U.S. patent application number 16/241161 was filed with the patent office on 2019-05-09 for gasoline particulate reduction using optimized port fuel injection plus direct injection.
The applicant listed for this patent is Ethanol Boosting Systems, LLC. Invention is credited to Leslie Bromberg, Daniel R. Cohn.
Application Number | 20190136790 16/241161 |
Document ID | / |
Family ID | 61687706 |
Filed Date | 2019-05-09 |
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United States Patent
Application |
20190136790 |
Kind Code |
A1 |
Bromberg; Leslie ; et
al. |
May 9, 2019 |
Gasoline Particulate Reduction Using Optimized Port Fuel Injection
Plus Direct Injection
Abstract
An optimized port plus direct injection (PFI+DI) fueling system
for reducing DI-generated particulates from a spark ignition
gasoline engine is disclosed. It uses information from a
computational model that includes piston wetting. Means for DI
particulate reduction include control of DI timing and duration as
a function of various parameters. Illustrative computational
results for decreasing particulates in various drive cycles are
presented. These calculations illustrate large potential
particulate reductions (e.g. 95%) that can be obtained relative to
DI operation alone. The optimized PFI+DI system could provide DI
generated particulate reduction, efficiency and cost advantages
relative to operation of a DI alone engine with a gasoline
particulate filter (GPF). Alternatively, it could be used in
combination with a GPF to ease GPF operation requirements and
provide additional particulate reduction. Techniques for reducing
piston wetting generation of particles from use of DI alone are
also described.
Inventors: |
Bromberg; Leslie; (Sharon,
MA) ; Cohn; Daniel R.; (Cambridge, MA) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Ethanol Boosting Systems, LLC |
Cambridge |
MA |
US |
|
|
Family ID: |
61687706 |
Appl. No.: |
16/241161 |
Filed: |
January 7, 2019 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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15713997 |
Sep 25, 2017 |
10227945 |
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16241161 |
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62399755 |
Sep 26, 2016 |
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62432140 |
Dec 9, 2016 |
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62477587 |
Mar 28, 2017 |
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62531935 |
Jul 13, 2017 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F01N 3/035 20130101;
F02D 35/027 20130101; Y02T 10/44 20130101; F02D 37/02 20130101;
F02D 2041/389 20130101; Y02T 10/40 20130101; F02D 41/1498 20130101;
F02D 41/401 20130101; F02D 2200/021 20130101; F02D 2200/08
20130101; F02D 41/402 20130101; F01N 2250/02 20130101; F02D
2200/1002 20130101; F02D 41/1448 20130101; F02D 41/2422 20130101;
F02D 41/3094 20130101 |
International
Class: |
F02D 41/40 20060101
F02D041/40; F02D 41/24 20060101 F02D041/24; F02D 37/02 20060101
F02D037/02; F01N 3/035 20060101 F01N003/035; F02D 41/14 20060101
F02D041/14; F02D 35/02 20060101 F02D035/02; F02D 41/30 20060101
F02D041/30 |
Claims
1. A fuel management system for a spark ignition engine that is
fueled with gasoline and uses port fuel injection and direct
injection; wherein the use of port fuel injection is such as to
reduce particulate emission by reducing the fraction of fuel in the
engine that is directly injected; and wherein during at least part
of a drive cycle the fraction of fuel that is provided by direct
injection is increased as torque is increased and wherein it is
matched to that needed for preventing knock as engine torque and
engine speed change; and wherein reduction of particulate emission
is obtained by use of port fuel injection to increase the fraction
of directly injected fuel that is introduced between a selected
first crank angle degree on the intake stroke and a selected second
crank angle degree on the compression stroke; and wherein
particulate emission decreases when the directly injected fuel is
introduced after the first crank angle and before the second crank
angle; and wherein at least one of the selected first and second
crank angles are determined by at least one input to a control
system.
2. The spark ignition engine of claim 1, wherein the selected first
crank angle is determined by at least one input to the control
system.
3. The spark ignition engine of claim 1, wherein the selected
second crank angle is determined by at least one input to the
control system.
4. The spark ignition engine of claim 1, wherein the selected first
crank angle is determined by temperature information that is used
by the control system.
5. The spark ignition engine of claim 1, wherein the selected
second crank angle is varied based on temperature information that
is used by the control system.
6. The spark ignition engine of claim 1, wherein the direct
injection is initiated within the crank angle range between the
first and second crank angle; and wherein the initiation is within
the first 25% of time that the piston is between the first and
second crank angle degrees.
7. The fuel management system of claim 1, wherein the selected
first and/or second crank angles are based on a look up table that
uses experimental data about particulate emissions.
8. The spark ignition engine of claim 1, wherein information from a
model for piston wetting is used.
9. The spark ignition engine of claim 1, wherein for at least part
of the first 100 seconds of engine operation, spark retard is used
to enable operation with port fuel injection alone and wherein
spark retard is less than 10 crank angle degrees.
10. The spark ignition engine of claim 1, wherein during at least
some time within the first 100 seconds of engine operation, the
fraction of fuel that is directly injected is matched to that
needed to prevent knock as torque varies and wherein the engine is
operated using port fuel injection alone for at least part of the
first 100 seconds of engine operation.
11. The spark ignition engine of claim 1 wherein during at least
some time within the first 100 seconds of engine operation, the
fraction of fuel that is directly injected is matched to that
needed to prevent knock as torque varies; and wherein the engine is
operated using port fuel injection for the first 100 seconds of
engine operation; and wherein a spark retard of less than 10 crank
angle degrees is employed during at least some time during the
first 100 seconds of engine operation.
12. A spark ignition engine that is fueled with gasoline and uses
direct injection; wherein reduction of particulate emission is
obtained by increasing a fraction of directly injected fuel that is
introduced between a selected first crank angle degree that occurs
on the intake stroke and a selected second crank angle degree that
occurs on the compression stroke; and wherein the selected first
and/or second crank angles are varied with changing
temperature.
13. The spark ignition gasoline engine of claim 12, wherein the
first crank angle occurs later when the temperature is lower.
14. The spark ignition gasoline engine of claim 12, wherein the
second crank angle occurs earlier when the temperature is
lower.
15. The spark ignition gasoline engine of claim 12, wherein the
first crank angle occurs later and the second crank angle occurs
earlier when the temperature is lower.
16. The spark ignition gasoline engine of claim 12, wherein
information about the first and second crank angles is used in a
lookup table for controlling injection timing.
17. The spark ignition gasoline engine of claim 12, wherein
information about the first and second crank angles is used in a
lookup table for controlling injection timing; and wherein the
crank angle limits are determined by laboratory measurements of
engine parameters.
18. The spark ignition gasoline engine of claim 12, wherein during
cold operation, the first crank angle is at least 10 degrees later
than it would otherwise be.
19. The spark ignition gasoline engine of claim 12, wherein during
cold operation, the first crank angle is at least 20 degrees later
than it would otherwise be.
20. The spark ignition gasoline engine of claim 12, wherein during
cold operation, the second crank angle is at least 10 degrees
earlier than it would otherwise be.
21. The spark ignition gasoline engine of claim 12, wherein the
start of direct injection is adjusted so as to begin after the
selected first crank angle and before the selected second crank
angle.
22. A spark ignition gasoline engine that is fueled with gasoline
and uses direct injection; and wherein reduction of particulate
emission is obtained by increasing a fraction of directly injected
fuel that is introduced between a selected first crank angle degree
that occurs on the intake stroke and a selected second crank angle
degree that occurs on the compression stroke; and wherein the
fraction of directly injected fuel that is introduced between the
first selected crank angle degree and a second crank angle degree
on the compression stroke increases with increasing fuel delivery
rate.
23. The spark ignition gasoline engine of claim 22, wherein the
fuel delivery rate is adjusted so as to increase the fraction of
directly injected fuel that is introduced between the first and
second crank angle degrees.
24. The spark ignition gasoline engine of claim 22, wherein a
higher fuel injection pressure is used for at a least some time
during a cold start period of the first 100 seconds of engine
operation.
25. The spark ignition engine of claim 22, wherein when the engine
temperature is lower, the selected first crank angle occurs later
than it otherwise would.
26. The spark ignition engine of claim 25, wherein port fuel
injection is also employed and is used so as to reduce particulate
emissions.
27. The spark ignition engine of claim 22, wherein the selected
first and/or second crank angles are determined by inputs from a
look up table.
28. The spark ignition engine of claim 22, wherein increased fuel
injection pressure is employed so as to increase the fraction of
fuel that is directly injected between the selected first and
second crank angles.
29. The spark ignition engine of claim 22, wherein increased fuel
injection pressure is employed so as to increase the fraction of
fuel that is directly injected between the selected first and
second crank angles and wherein an increase in fuel injection
pressure of 50% can increase the onset BMEP for particulate
emissions by at least 15% relative to the onset BMEP for
particulate emissions where a 50% higher fuel injection pressure is
not employed.
30. The spark ignition engine of claim 22, wherein a temporary fuel
injection pressure increase is employed during high power engine
operation.
31. The spark ignition engine of claim 22, wherein an electric fuel
pump is employed to temporarily increase fuel injector
pressure.
32. The spark ignition engine of claim 22, wherein the first and
second selected crank angles are varied based upon use of EGR.
33. The spark ignition engine of claim 22, wherein air preheating
is used to reduce particulate emissions.
Description
[0001] This application is a continuation of U.S. patent
application Ser. No. 15/713,997, filed Sep. 25, 2017, which claims
priority to U.S. Provisional Patent Application Ser. Nos.
62/399,755, filed Sep. 26, 2016; 62/432,140 filed Dec. 9, 2016;
62/477,587 filed Mar. 28, 2017; and 62/531,935, filed Jul. 13,
2017, the disclosures of which are incorporated in their
entireties.
BACKGROUND
[0002] Spark ignition (SI) gasoline engines that use direct
injection (DI), particularly those that are turbocharged, provide
important efficiency increases over port fuel injection (PFI)
engines and are rapidly increasing in use. The vaporization cooling
from direct injection provides increased resistance to knock,
thereby allowing operation at higher levels of turbocharging and/or
compression ratio.
[0003] However, particulate emissions from DI gasoline engines are
a health concern because they are small particles that can lodge in
the lungs and are more than 10 times greater in mass per mile
driven than emissions from PFI engines and are about 10-100 times
greater in number per miles driven.
[0004] Gasoline particulate mass and number are presently regulated
in Europe and more stringent regulations are expected. Regulations
are expected in the US and California in the near future.
[0005] There is a need for a robust and low cost means to greatly
reduce these particulate emissions with at least a near zero impact
on vehicle fuel efficiency and preferably a zero or positive impact
on vehicle fuel efficiency.
SUMMARY
[0006] Additional features of a PFI+DI system that is optimized for
reducing DI-generated PM emissions using a particulate generation
model are disclosed. These features include additional control
techniques for timing of direct injection. Illustrative
computational results for decreasing PM emissions from SI gasoline
engines operated in various drive cycles are also presented. These
calculations illustrate the very large potential reduction in
direct injection generated particulates (e.g. 98%) that can be
obtained.
[0007] The present invention also discusses how this optimized
PFI+DI system could remove the need for using a gasoline
particulate filter (GPF) by providing DI-generated particulate
reduction, reliability, efficiency and cost advantages relative to
use of GPF with a DI only engine. Alternatively, it could be used
in combination with a GPF to ease GPF operation requirements and/or
provide additional particulate reduction and robustness. Symbiotic
uses of GPF technology in combination with optimized PFI+DI are
described. In addition, the present invention also describes how
approaches used for reducing emissions for optimized PFI+DI can be
utilized in engines that are fueled with DI alone.
BRIEF DESCRIPTION OF THE DRAWINGS
[0008] For a better understanding of the present disclosure,
reference is made to the accompanying drawings, in which like
elements are referenced with like numerals, and in which:
[0009] FIG. 1 is a block diagram of a Control System for a PFI+DI
System according to one embodiment;
[0010] FIG. 2 shows illustrative contours of the percentage of fuel
that is provided by direct injection at borderline knock in a
gasoline engine fueled by PFI+DI as a function of brake mean
effective pressure and engine speed for a turbocharged gasoline
engine with a compression ratio of 12 and a downsizing of 30%;
[0011] FIG. 3A is a schematic of conditions for no piston wetting
(left) and piston wetting (right) by directly injected gasoline and
FIG. 3B shows piston position and threshold crank angles where
particulate generation occurs;
[0012] FIG. 4 is an illustrative calculation for the percentage of
directly injected gasoline that produces particulates as a function
of brake mean effective pressure (BMEP) and engine speed, using a
fuel injector operated with a 10 MPa fuel pressure;
[0013] FIG. 5 is an illustrative calculation for the percentage of
directly injected fuel that produces particulates as a function of
BMEP and engine speed, using operated with a 15 MPa fuel pressure;
and
[0014] FIG. 6 is an illustrative calculation of the percent of
total fuel that results in particulate production by direct
injection, in which the top line shows where 100% of the fuel is
directly injected and the fuel injector pressure is 10 MPa.
DETAILED DESCRIPTION
Basic Features
[0015] The basic control features of the optimized PFI+DI control
system are shown in FIG. 1. Inputs to controlling the amounts of
port and direct injected fuel include information for active
control using a knock detector and, when needed, open loop control
to determine the fraction of fuel that must be directly injected to
prevent knock; and information about the total amount of fuel used
at various values of torque and speed in the drive cycle.
[0016] Key parameters that are controlled can include spark retard
and/or other knock suppression techniques; direct injection pulse
timing relative to piston position in the intake and compression
strokes; and direct injector pulse width. The control system can
use closed and/or open loop control and employ look up tables. The
direct injection timing can depend on parameters that include
temperature, fuel injection pressure, engine torque and engine
speed. The fuel in the fuel tank can be gasoline; a low alcohol
concentration gasoline-alcohol mixture such as E10, or some other
fuel.
[0017] The fuel management system uses a layered set of synergistic
approaches: [0018] (1) Minimizing the fraction of fuel that is
directly injected while preventing knock; [0019] (2) Reducing
particulate emissions by reducing required knock resistance; [0020]
(3) Control of direct injection operation period and timing. Each
of these approaches is described in more detail below. Minimizing
the Fraction of Fuel that is Directly Injected while Preventing
Knock
[0021] The first approach robustly minimizes use of direct
injection so that it is limited to the fraction of total fuel that
is needed to prevent engine knock by vaporization cooling as the
engine operating parameters, particularly torque, change over a
drive cycle. Prevention of knock (unwanted self-ignition that can
damage the engine) enables use of smaller engines to provide the
same torque and power as larger engines and also operate at higher
compression ratio, thereby improving efficiency.
[0022] Since direct injection is not needed for knock prevention at
low torque, the engine can be fueled entirely by PFI in this
regime. PFI provides homogeneous air/fuel mixtures. At a certain
level of torque, some DI must be used to prevent knock and the
fraction of fuel that is provided by DI increases as the knock
suppression requirement increases with increasing load.
[0023] The control system matches the fraction of fuel provided by
DI (including zero or 100% use of DI) to that needed to prevent
knock as the torque and speed vary between zero and their highest
values. Using this technique over the torque and speed range of a
drive cycle minimizes the average fraction of fuel f.sub.1 that is
provided by direct injection over the drive cycle. Even if the
matching only occurs over a large fraction of the torque range from
zero to its highest value rather than 100% of the range (e.g. if
occurs in at least 80% of the torque range), there will be a large
reduction in the average fraction of fuel that is provided by
direct injection.
[0024] In the absence of effects from other approaches which are
mentioned below, the use of the first approach over a drive cycle
reduces particulate emissions from direct injection to a parameter
f.sub.1 times the amount of particulate emissions that would occur
if the engine were entirely fueled by direct injection. The amount
of particulates generated by direct injection is proportional to
the amount of direct injection that is employed at each value of
torque and speed. When the effects of the other approaches into
account, the fraction of particulates that are generated relative
to an engine using direct injection alone will be less than
f.sub.1.
[0025] The amount of reduction of DI-generated particulates using
the first approach alone can be calculated for various drive
cycles. The fraction of fuel that is directly injected is
determined by the requirement to prevent knock at various values of
brake mean effective pressure (BMEP), which corresponds to torque
per cylinder volume, and speed during the drive cycle.
[0026] To minimize use of direct injection, the fraction of fuel
used at various points in the drive cycle is matched to that needed
to prevent knock. It is substantially equal to the value needed to
prevent knock. As torque is reduced from its highest value, the
fraction of fuel provided by direct injection is continually
decreased so as to be substantially equal to the fraction needed to
prevent knock and reaches 0% at a certain value of torque.
[0027] The fraction of fuel that is provided by direct injection in
engine operation at various operating points during a driving cycle
is determined by a knock sensor and/or by a lookup table using a
calculation for the required fraction of directly injected fuel for
knock control at various values of torque, temperature and speed
for a given engine.
[0028] The information about the fraction of fuel that is required
to prevent knock is combined with information about the fraction of
time the engine is operated at various values of torque and speed
for a given drive cycle to determine the fraction of fuel, f.sub.1
that is directly injected over a drive cycle. f.sub.1 is higher for
an aggressive drive cycle such as the US06 cycle where more time is
spent at high BMEP where a higher fraction of directly injected
fuel is needed
[0029] FIG. 2 shows contours for the required DI gasoline fraction
(in percentage) for knock avoidance for a turbocharged engine with
an inlet manifold pressure of 1.5, operating at maximum brake
torque (MBT) timing, with a compression ratio of 12. This
combination of pressure and compression ratio has around the same
direct injection knock suppression requirement as an engine with an
inlet pressure of around 1.9 bar, a downsizing of around 45% and a
compression ratio of 9.5. FIG. 2 thus represents an illustrative
example of turbocharged direct injection gasoline engines.
[0030] This type of computationally generated information could be
used in a lookup table for the control system. The look up table
could use the combination of this computationally generated
information and engine measurements. It could provide information
of the fraction of fuel that must be directly injected to prevent
knock throughout the engine torque-speed operation space.
[0031] For a typical turbocharged gasoline direct injection engine
when at least 40% of the fuel is provided by direct injection at
the maximum load, the fuel management system would allow knock free
operation at about 2/3 of the maximum torque.
[0032] At the highest values of torque, the engine is required to
operate mostly on DI gasoline. Although higher speed reduces the
time for the autoignition that produces knock, the required DI is
more or less constant with engine speed at a given load, as a
result of the engine operating at higher power (and thus higher
temperatures) at the higher speeds.
[0033] Table 1 shows the relative amounts of DI-generated
particulates for the US06, urban UDDS and highway HWFET light duty
drive cycles using this approach. The percent DI-generated PM
reduction is also shown. For the US06 cycle and the representative
engine parameters used in FIG. 1, f.sub.1 and correspondingly the
relative amount of DI-generated particulates) is less than 0.3.
This fraction is reduced to less than around 0.2 when 100%
reduction of cold start DI-generated particulate emissions is taken
into account, as described later.
[0034] For typical urban cycles such as the UDDS cycle or less
aggressive highway cycles such as the HWFET, an illustrative value
of f.sub.1 is less than about 0.05.
TABLE-US-00001 TABLE 1 Illustrative reduction in DI-generated
particulates for various drive cycles by matching DI fraction of
fuel to that needed to prevent knock and by achieving 100%
reduction in direct-injection cold-start emissions by using 100%
port fuel injection. % reduction of DI-generated f particulates
USO6 <0.2 >80% UDDS <0.05 >95% HWFET <0.05
>95%
Reducing Particulate Emissions by Reducing Required Knock
Resistance
[0035] A second approach is to further reduce f.sub.1 by applying a
means to reduce the knock resistance that must be provided by
direct injection at given values of torque and speed throughout the
engine operating range. Techniques for doing this include spark
retard, upspeeding and variable valve timing. These techniques can
significantly reduce f.sub.1 but also reduce efficiency. They are
applied selectively so as to maximize the amount of additional
particulate reduction for a given amount of efficiency loss. Other
techniques that could be applied are variable valve timing;
internal exhaust gas recirculation (EGR); cooled EGR (especially at
high loads); and the substitution for open-valve port fuel
injection where direct injection would otherwise need to be used to
prevent knock.
[0036] One control feature is to provide a set amount of spark
retard in crank angle degrees for engine operation above torque (or
torque and speed) values where direct injection would otherwise be
required. The application of spark retard moves the contours of DI
fuel percentage in FIG. 2 upward.
[0037] Illustrative results are shown in Table 2 for operation in
the US06 driving cycle which, as previously mentioned, is a very
aggressive drive cycle that requires much more time at high torque
than other drive cycles. The amount of direct injection is
minimized by maximum use of port fuel injection without the onset
of knock.
TABLE-US-00002 TABLE 2 Illustrative effect of spark retard on
DI-generated particulates in the US06 cycle. Spark retard is in
crank angle (CA) degrees Spark retard 0 5 10 Efficiency, normalized
1 0.98 0.89 DI gasoline, normalized 1 0.62 0.31 Reduction in
DI-generated >80% >85% >90% particulates
[0038] The results in Table 2 are for an engine with a compression
ratio of 12 and downsizing of around 30%, and are normalized for
the case of no spark retard. The amount of particulate generation
from direct injection is reduced in proportion to the reduction in
f.sub.1 when only the first and second approaches are applied.
[0039] As shown in Table 2, increasing spark retard by .kappa.
crank angle (CA) degrees reduces f.sub.1 by nearly 40% along with a
small 2% relative decrease in efficiency. For a 10 CA degree spark
retard, there is a reduction of f.sub.1 of about 70% along with a
10% decrease in efficiency. The fuel management system can be
operated so as to insure that the spark retard that is used is no
greater than 10 CA degrees and preferably no larger than 5 CA
degrees.
[0040] Table 2 also shows percent reduction in DI-generated
particulates relative to 100% DI injection using the information in
Table 1. For the US06 drive cycle, the estimated percent reduction
is greater than 85% for an efficiency decrease of 2%. For the UDDS
and HWFET cycles, based on these calculations a reduction of
greater than 95% may be achievable without the use of spark retard.
For these cycles where there is very little need for DI for
controlling knock, there is no need for spark retard for its
minimization.
[0041] Another option for introducing spark retard is to only
deploy it at the higher end of the range at which direct injection
would otherwise be used.
[0042] An additional option for introducing spark retard is to
increase it with increasing DI fuel fraction. No change in spark
retard would be required when no direct injection is employed at
low values of torque. The spark retard would be increased with
increasing torque when direct injection is employed and could be
adjusted so as to be equal to the amount needed to prevent knock.
This approach could be employed to minimize the amount of
efficiency loss from its use.
[0043] The spark retard that is employed can be limited so as not
to generate temperatures that would damage a turbocharger. The need
for fuel enrichment to limit temperature could thus be avoided.
Fuel enrichment can result in increased hydrocarbon emissions.
[0044] In addition to reducing f.sub.1, the use of spark retard can
further reduce particulate emissions by decreasing the amount of
fuel that is directly injected (in contrast to the fraction of fuel
that is directly injected) and thus decreasing the duration of the
direct injection. The decrease in the duration of direct injection
enables an increase in the fraction of time that fuel is directly
injected that avoids piston wetting. This impact of spark retard is
described in the section on injection operation period and
timing.
[0045] Selective use of upspeeding or variable valve timing to
reduce particulate emission can also be employed in a way that is
analogous to the use of spark retard. Upspeeding reduces the DI
requirement faster than the increased speed reduces the window for
injection without piston impingement, and thus results in decreased
particulate emission during warmed engine operation as described
below.
[0046] The calculations for DI minimization and for use of spark
retard to enable a lower fraction of DI use are based on well
established information about engine operation.
Control of Direct Injection Operation Period and Timing
[0047] A third approach to decrease PM emissions, which uses a
computational model for optimization, is to reduce the duration
(length of time) of injection of directly injected fuel so that as
much of it as possible it falls in a window between two crank angle
limits where particulate production due to when piston wetting does
not occur.
[0048] Particulate generation from direct injection is mostly due
to piston wetting. In the computation model, which can be
applicable to cold start operation as well as warmed up operation,
it is assumed piston wetting occurs when the piston is high in the
cylinder close to the fuel injector, and the direct fuel injector
is open either earlier than a threshold crank angle .theta..sub.1
during the intake stroke, or later than another crank angle
.theta..sub.2 during the compression stroke, as shown in FIGS.
3A-B.
[0049] When the crank angle is such that the piston is sufficiently
low in the cylinder, liquid from the direct injection is not able
to reach the piston wall and produce particulates. In the model,
piston wetting does not occur for the time period during which
direct injection occurs within the time period between
.theta..sub.1 and .theta..sub.2, as shown in FIG. 3B. If the direct
injection entirely occurs between the time elapsed between the two
crank angles, according to the model there would be no particulate
production.
[0050] Since the time period for direct injection fueling depends
on the amount of directly injected fuel, prevention or reduction of
piston wetting can be obtained by limiting this parameter and by
starting injection when the first crank angle occurs. That is,
according to the model there is a window in time, or in crank angle
degrees, between .theta..sub.1 and .theta..sub.2 where there is no
impingement on the piston.
[0051] To obtain the greatest particulate reduction benefit, the
direct injection should be initiated at .theta..sub.1 (first crank
angle threshold) and, if not, preferably at a later time that is
less than 25% of the time elapsed between crank angle limits and
more preferably at a time that is less than 5% of the time between
crank angle degrees.
[0052] This approach provides a means to substantially reduce
particulate generation by controlling the amount of directly
injected fuel to be less than a threshold level at which the fuel
injection time period exceeds the time between the two crank angle
limits. When the amount of directly injected fuel is below this
level, particulate emissions from piston wetting by direct
injection can be reduced to a near zero level even when there a
significant fraction of fueling by direct injection in order to
prevent knock. The fuel injection can be provided by single pulse
or a series of shorter pulse during the fuel injection period.
[0053] Even when the fuel injection time period extends beyond the
threshold length at which piston wetting occurs, there will still
be a reduction in the amount of particulate production due to the
amount of fuel that was introduced into the cylinder during the
time that falls within the window for no piston wetting. For that
part of the fuel injection period that extends beyond the time
between the crank angle limits, the model indicates a linear rise
in the amount of particulate generation with increasing pulse
duration, corresponding to an increasing amount of fuel that is
directly injected. Even if the fuel injection period is twice the
window between the threshold crank angle degrees, according to the
model, there would still be around a factor of two reduction in
particulate emissions.
[0054] The computational model determines the relative amount of
direct injection-generated particulates (either in mass or number
density) by determining the percentage of directly injected fuel
that falls outside of the two crank angle thresholds. The
percentage of directly injected fuel that falls outside of the two
threshold crank angles depends on the fuel injection rate (which in
turn depends on the fuel injection pressure), the amount of
directly injected fuel and the time at which the fuel is injected.
The amount directly injected fuel depends on BMEP and speed.
[0055] The crank angle degree limits can be determined for various
engines by measurements in the laboratory. This information can be
used in a lookup table for controlling injection timing and related
engine and fuel injector parameters. Modeling of piston wetting
process can also be used to determine the crank angle thresholds
used in the creation of the look up table.
[0056] The crank angle limits, and thus the onset of piston
wetting, depend on the direct injection fuel penetration length
through the air-fuel mixture in the cylinder. This penetration
length will depend on viscosity of the gas in the cylinder, which
in turn depends on temperature and to a lesser amount, composition.
Decreased viscosity at low temperature at time of cold start
increases the penetration of the fuel droplets, increasing the
piston wetting. In addition, the lower temperatures at cold start
also reduces the rate of evaporation of the droplet, also
increasing the piston wetting (as compared to conditions of warmed
up operation). Increased penetration length results in a larger
value of .theta..sub.1 and a smaller value of .theta..sub.2.
[0057] Because of these dependences on temperature, varying the
time of the start of direct injection as a function of air
temperature, engine temperature and other factors that determine
direct injection fuel penetration length, including injector
pressure and manifold pressure, can be used to reduce the direct
injection particulate emissions.
[0058] At cold temperatures, the start of injection should be at a
later crank angle in order to achieve the substantial reduction in
piston wetting when the piston position is such that the path
length from the injector to the piston is greater than the
penetration length. Because of increased penetration of the
directly injected fuel, the start of direct injection, which can
correspond to the first selected crank angle, should be delayed by
at least 10 to 20 crank angle degrees to avoid piston wetting
depending on the injection pressure, the temperature and the motion
in the cylinder. Similarly, the end of direct injection, which can
correspond to the second selected crank angle should be advanced by
at least 10 to 20 crank angle degrees to avoid piston wetting.
Thus, the extent of the window in which there is greatly reduced
piston wetting can be reduced by 20 to 40 crank angle degrees.
[0059] A look up table can be used to control the start of
injection in crank angle degrees as a function of temperature and
other parameters. This approach could be used for engines that use
only direct injection as well as for engines that use optimized
PFI+DI.
[0060] At engine start-up, it is challenging to provide enough time
for fuel evaporation, and thus when direct injection is used, as
early start of injection as possible is desired, with the
limitation of the piston wetting. The engine operation can start
with port fuel injection can use as much port fuel injection as
possible during the cold start period including operation entirely
on port fuel injection, By the time that DI is needed for
controlling knock, the fuel pressure in the high pressure DI should
be built up.
[0061] Means of preheating the cylinder, such as air compression,
could be used to reduce the increased direct injection particulate
generation that occurs during cold start due to a longer fuel
penetration path and lower fuel vaporization relative to warmed up
operation. There are several ways of preheating the charge. One
possibility is to advance the timing of the exhaust valve, and in
particular, the exhaust valve closing. Thus, there is a relatively
hot, relatively high pressure gas in the cylinder at the time that
the inlet valve opens. The back flow into the inlet manifold can be
used to produce droplets from the pools in the inlet valves, and
the higher temperatures should help increase the rate of
evaporation for the PFI fuel, decreasing hydrocarbon emissions.
[0062] Advancing the exhaust valve by at least 30 and preferably at
least 40 crank angle degrees increases the temperature in the
cylinder by at least 30 K and preferably by at least 50 K. The
higher temperatures in the cylinder, during the next cycle can help
minimize the production of DI particulates by increased evaporation
rates of the DI spray and increased viscosity, which decreases the
penetration length. Temperature increases of around 50 C can be
achieved in this manner. Other means of preheating the air include
air preheaters, including electrical preheaters.
[0063] To provide illustrative calculations, our computational
model employs the experimental measurements by Keterer and Cheng
(On the Nature of Particulate Emissions from DISI Engines at Fast
Idle, SAE Int. J. Engines 7(2):986-994, 2014,
di:10.4271/2014-01-1368) of direct injection-generated particulates
as function of start of injection (SOI) as a basis for the selected
threshold crank angles .theta..sub.1 and .theta..sub.2. These crank
angles are consistent with rough calculations of the spray
penetration assuming 10 MPa injection and conventional nozzles for
the injector, at the conditions of a modern 2 liter engine with the
stock fuel injector.
[0064] It is not expected that the penetration will vary much with
either speed of the piston (although it affects the charge motion
in the cylinder) or the cylinder pressure. Generally in warmed up
operation, .theta..sub.1 should roughly be between 120 and 140
degrees BBDC (Before Bottom Dead Center) and .theta..sub.2 should
be between 70 and 80 degrees ABDC (After Bottom Dead Center). The
specifics depend on the injection pressure, the cylinder
parameters, temperature and to lesser degree, the cylinder
pressure, and cylinder charge motion.
[0065] FIG. 4 shows an illustrative calculation for the percentage
of directly injected fuel that produces particulates as a function
of BMEP and engine speed. These parameters are related to the
amount of directly injected fuel and thus the required injection
time period for given fuel injection rate. The fuel injection rate
is determined by fuel injection pressure. The injection time period
decreases with in increasing fuel injector rate which increases
with increasing fuel injection pressure. The fuel injection
pressure for FIG. 4 is 10 MPa; direct fuel injectors operated at
this pressure are in widespread use in gasoline engines.
[0066] As the engine speed increases, there is less time for
injection, and thus the fraction of the DI fuel that falls on the
piston increases at a given BMEP. It is assumed that the engine
operates at MBT (maximum brake torque) timing.
[0067] FIG. 5 shows the effect of a faster fueling rate using a
higher fuel injection pressure of 15 MPa (50% more than FIG. 4).
Higher pressure injection increases the amount of directly
injection fuel and the BMEP that can be employed within the window
without piston wetting, thereby decreasing the potential for piston
impingement. When the injection period is shorter than the window
for piston impingement, it is still desirable to inject early in
the window, as this results in more homogeneous charge as well as
lower peak temperatures. Comparing FIG. 4 and FIG. 5 shows that a
50% increase in the fuel injection pressure from 10 MPA to 15 MPA,
can increase the BMEP at which the onset of particulate emissions
occurs by at least 15%.
[0068] The combination of the first approach with the third
approach can be used to reduce the amount of fuel that is directly
injected and thus reduce the time period for direct injection. It
can thereby be employed to substantially increase the fraction of
directly injected fuel that occurs between the two crank angle
limits. The amount of direct injection-generated particulates is
minimized by matching the DI fuel fraction to substantially that
needed to prevent knock. The matching would occur throughout the
torque range or at least through a high fraction of the torque
range.
[0069] The combination of these effects can result in a large
decrease in the amount of warm-engine particulates with decreasing
torque. This can cause a further substantial decrease in the
fraction of particulates that are produced in a drive cycle
relative to the amounts produced if all of the fuel were to be
introduced by direct injection.
[0070] Relative to 100% direct injection, the direct
injection-generated particulates are reduced first by substitution
of port fuel injection for direct injection. This number is further
reduced by reducing the fraction of direct fuel injection that
produces particulates by increasing the fraction of direct
injection that occurs between the two crank angle limits.
[0071] To determine the fraction of fuel that is directly injected
and then impinges on the piston, the information in FIG. 2 is used
to determine the amount of directly injected fuel required for
knock suppression at various values of brake mean effective
pressure. This information is then combined with information that
relates the amount of directly injected fuel to the operation
period of direct injection and the time between the crank angle
limits.
[0072] Using the combination of information from FIG. 2 and FIG. 4,
FIG. 6 shows illustrative contours of the percentage of total fuel
that is impinging on the piston (and thus, producing particulates)
over the engine map for an engine with a compression ratio of 12
and which is downsized by 30%. The upper line in FIG. 6 indicates
where 100% of the fuel is directly injected.
[0073] Comparing FIG. 6 with FIG. 2, it can be seen that at each
value of torque and speed, the amount of direct injected generated
particulates that are produced can be reduced by a substantial
factor and in some cases, be eliminated all together by injection
time duration and timing control.
[0074] When the fuel introduced by port injection is substituted
for direct injection by an amount that is allowed by the
requirement to suppress knock, there is a large reduction in the
amount of directly injected fuel that is used and this causes an
additional decrease in the amount of particulates that are
generated at given values of BMEP and engine speed. In addition, an
increase in the injector fuel pressure by a factor of two would
cause a large shift of contours of constant particulate generation
to values of higher BMEP.
[0075] The fraction of total fuel over a given drive cycle that
impinges on the piston and produces particulates can be determined
by combining information about the amount of operation at various
points of torque and speed with the information in FIG. 6. Table 3
shows these results for a representative light duty truck with a
turbocharged 3.5 liter engine for the UDDS, US06 and HWFET cycles.
The fuel injector pressure is 10 MPa. The first row shows the
fraction of directly injected fuel that is used over the drive
cycle when DI is used and when PFI+DI is employed so that the
fraction of fuel that is directly injected matches that needed to
prevent knock. With no additional effects, this is equal to the
fraction of fuel over a drive cycle that produces direct injection
generated particulates, f.sub.1.
[0076] The second row in Table 3 shows the fraction of total fuel
that impinges on the piston taking the effect of injection timing
and pulse length into account using the information in FIG. 6. This
corresponds to the fraction of total fuel that results in direct
injection generated particulates, f.sub.2.
[0077] For the US06 cycle in Table 3, f.sub.1 is around 9%. This
number could be further reduced by use of spark retard or some
other knock suppression technique.
[0078] The 9% production of direct injection generated particulates
is for the warmed up part of the drive cycle. Assuming that one
third of direct injection generated particulates are produced at
cold start and that this number can be reduced to zero by 100% use
of port fuel injection (as discussed below), the direct injection
generated particulates could be reduced to around 6%.
[0079] Providing some margin, it would be reasonable to assume that
direct injection generated particulates could be reduced by at
least 90% over the US06 by matching the fraction of directly
injected fuel to that needed to prevent knock (without use of spark
retard or some other knock suppression mechanism). Even a greater
margin for 90% reduction could be obtained by use of spark retard
of some other knock suppression mechanism,
[0080] The reduction in particulate generation would be much
greater for the UDDS and HWFET drive cycles.
[0081] Taking the effect of injection timing and pulse length into
account the direct-injection particulate generation for the US06
cycle in Table 3 could be reduced by about a factor of 100 by using
PFI+DI instead of DI alone ("fraction wall wetting"=0.001 for
PFI+DI vs. 0.096 for DI). Even if the level of direct-injection
generated particulates using optimized PFI+DI is only a factor of
20 lower than that for use of DI alone in the same engine that
provides the same efficiency and performance for the US06 drive
cycle, this provides a large improvement over the 4 to 5 times
reduction from a gasoline particulate filter. The reductions
provided by optimized PFI+DI for less aggressive drive cycles are
even greater.
[0082] Additional reduction of direct-injection generated
particulate emissions could be obtained by temporarily operating at
a higher fuel injector pressure (e.g. 15-20 MPa) or using spark
retard in the warmed up part of the drive cycle.
TABLE-US-00003 TABLE 3 Illustrative relative direct
injection-generated particulate generation for the UDDS, US06 and
HWFET cycles for a turbocharged DI engine, the Ford 3.5 liter
ecoboost engine. The fuel injector pressure is 10 MPa. UDDS US06
HWFET DI PFI + DI DI PFI + DI DI PFI + DI Fraction of DI 1 0.005 1
0.088 1 <0.001 Fraction wall <0.001 0 0.096 0.001 0 0
wetting
[0083] In Table 3, the row for "the fraction of DI" shows relative
particulate generation for the case where particulate generation is
reduced by matching the fraction of fuel that is directly injected
to that needed to prevent knock. The row for "fraction wall
wetting" shows the fraction of total fuel that produces piston
wetting by impingement of directly injected fuel.
[0084] Additional reductions can be obtained by combining the knock
reducing techniques of the second approach (e.g. spark retard) with
both the first and third approaches. The application of these
techniques can be highly leveraged due to the very steep decrease
in DI particulate generation with decreasing torque. These
techniques can be applied so as to minimize any efficiency decrease
for a given amount of particulate reduction.
Control Techniques for Cold Start and Warmed Up Operation
[0085] During the cold start period, it could be advantageous for
the control system to be operated so as to be different for the
cold start versus the warmed up period of engine operation. The
direct-injection generated particulate emissions are much greater
during cold start than during the warmed up engine period. Roughly
one third to one half of the particulate emissions occur during
this period.
[0086] It can be especially important to minimize the use of direct
injection during cold start by matching the fraction of fuel that
is directly injected to that needed to prevent knock. Further
reduction can be obtained by use of spark retard.
[0087] A rough estimate is for the increased particulate cold start
period of 100 seconds. This cold start period is typically less
than 10% of the time of a typical drive cycle. Hence, the use of
spark retard can have a larger impact on direct-injection generated
particulate reduction relative to any resulting efficiency loss
averaged over a drive cycle.
[0088] Increased spark retard (e.g. in the. 5-10 CA degree range)
could be used to reduce direct injection use over the increased
particulate emissions cold start period down to a very low level
(less than 5% and preferably less than 2% of the total fuel use) or
to zero with zero or minimal impact (less than 1% and preferably
less than 0.5%) impact on overall drive cycle efficiency. The spark
retard could be varied with torque during cold start, and also in
warmed up operation, so as to limit its use based on the amount
needed to bring the DI fraction to zero or a desired very low
level, when the required spark retard is less than a specified
number (e.g. 10 CA degrees or less). If spark retard that is
required is larger than the specified number, the spark retard is
kept constant at the specified number.
[0089] Use of other means of reducing the need for direct injection
to suppress knock, such as variable valve timing and upspeeding can
also be employed.
[0090] Control of injection time period and timing could also be
used to further insure that the amount of particulate generation
from direct injection used is zero or at least at a very low level
during the cold start period for increased particulate production.
If direct injection is used, the start of direct injection can be
determined and/or varied as a function of temperature information
and other various inputs to the control system so as to avoid or
reduce operation before the selected first crank angle and after
the second threshold crank angle for particulate production from
wall wetting.
[0091] In addition, if direct injection is used during cold start,
temporary use of higher direct fuel injector pressure during part
or all of the cold start period could be employed to reduce or
eliminate direct injection-generated particulates.
[0092] The spark retard that is used in the increased particulate
generation cold start period could be varied as a function of
engine speed; torque; engine temperature; time after engine start;
and fraction of fuel that would otherwise be needed to be directly
injected to prevent knock. A look up table could be used to
determine the amount of spark retard that would meet the objective:
a combination of desired reduction in direct-injection generated
particulates (in mass or particulate number) and a limit on the
adverse effect on drive cycle efficiency. The look up table could
take the effects of the above means of reducing direct injection
generated particulates into account.
[0093] In both the cold start and the warmed up period of the drive
cycle, the effects of selectively used spark retard on reducing the
fraction of directly injected fuel and thus reducing the amount of
particulate generation from direct injection can be determined for
various control features. The control system can be operated so as
to meet various goals (e.g. maximizing particulate reduction per
amount of efficiency loss; and/or limiting efficiency loss to a
given amount).
[0094] As is the case for the cold start period, an additional
technique that can be used during warmed up period is to
temporarily increase the injection pressure so as to increase the
rate of fuel introduction. This reduces the required fuel injection
duration and increases the amount of directly injected fuel that
can be used before exceeding the threshold for particulate emission
from piston wetting.
[0095] Increasing the fuel pressure also results in smaller, higher
velocity aerosols. The smaller aerosols have increased viscous drag
(per unit mass) than larger aerosols, and evaporate faster, in part
compensating for the increase in penetration depth due to the
higher speed. The increased penetration depth decreases slightly
the window for injection without wall wetting, but this is a second
order effect, the main being the reduction in the total injection
duration.
[0096] This increase in the allowed amount of directly injected
fuel before the threshold for particulate emission from piston
wetting is reached increases the threshold brake mean effective
pressure (BMEP) at which particulate generation occurs.
[0097] The optimum use of a temporary increase in fuel injector
pressure could be determined from information about the percent of
operation time at various positions in the engine map and the
change in particulate generation at those positions when the fuel
injection pressure is increased.
[0098] The amount of time that this technique could be used during
a drive cycle would be limited so as to not cause injector
overheating or to exceed some other injector or fuel pump operation
limit.
[0099] The temporary increase in injector pressure could be used so
as to have the greatest impact on particulate reduction. Because
this technique would involve small amount of extra power use from
the engine and would only be used during a small fraction of the
drive cycle, it would have a negligible impact on engine
efficiency. Determination of the greatest impact on particulate
generation for a given time of temporary injection could be
employed to determine how to best use a temporary increase in
injector pressure.
[0100] In addition to a decreased direct injection duration, high
pressure injection also decreases the fuel aerosol droplet size,
increasing the evaporation rate. Increased evaporation rate
decreases the temperature in cylinder, as evaporative cooling of
the cylinder charge early in the compression stage is more
effective in decreasing peak temperatures in the cylinder than late
evaporative cooling. Higher pressure injection thus allows for
decreased amount of directly injection fuel required for preventing
knock, decreasing the potential for piston impingement.
[0101] As an example, the temporary increase in pressure could be
obtained by operation of a 10 MPa-15 MPa pressure fuel injector for
a limited time at 20 MPa or more. Today's high performance gasoline
fuel injectors are operating around 20 MPa.
[0102] The increased pressure could be applied based upon the
amount of direct injection fuel required and the amount of engine
power required and could be limited so as to prevent any adverse
effect on the injector system. The increased pumping power for the
temporary increased injector pressure would be small compared to
the pumping power used over a drive cycle.
[0103] Temporarily increased injector power could be especially
useful during high power engine operation in warmed up operation as
well as in cold start. During high power operation, a longer direct
injector pulse length would be needed than at lower power and could
thereby result in more direct injection being used outside of the
crank angle degree limits. Use of higher pressure operation during
this time could play an important role by reducing the direct
injection pulse length.
[0104] The change in injector pressure could be either a set amount
or a variable amount to provide the largest impact in reducing
particulate emissions.
[0105] The point or points in the drive cycle where the temporary
increase in fuel injector pressure would be used could be
determined by the engine power at different points, the amount of
time spent at these points for various drive cycles, and the impact
of the increase in injector pressure (linear reduction in
particulate emissions versus reducing the injector period to below
the threshold level).
[0106] Some options for fuel injector design and operation at
temporary operation at higher pressures are the use of an electric
pump that is driven harder during times of increased pressure
demand, or the use of pumps with a variable bypass for controlling
the pressure in the system, the bypass being reduced or closed
during the times of high pressure demand.
[0107] As fuel injector technology improves over time, the use of
even higher pressure temporary injection (e.g. 20 MPa or higher)
could provide a further decrease in the particulate generation
described in conjunction with the approaches described above.
[0108] Another control option is to vary the start of direct
injection as a function of the point in the engine torque-speed map
at which the engine is operating. A look up table could be used for
this control feature. The dependence of the start of injection
could be based on the model described here or on laboratory
measurements of particulate generation as a function of start of
injection at various torque-speed points or a combination of both.
The start of direct injection could also depend upon engine
temperature. Information about engine temperature, including
information from a temperature sensor could be employed.
[0109] This direct injection control feature could be especially
important for high power operation where the direct injection pulse
length could be sufficiently long that the some of the direct
injection would occur outside of the two crank angle limits. The
control system could vary the direct injection timing so as to
reduce or eliminate the amount of direct injection that would
otherwise occur outside the selected crank angle limits that are
selected based on experimental data and/or modeling. In this way
particulate emission from piston wetting by directly injected fuel
could be minimized. Further reduction of particulate emissions at
high power operation can be obtained by use of spark retard to
reduce the amount of direct injection needed to prevent knock and
substitute port fuel injection for direct injection.
[0110] Because of the effect of lower temperature operation in
reducing the extent of the crank angle extent between the two crank
angle limits at lower temperatures, especially during cold start
(e.g. the first 100 seconds of engine operation), variation of the
direct injection timing with temperature in addition to variation
with other parameters such as torque and speed can also be
especially important.
[0111] The techniques for reducing particulate generation from fuel
that is directly injected when optimized PFI+DI is employed can
also be used for engines that employ only DI. The control of direct
injection timing as a function of torque and speed so as to reduce
the amount of direct injection outside of the two crank angle
limits as well as control based on engine temperature can be
employed in engines that use direct injection alone as well as
engines that use optimized PFI+DI. Use of higher pressure, both at
a constant value and in temporary operation could also be utilized
for the case of direct injection alone. Temporary use of higher
pressure for DI only operation could be especially impactful during
cold start or during periods of high power operation.
[0112] Summary of Computational Model Results
[0113] The model calculations indicate that matching the fraction
of fuel that is provided by direct injection to that needed for
knock suppression along with a 100% or a very high (greater than
95%) use of port fuel injection during cold start could reduce the
amount of DI-generated gasoline particulates from a gasoline engine
operated in the US06 cycle by at least 85% and preferably at least
90%. For the UDDS and HWFET drive cycles, the amount of
DI-generated particulates could be reduced by more than 95%.
[0114] Use of timing of the direct injection to begin at the first
crank angle degree limit and keeping the injection period
sufficiently short would further insure a greater than 98%
reduction in direct injection particulates for the UDDS and HWFET
cycles. Our computational modeling indicates that it could provide
a reduction of at least 95% for the US06 cycle for a fuel injector
pressure of 10 MPa. Direct-injection generated particulates could
be reduced by at least 95% for all US test drive cycles.
[0115] Temporary or continuous higher direct injector pressure
(e.g. in the 15-25 MPa range) or use of increased spark retard
could provide greater margin for achieving this level of direct
injection generated particulate reduction or a larger reduction
factor of direct-injection generated particulates (in mass and in
number), such as a reduction of at least 98% in all US test drive
cycles.
[0116] For engines that use port fuel injection in addition to
direct injection for other purposes (e.g. for higher efficiency
and/or higher performance, which also use a substantial amount of
port fuel injection), the additional cost for operation so as to
greatly reduce particulate emissions can be very low.
[0117] Optimized PFI+DI technology could provide greater reductions
in DI-generated particulate emissions in gasoline engines than the
80-90% provided by gasoline particulate filter (GPF) technology
along with the advantages of lower cost; a simpler, more reliable
system; and no adverse effects of back pressure on efficiency. It
could thus be used in a DI gasoline engine instead of a GPF.
Use of Optimized PFI+DI with a GPF
[0118] As an alternative to removing the use of a GPF, the
optimized PFI+DI technology could be used along with a GPF for
greater particulate reduction and/or reduction of GPF cost and
detrimental impact on fuel efficiency.
[0119] For example, greater particulate reduction could be obtained
by a lower cost GPF which could also reduce the already very low
engine-out particulate emissions by a modest factor (for example,
less than a factor of four). It could be possible to use a filter
with thinner wall or decreased cell density in order to trade
reduced back pressure (thereby improving fuel economy) but with
decreased trapping efficiency of the particulate matter (addressed
by reduced engine-out particle emissions). The back pressure
efficiency loss might be reduced by at least a factor of 2 (e.g.
from a 2% efficiency loss to less than a 1% efficiency loss). Use
of this more modest GPF could also provide reductions in
particulates generated from port fuel injection.
[0120] For the US 06 drive cycle, the combination of modest GPF
with a particulate reduction factor of four and optimized PFI+DI
could potentially reduce tail pipe direct-injection generated
particulates by a factor of at least 50 relative to the DI
injection alone. This reduction is much greater than the reduction
factor of 5 to 10 using DI injection alone plus a state of the art
GPF.
[0121] The combination of engine operation with optimized PFI+DI
followed by exhaust treatment with a GPF can provide a greater
reduction of both tailpipe particulate number and mass than the GPF
alone. The reduction in mass is largely achieved through a greater
reduction of ultra fine particulates (>25 nanometer). There is
also at least 5 times greater reduction of nanoparticulates (<25
nanometers) than use of optimized PFI+DI alone.
[0122] In addition, the reduction of particulates by use of the
optimized PFI+DI operation prior to treatment by the GPF can be
important for avoiding GPF problems during prolonged driving
conditions (for example, more than 10 minutes) where the GPF does
not warm up sufficiently to burn the particulates. For example,
under repeated operation with cold engine conditions, soot in the
GPF can build up to an unsafe condition. When the engine warms up
sufficiently to burn the soot after this prolonged period of soot
accumulation, there is a possibility of a thermal runaway,
depending on the oxygen concentration in the exhaust.
[0123] By use of optimized PFI+DI operation to greatly decrease the
particulate production during light load operation, the potential
for heavy particulate buildup in the GPF is greatly decreased.
Using optimized PFI+DI to decrease the rate of generation of
particulates by factors of 10 to 100, a given gasoline particulate
trap could operate for a time 10 to 100 times longer for building
up comparable soot in the trap. As a consequence of this increased
time, the probability of obtaining the high temperature needed for
soot oxidation is substantially increased, thereby avoiding most
cases of soot buildup and a possible uncontrolled exothermal event
that can damage the trap.
[0124] An additional benefit with very low soot formation, is that
it is possible to have low temperature soot oxidation by NO.sub.2
that is present in the exhaust flow. Engine exhaust when it has
excess oxygen (for example, during an excursion to a rich condition
during the dithering of the air fuel ratio commonly used to control
stoichiometric operation in the engine) could convert a fraction of
the NO generated into NO.sub.2, which is a powerful oxidant. There
could be engine operating conditions when the NO.sub.2 present in
the unit is sufficient to eliminate the soot in the GPF. This
operation is enabled by the reduced generation of the soot by the
PFI+DI approach to control particulates.
[0125] It may be possible to convert some of the NO into NO.sub.2
in the upstream regions of a 4-way catalyst (a 3 way catalyst
combined with a GPF) used to eliminate soot in the downstream
region of the 4-way unit. NO.sub.2 particulate elimination could be
used either in engines that are dithering around stoichiometric
conditions, or engines that use lean-rich cycles for NO control
[See, for example, Parks, J., Storey, J., Prikhodko, V., Debusk, M.
et al., "Filter-Based Control of Particulate Matter from a Lean
Gasoline Direct Injection Engine," SAE Technical Paper
2016-01-0937, 2016, doi:10.4271/2016-01-0937].
[0126] The GPF design can be optimized in a number of ways under
conditions for the low soot operation enabled by engine operation
with optimized PFI+DI. These include adjusting the size (diameter,
length), the wall thickness, the cell density (numbers of cells in
the honeycomb per unit cross sectional area of the GPF), the
porosity, the pore size distribution, the catalyst loading or the
material of the GPF.
[0127] If the engine particulate generation is decreased by a
factor .eta., the wall area in the GPF can be decreased by .eta.
for comparable soot loading and the volume can be decreased by
.eta..sup.1.5. The catalyst loading may be decreased by .eta., for
comparable soot loadings in the case with and without PFI+DI
particulate control. For example, for a comparable soot loading in
the GPF, the catalyst loading could be reduced by a factor of more
than 3 while still providing greater than three times lower
tailpipe particulate mass emissions than in the case with use of
the GPF alone. The amount of NO.sub.2 generated by the engine and
catalyst upstream of the GPF can also be decreased by the same
factor of .eta..
[0128] Thus, the GPF design and operation can be optimized to
increase durability (mostly due to ease of operation at low
temperature), reduce cost (due to reduced size and decreased
catalyst loading), and with GPF optimized for further reduction of
particulate emissions, especially of nano particulates when used
with an engine that is operated with optimized PFI+DI.
[0129] Thus when used in combination with a GPF, the PFI+DI
approach for controlling particulates can both provide greater
particulate reduction and enable improved GPF performance and lower
exhaust treatment cost. In addition to the option of using a
smaller GPF, optimized GPF construction or optimized use of a GPF
combined with the PFI+DI, the optimized PFI+DI approach can be used
with a 4-way system, where a 3-way catalyst is combined with a GPF.
The decrease of the particulate generation by the use of PFI+DI
prevents the need for substantial increase in the size of the 3-way
catalyst when it is converted to a 4-way catalyst and thereby
enables use of the 4-way catalyst without the additional cost that
would otherwise occur. One can could be used instead of two
cans.
[0130] A range of GPF design parameters are available from
increasing the pore size of the GPF and thereby reducing back
pressure from use of the GPF. The pressure drop from a GPF is
dependent on the pore size, the frontal area of the GPF, the
porosity of the GPF, the catalyst loading and the wall
thickness/cell size. [See, for example, Ito, Y., Shimoda, T., Aoki,
T., Yuuki, K. et al., "Next Generation of Ceramic Wall Flow
Gasoline Particulate Filter with Integrated Three Way Catalyst,"
SAE Technical Paper 2015-01-1073, 2015,
doi:10.4271/2015-01-1073]
[0131] Increasing the mean pore size decreases the pressure drop.
For example, increasing the pore size from 12 to 16 microns
decreases the pressure drop by over a factor of 2, while increasing
the particulate number that escapes the GPF by a factor of 2. Since
the PFI+DI approach decreases substantially the particulate number,
a GPF with a larger pore size (and thus a smaller pressure drop)
can be used in combination with a PFI+DI particulate control
approach. A reduction in back pressure by at least 30% and
preferably at least 50% can be achieved relative to what would be
the case by using the combination of the optimized PFI+DI approach
with the a GPF with a larger pore size. It should be noted that the
particulate trapping efficiency in the GPF is not a strong function
of the wall thickness or the cell density.
[0132] Similarly, the pressure drop across a 4-way catalyst is
dependent on the catalyst loading (usually containing a precious
metal). The catalyst is deposited in the pores of the GPF,
resulting in increased pressure drop in the GPF. By using PFI+DI as
a first approach to controlling particulate mass or number, the GPF
with larger pores can be used with the same catalyst loading. The
GPF with larger pores with the same catalyst loading may be
insufficient to satisfy the regulatory requirements, but when
combined with PFI+DI the particulate emissions can be below the
regulatory requirement. The combination of the optimized PFI+DI
approach with a changed GPF can thus enable use of a GPF in a 4-way
catalyst with lower back pressure.
[0133] The pressure drop is also a strong function of the pore
volume per unit volume (the ratio between the pore volume to the
total volume of the GPF). Increased pore volume per unit volume
decreases substantially the pressure drop, but may result in
structural challenges. The impact of the catalyst loading on the
pressure drop is substantial and is especially relevant for a 4-way
unit. In this case, the use of PFI+DI allows the use of increased
pore volume per unit volume, with larger porosities, controlling
the particulate emissions by a combination of PFI+DI and GPF and
operating with lower back pressure and also lower tailpipe
emissions that with use of a GPF alone.
[0134] The buildup of soot on the GPF also results in a substantial
increase in pressure drop. In addition, the resulting pressure drop
associated with increased soot loading is much larger when the pore
volume per unit volume decreases. The soot implants itself within
the pores, decreasing the flow area and increasing the pressure
drop. By decreasing the rate of generation of the particulates with
PFI+DI, it is possible to avoid large buildup of soot in the
GPF.
[0135] It is possible to use the engine management system described
in this invention with a sensor in the GPF. By sensing the amount
of soot in the GPF, it is possible to control the use of spark
retard or heavy EGR under conditions when the GPF is relatively
clean, with little soot buildup; the engine management can be set
up not to minimize the particulate production, but rather to
increase the efficiency and to possibly decrease other regulated
emissions. As the soot builds up in the GPF, with an associated
increase in pressure drop or danger of GPF plug-up, spark retard or
heavy EGR can be used to decrease the generation of particulates,
allowing for the soot already in the GPF to eventually combust.
[0136] Monitoring of the soot content in the GPF it is possible to
adjust the rate of generation of particulates in the engine, while
controlling the particulate loading of the GPF. It is important to
be able to sense light soot loadings in the GPR, and thus,
alternatives to pressure sensors may be required, such as microwave
sensing of the GPF, or other techniques. In addition, as it is
possible to also monitor the amount of soot and/or ash in the 4-way
unit, it is possible to control engine operation for improved
performance of the 4-way unit. For example, when there is increased
ash (as for example, as the unit ages) in the GPF, filtration
efficiency for control of particulates improves. Thus, the fuel
management system can adjust the PFI+DI control (including
injection timing, PFI/DI split, EGR, spark retard, valve timing) in
order to improve efficiency/and or other emissions, while still
meeting the particulate emission requirement (through the increased
trapping efficiency of the ash-loaded GPF).
[0137] In the case of the use of the catalyst for converting some
of the NO to NO.sub.2 for control of the CO, hydrocarbon and/or
particulate emission, the monitor can control the state of the
3-way catalyst (controlling, for example, the engine stoichiometry
or additional injection of air), in order to provide improved
conversion of the regulated emissions (CO, HC, NOx and
particulates). The engine operation can be such that the air/fuel
ratio dithers (that is, brief excursions from stoichiometry), or
through lean/rich periods of engine operation (i.e., longer periods
of excursion).
[0138] Similarly, by having sensors in the 3-way catalyst or the
4-way catalyst that can sense the loading of the 3-way catalyst
(for NO control), it is possible to adjust the oxygen excursion in
the engine to increase the availability of free oxygen for
assisting the burning of the soot in the GPF or in the 4-way unit.
The minimum temperature for regeneration of the particulate filter
is a strong function of the excess oxygen, and having controlled
oxygen excursions enables regeneration of the particulate filter
under driving conditions where the temperature of the exhaust is
limited.
Increased Efficiency
[0139] The fuel management system for optimized PFI+DI system for
direct-injection generated particulate reduction can be optimized
for also meeting the goals of increasing efficiency relative to the
use of DI injection alone. The PFI+DI combination over a drive
cycle can be varied to accomplish an optimum combination of various
objectives. For example, the control system can be designed so as
to operate with port fuel injection alone or with a very small
amount of direct injection over the cold start period for
particulate emissions (e.g. less than 5% of total fuel) to maximize
particulate reductions; and to maximize efficiency over the warmed
up period of the drive cycle.
[0140] An attractive mode of operation during the warmed up part of
the drive cycle could be to optimize the PFI/total fuel ratio for
the highest efficiency while preventing knock and using this
optimization in combination with optimal direct injection timing
and pulse length.
[0141] The PFI+DI combinations for maximizing efficiency during the
warmed up period of the drive cycle are likely to involve
minimization of direct injection which is generally aligned with
particulate reduction. Minimization of direct injection increases
better mixing combustion stability and can allow use of increased
EGR (internal or external).
[0142] These factors could increase drive cycle efficiency by at
least 2% relative to using direct injection alone. Removing the
need for a state of the art GPF that provides at least an 80%
reduction in particulates and thus eliminating the resulting back
pressure could provide another 1 to 2% efficiency gain relative to
using a GPF. The combination of these effects could provide an
efficiency gain of at least 3% and preferably at least 4% relative
to a vehicle that used direct injection alone plus a GPF.
[0143] Another option is to use a GPF with reduced particulate
reduction and lower back pressure than present GPFs (e.g. a factor
of 4 or less reduction in direct injection generated particulates)
which have lower cost and less back pressure. The combination of
such a GPF with use of optimized PFI+DI could potentially provide
more than a 98% reduction in direct-injection generated
particulates, a factor of 4 reduction in port fuel-injection
generated particulates and an overall efficiency gain of at least
2% and preferably at least 3% relative to use of direct injection
alone and a GPF which provides a particulate reduction of 80 to
90%.
Use with EGR
[0144] In still another application, the use of the control system
for PFI+DI can be combined with EGR (engine gas recirculation) to
provide reduced particulates. The use of increased EGR can be
employed to reduce the piston wetting by changing the viscosity of
the gas, either through changes of composition or temperature.
Changes in EGR can be used to adjust the critical crank angles.
Internal EGR would increase the temperature of the charge,
affecting the penetration of the fuel spray, as described above.
External EGR would mostly affect the composition of the charge, as
well as the temperature during the compression stroke (because of
the impact on the ratio of specific heats, gamma). The addition of
EGR could also modify the charge motion. Combustion stability could
improve by providing some additional fuel in the region near the
spark.
[0145] Appropriate injection timing of the fuel, such that piston
wetting is avoided, would also improve the combustion stability,
thereby enabling increased EGR. Increased EGR at light load could
increase the efficiency of the engine, mostly due to decreased
throttling and decreased heat transfer to the walls due to the
dilution, while at heavy load, increased EGR could decrease the
need for directly injected fuel.
Use in Dual Fuel Engines
[0146] A further application is for the particulate reduction
approaches described here for the case of port injection of and
direct injection of the same fuel to be also employed in an optimal
manner for the dual fuel case where the direct injection uses a
high ethanol concentration fuel.
Uses of Computational Model for Better Fueling System Design and
Control
[0147] The computational model described here can be employed for
fuel management system design and for control of the fuel
management system using lookup table information provided by the
computational model. The model determines the effects of BMEP,
engine speed, direct injector pressure, drive cycle and other
parameters on direct-injection generated particulate levels. It
includes both a model for autoignition generation of knock and a
model for the impacts of direct injection pulse length and timing.
Effects of spark retard and other knock suppression techniques can
also be included. The computation model could be used for
optimizing operation of DI fueling alone as well as for the use of
PFI+DI.
[0148] Different control techniques for cold start operation can
additionally be incorporated as more information becomes available.
In addition, the use of PFI+DI to increase efficiency can be
included. The impacts of combinations of PFI+DI and gasoline
particulate filters can also be incorporated in the model. The
model can be used to determine direct-injection generated
particulate levels over a wide range of operating conditions
(including use of direct injection alone over the entire drive
cycle).
[0149] Computational tools, such as CFD, can be used to determine
the characteristics of the piston and injector that result in
non-wetting of the piston. Models of the injectors, including spray
pattern, droplet size and droplet velocity can be used to
investigate the droplet cloud, droplet evaporation and impact of
the injector on fluid motion in the cylinder. These CFD
calculations for spray penetration can be used to determine more
accurate values of .theta..sub.1 and .theta..sub.2 as a function of
engine conditions, such as cylinder pressure, engine speed and
engine temperature. The timing of direct injection, particularly
the start of direct injection, could then be changed as a function
of the parameters that change the values of .theta..sub.1 and
.theta..sub.2. This adjustment of timing can be used to reduce the
amount of direct injection that occurs before .theta..sub.1 and
after .theta..sub.2 and to thus reduce the particulate generation
from piston wetting that would otherwise occur.
[0150] The present disclosure is not to be limited in scope by the
specific embodiments described herein. Indeed, other various
embodiments of and modifications to the present disclosure, in
addition to those described herein, will be apparent to those of
ordinary skill in the art from the foregoing description and
accompanying drawings. Thus, such other embodiments and
modifications are intended to fall within the scope of the present
disclosure. Further, although the present disclosure has been
described herein in the context of a particular implementation in a
particular environment for a particular purpose, those of ordinary
skill in the art will recognize that its usefulness is not limited
thereto and that the present disclosure may be beneficially
implemented in any number of environments for any number of
purposes. Accordingly, the claims set forth below should be
construed in view of the full breadth and spirit of the present
disclosure as described herein.
* * * * *