U.S. patent application number 16/242185 was filed with the patent office on 2019-05-16 for gasoline particulate reduction using optimized port and direct injection.
The applicant listed for this patent is Ethanol Boosting Systems, LLC. Invention is credited to Leslie Bromberg, Daniel R. Cohn.
Application Number | 20190145341 16/242185 |
Document ID | / |
Family ID | 54869231 |
Filed Date | 2019-05-16 |
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United States Patent
Application |
20190145341 |
Kind Code |
A1 |
Cohn; Daniel R. ; et
al. |
May 16, 2019 |
Gasoline Particulate Reduction Using Optimized Port And Direct
Injection
Abstract
Additional approaches for the reduction of particulate emissions
in gasoline engines using optimized port+direct injection are
described. These embodiments include control of the amount of
directly injected fuel so as to avoid a threshold increase in
particulates due to piston wetting and reduction of cold start
emissions by use of air preheating using variable valve timing.
Inventors: |
Cohn; Daniel R.; (Cambridge,
MA) ; Bromberg; Leslie; (Sharon, MA) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Ethanol Boosting Systems, LLC |
Cambridge |
MA |
US |
|
|
Family ID: |
54869231 |
Appl. No.: |
16/242185 |
Filed: |
January 8, 2019 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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15691895 |
Aug 31, 2017 |
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16242185 |
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15208120 |
Jul 12, 2016 |
9840980 |
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15691895 |
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14840688 |
Aug 31, 2015 |
9441570 |
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15208120 |
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14391906 |
Oct 10, 2014 |
9435288 |
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PCT/US13/73334 |
Dec 5, 2013 |
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14840688 |
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61734438 |
Dec 7, 2012 |
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62044761 |
Sep 2, 2014 |
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62128162 |
Mar 4, 2015 |
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62079885 |
Nov 14, 2014 |
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Current U.S.
Class: |
123/478 ;
701/104 |
Current CPC
Class: |
F02D 2250/38 20130101;
F02D 2041/001 20130101; Y02T 10/40 20130101; F02D 41/047 20130101;
F02P 5/145 20130101; Y02T 10/18 20130101; F02D 41/0002 20130101;
F02D 13/0215 20130101; F02D 41/064 20130101; F02D 2013/0292
20130101; F02M 69/046 20130101; F02D 41/3094 20130101; F02D 35/027
20130101; F02P 5/152 20130101; Y02T 10/12 20130101; F02D 13/0242
20130101; Y02T 10/46 20130101; F02D 37/02 20130101; F02D 41/3023
20130101; F02P 5/1506 20130101 |
International
Class: |
F02D 41/30 20060101
F02D041/30; F02P 5/145 20060101 F02P005/145; F02D 41/06 20060101
F02D041/06; F02D 13/02 20060101 F02D013/02; F02D 41/04 20060101
F02D041/04; F02P 5/15 20060101 F02P005/15; F02M 69/04 20060101
F02M069/04; F02D 41/00 20060101 F02D041/00; F02D 37/02 20060101
F02D037/02; F02D 35/02 20060101 F02D035/02; F02P 5/152 20060101
F02P005/152 |
Claims
1. A spark ignition engine where temporary advanced closing of the
exhaust valve is used to provide hot gas which goes into the
manifold; and where a control system determines when the exhaust
valve timing returns from the advanced closing to a less advanced
closing.
2. The spark ignition engine of claim 1 where the temporary
advanced closing reduces particulate generation relative to what it
would have been if the temporary advanced exhaust valve timing were
not used.
3. The spark ignition engine of claim 1 where the temporary
advanced closing reduces hydrocarbon emission relative to what it
would have been if the temporary advanced exhaust valve timing were
not used.
4. The spark ignition engine of claim 1 where the exhaust valve
timing is temporarily advanced by between 30 and 60 crank angle
degrees.
5. The spark ignition engine of claim 1 where the control system
uses closed loop control where information is provided by a sensor
to determine when to stop the use of the advanced exhaust valve
timing.
6. The spark ignition engine of claim 1 where the control system
uses open loop control to determine when to stop using the advanced
exhaust valve timing.
7. The spark ignition engine of claim 1 where the advanced exhaust
valve closing is controlled using information about the engine
temperature.
8. The spark ignition engine of claim 1 where port fuel injection
is employed.
9. The spark ignition engine of claim 1 where direct fuel injection
employed.
10. The spark ignition engine of claim 1 where the temporary
advanced valve closing is used during cold start and where the hot
gas assists in evaporation of liquid fuel from the manifold.
11. A spark ignition engine where temporary advanced closing of the
exhaust valve is used during cold start to provide hot gas which
goes into the manifold; and where electricity is used to crank the
engine while the temporary closing of the exhaust valve is
employed.
12. The spark ignition engine of claim 11 where temporary advanced
closing of the exhaust valve is used for one engine cycle.
13. The spark ignition engine of claim 11 where the temporary
advanced closing of the exhaust valve is used for more than one
engine cycle.
14. The spark ignition engine of claim 11 where the power of the
electricity that is used to crank the engine is between 100 and
1000 watts.
15. The spark ignition engine of claim 11 where an electrically
powered oil pump is used.
16. The spark ignition engine of claim 11 where a bistable cam is
used.
17. The spark ignition engine of claim 11 where all electric valve
timing or electric assist valve timing is employed.
18. The spark ignition engine of claim 11 where a spark is not used
when the temporary advanced exhaust valve timing is employed.
Description
[0001] This application is divisional of U.S. patent application
Ser. No. 15/691,895 filed Aug. 31, 2017, which is a continuation of
U.S. patent application Ser. No. 15/208,120 filed Jul. 12, 2016
(now U.S. Pat. No. 9,840,980 issued Dec. 12, 2017), which is a
continuation of U.S. patent application Ser. No. 14/840,688 filed
Aug. 31, 2015 (now U.S. Pat. No. 9,441,570 issued Sep. 13, 2016),
which is a continuation-in part of U.S. patent application Ser. No.
14/391,906 filed Oct. 10, 2014 (now U.S. Pat. No. 9,435,288 issued
Sep. 6, 2016), which is a National Stage entry of PCT/US13/73334,
filed Dec. 5, 2013, which claims priority of U.S. Patent
Application Ser. No. 61/734,438 filed Dec. 7, 2012, the disclosures
of which are incorporated herein by reference in their entireties.
U.S. patent application Ser. No. 14/840,688 also claims priority of
U.S. Provisional Patent Application Ser. No. 62/044,761, filed Sep.
2, 2014, U.S. Provisional Patent Application Ser. No. 62/128,162,
filed Mar. 4, 2015 and U.S. Provisional Patent Application Ser. No.
62/079,885 filed Nov. 14, 2014, the disclosures of which are
incorporated herein by reference in their entireties.
BACKGROUND
[0002] There is increasing concern about particulate matter (PM)
emissions from gasoline engine vehicles. The concern is driven by
the substantially higher emissions of small particulates from spark
ignited gasoline powered vehicles that use direct injection (DI) of
gasoline into at least one of the engine cylinders as a liquid.
These small particulates lodge in the lungs and can be injurious to
human health.
[0003] Although direct injection increases engine efficiency and
performance by increasing knock resistance though evaporative
cooling, use of DI throughout a drive cycle substantially increases
the particulate emissions. Relative to conventional port fuel
injected (PFI) engines, the particle number when operating with
direct injection increases by factors of 10-100 over a drive cycle,
depending on the cycle and the engine operating conditions. The
emissions are especially concerning for engines that are
turbocharged and this would also be the case for supercharged
engines.
[0004] More stringent regulations on PM 2.5 (particulate matter
less than 2.5 microns in diameter) are planned for Europe and
anticipated in the US, including both EPA and California
regulations. The European regulations would apply to the number of
particulates as well as to the amount of particulate mass that is
emitted.
[0005] Therefore, techniques that improve engine performance while
minimizing particulate emissions would be beneficial.
SUMMARY OF INVENTION
[0006] Additional approaches for the reduction of particulate
emissions in gasoline engines are described. These embodiments
include control of the amount of directly injected fuel so as to
avoid a threshold increase in particulates due to piston wetting
and reduction of cold start emissions by use of air preheating
using variable valve timing.
BRIEF DESCRIPTION OF THE DRAWINGS
[0007] For a better understanding of the present disclosure,
reference is made to the accompanying drawings, which are
incorporated herein by reference and in which:
[0008] FIG. 1 is an illustrative model prediction of threshold BMEP
(brake mean effective pressure) for preventing direct injection
generation of particulates for warmed up engine conditions.
Operation below the line prevents particulates, operation above the
line generates particulates. Brake mean effective pressure
corresponds to torque for a given volume of the engine
cylinders.
[0009] FIG. 2 shows illustrative model calculations for particulate
emissions as function of BMEP at 2000 rpm.
[0010] FIG. 3 shows illustrative model calculations for particulate
matter generation in arbitrary units as a function of brake-mean
effective pressure (BMEP), for several engine speeds.
[0011] FIG. 4 shows particulate generation (arbitrary units) over
the engine operating map. Contours are shown for arbitrary units of
5 and 10.
[0012] FIG. 5 shows an illustrative fraction of gasoline that needs
to be directly injected in order to prevent knock for a
turbocharged engine with a compression ratio of 10.
[0013] FIG. 6 shows illustrative particulate matter (PM) generation
(in arbitrary units) for a turbocharged engine using PFI/DI and
minimization of the DI fuel, having a compression ratio=10.
[0014] FIG. 7 shows the DI fraction of fuel for an engine with a
manifold air pressure (MAP) of 1.7 bar (absolute), for both the
UDDS and US06 cycles as a function of time.
[0015] FIG. 8 shows an illustrative ratio between DI and PFI as a
function of torque, for a given engine speed, so as to constrain
particulate emissions and prevent knock while utilizing a high
fraction of DI at low torque. This is a representative scenario for
stratified direct injection.
[0016] FIG. 9 shows an engine control system for adjusting engine
operation and/or ratio of directly injected fuel and port-fuel
injected fuel to reduce particulate reduction with minimum drive
cycle efficiency reduction.
[0017] FIG. 10 shows normal (top) inlet and exhaust valve lifts and
advanced exhaust valve lift. The inlet conditions remain constant
for both.
[0018] FIG. 11 shows pressure (left) and temperature (right) for
conditions of normal valve timing (top) and advanced exhaust valve
timing (bottom). Note the increase in inlet temperature.
[0019] FIG. 12 shows mass flow rate through the inlet manifold
valve for the case of advanced exhaust valve timing.
[0020] FIG. 13 shows gas velocity across the inlet manifold for the
case of advanced exhaust valve timing.
DETAILED DESCRIPTION
[0021] As discussed in co-pending patent application WO2014/089304,
improved approaches to control particulate emissions from spark
ignited gasoline engines have been developed. These approaches
involve optimized use of port fuel injection in combination with
direct injection where the good mixing provided by port fuel
injection (PFI) produces much less particulate emissions than
direct injection. In these approaches, the fuel management system
minimizes the amount of direct injection by optimal use of port
injection while maximizing engine performance and efficiency
through the use of direct injection.
[0022] A basic approach that is used is increasing the fraction of
fuel that is introduced into the engine cylinders by direct
injection so that it is substantially equal to the amount needed to
suppress knock as the engine operating condition (torque, speed)
changes. Continual matching of fraction of fuel that is directly
injected to that needed to prevent knock throughout all the torque
range, or if not all, the high end of the torque range where direct
injection is needed to prevent knock, minimizes the amount of
directly injected fuel. When more knock resistance is called for,
the direct injection fraction is increased and when less knock
resistance is needed, the direct injection fraction is reduced. The
matching can follow the ups and down of higher torque operation
throughout the engine drive cycle. When direct injection is not
needed for knock control it can be set to zero. Closed loop control
using a knock detector together with open loop control using a look
up table that relates engine parameters to required knock
resistance can provide a highly responsive means of matching the
fraction of fuel that is directly injected so as to provide
required knock resistance as the torque changes.
[0023] The fuel management control system can be employed to
operate the engine with only port fuel injection or with both port
and direct injection or with direct injection alone depending on
engine conditions and on engine performance requirements.
[0024] In addition, the fuel management system can also further
reduce particulate emissions by making adjustments that reduce the
fraction of fuel that is directly injected during those portions of
the drive cycle when particulate emissions are especially high.
These portions of the drive cycle include cold start and certain
portions of the warmed-up engine part of the drive cycle. During
these portions of the drive cycle, adjustments are made so that the
fraction of fuel that is directly injected is lower than it would
otherwise be to avoid knock. The adjustments include increasing
spark retard and variable valve timing. They also can include
open-valve port fuel injection where open-valve port fuel injection
is used to provide vaporization cooling instead of direct
injection.
[0025] Measurements of particulate emissions have shown that they
are very high during a cold start period of the first 100 seconds
or so after the engine has been started. Minimizing the fraction of
fuel that is directly injected as torque increases and making
adjustments, such as increasing spark retard, can be especially
important throughout the entire torque and speed range during this
cold start portion of the drive cycle. Variable valve timing and/or
open-valve port fuel injection could also be used in this cold
start period of the drive cycle.
[0026] Measurements of particulate emissions during warmed up
engine operation (which occurs after the engine has been operated
for around 100 seconds or so) indicate that particulate emissions
are also especially high at high levels of torque and speed.
Optimized use of the fuel management system during the high torque
and high speed portion of the warmed up cycle can require a
different control approach than that used in cold start and certain
other transient conditions.
[0027] This disclosure describes additional approaches for
particulate reduction in both cold and warmed up engine
operation.
[0028] In order to provide a basis for the control system during
the warmed up engine portion of the drive cycle, a model for
particulate emissions based on piston wetting has been developed.
This model is then used to provide additional fuel management
control to further reduce particulate emissions. Although this
model has been developed for warmed up operation, it may also
provide some degree of applicability for cold start operation.
[0029] Other means for further optimizing the combined use of port
and direct injection for gasoline engine particulate reduction are
also described. They include control system operation optimized for
stratified direct injection. Stratified direct injection can
provide increased efficiency through dilute and open throttle
operation at low torque but also increases particulate
emissions
[0030] The combination of these approaches can make it possible for
drive cycle particulate emissions from a gasoline engine with
optimized direct+port fuel injection to be reduced to at least less
than 1.2 times the drive cycle particulate emissions for a
comparable PFI only fueled engine and to preferably be less than
1.1 times that of a comparable PFI only engine.
[0031] This could make it possible for environmental regulations to
be met without the cost, durability and efficiency reduction
concerns of adding a gasoline particulate filter (GPF) exhaust
aftertreatment system. Alternatively, this technology could be used
in combination with GPF exhaust treatment to substantially reduce
the cost, reliability, durability and efficiency drawbacks of
gasoline particulate filters. Further, this technology, in
combination with GPF treatment, could provide greater particulate
reduction than the use of a GPF system alone.
[0032] In this disclosure, additional aspects of the use of engine
air heating enabled by variable valve timing as a means to reduce
particulate emissions from port fuel injection are also
covered.
Wall Wetting Model for Particulate Matter Production
[0033] A simple heuristic model for particulate matter (PM)
production from wall wetting has been developed. This model
provides a means to assess the particulate suppression impacts from
changes in the combination of port and direct injection during
various times in the drive cycle.
[0034] It is believed that the increase in particulate emissions
from DI fueling is mainly due to liner or piston wetting, when fuel
droplets hit the surfaces and make a liquid film. There are means
of avoiding the spray from wetting the piston and/or the liner.
Smaller aerosols, with lower inertia-to-drag (thus, more attached
to the air flow and less likely to be separated by acceleration),
can be used. The increase in pressure required for atomizing the
droplets can result in increased penetration of the spray, but the
reduced tendency of the smaller droplets to separate from the flow
results in a lower wetting fraction of the fuel. The timing of
injection can also be adjusted. There is a tradeoff between
improved mixing early in the intake stroke and decreased distance
between the piston and the injector tip.
[0035] This heuristic model assumes that PM is generated as a
result of wall wetting, and that it is proportional to the amount
of fuel on the piston (in other words, liner wetting is not
assumed).
[0036] The amount of fuel that wets the piston depends on the
amount of fuel injected during that time when wall wetting is
likely. It takes time for the spray to penetrate to the location of
the piston, changing the overall flow pattern in the cylinder.
Hence, for short injection on-times, the spray pattern does not
make it to the piston and there is no wall wetting. With longer
injections, which is a consequence of longer injection times at
higher loads, the gas pattern changes (modified by the spray) and
there is piston wetting. At that point, the rate of fuel that hits
the piston is constant. Thus in this model, the amount of fuel on
the piston follows a displaced-linear relationship of the total
amount of fuel injected: there is a threshold torque, also known as
"load" or brake mean effective pressure (BMEP) below which it is
possible to inject all the needed fuel without fuel impingement,
followed by a linear growth until the highest torque (BMEP) is
reached.
[0037] There is a limiting crank angle during the intake that
avoids impingement. Beyond that crank angle during the intake,
there is no impingement. Similarly, there is a crank angle during
compression where impingement begins. There is no impingement
before that crank angle during the compression, and there is
impingement after it. If injection occurs between the two limits,
there is no impingement and very little or no soot formation.
[0038] FIG. 1 shows a model calculation for the maximum BMEP (brake
mean effective pressure) that results in avoidance of fuel
impingement (which corresponds to injection time between the two
crank-angles in the intake and compression strokes described
above), for an illustrative set of directly injected engine
parameters and fuel injection rates (assumed to be constant). The
BMEP threshold increases with decreasing engine speed, as there is
more time for injection that avoids fuel impingement. For a given
size engine, BMEP correspond to torque.
[0039] The model assumes that there is a sufficiently short on-time
for injection (in crank angle degrees) that prevents wall wetting
altogether. Since DI has approximately constant rate of fuel
injection (determined by the fuel pressure and the injector
characteristics), requiring less fuel results in decreased injector
on-time. In direct injection (as in the case of port-fuel
injection), the amount of fuel injected is controlled by adjusting
the injector on-time (using PWM, or Pulse Width Modulation). The
start of injection is a compromise between good mixture preparation
and preventing wall wetting (either cylinder lining or piston).
[0040] PM generation has been measured as a function of injection
timing (SOI-start of injection). The spray wets the piston/liner
either with very early injection in the intake stroke, or late
injection during the compression stroke. To minimize particulate
generation, injection should not occur earlier than what results in
piston wetting during the intake stroke or later than what results
in piston wetting during the compression stroke. There is a window
of piston location where direct injection does not result in piston
wetting. Under these conditions, with light load and low engine
speeds, there can be a wide range of SOI that results in minimal
particulate matter.
[0041] Particulates are mostly produced during transients
(acceleration) and at high load (and in cold start).
[0042] The particulate production is directly related to the amount
of impingement. Thus, at a constant engine speed, there is a torque
and a corresponding amount of direct injection below which there is
no PM generation, and PM generation increases linearly after this.
Consequently, reducing the amount of directly injected fuel causes
the particulate emissions to drop linearly and below a certain
value to drop very dramatically. This feature indicates that, by
minimizing the amount of direct injection by increasing it to
substantially only to the fraction of fueling needed to prevent
knock as torque is increased, there can be a large impact in
reducing particulate emissions.
[0043] FIG. 2 shows a calculation for particulate production for
the same engine parameters as in FIG. 1, operating at 2000 rpm,
with all the fuel introduced through the direct injector, as a
function of load. The threshold BMEP for production of particulates
is about 7 bar BMEP. Above that, the particulate production is
linear with injection until the end of injection, and thus is a
linear function of BMEP in excess of the threshold BMEP.
[0044] FIG. 3 shows the particulate matter generated using the
heuristic model for several engine speeds. For gasoline DI engines,
there is evidence of a correlation between the particulate number
density and the mass, as a result of a relatively consistent size
distribution of the particulate matter (particle mass is roughly
proportional to number concentration). Thus, for assessment of the
general implications of this model, it is assumed that FIG. 3 is
applicable to either mass or number density.
[0045] The results of the model for direct injection generated
particulate concentration in the engine exhaust are shown in FIG.
3. At the higher engine speeds and loads, the total mass and number
of particulate emission increases as the total fuel per injection
increases and the time allowed for non-wetting injection decreases.
According to the model, for a given engine speed, the change in
particulate matter generation with increasing BMEP is 0 up to a
given BMEP and then increases linearly with increasing BMEP. Since,
for a given size engine, BMEP corresponds to torque, FIG. 3 is also
a plot of particulate emissions as a function of torque and
speed.
[0046] The BMEP at which the onset of this change occurs
corresponds to a given amount of fuel that is directly injected
into the engine. Thus, the model indicates a flat dependence of
particulate emissions with increasing amount of fuel up to a given
amount of directly injected fuel followed by the onset of a linear
rise of particulates with an increasing amount of directly injected
fuel. Because the model is an approximation, a useful more general
description of this dependence of particulate matter on the amount
of directly injected fuel is that above a threshold level of the
amount of directly injected fuel, the particulate emissions undergo
a large percentage increase relative to the zero or near level
below the threshold level. This is referred to as a "threshold
increase".
[0047] FIG. 4 shows the particulate emission, with this model, over
the engine map. The PM generation is in arbitrary units. With this
information, different means of minimizing particulate emission may
be compared by comparing the engine maps, and then using the engine
maps to project implications of use of the approach over a driving
cycle.
[0048] Spray entrainment in the gas could modify the above model.
During direct injection, the spray is entrained in the gas flow,
which prevents impingement on the piston. As injection continues,
the gas flow is modified by interaction with the spray, and the
spray gets to the piston. As in the previous model, there is a time
during which there is no wall impingement. Afterwards, the spray
impingement in the piston increases linearly with load (and thus,
injected fuel and injection time). As the ratio between the fuel
injected and the density of air in the cylinder are constant, the
amount of fuel that impinges on the piston is constant relatively
independent on load. Thus, it is not expected that flow entrainment
issues would substantially modify the above model.
Combined Use of PFI and DI
[0049] The main benefit of using direct injection is generally to
increase knock resistance by vaporization cooling. This is
particularly important in turbocharged or supercharged engines. Use
of port fuel injection makes it possible to use direct injection
only in an amount needed to prevent knock. Optimized control of the
combination of port and direct injection can minimize the amount of
direct injection while providing the knock resistance where needed
to maximize engine performance and efficiency. Both closed loop
control using knock detection and other sensors and open loop
control can be employed.
[0050] The model described above can be used to determine
particulate emissions of combined PFI and DI operation that is
employed to minimize their generation by matching the DI use to the
amount needed to is the amount to prevent knock.
[0051] This is done using maps of the requirement for the fraction
of fuel that must be directly injected in order to prevent knock.
FIG. 5 shows a typical result, for a pressure boosted DI/PFI engine
using regular gasoline and with a maximum manifold air pressure
(MAP) of 1.7 bar and a compression ratio of 10. As shown in FIG. 5,
the fraction of directly injected fuel that is required to prevent
knock varies with both torque and speed. At a given torque, the
fraction of fuel that is directly injected decreases with
increasing speed. Thus, if 100% direct injection is required for
knock control at highest torque at low speeds, the fraction of fuel
that needs to be directly injected can be less than 100% at higher
speeds. The engine may operate with 100% of directly injected fuel
at high speeds and less than 100% at other speeds. Alternatively,
it may operate with less than 100% of directly injected fuel
throughout the highest torque regime with less directly injected
fuel used at high speeds.
[0052] It is assumed in FIG. 5 that the port fuel injected fuel is
introduced into the engine in the conventional way with the inlet
valve closed and hence produces essentially no vaporization
cooling. The results of the model can be used to heuristically
describe the particulate emissions behavior for both naturally
aspirated engines and for engines that are pressure boosted by
turbocharging or supercharging.
[0053] No effect of spark retard is assumed in FIG. 5. Increased
spark retard at a given brake mean effective pressure and speed
would reduce the fraction of fuel that would need to be directly
injected and thus reduce particulate emissions. Spark retard can be
selectively applied to minimize adverse effects on efficiency and
performance for a given amount of particulate emission.
[0054] Using the information from FIG. 4 that estimates the PM
generation over the map for a fully directly injected engine
together with that of FIG. 5, where only enough directly injected
fuel is used for knock avoidance, the PM generation over the engine
operation map can be calculated. By injecting a fraction of the
fuel by means of port fuel injection, the amount of directly
injected fuel decreases, and it can be possible to reduce the time
of injection below the maximum injection time allowed without fuel
impingement on the piston. Thus, a higher level of torque can be
used without exceeding the amount of directly injected fuel that
would cause the threshold level for fuel impingement and
particulate emission to be exceeded.
[0055] Thus, in addition to linearly reducing particulate emission
throughout the drive cycle by reducing the fraction of fuel
provided by direct injection, it is also possible to further reduce
it by a large amount, or possibly essentially eliminate it
altogether according the present model, over a substantial region
of the operating map. This is accomplished by the fuel management
system controlling the amount of fuel that is directly injected so
that it is kept below a selected level above which wall wetting
would increase in a non-linear way and there would be rapid
increase.
[0056] In addition to being used in combination with the
information in FIG. 4, the information in FIG. 5 can also be
employed to estimate the fraction of fuel that needs to be directly
injected (to prevent knock) over a drive cycle. For the UDDS drive
cycle, the fraction is around 1%. For the aggressive US 06 drive
cycle, the fraction is around 10%. For the combined city-highway
drive cycle, the fraction is around 5%. This drive cycle
information can be used to make a rough estimate of the relative
amount of particulate emissions for PFI+DI operation versus DI
operation in the case where particulate reductions from the model
for wall wetting are not taken into account. This could be a useful
way to estimate the effect of PFI+DI in reducing particulate
emissions during cold start.
[0057] For example, if the ratio of particulate emissions for DI to
particulate emissions for PFI is R, an estimate of the fractional
increase in particulate matter that is emitted is that it is equal
to (1-f)+(f*R), where f is the fraction of directly injected fuel
used in a drive cycle. Thus if R=10 and for the US 06 cycle, f=0.1,
the fractional increase in particulate emissions is 0.9+(10) (0.1),
or .about.2 greater than PFI emissions. The reduction in cold start
emissions can be further reduced by various adjustments, such as
spark retard or variable valve timing. For the UDDS and combined
city-highway drive cycles, the fractional increase without
adjustments would be much smaller. It can be possible to reduce the
amount of particulate emissions using a 100 second cold start
period by more than 80% without spark retard and by more than 90%
with spark retard relative to what it would be if only direct
injection were employed.
[0058] These estimates indicate that with minimization of the
fraction of fuel that is directly injected and with various
adjustments, the engine-out particulate emissions from engines
using optimized PFI+DI operation can be reduced to the low levels
approaching those of PFI engines. Optimized PFI+DI operation could
provide direct injection particulate reduction that is comparable
to that provided by gasoline particulate filters that have a
particulate removal efficiency of around 90%.
[0059] In addition to the wall wetting effect, there could also be
fuel vaporization effects that could also result in a rapid
increase in particulate emissions above certain combinations of
torque and speed. The fuel management system could also use
information regarding the instantaneous fueling rate and engine
speed as a basis for determining the values of torque and speed at
which control adjustment is needed.
[0060] FIG. 6 shows the results of the model for a turbocharged or
supercharged gasoline DI/PFI engine based on the information in
FIGS. 4 and 5. The BMEP is such as to enable a downsizing by a
factor of 1.7 over a naturally aspirated engine that does not use
direct injection. The compression ratio is 10. The amount of DI
fuel that is used is minimized by matching the amount to that
needed to prevent knock at a given value of torque and speed and
providing the rest of the fuel by PFI. No change in spark retard is
assumed. Minimization of particulate emissions is obtained by
matching the fraction of fuel that is directly injected to that
needed to prevent knock over at least the torque and speed range
where the amount of directly injected fuel would otherwise be
greater than the threshold level.
[0061] The model shows that the particulate emissions have a strong
dependence on BMEP, or equivalently torque, as a result of the
combination of the dependence of particulate emissions on the
amount of directly injected fuel, equivalent to a combination of
torque and speed, and the dependence of the fraction of fuel that
must be directly injected to prevent knock. For the set of
assumptions used in the model, it is possible to somewhat reduce PM
generation at the highest loads and altogether eliminate them
through most of the operating map.
[0062] The model has been used to determine emissions for the UDDS
and the US06 cycles. These model results are meant to serve more as
general guidance for optimizing particulate control rather
providing accurate numerical values for engine operation.
Particulate reduction can be tuned to the desired level by using
spark retard and other adjustments to obtain the desired
particulate reduction for actual engine operation. For a given
engine speed, there is a threshold torque above which a threshold
increase in particulate emissions occurs. This threshold torque can
be increased with spark retard. Other adjustments, such as variable
valve timing, can also be employed to increase the threshold
torque.
[0063] The results are shown in Table 1. The calculations also do
not include the PM generated during cold start, which is not
captured by the model. They also do not include the effect of
increased spark retard.
[0064] Based on the model, because UDDS is such a light load cycle,
it does not result in any particulate emission, even with use of DI
alone throughout the drive cycle. In contrast, there are relatively
large particulate emissions for the US 06 cycle using DI alone and
these emissions are substantially reduced by substitution of port
fuel injection for direct injection. Table 1 shows that particulate
emissions are reduced by more than 90% relative to the use of DI
alone. Use of spark retard could substantially further decrease the
particulate emissions for the US 06 cycle using PFI/DI. Based on
these model results for the case where spark retard is not employed
and the large impact that spark retard can have in reducing the
fraction of fuel that must be directly injected so as prevent
knock, the particulate emission level during warmed up engine
operation can be reduced by at least 90% relative to what it would
have been if only direct injection were utilized.
TABLE-US-00001 TABLE 1 Relative particulate matter number for DI
alone and PFI/DI for the UDDS and US06 drive cycles DI PFI/DI UDDS
0.00 0.00 US06 1 0.07
[0065] The fraction that must be directly injected in order to
prevent knock (borderline knock) is shown in FIG. 7 for both the
UDDS and for the US06 cycles, again for a downsizing of about 1.7.
The instantaneous fuel usage for the cycle is shown in FIG. 7, as a
function of time during the cycle (about 1400 s for the UDDS and
600 s for the US06). There is very little DI in the UDDS, and thus
very little impingement on the piston.
[0066] The model results show that because of the threshold effect,
minimizing use of direct injection by continually matching the
fraction of fuel that is directly injected to that required to
prevent knock throughout the range in which direct injection is
needed for knock prevention can have a large impact on particulate
reduction.
Control System Adjustments
[0067] In addition to reducing particulates by matching the
fraction of fuel that is directly injected to the requirement to
prevent knock as torque and speed change and thus minimizing its
use, there are a number of other control features that can further
reduce particulate emissions in warmed up engine operation.
[0068] Adjustments can be made at certain values of torque and
speed to reduce the fraction of fuel that is directly injected so
as to increase the torque at which particulate emission starts to
rapidly grow through the onset of piston wetting, as illustrated in
the model results in FIG. 6. This reduces the torque-speed region
in which piston wetting leads to substantial particulate emissions.
The adjustments reduce the direct injection fraction of the fuel
that is required. They include but are not limited to increased
spark retard, variable valve timing and/or open valve port fuel
injection.
[0069] The fuel management system can be operated so as to minimize
the drive cycle fuel efficiency decrease for a given amount of
drive cycle particulate reduction provided by use of an adjustment.
The adjustment would be used in certain ranges of torque and speed
where it would provide the greatest reduction in particulate
emissions for a given a given decrease in drive cycle
efficiency.
[0070] The level of the adjustment could be matched to the need to
prevent knock as torque is increased without increasing the amount
of fuel that is directly injected. For example, an increase in
spark retard could be employed at low torque end of a given torque
range and continuously increased as the torque increased so as to
prevent knock.
[0071] An adjustment could also be used to reduce particulate
emissions when piston wetting is occurring and there is a linear
dependence of particulate emissions on the amount of directly
injected fuel
[0072] Another adjustment that can be made is to temporarily
increase the pressure of the fuel in the injector. This enhances
this PM reduction opportunity by achieving several objectives.
Because of the higher fuel pressure, the delivery rate increases.
With shorter injection times, piston wetting can be avoided at all
but the highest loads and engine speeds. The increased fuel
pressure also results in smaller droplets. The smaller droplets
evaporate faster and are more likely to be entrained in the flow,
instead of separating (due to inertial forces) from the flow due to
centrifugal acceleration when the gas turns around before a solid
surface. The temporary increase in the pressure of the direct fuel
injector can be used when it has the greatest impact in reducing
particulate emissions. Examples are use in the high torque regime
and in the regime of high torque and speed.
Stratified Injection
[0073] Although the model results and control approaches that have
been described above are for directly injected fuel that is
injected in a uniform way into the engine cylinders, they can also
be applied to stratified direct injection. The stratified direct
injection is used to increase efficiency by facilitating dilute and
open throttle operation at low loads.
[0074] At low loads, the fuel could be supplied entirely by
stratified DI or mainly by stratified DI. Thus, at low loads, the
fraction of fuel provided by DI would be much greater than the zero
or small fraction of fuel by DI that is needed to prevent knock in
this low torque region. As the torque increases, knock would be
prevented by the use of the high fraction or complete use of DI.
However, at a sufficiently increased value of torque and speed, the
required injection time for the DI fuel becomes such that piston
wetting would occur unless some PFI displaces DI fuel, preventing
wall wetting. The fraction of fuel provided by direct injection
would then be reduced to reduce the amount of directly injected
fuel and prevent the wall wetting. At even higher loads, knock
constraint results in increased need of DI fuel, with the
consequence that limited wall wetting will occur and particulate
emissions will occur.
[0075] FIG. 8 shows the ratio of DI fuel to PFI fuel as a function
of torque for this type of fueling scenario for stratified direct
injection. At lower values of torque, the DI/PFI ratio is
determined by the constraint of reducing PM emissions. Above a
certain value of torque, the need to prevent knock is the dominant
constraint and the resulting higher amount of directly injected
fuel increases particulate emissions.
[0076] Another use of DI, other than for knock suppression, is its
use to very accurately meter the amount of fuel injected into the
cylinder during rapid changes while PFI is employed for steady
state or slowly changing engine operation. DI could be used for
optimizing control during transients where the improved fueling
metering allows for more precise delivery of the fuel, avoiding
need for enrichment usually required from PFI in order to achieve a
substantial increase in power. Particulate formation with PFI is
usually determined by periods of rich operation; it could be
minimized by DI during a transient, with slow adjustment of PFI/DI
split.
Cold Start
[0077] Reducing particulate emissions that occur during a cold
start period of 100 seconds or so after the engine has been started
can be made easier than in warmed up operation by two factors.
First, there will generally be a lower average directly injected
fraction of fuel that will be required to prevent knock in this
cold start portion of the drive cycle, since cold start driving
will be less aggressive (less high torque operation) in comparison
to driving when the engine is warmed up. A representative drive
cycle for cold start driving could be comparable to the UDDS cycle
where, as discussed previously, the average direct injection
fraction around is around 1% and the fractional increase in
particulate emissions relative to PFI operation would only be less
than 1%.
[0078] Minimizing the fraction of DI used to prevent knock as
torque increases and decreases during the cold start period for
very high particulate emissions could thus be sufficient to reduce
particulate emissions relative to use of DI alone to less than 20%
and preferably less than 10%.
[0079] Second, if this is not sufficient, the short time duration
of cold start and the resulting small impact on overall drive cycle
engine efficiency of an adjustment (such as spark retard) can allow
greater use of adjustments to further reduce the fraction of fuel
that is directly injected at a given value of torque and speed so
as to provide greater reduction in direct injection use and
particulate production from what would have otherwise have been the
case.
[0080] The fuel management system can be operated so that the
average fraction of fuel in the cylinder that is provided by direct
injection during the cold start period at which very high
particulate emissions occur is either limited to be less than a
selected value by minimization of use of direct injection as torque
is increased without a change in spark retard (or another
adjustment); or, if necessary, increased spark retard is introduced
to achieve this limitation by decreasing the fraction of directly
injected fuel that is needed to prevent knock. During a cold start
portion of the drive cycle of 100 seconds, the average fraction of
fuel in the cylinder that is introduced by direct injection is
controlled to be less than it is in warmed up operation.
[0081] The amount of spark retard can be controlled by look up
table or by closed loop control. The spark retard could be a
constant level during all or part of the drive cycle or can be
varied as torque changes.
[0082] The cold start periods, when the engine operation needs
adjustment for cold cylinders, cold exhaust treatment catalyst,
cold manifold and for very high levels of particulates, can occur
for different time durations. The time duration for an engine
adjustment, such the use of increased spark retard, that is
optimized for reducing cold start direct injection particulate
emissions can be determined by a set time, a look up table or by
engine sensors that monitor parameters that include, but are not
limited to, engine coolant temperature. The cold start period
during which very high particulate emissions occur is typically
around the first 100 seconds or so of engine operation. This cold
start period can be longer than the cold start period during which
the catalyst for exhaust treatment needs to be heated.
[0083] Using spark retard, it may be possible to use port fuel
injection alone or almost entirely during the 100 second or so cold
start period for very high particulate production. The fuel
management system will limit the amount of spark retard so that a
high level of spark retard will not cause misfire. A misfire
detector can be used as input for this control. Open loop control
using a lookup table can also be employed as well as closed loop
control using a knock detector. The fuel management system can
control the spark retard so that during at least some time during
cold start, the spark retard is increased so as to provide the
largest decrease in the fraction of fuel that is directly injected
without creating misfire.
[0084] Other adjustment or adjustments, such as variable valve
timing and/or open-valve port fuel injection, could also be
employed. The vaporization cooling knock resistance that is
provided by open-valve port fuel injection can be used as an
alternative to direct injection. The same fuel injector could be
used for both closed and open valve port fuel injection by changing
timing.
[0085] These adjustments would be made during at least a portion of
the cold start period, when the fuel management system matches the
fraction of fuel provided by direct injection to the increase
required to prevent knock as torque and speed change so as to
require more knock resistance and to match the decrease allowed as
torque and speed change so as to require less knock resistance.
[0086] The fueling scenario for the cold start period of 100
seconds or where higher particulate emissions occur, can use
introduction of fuel into at least one cylinder by port fuel
injection and where the fuel is introduced by the direct injection,
if needed, so as to prevent knock as the torque increases. As
torque increases, the fraction of fuel that is directly injected
can be increased so as to match the amount needed to prevent knock
(and thereby minimize the fraction of fuel that is directly
injected). Spark retard can also be increased during at least part
of this cold start period so as to prevent knock that would
otherwise occur. At the highest value of torque in the cold start
portion of the drive cycle, the engine could be operated with only
port injection or with a combination of port and direct injection
where the fraction of fuel provided by direct injection is reduced
by the use of increased spark retard and/or another adjustment,
such as variable valve timing. Variable valve timing can also be
used together with increased spark retard during at least part of
this cold start period. Spark retard and other adjustments may be
controlled based upon the torque and speed at which a large
threshold increase in particulate emissions occurs.
[0087] An additional adjustment that can be used during the 100
second or so cold start period for very high particulate emissions
is to temporarily increase the pressure of the direct injector
during at least part of this period so as to reduce particulate
emissions. The relatively short time period of the cold start
period could facilitate deployment of this adjustment.
[0088] A further adjustment is to limit the amount of pressure
boosting that is employed by a turbocharged or supercharged engine.
This reduces the required fraction of fuel that must be directly
injected.
[0089] Since there can still be substantial PM emissions from PFI
during the cold start portion of the drive cycle, it may also be
necessary to employ means to reduce these emissions. One approach
is to use air heating from engine compression as described below.
Air heating can also be employed to reduce particulate emissions
from direct injection.
Fuel Management System Optimization
[0090] A control system 100, shown in FIG. 9, can be employed to
control the adjustments so that the particulate emission levels
meet regulations, while keeping the decrease in efficiency over a
drive cycle below a selected value. The amounts of gasoline from
the fuel tank 110 that are injected by port fuel injectors 120 and
direct injectors 130 into the engine 140 are controlled by using
information that includes the amount of fuel that is directly
injected into the engine 140 and other inputs that include but are
not limited to direct injector pulse length and the detection of
misfire.
[0091] The control system 100 also controls various engine
operation adjustments that affect the amount of direct injection
that is required to prevent knock.
[0092] The control system 100 can also employ the adjustments so to
limit the efficiency decrease from the adjustments so that it is no
greater than a selected value and/or so that the performance
decrease is no more than a selected value. The control system 100
can employ a look up table that provides information on which
combination of adjustments (type of adjustment, what portion of the
drive cycle, and how much is employed) provides the lowest
efficiency reduction for a given particulate reduction. The engine
can be operated at times at above the threshold for particulate
production as well as below it.
[0093] The control system 100 can also be used to employ optimized
port+direct injection to obtain better engine efficiency by dilute
operation through greater use of EGR (internal or external) at low
loads by minimizing DI use. Minimizing DI use can also be used to
increase efficiency by extending the limit for operation with a
lean fuel/air mixture at low loads, thus providing an additional
way for providing dilute operation. The control system 100 can also
be used to provide better fueling control at high speeds and high
loads by minimizing the amount of direct injection at a given
torque and speed that is needed to prevent knock as the torque is
increased without compromising efficiency and performance.
[0094] The amount of direct injection is minimized by matching its
level to the amount needed to prevent knock as torque and speed are
changed. Variable valve timing can be used to increase knock
resistance, thereby reducing the fraction of fuel that is directly
injected, and can vary internal EGR level to increase
efficiency.
[0095] The approaches described herein can make it possible to use
a combination of port and direct injection to reduce particulate
emissions to a level that would meet stringent future regulations
without using a gasoline particulate filter. They could also make
it possible to meet this goal without requiring closed loop
feedback control from instantaneous measurement using a particulate
measurement sensor and or other use of a sensor for measuring
particulate emissions over time.
[0096] During the cold start period, the control system would
achieve sufficient reduction of particulate emissions by the
combination of matching the fraction of fuel that is directly
injected to the amount needed to prevent knock. This matching could
use closed loop control from a knock sensor and could also use open
loop control. The use of open loop control can be particularly
important during transients. In addition, spark retard would be
used to further reduce the amount of direct injection that is used.
The amount of increased spark retard could be controlled by a knock
detector and by a misfire detector.
[0097] The amount of spark retard would be limited by the
requirement not to misfire. Maximum spark retard could be used in a
preset fashion so as to minimize the amount of direct injection
that is employed. The length of this cold start operation could be
determined by a preset time or by measurement of engine
temperature. The use of spark retard could be determined by the
fraction of fuel that is directly injected. If this fraction
becomes too high based on a look up table that correlates
particulate emissions with the fraction of fuel that is directly
injected, spark retard is used. The look up table could be
determined by measurements of particulate emissions from a test
engine. It could be also be determined by the results of an engine
model.
[0098] During the warmed up engine portion of the drive cycle,
minimizing the fraction of fuel that is directly injected would be
achieved by matching of the fraction needed to prevent knock. The
matching would utilize closed loop control employing a knock
detector and could also use open loop control that employs a look
up table. Open loop control using a look up table could be
especially important during transients, which include rapid changes
in torque and during engine shutdown and restart. Spark retard and,
if needed, other adjustments such as variable valve control, could
be used to further reduce particulate emissions by reducing the
fraction of fuel that is directly injected. This would prevent the
amount of directly injected fuel from exceeding the threshold for
the amount of directly injected fuel to produce a large percentage
increase in particulate emission.
[0099] The amount of spark retard or another adjustment that is
used could be determined by the amount of directly injected fuel or
by parameters from which it is inferred and a look up table based
on the use of calibrated model for when the threshold occurs. The
amount of spark retard and perhaps other adjustments could be used
in an optimized way to minimize any decrease in efficiency and
performance for a given amount of particulate reduction.
[0100] Alternatively, during the warmed up portion of the drive
cycle, spark retard can be controlled with information about engine
torque and speed so as to increase the threshold torque at a given
speed at which a threshold increase in particulate emission would
otherwise occur.
[0101] An additional control feature is the timing of DI injection
with the constraints needed to prevent wall wetting. The DI
injection is set so that the start of injection (SOI) and the
End-of-Injection (EOI) are adjusted in order to prevent wall
wetting. In the case where the injection duration is less than that
the maximum injection duration that would result in wall wetting
(either because of early wall wetting during the intake stroke or
late wall wetting during the compression stroke), the injection
timing can be adjusted within the limiting times for wall wetting
avoidance. Earlier injection results in better mixing, while later
injection results in more stratified charge, which could be
beneficial for misfire or knock avoidance.
[0102] As an alternative to removing the need for a gasoline
particulate filter to meet requirement for reducing particulate
emissions, the port+direct injection fuel management system
described here could be used in combination with a gasoline
particulate filter to provide a greater reduction in particulate
emissions than could be obtained with the gasoline particulate
filter alone. Further, the combination could also reduce the cost
and mitigate reliability and efficiency reduction drawbacks of the
gasoline particulate filter system.
[0103] The combination of PFI and DI for particulate control can be
employed to reduce the requirements on the particulate reduction
required from the gasoline particulate filter; the degree of
control required by filter; the range of conditions under which it
is used; and its durability. They can also remove the need for
instantaneous monitoring of particulate emissions.
Engine Air Preheating Enabled by Variable Valve Timing
[0104] Further reduction in particulate emissions from optimized
port+direct injection can be obtained by reducing port fuel
injection particulate emissions, especially at cold start.
[0105] An important factor in cold start particulate matter
emissions is poor vaporization of the fuel. The air is cold, and so
are the cylinder walls (liner and piston), the inlet manifold and
the inlet valve. Vaporization could be improved and emissions
decreased if the air temperature could be increased. Providing an
air heater is not practical due to the transient nature, with the
heater being effective long after it is really needed.
[0106] Heating the air does not require much power, but if it is
heated through contact with a surface (i.e., a conventional
heater), the thermal mass of the heater dominates and results in a
long transient. Use of variable valve timing, for air pre-heating
by engine compression, is a means to address this issue. Engine
compression can provide a very effective means for air preheating.
This approach is enabled by progress that has been made in variable
valve timing.
[0107] It is possible to heat the cylinder charge during the cold
start to minimize emissions, including hydrocarbon gas emission as
well as particulate emissions. Cold gas temperature, as a result of
cold walls that are in contact with the gas charge, can result in
poor fuel evaporation and contribute to a large overfueling to
compensate for poor evaporation. The overfueling results in a
substantial fraction of the total hydrocarbon emissions over a
driving cycle, as well as substantial fraction of particulates.
[0108] The power required to heat the gas is not high; however,
because of finite heat capacity of heat exchangers, there is a
substantial delay in delivering energy to the gas by solid
elements. By the time that these elements are heated, the cold
start period may be over.
[0109] An alternative means of delivery the heating is through
compression heating of the gases. In the cylinder, during the gas
compression, there is also associated heating. If the valves have
sufficient authority for adjustment during the cold start period,
it is possible to heat the gases in the cylinder and recirculate
them back to the inlet manifold, where they can vaporize the fuel
as well as reduce the amount of throttling needed.
[0110] This approach has been modeled using an engine simulator.
FIG. 10 shows the valve lift as a function of the crank angle
during two conditions. The first one is normal values for the valve
timing. The second one is for a highly advanced exhaust valve
opening (and corresponding, early exhaust valve closing). For the
model, the compression ratio is assumed to be 9, and the engine
operating speed is 200 rpm. The inlet gas temperature, as well as
the reflux from the exhaust, is assumed to be 280 K.
[0111] In order to investigate the potential for gas heating, the
exhaust valve timing has been considerably advanced. The conditions
for gas temperature and cylinder pressure are shown in FIG. 11 for
both the normal exhaust valve timing as well as advanced timing.
During the exhaust part of the cycle, the exhaust valve closes
early, allowing the gas remaining in the cylinder to compress and
heat. The compressed warm gas then leaves the cylinder at high
velocity entering the inlet manifold during the early stages of the
opening of the inlet valve, and then reenters the cylinder during
the intake part of the cycle, after having vaporized some of the
fuel in the inlet manifold. The heating can be greater than 30 K.
In the case assumed in FIG. 11, the heating is about 50 K. Note
that in the first case with normal valve timing, the gas at the
bottom of the cycle is actually a little cooler than during the
inlet, due to heat exchange between the cylinder charge and the
cold cylinder walls.
[0112] Significant power is required to drive the compression. For
the case shown in FIG. 11 with normal exhaust valve timing, for a
cylinder with a 93 mm bore and 81 mm stroke, the power required for
the cycle is around 10 W (normal exhaust valve timing). For the
advanced exhaust valve timing used in the air preheating, it is
preferred that the cranking power is at least 100 watts and less
than 1000 watts. It is preferred that this cranking power would be
provided by a 12 volt battery that provides power for other
functions in a vehicle. However, it could be provided by an
additional 12 volt battery or by a higher voltage battery used in a
hybrid vehicle.
[0113] For the case in FIG. 11, heating of around 50 K is provided
by an advanced exhaust valve timing of 40 crank angle degrees
(CAD). The cranking power is 140 W, mostly going into compression
of the gas rather than heating. If instead the exhaust valve is
advanced 30 CAD, the power reduces to about 100 W, but the heating
is only about 20 K. It is preferred that the exhaust valve timing
is advanced by at least 30 crank angle degrees and less than 60
crank angle degrees.
[0114] The inlet manifold pressure is assumed to be 0.5 bar, while
the exhaust pressure is assumed to be 1 bar. It may be possible to
perform this non-combustion gas heating cycle for a single cycle
for each cylinder, followed by conventional cycle. Before the
non-combustion gas heating cycle, there is no fuel added to the
inlet manifold and the spark is either not used or ineffective as
there is no fuel. The second cycle is fired conventionally, with
fuel injected prior to the inlet valve opening. In this manner,
better air/fuel preparation can be achieved for port fuel injection
that will reduce particulate and hydrocarbon emissions during cold
start.
[0115] The technique could also be used for direct injection, with
direct injection occurring during the second cycle, with heated
gasses.
[0116] It may be possible to have several cycles of operation in
this mode, where a non-combustion gas heating cycle is followed by
a power cycle, before the engine fully warms up.
[0117] FIG. 12 shows the very large mass flux back from the
cylinder to the inlet manifold of hot gases. The flow velocity
across the valves is sonic, as choke flow is established for the
conditions shown in FIG. 10-13 (could be adjusted by adjusting the
valve lift profile and its timing to set the desired the flow
reversal process).
[0118] The compression of air in the cylinder results in heating
during the compression cycle or during the exhaust cycle, if the
exhaust valve is adjusted so that air is prevented from leaving the
cylinder. In this concept, variable valve timing with or without
variable valve lift is used during engine startup to draw air into
the cylinder and then sent back at higher temperature back into the
inlet manifold (reflux), where it can be used to assist in the
vaporization of fuel. In this non-combustion air heating cycle, no
fuel is introduced into the cylinders by the fueling system, and
the spark from the spark plug is not employed. The heated air
improves vaporization of fuel that is deposited on the inlet
valves, either prior to the engine combustion start-up or during
the fueling process (by port-fuel injection).
[0119] During the expansion cycle, there is cooling of the cylinder
charge. If the system were totally reversible (i.e., adiabatic),
the initial temperature of the air prior to the compression cycle
would be same as that at the end of the expansion cycle. Because of
losses to the walls, it is slightly lower. Indeed, FIG. 10 shows
that situation. With normal exhaust valve timing, note that the
temperature of the gas at bottom dead center after the expansion
stroke is substantially lower than the inlet temperature, actually
by as much as 30 K. In the case with advanced exhaust valve timing,
the temperature of the inlet manifold (320 K) is actually higher
than the outside temperature (280 K). The temperature during the
flow reversal could be as high as 500 K. It should be noted that
the calculations in FIGS. 10-13 are for "steady-state" calculation
to illustrate the potential for the technology, but even in the
case of the transients, the trends are similar.
[0120] This analysis shows that the temperature of the gas in the
region close to the inlet valve can be increased by over
150.degree. C., even in the case when the engine is already firing,
improving the evaporation of the fuel on the valves, as well as
large increase in the velocity of flow back gases (up to choke
conditions).
[0121] There are several embodiments for the engine air heating.
One preferred embodiment is where the exhaust valve timing is
substantially advanced. In this case, warmed-up air can be
reintroduced to the inlet manifold (flow reversal), as shown in
FIGS. 10-13. The amount of air that is re-fluxed can be controlled
by control of the valve timing with or without control of variable
valve lift and by exhaust valve timing. The amount of preheating is
controlled by the timing of the opening of the valve. The pressure
in-cylinder would be higher than the pressure in the inlet
manifold, and the warmed-up hot air would flow at high speed
through the valve opening (which could be controlled by variable
valve lift). The pressure differential across the inlet valve in
the case of FIGS. 10-13 is about 6 bar. It is not necessary to have
such a high value, as it results in increased power demand. Lower
pressure would correspond both to lower temperatures of the reverse
flow as well as lower power demand. The valve lift can also be
adjusted to control the velocity of the flow through the valve,
both during the time of reflux (when the cylinder charge is
returned to the inlet manifold) as well as during the intake, when
the air in the inlet manifold returns to the cylinder. High gas
speed would help vaporizing fuel film that is on the inlet valve.
There will be spray generated by the interaction, with droplets
hitting the walls of the inlet manifold, but that mass will be
re-admitted to the cylinder in subsequent cycles (or after the
inlet manifold has warmed up).
[0122] Since the engine is not sparking during the first cycle and
subsequent cycles that are used for warming, it would be possible
to operate the engine as a 2-cycle engine. This would require
substantial control of the valves, which may be impractical. On the
other hand, if the exhaust valve can be de-activated (remains
closed), then it would be possible to just adjust the inlet valve
timing so that the hot air can be flowed back to the inlet manifold
in both the compression cycle and what would have been the exhaust
cycle.
[0123] Although the above description does not involve sparking
during the compression gas heating cycle, it may also be possible
to use a spark during the compression gas heating cycle, in order
to achieve limited combustion from any fuel that is in the
cylinder, such as fuel left over from previous engine operation
(and thereby get additional heating from the partial combustion).
If the cylinder charge (with uncombusted fuel and with free oxygen)
is flushed back to the inlet manifold, then it could pick up
additional fuel and mix with some fresh oxygen. As the incompletely
combusted mixture is sent back to the inlet manifold, emissions are
minimized.
[0124] Although the description is for one compression gas heating
cycle, to be followed by sparking cycles, it is possible to repeat
the non-spark condition for two or more cycles, further improving
the likelihood of appropriate combustion during the first fully
sparking cycle, and also decrease the emissions during the overall
cold-start process. In the case of sparking cycle, adjusting the
exhaust valve can be used to reverse the flow of hot gases into the
inlet manifold, improving evaporation of the fuel, for a few cycles
during the cold start period. The temperatures would be higher
(because of combustion in the cylinder), and thus, lower amounts of
reverse flow would be required. At this point, the engine is
self-driving, and does not need externally supplied power. The
exhaust valve timing is adjusted as the engine combusts the
air/fuel mixture, as well as the cylinder and the inlet manifold
warming up.
[0125] Although the calculations described above assume that the
flow reversal is readmitted into the same cylinder where it left,
it may be possible to use flow reversal from some cylinders into
other cylinders of the engine, by appropriate adjustment of the
valve timing, inlet manifold pressure and other. It is assumed that
in the first cycle, the cylinders are at atmospheric pressure. With
reduced pressure in the inlet manifold, it is possible to have a
few of the cylinders provide substantial fraction of the air that
goes into the cylinders.
[0126] The calculations above are for cycles without combustion.
Limited number of combustion cycles could also be used with
advanced exhaust valve opening and closing. In addition to
providing hot gasses to improve evaporation of the fuel by flow
reversal in the inlet manifold, early exhaust valve opening allows
the rejection of high temperature gasses in the exhaust, assisting
in catalyst warming.
[0127] A different embodiment is where the exhaust valve is totally
de-activated (defined as remaining closed).
[0128] For those conditions where the exhaust valve is not
deactivated, it is possible to have the inlet valve open during the
beginning of the exhaust cycle (with exhaust valve closed), that
is, with very retarded exhaust valve opening. Alternatively, it may
be possible to achieve compression of the residual air in the
exhaust through early closing of the exhaust valve. In this case,
the residual air in the cylinder would be compressed, followed by
opening of the inlet valve.
[0129] The valve lift can be different for the inlet and exhaust
valves with the exhaust valve lift being smaller than the inlet
valve lift. It may be possible to adjust the exhaust valve lift
such that there is substantial compression of the gas, with a
substantial fraction of the warmed up gas going to the inlet
manifold, through very small lift of the exhaust valve during the
exhaust period. The valve lift may also be adjusted, being one set
of values during the startup phase and a different set of values
during the warm-up phase. Either the inlet valves or the exhaust
valves, or both sets of valves, could have variable valve lift.
[0130] The engine compression preheat of air can also useful for
combined port fuel injection-direct injection systems where port
fuel injection is substituted for direct injection in order to
reduce particulate emissions. For use of combined port fuel
injection and direct injection during cold start, even if the
amount of direct injection can be greatly reduced, the particulate
emissions from port fuel injection are still considerable. The use
of the engine compression heated air for better air/fuel
preparation and reduced particulate emissions during port fuel
injection can thus have a significant impact on DI engines where
port fuel injection is employed to reduce particulate
emissions.
[0131] In a preferred embodiment, there is no valve overlap during
the cold start process. Thus, even when both the inlet valve
opening and the exhaust valve closing are advanced, the inlet valve
opening should be advanced further than the exhaust valve closing
in order to allow hot air to leave through the inlet valve.
[0132] The first cycle in each cylinder, in which the air in the
cylinder has not combusted (i.e. the non-combustion gas heating
cycle), could be used for preheating the air. Multiple cycles
without fuel could also be used to condition the air before
introducing fuel into the cylinder. The number of non-combustion
cycles with different valve timing than in warmed up engine
operation can be controlled by open loop control using a look up
table or by closed loop control where information from sensors that
measure parameters that include engine temperature and various
types of emissions.
[0133] Once fuel is introduced, combustion gases could be
introduced into the inlet manifold, while maintaining advanced
exhaust valve opening. The hot gasses could be used for
facilitating fuel evaporation. However, a substantial level of
residuals would now be present in the cylinder. Nevertheless,
better fuel/air preparation and increased charge temperature would
mitigate the effect of the increased residuals. Alternatively,
after a fuel injection, the engine could be operated through an
entire cycle without fuel, in which case the gas flowing back
through the inlet valve is mostly air (as fuel was either absent or
minimally introduced into the cylinder during the previous
cycle).
[0134] A drawback of this approach is that the engine will not
start during the first cycle for all the cylinders, and it will
require some additional power as the engine is serving as an air
compressor/heater. However, the emissions would be substantially
reduced. Additional battery capability may need to be added for
some other applications. The approach is particularly well suited
for addressing emissions from hybrid vehicles with substantial
electrical power availability during start up. The engine start up
does not have to occur at the same time as the time when the
vehicle starts moving, as electrical drive could be used at that
time.
[0135] The system could be useful for reducing both hydrocarbon gas
emissions, as well as particulate formation during the cold start.
Fuel enrichment, needed in order to attempt appropriate ignition,
can be substantially reduced.
[0136] Since the reflux air/fuel is substantially warmer than under
normal start-up, it may be possible to combine port fuel injection
with direct injection in order to achieve appropriate mixture
formation. The warmer air could help vaporize the fuel that is
directly injected. The direct injection can be used for better
control of the air/fuel stoichiometry of the cylinder. This
approach, together with the substitution of port fuel injection for
direct injection, which can be increased by spark retard, can have
a large impact in reduced particulate emissions during cold start
from direct injection engines including engines that are
turbocharged or supercharged.
[0137] Multiple sparking can be used in order to gain increased
ignition during the cold start using variable valve timing.
[0138] An issue with applying valve timing and lift during cold
start is that most automakers utilize hydraulic fluids for these
adjustments and the valve timing/lift system is inoperative for a
short time during the engine start period (e.g. a few seconds and
for at least one second) as the oil pump needs time to build up the
oil pressure.
[0139] There are several options to address this issue and provide
early closing of the exhaust valve. The first one is to use either
all-electric valve timing, as with the engine valve VEL and others,
or just electric-valve timing assist. Although there are advantages
of variable lift, variable timing may be more important and could
be sufficient. Variable lift could, in conjunction with variable
valve timing, however, be used to control the flow out of the
cylinder into the manifolds, minimizing the power required for
compressing the gas (that is, to maintain appropriate pressure
differential between the cylinder and the inlet manifold). For
example, it is of interest to minimize the flow into the exhaust
manifold, and instead redirect a substantial fraction of it into
the inlet manifold. If the pressure differential is high, choke
flow conditions could be established for the reverse flow into the
inlet manifold. However, the reverse flow could be choked flow or
non-choked flow.
[0140] Alternatively, the valve timing could be adjusted so that
without oil pressure, the exhaust valve timing is advanced. Once
the oil pressure builds, and the engine starts to warm up, the
exhaust valve timing is adjusted to the "normal" position. A
disadvantage of this approach is that the hydraulics are operating
at higher pressures and power requirements than in the case when
the valve timing is adjusted so that little effort is needed during
warm-up conditions.
[0141] Another option is to design a cam that is bi-stable, with
two stable operating points that do not require much hydraulic
action, but that require hydraulic action for shifting from one
stable mode to the other, or to adjust the valve timing during
conventional operation.
[0142] A further option is to use an electrically driven oil pump.
Electrification of the vehicle has resulted in some manufacturers
making an electrically driven oil pump. Because the engine does not
have to be at operating speed (idle or higher), it is possible to
build the oil pressure in a much faster time scale, allowing for
hydraulic control of the valve timing during the initial phase of
engine startup. Conventional valve timing can be kept, with the
valves adjusted hydraulically. The electrically driven oil pump
could be used during the startup period time and not at other times
in the drive cycle.
[0143] The use of variable valve timing/lift can be used for
controlling particulate emissions both during the steady state as
well as during cold start. In particular, unlike diesel engines
where particulate emissions increase with increasing EGR (Exhaust
Gas Recirculation), in the case of gasoline direct injection
engines, the particulate emissions (both particle number and total
mass) is decreased by using EGR, either external EGR (cooled) or
hot EGR (internal). Internal EGR can decrease the particulate from
gasoline direct injection engines substantially more than external.
Thus, the use of variable valve timing, by adjusting valve overlap,
can significantly decrease the particulates (both mass and number).
The mechanism for the decrease of particulate emissions with
internal EGR could be both due to increased rate of fuel
evaporation (larger charge in the cylinder as well as higher
temperature), and decrease the penetration of the spray. Increased
tolerance to EGR by port fuel injection (which can provide higher
temperatures in the cylinder by avoidance of evaporation cooling of
the fuel) further decreases the particulate emissions from DI
engines. That is, increased port-fuel injection, with increased EGR
(and preferentially internal EGR), decreases the particulate
emissions from dual-injection engines.
[0144] In addition to engines that are fueled with gasoline, this
approach can be used for engines that are fueled by ethanol or
methanol where starting and cold start emissions can be of greater
concern than for gasoline engines or even natural gas. One
application is for engines using high concentrations of ethanol or
methanol, including reduction of formaldehyde gas emissions from
methanol. In addition, cold starting in natural gas engines could
also benefit from this approach.
[0145] The technique could be used in stationary engines as well as
on- and off-road vehicles.
[0146] The technique may also be used for diesel engine startup. In
this case, there is no throttling and there is no fuel on the inlet
manifold, but the heated gas should help ignitability of the diesel
fuel in subsequent cycles.
[0147] The engine compression air preheat can also help in
reluctant starting conditions. Although the engine is not started
on the first cycle or few cycles because fuel is not introduced,
the engine will have a very high probability of starting when fuel
is introduced.
[0148] The use or non-use of engine compression air preheat can be
determined by closed loop control using sensor input or open loop
control. The control system can use sensed or inferred information
that includes engine temperature and particulate emissions. When
engine compression preheat is no longer required, the control
system changes the valve timing and lift to values appropriate for
regular driving operation.
[0149] The use of compression for air heating for emissions
reduction could be particularly attractive for downsized engines
where the amount of power to turn over the engine is reduced.
[0150] In addition to car and truck engines, this approach could be
used in other spark ignition engines including but not limited to
lawnmower engines, boat engines, snowmobile engines, motor cycle
engines, aircraft engines and engines for electric power
generation.
[0151] It may be possible to use this approach in modification of
existing engines in vehicles and the other products mentioned above
as well as factory produced engines. If the engine management
system has sufficient authority to substantially adjust the exhaust
valve, this process could be used in present day vehicles as means
to reduce the cold emissions, both hydrocarbons and
particulates.
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