U.S. patent application number 09/799715 was filed with the patent office on 2001-08-30 for premixed charge compression ignition engine with optimal combustion control.
Invention is credited to Akinyemi, Omowoleola C., Durrett, Russ P., Flynn, Patrick F., Hunter, Gary L., Moore, Greg A., Mudd, Jackie M., Muntean, George G., Wagner, Julie A., Wright, John F., Zur Loye, Axel O..
Application Number | 20010017127 09/799715 |
Document ID | / |
Family ID | 25437267 |
Filed Date | 2001-08-30 |
United States Patent
Application |
20010017127 |
Kind Code |
A1 |
Flynn, Patrick F. ; et
al. |
August 30, 2001 |
Premixed charge compression ignition engine with optimal combustion
control
Abstract
A premixed charge compression ignition engine, and a control
system, is provided which effectively initiates combustion by
compression ignition and maintains stable combustion while
achieving extremely low nitrous oxide emissions, good overall
efficiency and acceptable combustion noise and cylinder pressures.
The present engine and control system effectively controls the
combustion history, that is, the time at which combustion occurs,
the rate of combustion, the duration of combustion and/or the
completeness of combustion, by controlling the operation of certain
control variables providing temperature control, pressure control,
control of the mixture's autoignition properties and equivalence
ratio control. The combustion control system provides active
feedback control of the combustion event and includes a sensor,
e.g. pressure sensor, for detecting an engine operating condition
indicative of the combustion history, e.g. the start of combustion,
and generating an associated engine operating condition signal. A
processor receives the signal and generates control signals based
on the engine operating condition signal for controlling various
engine components to control the temperature, pressure, equivalence
ratio and/or autoignition properties so as to variably control the
combustion history of future combustion events to achieve stable,
low emission combustion in each cylinder and combustion balancing
between the cylinders.
Inventors: |
Flynn, Patrick F.;
(Columbus, IN) ; Hunter, Gary L.; (Columbus,
IN) ; Zur Loye, Axel O.; (Columbus, IN) ;
Akinyemi, Omowoleola C.; (Columbus, IN) ; Durrett,
Russ P.; (Columbus, IN) ; Moore, Greg A.;
(Grammer, IN) ; Mudd, Jackie M.; (Columbus,
IN) ; Muntean, George G.; (Columbus, IN) ;
Wagner, Julie A.; (Columbus, IN) ; Wright, John
F.; (Columbus, IN) |
Correspondence
Address: |
NIXON PEABODY, LLP
8180 GREENSBORO DRIVE
SUITE 800
MCLEAN
VA
22102
US
|
Family ID: |
25437267 |
Appl. No.: |
09/799715 |
Filed: |
March 7, 2001 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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09799715 |
Mar 7, 2001 |
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09456382 |
Dec 8, 1999 |
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6230683 |
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09456382 |
Dec 8, 1999 |
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08916437 |
Aug 22, 1997 |
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60024515 |
Aug 23, 1996 |
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Current U.S.
Class: |
123/435 |
Current CPC
Class: |
F02B 19/14 20130101;
F02D 19/0665 20130101; F02D 19/081 20130101; Y02T 10/36 20130101;
F02D 13/0215 20130101; F02D 41/0025 20130101; F02B 2075/025
20130101; Y02T 10/12 20130101; F02D 13/028 20130101; F02B 2275/32
20130101; F02D 41/0087 20130101; F02M 26/33 20160201; F02P 19/00
20130101; F02B 37/18 20130101; Y02T 10/125 20130101; F02B 1/12
20130101; F02D 13/0269 20130101; Y02T 10/42 20130101; F02D 13/0234
20130101; F02M 26/08 20160201; F02D 35/023 20130101; Y02T 10/40
20130101; F02M 26/07 20160201; F02M 31/02 20130101; Y02T 10/142
20130101; F02M 26/04 20160201; F02M 31/20 20130101; Y02T 10/126
20130101; F02B 29/0406 20130101; F02D 2013/0292 20130101; Y02T
10/30 20130101; F02D 15/04 20130101; F02D 41/3029 20130101; F02D
41/3041 20130101; F02D 13/0203 20130101; F02D 13/0265 20130101;
F02M 26/43 20160201; F02B 3/06 20130101; F02D 35/028 20130101; F02D
19/0605 20130101; F02D 41/3035 20130101; F02B 37/04 20130101; F02D
41/38 20130101; F02D 19/0649 20130101; Y02T 10/128 20130101; F02M
26/28 20160201; F02D 41/3076 20130101; F02M 26/23 20160201; F02M
26/01 20160201 |
Class at
Publication: |
123/435 |
International
Class: |
F02D 041/14; F02M
007/00 |
Claims
We claim:
1. A premixed charge compression ignition internal combustion
engine, comprising: an engine body; a combustion chamber formed in
the engine body; an intake air system for delivering intake air,
including at least one of air and a mixture of air and fuel, to
said combustion chamber; combustion history control system for
controlling a combustion history of future combustion events to
reduce emissions and optimize efficiency, said combustion history
control system including a mixture autoignition property control
system for varying an autoignition property of the mixture, said
mixture autoignition property control system including a first fuel
supply for supplying a first fuel to the engine and a second fuel
supply connected to at least one of said intake air system and said
combustion chamber for supplying a second fuel to the engine, said
first fuel having a first autoignition property and said second
fuel having a second autoignition property different from said
first autoignition property; and a processor adapted to control an
amount of said second fuel delivered to said at least one of said
intake air system and said combustion chamber to optimize engine
operation.
Description
[0001] This is a divisional application of pending application Ser.
No. 09/456,382, filed Dec. 8, 1999, which is a divisional of
application Ser. No. 08/916,437, filed Aug. 22, 1997.
TECHNICAL FIELD
[0002] This invention relates generally to a compression ignition
engine arranged to internally burn a premixed charge of fuel and
air using autoignition to achieve reduced emissions while
maintaining the desired fuel economy.
BACKGROUND OF THE INVENTION
[0003] For well over 75 years the internal combustion engine has
been mankind's primary source of motive power. It would be
difficult to overstate its importance or the engineering effort
expended in seeking its perfection. So mature and well understood
is the art of internal combustion engine design that most so called
"new" engine designs are merely designs made up of choices among a
variety of known alternatives. For example, an improved output
torque curve can easily be achieved by sacrificing engine fuel
economy. Emissions abatement or improved reliability can also be
achieved with an increase in cost. Still other objectives can be
achieved such as increased power and reduced size and/or weight but
normally at a sacrifice of both fuel efficiency and low cost.
[0004] The challenge to contemporary designers has been
significantly increased by the need to respond to governmentally
mandated emissions abatement standards while maintaining or
improving fuel efficiency. In view of the mature nature of engine
design, it is extremely difficult to extract both improved engine
performance and emissions abatement from further innovations of the
basic engine designs commercially available today. Yet the need for
such innovations has never been greater in view of the series of
escalating emissions standards mandated for the future by the
United States government and other countries. Attempts to meet
these standards includes some designers looking for a completely
new engine design.
[0005] Traditionally, there have been two primary forms of
reciprocating piston or rotary internal combustion engines: diesel
and spark ignition engines. While these engine types have similar
architecture and mechanical workings, each has distinct operating
properties which are vastly different from each other. Diesel and
spark ignited engines effectively control the start of combustion
(SOC) using simple, yet distinct means. The diesel engine controls
the SOC by the timing of fuel injection. In a spark ignited engine,
the SOC is controlled by the spark timing. As a result, there are
important differences in the advantages and disadvantages of diesel
and spark-ignited engines. The major advantage that a spark-ignited
natural gas, or gasoline, engine has over a diesel engine is the
ability to achieve extremely low NOx and particulate emissions
levels. The major advantage that diesel engines have over premixed
charge spark ignited engines (such as passenger car gasoline
engines and lean burn natural gas engines) is higher thermal
efficiency. One key reason for the higher efficiency of diesel
engines is the ability to use higher compression ratios than
premixed charge spark ignited engines (the compression ratio in
premixed charge spark ignited engines has to be kept relatively low
to avoid knock). A second key reason for the higher efficiency of
diesel engines lies in the ability to control the diesel engine's
power output without a throttle. This eliminates the throttling
losses of premixed charge spark ignited engines and results in
significantly higher efficiency at part load for diesel engines.
Typical diesel engines, however, cannot achieve the very low NOx
and particulate emissions levels which are possible with premixed
charge spark ignited engines. Due to the mixing controlled nature
of diesel combustion a large fraction of the fuel exists at a very
fuel rich equivalence ratio which is known to lead to particulate
emissions. Premixed charge spark ignited engines, on the other
hand, have nearly homogeneous air fuel mixtures which tend to be
either lean or close to stoichiometric, resulting in very low
particulate emissions. A second consideration is that the mixing
controlled combustion in diesel engines occurs when the fuel and
air exist at a near stoichiometric equivalence ratio which leads to
high temperatures. The high temperatures, in turn, cause high NOx
emissions. Lean burn premixed charge spark ignited engines, on the
other hand, burn their fuel at much leaner equivalence ratios which
results in significantly lower temperatures leading to much lower
NOx emissions. Stoichiometric premixed charge spark ignited
engines, on the other hand, have high NOx emissions due to the high
flame temperatures resulting from stoichiometric combustion.
However, the virtually oxygen free exhaust allows the NOx emissions
to be reduced to very low levels with a three-way catalyst.
[0006] Relatively recently, some engine designers have directed
their efforts to another type of engine which utilizes premixed
charge compression ignition (PCCI) or homogeneous charge
compression ignition (HCCI), hereinafter collectively referred to
as PCCI. Engines operating on PCCI principles rely on autoignition
of a relatively well premixed fuel/air mixture to initiate
combustion. Importantly, the fuel and air are mixed, in the intake
port or the cylinder, long before ignition occurs. The extent of
the mixture may be varied depending on the combustion
characteristics desired. Some engines are designed and/or operated
to ensure the fuel and air are mixed into a homogeneous, or nearly
homogeneous, state. Also, an engine may be specifically designed
and/or operated to create a somewhat less homogeneous charge having
a small degree of stratification. In both instances, the mixture
exists in a premixed state well before ignition occurs and is
compressed until the mixture autoignites. Importantly, PCCI
combustion is characterized in that: 1) the vast majority of the
fuel is sufficiently premixed with the air to form a combustible
mixture throughout the charge by the time of ignition and
throughout combustion; and 2) combustion is initiated by
compression ignition. Unlike a diesel engine, the timing of the
fuel delivery, for example the timing of injection, in a PCCI
engine does not strongly affect the timing of ignition. The early
delivery of fuel in a PCCI engine results in a premixed charge
which is very well mixed, and preferably nearly homogeneous, thus
reducing emissions, unlike the stratified charge combustion of a
diesel which generates higher emissions. Preferably, PCCI
combustion is characterized in that most of the mixture is
significantly leaner than stoichiometric to advantageously reduce
emissions, unlike the typical diesel engine cycle in which a large
portion, or all, of the mixture exists in a rich state during
combustion.
[0007] An engine operating on PCCI combustion principles has the
potential for providing the excellent fuel economy of the diesel
engine while providing NOx and particulate emissions levels that
are much lower than that of current spark-ignited or diesel engine.
For example, U.S. Pat. No. 4,768,481 to Wood discloses a process
and engine that is intended to use a homogeneous mixture of fuel
and air which is spontaneously ignited. A controlled rate of
combustion is said to be obtained by adding exhaust products to the
air-fuel mixture. A combustion chamber is connected to the engine
cylinder and fuel gas is supplied to the chamber via a check valve.
A glow plug is positioned between the combustion chamber and the
cylinder. The mixture entering the combustion is heated by the glow
plug and by the hot walls of the combustion chamber. The mixture
ignites due to the increase in temperature and the increase in
pressure resulting from compression. The Wood patent is
specifically directed to a two-stroke engine, but generally
mentions that the technology could be applied to a four-stroke
engine. However, this reference fails to discuss how the exhaust
gas recirculation and glow plug would be controlled to optimize the
start of combustion and to maintain the optimal start, and
duration, of combustion, as load and ambient conditions change. A
practical embodiment of this engine is unlikely to be capable of
effectively controlling and maintaining PCCI combustion without
additional controls.
[0008] U.S. Pat. No. 5,535,716 issued to Sato et al., discloses a
compression ignition type engine which greatly reduces NOx
emissions by introducing an evaporated fuel/air mixture into the
combustion chamber during the intake event and early in the
compression event for self-ignited combustion later in the
compression event. The amount of NOx emissions produced by this
engine is about one-thirtieth of that produced by a diesel engine.
These principles are also set forth in SAE Technical Paper No.
960081, Aoyama, T. et al., "An Experimental Study on
Premixed-Charge Compression Ignition Gasoline Engine", Feb. 26,
1996. However, these references do not specifically discuss
controlling the timing of the start of combustion and the rate of
combustion. Moreover, the engine disclosed in these references only
uses the heat generated by compression to ignite the charge,
without the use of any preheating. Also, these references do not
suggest the controls, nor the manner of operating the controls,
necessary to maintain stable combustion. Also, these references
only disclose the use of gasoline.
[0009] U.S. Pat. No. 5,467,757 issued to Yanagihara et al.,
discloses a direct injection compression-ignition type engine in
which fuel is injected into a combustion chamber during the intake
stroke or compression stroke, before 60 degrees BTDC of the
compression stroke, so as to reduce the amount of soot and NOx
generated to substantially zero. These advantages are achieved by
considerably enlarging the mean particle size of the injected fuel
from the mean particle size used in conventional combustion
processes to prevent the early vaporization of injected fuel after
injection and by making the injection timing considerably earlier
than conventional injection timing to ensure a uniformed fusion of
the injected fuel in the combustion chamber. However, this
reference nowhere suggests a manner of actively controlling the
combustion history, such as the timing of the start of combustion
and/or the duration of combustion.
[0010] Researchers have used various other names to refer to PCCI
combustion. For example, Onishi, et al. (SAE Technical Paper No.
790501, Feb. 26-Mar. 2, 1979) called it "ATAC", which stands for
"Active Thermo-Atmosphere Combustion." Noguchi, et al. (SAE
Technical Paper No. 790840, Sep. 10-13, 1979) called it "TS", which
stands for "Toyota-Soken", and Najt, et al. (SAE Paper No. 830264,
1983) called it "CIHC", which stands for "compression-ignited
homogeneous charge."
[0011] Onishi, et al., worked with two-stroke engines. They found
that PCCI combustion (ATAC) could be made to occur in a two-stroke
engine at low load over a wide speed range. Combustion stability
was much better than in the standard engine and there were
significant improvements in fuel economy and exhaust emissions.
Schlieren photography of the combustion was carried out with
results quite similar to those obtained in their combustion
studies. It was found that combustion was initiated at many points
in the combustion chamber. However, there were small time
differences between the start of combustion of these many points.
Also, the combustion reactions were found to require a relatively
long time compared to conventional spark-ignited flame propagation.
To attain PCCI combustion, the following conditions were found to
be important. The quantity of mixture and the air/fuel ratio
supplied to the cylinder must be uniform from cycle to cycle. The
scavenging "directivity" and velocity must have cyclic regularity
to ensure the correct condition of the residual gases remaining in
the cylinder. The temperature of the combustion chamber walls must
be suitable. The scavenging passage inlet must be located at the
bottom of the crankcase. It was found that at very light loads,
PCCI was not successful because charge temperatures were too low.
At very high loads, PCCI was not successful because the residual
gas quantity was too low. In between these regions, PCCI combustion
was successful.
[0012] Noguchi also obtained PCCI combustion in a two-stroke
engine. Very stable combustion was observed, with low emissions of
hydrocarbons (HC) and improved fuel consumption. Operation in PCCI
mode was possible between 800 and 3200 rpm and air/fuel ratios
between 11 and 22. Delivery ratios of up to 0.5 could be achieved
at idle conditions. They observed that combustion could start at
lower temperatures and pressures than those required for
conventional diesel combustion. The combustion behavior was
different from that of conventional spark-ignited combustion.
Ignition occurred at numerous points around the center of the
combustion chamber and the flame spread rapidly in all directions.
The combustion duration was shorter than that of conventional
combustion. It was proven that ignition kernels were not generated
from contaminants deposited on the combustion chamber walls
(generally presumed to be the cause of "run-on" phenomena in
conventional gasoline engines). To gain a better understanding of
the combustion, they set up an experimental apparatus for detecting
radicals in the combustion chamber. It was found that the radicals
showed higher peaks of luminous intensity that disappeared at an
earlier time than with conventional spark-ignited combustion. In
the case of conventional spark-ignition combustion, all the
radicals such as OH, CH, C.sub.2, H, and CHO, HO.sub.2, O were
observed at almost the same crank angle. However, with PCCI
combustion, CHO, HO.sub.2 and O radicals were detected first,
followed by HC, C.sub.2, and H radicals, and finally the OH
radical.
[0013] Najt, et al. were able to achieve PCCI combustion in a
four-stroke engine. They used a CFR single-cylinder engine with a
shrouded intake valve. Several compression ratios were tried, and
it was found that, although higher ratios would allow combustion at
lower charge gas temperatures, they also resulted in excessively
fast heat release rates. While a compression ratio of 7.5:1 was
satisfactory, a compression ratio of 10:1 was not. Intake
temperatures were in the range of 480.degree. K to 800.degree. K.
Their average energy release rates were considerably higher than
those measured by Onishi and Noguchi.
[0014] SAE Paper No. 960742, entitled "Improving the Exhaust
Emissions of Two-Stroke Engines by Applying the Activated Radical
Combustion", Ishibashi, Y. et al., 1996, is noted as disclosing yet
another study of PCCI combustion in a two-stroke engine.
[0015] Although Onishi et al., Noguchi et al., Najt et al. and
Ishibashi, et al. have made significant progress in understanding
PCCI combustion, these references fail to suggest a practical PCCI
engine having a control system capable of maintaining stable,
efficient PCCI combustion with low emissions by controlling the
time at which combustion occurs, the duration of combustion, the
rate of combustion and/or the completeness of combustion.
Specifically, these references do not suggest a PCCI engine and
control system capable of effectively controlling the start of
combustion. Moreover, these references do not suggest a system
capable of actively enhancing the engine startability and achieving
combustion balancing between the cylinders in a multi-cylinder
engine.
[0016] SAE Technical Paper No. 892068, entitled "Homogeneous-Charge
Compression Ignition (HCCI) Engines", Thring, R., Sep. 25, 1989,
investigated PCCI operation of a four stroke engine. The paper
found that PCCI required high exhaust gas recirculation (EGR) rates
and high intake temperatures. It was shown that PCCI combustion
produces fuel economy results comparable to a direct injection
diesel engine and, that under favorable conditions, i.e.
equivalence ratio of 0.5 and EGR rate of 23%, produces very low
cyclic irregularity. This study also concluded that before PCCI can
be made practical, it will be necessary to operate an engine in the
PCCI mode without the need to supply large amounts of heat energy
to the intake. The paper suggests two possibilities: the use of
heated surfaces in the combustion chamber and the use of
multi-stage turbocharging without intercoolers. However, although
this paper suggests further investigating the effects of EGR and
intake temperature on the timing of the start of combustion, this
paper fails to disclose a system for effectively achieving active
control of the start and duration of combustion.
[0017] U.S. Pat. No. 5,476,072 to Inventor discloses another
example of a PCCI engine which includes a cylinder head design that
prevents excessive stresses and structural damage that PCCI engines
inherently tend to cause. Specifically, the head includes a movable
accumulator piston which moves to limit the peak cylinder pressure
and temperature. However, control over the movement of the piston
is merely passive and, therefore, this engine is unlikely to
effectively stabilize combustion. Moreover, this reference nowhere
suggests controlling the timing at which rapid combustion occurs,
nor how such control could be accomplished.
[0018] An October 1951 publication entitled "Operating
directions--LOHMANN BICYCLE MOTOR" discloses a two-stroke engine
operating on PCCI combustion principles. Compression ratio is
continuously adjustable based on outside temperature, fuel, speed
and load. However, this engine requires the operator control the
compression ratio manually. Therefore, this engine could not
provide effective active control of combustion to ensure efficient
combustion with low emissions throughout all operating conditions.
Also, manual adjustment of compression ratio alone, without
automatic temperature, equivalence ratio and/or autoignition
property control, will not result in stable, optimized combustion
throughout all operating conditions.
[0019] Conventional "dual fuel" engines operate on both a gaseous
fuel mixture and diesel fuel. However, conventional dual fuel
engines utilize the timing of the injection of diesel fuel to
control the SOC of the fuel/air mixture received from the intake
duct. In order to achieve this result, dual fuel engines inject the
diesel fuel at approximately top dead center. In addition, the
quantity of diesel fuel injected in a dual fuel engine is
sufficient to ensure that the gaseous fuel in the combustion
chamber ignites and burns virtually completely. As a result, dual
fuel engines produce emissions similar to most conventional diesel
and natural gas engines. In particular, in known dual fuel engines
using diesel fuel and natural gas at high load, only a small amount
of diesel fuel is required to start ignition and the emissions
produced would be similar to a spark ignited natural gas engine.
Under other conditions when substantial diesel fuel is injected,
the emissions produced would be similar to a conventional diesel
engine.
[0020] Consequently, there is a need for an engine operating on
PCCI principles which includes a combustion control system capable
of effectively controlling the timing of the start of combustion or
location of combustion, and the rate or duration of combustion
during engine operation.
SUMMARY OF THE INVENTION
[0021] A general objective of the subject invention is to overcome
the deficiencies of the prior art by providing a practical PCCI
engine and a control system for effectively and efficiently
operating the PCCI engine.
[0022] Another object of the present invention is to provide a PCCI
engine and control scheme for controlling the engine in a manner to
optimally minimize emissions, especially oxides of nitrogen and
particulate emissions, while maximizing efficiency.
[0023] Yet another object of the present invention is to provide a
PCCI engine and control system for optimally controlling the
combustion history of subsequent combustion events to effectively
control the combustion event.
[0024] Still another object of the present invention is to provide
a PCCI engine and control system for effectively controlling PCCI
combustion in such a manner to achieve acceptable cylinder pressure
while minimizing combustion noise.
[0025] A further object of the present invention is to provide a
PCCI engine and control system which operates to actively control
the combustion history of future combustion events during engine
operation by sensing an engine operating condition indicative of
the combustion history.
[0026] A still further object of the present invention is to
provide a PCCI engine and control system which effectively controls
various engine operating control variables to control the time at
which the combustion event occurs during the compression and
expansion events of the engine.
[0027] Yet another object of the present invention is to provide a
PCCI engine and control system which effectively ensures that
combustion occurs at an appropriate crank angle during the engine
cycle to ensure stable combustion, low emissions, acceptable
pressure levels and optimum efficiency.
[0028] Another object of the present invention is to provide a PCCI
engine and control system which effectively controls the
temperature, pressure, equivalence ratio and/or air/fuel mixture
autoignition properties to precisely control the timing of the
start of combustion.
[0029] A still further object of the present invention is to
provide a PCCI engine and control system which effectively achieves
continuous, stable PCCI combustion while achieving acceptable
cylinder pressures and the desired brake mean effective
pressure.
[0030] Yet another object of the present invention is to provide a
PCCI engine and control system which effectively controls the
commencement of combustion and the combustion rate so as to ensure
that substantially all of the combustion process occurs within an
optimal crank angle limit, i.e. 20 degrees BTDC through 35 degrees
ATDC, while minimizing emissions and maximizing efficiency.
[0031] Another object of the present invention is to provide a PCCI
engine which can be easily started.
[0032] Still another object of the present invention is to provide
a multi-cylinder PCCI engine and control system which effectively
minimizes variations in the combustion events of the cylinders.
[0033] Yet another object of the present invention is to provide a
multi-cylinder PCCI engine and control system which effectively
controls the start of combustion to achieve stable, low emission,
efficient combustion throughout exposure to changes in engine load
and ambient conditions.
[0034] Another object of the present invention is to provide a
control system for a PCCI engine which effectively detects or
senses the start of combustion to provide feedback control and then
controls the operating conditions of the engine to optimize the
start of combustion.
[0035] Still another object of the present invention is to provide
a PCCI engine and control system which effectively minimizes the
unburned hydrocarbon and carbon monoxide emissions.
[0036] The above objects and others are achieved by providing a
premixed charge compression ignition internal combustion engine,
comprising an engine body, a combustion chamber formed in the
engine body and combustion history control system for controlling a
combustion history of future combustion events to reduce emissions
and optimize efficiency. The combustion history control system
includes at least one of a temperature control system for varying
the temperature of the mixture of fuel and air, a pressure control
system for varying the pressure of the mixture, an equivalence
ratio control system for varying an equivalence ratio of the
mixture and a mixture autoignition property control system for
varying an autoignition property of the mixture. The engine further
includes an operating condition detecting device for detecting an
engine operating condition indicative of the combustion history and
generating an engine operating condition signal indicative of the
engine operating condition, and a processor for receiving the
engine operating condition signal, determining a combustion history
value based on the engine operating condition signal, and
generating one or more control signals based on the combustion
history value. The one or more control signals are used to control
at least one of the temperature control system, the pressure
control system, the equivalence ratio control system and the
mixture autoignition property control system to variably control
the combustion history of future combustion events.
[0037] The engine operating condition detecting device may include
a start of combustion sensor for sensing the start of combustion
and generating a start of combustion signal. Also, the combustion
history value may be determined based on the start of combustion
signal. The engine operating condition detecting device may be a
cylinder pressure sensor.
BRIEF DESCRIPTION OF THE DRAWINGS
[0038] FIG. 1a is a schematic diagram of one embodiment of the
present invention showing a single cylinder of the engine of FIG.
1b and associated control system;
[0039] FIG. 1b is a schematic diagram of a multi-cylinder engine of
the present invention;
[0040] FIG. 2 is a graph showing cylinder pressure and heat release
rate as a function of crank angle for the PCCI engine of the
present invention;
[0041] FIG. 3 is a graph showing the apparent heat release rate as
a function of crank angle for several different engine operating
conditions;
[0042] FIG. 4a is a graph showing knock intensity as a function of
time for a given set of operating conditions;
[0043] FIG. 4b is a graph showing gross indicated mean effective
pressure (GIMEP) as a function of time;
[0044] FIG. 4c is a graph showing peak pressure as a function of
time for the same conditions of FIGS. 4a and 4b;
[0045] FIG. 5 is a graph showing apparent heat release rate as a
function of crank angle and illustrating the increase in the heat
release rate duration as the combustion or heat release location or
timing is retarded;
[0046] FIG. 6 is a graph showing cylinder pressure as a function of
crank angle and illustrating the decrease in peak cylinder pressure
as the heat release rate retards;
[0047] FIG. 7a is a graph showing GIMEP as a function of intake
manifold temperature for two different engine speed cases;
[0048] FIG. 7b is a graph showing the coefficient of variation of
GIMEP as a function of intake manifold temperature for two
different engine speed cases;
[0049] FIG. 7c is a graph showing peak cylinder pressure as a
function of intake manifold temperature for two different engine
speeds;
[0050] FIG. 7d is a graph showing the start of combustion as a
function of intake manifold temperature for two different engine
speeds;
[0051] FIG. 7e is a graph showing heat release duration in crank
angle degrees as a function of intake manifold temperature for two
different engine speeds;
[0052] FIG. 7f is a graph showing heat release duration in time as
a function of intake manifold temperature for two different engine
speeds;
[0053] FIG. 7g is a graph showing gross indicated thermal
efficiency as a function of intake manifold temperature for two
different engine speeds;
[0054] FIG. 7h is a graph showing fuel specific hydrocarbons as a
function of intake manifold temperature for two different engine
speeds;
[0055] FIG. 7i is a graph showing fuel specific carbon monoxide as
a function of intake manifold temperature for two different engine
speeds;
[0056] FIG. 7j is a graph showing fuel specific oxides of nitrogen
emissions as a function of intake manifold temperature for two
different engine speeds;
[0057] FIG. 7k is a graph showing noise as a function of intake
manifold temperature for two different engine speeds;
[0058] FIG. 8 is a graph showing apparent heat release rate as a
function of crank angle for three different intake manifold
temperatures;
[0059] FIG. 9 is a graph showing both the start of combustion and
combustion duration as a function of wall temperature;
[0060] FIG. 10 is a graph showing both the start and end of
combustion as a function of crank angle for a given time period,
and GIMEP for the same time period, wherein a glow plug is
cycled;
[0061] FIG. 11 is a graph showing the apparent heat release rate as
a function of crank angle for the glow plug transient of FIG.
10;
[0062] FIG. 12 discloses one embodiment of an end cylinder
compensating system of the present invention for providing
cylinder-to-cylinder temperature control;
[0063] FIG. 13 is a schematic diagram of a second embodiment of the
end cylinder compensating device for providing cylinder-to-cylinder
temperature control;
[0064] FIG. 14 is a graph showing the effects of changing intake
and exhaust valve opening and closing events on top dead center
(TDC) temperature;
[0065] FIG. 15 is a graph showing the effects of changing intake
and exhaust valve opening and closing events, and variable
compression ratio, on the residual mass fraction and temperature at
top dead center;
[0066] FIG. 16 is a graph showing both cylinder pressure and heat
release as a function of crank angle for different exhaust valve
lash settings;
[0067] FIG. 17 is a graph showing the effects of varying exhaust
gas recirculation (EGR) on the location of the heat release rate
relative to the crank angle and the effect of variations in EGR on
the magnitude of the heat release rate;
[0068] FIG. 18 is a graph showing the effect of varying the EGR
rate on the timing of the start of combustion;
[0069] FIG. 19 is a schematic of an improved engine of the present
invention having one cylinder operating under PCCI conditions to
optimize the use of EGR;
[0070] FIG. 20 is a graph showing the effects of changing
compression ratio on the temperature at top dead center;
[0071] FIG. 21 is a graph showing the start of combustion as a
function of intake manifold temperature and the effects of changing
the compression ratio on the start of combustion and intake
manifold temperature;
[0072] FIG. 22a is a partial cross sectional view of one cylinder
of the PCCI engine of the present invention including one
embodiment of a compression ratio varying device;
[0073] FIG. 22b is a partial cross sectional view of one cylinder
of the PCCI engine of the present invention showing a second
embodiment of a compression ratio varying device;
[0074] FIG. 22c is a partial cross sectional view of one cylinder
of the present PCCI engine showing a third embodiment of the
compression ratio varying device;
[0075] FIG. 22d is a partial cross sectional view of a single
cylinder of the present PCCI engine showing a fourth embodiment of
the compression ratio varying device of the present invention;
[0076] FIG. 23 is a schematic diagram of an opposed piston PCCI
engine of the present invention including a variable phase shifting
mechanism for varying the compression ratio;
[0077] FIG. 24 is a side view of the differential mechanism used in
the variable phase shifting mechanism of FIG. 23;
[0078] FIG. 25 is a graph showing compression ratio as a function
of the degrees out of phase of two pistons in the opposed piston
engine, for example, of FIG. 23 illustrating various compression
ratio settings;
[0079] FIG. 26 is a graph showing cylinder volume as a function of
crank angle of a reference piston in an opposed piston PCCI engine
which shows that the compression ratio decreases as the pistons
become more out of phase;
[0080] FIG. 27 is a graph showing the effects of changing intake
and exhaust valve opening and closing events, and varying the
compression ratio, on the percent of baseline airflow rate and the
TDC temperature;
[0081] FIG. 28 is a graph showing the effects of changes and intake
in exhaust valve opening and closing events, and varying the
compression ratio, on the diesel equivalent brake specific fuel
consumption and TDC temperature;
[0082] FIG. 29 is a graph showing the effects of changes and intake
in exhaust valve opening and closing events, and variations in
compression ratio, on peak cylinder pressure and TDC
temperature;
[0083] FIG. 30 is a graph showing the effects of water injection on
intake manifold temperature and temperature at top dead center;
[0084] FIG. 31a is a graph showing the combustion duration in crank
angle degrees as a function of intake manifold pressure (IMP);
[0085] FIG. 31b is a graph showing combustion duration in time as a
function of IMP;
[0086] FIG. 31c is a graph showing the effect of changes in IMP on
the magnitude and timing or location of the heat release rate;
[0087] FIG. 31d is a graph showing the start of combustion timing
and crank angle degrees as a function of IMP;
[0088] FIG. 31e is a graph showing fuel specific hydrocarbons as a
function of IMP;
[0089] FIG. 31f is a graph showing GIMEP as a function of IMP;
[0090] FIG. 31g is a graph showing gross indicated thermal
efficiency as a function of IMP;
[0091] FIG. 31h is a graph showing fuel specific carbon monoxide as
a function of IMP;
[0092] FIG. 31i is a graph showing fuel specific oxides of nitrogen
emissions as a function of IMP;
[0093] FIG. 31j is a graph showing the coefficient of variation of
GIMEP as a function of IMP;
[0094] FIG. 31k is a graph showing the peak cylinder pressure as a
function of IMP;
[0095] FIG. 31l is a graph showing noise as a function of IMP;
[0096] FIG. 31m is a graph showing the effects of increasing IMP on
peak cylinder pressure and GIMEP;
[0097] FIG. 32 is a graph showing the effect of various trace
species on a start of combustion and temperature;
[0098] FIG. 33 is a graph showing the effects of additional amounts
of ozone on advancing the start of combustion;
[0099] FIG. 34 is a graph showing the effect of varying the type of
fuel used in the present PCCI engine on the start of combustion
wherein the increase in temperature indicates the start of
combustion;
[0100] FIG. 35 is a graph showing the apparent heat release
duration as a function of equivalence ratio;
[0101] FIG. 36 is a graph showing the start of combustion in crank
angle degrees as a function of equivalence ratio;
[0102] FIG. 37 is a graph showing the effects of variations in
equivalence ratio on the start of combustion wherein an increase in
temperature indicates the start of combustion;
[0103] FIG. 38 is a graph showing the effects of variations in the
equivalence ratio on the magnitude and timing, or location, of the
heat release rate;
[0104] FIG. 39 is a graph showing the effects of equivalence ratio
on the compressor pressure ratio and the compressor outlet
temperature;
[0105] FIG. 40 is a graph showing the effects of varying the
equivalence ratio on the brake specific fuel consumption;
[0106] FIG. 41 is a graph showing the differences in pumping mean
effective pressure and GIMEP for two differently sized turbine
casings;
[0107] FIG. 42 is a graph showing the diesel equivalent BSFC and
BMEP for two differently sized turbine casings;
[0108] FIG. 43 is a graph showing the turbine rotor speed and
intake manifold pressure for two differently sized turbine
casings;
[0109] FIG. 44 is a graph showing the fuel specific oxides of
nitrogen emissions for PCCI combustion with various fuels in
comparison to a typical compression ignition diesel engine;
[0110] FIG. 45 is a graph showing emissions as a function of engine
speed;
[0111] FIG. 46 is a graph showing emissions as a function of
temperature at bottom dead center;
[0112] FIG. 47 is a graph showing fuel specific carbon monoxide as
a function of end of combustion flame temperature;
[0113] FIGS. 48a-50b are partial cross sectional views of a single
cylinder of the PCCI engine of the present invention showing an
alternative embodiment including various crevice minimizing
features; and
[0114] FIG. 51 is a graph showing the effects of various
percentages of diesel pilot injections on the heat release rate
location and shape.
DETAILED DESCRIPTION OF THE INVENTION
[0115] The present invention is directed to an improved premixed
charge compression ignition (PCCI) engine and control scheme for
controlling the engine in a manner to optimally minimize emissions
while maximizing efficiency. For the purposes of this application,
PCCI refers to any engine or combustion process in which: 1) the
vast majority of the fuel is sufficiently premixed with the air to
form a combustible mixture throughout the charge by the time of
ignition and throughout combustion; and 2) combustion is initiated
by compression ignition. PCCI also refers to any compression
ignition engine or combustion process in which the fuel and air are
premixed long before ignition. As a result, the timing of injection
of the fuel in the PCCI engine does not affect the timing of
ignition of the fuel/air mixture. Also, it should be understood
that PCCI is meant to encompass homogeneous charge compression
ignition (HCCI) engines and processes wherein the mixture exists in
a homogeneous, or nearly homogeneous state, at the start of
combustion. In the present invention, the fuel/air mixture is
thoroughly mixed to form a very lean homogeneous mixture, or is
mixed in a manner to form a less homogeneous mixture with a desired
air/fuel stratification, to ensure relatively even, low flame
temperatures which result in extremely low oxides of nitrogen (NOx)
emissions. It should be understood the some engines operate under
PCCI conditions continuously while other engines may operate under
PCCI conditions for only a limited period of operation either by
design or inadvertently.
[0116] Applicants have recognized that the key to producing a
commercially viable PCCI engine lies in the control of the
combustion history of subsequent or future combustion events in
such a manner so as to result in extremely low NOx emissions
combined with very good overall efficiency, combustion noise
control and with acceptable cylinder pressure. The combustion
history may include the time at which combustion occurs (combustion
timing), the rate of combustion (heat release rate), the duration
of combustion and/or the completeness of combustion. Applicants
have determined that the combustion history, and especially the
combustion timing, is sensitive to, and varies depending on, a
variety of factors including changes in load and ambient
conditions. The engine and control system of the present invention
operates to actively control the combustion history of future
combustion events during engine operation to ensure the desired
combustion and engine operation is maintained. In the preferred
embodiment, the present engine and control system controls the
combustion timing during the compression and expansion events of
the engine.
[0117] FIGS. 1a and 1b illustrates the PCCI engine and control
system of the present invention, indicated generally at 10. FIG. 1a
shows a single engine cylinder 12 of the multi-cylinder
reciprocating piston type engine shown in FIG. 1b. Of course, the
PCCI control system of the present invention could be used to
control PCCI combustion in an engine having only a single cylinder
or any number of cylinders, for example, a four, six, eight or
twelve cylinder internal combustion engine. In addition, although
the present PCCI control system is primarily discussed with
reference to a four stroke engine, the present control system could
be applied to a two stroke engine. Also, the PCCI system of the
present invention may be adapted for use on any internal combustion
engine having compression, combustion and expansion events,
including a rotary engine and a free piston engine.
[0118] As shown in FIG. 1a, a piston 14 is reciprocally mounted in
the cylinder to form a combustion chamber 13. The piston transmits
forces generated by a combustion event into a conventional engine
drive system. Referring to FIGS. 1a and 1b, an intake air system 23
including an intake manifold 15 supplies intake air, or an air/fuel
mixture to a respective intake port 26 associated with each
cylinder 12. Likewise, an exhaust gas system 27 including an
exhaust manifold 17 receives exhaust gases flowing from exhaust
ports 31. One or more intake valves, such an intake valve 19 and
one or more exhaust valves, such as exhaust valve 21, are moved
between open and closed positions by a conventional valve control
system, or a variable valve timing system, to control the flow of
intake air or air/fuel mixture into, and exhaust gases out of, the
cylinder, respectively.
[0119] The PCCI system 10 includes a combustion sensor 16 for
sensing or detecting an engine operating condition indicative of
the combustion history and generating a corresponding signal 18. In
the preferred embodiment, sensor 16 permits effective combustion
control capability by detecting an engine operating condition or
parameter directly related to, or indicative of, the time at which
the combustion event occurs during the compression and/or expansion
strokes, i.e. preferably the start of combustion (SOC). For
example, a cylinder pressure sensor may be provided on any or all
engine cylinders for sensing, on a cycle-by-cycle basis, the SOC.
In this case, the sensor 16 also provides other engine condition
data, such as the combustion rate, combustion duration, combustion
event or heat release location and end of combustion data, any one
of which may be used instead of the start of combustion data. Any
conventional means for detecting the start of combustion may be
used, for example, by sensing the very rapid increase in the
cylinder pressure. Other forms of sensors could be used including
accelerometers, ion probes, optical diagnostics, strain gages
and/or fast thermocouples in the cylinder head, liner or piston.
Also, torque or RPM sensors could be used to detect changes in
engine torque and RPM associated with each combustion event.
Alternatively, or additionally, an emissions sensor could be used
to detect emissions having a known correlation to the completeness
of combustion.
[0120] Sensor 16 provides feedback control to an electronic control
unit 20 (ECU). ECU 20 receives signal 18, processes the signal and
determines an actual combustion history value, i.e. start of
combustion value. The actual combustion history value is then
compared to a predetermined desired combustion history value
obtained, for example, from a look-up table. Based on the
comparison of the actual combustion history value to the desired
combustion history value, ECU 20 then generates a plurality of
output signals, indicated at 22, for variably controlling
respective components of the system so as to effectively ensure, in
the preferred embodiment, that the SOC and completion of combustion
occur between 20 degrees before top dead center (BTDC) during the
compression stroke and 35 degrees after top dead center (ATDC)
during the power stroke of the piston thereby minimizing NOx
emissions while maximizing engine efficiency. The PCCI combustion
control scheme is most preferably implemented in software contained
in ECU 20 that includes a central processing unit such as a
micro-controller, micro-processor, or other suitable
micro-computing unit.
[0121] As discussed herein, PCCI system 10 may include various
components for optimizing the combustion event. The objectives of
the present system, i.e. low oxides of nitrogen (NOx) emissions,
high efficiency, etc, may be achieved using any one of the various
control components, or any combination of the components. In
particular, as shown in FIG. 1b, a compressor 24 may be provided
along an intake air system 23 upstream of intake manifold 15 for
varying the boost intake pressure. Compressor 24 may be driven by
any conventional means, such as an exhaust gas driven turbine 25. A
bypass circuit 33 including a waste gate valve 43 may be provided
in a conventional manner. A second compressor or supercharger 58
may be provided upstream of compressor 24. Supercharger 58 is
mechanically driven by the engine drive system. A charge air cooler
28 may also be provided downstream of compressor 24. Also, an
intake air heater 30 (such as a burner, heat exchanger or an
electric heater) may be provided, for example, after cooler 28 as
shown in FIG. 1b, or alternatively, upstream of compressor 24.
Also, an individual heater 29 may be provided in the intake port 26
associated with each cylinder 12 to provide quicker control of the
intake manifold temperature for each cylinder to enhance both
individual cylinder combustion control and balancing of combustion
between the cylinders. Compressor 24, cooler 28 and heater 30 each
include control devices for varying the effect of the particular
component on the pressure/temperature of the intake air or mixture.
For example, a bypass valve or waste gate 43 could be used to
regulate the amount of exhaust gas supplied from the associated
exhaust system, which is connected to an exhaust duct 31, to
turbine 25 thereby varying the intake pressure as desired.
Similarly, a control valve could be provided in the cooling fluid
flow path supplied to cooler 28 to permit variable control of the
cooling effect of cooler 28. Likewise, various types of variable
controls could be used to vary the heating effect of heater 30.
Output signals 22 from ECU 20 are supplied to the various control
devices to control compressor 24, cooler 28 and heater 30 so as to
variably control the pressure and temperature of the intake air or
mixture preferably on a cycle-by-cycle basis.
[0122] In addition, the PCCI system 10 may include a plurality of
fuel supplies 32 and 34 for supplying fuels having different
autoignition properties (for example, different octane or methane
ratings, or activation energy levels) into the intake air flow.
Fuel control valves 39 and 41 are used to control the amount of
each fuel supply 32, 34 delivered, respectively. For example, fuel
may be supplied along the intake air path between cooler 28 and air
heater 30 as shown in FIG. 1b. Of course, fuel could be introduced
at various locations along the intake of the engine, such as
upstream of the cooler, e.g. upstream of the compressor.
Alternatively, the fuel could be injected, by for example an
injector 35, into the respective intake duct 26 associated with
each cylinder, as shown in FIG. 1a.
[0123] The present PCCI system 10 also importantly includes a
variable compression ratio means 38 for varying the compression
ratio so as to advantageously advance or retard the combustion
event as desired. For example, variable compression ratio means 38
may be in the form of a control mechanism for varying the shape of
the combustion chamber or height of the piston to vary the
effective compression ratio. The effective compression ratio could
also be varied by varying the timing of closing of intake valve 19
as discussed more fully hereinbelow. The variations in the timing
of opening and closing of the intake and exhaust valves may be
accomplished using any conventional variable valve timing actuator
system capable of receiving signals from ECU 20 and effectively
varying the opening and/or closing of the valves in accordance with
the principles set forth hereinbelow.
[0124] In addition, in-cylinder diluent injection may be
accomplished using an injector 40 for injecting a gas or liquid,
e.g. air, nitrogen, carbon dioxide, exhaust gas, water, etc., into
the cylinder to vary the temperature and the temperature
distribution in the cylinder so as to control the combustion event.
Similarly, a diluent may be injected into intake duct 26 using, for
example, an injector 42.
[0125] The present PCCI system may also include a fuel injector 36
for injecting fuel 37, e.g. diesel fuel, directly into the
combustion chamber. Fuel 37 would be injected either early in the
compression event, preferably approximately between 180 degrees and
60 degrees BTDC, as described below, or later in the compression
event near TDC.
[0126] By injecting the fuel 37 early in the compression event, it
is much more thoroughly mixed with the fuel/air mixture received
from the intake duct than would be the case for a diesel engine,
thus ensuring a more desirable combustion process, in particular
the fuel will burn at a leaner equivalence ratio which results in
much lower NOx emissions. The start or initiation of the combustion
(SOC) of the fuel/air mixture received from the intake duct may be
varied by controlling the quantity of fuel 37 injected. For
instance, an earlier combustion event may be achieved by increasing
the quantity of fuel 37 while the timing of the combustion event
may be delayed by decreasing the quantity of fuel 37 injected.
[0127] By injecting the fuel 37 later in the compression stroke,
that is near TDC, conventional diesel fuel injection systems can be
used. This approach could be combined with the introduction of one
or more additional types of fuel in the intake manifold to achieve
a PCCI mode of operation. In particular, the fuel injected into the
intake manifold would have a higher excess air ratio. The excess
air ratio is the actual air-fuel ratio of the engine divided by the
air-fuel ratio at stoichiometric conditions. For the very lean
excess air ratio, combustion along a flame front is impossible.
However, autoignition is possible thereby allowing combustion of a
mixture that would be too lean to burn in a typical spark-ignited
engine. Applicants have determined that PCCI combustion does not
initiate at, and propagate out from, a single location. On the
contrary, the results show that combustion includes multiple
ignition sites distributed throughout the combustion chamber.
[0128] For efficient, low emission PCCI combustion, it is important
to have combustion occur during an appropriate crank angle range
during the engine cycle. If combustion starts too early, cylinder
pressures will be excessively high and efficiency will suffer. If
combustion starts too late, then combustion will be incomplete
resulting in poor HC emissions, poor efficiency, high carbon
monoxide (CO) emissions, and poor stability. Applicants have
determined that the timing of the SOC and the combustion rate, and
therefore combustion duration, in a PCCI engine primarily depend on
the temperature history; the pressure history; fuel autoignition
properties, e.g. octane/methane rating or activation energy, and
trapped cylinder charge air composition (oxygen content, EGR,
humidity, equivalence ratio etc.). The present invention presents a
structured approach to affecting these variables in such a way that
the start of combustion and/or the combustion rate (heat release
rate) can be controlled through various combinations of features
discussed more fully hereinbelow.
[0129] The various control features for controlling the start of
combustion and the combustion rate are controlled/varied to ensure
optimum combustion throughout engine operating conditions so as to
achieve low NOx emissions and high efficiency. Application of these
control features will cause combustion to occur within a preferred
crank angle range relative to the top dead center position of the
engine piston. Specifically, applicants have recognized that
substantially all of the combustion event should occur between 20
crank angle degrees BTDC and 35 crank angle degrees ATDC. Also,
combustion would be initiated, preferably between 20 crank angle
degrees BTDC and 10 crank angle degrees ATDC, and ideally,
approximately between 10 degrees BTDC and 5 degrees ATDC. In
addition, the duration of the combustion event will typically
correspond to a crank angle in the range of 5-30 crank angle
degrees. Preferably, however, one or more of the control features
listed below will be controlled to prolong the duration of
combustion to approximately 30-40 degrees to achieve desirable peak
cylinder pressures and reduced noise. Thus, optimal control of one
or more of the following features will effectively control the
start of combustion and/or the rate of combustion such that
substantially all of the combustion event occurs between 20 crank
angle degrees BTDC and 35 crank angle degrees ATDC. Of course,
there may be conditions under which the start of combustion occurs
outside the above-stated crank angle range and/or the duration of
combustion in the PCCI engine occurs over a broader crank angle
range, or may extend beyond the limit described above.
[0130] Applicants have shown that stable, efficient PCCI combustion
can be achieved with most of the heat release occurring after TDC.
For example, as shown in FIG. 2, the centroid of heat release may
be positioned at 5.degree. ATDC. Applicant have determined that, at
light load and lean conditions, as shown in FIG. 3, heat release
duration may be in the range of approximately 21.5-25 crank angle
degrees.
[0131] As shown in FIGS. 4a, 4b and 4c, applicants have determined
that with an engine running close to its misfire limit, the SOC and
end of combustion (EOC) progressively retard and heat release
duration lengthens. Gross indicated mean effective pressure (GIMEP)
passes through a maximum as the SOC retards to after TDC.
Meanwhile, the knock intensity and peak cylinder pressure (PCP)
decrease substantially close to the misfire limit, while GIMEP
remains acceptable. As shown in FIG. 5, the peak heat release rate
also decreases and the heat release duration increases as the
misfire limit is approached. Meanwhile, as shown in FIG. 6, the
peak cylinder pressure decreases as the heat release rate retards.
Clearly, the engine cannot sustain this reaction process without
providing the appropriate controls discussed herein. Applicants
have determined that the best operating point occurs with the SOC
occurring a few degrees after TDC. Certainly, improving the
PCP-GIMEP tradeoff for PCCI combustion requires a SOC after TDC. As
a result, it is clear that variable, active control is necessary to
maintain the SOC and duration of combustion at the desired location
and at the desired length, respectively, to achieve effective,
efficient PCCI combustion.
[0132] Variation in the SOC, between sequential combustion events
in a single cylinder engine and between cylinders in a
multi-cylinder engine, is due to the sensitivity of PCCI combustion
to the pressure and temperature history leading up to the
particular combustion event. Very small variations in the
compression ratio, the amount of trapped residual, wall
temperatures, etc. have a significant effect on the pressure and
temperature history. The present PCCI engine and method of
operating the engine include control variables/features capable of
compensating for, and controlling, these variations to achieve
optimum PCCI combustion.
[0133] Generally, the control variables, which can be used to
effectively control the commencement of combustion and the
combustion rate so as to ensure that substantially all of the
combustion process occurs within the optimal crank angle limit,
i.e. 20 degrees BTDC through 35 degrees ATDC while minimizing
emissions and maximizing efficiency, may be classified in four
categories of control: temperature control; pressure control;
control of the mixture's autoignition characteristic; and
equivalence ratio control.
[0134] Temperature Control
[0135] The temperature of the in-cylinder air/fuel mixture
(in-cylinder temperature) plays an important role in determining
the start of combustion. The in-cylinder temperature may be varied
to control the start of combustion by varying certain key control
features, such as compression ratio (CR), intake manifold
temperature (IMT), exhaust gas recirculation (EGR), residual mass
fraction (RMF), heat transfer and temperature stratification.
[0136] Applicants have determined that intake manifold temperature
(IMT) has a significant effect on propane-fueled PCCI combustion.
During two of Applicants' studies, engine speed, equivalence ratio
(.PHI.) and intake manifold pressure (IMP) were held constant while
IMT was swept through the practical operating range. The lowest IMT
was limited by unstable operation and the highest IMT was limited
by maximum allowable peak cylinder pressure (PCP). The conditions
of the first and second studies, respectively, included engine
speed=1200 rpm and 2000 rpm; equivalence ratio=0.30 and 0.24; and
IMP=3.3 bar and 4.1 bar. As shown in FIGS. 7a and 7b, increasing
IMT resulted in increased GIMEP and a decreased coefficient of
variation (CoV) of GIMEP. Also, increasing IMT increased the PCP as
shown in FIG. 7c, while advancing the SOC and decreasing combustion
duration (FIGS. 7d-7f). Increasing IMT also increased gross
indicated thermal efficiency (FIG. 7g) and the estimated noise
(FIG. 7k). With respect to emissions, increasing IMT decreased FSHC
emissions (FIG. 7h), decreased fuel specific carbon monoxide (FSCO)
emissions (FIG. 7i), but had no observable effect on FSNOx (FIG.
7j).
[0137] In summary, Applicants have determined that small changes in
IMT have large effects on many aspects of propane-fueled PCCI
combustion. By varying the intake temperature, the combustion event
can be advanced or retarded. Increasing the intake temperature will
advance the start of combustion; decreasing the intake temperature
will retard the start of combustion, as shown graphically in FIG.
8. This temperature control may be accomplished using heat
exchangers or burners. For example, a charge air cooler may be
positioned along the intake manifold. A burner or heater in
combination with a cooler offers exceptional intake temperature
control. The exhaust products of the burner may be directly mixed
with the intake air, the burner could use the intake air directly
for its air supply, or the heat generated by the burner could be
added to the intake air through a heat exchanger. The heat
exchanger may use waste heat in engine coolant or exhaust gases to
heat the intake air. Also, rapid control of IMT can be achieved by
using a charge air cooler bypass. A regenerator (similar to that
used in a Stirling engine) could be used to recover and transfer
exhaust heat into the intake air through a heat exchanger thereby
controlling the intake temperature. In addition, IMT could be
varied by injecting fuel into the manifold in different phases,
e.g. as a liquid or a gas. The change in the heat required for
vaporization of a liquid fuel would reduce IMT. Of course,
different types of fuels would have different effects on IMT.
[0138] Applicants have also determined how residual and intake
temperature, boost and combustion chamber and port wall heat
transfer, affect in-cylinder bulk temperature throughout intake and
compression, and also the effect on spatial temperature
distribution at TDC. Specifically, Applicants compared the intake
and compression events for an engine running on an air and propane
mixture. Applicants determined that the temperature at the SOC is
also determined in part by the reheating of the intake charge by
existing heat energy. For the purposes of this application, reheat
is defined as T(average in-cylinder@intake valve closing
(IVC))-T(average intake manifold), that is, the difference between
intake manifold temperature, i.e. temperature assigned at the inlet
to the port and the in-cylinder bulk temperature at IVC. Applicants
determined that reheat starts in the port and continues
in-cylinder. Moreover, 56% of reheat was due to wall heat transfer
and 44% due to mixing and boost for the condition examined.
Clearly, heat transfer is very important in determining reheat.
[0139] One study that elucidates the importance of the wall
temperatures on the in-cylinder heat transfer is the following. In
comparing the firing cylinder to the misfiring cylinder, it was
noted that the misfiring cylinder's reheat was 63% of the firing
case (27 vs 43 K). Lower wall temperatures for a misfiring cylinder
compared to a firing cylinder are the main reason for its lower
in-cylinder temperatures. The firing cylinder had a TDC in-cylinder
temperature 46 K higher than a misfiring cylinder, compared to a 16
K higher temperature at IVC. If compression were done adiabatically
for each case, the temperature difference at TDC would have been
.about.35 K given the initial 16 K difference. Therefore, .about.11
K (46-35 K) temperature loss from IVC to TDC is due to cooler
misfiring wall temperatures. Interestingly, although walls heat the
in-cylinder gases for the majority of the intake and compression
event, relatively fast rates of heat transfer out of the gas near
TDC compression can result in cooler in-cylinder contents than if
there were no heat transfer at all. Also, mass flow rate decreased
7.5% due to heat transfer when comparing a normally firing cylinder
with wall heat transfer to a firing cylinder with adiabatic walls,
primarily due to the density effect.
[0140] Referring to FIG. 9, with respect to the effect of wall
temperatures, i.e. piston temperature, head temperature, and liner
temperature, on the SOC, Applicants have determined that as wall
temperatures are increased, SOC becomes more advanced. The
increased surface temperatures cause lower heat transfer to the
combustion chamber surfaces thereby advancing combustion.
Applicants have shown that with wall temperature varying from 255
to 933 K and all other parameters kept constant (IMT= 342 K,
reheat=43 K, .phi.=0.24), the mixture did not ignite with a wall
temperature below 400 K. From about 400 K to 550 K combustion
duration increases as a larger percent of the fuel burns. Above 550
K all the fuel burns and the combustion duration decreases with
increasing temperature. Varying in-cylinder surface temperatures
can be achieved by varying the cooling effect of the engine coolant
and/or the lubricating oil on the cylinder/piston assembly.
Although cylinder wall temperature may be difficult to use as a
lever for effectively controlling SOC, cylinder wall temperatures
are one of the parameters considered when controlling SOC,
particularly for starting or transient operation. Applicants have
shown that there is a region of operating conditions where there
are two stable solutions: one without combustion and cool walls,
and one with combustion and hot walls. Also, varying the surface to
volume ratio in the combustion chamber can change the heat transfer
and, therefore, can be used to control the combustion.
[0141] By comparing a normally firing cylinder with wall heat
transfer to a firing cylinder with adiabatic walls, wall heat
transfer is seen to be the major contributor to spatial temperature
distribution at TDC. Spatial temperature distribution is defined as
the manner in which the temperature varies throughout a region, be
it in the port, or in the cylinder at a particular crank angle. By
varying the in-cylinder temperature distribution, the start of
combustion and/or the overall combustion rate can be positively
affected. One way to vary in-cylinder temperature distribution is
to use split intake ports arranged so that some of the incoming
air/fuel mixture is warmer/colder than the rest of the incoming
mixture. Another method is to introduce hot spots in the cylinder
or to use a glow plug 44 (FIG. 1a). Also, in-cylinder temperature
distribution may be controlled by varying the temperature of the
combustion chamber walls (e.g. the wall temperature of the cylinder
liner, piston and/or engine head) by varying, for example, the
temperature of the engine coolant, the temperature of the engine
oil or the rate of cooling of the combustion chamber walls. As
shown in FIG. 1b, the temperature of the engine coolant may be
varied by controlling the flow through a coolant heat exchanger 46
positioned in the engine coolant circuit 47 by varying the flow
through a bypass circuit 48 using a bypass valve 50. It was
determined that wall heat transfer has similar impact on spatial
temperature distribution for both firing and misfiring cylinders.
Similarly, applicants also determined how residual temperature and
wall heat transfer affect in-cylinder temperature distribution
throughout intake and compression. The determination included three
studies of the intake and compression events of an air and propane
mixture. These studies revealed that, during most of intake and
compression, hot residual is the main source of spatial temperature
variation. However, near TDC compression, residual history is of
minor importance compared to heat transfer with the walls in
setting up temperature variations in the combustion chamber. As a
result, it is believed that to promote a combustion event that uses
more of the fuel that is available, fuel may be introduced in such
a way that at SOC, fuel and air exist in proper proportion in
regions where the temperature field is adequate to sustain
combustion. Two areas where the temperature field is inadequate to
sustain combustion are in the crevices and adjacent cooled
surfaces. It is therefore desirable to keep the fuel away from both
the crevices and cooled surfaces.
[0142] Clearly, heat transfer into the in-cylinder mixture
increases the temperature of the in-cylinder mixture thus advancing
SOC. Applicants have shown that a glow plug can be used to
effectively control the SOC to a small degree. As shown in FIG. 10,
once the glow plug is turned off, the SOC and EOC retard slightly.
Also, GIMEP decreases significantly since less fuel is being
burned. The decrease in the amount of fuel being burned also
results in a decrease in the heat release rate as shown in FIG. 11.
Between cycles #1 and #100, the glow plug was turned off and
remained off until a time between cycles #300 and #400, at which
point it was turned back on. Perhaps most importantly, when the
glow plug is turned off, the start of rapid combustion is
significantly delayed without an increase in duration, which in
combination with the decrease in heat release rate, causes the
cumulative heat release to decrease. Thus, glow plug 44 (FIG. 1b)
could be used to positively control combustion to a limited
degree.
[0143] In any practical reciprocating engine, heat will be lost
from the combustion chamber during the compression process. The
heat loss depends upon many factors, but primarily upon engine
speed and the temperature difference between inside and the outside
of the cylinder. This heat transfer during the compression process
becomes a problem for diesel engines during cold ambient starts as
combustion can be difficult to initiate and sustain in cylinders
where the combustion chamber surfaces are cold. Typically, the
cylinders located at the ends of each bank of cylinders run the
coldest and are the least likely to fire. It is quite common under
such conditions for the charge in the end cylinders to fail to
combust due to excessive heat exchange with the colder cylinder
walls. With diesel engines, however, once all the cylinders warm
up, combustion is quite consistent and much less dependent on
combustion chamber surface temperatures.
[0144] With PCCI, the combustion process is initiated by obtaining
a certain pressure and temperature "history". Thus, as discussed
hereinabove, the PCCI combustion process is strongly dependent
upon, and sensitive to, the surface temperatures of the combustion
chamber. The present PCCI engine may include an end cylinder
compensating means for achieving desired combustion chamber surface
temperatures in the end cylinders to ensure better
cylinder-to-cylinder temperature control thereby increasing the
likelihood of stable combustion and very low NOx emissions. The end
cylinder compensating means may include a system for reducing the
effective cooling of specific cylinders, such as reducing piston
cooling nozzle flow; increasing coolant temperature; or reducing
coolant flow rate. Specifically, referring to FIG. 12, the end
cylinder compensating means may include an oil flow control system
70 including oil flow control valves 72 positioned in branch flow
passages 74 delivering cooling oil to piston cooling nozzles 76
from an oil pump 78. Thus, control valves 72 can be controlled to
vary the flow of cooling oil to the piston assemblies to vary the
temperature of the piston and thus favorably influence the
in-cylinder temperature. Alternatively, flow restrictions could be
used instead of valves 72, or the nozzles 76 associated with the
end cylinders may be designed with a smaller effective flow area
than the remaining nozzles to permanently reduce the flow to these
piston cooling nozzles. In addition, if more than one nozzle 76 is
provided as shown in FIG. 1a, the number of nozzles operating could
be varied by controlling the respective control valves associated
with each nozzle.
[0145] Referring to FIG. 13, end cylinder compensating means may
include an engine coolant flow control system 80 including a
coolant pump 81 and coolant flow control valves or restrictions 82
positioned in branch passages 84 leading to the end cylinders 86 of
the engine 88. The valves 82 are operated to reduce the flow of
cold coolant delivered from a radiator 90. Also, control valves 92,
positioned in hot coolant return passages 94, are used to control
the flow of higher temperature coolant, bypassing radiator 90, and
delivered directly to the end cylinders. These systems all function
to control the flow of coolant to the end cylinders to compensate
for the fact that they are cooled more by the ambient surroundings
so that the total cooling to each end cylinder is equal to each of
the other cylinders. These systems can be used to assist in
cylinder warm-up to improve engine startability and to provide
enhanced control of cylinder combustion and cylinder-to-cylinder
balancing.
[0146] The end cylinder compensating means may, alternatively, or
additionally, include end cylinders having an effective compression
ratio nominally greater than the other cylinders to offset the
extra heat loss. This compression ratio could be designed into the
end cylinders so that the end cylinder compression temperature is
equal to the middle cylinders. This approach is advantageous from a
performance perspective since end cylinder combustion chamber
surface temperatures would be enhanced for both start-up as well as
warmed-up operation. This compression ratio difference may
alternatively be accomplished through the camshaft valve lobe
phasing. In this scenario, the end cylinders would have intake
valve closing (IVC) near bottom dead center (BDC) so that the
effective compression ratio (CR) is approximately equal to the
geometric CR. The middle cylinders could then have a retarded IVC
which would produce a lower nominal effective CR than the end
cylinders. The effect of varying the compression ratio on PCCI
combustion is discussed more fully hereinbelow.
[0147] One of the biggest challenges with premixed charge,
compression ignition (PCCI) engine technology is in the placement
of the heat release profile. Start of combustion with standard
diesel or spark ignition engines is controlled with injection
timing or spark timing. With PCCI engines, the start of combustion
is dictated by the in-cylinder temperatures and pressures. As SOC
timings near TDC (and after) are approached on the PCCI engine, the
sensitivity to small geometric and/or operational variations in
temperatures, pressures, etc. increase dramatically. As retarded
heat release profiles are sought for PCCI engines (to minimize peak
cylinder pressures and improve efficiency), the risk of misfire or
partial burn increases dramatically. This is due to the fact that
the cylinder temperatures decrease after top dead center due to the
expansion of the charge. If autoignition has not yet occurred by
TDC, autoignition will not likely occur much after top dead center.
This problem is further aggravated if one cylinder begins to
misfire. The misfiring cylinder cools down making it even more
likely that the misfiring will continue.
[0148] In a multi-cylinder engine variations inevitably exist
between cylinders with respect to compression ratio, wall
temperatures, reheat and residual mass fraction. This variability
makes it quite difficult to operate a PCCI engine with the desired
retarded combustion timing while maintaining optimum combustion
without having individual cylinders (which happen to be running
slightly cool) begin to misfire.
[0149] Applicants have determined that manipulating valve events
can have a significant effect on the temperature at TDC and
therefore is an effective tool for controlling the start of
combustion as suggested by analytical results shown in FIG. 14.
Specifically, referring to Table I, varying valve events has the
following effects:
1TABLE I modified effect of advancing valve effect of retarding
valve event baseline timing relative to baseline timing relative to
baseline EVC -357.degree. traps hot residual which exhaust blown
back into advances SOC intake which advances SOC EVO 135.degree. no
effect no effect IVC -167.degree. Miller cycle - lowers at these
particular conditions, effective CR which retards retarding
slightly improves SOC breathing; retarding further reduces
effective CR which retards SOC IVO 341.degree. allow hot exhaust to
flow into restricts flow from intake intake which advances SOC
manifold which has minimal effect on SOC
[0150] As shown in FIG. 15, exhaust valve closing (EVC) plays a
significant role in determining the amount of combustion products
that remain in, or are made available to, the combustion chamber
from one combustion event to the next event, i.e. the residual mass
fraction (RMF). The residual exists at a higher temperature than
the incoming charge and therefore heats the charge for the next
combustion event. Thus, the timing of exhaust valve closing can be
used to adjust the in-cylinder temperature and therefore
controlling the SOC. In order to "heat up" a cold cylinder (e.g.
one that is beginning to misfire) the residual mass fraction can be
increased in the individual cylinder by an early exhaust valve
closing event. These hot residuals will increase the reheat of the
incoming charge and tend to advance the start of combustion
thereby, for example, restoring a misfiring cylinder. As shown in
FIG. 15, advancing EVC traps hot residual in the cylinder while
retarding EVC allows hot exhaust to be blown back into the cylinder
(in this case, exhaust manifold pressure (EMP)>IMP). The
baseline EVC is the optimum of these two effects: trapping the
minimum amount of residual and resulting in the lowest TDC
temperature. Similarly, advancing IVO allows some of the hot
residual in the cylinder to be blown back in to the intake, again
because EMP>IMP, causing the TDC temperature to increase.
Lowering compression ratio, discussed more fully hereinbelow, by,
for example, advancing IVC, will also increase residual in the
cylinder, but to a lesser extent. Adjusting the timing of exhaust
valve closing may also be used to effectively compensate for the
small geometric and operational variations between the cylinders to
permit the engine to be "tuned" cylinder-to-cylinder. Any other
means for effectively increasing or decreasing the RMF may be used
to advance or retard the SOC, respectively.
[0151] One method of implementing this strategy has been
successfully tested on a multi-cylinder PCCI engine. This technique
involved the increase of the exhaust valve lash setting. Opening up
the lash effectively closes the exhaust valve early and advances
the start of combustion as desired. Applicants have determined that
reducing the exhaust valve event by 10 degrees leads to slightly
higher surface temperatures and 22 degree warmer inlet
temperatures. Given the dramatic effect that 22 degree IMT swings
have on combustion (FIGS. 7c-7f), this method would indicate a
potential for tuning the multi-cylinder engine with valve lash
adjustments. As shown in FIG. 16, shortening the duration that an
exhaust valve is open by increasing the lash does indeed advance
combustion. Ultimately, cylinder-to-cylinder variations can be
controlled passively by any means which can adjust the static
exhaust valve closing. It could also be controlled actively if it
is coupled with some diagnostic measurements. If control exists on
all cylinders then this technique could also be used to effect the
overall start of combustion within the engine.
[0152] Another method of controlling in-cylinder temperature by
controlling the residual mass fraction (RMF) is to compress a
pocket of residual gas from the previous cycle in a chamber
positioned separate from the incoming charge. The proportion of
trapped residual to fresh charge can be manipulated by the size of
such a chamber. The mass of hot exhaust could be as large as
(1/2)(1/CR) and therefore. {fraction (1/30)} of the chamber mass if
all the TDC volume is in such a chamber. The structure of such a
chamber will have to be managed to make at least a portion of the
hot gas survive the compression process without completely mixing
with the incoming charge. If the trapped exhaust is mixed very
early in the compression process, the high temperature required to
initiate the fast reactions will not be reached. The timing of
flows into and out of such a chamber may help manage the timing of
the beginning of rapid energy release in cylinder. Additional
sources of local heat input may be able to supply such a fast
reaction initiation. This might be a heated glow plug or a
thermally isolated mass.
[0153] The residual mass fraction is also sensitive to the exhaust
manifold back pressure (EMP). By increasing EMP relative to IMP,
the residual mass fraction can be increased thus increasing the
temperature of the charge which, in turn, advances combustion.
Applicants have determined that raising EMP does have the expected
result of advancing SOC. However, applicant also showed that SOC
advanced only by about 4.degree. with a 3 bar increase in EMP for a
four cycle engine. Applicants have determined that the increase in
temperature is nearly linear with increase in EMP, with all other
things being held constant. For a 1 bar increase in EMP,
temperature at TDC increased about 10 K. Therefore, considering the
practical range of EMP, controlling EMP seems to be a relatively
weak lever in controlling SOC on a four cycle engine. Moreover, a
very substantial BSFC penalty is paid when using EMP to increase
TDC temperature in a four cycle engine. The BSFC would be
significantly higher than using either exhaust valve closing or
variable compression ratio. Although the effect of increasing EMP
is the same as advancing EVC, i.e. trapping more hot residual mass
in the cylinder, the BSFC is much higher because, when EMP is
increased, the piston has to work against that pressure through the
entire exhaust stroke. If the engine has turbomachinery, then
further complications would arise with trying to use EMP to control
SOC. However, using an exhaust restriction may still be viable on a
two-cycle engine.
[0154] Another important way to control intake temperature is by
using hot exhaust gas recirculation (EGR). As shown in FIG. 1b, a
high pressure EGR circuit 54 may be used to direct hot exhaust gas
from upstream of turbine 25 into the intake system 23. EGR circuit
54 includes a high pressure EGR control valve 60 for controlling
the recirculation of exhaust gas. A low pressure EGR circuit 62 and
control valve 64 may be used to direct a flow of low pressure EGR
from downstream of turbine 25 into the intake system 23. Applicants
have shown that EGR is especially effective in increasing the
intake manifold temperature when introduced upstream of the
compressor 24 (assuming the effect of adding EGR is not cancelled
by additional charge air cooling). Exhaust gas recirculation (EGR)
has more utility in PCCI engines because the exhaust gas of such an
engine will contain less particulates and thus the exhaust gas can
be recirculated to the ideal upstream location (intake of
compressor of turbocharger). The intake of the compressor is the
best location because the pressure differential is almost always
favorable. The fresh intake air and hot EGR mixture will get
compressed by the compressor thereby providing heating and mixing.
By introducing the EGR upstream of the compressor and increasing
the compressor inlet temperature, the result is a much higher
compressor outlet temperature than if the EGR is introduced after
the compressor. Introducing EGR into the intake of the compressor
is very difficult in normal diesel engines because the particulates
in the exhaust gases of the engine "gum up" the compressor. In a
PCCI engine, however, the virtually particulate free exhaust could
be introduced upstream of the compressor without significant
problems. Also, as shown in FIGS. 16, 17 and 18, applicants have
determined that, regardless of the technique used to introduce
exhaust products, e.g. EGR, RMF, etc., by adding exhaust products
while maintaining the temperature of the charge by, for example,
injecting a cooling diluent, such as air and/or water, the
combustion rate can be slowed thus increasing the combustion
duration, retarding combustion and decreasing the amount of heat
release.
[0155] Referring to FIG. 19, an improved engine 100 is shown which
benefits from the PCCI engine and control system of the present
invention by operating a limited number of a plurality of cylinders
in a PCCI mode while operating the remainder of the cylinders in a
diesel mode. Specifically, for example, five cylinders 102 in a six
cylinder engine may be operated in the diesel mode while one
cylinder 104 is operated in a PCCI mode. This engine also includes
an EGR system 106 associated only with the PCCI cylinder 104 and
separate from an exhaust system 108 associated with the diesel
cylinders 102. The pressure of the piston in the PCCI cylinder 104
is used to force the exhaust gas into the intake system. The EGR
system 106 includes an EGR cooler 110 utilizing, for example,
engine coolant, which cools PCCI exhaust gas before recirculating
the gas to the upstream side of a compressor 105. Of course, the
exhaust gas could be delivered to the intake manifold 112 serving
only diesel cylinders 102. A well known problem confronted in the
use of EGR in diesel engines is the excessive amounts of
particulates and NOx present in diesel engine exhaust gas. The
improved engine 100 permits a diesel engine to benefit from EGR
while substantially avoiding the drawbacks associated with heavy
particulate diesel exhaust thereby providing a less complex and
costly system. For example, as discussed hereinabove, the PCCI EGR
from cylinder 104 could more easily be introduced upstream of the
compressor without fouling the compressor. Also, the low NOx
emissions of the PCCI EGR reduce the formation nitric acid thereby
reducing corrosion in the engine. Applicants have shown that the
engine of FIG. 19 lowers the brake specific NOx emissions while
only negligibly increasing the brake specific fuel consumption.
[0156] Perhaps one of the most effective control features for
varying the temperature at TDC and therefore the SOC is variable
control of the compression ratio (CR) of a cylinder. By varying the
effective or the geometric compression ratio, both the temperature
and the pressure histories can be controlled. Increasing the
compression ratio advances the combustion event. Decreasing the
compression ratio retards it. For certain purposes, the compression
ratio may range from 24:1 (to promote cold starting) to 12:1 (to
permit control over the start of combustion and limit the peak
combustion pressures). The range of compression ratios would depend
on, among other factors, the type of fuel used (more specifically
its autoignition properties), for example, natural gas or propane.
Applicants have determined the effect of compression ratio on PCCI
combustion. For example, referring to FIG. 20, applicants have
shown that varying the compression ratio is a large lever in
changing in-cylinder temperature and therefore SOC. As shown in
FIG. 21, applicants have shown that variations in compression ratio
significantly affects the location of the SOC relative to TDC.
[0157] The compression ratio can be varied by varying the geometric
compression ratio, i.e. using a control mechanism to vary the
physical dimensions/shape of the combustion chamber. The present
invention includes a compression ratio varying device 38 for
varying the geometric or the effective volume of the combustion
chamber during engine operation to achieve a desired SOC. The
compression ratio varying device may be a mechanical device for
causing compression heating of the charge near TDC by changing the
geometric volume of the combustion chamber. As shown in FIGS.
22a-22d, the compression ratio varying device may include a movable
auxiliary piston or plunger which moves to extend into the
combustion chamber at a crank angle near TDC to decrease the
combustion chamber volume thereby increasing the compression ratio
and heating the charge sufficiently to allow ignition to start. The
key function of the plunger is to displace some of the charge near
TDC. Therefore, the shape and location of the plunger in the
combustion chamber will not be critical to its function, except to
the extent that the plunger affects the crevice volume.
[0158] The size of the plunger will be based on the desired
compression ratio control range and may be estimated by the
following example:
Swept volume (displacement) per cylinder=1,000 cc1 l.
TDC clearance volume=100 cc
Compression ratio=(1000 cc+100 cc)/100.0 cc=11.0
[0159] If the plunger volume=30 cc, then the effective compression
ratio with plunger fully extended=(1000 cc+100 cc)/(100 cc-30
cc)=15.7.
[0160] For a given set of conditions, the modified compression
ratio should be sufficient to allow a large enough increase in
temperature and pressure to cause compression ignition for a
fuel/air mixture that would not ignite without the plunger. Of
course, the engine's compression ratio and the size of the plunger
are easily changed during the design stage of the engine. Also,
different fuels and intake temperatures could require different
plunger sizes and compression ratios.
[0161] As shown in FIG. 22a, the plunger 150 may be positioned in a
bore 152 in the cylinder head 154 and operated by a cam 156 rotated
in predetermined timed relationship to the movement of the engine
piston 158. A retraction spring 160 biases the plunger toward cam
156 to increase the size of combustion chamber 162. This particular
arrangement is advantageous in that cam driven plunger 150 can put
work back into the camshaft as the plunger retracts. Also, some of
the work that plunger 150 does on the charge can be extracted by
the engine piston, as long as plunger 150 does not retract until
late in the expansion stroke, or after the expansion stroke.
[0162] Alternatively, referring to FIG. 22b, a plunger 170 may be
hydraulically operated by a pressurized supply of fluid, e.g. fuel,
delivered to a chamber 174 by a hydraulic circuit 172 connected to,
for example, a jerk pump or common rail system. FIG. 22c,
illustrates another hydraulically actuated embodiment in which a
plunger 180 is assisted by a spring 182, positioned in a chamber
184 formed adjacent one end of plunger 180, to allow energy to be
stored in the spring. In this system a retaining mechanism, e.g.
hydraulic, electromagnetic or mechanical, (not shown) maintains the
plunger in the unextended position. When the piston is near TDC, a
hydraulic fluid supply system 186 forces plunger 180 down (at this
point the retaining system no longer holds the plunger). This
downward motion is heavily assisted by spring 182. After
combustion, plunger 180 moves back up recompressing spring 182
thereby returning energy to the spring. To optimize this energy
extraction process, the hydraulic chamber 184 bleeds down at a rate
controlled by a valve 188.
[0163] FIG. 22d illustrates yet another embodiment in which a
spring 190, biasing a plunger 192 into the extended position, is
strong enough to overcome the gas pressure in the combustion
chamber before combustion. Near TDC, a bleed down valve 194,
connecting a chamber 196 is opened and the spring 190 pushes
plunger 192 into the extended position in the combustion chamber
162 causing the charge to ignite and the pressure in the combustion
chamber 162 to increase. As a result, plunger 192 is pushed back up
against spring 190. If needed, a high pressure supply 200 supplies
hydraulic fluid to chamber 196 to ensure plunger 192 moves back up
into the retracted position. If the gas pressure is sufficient to
move the plunger back up into the retracted position, a low
pressure hydraulic fill supply 202, including a one-way valve 204,
may be used to fill the chamber 196 below plunger 192.
[0164] The compression ratio may also be varied by providing an
opposed piston engine design having variable phase shifting to
permit the compression ratio to be varied during operation by
changing the phase of rotation between two crankshafts. The opposed
piston engine may be of the type disclosed in U.S. Pat. No.
4,010,611 or of the interconnected cylinder type with variable
phasing as disclosed in U.S. Pat. No. 4,955,328, the entire
contents of both of these references being hereby incorporated by
reference. Alternatively, referring to FIG. 23, the compression
ratio could be varied using a phase shifting mechanism 210
including a conventional differential assembly 211 connected
between an input shaft portion 212 of one of the crankshafts 214,
216 associated with respective pistons 218, 220 and an output shaft
portion 222 of the same crankshaft 214 to permit the portions of
the crankshaft to be rotatively shifted relative to one another.
Crankshafts 214 and 216 are connected via a conventional gear
assembly 223 for transferring power to a driven shaft 225. As shown
in FIG. 24, the differential 211 includes a ring gear 224 mounted
on one end of input shaft portion 212, an arm 226 extending from
ring gear 224 and a gear assembly 227 mounted on the opposing ends
of shafts portions 212, 222. A rotator mechanism 228, including a
pinion gear 230, is operatively connected to ring gear 224 to
rotate the ring gear when a change in the phasing between the
crankshafts is desired. As long as ring gear 224 remains
stationary, shafts portions 212, 222 remain in phase. When ring
gear 224 is rotated by rotating pinion gear 230, arm 226 rotates
causing a change in the phasing between shaft portions 212, 222.
The rotator mechanism 228 would, therefore, be used to adjust the
relative phasing of the input shaft to the output shaft, thereby
adjusting the phasing of the two crankshafts and the compression
ratio. In addition, two crankshafts per cylinder could be used to
eliminate the inherent side thrust imparted by the crankarm in the
single crankshaft design. The effect of the maximum possible
compression ratio on the sensitivity to CR on phasing should be
noted. It might be advantageous to have a geometry where the
pistons interfere with each other at "zero" phasing. Of course,
this set up would operate with non-zero phasing all the time.
[0165] Applicants have determined how the change in phasing of an
opposed piston engine changes the compression ratio. This effort
includes three studies as shown in FIG. 25. In the first, when the
two pistons were in phase, i.e. both pistons reach TDC at the same
time, the compression ratio was 25:1. In the second, when the
pistons were in phase they would come together and just touch at
TDC. With a flat top piston there would be no volume between the
pistons and, assuming no crevice volume, the compression ratio
would become infinite. The third case assumes negative interference
so that the pistons would come in contact while out of phase to
some degree. For this case, the overlap was about 10% of the stroke
causing the pistons to contact at 46.degree. out of phase. Of
course, engine geometry (bore, stroke, connecting rod length) will
also effect CR versus phasing; these values were kept constant in
this study.
[0166] These results indicate that the compression ratio could be
varied over a very large range using an opposed piston arrangement
with variable phasing. Also, the slope of the change in compression
ratio with phasing depends on the amount of clearance or negative
clearance between the pistons at TDC with 0.degree. phasing. Thus,
in a practical application, it would be desirable to strike a
balance between the range of phasing needed to cover the desired
range of compression ratio and the precision with which the phasing
needs to be controlled, i.e. the slope of the curve in FIG. 25
should be optimized. Thus, ideally, the slope of the curve would be
steep enough that the desired range of compression ratio could be
achieved within a limited amount of phasing, and not so steep that
the phasing needs to be too precise.
[0167] Referring to FIG. 26, it is very clear that as the pistons
become more and more out of phase that the compression ratio
decreases. It is also clear that there is very little change in
shape of the cylinder volume versus crank angle curve for phasing
angles less than about 120.degree.. As a result, the variation in
phasing can be used to control compression ratio over a large range
without any affect in the cylinder volume versus crank angle. An
opposed piston system with variable phasing clearly provides the
desired flexibility to achieve a broad range of compression ratio
values.
[0168] The effective compression ratio may be varied with variable
valve timing. Specifically, as shown in Table I, advancing of the
intake valve closing lowers the effective CR while significant
retarding of the IVC also reduces effective CR. However, changing
valve events can have a very large effect on the breathing of an
engine, and thus the air/fuel ratio, in comparison to varying the
geometric compression ratio (assuming that the fuel flow rate is
held constant). The steepest change in airflow with TDC temperature
is when IVC is changed. As IVC becomes earlier, the TDC temperature
is lowered, but airflow is severely restricted possibly undesirably
changing the equivalence ratio. In this case, an increase in boost
accompanying earlier IVC could be used to maintain a constant air
flow rate. Similarly with EVC, as EVC is changed the amount of
residual trapped in cylinder changes, and therefore breathing is
affected. The slope of the IVC line is roughly twice that of EVC
and IVO while varying the geometric compression ratio does not have
an effect on airflow. In terms of changing TDC temperature without
effecting airflow, variable geometric compression ratio appears to
be the most effective of the control features.
[0169] Referring to FIG. 28, changing any of the valve events or
compression ratio has a definite effect on BSFC. In order to get
the best BSFC, increasing compression ratio would be a better
choice than changing exhaust valve closing when a higher
temperature is needed. A very large BSFC penalty would be paid if
EVC is advanced to increase the temperature at TDC. If a lower
temperature is needed, advancing IVC is the best method whereas
varying the geometric compression ratio could also be an option
since it results in only somewhat higher BSFC.
[0170] Applicants have also determined, as expected, that changing
the effective compression ratio has a large effect on peak cylinder
pressure, as shown in FIG. 29. IVC has an almost identical curve as
VCR, confirming the fact that changing IVC really changes the
effective compression ratio. Because the heat release starts
5.degree. ATDC in this case, the cylinder pressure trace appears to
be "double-humped": the first peak at TDC is due to compression;
the second peak after TDC is due to combustion. The appearance of
two slopes of VCR and IVC lines is due to the absolute peak
cylinder pressure occurring either on the combustion hump
(CR<18) or the compression hump (CR>18). In order to increase
the temperature at TDC from the baseline without undesirably
affecting peak cylinder pressure, changing EVC or IVO would be the
best strategy. However, this strategy may result in an undesirable
increase in BSFC (FIG. 28) and may also change the engine breathing
(FIG. 27).
[0171] Applicants have also determined that very high compression
ratios are needed for combustion at low intake temperatures. For
example, it has been found that at intake temperatures of 0, 20,
and 40.degree. F., no combustion occurs when the corresponding
compression ratios are below 35, 33, and 30, respectively. At
warmed up conditions, the desired compression ratio is
approximately 15, which means that a change of approximately 20
compression ratios would be needed to cover these conditions. Due
to the very high compression ratios required under these
conditions, peak cylinder pressures are also high and in some cases
greater than 200 bar. As a result, intake air heaters and/or some
other method of starting in cold conditions may be more practical
than using variable compression ratio alone. Also, maintaining a
lower compression ratio will allow a higher GIMEP to be achieved
before hitting the peak cylinder pressure limit.
[0172] Another method of controlling the temperature is to
introduce water into the intake manifold or directly into the
cylinder. Applicants have shown that when the nitrogen in the
intake air is completely replaced with water, the water will likely
result in a lower flame temperature (205 K lower) due to
dissociation. Also, in applicants' study, the ignition delay
increased slightly (by 0.04 msec) and the peak reaction rate
dropped by about 50%. Also, when water was added into the intake
manifold, e.g. water fumigation, the chemical effect, although
small, is to slightly retard the SOC. However, liquid water
injection into the intake manifold effectively cools the intake
manifold due to the vaporization of the liquid to steam. As a
result, IMT and TDC temperatures are significantly decreased as
shown in FIG. 30. The impact of water injection on temperature at
TDC is mostly due to the decrease in IMT, not due to the change in
the ratio of specific heats. The effect on IMT should be viewed as
an upper limit.
[0173] It should be noted that applicants have shown that PCCI can
be maintained without adverse thermal effects on the piston 14
(FIG. 1a). Even though PCCI combustion can create knock intensity
levels 10-20 times higher than the safe level experienced in
spark-ignited engines, both aluminum and steel pistons do not reach
excessive temperature levels. In applicants' preferred embodiment,
the temperatures resulting from autoignition in PCCI combustion are
much lower than the temperatures experienced in spark-ignited
engines since, in applicants' preferred embodiment, PCCI combustion
operates under such lean conditions. Pressure Control
[0174] The SOC may also be controlled by controlling the pressure
in the combustion chamber. One way of controlling in-cylinder
pressure is to use a compression ratio varying device to vary the
pressure in the combustion chamber. Although varying the
compression ratio ultimately varies both the pressure and
temperature of the charge, the pressure is directly changed. An
increase in the compression ratio will tend to increase the
pressure at TDC, and a decrease in compression ratio will decrease
pressure at TDC. Applicants have shown that increasing the
in-cylinder pressure advances the start of combustion and
decreasing the in-cylinder pressure retards the SOC. Any of the
compression ratio varying devices discussed hereinabove with
respect to temperature control may be used.
[0175] A second way of controlling the in-cylinder pressure is to
vary to the intake manifold, or boost, pressure (IMP). The timing
of the SOC has been shown to be a function of pressure. Applicants
have determined the effects of varying IMP on combustion and engine
operation. The engine conditions for one engine study were 1200
RPM, 355.7K<IMT<357.4K, 0.256<.PHI.<0.263. IMP was
varied. Maintaining these conditions while increasing IMP required
increasing air flow and fuel flow. FIGS. 31a and 31b show that the
duration of heat release decreases as IMP increases both in the
crank angle domain and the time domain. FIG. 31d shows that SOC
occurs earlier as IMP increases. FIG. 31c, showing results from
another study, clearly indicates that increasing the boost pressure
significantly advances the heat release event. FIG. 31e shows that
FSHC emissions decrease as IMP increases, indicating more complete
combustion. FIG. 31f shows that GIMEP increases as IMP increases,
mostly due to the increase in complete combustion, and, to a lesser
extent, more fuel. FIG. 31g shows that gross indicated thermal
efficiency increases as IMP increases, partly due to more complete
combustion. FIG. 31h shows that FSCO emissions decrease as IMP
increases, apparently due to more complete combustion. FIG. 31i
shows that FSNOx emissions are not significantly affected by IMP.
FIG. 31j shows that coefficient of variation (COV) of GIMEP
decreases as IMP increases. FIG. 31k shows that PCP increases as
IMP increases. FIG. 31l shows that estimated noise increases as IMP
increases. FIG. 31m shows that as IMP increases, smaller gains in
GIMEP cause larger rises in PCP. This effect is due to the earlier
SOC that occurs as IMP increases.
[0176] One study varied the pressure at BDC of the compression
stroke. The study was performed using a compression ratio of
14.5:1, an engine speed of 1200 rpm, a BDC compression temperature
of 389 K, an equivalence ratio of 0.3285, and no heat transfer. The
fuel used was propane and the pressure at BDC was varied while all
other parameters were held constant. This study clearly revealed
that as pressure at BDC increases, he SOC becomes earlier. In
addition, for BDC pressures less than 1.75 bar, less than 10% of
the fuel energy was released, while for BDC pressures greater than
P=1.75 bar, virtually all of the fuel energy was released. This
indicates that the combustion is highly sensitive to changes in
pressure. At very low pressures, very little of the fuel burns,
leading to high FSHC emissions. Since none of the fuel is burning
at these low pressures, no carbon monoxide is produced. As the
pressure increases (while maintaining IMT constant), a higher
percentage of the fuel is burned, which leads to decreased
production of carbon monoxide and lower FSHC. Above a certain
critical pressure, all of the fuel burns completely, leading to
extremely low FSHC and FSCO emissions. Also, a very small change in
BDC pressure leads to a very large change in peak cycle temperature
(PCT). The results of the simulation indicate that at low peak
cycle pressures (PCP), the fuel does not burn. Hence, the pressure
peaks at the isentropic compression. As pressure is increased, a
higher percentage of the fuel energy is released, causing the
cylinder pressure to rise above the isentropic compression
pressure. As pressure increases further, all of the fuel energy is
being released and further increases in pressure raise the PCP due
to isentropic effects.
[0177] Clearly, varying IMP can be an effective way of controlling
the SOC and the duration of combustion. Increasing the IMP tends to
advance SOC while decreasing the duration of heat release.
Likewise, decreasing the IMP tends to retard SOC while increasing
the duration of heat release. In a typical application, for a
constant torque condition, the fuel flow rate would remain
virtually constant, and the boost pressure would be increased to
advance the start of combustion or decrease the boost to retard the
start of combustion. For example, an air compressor, a
turbocharger, a supercharger such as driven by an engine power
take-off, or an electrically powered compressor, could be used. For
a given power level, and, therefore, for a given fuel flow rate,
there typically exists a preferred intake pressure and temperature.
At very low loads, it may be desirable to control the intake
manifold pressure with a throttle 53 (FIG. 1a) in the same way that
the intake pressure is controlled on a current production spark
ignited engine. Throttle 53 would also be used when operating a
multi-mode PCCI engine in a spark ignited mode as described
hereinbelow. Of course, a throttle could alternatively be located
at other locations in the intake system, such as in the intake
manifold. Air/Fuel Mixture Autoignition Properties
[0178] Another strategy for controlling the start and duration of
combustion is to vary the air/fuel mixture autoignition properties.
The autoignition properties of the air/fuel mixture may be
controlled by injecting a gas, e.g. air, oxygen, nitrogen, ozone,
carbon dioxide, exhaust gas, etc., into the air or air/fuel mixture
either in the intake system, e.g. preferably in the port using, for
example, injector 42, or in the cylinder directly using, for
example, injector 40, thereby providing control over the start of
combustion and the combustion rate.
[0179] Applicants have examined the effect of adding reactive
species to the air/fuel mixture on the combustion process. One
study was performed using an equivalence ratio of 0.3, a
temperature at BDC of 389 K, pressure at BDC of 3 bar, and propane
as the fuel. The compression ratio was 14.5, and the engine speed
was 1800 RPM. The engine geometry used was for a Cummins C series
engine. The nitrogen, oxygen, and fuel mole fractions were held
constant at 0.771, 0.216, and 0.0123, respectively, for all cases.
The mole fraction for the reactive species added was 0.000411 for
all cases. The reactive species examined were H.sub.2,
H.sub.2O.sub.2, OH, CO, O, HO.sub.2, H, and O.sub.3. FIG. 32 shows
the temperature versus crank angle. Although CO and H.sub.2
advanced the SOC by less than 0.5 crank angle degrees, all other
species significantly advanced the SOC, with O.sub.3 (ozone)
causing the largest change in the SOC. Therefore, small
concentrations of most common radicals will cause significant
changes in the SOC.
[0180] Thus, applicants have determined that the addition of very
small quantities of ozone advances the SOC by significant amounts.
Applicants have also shown that virtually all of the ozone will be
consumed by the combustion process and that the change in the SOC
will diminish as the amount of ozone added increases. Specifically,
FIG. 33 illustrates the effects of additional ozone on advancing
the SOC. The increase in temperature indicates the start of the
combustion event.
[0181] Given the significant effect additional ozone has on the
SOC, ozone can be used in several ways to advantageously control
the combustion in a PCCI engine. First, by adding different amounts
of O.sub.3 to the intake ports, one, several, or all cylinders
could have their SOC adjusted. Second, adding O.sub.3 to the intake
could be used as a cold starting aid for PCCI and diesel engines.
Third, adding O.sub.3 to the exhaust of an engine would allow a
catalyst to light earlier thus possibly significantly reducing cold
start emissions on catalyst-equipped spark ignited engines, diesel
engines and PCCI engines. 03 could be produced "on board" through a
simple electrochemical reaction. Ozone generators are commercially
available. Also, the ignition delay of a diesel engine could be
reduced by adding O.sub.3 to the intake. This would reduce the
premixed burn fraction which would then lower NOx emissions and
reduced noise.
[0182] Applicants have shown that increasing the oxygen
concentration advances the SOC. However, applicants have determined
that oxygen enrichment from 20.7 percent to 21.65 percent will
advance the SOC by less than one crank angle degree, and oxygen
enrichment from 20.7 percent to 23.7 percent will advance the SOC
by less than 1.5 crank angle degrees. Therefore, combustion may be
controlled to a limited degree by modifying the oxygen
concentration of the intake air. This may be done by adding oxygen
(or an oxygen rich gas mixture) to the intake or by selectively
removing nitrogen from the intake air (using a membrane for
example). Applicants have also shown that increasing the percent of
nitrogen in the intake charge from 78.6 percent to 80.6 percent
resulted in the retardation of the SOC by less than 2 crank angle
degrees at 1800 rpm. It was also noted the same percentage increase
of N.sub.2 in the fresh charge lowers the FSNOx from 0.144 to 0.048
grams of NOx per Kg of fuel.
[0183] Another method of varying the effect of oxygen on the
combustion process is to dilute the mixture with EGR. In one study,
an engine EGR system was plumbed from the exhaust manifold to the
compressor inlet. Because the EGR is mixed in upstream of the
aftercooler, and in the present study, the aftercooler exit
temperature was controlled and held fixed, the EGR should not have
significantly effected the temperature at SOC. During this study,
fuel rate and intake manifold temperature were held constant. As
the EGR rate was increased, exhaust manifold pressure decreased,
which in turn decreased air flow on this turbocharged engine. The
fuel rate was held constant, so the fresh equivalence ratio
increased. In spite of the increased equivalence ratio, SOC
retarded as the EGR rate increased, most likely due to the diluent
effect of the EGR. As expected, SOC retarded as the EGR rate
increased. However, as EGR rate increased, CO and HC emissions also
increased. Also, as EGR rate increased, the spread in SOC between
cylinders increased. In a similar study, the SOC was held constant
by adjusting IMT. As the EGR rate was increased, exhaust manifold
pressure decreased, which in turn decreased air flow. The fuel rate
was held constant thus causing the equivalence ratio to increase.
In addition, as the EGR rate increased from about 7 to 13% EGR,
there was a sharp rise in the cylinder to cylinder variation in
SOC. Ultimately, a higher IMT was required to maintain constant SOC
as the EGR rate increased, in spite of an increase in equivalence
ratio. This requirement was due to the diluent effect of increased
EGR on the intake air.
[0184] Another technique for modifying the autoignition properties
of the air/fuel mixture to control SOC and the duration of
combustion is to vary the octane, methane or cetane number of the
charge by, for example, by providing two or more fuels have
different octane, methane or cetane numbers. The fuel supply can be
either selectively switched between the fuels or the fuels can be
mixed. This technique makes it possible to retard or advance the
combustion event. For example, a fuel which tends to autoignite
more readily (lower octane or methane number, or higher cetane
number) could be controllably mixed with a fuel that tends to
autoignite less readily (or a fuel that ignites at a high
temperature and a fuel that ignites at a low temperature could be
used) to enable direct control over the timing of ignition and rate
of combustion by changing the ratio of the fuels that are present
in the combustion chamber during the combustion event. As shown in
FIG. 34, propane, octane and heptane have significantly different
effects on the SOC. The same effect may be achieved by using a fuel
additive, such as a controlled amount of propane, ethane, or other
hydrocarbons, such as engine lubricating oil, that change the
autoignition properties of the fuel to advance or retard the start
of combustion. Of course, any method that changes the fuel's
octane/methane number or the activation energy of the fuel can be
used to advance/retard combustion. Applicants have determined that
there is a significant sensitivity of start of combustion to octane
number. This effect was independent of intake manifold temperature.
Moreover, in one study, the start of combustion was retarded
approximately 7.degree. for an increase in octane number from 80 to
100.
[0185] Achieving dynamic control over individual cylinder
combustion in a multi-cylinder PCCI engine will be critical to
achieving improved combustion. Since many of the gases/fluids
discussed hereinabove, e.g. fuel, ozone, oil, water, etc. have now
been shown to significantly affect the SOC and/or rate of
combustion, these additives can be used to advantageously balance
combustion between the cylinders in a multi-cylinder engine running
on PCCI principles. For example, by injecting a liquid or gas
diluent, such as a less reactive fuel, water, uncooled or cooled
exhaust products, air and/or nitrogen either into the intake air or
directly into the charge in the cylinder, the SOC can be retarded.
Also, by injecting, for example, a more reactive fuel, ozone, oil
and/or oxygen into the charge the SOC can be advanced. FIG. 1b
illustrates one system for balancing combustion between cylinders
of a multi-cylinder engine. The system uses port injection of fuel
on the engine with two types of supplies per cylinder--supply 32 to
inject liquid fuel and supply 34 to inject gaseous fuel. Although
the supplies 32 and 34 are shown feeding into a single passage for
delivery to the intake port, the supplies may include separate
delivery passages connected to the intake port at different
locations. Liquid fuel will decrease intake charge temperature by
the heat of vaporization of the liquid fuel. The temperature at TDC
compression, and therefore SOC, can be controlled by varying the
amount of liquid versus gaseous fuel. Also, it should not matter if
the liquid vaporizes in the port or during compression. The gaseous
and liquid fuel can be the same fuel in different states, e.g.
propane, or different fuels, e.g. gaseous natural gas and liquid
gasoline, such as indolene. It is important that the port injection
system have good separation between cylinders and sequential (timed
to intake event) injection is likely to be required. During
operation, a cylinder that is "going out" would be given more
gaseous fuel and a cylinder that is "too hot" would be given more
liquid. This method can be used to achieve about a 20 degree
temperature difference. One of the supplies could be lubricating
oil or ozone while the other supply could be a fuel having a high
resistance to ignition, e.g. a high octane number, to permit the
SOC to be effectively controlled by varying the amount of oil or
ozone added to the mixture. Also, by using the engine's lubricating
oil supply, or using ozone created by the engine during operation,
an additional supply of fuel/additive can be avoided.
[0186] Equivalence Ratio
[0187] Another control variable that applicants have shown can be
effectively used to control the SOC and combustion duration or heat
release rate is the equivalence ratio (p of the fuel/air mixture.
Equivalence ratio is equal to fuel/air ratio divided by
stoichiometric fuel/air ratio (if<1, fuel deficient; if>1,
fuel excess). Combustion needs to be slowed down in a PCCI engine
because fast combustion leads to high noise, lowered efficiency and
high peak cylinder pressure. If different temperatures and/or
equivalence ratios can be achieved throughout the charge of
air/fuel at or near point of ignition, the resulting rate of
combustion will possibly be slowed down thus advantageously
lengthening the duration of combustion. The equivalence ratio could
be increased by increasing the fuel flow to the cylinder without a
corresponding increase in intake air flow, or by decreasing the
intake air flow. The equivalence ratio could be lowered by
decreasing the fuel flow to the cylinder without a corresponding
decrease in air flow, or increasing the air flow rate. Variations
in the quantity of fuel delivered to a cylinder is varied by
controlling the operation of fuel control valves 39, 41, and/or
fuel injectors 35, 36 in a known manner. The variations in the air
flow rate could be achieved by, for example, variably controlling
compressor 24 to vary boost pressure.
[0188] To test the lower limit for equivalence ratio, applicants
conducted engine studies to determine whether acceptable PCCI
combustion could be obtained with an extremely lean mixture. The
results indicate that very stable combustion can be achieved at an
extremely lean equivalence ratio of 0.05 while obtaining a heat
release duration of approximately 30 degrees. Also, as shown in
FIGS. 35 and 36, the results indicated that the start of combustion
advances and the apparent heat release duration decreases as
equivalence ratio increases, i.e. air/fuel mixture becomes richer.
Applicants have clearly shown, as indicated in FIG. 37, where
cylinder temperature increases indicate the heat release event.
Moreover, referring to FIG. 38, the apparent heat release duration
becomes longer as equivalence ratio decreases, i.e. air/fuel
mixture becomes leaner. Also, applicant have shown that for a four
stroke engine that both peak cylinder pressure and GIMEP increase
as equivalence ratio becomes richer. With respect to a two-stroke
engine, applicant have determined that as equivalence ratio
increases, GIMEP increases.
[0189] Studies were also conducted to investigate whether the
equivalence ratio affects the amount of fuel burned in PCCI
combustion. The results indicated that as equivalence ratio becomes
richer, the percentage of fuel energy showing up as apparent heat
released increases at first and then levels off near 80%. This
number can never reach 100% because of heat transfer. With respect
to emissions, as equivalence ratio becomes richer, fuel specific
hydrocarbon emissions decrease. In addition, as equivalence ratio
became richer, average noise levels increased, and GIMEP increased.
As equivalence ratio becomes richer, the average knock intensity
increases. As equivalence ratio became richer, the cycle-to-cycle
combustion variation, as measured by the coefficient of variation
(COV) of GIMEP, generally decreased. In fact, the COV's of GIMEP,
for the conditions of the study, stayed below the combustion
stability limit (in this case defined as 5%), where a COV above the
limit indicates unacceptable stability.
[0190] Studies were performed to determine the effect that
variations in equivalence ratio have on thermal efficiency in PCCI
combustion. An equivalence ratio study was performed while matching
the following parameters: speed, IMT, IMP, engine oil temperature,
and engine water temperature. Equivalence ratio was increased by
holding air flow constant and increasing fuel flow to the engine.
As fuel flow increased and equivalence ratio became richer, gross
indicated thermal efficiency increased at first and finally leveled
off. Engine work output increased with respect to increased fuel
flow as more fuel was burned. At the leaner equivalence ratios, a
significant amount of fuel is left unburned. At the richer
equivalence ratios, the percentage of fuel that is being burned
levels off as noted hereinabove, and the gross indicated thermal
efficiency levels off because the increase in engine output is
being offset by the additional fuel input.
[0191] In addition, an engine study was conducted with the engine
cycle running from bottom dead center of the compression stroke to
BDC of the expansion stroke. The study was conducted using a
compression ratio of 14.5:1, an engine speed of 1200 RPM, a BDC
compression temperature of 389 K, pressure at BDC of 4.31 bar, and
no heat transfer. The fuel used was propane. The equivalence ratio
was varied while all other parameters were held constant. It was
discovered that the percent of energy released slowly tapered off
as the equivalence ratio drops below 0.15. This data indicates that
for a given temperature and pressure, there is a lower limit to the
equivalence ratio of a mixture that will burn completely. Also, it
was shown that FSCO emissions are very high at equivalence ratios
below 0.15. This data indicates that only a small amount of the
fuel burns to completion at these low equivalence ratios for this
temperature and pressure. In addition, the FSHC decrease slightly
as the equivalence ratio is varied from 0.05 to 0.4. Thus, most of
the fuel reacts regardless of the equivalence ratio. It was also
shown that SOC occurs earlier as equivalence ratio increases. The
study showed that peak cylinder temperature gradually increases as
equivalence ratio is increased showing the increased amount of
energy available to be released. Peak cylinder pressure (PCP)
gradually increases as equivalence ratio is increased showing the
increased amount of energy available to be released. At equivalence
ratios greater than or equal to 0.18, virtually all of the
available fuel energy is released, leading to a nearly linear
increase in PCP as equivalence ratio increases.
[0192] Applicants have determined that it may be possible, although
not necessarily desirable, to maintain PCCI combustion at very rich
equivalence ratios, e.g. .5, if IMP and IMT are sufficiently low to
prevent the peak cylinder pressure limit from being exceeded. It
will be difficult to start an engine at the low boost and IMT
levels needed for maintaining low cylinder pressures at such rich
equivalence ratios. The very advanced heat release, loud knock, and
combustion roughness make running at this condition undesirable. A
lower CR for retarding SOC may improve these aspects.
[0193] Also, by varying the level of charge stratification, the
temperature and equivalence ratio distribution can be altered to
permit control of the combustion rate and/or the start of
combustion. An auxiliary combustion chamber concept may be a
mechanism for achieving the desired stratification, thereby
enabling better control over the start of combustion. For example,
conventional auxiliary combustion chamber designs typically used on
small engines having indirect injection (IDI), and large spark
ignited engines using natural gas fuel, could be used.
[0194] In order to operate under the desired lean conditions for
optimal PCCI combustion, substantial air flow must be provided to
the intake manifold. A turbocharger could provide the needed air
flow for a multi-cylinder PCCI engine. Applicants' original target
was to reach an equivalence ratio of 0.40 or leaner. Referring to
FIG. 39, applicants have shown that operating at leaner than an
equivalence ratio of 0.29 would violate the compressor pressure
ratio limit of the available turbocharger. Applicant determined
that turbine pressure ratios are very high at lean equivalence
ratios. As a result, the exhaust manifold pressure is very high
which causes a large BSFC penalty. Because of the relatively cool
exhaust temperatures produced by PCCI combustion, very small
turbine cases are needed which result in high exhaust manifold
pressures.
[0195] Applicants have determined that it would be desirable to
operate under slightly leaner conditions than the original target.
At an equivalence ratio less than 0.4, a smaller turbine casing was
used to decrease the compressor pressure ratio and exhaust manifold
pressure ratio, but a high BSFC penalty is paid, as shown in FIG.
40. FIGS. 41 and 42 illustrate the higher PMEP losses with the
smaller turbine casing and the higher BSFC. Also, with the smaller
turbine casing, the rotor speed is much higher and, in fact, near
the limit on rotor speed as seen in FIG. 43 (rotor speed limit
120-125 k rpm range). Applicants discovered that there is a lower
limit on the size of the turbine casing used due to the losses
incurred with the high back pressure and with reaching the rotor
speed limit.
[0196] In order to avoid this problem with the high back pressure
and rotor speed limiting airflow, one possible solution is to use a
mechanically driven supercharger in conjunction with a
turbocharger. The supercharger would be upstream of the compressor
so that the turbine bears less of a burden for producing boost.
Some BSFC penalty would be incurred for the shaft work absorbed by
the supercharger; however, the BSFC penalty is less than the very
high penalty incurred with the very small turbine. Because the
supercharger is driven mechanically from the shaft, there should be
no trouble getting the desired air flow. The turbine then can be
sized somewhat larger, and should not approach the speed limit and
should not have extremely high back pressure.
[0197] Applicants have also determined the effect of engine speed
on SOC. The time of autoignition depends on the temperature and
pressure histories. By changing the engine speed, these histories
are changed. It is possible to advance the combustion event by
reducing the engine speed, and to retard the combustion event by
increasing the engine speed. Specifically, a 75% increase in engine
speed, from 1000 to 1750 resulted in a 1.5% increase in the start
of combustion pressure and a 2.8% increase in the start of
combustion temperature. In addition, a 75% increase in engine speed
decreased the heat release rate duration by 0.81 ms (only a 23%
decrease) which corresponds to an increase in heat release duration
of 1.7 crank angle degrees (only an 8% increase). Given this
minimal impact of engine speed on the SOC and heat release, and the
inability to effectively vary engine speed in many practical engine
applications, engine speed is not viewed as an effective combustion
control variable. However, one example where engine speed could be
used to provide some control over combustion is in an application
where the engine drives a generator or alternator.
[0198] As discussed hereinabove, the foregoing control variables
are used to control the SOC and the duration of combustion to
achieve optimum PCCI combustion. One key consequence of efficient,
optimum combustion is reduced emissions. Applicants have shown that
a PCCI engine can achieve NOx emission levels that are well below
any other NOx emission levels ever demonstrated by applicants using
diesel and natural gas engines, and well below future emissions
standards as shown in FIG. 44. The use of propane as the fuel
resulted in the lowest NOx emissions relative to diesel fuel and
gasoline.
[0199] Applicants have also determined the effect of the control
variables and other factors on emissions of a PCCI engine. Engine
speed has little effect on the quantity of NOx emissions. Although
a 75% increase in engine speed approximately tripled the FSNOx, the
levels of NOx emissions produced were still extremely low. Also, as
equivalence ratio becomes richer, fuel specific NOx generally
increases, but still remains at extremely low levels. Referring to
FIG. 45, applicants have determined that engine speed appears to
affect FSCO and FSHC emissions more significantly. As shown, below
a certain critical speed, virtually all of the fuel burns, FSHC are
low and FSCO is low. Just above the critical speed, the fuel
partially burns, resulting in higher FSCO emissions. As engine
speed continues to increase, the percentage of the fuel that burns
continues to drop, resulting in lower FSCO emissions. These
emissions also vary as the temperature at BDC varies. Referring to
FIG. 46, at very low temperatures, very little of the fuel burns,
leading to high FSHC emissions. Since none of the fuel is burning
at these low temperatures, no carbon monoxide is produced. As the
temperature increases, a higher percentage of the fuel is burned,
which leads to increased production of carbon monoxide and lower
FSHC. Finally, above a certain critical temperature, all of the
fuel burns completely, leading to extremely low FSHC and FSCO
emissions. In fact, as shown in FIG. 47, applicants have shown that
all data points with end of combustion flame temperatures above
1600 K had acceptable CO emissions. It has been shown that both
high temperature and the hydroxyl radical (OH) are critical for the
desired oxidation of CO. Importantly, as equivalence ratio becomes
richer, fuel specific CO decreases, while the concentration of
CO.sub.2 in the exhaust increases. In one study, all points taken
with an equivalence ratio <0.2 had CO emissions above the EPA CO
limit.
[0200] As equivalence ratio becomes richer, fuel specific HC
decreases. Clearly, unburned hydrocarbons (UHC) are a key concern
for PCCI engines since reducing unburned hydrocarbons is essential
to the commercial feasibility of a PCCI engine. Applicants have
determined that UHC, and CO, is formed in small crevices positioned
in the components forming the combustion chamber, i.e above the top
ring of the piston between the piston and the liner; between the
cylinder head and the cylinder liner; and around the components
mounted in the cylinder head. The crevices prevent the volume of
mixture in the crevice from reaching a sufficiently high
temperature necessary for burning of the HC and oxidation of the
CO. For example, applicants have shown that similar pistons with
different crevice volumes have different UHC levels. The present
PCCI engine may include one of several designs to minimize UHC. The
present crevice minimizing designs result in a low crevice volume;
keep the fuel away from any existing crevices; or cause the mixture
in the crevice volume to burn appropriately. The designs shown in
FIGS. 48a and 48b are most easily implemented in a ported two
stroke cycle engine. Referring to FIG. 48a, in one embodiment, the
engine has a single piece head and liner combination 300, although
a two-piece system could be used. Just above the top ring 302 (at
TDC), the bore 304 increases to eliminate the crevice around the
top land 306 of piston 308. There are no crevices in the cylinder
head, as it is a single piece without valves, gaskets, etc.
[0201] Referring to FIG. 48b, a second embodiment of the crevice
minimizing design may similarly include a one-piece head and liner
310. However, in this embodiment, the piston 312 has a very
aggressive cutback 314 forming the top land to enlarge the crevice
volume 316 between the top land and the liner. The crevice volume
316 is now so large that it will no longer quench combustion in
this area thus allowing fuel in this volume to burn resulting in
reduced UHC. FIG. 49 illustrates yet another embodiment including a
cup or chamber 320 formed in the cylinder head 322 of the engine.
The fuel injector 324 is positioned to inject fuel directly into
cup 320 early in the compression stroke. Because air is pushed into
cup 320, the fuel does not exit the cup. After compression ignition
occurs, the products can pass through the relatively large passage
or throat 326 between the cup 320 and the main cylinder 328. The
fuel is well mixed because of the turbulence of the air entering
the cup. Because there are no crevices in the cup and because the
fuel does not leave the cup until after combustion is completed,
UHC are extremely low. The cup could easily be coated with a
thermal barrier coating to reduce heat losses.
[0202] FIGS. 50a and 50b illustrate a cup design for a four stroke
engine. The exhaust and intake valves 330 are arranged around a cup
332 in the head 334. Cup 332 may be positioned directly above the
combustion chamber 336 as shown in FIG. 50a or offset to allow more
room for the valves 330 as shown in FIG. 50b. Another possibility
is to include a small auxiliary valve in the cup to allow the
products to exit the cup more efficiently. This valve could open
after the main exhaust valve opens so that the auxiliary exhaust
valve in the cup would not open against a high pressure. In this
case, the auxiliary exhaust valve could be electronically operated.
The timing of opening and closing of this valve could be used to
vary the residual mass fraction which would allow control over the
SOC using this auxiliary valve. Also, an opposed piston engine, as
discussed hereinabove, may be used to substantially reduce the
crevice volume by avoiding a cylinder head and the associated
crevices.
[0203] Now referring to FIG. 1a, another embodiment of the present
invention for reducing emissions is disclosed. Specifically, this
embodiment controls UHC and CO by heating the upper portion of the
cylinder liner 49 to cause oxidation of the charge in the crevices.
A heater 51 is incorporated into the upper part of the liner. The
heater could be any type of heater capable of effectively producing
heat, such as an electrical resistance heater. The heater heats the
gas in the crevice above the top ring when the piston nears TDC.
This heating will cause the gas to be less dense resulting in a
smaller mass of charge remaining in the crevice. The charge leaving
the crevice will be at a higher temperature due to the heating thus
increasing the tendency of the charge to react and form CO.sub.2
instead of CO and UHC.
[0204] Also, a glow plug may be used to heat the combustion gases
to reduce emissions by enabling a larger portion of the crevice
volume to be burned. It has been determined by applicants that a
glow plug will have only a slight impact on the SOC. Since SOC
changes only slightly when the glow plug is turned on, it does not
appear that the glow plug is initiating combustion. It is more
likely that when the glow plug, which was located in a spacer
plate, is turned on, it gradually warms up the gas in the crevice
volume. This increase in temperature is sufficient to speed up the
onset of rapid combustion, and burn more of the fuel than would
have been burned without the glow plug on, resulting in a slight
increase in GIMEP.
[0205] The present engine, as shown in FIGS. 1a and 1b, may also be
operated as a multi-mode engine which changes modes of operation
based on the operating conditions or needs of the particular
application. For instance, the engine may be operated on diesel
fuel only as a conventional diesel engine, as a modified diesel
engine with diesel fuel being injected earlier in the compression
event than the conventional diesel engine, as a spark-ignited
engine using spark plug 56 (FIG. 1a) or as a PCCI engine. This type
of spark ignited/compression ignited DI (direct injection) variable
compression ratio engine provides a combination of low emissions,
high power density, and ease of starting.
[0206] This engine operates in the following different modes
depending on the current operating conditions/needs of the
engine.
[0207] 1) Medium compression ratio (.about.10:1), early injection
(fuel injected during intake stroke or very early in the
compression stroke) nearly homogeneous:
[0208] a) Overall lean mixture, spark ignited--allows low NOx, high
brake mean effective pressure (BMEP) operation, as well as medium
BMEP operation.
[0209] b) Stoichiometric mixture, spark ignited--allows high NOx
high BMEP transient operation, as well as low NOx operation with a
3 way catalyst.
[0210] 2) High compression ratio (.about.15:1), early injection,
nearly homogeneous, very lean (.phi.<0.5), compression
ignition--allows very low NOx medium BMEP and low BMEP
operation.
[0211] 3) High compression ratio (.about.15:1), late injection,
stratified charge:
[0212] a) Spark ignited--allows medium NOx, medium BMEP unthrottled
operation, and low BMEP operation.
[0213] b) Compression ignition--allows medium NOx medium and low
BMEP unthrottled operation.
[0214] 4) Low compression ratio (.about.8:1), early injection,
nearly homogeneous, spark ignited:
[0215] a) Lean burn--allows very high BMEP operation.
[0216] b) Stoichiometric'allows very high BMEP operation.
[0217] 5) Medium compression ratio (.about.10:1), late injection,
stratified charge, spark ignited--allows medium NOx, medium and low
BMEP, and high BMEP operation.
[0218] 6) Very high compression ratio (.about.20:1), lean burn,
early injection, nearly homogeneous, compression ignition--allows
the engine to be started in PCCI mode.
[0219] The key here is to take full advantage of the variable
compression ratio. Starting of the engine can be achieved with
spark ignition at a lower compression ratio and then transitioning
to high compression ratio, lean PCCI operation for low NOx. For
less severe (e.g. not as cold) conditions, engine starting could be
achieved directly with very high compression ratio PCCI operation.
At low and medium loads, the engine can operate in a PCCI mode as
the compression ratio is adjusted to keep the start of combustion
near the optimum crank angle. For high load requirements, the
air/fuel ratio can be enriched, compression ratio lowered, and the
engine can be spark ignited. In order to handle sudden transients,
the engine may go into one of the late injection modes where richer
air/fuel ratios are possible without engine damage.
[0220] In the multi-mode engine, ECU 20 (FIG. 1) functions with a
control strategy for controlling the various controlled features of
the engine to effectively switch between, and operate in, the
different modes in order to achieve a variety of objectives. For
example, the multi-mode engine achieves low NOx emissions in the
PCCI mode while enhancing startability by providing a high
compression ratio or spark ignition. In addition, the engine can
achieve a high cylinder pressure at high BMEP by switching to a
lower compression ratio spark-ignited mode. The multi-mode engine
also permits stable combustion to occur after switching to late
injection which results in a stratified charge by rapidly adjusting
the compression ratio. Also, fuel consumption can be effectively
controlled using high compression, PCCI operation and stratified
charge operation requiring no throttling which have excellent
thermal efficiency. This operation also improves transient response
by going from PCCI to late injection, stratified charge to suddenly
enrich the mixture. This multi-mode engine can also effectively
minimize knock, and therefore knock damage, by effectively
operating at lean PCCI or stratified charge or low compression
ratio, lean burn or stoichiometric conditions. Of course, the
engine operates to effectively control the start of combustion
during PCCI operation by varying, for example, as discussed
hereinabove, the temperature and/or the equivalence ratio and/or
the pressure and/or the air/fuel mixture autoignition properties.
This engine could run on a variety of fuels like gasoline or diesel
fuel.
[0221] Another operating mode is dual injection in which an early
injection is used to create a lean charge for PCCI operation. A
second, late injection then adds a small amount of stratified fuel
which can be either spark or compression ignited to help ignite the
remaining fuel. This mode is similar to diesel pilot operation but
would only be used during transition between the different modes of
operation or during engine starting. Applicants have studied the
effects of diesel pilot operation on emissions. FIG. 51 shows a
comparison of the normalized heat release rate versus crank angle
for the three different diesel pilot injection quantities into a
PCCI engine operating on propane. A micro-pilot injection of 0.1%
resulted in good heat release placement with no measurable increase
in FSNOx. A diesel pilot of an amount estimated to supply 3.6% of
the fuel energy resulted in a heat release curve having
substantially the same shape as the previous case. The SOC is
slightly more advanced than that of the 0.1% case despite a lower
IMT and constant equivalence ratio. Also, FSNOx emissions have
increased over the 0.1% case from zero to 3.9 g/kg. The final
curves illustrates the heat release for a case with 18% of the fuel
energy coming from the diesel pilot. The heat release rate curve is
shaped the same as the classic diesel heat release rate curve with
a premixed burn spike and a diffusion burn region. Also, the FSNOx
(15.3 g/kg) and FSHC (478 g/kg) are significantly higher than in
the cases with smaller diesel pilots.
[0222] With respect to diesel pilot injection, as the percentage of
fuel energy from the pilot increases, the start of combustion (SOC)
becomes more advanced, despite the lowering of IMT and a constant
equivalence ratio. This earlier SOC is caused by the diesel fuel
autoigniting earlier than the propane. As the percentage of pilot
increases, the heat released by the pilot during the compression
stroke increases, leading to higher temperatures earlier in the
cycle. Higher temperatures increase the chemical reaction rates of
reactions involving propane, leading to earlier autoignition of the
propane. Therefore, extremely low NOx levels and good heat release
placement can be achieved when using a very small diesel pilot or
micropilot, preferably less than 4% of the total fuel energy.
[0223] Applicants have also studied the control of noise associated
with PCCI combustion. Level of noise generated by PCCI combustion
is related to the knock intensity. Thus, as knock intensity is
decreased, noises decreases. As shown in FIGS. 4a, 4c and 6,
lowering cylinder pressure, for example, by retarding the SOC,
substantially decreases the knock intensity and, therefore, noise.
The present engine and control system permits continuous PCCI
combustion with minimal noise by avoiding excessive peak cylinder
pressures while maintaining the required cylinder pressure
necessary for efficient, low emission PCCI combustion and the
desired power output.
[0224] The control system of the present invention operates to
actively and variably control the mixture's temperature, pressure,
autoignition characteristic and equivalence ratio to ensure that
the combustion event occurs between 20 crank angle degrees BTDC and
35 crank angle degrees ATDC. The control system achieves this
function by using combustion sensor 16, e.g. pressure sensor, to
signal the start of combustion or the location of the heat release
event for each cycle. Also, ECU 20, which receives the signals from
sensor 16, determines whether the SOC is occurring within a
predetermined crank angle range and determines whether the duration
of combustion is within a predetermined desired crank angle range.
One conventional way for the ECU to determine the optimum SOC would
be to use a look-up table. If the SOC and/or the duration of
combustion are outside the predetermined crank angle range, then
ECU 20 determines the appropriate control variable or variables to
adjust, and generates and sends the appropriate signal 22 to the
chosen control mechanism or mechanisms, e.g. air cooler 28, heater
30, glow plug 44, fuel control valves 39, 41, variable compression
ratio device 38, etc., as discussed hereinabove. The control
variables are varied as required to maintain the timing of the
start of PCCI combustion preferably between 20 crank angle degrees
BTDC and 10 crank angle degrees ATDC, and to maintain the duration
of combustion in the range of 5-30 crank angle degrees.
[0225] Applicants have determined that, in order to initiate and
maintain PCCI combustion upon start-up in a cold engine, the
conditions in the cylinders, e.g. temperature and/or pressure, must
be actively influenced. For example, the intake air temperature
could be raised using heater 30 and/or a glow plug 44, and/or the
in-cylinder walls heated using a cylinder wall heater 51 and/or an
engine coolant/lubricating oil heater. Also, the in-cylinder
pressure and temperature could be increased using variable
compression ratio device 38. Another effective control feature for
enhancing startability is to add small amounts of ozone to the
intake air supply using injector 42, or into the cylinder using
injector 40. Alternatively, or additionally, one of the fuel
supplies could have a high autoignition property, e.g. low octane
number. Also, the engine may be operated in a non-PCCI, for
example, as a spark-ignition, dual fuel or diesel engine, during
starting of the engine. One or a combination of these controls are
varied, in accordance with the principles discussed hereinabove
with respect to each control feature, to cause PCCI combustion to
occur. As the engine starts, the ECU will monitor the start of
combustion and duration of combustion by receiving combustion data,
e.g. pressure signals, from sensor 16 throughout engine
operation.
[0226] Once the engine is warmed up, the SOC and duration of
combustion will vary due to the sensitivity of PCCI combustion to
the temperature and pressure history. Small variations in the
numerous factors affecting temperature and pressure history, such
as combustion chamber wall temperature, IMT, equivalence ratio,
IMP, etc. result in significant variation in the SOC and the
duration of combustion. During operation, the control system of the
present invention will vary one or more of the control variables,
that is, temperature, pressure, air/fuel mixture autoignition
properties and/or equivalence ratio, using the various control
mechanisms discussed hereinabove, in such a manner to maintain the
SOC and duration of combustion in the desired ranges. For example,
applicants have shown that SOC can be advanced from 5.degree. ATDC
to 0.5.degree. BTDC by increasing the IMT from 184.degree. F. to
195.degree. F., as shown in FIG. 8. Applicants have also shown that
increasing CR, which raises the in-cylinder temperatures, can be
used to advance SOC. For example, FIG. 21 shows that increasing CR
from 14:1 to 22:1 advanced the SOC from 2.degree. ATDC to
13.degree. BTDC when the equivalence ratio was 0.35 and IMT was 380
K. In addition, applicants have shown that increasing RMF to raise
the temperature of the charge also can be used to advance SOC. When
RMF was increased by adjusting exhaust valve lash from 0.025" to
0.046", the SOC advanced from 6.4.degree. ATDC to 1.7.degree. ATDC,
as shown in FIG. 16. Heat transfer to the charge, whether from
active heating elements or hot surfaces such as the combustion
chamber walls, has also been shown to advance SOC. Applicants have
also shown that, with a glow plug installed in the combustion
chamber, the SOC retarded from 0.6.degree. ATDC to 1.5.degree. ATDC
after the glow plug had been turned off, as shown in FIG. 11.
Applicants have determined, as shown in FIG. 9, that increasing
combustion chamber wall temperatures from 400K to 933K can advance
the SOC from 7.degree. ATDC to 14.degree. BTDC.
[0227] With respect to pressure control, increasing IMP serves to
advance the SOC. FIG. 31c, for example, shows that increasing IMP
on the single cylinder engine from 52 psia to 57 psia caused the
SOC to advance from 3.7.degree. ATDC to 1.50 BTDC. Any method of
affecting cylinder pressure, such as varying compression ratio or
changing valve timing, both illustrated above, can be used to
control SOC.
[0228] With respect to equivalence ratio, applicants have
determined, as shown in FIG. 38, show that increasing equivalence
ratio from 0.30 to 0.33 by increasing fuel flow to the engine
advanced the SOC from 5.5.degree. ATDC to 2.0.degree. ATDC. Also,
varying the autoignition properties of the air/fuel mixture by the
addition of reactive species or even diluent can affect SOC.
Applicants have shown that for the case shown in FIG. 33,
increasing the amount of ozone added to the charge from 0 to 36
g/kg of fuel had the effect of advancing SOC from 1.degree. ATDC to
12.5.degree. BTDC. In one study where diesel fuel was used in a
pilot injection to initiate SOC in an air-propane mixture, the
amount of pilot used affected SOC. For example, when pilot quantity
was increased from approximately 0.1% to 18% of the total fuel
energy, the SOC advanced from 2.degree. ATDC to 10.degree. BTDC. In
one study, EGR was used as a diluent to retard SOC while holding
IMT constant with an aftercooler. As shown in FIG. 17, when EGR
rate was increased from 2.9% to 8.0%, the SOC retarded from
1.2.degree. ATDC to 4.2.degree. ATDC. Applicants have shown that
increasing the air/fuel mixture's resistance to autoignition by
increasing octane number, for example, can be used to retard SOC.
Also, applicants have shown that when octane number was increased
from 80 to 100, the SOC retarded from 14.degree. BTDC to 7.degree.
BTDC for a case where IMT plus reheat was 311K.
[0229] Of course, any of these control variables could be adjusted
in the opposite direction from the above examples to achieve the
opposite effect on SOC if necessary. For example, rather than
increasing IMT to advance SOC, IMT could be decreased to retard
SOC. Also, the magnitudes of such variations would be increased or
decreased as necessary to maintain the desired SOC.
[0230] Applicants have shown that the combustion or heat release
duration can be affected by varying different parameters. As SOC is
retarded, the heat release duration increases. For example, FIG. 8
shows that as SOC is retarded, by reducing IMT from 195 degrees F
to 184 degrees F, the duration increases from approximately 6
degrees to approximately 24 degrees. Similarly, increasing the
equivalence ratio decreases the heat release duration. Applicants
also believe that increasing the degree of temperature and
equivalence ratio stratification of the charge increases the heat
release duration. However, given the difficulty of measuring the
degree of temperature or equivalence ratio stratification more work
is needed to quantify the level of stratification.
[0231] Of course, given the relationship between SOC and duration,
any control strategy that retards SOC should also increase the
duration. By maintaining the SOC and the duration of combustion in
the desired ranges while controlling the equivalence ratio to
ensure lean burn conditions, the control system minimizes NOx
emissions. Also, the present engine design, also reduces UHC and CO
emissions by minimizing the crevices in the cylinder thereby
minimizing the unburned gases as shown in FIGS. 48a-50b.
[0232] During operation, balancing the combustion processes between
the cylinders of the engine of FIG. 1b can be accomplished by
varying any of the control variables used to control the SOC, as
discussed hereinabove. The ECU 20 compares the SOC and duration of
combustion data provided by sensor 16 for each cylinder. When the
data indicates that the SOC and/or duration of combustion of one or
more cylinders is occurring outside a predetermined crank angle
range, the ECU will determine the appropriate control variable or
variables most effective for the given operating conditions and
generates a control signal for controlling the control variable to
cause the SOC and/or duration of combustion to adjust so as to fall
within the desired range. Applicants have determined that cylinder
balancing is best achieved by controlling equivalence ratio, adding
ozone to the mixture, controlling individual heaters associated
with each cylinder intake port, varying compression ratio using
device 38 or variable valve timing, adding oil via pilot injection
or port fuel injection, port injection of water and/or any of the
methods discussed hereinabove for varying EGR or RMF. Any of these
or other forms of combustion control could be used alone, or in a
variety of combinations, to enhance combustion balancing control.
For example, the combustion control provided by the multiple
fuel/additive system described hereinabove could be enhanced by
providing variable valve timing and/or combustion chamber surface
temperature cooling, e.g. engine coolant, or piston cooling nozzle
control. Also, one or more glow plugs 44 (FIG. 1a) may be used as
an inexpensive, easy method of achieving at least partial control
over combustion balancing between the cylinders. It may also be
possible to control the EGR rate for each cylinder in order to
balance combustion quality.
INDUSTRIAL APPLICABILITY
[0233] The present PCCI engine and control system may be used in
any stationary or nonstationary power plant, including any
automotive, industrial, marine or military application. The present
PCCI engine and control system is especially advantageous in any
power generation application where low emissions are desirable.
* * * * *