U.S. patent application number 10/108246 was filed with the patent office on 2002-10-31 for hydraulic valve actuation systems and methods.
Invention is credited to Babbitt, Guy Robert, Raimao, Miguel Angelo, Turner, Christopher Wayne.
Application Number | 20020157623 10/108246 |
Document ID | / |
Family ID | 24931263 |
Filed Date | 2002-10-31 |
United States Patent
Application |
20020157623 |
Kind Code |
A1 |
Turner, Christopher Wayne ;
et al. |
October 31, 2002 |
HYDRAULIC VALVE ACTUATION SYSTEMS AND METHODS
Abstract
Hydraulic engine valve actuation systems and methods for
internal combustion engines. The systems utilize a proportional
valve to regulate the flow of a working fluid to and from a
hydraulic actuator controlling the engine valve position. The
position of the proportional valve is controlled by high speed
valves to control various engine valve parameters, including engine
valve takeoff and landing velocities. Returning all valves to a
known starting position between engine valve events avoids
accumulation of errors in proportional valve positioning.
Embodiments using spool valves for the high speed valves and the
proportional valve, and spring return and hydraulic return for the
engine valve, are disclosed. A specially shaped spool in the
proportional valve provides enhanced control over the engine valve
operation. Various further alternate embodiments are disclosed.
Inventors: |
Turner, Christopher Wayne;
(Fort Collins, CO) ; Raimao, Miguel Angelo;
(Manitou Springs, CA) ; Babbitt, Guy Robert;
(Colorado Springs, CO) |
Correspondence
Address: |
BLAKELY SOKOLOFF TAYLOR & ZAFMAN
12400 WILSHIRE BOULEVARD, SEVENTH FLOOR
LOS ANGELES
CA
90025
US
|
Family ID: |
24931263 |
Appl. No.: |
10/108246 |
Filed: |
March 25, 2002 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
10108246 |
Mar 25, 2002 |
|
|
|
09729487 |
Dec 4, 2000 |
|
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Current U.S.
Class: |
123/90.12 |
Current CPC
Class: |
F02B 2275/32 20130101;
F01L 2800/00 20130101; F01L 9/10 20210101; Y02T 10/12 20130101 |
Class at
Publication: |
123/90.12 |
International
Class: |
F01L 009/02 |
Claims
What is claimed is:
1. Apparatus for opening an engine valve comprising: a hydraulic
actuator disposed with respect to the valve to encourage the valve
toward a valve open position by the pressure of a fluid in the
hydraulic actuator; a proportional spool valve having a spool
hydraulically moveable between a first position coupling a source
of fluid under a first pressure to the hydraulic actuator and a
second position coupling the hydraulic actuator to a reservoir of
fluid under a second pressure, the second pressure being less than
the first pressure; electrically controlled valving hydraulically
controlling the position of the spool between the first and second
positions; and, a valve return returning the valve to a closed
position.
2. The apparatus of claim 1 wherein the spool valve has a third
position between the first and second positions blocking the source
of fluid under a first pressure from the hydraulic actuator
blocking the hydraulic actuator from the reservoir of fluid under
the second pressure.
3. The apparatus of claim 1 wherein the second pressure is
atmospheric pressure.
4. The apparatus of claim 1 wherein the spool is stepped in
diameter to shape the area versus spool position for flow between
the source of fluid under the first pressure and the hydraulic
actuator, and the hydraulic actuator to the reservoir of fluid
under the second pressure.
5. The apparatus of claim 1 wherein the valving comprises double
solenoid latching spool valves.
6. A method of opening an engine valve comprising: providing a
hydraulic actuator disposed with respect to the valve to encourage
the valve toward a valve open position by the pressure of a fluid
in the hydraulic actuator; coupling the hydraulic actuator to a
proportional spool valve having a spool hydraulically moveable
between a first position coupling a source of fluid under a first
pressure to the hydraulic actuator and a second position coupling
the hydraulic actuator to a reservoir of fluid under a second
pressure, the second pressure being less than the first pressure;
and, hydraulically controlling the position of the spool between
the first and second positions by electrically controlled valving.
Description
CROSS-REFERENCE TO RELATED APPLICATION
[0001] This application is a continuation of application Ser. No.
09/729,487, filed Dec. 4, 2000, entitled "Hydraulic Valve Actuation
Systems and Methods."
BACKGROUND OF THE INVENTION
[0002] 1. Field of the Invention
[0003] The present invention relates to the field of hydraulic
valve actuation for internal combustion engines.
[0004] 2. Prior Art
[0005] At the present time, piston-type internal combustion engines
of interest to the present invention are currently widely used in
automobiles, trucks, buses and various other mobile and stationary
power systems. Such engines include the common gasoline and diesel
engines, as well as similar engines operating from different fuels
such as liquid propane. These engines commonly utilize intake and
exhaust valves that are spring loaded to the closed position and
which are directly or indirectly opened at appropriate times by a
camshaft driven from the engine crankshaft. In a two-cycle engine
such as a two-cycle diesel engine, the camshaft will rotate in
synchronism with the engine crankshaft, though in a four-cycle
engine, the camshaft is driven through a two-to-one reduction drive
system (gear or chain or belt, etc.) to rotate at one-half the
engine crankshaft speed.
[0006] Camshaft actuation of engine valves historically has had a
number of advantages, resulting in its relatively universal use in
such engines for many decades. These advantages include high
reliability, particularly given the current level of development of
such cam actuated valve systems. Cam actuation is also relatively
cost effective, again given the state of development and quantities
in which it is produced. Cam actuation also has the advantage of
allowing shaping the cam to provide a smooth curve defining valve
position versus camshaft angle. This results in a rather low
velocity takeoff and initial valve opening, as well as a rather low
velocity valve final closing at low engine speeds, resulting in
minimum noise being generated. It also results in faster valve
opening and valve closing at higher engine speeds as required to
maintain the same valve timing throughout the engine speed
operating range.
[0007] Cam actuated valve systems also have certain limitations
which are becoming of increasing concern. In particular, optimal
valve timing is not fixed throughout the engine operating range.
For instance, valve timing for maximum power at one engine speed
will not be the same as valve timing for maximum power at another
engine speed. Accordingly, the classic cam operated valve systems
utilize a compromise valve timing, providing reasonable performance
over a reasonable range of engine operating conditions while being
less than optimal for most, if not at all, these conditions.
Further, valve timing for maximum power at any engine speed may not
be optimal from an engine emissions standpoint. Optimum valve
timing at any given engine speed may need to be dependent on engine
loading, and perhaps other parameters, such as air temperature, air
pressure, engine temperature, etc.
[0008] Recently, mechanisms have been introduced to attempt to make
up for some of the limitations in the fixed timing cam operated
valve systems. These mechanisms include mechanisms for varying
valve timing (but not valve opening duration in terms of camshaft
angle) with engine speed, as well as mechanisms for also increasing
the valve open duration. However, such mechanisms tend to be
complicated, open the valve a fixed distance under all engine
operating speeds and are limited in the number and range of
variables for which valve operation may begin to be optimized.
[0009] Recently various hydraulic systems for valve actuation have
been proposed. These systems offer the potential of more flexible
control of valve actuation parameters over the range of the various
engine operating parameters. The present invention is an
improvement on these systems.
BRIEF SUMMARY OF THE INVENTION
[0010] Hydraulic engine valve actuation systems and methods for
internal combustion engines. The systems utilize a proportional
valve to regulate the flow of a working fluid to and from a
hydraulic actuator controlling the engine valve position. The
position of the proportional valve is controlled by high speed
valves to control various engine valve parameters, including engine
valve takeoff and landing velocities. Returning all valves to a
known starting position between engine valve events avoids
accumulation of errors in proportional valve positioning.
Embodiments using spool valves for the high speed valves and the
proportional valve, and spring return and hydraulic return for the
engine valve, are disclosed.
[0011] To provide enhanced control over the engine valve operation,
a specially shaped spool in the proportional valve may be used to
shape the flow areas versus spool position. This allows more
gradual restricting of the flow areas versus spool movement over
selected portions of the possible spool positions, diminishing the
effect of small errors in spool position in such regions without
inhibiting the maximum flow areas when the spool is at its maximum
positions.
[0012] Various further alternate embodiments are disclosed.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] FIG. 1 is a block diagram of an exemplary configuration of a
system in accordance with the present invention.
[0014] FIG. 2 is a diagram illustrating the general structure and
function of the three-way proportional spool valve 24 of FIG.
1.
[0015] FIG. 3 is a perspective view of the spool 38 of the
proportional valve of FIG. 2.
[0016] FIG. 4 is an expanded view of an edge of the center land of
the spool 38 of FIG. 3.
[0017] FIGS. 5 and 6 are graphical representations of the flow area
versus spool position provided by the proportional valve 24 between
the high pressure rail and the chamber 26 of the valve actuator,
and between the chamber 26 and the vent 37, respectively.
[0018] FIG. 7 is a cross sectional view of an engine valve actuator
consisting of two concentric pistons that may be used with the
present invention.
[0019] FIG. 8 is a diagram of an embodiment of the present
invention that controls a hydraulically returned engine valve using
a closed center 3-way proportional valve.
[0020] FIG. 9 is a diagram of an embodiment of the present
invention that controls a hydraulically returned engine valve using
a closed center 4-way proportional valve.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
[0021] The present invention is a hydraulic valve operating system
for operating one or more intake valves or one or more exhaust
valves in a piston-type internal combustion engine, which provides
full flexibility in valve timing, valve duration, extent of
opening, and valve opening and closing velocity. Operation over the
desired range of these and other parameters may be controlled, and
more importantly optimized, for all engine operating conditions.
Such optimization may also include incrementally adjusting the
valve operation based on the valve operation during a previous
valve operating cycle. This is achieved by controlling the position
of a proportional valve by the use of pilot valves to control the
operating parameters of an intake or exhaust valve. In that regard,
a reference herein and in the claims to an "intake valve" or an
"exhaust valve," unless otherwise made clear by the context in
which the phrase is used, shall mean one or more intake valves for
a cylinder of an internal combustion engine, or one or more exhaust
valves of a cylinder of an internal combustion engine. Exemplary
embodiments of this system, sometimes referred to herein as a
"two-stage" system, are hereafter described in detail.
[0022] First referring to FIG. 1, a block diagram of an exemplary
configuration of a system in accordance with the present invention
may be seen. The system illustrated in FIG. 1 may be used to
actuate an intake or an exhaust valve. This 2-stage system consists
of 2 miniature 2-way digital latching spool valves 20 and 22
coupled to control the position of a 3-way proportional spool valve
24. The proportional spool valve, in turn, controls the flow area
into, and out of, a control volume 26. This control volume acts on
an actuator 28 to regulate the position of the engine valve 30. In
this embodiment, a spring return 32 is utilized for valve closing,
though embodiments with hydraulic valve closing may also be used,
as shall be subsequently described.
[0023] The 2 miniature 2-way digital latching spool valves 20 and
22 (referred to herein as pilot valves) may preferably be identical
valves, preferably in accordance with the 2 way valves disclosed in
U.S. Pat. No. 5,640,987 entitled Digital Two, Three, and Four Way
Solenoid Control Valves, issued Jun. 24, 1997, the disclosure of
which is incorporated herein by reference. Such valves are double
solenoid, high speed, magnetically latching spool valves, that as
used in the present invention, are operable between two positions.
The first position couples a first port to a second port for fluid
communication between the two ports, and the second position blocks
fluid communication between the first and second ports. While other
types of valves could be used, such as poppet valves, valves
generally of the type disclosed in the above referenced patent are
preferred because of their very high speed for good control, and
low energy consumption because of such capabilities as their
magnetic latching, and the ability to sense completion of
actuation, if used, to minimize heating above the already
relatively warm environment in which they operate. (See U.S. Pat.
Nos. 5,720,261 and 5,954,030.)
[0024] In the embodiment of the present invention of FIG. 1, valve
20 allows fluid flow from fluid line 34 to a drain line or
reservoir 37 (at a relatively low pressure, such as atmospheric
pressure) when in its first position, and blocks fluid flow from
fluid line 34 to the drain 37 when in its second position. Valve 22
allows fluid flow from a low pressure rail 36 to the fluid line 34
when in its first position, and blocks fluid flow from the low
pressure rail 36 to the fluid line 34 when in its second position.
Check valve 23 is optional, and is normally closed, as the
differential pressure on the check valve normally will not be in a
direction to open the valve. Its presence however, will help damp
transient pressure fluctuations and recover energy in the pressure
fluctuations.
[0025] Now referring to FIG. 2, a diagram illustrating the general
structure and function of the three-way proportional spool valve 24
of FIG. 1 may be seen. The proportional spool valve includes a
spool 38 within an internal housing 40 which fits within an
external housing assembly (not shown) with O-rings in O-ring
grooves 42 to separate the regions of ports 1, 2 and 3 from each
other and from the ends of the internal housing 40. In that regard,
the outer housing assembly, in addition to having the associated
fluid connections, also includes internal annual grooves adjacent
each of the regions identified as ports 1, 2 and 3 in FIG. 2, each
to act as a manifold region for the holes through the internal
housing 40 for fluid communication with a respective one of the
inner regions 44, 46 and 48 in the internal housing 40,
respectively. Fluid communication from each of the ports to the
associated inner region 44, 46 or 48 is provided in the exemplary
embodiment not only by through holes 50, but also by cooperatively
disposed orthogonal through holes 52 associated with each of the
ports.
[0026] As schematically illustrated in FIG. 2, the spool 38 is
positioned within the internal housing 40 by fluid pressures acting
on a piston at the left end of the spool having an effective area
Al and a piston at the right side of the spool having an effective
area of A.sub.2. As specifically illustrated in FIG. 2, the spool
38 is shown in its extreme right position, referred to herein as
its first position, as defined by stops on the travel of either the
pistons actuating the spool or stops acting on the spool itself. In
this position, the spool 38 is blocking fluid communication between
ports 3 and 2 and is allowing fluid communication between ports 2
and 1. Obviously, when the spool is at its left-most position,
referred to herein as its second position, fluid communication
between ports 1 and 2 is blocked and fluid communication between
ports 2 and 3 is enabled.
[0027] Normally in a spool valve, by way of example in the two
miniature, two-way digital latching spool valves 20 and 22 of FIG.
1, fluid communication between two adjacent ports will be blocked
when the spool is in one position and during the initial motion of
the spool toward the other position. However, once the relief on
the spool associated with the land in the housing separating the
regions coupled to the two adjacent ports starts to bridge the
land, a flow area between the regions coupled to the two ports is
established, that flow area increasing linearly with further motion
of the spool. Because that flow area is a peripheral flow area of
the full diameter of the spool, once opening starts, a relatively
large flow area between the two ports will be opened with only a
relatively small further motion of the spool.
[0028] However, in the three-way proportional spool valve 24 (FIG.
1), some of the details of which are illustrated in FIG. 2, this
change in flow area versus spool position is purposely modified to
reshape the flow area versus spool position. In the exemplary
embodiment, this is accomplished in the manner illustrated in FIGS.
3 and 4. In that regard, FIG. 3 is a perspective view of the spool
38 and FIG. 4 is an expanded view of an edge of the center land of
the spool 38. As may be seen in FIG. 3, the center land on the
spool has a plurality of kerfs 54 equally spaced around each end of
the center land, which kerfs begin to open a controlled flow area
with spool position prior to the edge of the land on the spool
reaching the edge of the land on the internal housing, the normal
position for a spool valve flow area starting to be
established.
[0029] In addition, as may be seen in FIG. 4, small steps are
ground in the center land of the spool of the three-way
proportional spool valve of the exemplary embodiment. Thus, while
the spool has an outer diameter Do having a close sliding fit
within the inner diameter of the internal housing, each end of the
center land of the exemplary spool has additional diameters
D.sub.1, D.sub.2 and D.sub.3, where D.sub.3 is less than D.sub.2,
D.sub.2 is less than D.sub.1 and D.sub.1 is less than D.sub.0. This
provides a non-linear variation in flow area versus spool position
during the opening and closing of the fluid communication between
adjacent ports, as illustrated in FIGS. 5 and 6. These figures
illustrate the flow area between ports 1 and 2, and ports 2 and 3,
respectively, versus the position of the spool in the three-way
proportional spool valve. As may be seen in FIG. 5, when the spool
is at the right-most position, the flow area between ports 1 and 2
is a maximum, initially decreasing at a relatively high rate for
the initial motion of the spool to the right, then decreasing in
rate for another part of the motion, then decreasing at a further
reduced rate to a substantially zero flow area for the rest of the
spool motion, essentially blocking communication between ports 1
and 2 when approximately 40% of the spool motion has been achieved.
In comparison, FIG. 6 shows the flow area between ports 2 and 3,
which is a mirror image of FIG. 5.
[0030] It will be noted from FIGS. 5 and 6 that in the exemplary
embodiment of the present invention, the reduction in flow area on
the initial valve closing motion of the spool occurs at a high rate
with respect to spool position, decreasing in change in flow rate
with an increasing position of the spool until the flow area goes
to substantially zero when less than half of the spool motion has
been achieved, thereby substantially altering the flow area versus
spool position characteristic of a conventional spool valve. Also,
because the flow area goes to substantially zero before one-half of
the maximum spool travel has been achieved, fluid communication
between both ports 1 and 2, and ports 2 and 3, is disabled or
blocked when the spool is approximately centered within its travel
range. For the specific exemplary embodiment illustrated, the
substantial blockage between both ports 1 and 2, and ports 2 and 3,
occurs whenever the spool's position is anywhere between
approximately 40% of its travel and 60% of its travel. Obviously
other shaping of the flow areas, or no shaping may be used if
desired, though preferably some shaping will be used to diminish
the effect of small errors in spool position in the restricted
regions without inhibiting the maximum flow areas when the spool is
at its maximum positions.
[0031] Referring again to FIG. 1, it may be seen that fluid in the
low pressure rail 36, which may have a pressure, by way of example,
of 20 to 50 bar, is coupled to the right side of the three-way
proportional spool valve 24 to act on the area A.sub.2 (FIG. 2) of
a piston encouraging the spool to its left-most position.
[0032] Assuming spool valve 22 is open and spool valve 20 is
closed, pressure in the low pressure rail 36 is communicated to
line 34, and thus acts on area A.sub.1 of the piston actuating the
spool of the proportional spool valve (FIGS. 1 and 2). Because the
area A.sub.1 is larger than the area A.sub.2, the spool of the
proportional spool valve is forced to its right-most position,
coupling port 1 and port 2 to couple chamber 26 to vent, allowing
the valve return spring 32 to force the valve 30 to the closed
position. Preferably area A.sub.1 is approximately twice area
A.sub.2 so that A.sub.1-A.sub.2 A.sub.2.
[0033] If the two-way valve 22 is closed and the two-way valve 20
is open, line 34 will be vented to the drain 37, so that the
pressure acting on piston area A.sub.1 (FIG. 2) of the three-way
proportional spool valve will be substantially zero. The pressure
acting on area A.sub.2 of the spool valve, however, will be equal
to the pressure of the low pressure rail 36, thereby creating an
unbalanced force on the spool to force the spool to its left-most
position. In this position, port 2 is in fluid communication with
port 3, communicating the pressure in the high pressure rail 56 to
control volume 26 to force the valve 30 open.
[0034] If, by way of example, valve 30 is half open and spool
valves 20 and 22 are both closed, then port 2 of the proportional
spool valve will be isolated from both ports 1 and 3, so that the
fluid in the control volume 26 is trapped, maintaining the valve 30
at its present position. Finally, since the two-way spool valves 20
and 22 are very high speed valves, they may be controlled in such
as manner as to rapidly controllably place the spool of the
proportional spool valve at any desired location within the
extremes of its travel, and thus variably control the flow rate of
fluid into or out of the control volume 26. This, in turn, allows
full control of the operating parameters of the valve 30, such as
the extent of opening, the timing and duration of opening, the
velocity profile of the opening and closing of the valve (which
profiles can be different from each other and/or vary with engine
operating conditions), and the final valve closing velocity with
engine rpm. This allows a relatively low velocity valve closing at
low engine rpm for low noise operation, while still allowing the
closing velocity to be increased with engine rpm, as necessary for
higher engine operating speeds.
[0035] The fluid used in the exemplary embodiment in the low
pressure rail, the high pressure rail and passed to drain is engine
operating oil, though other fluids may be used if desired. Since
the flow rates in the control system for valve 30 will vary with
various parameters, such as oil viscosity, and thus oil
temperature, and the pressure of the low pressure rail and the high
pressure rail, operation of the valve control system of FIG. 1 must
reasonably compensate for such variations. As a first order
approximation, these variations may be reasonably modeled so that
the control system as shown in FIG. 1 can reasonably vary operating
durations of valves 20 and 22 to at least approximate the desired
profile of the proportional valve spool position with engine
crankshaft angle, given the existing engine operating parameters
(speed, engine load, fuel temperature, air temperature, engine oil
temperature, atmospheric pressure, etc.).
[0036] In the exemplary embodiment, a small Hall effect sensor 58
is positioned adjacent actuator 28 for the valve 30 so as to
provide a feedback signal to the controller. Thus valve motion
during a valve operating cycle may be monitored and used to control
the operation of the valves 20 and 22 for that valve operating
cycle, and/or to make corrections in the next valve operating cycle
to more accurately achieve optimum valve operation for that valve
operating cycle. In that regard, more optimum operation may be
determined in any of various ways, including better compliance to a
predetermined valve position profile versus engine crank angle as
predetermined for the then existing engine operating conditions and
ambient conditions, or as determined by the effect of incremental
changes on one or more engine performance characteristics for the
change in valve operation just made, or a combination of both.
[0037] In the event two (or more) valves are being actuated in
unison by a single proportional valve 24, a sensor such as a
position sensor (Hall effect sensor or other position sensor) may
be used on only one of the valves, or on both valves, the sum of
the signals providing a better average indication of the position
profile of the two valves and the difference in the signals
providing fault detection, such as a sticky valve. While a position
sensor(s) is preferred, other types of sensors could be used, such
as a velocity sensor, as the integration times to convert to
position are short. In that regard, at the end of each valve
operating cycle, the control valve 22 is actuated to couple line 34
to the low pressure rail 36 and control valve 20 is actuated to
decouple line 34 from the drain 37 to bring the spool 38 to the
stop at the position shown schematically in FIG. 1. This provides
predetermined spool and pilot valve starting points for each valve
operating cycle so that errors in the spool valve position do not
accumulate, valve operating cycle to valve operating cycle. If
desired, a sensor may also be used to sense the position of the
proportional spool valve spool 38, though this is not
preferred.
[0038] Thus the two miniature latching valves 20 and 22 (sometimes
referred to herein as pilot valves) control the position of the
proportional valve 24. Specifically, the supply pilot valve 22
allows fluid to flow between a low-pressure rail 32 (approximately
20-50 bar) and a first piston used to move the proportional 3-way
valve. The vent pilot valve 20 will allow fluid to flow from the
piston to a vent at atmospheric pressure. Using these pilot valves,
the position of the proportional valve can be changed quickly and
accurately. The position of the proportional valve can be
infinitely varied throughout 3 flow states noted in FIGS. 5 and 6,
namely:
[0039] State 1: The high pressure fluid from the high pressure rail
56 (approximately 100-240 bar) is allowed to flow from the high
pressure rail to a control volume 26 above the engine valve
actuation piston.
[0040] State 2: The spool 38 of the proportional valve is centered
between its hard stops, trapping fluid in the control volume above
the engine valve actuation piston and creating a hydraulic
lock.
[0041] State 3: The fluid in the control volume 26 above the engine
valve actuation piston is vented to atmospheric pressure.
[0042] As the proportional valve moves from state 2 to state 3, the
area through which high-pressure fluid from the high pressure rail
56 can flow into the control volume 26 above the engine valve
actuation piston increases nonlinearly. (See FIG. 6). Similarly, as
the proportional valve moves from state 2 to state 1, the area
through which fluid can flow out of the control volume 26 above the
engine valve actuation piston to drain 37 increases nonlinearly
(See FIG. 5). Thus the geometry of the proportional spool valve has
been designed with regions of high and low gain. The low gain
region provides fine control for take-off and seating velocities,
while the high gain region provides the large flow area required to
achieve maximum engine valve velocities. This facilitates more
accurate control of the engine valve during seating and take-off.
These areas need more accuracy so that proper seating velocities
and valve overlap are achieved throughout the full range of engine
speed and temperature.
[0043] To better describe the function of the exemplary hydraulic
system, the following description traces the system through a
complete engine valve operating cycle, mimicking results from a
nodal hydraulics simulation. The specific simulation model used
100.degree. C. 0W30 synthetic motor oil at an engine speed of 6000
rpm, though simulations at lower engine speeds have also been
run.
[0044] An exemplary valve event may be described as follows.
Initially the supply pilot valve 22 is open and the vent pilot
valve 20 is closed (as illustrated in FIG. 1). This keeps the
proportional valve spool in the venting (rightmost) position (State
3, FIGS. 5 & 6). Specifically, the flow area between engine
valve actuation piston control volume 26 and vent 37 is at a
maximum (state 3, FIG. 6) and the area between engine valve
actuation piston control volume and the high-pressure rail is
closed (state 3, FIG. 5). As a result, the engine valve is forced
closed against its seat by the return spring 32.
[0045] To initiate valve opening, the supply pilot valve 22 is
opened and the vent pilot valve 20 is closed. This allows fluid to
flow from the control volume of the proportional spool valve to
vent. As a result, the proportional spool begins to move from state
3. The vent pilot valve 20 is left open long enough for the
proportional spool to pass through state 2 and into state 1.
However, the proportional valve is only allowed to travel until
just a small flow area in the low gain region of state 1 is open
between the high-pressure rail (FIG. 5) and engine valve actuation
piston control volume 26. This results in a slow take-off of the
engine valve. The speed of this take-off will vary depending on
where the proportional valve is stopped. Then the vent pilot valve
20 is opened once again so that the proportional spool moves to a
position that opens a larger flow area between the high pressure
rail and the engine valve actuation piston's control volume. This
results in a rapid opening of the engine valve after the initial
slow takeoff.
[0046] The engine valve now must stop at the desired lift, in this
particular example, 11 mm. To do this, the proportional spool will
be moved to state 2 in which the control volume above the engine
valve is hydraulically locked. This is achieved by closing the vent
pilot valve 20 and opening the supply pilot valve 22 for the
required amount of time. The engine valve will stay in this
position until it is commanded to return. At this point the kinetic
energy in the engine valve is fully converted into potential energy
of the fluid in the control volume and the engine valve return
spring. This trade off between kinetic and potential energy occurs
several times while the control volume is hydraulically locked,
which can result in a slight oscillation of the engine valve
position. To reduce this effect and to recover some of the kinetic
energy in the proportional valve spool, a check valve could also be
placed between the control volume 26 of the engine valve actuator
and the high-pressure rail 56 in order to damp out any high
pressure spikes that may occur during operation.
[0047] Next, the supply pilot valve 22 will be opened again long
enough to move the proportional valve to the high gain region of
state 3, and then closed, at least before vent pilot valve 20 is
again opened. To reiterate, at this point the flow area between the
engine valve control volume 26 and vent 37 is a maximum. Therefore,
the engine valve will accelerate very quickly toward its seat via
the stored energy in the valve spring.
[0048] In order to seat the valve at the desired velocity, the flow
area that connects the engine valve control volume 26 and vent 37
must be restricted. This can be achieved by once again opening the
vent pilot valve 20 for a short period to reposition the
proportional valve to a low gain in state 3. This seating velocity
will change depending on where the proportional valve is stopped in
this region.
[0049] This completes one engine valve cycle. In order to prepare
the system for the next event, all components will be repositioned
to their initial conditions. The only component that is out of
position is the proportional valve. The supply pilot valve 22 is
again opened, returning the proportional valve to a position of
maximum flow area in state 3. This reestablishes a reference point
at the beginning of each valve event, so that errors in
proportional valve positioning do not accumulate, one valve cycle
to another. In this way, the seating velocities desired at
different engine speeds, loads and temperatures can be achieved by
changing the position at which the proportional spool dwells. This
can be facilitated further by varying the pressure in the
low-pressure rail if desired, thus accomplishing finer control of
the proportional spool valve.
[0050] In a simulation of the system described above, an engine
valve actuator consisting of two concentric pistons was used, as
illustrated in FIG. 7. Instead of using one actuator with a
relatively large area exposed to pressure to drive the engine valve
through its entire stroke, the large piston (boost piston 60) is
used only initially to achieve peak velocities before reaching a
mechanical stop, while the remainder of the stroke is accomplished
using a smaller telescoping piston (drive piston 62). Specifically,
when the engine valve, particularly an exhaust valve, initiates
lift from its seat, in-cylinder pressure remains substantial. In
addition, maximum engine valve acceleration is also required at
this time. As a result, a greater force is needed to actuate the
engine valve through the beginning of its stroke while a much lower
force is required for the remainder of the stroke. The present
invention system does not rely on using the two concentric piston
design, as it will also function if just one actuator is used.
However, the two concentric piston design requires less fluid from
the high pressure rail for each valve cycle, and thus requires less
energy for valve operation.
[0051] One can also use a hydraulically returned engine valve as
opposed to a spring returned valve, as illustrated in FIG. 8. In
this embodiment, the valve actuator comprises a piston 64 having a
cross-sectional area A.sub.1 on a piston rod 66 having a
cross-sectional area A.sub.2. Chamber 68 is permanently coupled to
the high pressure rail, and chamber 70 is switchable by the
proportional valve between the high pressure rail and the vent.
Consequently, the maximum valve opening force is equal to the
pressure of the high pressure rail times A.sub.2 and the maximum
valve closing force is equal to the pressure of the high pressure
rail times A.sub.1-A.sub.2. Because of the functional relationship
between spring force and stroke, one can achieve essentially the
same valve dynamics as a hydraulically returned valve with a
smaller diameter actuator.
[0052] With a return spring, the spring closing force is at a
minimum when one desires a large opening force for maximum
acceleration against peak cylinder pressure. With hydraulic return,
the closing force of a hydraulically returned engine valve is
constant and therefore will be higher than that of a spring when
the valve is seated. Therefore, the force characteristic of a
mechanical spring is more desirable for returning the engine valves
than a single piston return mechanism.
[0053] Instead of using a closed center 3-way proportional valve,
the hydraulically returned system can also be constructed using a
closed center 4-way proportional valve (FIG. 9). Like the closed
center 3-way proportional valve, its position can also be
infinitely varied throughout 3 flow states.
[0054] State 1: The high pressure fluid is allowed to flow from the
high pressure rail to a control volume 70 above the engine valve
actuation piston 64 while the fluid in chamber 68 acting below the
engine valve actuation piston 64 is vented to tank.
[0055] State 2: The proportional valve is centered between its hard
stops, trapping fluid in the control volume 70 above the engine
valve actuation piston and in the control volume 68 below the
engine valve actuation piston, thus creating a hydraulic lock.
[0056] State 3: The fluid in the control volume 70 above the engine
valve actuation piston is vented to atmospheric pressure while high
pressure fluid is allowed to flow from the high pressure rail to
the control volume 68 below the engine valve actuation piston
64.
[0057] As the proportional valve moves from state 2 to state 1 the
area through which high-pressure fluid can flow into the control
volume 70 above the engine valve actuation piston 64 increases
nonlinearly (similar to FIGS. 5 & 6). At the same time the flow
area between the fluid below the engine valve actuation piston 64
and tank increases nonlinearly. Similarly, as the proportional
valve moves from state 2 to state 3, the area through which fluid
can flow out of the control volume 70 above the engine valve
actuation piston 64 to tank increases nonlinearly. At the same time
the flow area between the fluid below the engine valve actuation
piston 64 and the high-pressure rail increases nonlinearly.
[0058] In all the systems described, the proportional valve uses
hydraulic force to oppose the pressure in its control volume.
Alternatively, the proportional valve can also use a spring to
supply part or all of the opposing force.
[0059] Any of the 3-way proportional valve systems can take
advantage of the recovery systems known in the art. The
low-pressure rail used for actuating the 3 or 4 way proportional
valve can be used for the low-pressure source of the recovery
system if that system is implemented.
[0060] The present invention has many advantages for both diesel
and gasoline engines, as well as similar engines powered with
alternate fuels. These advantages include:
[0061] Infinitely variable engine valve timing for both opening and
closing times.
[0062] Infinitely variable engine valve lift from the valves seat
to its maximum lift position.
[0063] Infinitely variable valve open and/or close time
duration.
[0064] The proportional 3 or 4 way valve has low gain flow regions
for fine control at valve take-off and seating. It also has high
gain flow regions for maximum flow allowing increased speed of the
engine valve so that airflow into the engine cylinders can be
maximized.
[0065] The system can allow the engine valve profile to be
non-symmetric.
[0066] The system is capable of an infinitely varying the slew rate
of the engine valve independent of rail pressure.
[0067] The system does not require a slow take-off and landing.
Specifically, the valve can begin opening with maximum acceleration
or seat at maximum velocity if desired.
[0068] The system does not need a lash adjustment system,
specifically:
[0069] the system is unaffected and can compensate for the growth
of engine components (specifically valve train components) due to
thermal expansion,
[0070] the system is unaffected and can compensate for engine valve
recession due to wear of the valve seat and the engine valve,
and
[0071] the system is unaffected and can compensate for tolerance
stack up between valve train components resulting from initial
assembly and manufacturing tolerances.
[0072] The system can compensate for varying working fluid
viscosity due to temperature, age, etc.
[0073] The system can optimize the amount time at which air is
metered into the engine cylinder thus optimizing the combustion
event at the full spectrum of engine operating conditions resulting
in:
[0074] maximum power,
[0075] lower emissions,
[0076] reduced emissions by controlling fuel/air mixing,
[0077] reduced heat rejection by reduction of unnecessary in
cylinder air motion, and
[0078] high BMEP combustion schemes to improve catalyst light-off,
reduce startup emissions.
[0079] The system can be operated in such a way that engine braking
will result, specifically by shutting off the injector during
braking and opening the exhaust valve at the top of the compression
stroke to dissipate the compression energy.
[0080] The engine cycle can be varied to allow for:
[0081] 2 stroke operation.
[0082] Multiple stroke operation (such as, by way of example
2-stroke to 4-stroke, 4-stroke to 6-stroke or 8-stroke operation,
etc.) by eliminating one or more pairs of strokes from the normal
engine operating cycle, with the valves being controlled during
these pairs of strokes for minimum energy loss and/or other
considerations.
[0083] The system can allow for internal exhaust gas recirculation
(EGR). As a result the EGR valve can be removed.
[0084] Variable compression ratio.
[0085] Miller cycle operation--Maximum cylinder pressure control
with high expansion ratio for maximum thermodynamic efficiency.
[0086] Atkinson cycle operation.
[0087] Reduced heat rejection by reduction of unnecessary in
cylinder air motion.
[0088] High BMEP combustion schemes to improve Catalyst light-off,
reduce startup emissions.
[0089] Improved Cranking and Cold Start.
[0090] Reduced white smoke and diesel "fuel" smell during
startup/cold temperature idle/high altitude operation
[0091] High altitude compensation.
[0092] Variable torque curves to better fit duty/drive cycle of
vehicle.
[0093] Increased torque at low speeds for better driveability,
potential vehicle fuel economy improvements.
[0094] The system will operate more efficiently with a sequentially
apportioned pump.
[0095] The low-pressure rail can be replaced with an accumulator
that is supplied by the return flow of the engine valve
actuator.
[0096] Because the engine valve motion can be varied so that air
can be throttled at the engine valve, the throttle body can be
eliminated.
[0097] Operation of the turbo can be optimized at all engine
operating conditions.
[0098] Cylinder deactivation for improved vehicle fuel economy.
[0099] This 2-stage system has the capability of satisfactorily
controlling engine valves at very high engine speeds (from idle
speeds to 10,000 RPM). In addition, the critical regions of valve
take-off and seating can be controlled with accuracy and precision
while providing the features of infinitely variable valve timing,
duration and lift. The system also has the capability of
significantly increasing the amount of air that can be supplied to
an engine's cylinders throughout the full range of engine speed by
adjusting valve timing and duration to maximize the dynamic effects
of flow into and out of the combustion chamber at all engine
speeds.
[0100] While an exemplary embodiment and various alternate
embodiments of the present invention have been disclosed herein, it
will be obvious to those skilled in the art that various changes in
form and detail may be made therein without departing from the
spirit and scope of the invention.
* * * * *