functions, an internal EGR function can be created by
commanding a second set of exhaust valve opening and closing events
during the intake sequence. Similarly, the intake valve may be
opened and closed several times during the intake or exhaust
sequence to promote scavenging and later to follow the piston to
promote intake volumetric optimization, and intake and exhaust
valves may be dithered to control engine throttling and
braking.
[0087] As a further indication of the benefits of the invention,
one intake port would be designed for high swirl (lower volume)
while a second intake port would be designed for high volume (lower
swirl). During throttled conditions, only the high swirl port would
be used to optimize combustion efficiency. If exhaust valves are
provided as different sizes, the smaller would be opened first so
as to substantially lower cylinder pressure prior to opening the
second exhaust valve. When both valves are of equal size, either
valve could be opened ahead of the second to again lower cylinder
pressure before opening the second valve. This sequencing may allow
the use of smaller valve actuators and certainly reduced energy to
operate the second valve. Engines with both multiple intake and
exhaust valves can be made to operate under higher conditions of
swirl. Although paired intake and exhaust valves may be of equal
size, swirl is maximized by having different sized valves and
properly sequencing them. Refer to FIG. 19. Sequence is as
follows:
[0088] a. #1 Intake valve 184 opens (largest valve)
[0089] b. #2 Intake valve 186 opens (smaller valve)
[0090] c. #2 Intake valve 186 closes
[0091] d. #1 Intake valve 184 closes
[0092] e. Compression and power stroke take place.
[0093] f. #4 Exhaust valve 190 opens (smaller valve w/less surface
area)
[0094] g. #3 Exhaust valve 188 opens (larger valve w/more
volume)
[0095] h. #3 Exhaust valve 188 closes
[0096] i. #1 Intake valve 184 opens (overlap begins)
[0097] j. #4 Exhaust valve 190 closes (overlap ends)
[0098] The invention can also effectively use a bridge in the
combustion chamber to assist swirl. In addition to valve size and
sequencing to promote higher swirl, the upper combustion chamber
may incorporate a "bridge" effectively separating the intake side
from the exhaust side in the dome of the combustion chamber. With
the "bridge" in place, gases would be better directed to flow in a
"swirl" pattern as shown in FIG. 19.
[0099] Using the invention, engines having multiple intake or
exhaust valves could be start sequenced having only one intake and
one exhaust valve operating. The invention permits reprogramming to
allow reverse engine rotation by simply inverting one input wire
pair. Reverse operation is advantageous to operation of marine
equipment having dual outdrives or T-drives, since vehicle
torsional accelerations are canceled by reverse rotational engines.
This feature would also eliminate the need for reverse gear(s) in
the transmission since forward gears would be used to operate in
either vehicle direction. This provides an opportunity for multiple
reverse gears without added hardware.
[0100] It will be apparent from the foregoing that while particular
forms of the invention have been illustrated and described, various
modifications can be made without departing from the spirit and
scope of the invention. Accordingly, it is not intended that the
invention be limited, except as by the appended claims.
* * * * *