U.S. patent application number 11/903729 was filed with the patent office on 2008-02-28 for method and apparatus for optimized combustion in an internal combustion engine utilizing homogeneous charge compression ignition and variable valve actuation.
This patent application is currently assigned to Delphi Technologies, Inc.. Invention is credited to Philip J. G. Dingle, Mark C. Sellnau.
Application Number | 20080047509 11/903729 |
Document ID | / |
Family ID | 36638937 |
Filed Date | 2008-02-28 |
United States Patent
Application |
20080047509 |
Kind Code |
A1 |
Sellnau; Mark C. ; et
al. |
February 28, 2008 |
Method and apparatus for optimized combustion in an internal
combustion engine utilizing homogeneous charge compression ignition
and variable valve actuation
Abstract
A valvetrain system mechanization for an internal combustion
engine using compression ignition, including homogeneous charge
compression ignition, having two intake and one or more exhaust
valves per cylinder. The valves are operated by dual overhead
camshafts having two-step cams. The intake and exhaust camshafts
are provided with phasers for varying the opening and closing of
the intake and exhaust valves. A two-step roller finger follower is
disposed for each valve between the cam lobes and the valve stem.
The two sets of intake and exhaust valves are controlled by
separate oil control valves. Swirl of gases may be introduced by
mismatching the lifts of the valves. The valve opening times,
closing times, lifts, fuel injection, compression ratio, and
exhaust gas recirculation may be varied to optimize combustion
conditions for a range of engine operating modes.
Inventors: |
Sellnau; Mark C.;
(Bloomfield Hills, MI) ; G. Dingle; Philip J.;
(Rochester, MI) |
Correspondence
Address: |
DELPHI TECHNOLOGIES, INC.
M/C 480-410-202
PO BOX 5052
TROY
MI
48007
US
|
Assignee: |
Delphi Technologies, Inc.
|
Family ID: |
36638937 |
Appl. No.: |
11/903729 |
Filed: |
September 24, 2007 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
11027109 |
Dec 30, 2004 |
7308872 |
|
|
11903729 |
Sep 24, 2007 |
|
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Current U.S.
Class: |
123/90.15 ;
464/1 |
Current CPC
Class: |
Y02T 10/128 20130101;
F02D 2041/001 20130101; Y02T 10/12 20130101; F01L 1/08 20130101;
F01L 1/185 20130101; F01L 2001/186 20130101; F01L 2820/01 20130101;
F02D 41/3035 20130101; F01L 2305/00 20200501; F01L 13/0036
20130101; F02B 31/02 20130101; Y02T 10/146 20130101; F02D 2013/0292
20130101; F02D 13/0207 20130101; F01L 1/2405 20130101; F02B 1/12
20130101; F02D 13/0261 20130101; F02B 37/12 20130101; Y02T 10/18
20130101; F02D 13/0257 20130101 |
Class at
Publication: |
123/090.15 ;
464/001 |
International
Class: |
F16D 3/10 20060101
F16D003/10 |
Claims
1-23. (canceled)
24. In an internal combustion engine using compression ignition
having a plurality of cylinders, and having a first and second
intake valve and at least one exhaust valve for each cylinder, and
having a variable valve actuator for controlling said first and
second intake valves of at least one cylinder, and having an
exhaust camshaft phaser, a method for operating said engine in a
predetermined operating mode, comprising the steps of: a) adjusting
the setting of said exhaust phaser to provide a predetermined
opening and closing timing of said exhaust valves; and b) adjusting
the setting of said valve actuator for controlling said first and
second intake valves of at least one cylinder to provide a
predetermined schedule of intake activation and deactivation.
25. A method in accordance with claim 24 wherein said engine
further having means for recirculating exhaust gas from an engine
exhaust system into said cylinders and having means for varying
fuel injection timing into said cylinders, said method including
the further steps of: a. adjusting the setting of said exhaust gas
recirculation means to provide a predetermined amount of exhaust
gas recirculated into the cylinders; and b. adjusting the time and
amounts of fuel injection into the cylinders according to a
predetermined fuel injection schedule.
26. A method in accordance with claim 24 wherein said engine
includes a first and second exhaust valve for each cylinder and a
variable valve actuator controlling said first and second exhaust
valves of at least one cylinder, said method including the further
step of adjusting the setting of said variable valve actuator for
controlling activation and deactivation of said first and second
exhaust valves.
27. A method in accordance with claim 24 wherein said engine
includes an intake cam phaser, said method including the further
step of adjusting the setting of said intake cam phaser to provide
a predetermined opening and timing of said intake valves.
28. A method in accordance with claim 26 wherein said engine
includes an intake cam phaser, said method including the further
step of adjusting the setting of said intake cam phaser to provide
a predetermined opening and timing of said intake valves.
29. A method in accordance with claim 28 wherein said predetermined
operating mode is Cold Start, comprising the steps of: a) advancing
said intake phaser to provide early opening and early closing of
said intake valves and high effective compression ratio; b) setting
said first intake valves on low lift and setting said second intake
valves on zero lift to provide in-cylinder charge swirl; c)
advancing said exhaust phaser to provide early opening and early
closing of said exhaust valves to warm said engine exhaust system
and to trap exhaust gases for cylinder heating; and d) setting said
first and second exhaust valves on normal lift.
30. A method in accordance with claim 28 wherein said predetermined
operating mode is Warm Idle, comprising the steps of: a) advancing
said intake phaser and retarding said exhaust phaser to reduce
compression ratio and increase effective expansion ratio; b)
setting said first intake valves on high lift and setting said
second intake valves on zero lift to provide in-cylinder charge
swirl; and c) setting said first and second exhaust valves on
normal lift.
31. A method in accordance with claim 28 wherein said predetermined
operating mode is Low Load, comprising the steps of: a) advancing
said intake phaser and retarding said exhaust phaser to provide an
intermediate compression ratio and a high effective expansion
ratio; b) setting said first intake valves on high lift and setting
said second intake valves on zero lift to provide in-cylinder
charge swirl; and c) setting said first and second exhaust valves
on normal lift.
32. A method in accordance with claim 28 wherein said predetermined
operating mode is Medium Load, comprising the steps of: a)
retarding said intake phaser to reduce effective compression ratio;
b) setting said first intake valves on high lift and setting said
second intake valves on zero lift to provide in-cylinder charge
swirl; c) retarding said exhaust phaser to provide a high effective
expansion ration; and d) holding said first exhaust valves open
during part of the induction stroke to provide internal exhaust gas
recirculation and in-cylinder exhaust gas swirl.
33. A method in accordance with claim 32 wherein said engine
includes a means for varying fuel injection timing into said
cylinders, said method including the further step of pulsing said
fuel injection means a plurality of times beginning ahead of top
dead center of the compression stroke of said engine such that
injected fuel is substantially premixed before combustion.
34. A method in accordance with claim 28 wherein said predetermined
operating mode is Peak Torque, comprising the steps of: a)
retarding said intake phaser to reduce effective compression ratio;
b) setting said first and second intake valves on high lift to
increase in-cylinder combustion air; and c) fully advancing said
exhaust phaser to provide early exhaust valve opening.
35. A method in accordance with claim 28 wherein said predetermined
operating mode is Peak Power, comprising the steps of: a) retarding
said intake phaser to reduce effective compression ratio below that
for Peak Torque mode; b) setting said first and second intake
valves on high lift to increase in-cylinder combustion air; and c)
fully advancing said exhaust phaser to provide early exhaust valve
opening.
36. A method in accordance with claim 28 wherein said predetermined
operating mode is Acceleration Transient Operating Mode, comprising
the steps of: a) fully advancing said intake phaser to increase
in-cylinder combustion air; b) setting said first intake valves on
high lift and setting said second intake valves on zero lift to
provide in-cylinder charge swirl; c) fully advancing said exhaust
phaser to provide early exhaust valve opening to rapidly accelerate
said turbocharger; d) holding said first and second exhaust valves
open during part of the opening of said first and second intake
valves to provide internal exhaust gas recirculation; and
Description
TECHNICAL FIELD
[0001] The present invention relates to combustion in internal
combustion engines; more particularly, to a variable valve
actuation (VVA) and phasing system for a compression-ignited
engine; and most particularly, to a simple method and apparatus
incorporating variable valve actuation and camshaft phasing for
optimizing and controlling compression-ignited combustion, and
especially homogeneous charge compression ignition (HCCl) (also
known as controlled autoignition) combustion, to reduce emissions
of nitrogen oxides (NO.sub.x) and particulate matter (PM), improve
engine performance, and increase fuel efficiency over a range of
engine operating modes.
BACKGROUND OF THE INVENTION
[0002] Advanced combustion processes are being developed in the
engine combustion arts to reduce emissions of NO.sub.x and PM from
both spark-ignited (SI) and compression-ignited (CI) engines,
especially light-duty automotive engines. As used herein,
"spark-ignited" and "spark ignition" refers to any internal
combustion engine wherein ignition of a compressed combustible
mixture of fuel and air in an engine cylinder occurs principally
because of an electric discharge formed in the midst of the
compressed combustible mixture. "Compression-ignited" and
"compression ignition" refers to any internal combustion engine
wherein ignition of a compressed combustible mixture of fuel and
air occurs principally because some or all of the components of the
compressed combustible mixture have been adiabatically compressed
in a cylinder to a temperature at or above the spontaneous ignition
temperature of the mixture. Thus, as used herein,
"compression-ignited" and "compression ignition" should be taken to
mean not only conventional prior art diesel ignition, wherein fuel
is injected into a compressed air charge at substantially the top
of the compression stroke to form a non-homogeneous mixture, but
also all other compression-type ignition including but not limited
to homogeneous charge compression ignition (HCCI), controlled
autoignition (CAI), and premixed diesel (PMD) ignition.
[0003] HCCI-diesel is homogeneous charge CI using diesel fuel. It
is mixed mode in that this engine must revert to conventional
diesel combustion in some conditions. Fuel may be injected early to
foster mixing. HCCI-diesel is controlled using VVA and other means
such that the charge is diluted and combustion temperatures are
low. This produces combustion with low soot and low NOx.
[0004] HCCI-gasoline is also homogeneous charge compression
ignition but it uses gasoline. For certain conditions such as idle
and higher loads, it reverts to conventional spark ignition. The
combustion process is similar to HCCI-diesel but a lower pressure,
and a lower cost injection system is utilized.
[0005] HCCI, whether HCCI-diesel or HCCI-gasoline, is known to be
chemically-kinetically controlled. These combustion processes
require special in-cylinder conditions, operate in critical ranges,
and generally are difficult to control. If successful, however,
HCCI promises drastically reduced emissions that may satisfy future
US and European emissions standards. Avoidance of high cost and
complexity is a significant challenge for many advanced engine
concepts.
[0006] Advanced engines are foreseen and in development in the
engine arts which may use diesel fuel, gasoline, mixtures thereof,
or other specialty fuels. Fuel may be port-injected and/or
cylinder-injected to foster homogeneous charge compression ignition
(HCCI) or "controlled autoignition" (CAI) to provide controlled,
low-temperature burning of the fuel. Hybrid engines may, for
example, utilize homogeneous charge compression ignition (HCCI) in
some operating modes and spark-ignition (SI) or conventional
compression-ignition (CI) in other modes. Thus, the scope of the
present invention applies to both conventional combustion modes and
advanced premixed combustion modes for gasoline-fueled and
diesel-fueled engines. A main need in the art is a simple solution
to satisfy HCCI requirements while reducing cost and complexity of
the overall powertrain.
[0007] Because combustion initiation is chemically-kinetically
controlled, one problem of HCCI systems is that parameters other
than fuel injection timing (as for conventional diesel engines)
must be controlled in order to control combustion initiation.
Mixture compression temperature and exhaust gas recirculation (EGR)
level are two such control parameters. The Miller Cycle, known in
the prior art, with variable late intake valve closing (LIVC) and
turbo compounding can be used to control compression temperature
over the engine operating range. The Miller Cycle with LIVC
provides independent control of compression ratio (CR) and
expansion ratio (ER), with CR generally lower than ER.
[0008] Currently, light-duty diesel engines operate with fixed
geometric compression ratios of about 18:1 to about 22:1 in order
to achieve good cold-starting characteristics. However, once warmed
up, such compression ratios are excessively high and reduce
thermodynamic efficiency due to high heat losses. Such high
compression ratios contribute to higher peak cycle temperatures,
which exacerbates NO.sub.x production. As with other internal
combustion engines, the diesel engine operates within a peak
cylinder pressure constraint that is dictated by structural
strength considerations. In other words, higher CR demands a more
massive, heavy, and expensive engine than lower CR.
[0009] Thus, a further need in the CI art is an engine having a
lower CR when warmed to permit lower peak cylinder temperatures and
pressures.
[0010] Variable LIVC is useful to provide high effective
compression ratio (ECR) for good cold start characteristics, while
providing lower ECR and lower compression temperatures for
warmed-up operation. In some applications, the use of variable LIVC
can eliminate the need for prior art glow plugs (GP) for engine
starting. GP elimination has at least one important side benefit in
that real estate is freed up on the engine head for a flush-mounted
cylinder pressure transducer. Cylinder-pressure-based control may
be necessary for optimal HCCI systems and is currently under study
in the engine arts.
[0011] As noted above, EGR is an important control parameter for
HCCI. Generally, HCCI systems require relatively high levels of
exhaust gas recirculation, as high as 50% to 70%, and the
combustion process is sensitive to small changes in EGR level. EGR
can be used to control both combustion initiation and combustion
burn rates, while also lowering flame temperatures for reduced
NO.sub.x emissions. Prior art external EGR systems, wherein exhaust
gas is metered from the exhaust manifold into the intake manifold,
offer the advantage of cooling the exhaust gas for reduced
NO.sub.x, but these systems are bulky, expensive, and slow to
respond.
[0012] Thus, a still further need in the art is means for rapid
control of EGR introduction into the firing chamber for good
transient response of advanced HCCI systems.
[0013] Another requirement of HCCI systems is in-cylinder swirl of
intake gases to provide effective mixing of injected fuel and air.
Swirl can be produced in the prior art by swirl ports, but such
devices limit full-load airflow and engine power. Alternatively,
port deactivation (PDA) can provide swirl by blocking one intake
port of a two-intake-valve system with a butterfly valve, barrel
valve, or slider valve. PDA systems are used widely in the prior
art but they suffer from deposit accumulation downstream of the PDA
device and may introduce a flow loss due to shafts or other
mechanisms blocking airflow. Valve deactivation (VDA) is a
preferred alternative method that involves deactivating one of the
two intake valves. This method is advantageous because it avoids
possible deposit problems and airflow restrictions.
[0014] Thus, a still further need in the art is a simple,
inexpensive mechanism for providing valve deactivation of intake
valves.
[0015] A desirable feature for advanced HCCI combustion systems is
the ability to rapidly heat both the combustion chamber and the
exhaust gas catalyst(s) during a cold start. Rapid heating of the
combustion chamber walls can improve combustion within the first
few seconds of operation during which the catalyst is inactive.
Heating the catalyst more quickly shortens the time to catalyst
light off and thereby shortens the period of uncatalyzed
emissions.
[0016] One known method for heating the cylinder during a cold
start is early exhaust valve closing (EEVC) through which hot
burned gases are trapped in the cylinder just prior to fuel
injection. This can be combined with increased effective
compression ratio (ECR) by closing the intake valve near bottom
dead center (late intake valve closing (LIVC)) for increased
compression temperatures and good cold start characteristics.
[0017] One known method to accelerate heating the exhaust catalyst
is early exhaust valve opening (EEVO), which effectively blows down
the hot cylinder gases before expansion is complete. This can be
combined with late combustion phasing for additional exhaust
temperature increases.
[0018] Thus, a still further need in the art is a simple means for
EEVC and EEVO.
[0019] Another problem with CI combustion in general, and HCCI
combustion in particular, is the extremely low exhaust temperatures
that are typically encountered for warmed-up conditions. Exhaust
temperatures may drop below temperatures at which the catalyst is
active, for example, while the engine is idling or at partial load.
Exhaust temperatures below 150.degree. C. are known to limit
catalyst conversion efficiency.
[0020] As is known in the prior art, exhaust temperatures can be
increased without a loss of engine efficiency by reducing trapped
air mass in the cylinder. This enables lower air-fuel ratios at any
fueling level and increases exhaust temperature. Some CI engines
incorporate a throttle to reduce trapped air mass but throttling
reduces engine efficiency. Alternatively, control of trapped air
mass is possible by variable LIVC and variable charging using a
variable nozzle turbocharger (VNT). Another method to increase
exhaust temperature is cylinder deactivation (CDA), which
effectively increases load factors in the remaining firing
cylinders. Both approaches can be achieved by variable valve
actuation mechanisms.
[0021] Thus, a still further need in the art is a simple means for
providing variable valve actuation in a CI or HCCI engine.
[0022] Finally, while current diesel engines exhibit generally good
low-speed torque compared to modern Si engines, greater levels of
low-speed torque are highly valued in the industry. It is also
desirable that such low-speed torque be available on demand without
the effects of turbocharger lag and other transient effects. If the
intake plenum is charged with fresh air plus
conventionally-recirculated exhaust gas, several combustion cycles
may be necessary to purge the EGR from the system, and this can
contribute to poor or delayed torque response. EEVO, while useful
for cold starts as described above, can also be used to quickly
accelerate the turbocharger for improved torque response. To
minimize prior art external EGR delay problems, at least a portion
of the EGR can be delivered internally by VVA.
[0023] Thus, a still further need is a simple means for providing
variable valve actuation in an HCCI engine to match valve events to
combustion requirements for various operating conditions.
[0024] It is a principal object of the present invention to reduce
emissions of oxides of nitrogen and particulates from an HCCI
engine.
[0025] It is a further object of the present invention to improve
cold start characteristics of an HCCI engine.
[0026] It is a still further object of the present invention to
improve the performance characteristics of an HCCI engine.
[0027] It is a still further object of the present invention to
reduce the cost and complexity of an HCCI engine.
SUMMARY OF THE INVENTION
[0028] Briefly described, a valvetrain system mechanization for a
CI engine includes two engine intake valves and two engine exhaust
valves per cylinder. The valves are operated by dual overhead
camshafts (intake and exhaust) having "three-lobe" two-step cams
for each valve wherein a central lobe is a high-lift or low-lift
lobe and a pair of lateral lobes outboard of the central lobe are
correspondingly low-lift or high-lift lobes. The lift of the
high-lift lobe is greater than the lift of the low-lift lobes at
any cam position. A conventional two-step roller finger follower
(RFF) for varying selectively the lift of a valve between high-lift
and low-lift by selective application or withholding of
high-pressure oil to each RFF is disposed for each intake valve and
each exhaust valve between the cam lobes and the valve stem.
[0029] Two separate oil control valves (OCV) independently control
the pressurized oil supply to all the first and second intake valve
RFFs, respectively, via two separate intake oil supply galleries in
the engine head. Likewise, two additional separate oil control
valves independently control the pressurized oil supply to all the
first and second exhaust valve RFFs, respectively, via tow separate
exhaust oil supply galleries in the engine head. Thus, in-cylinder
swirl of intake and/or exhaust gases may be introduced as desired
by differentially lifting the two intake valves and/or the two
exhaust valves. In-cylinder swirl is known in the diesel art as a
useful adjunct to HCCI combustion. First intake RFFs are activated
and deactivated by a first intake controllable OCV, and second
intake RFFs are activated and deactivated by a second intake
controllable OCV. First exhaust RFFs are activated and deactivated
by a first exhaust controllable OCV, and second exhaust RFFs are
activated and deactivated by a second exhaust controllable OCV.
Thus the lift of all the intake valves may be selected between high
lift and low lift; and, independently, the lift of all the exhaust
valves may be selected between high lift and low lift. Preferably,
the low lift for one of the intake valves is zero lift, or total
valve deactivation.
[0030] Preferably, the intake camshaft and the exhaust camshaft are
each provided with a conventional camshaft phaser for controllably
varying the opening and closing points of the intake and exhaust
valves during a 720.degree. combustion cycle of the crankshaft.
[0031] A programmable engine control module (ECM) is provided by
means of which the valve opening times, valve closing times, valve
lifts, number and timing of fuel injection pulses, effective
compression ratio, and exhaust gas recirculation may be varied
independently to provide near-optimal combustion conditions for a
wide range of engine operating modes. Combinations of conditions
are provided for up to seven different engine-operating modes:
[0032] 1. Cold Start
[0033] 2. Warm Idle
[0034] 3. Low Load
[0035] 4. Medium Load
[0036] 5. Peak Torque
[0037] 6. Peak Power
[0038] 7. Acceleration Transient
BRIEF DESCRIPTION OF THE DRAWINGS
[0039] The present invention will now be described, by way of
example, with reference to the accompanying drawings, in which:
[0040] FIG. 1 is a schematic drawing of a four-cylinder
compression-ignited engine in accordance with the invention;
[0041] FIG. 2 is an elevational cross-sectional view of a two-step
variable valve actuation valve train in accordance with the
invention;
[0042] FIG. 3 is an isometric view of the roller finger follower
shown in FIG. 2, showing a central roller and outboard deactuation
followers having slider surfaces;
[0043] FIG. 3a is an isometric view of an outboard deactuation
follower like those shown in FIG. 3 but equipped with a roller;
[0044] FIG. 4 is a graph of effective compression ratio (ECR) as a
function of crankshaft angle after top dead center, for a variety
of intake valve opening durations as governed by an intake camshaft
phaser;
[0045] FIG. 5 is a graph of effective expansion ratio (EER) as a
function of crankshaft angle after top dead center, as governed by
an exhaust camshaft phaser;
[0046] FIG. 6 is a graph showing exemplary valve lift as a function
of crankshaft angle for exhaust valves and intake valves during an
exemplary two-revolution cycle of a prior art conventional diesel
engine;
[0047] FIG. 7 is a chart showing operational states of engine
components for the engine shown in FIGS. 1 through 3 for seven
different Operating Modes;
[0048] FIG. 8 is a graph showing definitions of the Operating Modes
shown in FIG. 7 as a function of engine speed and engine load;
[0049] FIGS. 9a,9b through 15a, 15b show exemplary valve lift and
timing as a function of crank position and also schematically for
an individual engine cylinder for the seven Operating Modes shown
in FIG. 7.
[0050] FIG. 16 is a chart showing six alternative engine
configurations for meeting various improved performance objectives
at differing engine costs and complexities;
[0051] FIGS. 17 through 22 are individual schematic drawings of the
six engine alternatives shown in FIG. 16; and
[0052] FIG. 23 is a graph showing exemplary valve lift profiles for
a three-step VVA system.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0053] In the following discussion, the mechanisms and strategies
for varying valve lift and valve timing for a compression-ignited
(CI) engine in accordance with the invention are presented in the
context of light duty engines which may be operated in either
conventional diesel ignition (DI) mode or homogeneous charge
compression ignition (HCCI) mode, or combinations thereof depending
upon engine speed and load. However, it should be understood that
such novel mechanisms and strategies may also be applied to medium
duty and heavy duty CI-type engines. Further, engines comprehended
by the invention may be hybrids that utilize compression ignition
under some operating conditions and spark ignition under other
operating conditions. Further, such engines are not limited to
diesel fuel and may be fueled by other fuels such as gasoline
and/or specialty fuels. The term "specialty fuels" as used herein
is defined to include blended fuels, as needed, to provided the
desired autoignition properties in the fuel for use by the range of
internal combustion engines described above.
[0054] Referring to FIGS. 1 through 4, a schematic drawing for
valvetrain system mechanization 10 for a compression ignition
engine 12 includes two engine intake valves 14, designated I1 and
I2, and two engine exhaust valves 16, designated E1 and E2, per
cylinder 18. The valves are operated by dual overhead camshafts
(intake and exhaust) 20,22 having "three-lobe" two-step cams 24 for
each valve wherein a central lobe 26 is a low-lift lobe and a pair
of lateral lobes 28 outboard of the central lobe are high-lift
lobes. A conventional two-step roller finger follower (RFF) 30 for
varying selectively the lift of a valve between a high-lift and a
lower lift, which may be zero lift, is disposed for each intake
valve and each exhaust valve between the cam lobes and the valve
stem 25. RFF 30 may have either sliding 35 or rolling 37 contact
elements with cam lobes 26,28; a currently preferred embodiment
includes rollers 39,41 in both the high-lift and low-lift elements
of RFF 30. An exemplary valve train and two-step RFF suitable for
use in method and apparatus of the invention is disclosed in U.S.
Pat. No. 6,668,779, the relevant disclosure of which is
incorporated herein by reference. Other two-step RFFs as may be
known in the prior art, having variable lift capability of either
the outboard contact elements or the central element, may also be
suitable.
[0055] The RFFs for the I1 intake valves are activated and
deactivated by a first controllable oil valve 32; the RFFs for the
I2 intake valves are activated and deactivated by a second
controllable oil valve 34; the RFFs for the E1 exhaust valves are
controlled by a third oil control valve 36; and the RFFs for the E2
exhaust valves are controlled by a fourth oil control valve 38.
Each oil control valve is connected to its respective RFFs via
galleries formed in the engine head which terminate, for example,
in hydraulic lash adjusters 33 which pivotably support the RFFs and
through which actuating oil is supplied to the RFFs. The lifts of
the I1 intake valves, I2 intake valves, E1 exhaust valves, and E2
exhaust valves may be selected independently between high lift and
low lift. Thus, there are 2.sup.4=16 possible combinations of
cam/valve actuations for the engine.
[0056] Preferably, for robust valve train operation over the engine
lifetime, a dedicated oil lubrication system 31 is provided wherein
contamination-free oil is circulated around the cylinder head and
valve train system only, while carbon-contaminated lube oil is
confined to the crankcase and remainder of the engine.
[0057] Preferably, the cam profiles for the I1 valves differ from
the cam profiles for the I2 valves to provide desired performance
combinations as described below. Similarly, the cam profiles for
the E1 valves differ from the cam profiles for the E2 valves.
Further, for some valves requiring total deactivation in some
operating modes, the low-lift lobes 26 may be provided as zero-lift
lobes. Still further, for a special case, an additional port
deactivation (PDA) valve (not shown) as is known in the prior art
may be combined with VVA to shut off the port (equivalent of
zero-lift) when required.
[0058] The length of valve opening for both high and low lift may
also be varied for each valve as desired, by appropriate grinding
of the respective cam lobe. In addition, exhaust valves may have a
prolonged opening (delayed closing), which overlaps the intake
valve opening, by means of a suitably-shaped "bump" on the
appropriate cam lobe. This may assist in providing a predetermined
exhaust gas dilution of the incoming charge.
[0059] The intake camshaft 20 and the exhaust camshaft 22 are each
provided with a conventional camshaft phaser 40,42, respectively,
which may be either a spline-type phaser or a vane-type phaser, the
latter being currently preferred, for controllably varying the
opening and closing points of the intake and exhaust valves during
a 720.degree. combustion cycle of the crankshaft. Each phaser is
controlled independently by a phaser oil control valve 44,46,
respectively, supplied by oil system 31.
[0060] An exhaust gas recirculation valve 48 is provided for
recirculating in known fashion a portion of the exhaust gas from
the engine exhaust stream 50 into the engine intake manifold 52 as
may be desired.
[0061] A fuel system 54 comprising a common fuel rail 56 or other
HCCI-compatible fuel system for supplying fuel to cylinder fuel
injectors 58 is provided in known fashion.
[0062] A programmable engine control module (ECM) 60 is provided by
means of which the valve opening times, valve closing times, valve
lifts, number and timing of fuel injection pulses, effective
compression ratio, effective expansion ratio, and exhaust gas
recirculation may be varied independently to provide near-optimal
combustion conditions for a wide range of engine operating
modes.
[0063] It should be noted that the individual engine components as
just described need not in themselves be novel. In fact, it is an
important consideration of the present invention that existing
technology may be adapted and novelly combined, thereby minimizing
the cost and complexity of providing much-improved engine
operation.
[0064] Referring to FIG. 4, the effective compression ratio (ECR)
may be varied by varying the intake valve timing by means of an
intake camshaft phaser (ICP). Such variation is shown for a variety
of valve lift durations. By combining a phaser angular range of 50
crank angle degrees (CAD) with a lift duration of between 200 and
300 CAD, compression ratios between 4 and 19 are possible.
[0065] Referring to FIG. 5, the effective expansion ratio (EER) may
be varied by varying the exhaust valve opening timing by means of
an exhaust camshaft phaser (ECP). Over a phaser angular range of 50
crank angle degrees, EER may be varied between about 13 and greater
than 18.
[0066] Referring to FIG. 6, exemplary reference opening curves for
the intake and exhaust valves as a function of crankshaft rotation
are shown for a conventional prior art diesel engine not equipped
with either variable valve actuation means or camshaft phasers.
During the discussion below, the terms "early" and "late" should be
understood to be with reference to the valve openings and closing
shown for the prior art diesel engine.
[0067] Referring to FIG. 7, the operational states of various
engine components are shown in tabular form for seven Operating
Modes of an HCCI engine as described above in accordance with the
invention. In a currently preferred embodiment, the intake and
exhaust valve cams are ground as follows to produce the desired
valve lifts, with respect to the prior art reference engine:
[0068] I1: high-lift cam, long; low-lift cam, short.
[0069] I2: high-lift cam, long; low-lift cam, zero lift.
[0070] E1: high-lift cam, long; low-lift cam, normal.
[0071] E2: high-lift cam, long; low-lift cam, normal.
[0072] Following are the Operating Modes, which define methods of
operating a CI engine in accordance with the invention; in the
valve opening curves, the crank angle degree (CAD) positions are
defined as 180.degree.=BDC, 360.degree.=TDC, and 540.degree.=BDC.
Referring to FIG. 8, the engine performance regions of the
Operating Modes are approximately defined in terms of engine speed
(RPM) and engine load (brake mean effective pressure, BMEP). Note
that in the following Modes the optimal values for valve opening
durations, valve lifts and cam lift profiles (including the exhaust
cam "bumps"), and exhaust and intake phaser advances and retards
may differ for specific engine applications and thus may each
require specific configurations determined through simulations
and/or experimentation within the capabilities of one of ordinary
skill in the art.
[0073] Mode 1, Cold Start (CS) The entire engine is at ambient
temperature. Referring to FIGS. 7, 9a, and 9b, a strategy for
starting a cold engine includes valve actuations to maximize heat
generation to rapidly warm the cylinders (and cylinder gases) and
the exhaust system including the exhaust catalyst, as follows:
[0074] A high effective compression ratio (ECR approx. 19) for
maximum compression temperature; thus, both intake valves are at
low lift and are closed near or at bottom dead center. [0075] Early
exhaust valve opening (EEVO) which effectively blows down the hot
cylinder gases before expansion is complete, providing a low
effective expansion ratio (EER) and early exhaust valve closing
(EEVC) which traps hot burned gases in the cylinder; thus, both
exhaust valves are on normal exhaust lift profiles; the exhaust cam
phaser is fully advanced, preferably to about 105.degree. and
preferably as a default position; and the intake cam phaser is
fully advanced, preferably to about 330.degree. and preferably as a
default position, such that exhaust/intake valve open overlap (OL)
is effectively zero. [0076] Conventional combustion strategy; thus,
fuel is injected conventionally, late in the compression stroke.
[0077] Intake swirl to maximize dispersal of injected fuel; thus,
the I2 valve is deactivated (zero lift lobe); I1 valve is on short
lift. [0078] No external exhaust gas recirculation (EGR) from the
engine exhaust system. In many applications, and especially in
warmer climates, the conditions of Mode 1 permit the elimination of
a conventional glow plug in the engine.
[0079] In Cold Start Mode, it may be further desirable to delay
some or all fuel injection until the piston has passed TDC at the
end of the compression stroke, resulting in reduced combustion
efficiency and providing a hot exhaust gas into the exhaust
catalyst, especially when the exhaust valves are opened early as
well. Of course, the injection must be timed to be early enough in
the power stroke that the compressed air is still hot enough to
cause ignition.
Mode 2, Warm Idle (WI) The engine is warm and stable. Referring to
FIGS. 7, 10a and 10b, a strategy for operating a warm engine at
idle includes;
[0080] A lowered ECR, about 15-16. Both exhaust valves are at
normal (low) lift and are closed near or at bottom dead center (EER
.about.18). [0081] Full expansion for best efficiency; thus, the
exhaust cam phaser is retarded. [0082] Intake swirl to maximize
dispersal of injected fuel and promote high thermodynamic
efficiency; thus, the I2 valve is deactivated (zero lift cam lobe);
I1 valve is on long lift (high-lift cam lobe). [0083] No external
EGR. Mode 3, Low Load (LL) The engine is under a low load rather
than simply idling. Referring to FIGS. 7, 11a and 11b, it is seen
that the operating conditions are substantially the same as for
Mode 2. Because the exhaust is relatively cold, the exhaust
catalyst may be warmed somewhat by optionally advancing the exhaust
phaser to advance the exhaust valve opening (EEVO) if desired with
some impact on thermal efficiency.
[0084] Mode 4, Medium Load (ML) This is a condition representative
of Federal vehicle performance testing. It is important to minimize
emissions and maximize fuel economy, which requires close control
of compression ratio, expansion ratio, exhaust gas dilution, and
premixed combustion conditions. Referring to FIGS. 7, 12a and 12b,
a strategy for operating an engine at medium load includes [0085]
Low ECR (approx. 10-11) and compression temperature control; thus
the I1 valve is on long lift and the intake valve phaser is
variably retarded such that the I1 valve does not close until after
BDC. [0086] Intake swirl to maximize dispersal of injected fuel and
promote high thermodynamic efficiency; thus the I2 valve is
deactivated (zero lift cam lobe). [0087] High internal EGR and
exhaust swirl; thus the E2 valve is on normal (low) lift and the E1
valve is on long (high) lift and the exhaust phaser is retarded.
[0088] Precise combustion temperature control; thus the exhaust
phaser is retarded. The trailing "bump" on the cam lift profile of
the E1 valve occurs after TDC such that exhaust gas is inducted
into the cylinder to dilute the intake charge during exhaust valve
opening; external EGR is also added to help control compression
(and thus ignition) temperature. [0089] High in-cylinder swirl in
both the intake and exhaust streams to provide good mixing of
injected fuel and air; thus the I2 valve is deactivated causing
intake swirl as in the first three modes, and the late closing of
the E1 valve causes synergistic exhaust swirl from re-induction of
exhaust gas into the cylinders from the exhaust manifold. [0090]
Homogeneous charge compression ignition via progressive and
repeated fuel injection during the compression stroke; thus each
fuel injector is pulsed a plurality of times, beginning relatively
early in the compression stroke to provide a smooth and progressive
combustion uniformly distributed over the entire combustion
chamber. It should be noted that the highly retarded exhaust valve
closing nearly causes the valves to collide with the ascending
piston (curve 80). To guard against catastrophic collision and
engine destruction in event of a system failure, it can be
desirable to provide room for the opened valves within the mixing
bowl in the piston, or to recess the valve seats into the head by a
distance sufficient to prevent such collision. Mode 5, Peak Torque
(PT) The engine is under maximum torque load. Referring to FIGS. 7,
13a and 13b, a strategy for operating an engine at peak torque
includes [0091] Maximum boost for maximum volumetric efficiency;
both I1 and I2 valves are on their long lift (high lift) cam
profiles with phaser positions optimized to provide full engine
breathing. [0092] The exhaust valves are opened early (EEVO) by
advancing the exhaust phaser (ECP) to minimize exhaust pumping
losses. [0093] No external EGR. [0094] Normal fuel injection.
[0095] Valve overlap is controlled by timing of the phasers. Mode
6, Peak Power (PP) Referring to FIGS. 7, 14a and 14b, Mode 6 is
similar to Mode 5 (PT), except that the intake valve phaser is
further retarded to improve breathing characteristics at maximum
rated engine speed. Mode 7, Acceleration Transient (AT). The
transient condition in which the engine is launched from very low
loads and speeds to higher power output. Referring to FIGS. 7, 15a
and 15b, a strategy for rapidly accelerating an engine includes
[0096] Increasing the trapped charge mass; thus the I1 valve is on
its long profile and the intake phaser is advanced such that the
ECR is approx. 15-16. [0097] Swirl for improved combustion; thus I2
is deactivated. [0098] Accelerating the turbocharger; thus both E1
and E2 are on their respective long profiles and the exhaust phaser
is advanced rapidly and momentarily for EEVO. [0099] Fast reduction
of exhaust gas in the intake manifold; thus, the timing of the
phasers is controlled such that the advance in the intake phaser
lags behind the advance in the exhaust phaser. [0100] Combustion
may be either normal or premixed.
[0101] The above methods and benefits are predicated on an engine
fully equipped as shown in FIGS. 1 through 3. However, significant
(albeit reduced) benefits can be obtained from simpler, less
expensive alternative engine configurations having reduced hardware
requirements, and thus reduced cost and complexity.
[0102] Referring to FIG. 16, six alternative engine configurations
are shown, the first five being simpler than engine 12, and their
eliminated hardware and consequent lost functionalities are
summarized. The alternative configurations are shown and discussed
in more detail in FIGS. 17 through 22.
[0103] Referring to FIG. 17, in a first alternative system 100, the
intake valve system is unchanged from the parent configuration in
FIG. 1, but the exhaust valve system employs a single oil control
valve for all the exhaust valves, thus eliminating one oil control
valve and a separate oil gallery on the exhaust side. The
functional cost is that exhaust swirl capability is lost under high
levels of internal EGR, as benefits Mode 4, thus reducing
combustion efficiency.
[0104] Referring to FIG. 18, in a second alternative system 200,
two-step functionality is provided for only the intake valves by
eliminating two (exhaust) oil control valves, all of the exhaust
two-step RFFs, and an additional oil gallery on the exhaust side.
The functional cost is loss of a high level of internal EGR
produced in the parent configuration by re-induction of exhaust gas
from the exhaust system, and the concomitant loss of high levels of
exhaust swirl. The net effect is lower overall swirl levels,
reducing the efficiency of premixed combustion modes.
[0105] Referring to FIG. 19, in a third alternative system 300, the
intake valves are functionally paired through a single oil control
valve, as are the exhaust valves, for each cylinder. Two oil
control valves and two oil galleries are eliminated. Two-step
functionality is provided for all the intake valves and all the
exhaust valves. Significant loss of in-cylinder swirl can be
partially restored by providing individual port deactivation (PDA)
as in the prior art, at some additional cost and complexity.
[0106] Referring to FIG. 20, in a fourth alternative system 400 for
diesel operation with conventional combustion, all two-step
functionality is removed for the exhaust valves, and the intake
valves are functionally paired through a single oil control valve
as in alternative system 300. Three oil control valves, eight
two-step RFFs, and three oil galleries are eliminated. Lost is
capability for high internal EGR and exhaust swirl.
[0107] Referring to FIG. 21, in a fifth alternative system 500,
cylinder deactivation is provided for half the cylinders in the
engine, and phasing is provided for only the exhaust camshaft. This
arrangement eliminates two oil control valves, eight two-step RFFs,
two oil galleries, and the intake camshaft phaser and its oil
control valve. The lost functionalities include the cold start
compression ratio benefit; variable late intake valve closing for
effective compression ratio control; capability for high internal
EGR; valve deactivation for intake swirl, requiring a port
deactivation valve; and exhaust swirl capability for high levels of
internal exhaust gas recirculation.
[0108] Referring to FIGS. 22, 23, in a sixth alternative system
600, a three-step VVA mechanism provides a still greater range of
lift combinations and valve timing. Three-step variable valve
actuation of a gasoline-fueled spark-ignited engine is disclosed,
for example, in U.S. Pat. No. 6,810,844, the relevant disclosure of
which is incorporated herein by reference. Intake oil control valve
32a is provided with a second outlet for supplying a second oil
feed via a second intake valve oil gallery 602, as required for
operation of the three-step RFFs. Similarly, exhaust oil control
valve 36a is provided with a second outlet for supplying a second
oil feed via a second exhaust valve oil gallery 604.
[0109] The three separate lifts shown in FIG. 23 have substantially
the same opening timing but variable valve closing; however, it is
obvious that, by appropriate grinding of the three independent cam
lobes providing the three independent lifts, the openings and
closings and amplitudes of the lifts may be configured
independently to satisfy a wide variety of CI engine operational
requirements.
[0110] In addition, two-step functionality plus variable cylinder
deactivation can be provided by use of three-step valvetrain
mechanism 600 wherein the low-lift profile is zero lift (base
circle) for cylinders intended for cylinder deactivation (e.g.,
cylinders number 2 and 3). When those cylinders are deactivated,
the resulting higher load factors in the active cylinders result in
desirably higher exhaust temperatures. The three-step VVA mechanism
provides equal lift to pairs of the intake valves, so that intake
swirl is not provided; however, individual port deactivation can be
used to generate intake swirl control.
[0111] The two-step VVA strategies and methods described above in
accordance with the invention, when combined with a turbocharged
Miller Cycle in a compression-ignited engine, and especially in a
homogenous charge compression-ignited engine: [0112] 1. Provide
control of compression temperature and charge dilution over wide
ranges as required by compression ignition. [0113] 2. Enable fast
control of the more critical HCCI process (i.e., internal
residuals, compression ratio, expansion ratio). [0114] 3. Improve
cold starting and may enable elimination of glow plugs. [0115] 4.
Improve full-load output of the engine across the engine speed
range, especially at low speeds. [0116] 5. Improve transient
response of the powertrain by accelerating the turbocharger. [0117]
6. Permit cylinder deactivation to also be performed in addition to
the above-described functions when three-step RFFs and three-lobe
cams are employed. Taken together, these strategies, methods, and
apparati reduce the NO.sub.x and particulate emissions of advanced
diesel engines while greatly improving engine output and transient
response.
[0118] While the invention has been described by reference to
various specific embodiments, it should be understood that numerous
changes may be made within the spirit and scope of the inventive
concepts described. Accordingly, it is intended that the invention
not be limited to the described embodiments, but will have full
scope defined by the language of the following claims.
* * * * *