U.S. patent application number 13/359532 was filed with the patent office on 2012-08-02 for variable force valve spring.
This patent application is currently assigned to SCUDERI GROUP, LLC. Invention is credited to Riccardo Meldolesi, Joseph Paturzo, John Schwoerer.
Application Number | 20120192817 13/359532 |
Document ID | / |
Family ID | 46576285 |
Filed Date | 2012-08-02 |
United States Patent
Application |
20120192817 |
Kind Code |
A1 |
Meldolesi; Riccardo ; et
al. |
August 2, 2012 |
VARIABLE FORCE VALVE SPRING
Abstract
Devices and related methods are disclosed that generally involve
variable force valve springs for controlling the motion of an
engine valve. The force exerted by the valve spring can be adjusted
by altering the pressure at which a fluid is supplied to a fluid
chamber thereof, by altering the volume of the fluid chamber,
and/or by changing the aggregate surface area over which fluid
pressure is coupled to the engine valve. Associated fluid control
systems are also disclosed herein, as are various methods for
adjusting the force of a valve spring based on a variety of engine
parameters, such as engine speed, engine load, and/or a combination
thereof.
Inventors: |
Meldolesi; Riccardo; (West
Sussex, GB) ; Schwoerer; John; (Storrs, CT) ;
Paturzo; Joseph; (Avon, CT) |
Assignee: |
SCUDERI GROUP, LLC
West Springfield
MA
|
Family ID: |
46576285 |
Appl. No.: |
13/359532 |
Filed: |
January 27, 2012 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
61436739 |
Jan 27, 2011 |
|
|
|
Current U.S.
Class: |
123/90.14 ;
29/888.03 |
Current CPC
Class: |
F01L 3/22 20130101; Y10T
29/49247 20150115; F01L 1/465 20130101; F02B 33/22 20130101; F01L
2003/258 20130101; F01L 3/10 20130101 |
Class at
Publication: |
123/90.14 ;
29/888.03 |
International
Class: |
F01L 9/00 20060101
F01L009/00; B21K 1/76 20060101 B21K001/76 |
Claims
1. An engine, comprising: a pneumatic valve spring coupled to an
engine valve; and a control module configured to adjust a spring
force of the pneumatic valve spring based on at least one of engine
speed and engine load.
2. The engine of claim 1, wherein the control module is configured
to increase the spring force when the engine speed exceeds a first
speed threshold.
3. The engine of claim 1, wherein the control module is configured
to increase the spring force when the engine load exceeds a first
load threshold.
4. The engine of claim 1, wherein the control module is configured
to decrease the spring force when the engine speed is below a first
speed threshold.
5. The engine of claim 1, wherein the control module is configured
to decrease the spring force when the engine load is below a first
load threshold.
6. The engine of claim 2, wherein the first speed threshold is 3000
rpm.
7. The engine of claim 2, wherein the first speed threshold is four
times higher than an idle speed of the engine.
8. The engine of claim 3, wherein the first load threshold is 50
percent of a maximum load of the engine.
9. The engine of claim 3, wherein the first load threshold is three
times higher than an idle load of the engine.
10. A valve spring system, comprising: a vessel having a piston
reciprocally disposed therein; an engine valve coupled to the
piston; a first source of pressurized air having a first pressure;
a second source of pressurized air having a second pressure that is
greater than the first pressure; and a control valve configured to
adjust a force exerted on the piston by controlling a first fluid
communication between the vessel and the first source, and a second
fluid communication between the vessel and the second source.
11. The valve spring system of claim 10, wherein the engine valve
is actuated by a lost motion system.
12. The valve spring system of claim 10, wherein the engine valve
is outwardly-opening.
13. The valve spring system of claim 10, wherein the engine valve
is a crossover valve in a split-cycle engine.
14. The valve spring system of claim 13, wherein the second source
is supplied from a crossover passage of the split-cycle engine.
15. The valve spring system of claim 13, wherein a first force
exerted on the engine valve by a crossover charge in the crossover
passage and a second force exerted on the engine valve by the
piston together exceed a combustion force exerted on the engine
valve.
16. The valve spring system of claim 10, wherein the first pressure
is between about 1 bar and about 10 bar.
17. The valve spring system of claim 10, wherein the second
pressure is between about 20 bar and about 85 bar.
18. The valve spring system of claim 10, further comprising a
control module configured to actuate the control valve.
19. An engine comprising the valve spring system of claim 10.
20. A valve spring system, comprising: a cap defining a first bore
therein; a piston coupled to an engine valve, the piston being
slidably disposed within the first bore such that the piston and
the cap define a first fluid chamber; an outer housing defining a
second bore therein, the cap being slidably disposed within the
second bore such that the cap and the outer housing define a second
fluid chamber; and a fluid port configured to selectively release
pressurized fluid from the second fluid chamber, thereby allowing
the cap to slide relative to the outer housing such that a volume
of the first fluid chamber is increased.
21. The valve spring system of claim 20, wherein the cap includes
at least one aperture in fluid communication with the first fluid
chamber.
22. The valve spring system of claim 21, wherein the aperture
remains in fluid communication with a pressurized fluid supply line
throughout a full engine cycle.
23. The valve spring system of claim 21, wherein fluid disposed in
the first fluid chamber is forced out of the first fluid chamber
through the aperture when the engine valve is open.
24. The valve spring system of claim 21, wherein the first fluid
chamber is filled with fluid through the aperture when the engine
valve is closed.
25. The valve spring system of claim 20, wherein the pressure in
the first fluid chamber is substantially equal to the pressure in
the second fluid chamber when the engine valve is closed for at
least 100 degrees crank angle.
26. A method of varying a spring force exerted on an engine valve
by the valve spring system of claim 20, comprising: coupling the
fluid port to a vent to release pressure from the second fluid
chamber and increase the spring force exerted on the engine valve;
and coupling the fluid port to a pressurized fluid supply to
increase pressure in the second fluid chamber and decrease the
spring force exerted on the engine valve.
27. The method of claim 26, wherein the spring force is increased
when a load of the engine exceeds a first predetermined threshold
and is decreased when the load of the engine is below a second
predetermined threshold.
28. The method of claim 26, wherein the spring force is increased
when a speed of the engine exceeds a first predetermined threshold
and is decreased when the speed of the engine is below a second
predetermined threshold.
29. A valve spring system, comprising: a first housing defining a
first bore therein in which a first piston is reciprocally
disposed, the first piston being coupled to a valve stem of an
engine valve; a second housing defining a second bore therein in
which a second piston is reciprocally disposed, the second piston
being coupled to an extension stem; a third housing defining a
hydraulic plenum, the hydraulic plenum being in fluid communication
with a proximal end of the valve stem, a distal end of the
extension stem, and a control valve; wherein the plenum is in fluid
communication with an accumulator when the control valve is opened
such that the valve stem is movable independently from the
extension stem; and wherein the plenum is sealed when the control
valve is closed such that movement of the valve stem requires
movement of the extension stem.
30. The valve spring system of claim 29, wherein the first housing,
the second housing, and the third housing are formed
integrally.
31. The valve spring system of claim 29, wherein a first fluid
chamber defined by the first housing and the first piston is
supplied with pressurized fluid through a first aperture formed in
the first housing and wherein a second fluid chamber defined by the
second housing and the second piston is supplied with pressurized
fluid through a second aperture formed in the second housing.
32. The valve spring system of claim 31, wherein fluid disposed in
the first fluid chamber is forced out of the first fluid chamber
through the first aperture when the engine valve is open.
33. The valve spring system of claim 31, wherein fluid disposed in
the second fluid chamber is forced out of the second fluid chamber
through the second aperture when the engine valve is open and the
control valve is closed.
34. A method of varying a spring force exerted on an engine valve
by the valve spring system of claim 29, comprising: closing the
control valve to increase the aggregate surface area over which
fluid pressure is coupled to the engine valve and thereby increase
the spring force; and opening the control valve to decrease the
aggregate surface area over which fluid pressure is coupled to the
engine valve and thereby decrease the spring force.
35. The method of claim 34, wherein the spring force is increased
when a load of the engine exceeds a first predetermined threshold
and is decreased when the load of the engine is below a second
predetermined threshold.
36. The method of claim 34, wherein the spring force is increased
when a speed of the engine exceeds a first predetermined threshold
and is decreased when the speed of the engine is below a second
predetermined threshold.
37. A method of actuating a valve of an engine, comprising:
increasing a spring force of a valve spring associated with the
valve based on an increase in an engine parameter; and decreasing
the spring force of the valve spring based on a decrease in the
engine parameter.
38. The method of claim 37, wherein the engine parameter is
selected from the group consisting of engine speed, engine load,
engine temperature, throttle position, and engine age.
39. The method of claim 38, wherein the spring force is increased
by increasing a pressure supplied to a fluid chamber of the valve
spring.
40. The method of claim 38, wherein the spring force is increased
by decreasing a volume of a fluid chamber of the valve spring.
41. The method of claim 38, wherein the spring force is increased
by increasing an aggregate surface area over which fluid pressure
is coupled to the engine valve.
Description
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] This application claims the benefit of priority of U.S.
Provisional Patent Application No. 61/436,739, filed on Jan. 27,
2011, the entire contents of which are incorporated herein by
reference.
FIELD
[0002] The present invention relates to methods and devices for
actuating an engine valve. More particularly, the invention relates
to variable force valve springs and methods related thereto.
BACKGROUND
[0003] Internal combustion engines generally include one or more
valves for controlling the flow of air and fuel through the engine.
These valves are usually actuated by a mechanical cam. For example,
a rotating shaft having a teardrop-shaped cam lobe can be
configured to impart motion to the valve, either directly or via
one or more intermediate elements. As the shaft rotates, the
eccentric portion of the cam lobe imparts a linear motion to the
valve over a range of the shaft's rotation.
[0004] It is often necessary to bias the valve towards a particular
position (i.e., opened or closed). In conventional internal
combustion engines, the valves are generally inwardly-opening
(i.e., opening into the combustion chamber and toward the piston)
and are biased to a closed position by one or more mechanical valve
springs. One problem with this system is that there is no way to
alter or control the amount of force required to compress the valve
spring. The mechanical springs are generally chosen such that they
have sufficient fitted force and are stiff enough to handle the
most demanding of the conditions the engine will experience over
its operating range, which may only be encountered in very rare
situations. In other words, the valve spring is far stronger than
what is required for the majority of the engine's operating range.
In conventional internal combustion engines, a stronger valve
spring produces higher rocker shaft and cam shaft bearing friction,
contributing to the parasitic losses of the engine and reducing
overall engine efficiency.
[0005] Traditional mechanical valve springs are also poorly suited
for "lost-motion" valve train systems. A lost-motion system is
generally a system in which a lost-motion valve train element is
selectively actuated to operatively disconnect the valve from a cam
during a portion of the cam's rotation. In a valve train having no
lost-motion system, some of the energy exerted to compress the
valve spring is returned to the cam and ultimately to the rotation
of the engine when the spring expands against the closing ramp of
the cam. In lost-motion systems, however, the valve can be closed
earlier than what is called for by the cam by operatively
disconnecting the valve from the cam. In this case, the energy
stored in the compressed valve spring of an open valve is not
returned to the cam, and is instead applied only to the valve as it
closes. Since none of the energy is returned to the cam, the energy
spent to compress the valve spring in the first place is
essentially wasted.
[0006] Accordingly, there is a need for improved valve springs and
related methods.
[0007] For purposes of clarity, the term "conventional engine" as
used in the present application refers to an internal combustion
engine wherein all four strokes of the well-known Otto cycle (the
intake, compression, expansion and exhaust strokes) are contained
in each piston/cylinder combination of the engine. Each stroke
requires one half revolution of the crankshaft (180 degrees crank
angle ("CA")), and two full revolutions of the crankshaft (720
degrees CA) are required to complete the entire Otto cycle in each
cylinder of a conventional engine.
[0008] Also, for purposes of clarity, the following definition is
offered for the term "split-cycle engine" as may be applied to
engines disclosed in the prior art and as referred to in the
present application.
[0009] A split-cycle engine generally comprises:
[0010] a crankshaft rotatable about a crankshaft axis;
[0011] a compression piston slidably received within a compression
cylinder and operatively connected to the crankshaft such that the
compression piston reciprocates through an intake stroke and a
compression stroke during a single rotation of the crankshaft;
[0012] an expansion (power) piston slidably received within an
expansion cylinder and operatively connected to the crankshaft such
that the expansion piston reciprocates through an expansion stroke
and an exhaust stroke during a single rotation of the crankshaft;
and
[0013] a crossover passage interconnecting the compression and
expansion cylinders, the crossover passage including at least a
crossover expansion (XovrE) valve disposed therein, but more
preferably including a crossover compression (XovrC) valve and a
crossover expansion (XovrE) valve defining a pressure chamber
therebetween.
[0014] A split-cycle air hybrid engine combines a split-cycle
engine with an air reservoir and various controls. This combination
enables the engine to store energy in the form of compressed air in
the air reservoir. The compressed air in the air reservoir is later
used in the expansion cylinder to power the crankshaft. In general,
a split-cycle air hybrid engine as referred to herein
comprises:
[0015] a crankshaft rotatable about a crankshaft axis;
[0016] a compression piston slidably received within a compression
cylinder and operatively connected to the crankshaft such that the
compression piston reciprocates through an intake stroke and a
compression stroke during a single rotation of the crankshaft;
[0017] an expansion (power) piston slidably received within an
expansion cylinder and operatively connected to the crankshaft such
that the expansion piston reciprocates through an expansion stroke
and an exhaust stroke during a single rotation of the
crankshaft;
[0018] a crossover passage (port) interconnecting the compression
and expansion cylinders, the crossover passage including at least a
crossover expansion (XovrE) valve disposed therein, but more
preferably a crossover compression (XovrC) valve and a crossover
expansion (XovrE) valve defining a pressure chamber therebetween;
and
[0019] an air reservoir operatively connected to the crossover
passage and selectively operable to store compressed air from the
compression cylinder and to deliver compressed air to the expansion
cylinder.
[0020] FIG. 1 illustrates one exemplary embodiment of a prior art
split-cycle air hybrid engine. The split-cycle engine 100 replaces
two adjacent cylinders of a conventional engine with a combination
of one compression cylinder 102 and one expansion cylinder 104. The
compression cylinder 102 and the expansion cylinder 104 are formed
in an engine block in which a crankshaft 106 is rotatably mounted.
Upper ends of the cylinders 102, 104 are closed by a cylinder head
130. The crankshaft 106 includes axially displaced and angularly
offset first and second crank throws 126, 128, having a phase angle
therebetween. The first crank throw 126 is pivotally joined by a
first connecting rod 138 to a compression piston 110 and the second
crank throw 128 is pivotally joined by a second connecting rod 140
to an expansion piston 120 to reciprocate the pistons 110, 120 in
their respective cylinders 102, 104 in a timed relation determined
by the angular offset of the crank throws and the geometric
relationships of the cylinders, crank, and pistons. Alternative
mechanisms for relating the motion and timing of the pistons can be
utilized if desired. The rotational direction of the crankshaft and
the relative motions of the pistons near their bottom dead center
(BDC) positions are indicated by the arrows associated in the
drawings with their corresponding components.
[0021] The four strokes of the Otto cycle are thus "split" over the
two cylinders 102 and 104 such that the compression cylinder 102
contains the intake and compression strokes and the expansion
cylinder 104 contains the expansion and exhaust strokes. The Otto
cycle is therefore completed in these two cylinders 102, 104 once
per crankshaft 106 revolution (360 degrees CA).
[0022] During the intake stroke, intake air is drawn into the
compression cylinder 102 through an inwardly-opening (opening
inward into the cylinder and toward the piston) poppet intake valve
108. During the compression stroke, a compression piston 110
pressurizes the air charge and drives the air charge through a
crossover passage 112, which acts as the intake passage for the
expansion cylinder 104. The engine 100 can have one or more
crossover passages 112.
[0023] The volumetric (or geometric) compression ratio of the
compression cylinder 102 of the split-cycle engine 100 (and for
split-cycle engines in general) is herein referred to as the
"compression ratio" of the split-cycle engine. The volumetric (or
geometric) compression ratio of the expansion cylinder 104 of the
engine 100 (and for split-cycle engines in general) is herein
referred to as the "expansion ratio" of the split-cycle engine. The
volumetric compression ratio of a cylinder is well known in the art
as the ratio of the enclosed (or trapped) volume in the cylinder
(including all recesses) when a piston reciprocating therein is at
its bottom dead center (BDC) position to the enclosed volume (i.e.,
clearance volume) in the cylinder when said piston is at its top
dead center (TDC) position. Specifically for split-cycle engines as
defined herein, the compression ratio of a compression cylinder is
determined when the XovrC valve is closed. Also specifically for
split-cycle engines as defined herein, the expansion ratio of an
expansion cylinder is determined when the XovrE valve is
closed.
[0024] Due to very high volumetric compression ratios (e.g., 20 to
1, 30 to 1, 40 to 1, or greater) within the compression cylinder
102, an outwardly-opening (opening outwardly away from the cylinder
and piston) poppet crossover compression (XovrC) valve 114 at the
crossover passage inlet is used to control flow from the
compression cylinder 102 into the crossover passage 112. Due to
very high volumetric compression ratios (e.g., 20 to 1, 30 to 1, 40
to 1, or greater) within the expansion cylinder 104, an
outwardly-opening poppet crossover expansion (XovrE) valve 116 at
the outlet of the crossover passage 112 controls flow from the
crossover passage 112 into the expansion cylinder 104. The
actuation rates and phasing of the XovrC and XovrE valves 114, 116
are timed to maintain pressure in the crossover passage 112 at a
high minimum pressure (typically 20 bar or higher at full load)
during all four strokes of the Otto cycle.
[0025] At least one fuel injector 118 injects fuel into the
pressurized air at the exit end of the crossover passage 112 in
coordination with the XovrE valve 116 opening. Alternatively, or in
addition, fuel can be injected directly into the expansion cylinder
104. The fuel-air charge fully enters the expansion cylinder 104
shortly after the expansion piston 120 reaches its top dead center
("TDC") position. As the piston 120 begins its descent from its TDC
position, and while the XovrE valve 116 is still open, one or more
spark plugs 122 are fired to initiate combustion (typically between
10 to 20 degrees CA after TDC of the expansion piston 120).
Combustion can be initiated while the expansion piston is between 1
and 30 degrees CA past its TDC position. More preferably,
combustion can be initiated while the expansion piston is between 5
and 25 degrees CA past its TDC position. Most preferably,
combustion can be initiated while the expansion piston is between
10 and 20 degrees CA past its TDC position. Additionally,
combustion can be initiated through other ignition devices and/or
methods, such as with glow plugs, microwave ignition devices, or
through compression ignition methods.
[0026] The XovrE valve 116 is then closed before the resulting
combustion event enters the crossover passage 112. The combustion
event drives the expansion piston 120 downward in a power stroke.
Exhaust gases are pumped out of the expansion cylinder 104 through
an inwardly-opening poppet exhaust valve 124 during the exhaust
stroke.
[0027] With the split-cycle engine concept, the geometric engine
parameters (i.e., bore, stroke, connecting rod length, compression
ratio, etc.) of the compression and expansion cylinders are
generally independent from one another. For example, the crank
throws 126, 128 for the compression cylinder 102 and expansion
cylinder 104, respectively, have different radii and are phased
apart from one another with TDC of the expansion piston 120
occurring prior to TDC of the compression piston 110. This
independence enables the split-cycle engine to potentially achieve
higher efficiency levels and greater torques than typical
four-stroke engines.
[0028] The geometric independence of engine parameters in the
split-cycle engine 100 is also one of the main reasons why pressure
can be maintained in the crossover passage 112 as discussed
earlier. Specifically, the expansion piston 120 reaches its top
dead center position prior to the compression piston 110 reaching
its top dead center position by a discrete phase angle (typically
between 10 and 30 crank angle degrees). This phase angle, together
with proper timing of the XovrC valve 114 and the XovrE valve 116,
enables the split-cycle engine 100 to maintain pressure in the
crossover passage 112 at a high minimum pressure (typically 20 bar
absolute or higher during full load operation) during all four
strokes of its pressure/volume cycle. That is, the split-cycle
engine 100 is operable to time the XovrC valve 114 and the XovrE
valve 116 such that the XovrC and XovrE valves 114, 116 are both
open for a substantial period of time (or period of crankshaft
rotation) during which the expansion piston 120 descends from its
TDC position towards its BDC position and the compression piston
110 simultaneously ascends from its BDC position towards its TDC
position. During the period of time (or crankshaft rotation) that
the crossover valves 114, 116 are both open, a substantially equal
mass of gas is transferred (1) from the compression cylinder 102
into the crossover passage 112 and (2) from the crossover passage
112 to the expansion cylinder 104. Accordingly, during this period,
the pressure in the crossover passage is prevented from dropping
below a predetermined minimum pressure (typically 20, 30, or 40 bar
absolute during full load operation). Moreover, during a
substantial portion of the intake and exhaust strokes (typically
90% of the entire intake and exhaust strokes or greater), the XovrC
valve 114 and XovrE valve 116 are both closed to maintain the mass
of trapped gas in the crossover passage 112 at a substantially
constant level. As a result, the pressure in the crossover passage
112 is maintained at a predetermined minimum pressure during all
four strokes of the engine's pressure/volume cycle.
[0029] For purposes herein, the method of opening the XovrC 114 and
XovrE 116 valves while the expansion piston 120 is descending from
TDC and the compression piston 110 is ascending toward TDC in order
to simultaneously transfer a substantially equal mass of gas into
and out of the crossover passage 112 is referred to herein as the
"push-pull" method of gas transfer. It is the push-pull method that
enables the pressure in the crossover passage 112 of the engine 100
to be maintained at typically 20 bar or higher during all four
strokes of the engine's cycle when the engine is operating at full
load.
[0030] The crossover valves 114, 116 are actuated by a valve train
that includes one or more cams (not shown). In general, a
cam-driven mechanism includes a camshaft mechanically linked to the
crankshaft. One or more cams are mounted to the camshaft, each
having a contoured surface that controls the valve lift profile of
the valve event (i.e., the event that occurs during a valve
actuation). The XovrC valve 114 and the XovrE valve 116 each can
have its own respective cam and/or its own respective camshaft. As
the XovrC and XovrE cams rotate, eccentric portions thereof impart
motion to a rocker arm, which in turn imparts motion to the valve,
thereby lifting (opening) the valve off of its valve seat. As the
cam continues to rotate, the eccentric portion passes the rocker
arm and the valve is allowed to close.
[0031] For purposes herein, a valve event (or valve opening event)
is defined as the valve lift from its initial opening off of its
valve seat to its closing back onto its valve seat versus rotation
of the crankshaft during which the valve lift occurs. Also, for
purposes herein, the valve event rate (i.e., the valve actuation
rate) is the duration in time required for the valve event to occur
within a given engine cycle. It is important to note that a valve
event is generally only a fraction of the total duration of an
engine operating cycle (e.g., 720 degrees CA for a conventional
engine cycle and 360 degrees CA for a split-cycle engine).
[0032] The split-cycle air hybrid engine 100 also includes an air
reservoir (tank) 142, which is operatively connected to the
crossover passage 112 by an air reservoir tank valve 152.
Embodiments with two or more crossover passages 112 may include a
tank valve 152 for each crossover passage 112, which connect to a
common air reservoir 142, or alternatively each crossover passage
112 may operatively connect to separate air reservoirs 142.
[0033] The tank valve 152 is typically disposed in an air tank port
154, which extends from the crossover passage 112 to the air tank
142. The air tank port 154 is divided into a first air tank port
section 156 and a second air tank port section 158. The first air
tank port section 156 connects the air tank valve 152 to the
crossover passage 112, and the second air tank port section 158
connects the air tank valve 152 to the air tank 142. The volume of
the first air tank port section 156 includes the volume of all
additional recesses which connect the tank valve 152 to the
crossover passage 112 when the tank valve 152 is closed.
Preferably, the volume of the first air tank port section 156 is
small relative to the second air tank port section 158. More
preferably, the first air tank port section 156 is substantially
non-existent, that is, the tank valve 152 is most preferably
disposed such that it is flush against the outer wall of the
crossover passage 112.
[0034] The tank valve 152 may be any suitable valve device or
system. For example, the tank valve 152 may be a pressure-activated
check valve, or an active valve which is activated by various valve
actuation devices (e.g., pneumatic, hydraulic, cam, electric, or
the like). Additionally, the tank valve 152 may comprise a tank
valve system with two or more valves actuated with two or more
actuation devices.
[0035] The air tank 142 is utilized to store energy in the form of
compressed air and to later use that compressed air to power the
crankshaft 106. This mechanical means for storing potential energy
provides numerous potential advantages over the current state of
the art. For instance, the split-cycle air hybrid engine 100 can
potentially provide many advantages in fuel efficiency gains and
NOx emissions reduction at relatively low manufacturing and waste
disposal costs in relation to other technologies on the market,
such as diesel engines and electric-hybrid systems.
[0036] The engine 100 typically runs in a normal operating or
firing (NF) mode (also commonly called the engine firing (EF) mode)
and one or more of four basic air hybrid modes. In the EF mode, the
engine 100 functions normally as previously described in detail
herein, operating without the use of the air tank 142. In the EF
mode, the air tank valve 152 remains closed to isolate the air tank
142 from the basic split-cycle engine. In the four air hybrid
modes, the engine 100 operates with the use of the air tank
142.
[0037] The four basic air hybrid modes include:
[0038] 1) Air Expander (AE) mode, which includes using compressed
air energy from the air tank 142 without combustion;
[0039] 2) Air Compressor (AC) mode, which includes storing
compressed air energy into the air tank 142 without combustion;
[0040] 3) Air Expander and Firing (AEF) mode, which includes using
compressed air energy from the air tank 142 with combustion;
and
[0041] 4) Firing and Charging (FC) mode, which includes storing
compressed air energy into the air tank 142 with combustion.
[0042] Further details on split-cycle engines can be found in U.S.
Pat. No. 6,543,225 entitled Split Four Stroke Cycle Internal
Combustion Engine and issued on Apr. 8, 2003; and U.S. Pat. No.
6,952,923 entitled Split-Cycle Four-Stroke Engine and issued on
Oct. 11, 2005, each of which is incorporated by reference herein in
its entirety.
[0043] Further details on air hybrid engines are disclosed in U.S.
Pat. No. 7,353,786 entitled Split-Cycle Air Hybrid Engine and
issued on Apr. 8, 2008; U.S. Patent Application No. 61/365,343
entitled Split-Cycle Air Hybrid Engine and filed on Jul. 18, 2010;
and U.S. Patent Application No. 61/313,831 entitled Split-Cycle Air
Hybrid Engine and filed on Mar. 15, 2010, each of which is
incorporated by reference herein in its entirety.
SUMMARY
[0044] Devices and related methods are disclosed that generally
involve variable force valve springs for controlling the motion of
an engine valve. The force exerted by the valve spring can be
adjusted by altering the pressure at which a fluid is supplied to a
fluid chamber thereof, by altering the volume of the fluid chamber,
and/or by changing the aggregate surface area over which fluid
pressure is coupled to the engine valve. Associated fluid control
systems are also disclosed herein, as are various methods for
adjusting the force of a valve spring based on a variety of engine
parameters, such as engine speed, engine load, and/or a combination
thereof.
[0045] In one aspect of at least one embodiment of the invention,
an engine is provided that includes a pneumatic valve spring
coupled to an engine valve and a control module configured to
adjust a spring force of the pneumatic valve spring based on at
least one of engine speed and engine load.
[0046] In another aspect of at least one embodiment of the
invention, a valve spring system is provided that includes a vessel
having a piston reciprocally disposed therein, an engine valve
coupled to the piston, a first source of pressurized air having a
first pressure, and a second source of pressurized air having a
second pressure that is greater than the first pressure. The system
also includes a control valve configured to adjust a force exerted
on the piston by controlling a first fluid communication between
the vessel and the first source and a second fluid communication
between the vessel and the second source.
[0047] In another aspect of at least one embodiment of the
invention, a valve spring system is provided that includes a cap
defining a first bore therein and a piston coupled to an engine
valve, the piston being slidably disposed within the first bore
such that the piston and the cap define a first fluid chamber. The
system also includes an outer housing defining a second bore
therein, the cap being slidably disposed within the second bore
such that the cap and the outer housing define a second fluid
chamber, and a fluid port configured to selectively release
pressurized fluid from the second fluid chamber, thereby allowing
the cap to slide relative to the outer housing such that a volume
of the first fluid chamber is increased.
[0048] In another aspect of at least one embodiment of the
invention, a valve spring system is provided that includes a first
housing defining a first bore therein in which a first piston is
reciprocally disposed, the first piston being coupled to a valve
stem of an engine valve. The system also includes a second housing
defining a second bore therein in which a second piston is
reciprocally disposed, the second piston being coupled to an
extension stem. The system also includes a third housing defining a
hydraulic plenum, the hydraulic plenum being in fluid communication
with a proximal end of the valve stem, a distal end of the
extension stem, and a control valve. The plenum is in fluid
communication with an accumulator when the control valve is opened
such that the valve stem is movable independently from the
extension stem, and the plenum is sealed when the control valve is
closed such that movement of the valve stem requires movement of
the extension stem.
[0049] In another aspect of at least one embodiment of the
invention, a method of actuating a valve of an engine is provided
that includes increasing a spring force of a valve spring
associated with the valve based on an increase in an engine
parameter, and decreasing the spring force of the valve spring
based on a decrease in the engine parameter.
[0050] The present invention further provides devices, systems, and
methods as claimed.
BRIEF DESCRIPTION OF THE DRAWINGS
[0051] The invention will be more fully understood from the
following detailed description taken in conjunction with the
accompanying drawings, in which:
[0052] FIG. 1 is a schematic cross-sectional view of a prior art
split-cycle air hybrid engine;
[0053] FIG. 2A is a schematic view of one embodiment of a valve
train according to the present invention in which a valve is
closed;
[0054] FIG. 2B is a schematic view of the valve train of FIG. 2A in
which the valve is opened;
[0055] FIG. 3A is a schematic view of one embodiment of a fluid
control system according to the present invention, shown with a
spool valve in a first position;
[0056] FIG. 3B is a schematic view of the fluid control system of
FIG. 3A, shown with a spool valve in a second position;
[0057] FIG. 3C is a schematic view of the fluid control system of
FIGS. 3A and 3B, shown with a spool valve in a third position;
[0058] FIG. 4 is a schematic view of another embodiment of a fluid
control system according to the present invention;
[0059] FIG. 5A is a schematic view of one embodiment of a valve
train according to the present invention in which an
inwardly-opening valve is closed;
[0060] FIG. 5B is a schematic view of the valve train of FIG. 5A in
which the inwardly-opening valve is opened;
[0061] FIG. 6A is a schematic view of another embodiment of a valve
spring according to the present invention shown in a reduced
filling volume configuration;
[0062] FIG. 6B is a schematic view of the valve spring of FIG. 6A
shown in an increased filling volume configuration;
[0063] FIG. 7 is a schematic view of another embodiment of a valve
spring according to the present invention;
[0064] FIG. 8A is a plot of required spring pressure as a function
of engine speed for one embodiment of an engine according to the
present invention;
[0065] FIG. 8B is a plot of required spring pressure normalized to
required idle spring pressure as a function of engine speed for one
embodiment of an engine according to the present invention;
[0066] FIG. 9A is a series of plots of spring force as a function
of engine speed for various embodiments of the present invention;
and
[0067] FIG. 9B is a series of plots of spring force as a function
of engine load for various embodiments of the present
invention.
DETAILED DESCRIPTION
[0068] Certain exemplary embodiments will now be described to
provide an overall understanding of the principles of the
structure, function, manufacture, and use of the devices and
methods disclosed herein. One or more examples of these embodiments
are illustrated in the accompanying drawings. Those skilled in the
art will understand that the devices and methods specifically
described herein and illustrated in the accompanying drawings are
non-limiting exemplary embodiments and that the scope of the
present invention is defined solely by the claims. The features
illustrated or described in connection with one exemplary
embodiment may be combined with the features of other embodiments.
Such modifications and variations are intended to be included
within the scope of the present invention.
[0069] Although certain methods and devices are disclosed herein in
the context of a split-cycle engine and/or an air hybrid engine, a
person having ordinary skill in the art will appreciate that the
methods and devices disclosed herein can be used in any of a
variety of contexts, including, without limitation, non-hybrid
engines, two-stroke and four-stroke engines, conventional engines,
diesel engines, etc.
[0070] In order to operate an engine at maximum efficiency, it is
desirable to vary the force applied by the valve spring or springs
acting upon the various engine valves. For example, efficiency
gains can be realized by varying the spring force based on engine
speed. In split-cycle engines (such as the engine 100 detailed
above), the dynamic actuation of the crossover valves (i.e., 114,
116) is very demanding. This is because the crossover valves must
achieve sufficient lift to fully transfer the fuel-air charge in a
very short period of crankshaft rotation (typically in a range of
about 30 to 45 degrees CA) relative to that of a conventional
engine, which normally actuates the valves for a period of
approximately 180 degrees CA. As a result, the crossover valves are
required to actuate about four to six times faster than the valves
of a conventional engine. Thus, the valve train and valve springs
associated with these valves must be capable of relatively fast
actuation rates. As engine speed increases, the speed of the valves
likewise increases, generating higher valve acceleration and higher
valve inertia. As a result, the force of the valve spring required
to overcome the valve inertia increases with engine speed.
Efficiency can therefore be improved by using a low spring force
when the engine is operating at a low speed, and using a high
spring force when the engine is operating at a high speed. Changes
in engine speed are relatively slow, and therefore the valve spring
force can be adjusted gradually to compensate for changes in engine
speed.
[0071] Efficiency gains can also be realized by adjusting valve
spring force based on engine load. In the split-cycle engine 100
detailed above, the crossover valves 114, 116 are outwardly-opening
and therefore the valve springs associated therewith must exert
enough force to hold the crossover valves closed during at least
the peak pressure portion of the compression and expansion strokes
of the engine. Since the peak combustion pressure in the expansion
cylinder increases with engine load, the force requirement of the
XovrE valve spring likewise increases. Efficiency can therefore be
improved by using a low spring force when the engine is operating
under a relatively low load, and using a high spring force when the
engine is operating at a relatively high load. Changes in engine
load can occur almost instantaneously, for example when the
throttle is suddenly opened to a full throttle position or suddenly
closed. Accordingly, the valve spring force must likewise be
adjustable almost instantaneously in order to successfully vary
spring force based on load. A variety of adjustable valve springs
is disclosed herein, including some that are gradually adjustable,
some that are substantially instantaneously adjustable, and some
that are both gradually and substantially instantaneously
adjustable.
[0072] FIGS. 2A-2B illustrate one exemplary embodiment of a valve
train 200 that includes a variable force valve spring 208. Although
the operation of the valve train 200 is described with respect to
an outwardly-opening poppet valve, it will be appreciated that it
can be readily adapted for use with almost any engine valve,
including inwardly-opening poppet valves.
[0073] As shown in FIG. 2A, the valve train 200 generally includes
a cam 202, a rocker 204, a valve 206, and a variable force valve
spring 208. The valve train 200 also includes one or more
associated support elements which, for purposes of brevity, are not
illustrated.
[0074] The valve 206 comprises a valve head 210 and a valve stem
212 extending vertically from the valve head 210. A valve adapter
assembly 214 is disposed at the tip of the stem 212 opposite the
head 210 and is securely fixed thereto. The valve adapter assembly
214 includes an adapter portion 215 and a collet 217 disposed
therein.
[0075] The rocker 204 comprises a forked rocker pad 220 at one end,
which straddles the valve stem 212 and engages the underside of the
valve adapter assembly 214. Additionally, the rocker 204 includes a
solid rocker pad 222 at an opposing end, which slidably contacts
the cam 202. The rocker 204 pivots about a rocker shaft 224
extending therethrough. The forked rocker pad 220 of the rocker 204
contacts the valve adapter assembly 214 of the outwardly-opening
poppet valve 206 such that a downward direction of the rocker pad
222 caused by the actuation of the cam 202 translates into an
upward movement of the rocker pad 220, which in turn opens the
valve 206.
[0076] The cam 202 is a "dwell cam," which as used herein is a cam
that includes a dwell section (i.e., a section of the eccentric
portion of the cam having a constant radius) of at least 5 degrees
CA. For purposes herein, the dwell section is referred to as being
part of the eccentric portion of the cam, even though the dwell
section is concentric with the base circle portion of the cam in
the illustrated embodiment. In the illustrated embodiment, the
dwell cam 202 rotates clockwise (in the direction of the arrow Al).
As an eccentric portion 226 of the cam 202 contacts the rocker 204,
the rocker 204 rotates about the rocker shaft 224 to lift the valve
206 off of its seat 216.
[0077] As will be appreciated by those having ordinary skill in the
art, the valve train 200 can optionally include a lost-motion
system for selectively operatively disconnecting and connecting the
rocker from the cam, the rocker from the valve, etc. For example,
the rocker 204 can be mounted on a collapsible hydraulic tappet
that can be selectively drained of and filled with hydraulic liquid
to dynamically change the pivot height of the rocker 204. Other
examples of lost-motion systems are described at length in U.S.
application Ser. No. 13/359,521, entitled "LOST-MOTION VARIABLE
VALVE ACTUATION SYSTEM WITH CAM PHASER" filed on an even date
herewith, which is hereby incorporated by reference in its
entirety.
[0078] The valve spring 208 biases the valve 206 towards a closed
position and holds the valve head 210 securely against the valve
seat 216 when the valve 206 is fully closed. The valve spring 208
also provides sufficient closing force to collapse a lost-motion
valve train element when valve closing control is requested. As
shown, the valve spring 208 includes an outer housing 228 that
defines a cylindrical vessel 230 therein. A piston 232 coupled to
the engine valve 206 is reciprocally positioned within the vessel
230. The piston 232 is directly coupled to the engine valve 206
(i.e., to the valve stem 212) or is coupled thereto via one or more
intermediate elements, including for example the valve adapter
assembly 214. In one embodiment, the piston has a diameter of about
13 mm. The piston 232 includes one or more sealing features formed
on or mounted to a sidewall thereof to facilitate a fluid-tight
sealing engagement between the piston 232 and the housing 228. It
will be appreciated that the term "fluid" as used herein includes
compressible gasses, such as air, nitrogen, and the like. In this
manner, the piston 232 and housing 228 together define a
variable-volume fluid chamber 234. The housing 228 includes a fluid
port 236 for supplying or removing fluid to or from the chamber
234. In FIG. 2A, the valve 206 is shown in a closed position. When
the valve is so-positioned, the piston 232 is substantially at the
bottom of its stroke within the vessel 230 and the chamber 234 has
a maximum volume.
[0079] As shown in FIG. 2B, when the cam 202 rotates clockwise, as
a camshaft to which it is mounted is driven by rotation of the
engine's crankshaft, the eccentric portion 226 of the cam 202
imparts a downward motion to the rocker 204. This results in a
counter-clockwise rotation of the rocker 204, which in turn is
effective to lift the valve 206 off of the seat 216, thereby
opening the valve. As the valve 206 is lifted, the piston 232
coupled thereto slides upwards within the vessel 230 of the valve
spring 208, reducing the volume of the chamber 234. Any fluid
disposed in the chamber 234 is either compressed or forced out of
the chamber 234 through the fluid port 236. When the valve 206 is
fully opened as shown in FIG. 2B, the piston 232 is substantially
at the top of its stroke within the vessel 230 and the chamber 234
has a minimum volume.
[0080] FIGS. 3A-3B illustrate one embodiment of a fluid control
assembly 300 for controlling a spring force of the valve spring
208. The fluid control assembly generally includes a control valve
338, a high pressure source 340, and a pressure regulator 344 that
supplies a low pressure source 345. In the illustrated embodiment,
the control valve 338 is in the form of a high-speed,
solenoid-actuated spool valve, though a variety of valves known in
the art can also be employed without departing from the scope of
the present invention. The high pressure source 340 is coupled via
a first fluid line 341 to a first input 346 of the control valve
338. The high pressure source 340 is supplied by an air compressor
operatively coupled to the engine in which the valve spring 208 is
installed. The high pressure source 340 can also be supplied from
the crossover passage or the air hybrid reservoir of the engine in
which the valve spring 208 is installed. U.S. Patent Publication
No. 2010/0282225 entitled "AIR SUPPLY FOR COMPONENTS OF A
SPLIT-CYCLE ENGINE," filed on May 7, 2010, provides further details
on generating a high pressure source from the crossover passage of
a split-cycle engine, and is incorporated herein by reference in
its entirety. In some embodiments, the high pressure source has a
pressure of at least about 20 bar, at least about 30 bar, and/or at
least about 40 bar.
[0081] The high pressure source 340 is also fed through the
pressure regulator 344, where it is converted into a comparatively
low pressure source 345 (e.g., a second fluid line, a low pressure
reservoir, or the like) and coupled to a second input 348 of the
control valve 338. The regulator 344 can supply the low pressure
source 345 at a variety of pressures, and can optionally be
adjustable, electronically or otherwise. In some embodiments, the
low pressure source 345 can have a pressure of less than about 1
bar, less than about 5 bar, and/or less than about 20 bar. For
example, the low pressure source can have a pressure between about
1 bar and about 20 bar, between about 1 bar and about 10 bar,
between about 5 bar and about 10 bar, etc.
[0082] The control valve 338 also includes an output 352 coupled to
the fluid port 236 of the valve spring 208 shown in FIGS.
2A-2B.
[0083] In use, an electric current is selectively applied to a
solenoid disposed within a solenoid body 366 of the control valve
338 to impart a linear pushing or linear pulling motion to a spool
356. The electric current is selectively supplied from or at the
request of a control module (i.e., an engine control computer). The
control module or engine control computer includes a microprocessor
coupled to a memory and executes one or more programs for
controlling various aspects of an engine's operation, including,
without limitation, engine valve timing, valve spring rate, air
hybrid operating mode, etc. The spool 356 is sized to reciprocate
within the body 358 of the control valve 338 and to form a sealing
engagement with an interior surface thereof. The spool 356 includes
first and second end portions 360, 362 that form a seal with the
surrounding valve body 358. The first and second end portions 360,
362 are joined by a connecting element 364 which has a smaller
diameter than the first and second end portions 360, 362 such that
fluid can flow between the connecting element 364 and the interior
surface of the control valve body 358. The connecting element 364
can also include one or more apertures to facilitate fluid flow
therethrough.
[0084] In FIG. 3A, the spool 356 is illustrated in a first position
in which the first input 346 of the control valve 338 is in fluid
communication with the output 352. Accordingly, the high pressure
source 340 is in fluid communication with the fluid port 236 of the
valve spring 208. Since the second input 348 is blocked by the
spool 356, the fluid chamber 234 of the valve spring 208 is
supplied only by the high pressure source 340. In this position of
the control valve 338, the valve spring 208 can be said to have a
"high" spring force, as the valve spring 208 will bias the engine
valve 206 towards a closed position under the relatively high
pressure of the high pressure source 340. In this position, more
force and more energy is required to open the engine valve 206 and
to hold the engine valve 206 open. In addition, this position
causes the valve 206 to close faster.
[0085] Referring now to FIG. 3B, the solenoid can be actuated to
push the spool 356 downwards away from the solenoid body 366. As
the spool 356 is advanced, the first and second inputs 346, 348 are
blocked by the spool 356, thereby isolating the output 352 and
sealing the fluid chamber 234.
[0086] As shown in FIG. 3C, the spool 356 can be advanced further
downwards away from the solenoid body 366. In this position, the
second input 348 of the valve 338 is placed in fluid communication
with the output 352 while the first input 346 is blocked by the
spool 356. In this position, the fluid chamber 234 of the valve
spring 208 is charged with low pressure fluid from the pressure
regulator 344. In this position of the control valve 338, the valve
spring 208 can be said to have a "low" spring force, as the valve
spring 208 will bias the engine valve 206 towards a closed position
under the relatively low pressure of the regulated pressure source
345. In this position, considerably less force and less energy is
required to open the engine valve 206 and to hold the engine valve
206 open than would be required with the control valve 338 in the
position shown in FIG. 3A. In addition, this position causes the
valve 206 to close more slowly than when the control valve 338 is
in the position shown in FIG. 3A.
[0087] In use, the control valve 338 is selectively actuated (e.g.,
under the control of an engine computer or control module) to vary
the spring force of the valve spring 208. For example, the valve
spring 208 can be set to a "high" spring force by simply holding
the control valve 338 in the position shown in FIG. 3A. Similarly,
the valve spring 208 can be set to a "low" spring force by holding
the control valve 338 in the position shown in FIG. 3C.
Alternatively, the spring force of the valve spring 208 can be
adjusted to any of a variety of levels between the "high" and "low"
settings by manipulating the control valve 338. For example, the
pressure in the fluid chamber 234 can be maintained at a particular
level by positioning the control valve 338 as shown in FIG. 3B to
effectively isolate and seal off the fluid chamber 234. To raise
the pressure in the fluid chamber 234 slightly from the current
level, the control valve 338 can be moved briefly to the position
shown in FIG. 3A. Once the desired pressure is reached, which in
some embodiments can be indicated by a pressure sensor configured
to measure the pressure within the fluid chamber 234, the control
valve can be moved back to the position shown in FIG. 3B to
maintain the valve spring 208 at the new pressure level. Likewise,
to lower the pressure in the fluid chamber 234 slightly from the
current level, the control valve 338 can be moved briefly to the
position shown in FIG. 3C. Once the desired pressure is reached
(e.g., as indicated by an output of a pressure sensor), the control
valve 338 can be moved back to the position shown in FIG. 3B to
maintain the valve spring 208 at the new pressure level. Thus, by
momentarily moving the control valve 338 up or down from the
position shown in FIG. 3B, slight increases or decreases in valve
spring pressure can be effected. These changes can be dictated by
the engine control computer, based at least in part on a feedback
loop provided by a pressure sensor configured to measure the
pressure within the fluid chamber 234.
[0088] FIG. 4 illustrates another embodiment of a fluid control
assembly 400 for controlling the spring force of the valve spring
208. The fluid control assembly 400 generally includes a high
pressure source 440, a continuously-adjustable pressure regulator
444, and a release valve 468. The pressure regulator 444 is
configured to convert the high pressure source 440 to a
pressure-regulated output source 445 having any of a variety of
pressures in a prescribed range. The regulator 444 can be
configured to generate an output source of fluid at a pressure of
about 1 bar to about 100 bar, about 1 bar to about 85 bar, about 1
bar to about 20 bar, and/or about 10 bar to about 20 bar. The
regulator 444 can be electronically adjustable, and/or can be
hydraulically, magnetically, mechanically, and/or manually
adjustable. It will be appreciated that the regulator 444 permits
the pressure supplied to the fluid chamber 234 of the valve spring
208 to be varied across a wide range of pressures. This
advantageously permits dynamic control of the spring force of the
valve spring 208. The release valve 468 can be selectively actuated
to release pressurized fluid stored in the fluid chamber 234. For
example, the release valve 468 can be electronically actuated and
can be configured to open briefly when instructed by an engine
control computer or control module to release the pressurized
charge from the fluid chamber 234. In one embodiment, the release
valve 468 is an integral component of the regulator 444 and is
configured to release stored pressure automatically when the
regulator 444 is actuated to reduce the output pressure.
[0089] In use, the regulator 444 is selectively adjusted (e.g.,
under the control of an engine computer or control module) to vary
the spring force of the valve spring 208.
[0090] FIGS. 5A-5B illustrate another embodiment of a valve spring
according to the present invention for use with inwardly-opening
engine valves (i.e., engine valves that open into the cylinder and
towards the piston). As shown, the valve spring 508 biases the
valve 506 towards a closed position and holds the valve head 510
securely against a valve seat 516 when the valve 506 is fully
closed. The valve spring 508 also provides sufficient closing force
to collapse a lost-motion valve train element when valve closing
control is requested. The valve spring 508 includes an outer
housing 528 that defines a cylindrical vessel 530 therein. A piston
532 coupled to the engine valve 506 is positioned within the vessel
530 such that it reciprocates therein. The piston 532 can be
directly coupled to the engine valve 506 (i.e., to the valve stem
512) or can be coupled thereto via one or more intermediate
elements. The piston 532 includes one or more sealing features
formed on or mounted to a sidewall thereof to facilitate a
fluid-tight sealing engagement between the piston 532 and the
housing 528. In this manner, the piston 532 and housing 528
together define a variable-volume fluid chamber 534. The housing
528 includes a fluid port 536 for supplying or removing fluid to or
from the chamber 534. The piston 532 also includes an engagement
portion 533 extending therefrom for engaging the cam 502 or an
intermediate valve train element. In FIG. 5A, the valve 506 is
shown in a closed position. When the valve is so-positioned, the
piston 532 is substantially at the top of its stroke within the
vessel 530 and the chamber 534 has a maximum volume.
[0091] As shown in FIG. 5B, when the cam 502 rotates clockwise as a
camshaft to which it is mounted is driven by rotation of the
engine's crankshaft, an eccentric portion 526 of the cam 502
imparts a downward motion to the piston 532, which in turn is
effective to lift the valve 506 off of the seat 516. As the valve
506 is lifted, the piston 532 slides downwards within the vessel
530 of the valve spring 508, reducing the volume of the chamber
534. Any fluid disposed in the chamber 534 is either compressed or
forced out of the chamber 534 through the fluid port 536. When the
valve 506 is opened as shown in FIG. 5B, the piston 532 is
substantially at the bottom of its stroke within the vessel 530 and
the chamber 534 has a minimum volume.
[0092] It will be appreciated that any of the fluid control systems
described herein (e.g., the fluid control systems 300, 400) can be
coupled to the fluid port 536 as described above to dynamically
control the spring force of the valve spring 508 and thereby
improve the efficiency of an engine in which it is installed. For
example, the fluid control systems 300, 400 can be used to
selectively vary the pressure of a fluid charge in the chamber 534,
thereby varying the force of the valve spring 508.
[0093] FIGS. 6A-6B illustrate another embodiment of a valve spring
608 according to the present invention in which the filling volume
of the valve spring's fluid chamber can be altered to change the
peak spring force of the valve spring at peak valve lift, even
though the filling pressure (pressure to which the valve spring's
fluid chamber is filled when the valve is in its closed position on
its valve seat) is held substantially constant. The valve spring
608 biases an engine valve 606 towards a closed position and holds
the valve 606 securely against a valve seat 616 when the valve 606
is fully closed. The valve spring 608 also provides sufficient
closing force to collapse a lost-motion valve train element when
valve closing control is requested. As shown, the valve spring 608
includes a sliding cap 628 that defines a cylindrical vessel 630
therein. A piston 632 coupled to the engine valve 606 is
reciprocally positioned within the vessel 630. In addition, the
sliding cap 628 is reciprocally disposed within a cylindrical
vessel 631 formed in an outer housing 629. In this manner, the
piston 632 and sliding cap 628 together define a first fluid
chamber 634. An aperture 636 is provided in the sliding cap 628 for
supplying or removing fluid to or from the first fluid chamber 634.
In some embodiments, the fluid supplied to or removed from the
first fluid chamber 634 is a substantially compressible fluid
(e.g., air or nitrogen). The sliding cap 628 and the outer housing
629 together define a second fluid chamber 635. The outer housing
629 includes a fluid port 637 for supplying or removing fluid to or
from the second fluid chamber 635. The fluid supplied to or removed
from the second fluid chamber 635 can be a substantially
compressible fluid (e.g., air or nitrogen), a substantially
incompressible fluid (e.g., oil), or some combination thereof.
[0094] In use, the first fluid chamber 634 remains coupled to a
pressurized fluid supply via the aperture 636. In one embodiment,
the filling pressure (e.g., the pressure in the first fluid chamber
634 when the valve 606 is closed long enough to equalize the supply
pressure and the pressure in the first fluid chamber 634) is about
20 bar.
[0095] In FIG. 6A, the valve spring 608 is shown in a "high" force
configuration in which the amount of peak pressure, and therefore
peak spring force exerted by the valve spring 608 when the valve
606 is in its peak lift position (i.e., peak valve lift) to close
the engine valve 606, is greatest. In this configuration, the
second fluid chamber 635 is pressurized such that the force pushing
the sliding cap 628 down with respect to the valve 606 (e.g.,
towards the valve seat 616) exceeds the force pushing the sliding
cap 628 up with respect to the valve 606 (e.g., away from the valve
seat 616). As a result, the sliding cap 628 is moved to or held at
the bottom of its stroke within the outer housing 629 against a
mechanical stop (not shown). It will be appreciated that the
pressure in the second fluid chamber 635 need not necessarily be
greater than the pressure in the first fluid chamber 634 for this
to occur, as the surface area of the sliding cap 628 that is
exposed to the second fluid chamber 635 exceeds the surface area of
the sliding cap 628 that is exposed to the first fluid chamber 634.
Thus, if both the first and second fluid chambers 634, 635 are
pressurized to the same pressure (e.g., 20 bar), the sliding cap
628 will still move to, or be held at, the bottom of its stroke
within the outer housing 629. In this configuration, the first
fluid chamber 634 has a minimum filling volume (e.g., the volume of
the first fluid chamber 634 when the valve 606 is fully
closed).
[0096] When the valve 606 is opened, the piston 632 is forced
upwards relative to the sliding cap 628, causing the pressure in
the first fluid chamber 634 to increase. As a result, fluid in the
first fluid chamber 634 flows back through the aperture 636 and
into the fluid supply. The aperture 636 is sized to choke the flow
of fluid back into the supply such that enough fluid is maintained
in the first fluid chamber 634 to maintain the requisite closing
pressure. In one embodiment, the peak pressure in the first fluid
chamber 634 when the valve 606 is opened is 45 bar and the supply
pressure is 20 bar.
[0097] When the valve 606 is no longer held open by the valve
train, the pressure in the first fluid chamber 634 forces the
piston 632 downwards relative to the sliding cap 628, thereby
closing the valve 606. As the piston 632 slides downwards and the
volume of the first fluid chamber 634 increases, the first fluid
chamber 634 is refilled through the aperture 636 until the pressure
therein matches the supply pressure. In other words, the aperture
636 is sized such that the chamber pressure and the supply pressure
equalize during the period of cam rotation during which the valve
606 is closed.
[0098] When desired, the peak spring force exerted by the valve
spring 608 at peak valve lift can be decreased by transitioning the
valve spring 608 to the configuration shown in FIG. 6B. In FIG. 6B,
the valve spring 608 is shown in a "low" force configuration in
which the amount of peak pressure, and therefore force exerted by
the valve spring 608 at peak valve lift, to close the engine valve
606, is at a minimum. In this configuration, the pressure in the
second fluid chamber 635 is released, allowing the sliding cap 628
to slide upwards relative to the valve 606 (e.g., away from the
valve seat 616) to the top of its stroke within the outer housing
629 against a mechanical stop (not shown). The pressure in the
second fluid chamber 635 can be released in a variety of ways, for
example by switching one or more control valves such that the fluid
port 637 is disconnected from the pressurized supply and instead
connected to a lower pressure source or vented to ambient. When the
pressure in the second fluid chamber 635 is released, the sliding
cap 628 is urged upwards by the pressure in the first fluid chamber
634 and/or by an optional mechanical bias spring (not shown). With
the sliding cap 628 positioned at the top of its stroke within the
outer housing 629, the first fluid chamber 634 has a maximum
filling volume and the force exerted by the valve spring 608 at
peak valve lift is at a minimum. When it is desired to again
increase the force exerted by the valve spring 608 at peak valve
lift, the second fluid chamber 635 is re-pressurized, returning the
valve spring 608 to the configuration shown in FIG. 6A.
[0099] Since the lift of the engine valve 606 remains the same,
increasing the filling volume of the first fluid chamber 634 is
effective to decrease the ratio by which the fluid disposed therein
must be compressed when the valve is opened. In other words, even
if the supply pressure to the chamber 634 is kept constant, the
spring force at peak valve lift is decreased as the filling volume
is increased. Thus, the force exerted by the valve spring 608 at
peak valve lift is a function of the filling volume of the first
fluid chamber 634, and decreasing the filling volume (as shown in
FIG. 6A) increases the force exerted by the valve spring 608 at
peak valve lift, while increasing the filling volume (as shown in
FIG. 6B) decreases the force exerted by the valve spring 608 at
peak valve lift.
[0100] It will be appreciated that, in the embodiment of FIGS.
6A-6B, pressure can be supplied to or removed from the second fluid
chamber 635 substantially instantaneously (e.g., within one engine
cycle). Accordingly, the spring force can be altered substantially
instantaneously, which as noted above is desirable in situations
where the spring force must be altered to reflect the engine
load.
[0101] It can also be desirable, however, to adjust the spring
force gradually such that it is proportional to parameters that
also change gradually, such as engine speed. (As noted above,
engine speed generally changes gradually due to the inertia of the
flywheel and/or vehicle in which the engine is installed). Thus, in
the embodiment of FIGS. 6A-6B, the spring pressure can also be
adjusted gradually by adjusting the supply pressure supplied to the
first and second fluid chambers 634, 635 (e.g., using the fluid
control devices 300, 400 discussed above).
[0102] FIG. 7 illustrates another embodiment of a valve spring 708
according to the present invention in which the spring force can be
adjusted by increasing or decreasing the aggregate surface area
over which fluid pressure is coupled to the engine valve 706. The
valve spring 708 biases the valve 706 towards a closed position and
holds the valve head 710 securely against a valve seat 716 when the
valve 706 is fully closed. The valve spring 708 also provides
sufficient closing force to collapse a lost-motion valve train
element when valve closing control is requested. As shown, the
valve spring 708 includes first and second outer housings 728A,
728B that define respective cylindrical vessels 730A, 730B therein.
First and second pistons 732A, 732B are reciprocally positioned
within the vessels 730A, 730B, respectively, thereby defining first
and second fluid chambers 734A, 734B, respectively. The first and
second fluid chambers 734A, 734B are coupled to a pressurized
supply via first and second apertures 736A, 736B. The first piston
732A is directly coupled to the valve stem 712 of the engine valve
706 by a collet 717. The second piston 732B is coupled to an
extension stem 719.
[0103] The proximal end 721 of the valve stem 712 and the distal
end 723 of the extension stem 719 are slidably received in a third
housing 728C that defines a hydraulic plenum 725. A stop 701 is
provided in the third housing 728C to limit the travel of the
extension stem 719 in the valve closing direction. It will be
appreciated that the stop 701 can also be positioned elsewhere in
the system, such as beneath the second piston 732B. The plenum 725
is coupled via a control valve 727 to a hydraulic accumulator 715,
which is coupled via a check valve 703 to a low pressure fluid
supply 705 (e.g., the engine oil supply at a pressure of 2-4 bar).
The control valve 727 is actuated under the direction of an engine
control computer using any of a variety of actuation techniques
known in the art, including hydraulic, electrical, pneumatic,
mechanical, and/or magnetic actuation. The check valve 703 allows
one-way flow of fluid from the fluid supply 705 to the accumulator
715, for example to provide make-up for piston leakage.
[0104] In use, the spring force of the valve spring 708 can be
varied by opening and closing the control valve 727. When a reduced
spring force is required, the control valve 727 is opened, placing
the accumulator 715 in fluid communication with the plenum 725.
When the engine valve 706 is opened, the valve stem 712 slides
further into the plenum 725, forcing hydraulic fluid through the
control valve 727 and into the accumulator 715. Since the force
required to fill the accumulator 715 is less than the force
required to move the second piston 732B relative to the second
housing 728B, the extension stem 719 does not move and instead
remains at the bottom of its stroke within the plenum 725.
Accordingly, only the pressure in the first fluid chamber 734A and
the relatively small pressure supplied by the accumulator 715 act
on the engine valve 706. When the engine valve 706 is subsequently
closed and the valve stem 712 is partially withdrawn from the
plenum 725, the extension stem 719 rests against the stop 701 and
the accumulator 715 forces hydraulic fluid back through the open
control valve 727 to refill the plenum 725.
[0105] When an increased spring force is required, the control
valve 727 is closed, locking a volume of hydraulic fluid in the
plenum 725. When the engine valve 706 is opened and the valve stem
712 is advanced into the plenum 725, the relatively incompressible
hydraulic fluid can no longer escape through the control valve 727
into the accumulator 715, and instead the extension stem 719 is
forced upwards. As a result, the pressure in both the first and
second fluid chambers 734A, 734B acts on the valve 706, effectively
increasing the spring force.
[0106] Since the spring force is a function of the aggregate
surface area over which fluid pressure is coupled to the valve 706,
the degree to which the spring force is adjusted can be controlled
in a variety of ways. For example, the size of the second piston
732B and/or the second fluid chamber 734B can be varied to adjust
the degree to which the spring force changes when the second piston
732B is engaged. Alternatively, or in addition, the plenum 725 can
be coupled to a plurality of extension stems, each having its own
respective control valve and associated piston/fluid chamber
combination. In such embodiments, the number of control valves that
are opened or closed can be controlled to provide more granularity
in spring force adjustment. For example, when only a small increase
in spring force is required, a first control valve can be closed to
couple the engine valve to a single extension stem. If more spring
force is required, two or more control valves can be closed to
couple the engine valve to multiple extension stems.
[0107] FIGS. 8A and 8B illustrate valve spring pressure
requirements as a function of engine speed for an exemplary
split-cycle engine. As shown in FIG. 8A, the pressure in the valve
spring required to overcome the momentum of the valve need only be
about 1-2 bar at the 660 rpm idle speed of the engine. As engine
speed increases, however, the valve momentum increases and the
valve spring pressure required to overcome that momentum also
increases. For example, at 4000 rpm, 20 bar of valve spring
pressure is required to maintain the desired valve opening and
closing characteristics. FIG. 8B shows the spring pressure
requirement normalized to the idle spring pressure requirement. In
other words, the spring pressure requirement at 2000 rpm is 10
times the spring pressure requirement at idle, the spring pressure
requirement at 3000 rpm is 20 times the spring pressure requirement
at idle, etc.
[0108] The methods and devices disclosed herein can be used to
adjust the spring force of a valve spring based on a variety of
parameters, including engine speed, engine load, throttle position,
engine temperature, ambient temperature, ambient pressure, intake
air pressure, intake air temperature, exhaust temperature, engine
age, and the like. The spring force can also be adjusted based on a
manual setting by a user (e.g., a control switch on a vehicle's
dashboard that permits a user to select an "economy mode" or a
"sport mode" of engine operation). FIGS. 9A-9B illustrate various
exemplary embodiments of methods for varying the spring force of a
valve spring based on such parameters.
[0109] FIG. 9A illustrates spring force as a function of engine
speed when various valve spring systems disclosed herein are used
to adjust the spring force based on the speed of the engine. The
first plot 900 illustrates the spring force when the valve spring
208 of FIGS. 2A-2B is used with the fluid control system 300 of
FIGS. 3A-3C in a first mode of operation. As shown, the control
valve 338 remains in the position shown in FIG. 3C such that the
fluid chamber 234 of the valve spring 208 is in fluid communication
with the regulated low pressure source 345 until a threshold engine
speed is reached (about 3000 rpm in the illustrated embodiment).
When the threshold is reached, the control valve 338 is actuated to
switch from the low pressure source 345 to the high pressure source
340 to increase the spring force of the valve spring 208. In one
embodiment, a crank position sensor is employed to detect the speed
of the engine. The crank position sensor is coupled to an engine
control computer or control module, which in turn selectively
actuates the control valve 338 based on the input engine speed. It
will be appreciated that any engine speed between the engine's idle
speed and the engine's redline speed can be used for the threshold
speed. For example, the threshold can be about 1000 rpm, about 3000
rpm, and/or about 5000 rpm. It will also be appreciated that the
fluid control system 300 can be used to set the spring forces used
when engine speed is below or above the threshold to any of a
variety of levels intermediate to the high pressure source 340 and
the low pressure source 345, as described above.
[0110] The first plot 900 also typifies the spring force when the
valve spring 508 of FIGS. 5A-5B, the valve spring 608 of FIGS.
6A-6B, or the valve spring 708 of FIG. 7 is used. For example, when
the predetermined speed threshold is reached, the fluid port 637 of
the valve spring 608 of FIGS. 6A-6B can be coupled to a pressurized
source. This allows pressure to build in the second fluid chamber
635, thereby reducing the volume of the first fluid chamber 634 and
generating a sudden increase in spring force of the valve spring
608. Similarly, the normally-open control valve 727 of the valve
spring 708 of FIG. 7 can be closed when the speed threshold is
reached, effectively coupling the valve stem 712 to the extension
stem 719 and rapidly increasing the spring force of the valve
spring 708.
[0111] The second plot 902 illustrates the spring force when the
valve spring 208 of FIGS. 2A-2B is used with the fluid control
system 400 of FIG. 4. As shown, the regulator 444 remains set to a
minimum output pressure 445 when the engine is idling and/or
operating below a certain threshold speed (about 1200 rpm in the
illustrated embodiment). As the engine speed increases beyond that
threshold, the regulator 444 is adjusted to increase the output
pressure 445 supplied to the fluid chamber 234 of the valve spring
208, thereby increasing the spring force of the valve spring. The
output pressure 445 is increased to a degree commensurate with the
present engine speed until a second threshold is reached (about
4800 rpm in the illustrated embodiment), at which point the
regulator 444 is set to the maximum output pressure. When the
engine speed decreases, the release valve 468 is selectively
actuated to release pressure from the fluid charge in the chamber
234, while the regulator 444 is dialed down to a lower output
pressure. As described above, a crank position sensor can be
employed to detect the speed of the engine. The crank position
sensor is coupled to an engine control computer or control module,
which in turn selectively actuates the release valve 468 and/or the
regulator 444 based on the input engine speed.
[0112] The second plot 902 is also illustrative of the spring force
when the valve spring 208 of FIGS. 2A-2B is used with the fluid
control system 300 of FIGS. 3A-3C in a second mode of operation, or
when the valve spring 508 of FIGS. 5A-5B, the valve spring 608 of
FIGS. 6A-6B, or the valve spring 708 of FIG. 7 is used. For
example, the control valve 338 of the fluid control system 300 of
FIGS. 3A-3C can be manipulated to gradually increase the pressure
in the fluid chamber 234 when a speed threshold is reached. By way
of further example, when a speed threshold is reached, the supply
pressure to the first and second fluid chambers 634, 635 of the
valve spring 608 of FIGS. 6A-6B can be gradually increased to
increase the spring force. Similarly, the supply pressure to the
first and second fluid chambers 734A, 734B of the valve spring 708
of FIG. 7 can be gradually increased to increase the spring force
when a speed threshold is reached.
[0113] It will thus be appreciated that according to the methods of
FIG. 9A, the same valve spring can have a relatively low spring
force when the engine is idling or operating at low speed, yet also
have a relatively high spring force when the engine is operating at
high speeds. It is thus possible to save energy that would
otherwise be wasted compressing a stiff valve spring when the
engine is operating at a low speed, without compromising the valve
train's ability to perform at higher engine speeds. The illustrated
thresholds and ramp rate are merely exemplary, and any of a wide
range of values can be selected for these parameters without
departing from the scope of the present invention.
[0114] As shown in FIG. 9B, the principles discussed above for
varying spring force with respect to engine speed can be applied to
vary spring force based on engine load. For example, the plot 904
illustrates spring force as a function of engine load when the
valve spring 208 of FIGS. 2A-2B is used with the fluid control
system 300 of FIGS. 3A-3C in a first mode of operation. As shown,
the control valve 338 remains in the position shown in FIG. 3C such
that the fluid chamber 234 of the valve spring 208 is in fluid
communication with the regulated low pressure source 345 until a
threshold engine load is reached (about 45% of the maximum engine
load in the illustrated embodiment). When the threshold is reached,
the control valve 338 is actuated to switch from the low pressure
source 345 to the high pressure source 340 to increase the spring
force of the valve spring 208. In one embodiment, the engine load
is estimated based on readings from one or more sensors, which can
include an intake air flow sensor, a throttle position sensor, an
engine speed sensor, and/or a vehicle speed sensor. Any or all of
these sensors can be coupled to an engine control computer or
control module, which in turn selectively actuates the control
valve 338 based on the input sensor readings. It will be
appreciated that any engine load or range of engine loads in the
range of 0% to 100% can be used for the threshold engine load. The
engine load threshold can also be specified relative to the engine
load at idle. For example, the engine load threshold can be about
twice the idle load, about three times the idle load, about four
times the idle load, etc. It will also be appreciated that the
fluid control system 300 can be used to set the spring forces used
when engine load is below or above the threshold to any of a
variety of levels intermediate to the high pressure source 340 and
the low pressure source 345, as described above.
[0115] The first plot 904 also typifies the spring force when the
valve spring 508 of FIGS. 5A-5B, the valve spring 608 of FIGS.
6A-6B, or the valve spring 708 of FIG. 7 is used. For example, when
the predetermined load threshold is reached, the fluid port 637 of
the valve spring 608 of FIGS. 6A-6B can be coupled to a pressurized
source. This allows pressure to build in the second fluid chamber
635, thereby reducing the volume of the first fluid chamber 634 and
generating a sudden increase in spring force of the valve spring
608. Similarly, the normally-open control valve 727 of the valve
spring 708 of FIG. 7 can be closed when a load threshold is
reached, effectively coupling the valve stem 712 to the extension
stem 719 and rapidly increasing the spring force of the valve
spring 708.
[0116] The second plot 906 illustrates the spring force when the
valve spring 208 of FIGS. 2A-2B is used with the fluid control
system 400 of FIG. 4. As shown, the regulator 444 remains set to a
minimum output pressure when the engine is idling and/or the engine
load is below a first threshold (about 20% load in the illustrated
embodiment). As the engine load increases beyond that threshold,
the regulator 444 is adjusted to increase the output pressure
supplied to the fluid chamber 234 of the valve spring 208, thereby
increasing the spring force of the valve spring. The output
pressure is increased to a degree commensurate with the present
engine load until a second threshold is reached (about 80% engine
load in the illustrated embodiment), at which point the regulator
444 is set to the maximum output pressure. When the engine load
decreases, the release valve 468 is selectively actuated to release
pressure from the fluid charge in the chamber 234, while the
regulator 444 is dialed down to a lower output pressure. As
described above, the engine load can be estimated based on readings
from one or more sensors, which can include an intake air flow
sensor, a throttle position sensor, an engine speed sensor, and/or
a vehicle speed sensor. Any or all of these sensors can be coupled
to an engine control computer or control module, which in turn
selectively actuates the release valve 468 and/or the regulator 444
based on the input sensor readings.
[0117] The second plot 906 is also illustrative of the spring force
when the valve spring 208 of FIGS. 2A-2B is used with the fluid
control system 300 of FIGS. 3A-3C in a second mode of operation, or
when the valve spring 508 of FIGS. 5A-5B, the valve spring 608 of
FIGS. 6A-6B, or the valve spring 708 of FIG. 7 is used. For
example, the control valve 338 of the fluid control system 300 of
FIGS. 3A-3C can be manipulated to gradually increase the pressure
in the fluid chamber 234 when a load threshold is reached. By way
of further example when a load threshold is reached, the supply
pressure to the first and second fluid chambers 634, 635 of the
valve spring 608 of FIGS. 6A-6B can be gradually increased to
increase the spring force. Similarly, the supply pressure to the
first and second fluid chambers 734A, 734B of the valve spring 708
of FIG. 7 can be gradually increased to increase the spring force
when a load threshold is reached.
[0118] It will thus be appreciated that according to the methods of
FIG. 9B, the same valve spring can have a relatively low spring
force when the engine is idling or operating under a low load, yet
also have a relatively high spring force when the engine is
operating under a high load. It is thus possible to save energy
that would otherwise be wasted compressing a stiff valve spring
when the engine is operating under a low load, without compromising
the valve train's ability to perform under higher engine loads. The
illustrated thresholds and ramp rate are merely exemplary, and any
of a wide range of values can be selected for these parameters
without departing from the scope of the present invention.
[0119] The engines disclosed herein can be configured to operate
reliably over a broad range of engine speeds. In certain
embodiments, engines according to the present invention are capable
of operating at a speed of at least about 4000 rpm, and preferably
at least about 5000 rpm, and more preferably at least about 7000
rpm.
[0120] Although the invention has been described by reference to
specific embodiments, it should be understood that numerous changes
may be made within the spirit and scope of the inventive concepts
described. For example, one or both of the crossover valves can be
actuated by a cam having no dwell section or can be actuated using
a cam-less system. Also, one or both of the crossover valves can be
inwardly-opening. There can also be more than two crossover valves,
and more than one crossover passage. The intake and exhaust valves,
and any other valve in the engine for that matter, can also include
a valve spring as described herein with respect to the crossover
valves. The engines disclosed herein are not limited to having only
two cylinders. Accordingly, it is intended that the invention not
be limited to the described embodiments, but that it have the full
scope defined by the language of the following claims.
* * * * *