Proceedings of the ASME 2016 Internal Combustion Engine Division Fall Technical Conference
ICEF2016
October 9-12, 2016, Greenville, South Carolina, USA
1 Copyright © 2016 by ASME
ICEF2016-9349
EFFICACY OF ADD-ON HYDROUS ETHANOL DUAL FUEL SYSTEMS TO REDUCE NOX
EMISSIONS FROM DIESEL ENGINES
Jeffrey T. Hwang, Alex J. Nord, William F. Northrop*
University of Minnesota
Department of Mechanical Engineering
Minneapolis, MN 55455
ABSTRACT
Aftermarket dual-fuel injection systems using a variety of
different fumigants have been proposed as alternatives to
expensive after-treatment to control NOX emissions from
legacy diesel engines. However, our previous work has shown
that available add-on systems using hydrous ethanol as the
fumigant achieve only minor benefits in emissions without re-
calibration of the diesel fuel injection strategy. This study
experimentally re-evaluates a novel aftermarket dual-fuel port
fuel injection (PFI) system used in our previous work, with the
addition of higher flow injectors to increase the fumigant
energy fraction (FEF), defined as the ratio of energy provided
by the hydrous ethanol on a lower heating value (LHV) basis
to overall fuel energy. Results of this study confirm our earlier
findings that as FEF increases, NO emissions decrease, while
NO2 and unburned ethanol emissions increase, leading to no
change in overall NOX. Peak cylinder pressure and apparent
rates of heat release are not strongly dependent on FEF,
indicating that in-cylinder NO formation rates by the
Zel’dovich mechanism remains the same. Through single zone
modeling, we show the feasibility of in-cylinder NO
conversion to NO2 aided by unburned ethanol. The modeling
results indicate that NO to NO2 conversion occurs during the
early expansion stroke where bulk gases have temperature in
the range of 1150 -1250 K. This work conclusively proves that
aftermarket dual fuel systems for fixed calibration diesel
engines cannot reduce NOX emissions without lowering peak
temperature during diffusive combustion responsible for
forming NO in the first place.
INTRODUCTION
Diesel engines are known for reliability, durability, low
manufacturing cost and high power density. Given their
longevity, legacy diesels regulated to older emissions levels
will continue to be used in practice for decades to come. New
diesel engines have ~27% lower NOX emissions than engines
of a decade ago [1] in part due to selective catalytic reduction
(SCR) aftertreatment systems. Although aftertreatment is an
effective method for reducing emissions, in-cylinder
techniques are also attractive to reduce SCR urea dosing rate
requirements or to possibly eliminate the need for NOX
aftertreatment altogether.
In-cylinder NO is primarily formed during combustion
through a combination of chemical pathways including the
extended Zel’dovich, prompt (Fenimore) and N2O
mechanisms [2–4]. NO in diesel engines mainly arises through
the thermally controlled Zel’dovich mechanism in lean to
stoichiometric regions found near the periphery of the
diffusive flame front. NO is oxidized to NO2 and
concentrations “freeze” short of thermodynamic equilibrium
soon after the end of injection and mixing of burned gases [5]
in the expansion stroke.
Low temperature combustion modes like dual fuel
reactivity controlled compression ignition (RCCI) have been
shown to simultaneously limit in-cylinder NOX and soot
production over a wide speed and load range [6–8]. RCCI uses
fumigation of a low reactivity fuel, like gasoline into the
intake manifold and early direct injection of a high reactivity
fuel like diesel to avoid high temperature NOX formation
regions found in conventional diesel combustion. Various
fumigants have been investigated for RCCI including
hydrogen, gasoline, hydrous ethanol, and natural gas [9–14]
and all have similar impacts on avoiding in-cylinder NOX
formation.
Although RCCI is an attractive method for in-cylinder
emissions reduction, it must be implemented in new engines
due to the requirement for significant modifications to engine
hardware and software. To date, manufacturers have chosen to
employ NOX aftertreatment like SCR to meet stringent
emissions standards for new engines and rely less on advanced
in-cylinder techniques like dual fuel RCCI.
For legacy diesel engines regulated to older emissions
standards, add-on SCR and lean NOX trap aftertreatment
systems have been marketed to meet new in-use NOX
regulations [15]. Dual fuel retrofit kits are available that also
claim to reduce NOX emissions without aftertreatment while
also substituting diesel fuel for lower carbon fuels like
compressed natural gas or partially renewable ethanol [16,17].
These aftermarket systems incorporate a separate fuel system
and fumigate the secondary fuel directly into the intake
2 Copyright © 2016 by ASME
plumbing, retaining the stock engine calibration and hardware.
Therefore, these systems cannot achieve the emissions
reductions possible with RCCI since they do not change the
diesel fuel injection strategy. Fumigation for diesel engines
has a long history [18]. Previous work has examined
advancing diesel injection timing of a mechanically injected
engine with fumigation of hydrous ethanol [19] to increase the
diesel replacement quantity but this strategy is not applicable
to modern electronically controlled diesel engines.
Our previous work investigated using hydrous ethanol as
the fumigant in an add-on configuration both with an existing
commercial fumigation system [16] and with a novel port-
injection system [20]. This work along with other published
literature on aftermarket dual fuel systems [9,13,21,22] show
that NOX is not appreciably reduced with increased fumigant
energy fraction (FEF), defined as the ratio of fumigant lower
heating value to overall input fuel lower heating value. Others
have shown that when the water content of hydrous ethanol
exceeds 50%, NOX can be mitigated through intake charge
cooling, lowering peak combustion temperatures in the
diffusive flame [2,23].
Although overall NOX emissions do not significantly
decrease for add-on dual fuel systems, our work and research
by others have shown that the NO2 to NOX fraction increases
with increasing FEF [2,14,16,20]. This suggests that NO
formation during combustion remains unchanged for dual fuel
operation but that NO oxidation to NO2 occurs with more
fumigation.
Increased unburned hydrocarbon emissions also increase
with FEF, which have been implicated in the NO to NO2
conversion process. HO2 radicals formed within the cylinder
from the oxidation of intermediate species such as CH3CHOH
have been shown to be responsible for the conversion of NO
to NO2 [3,4]. At high temperatures, HO2 is unstable and
unable to react, however as temperatures decrease, the HO2
radical becomes stable and begins to promote the reaction of
NO to NO2. Hori et al. illustrated a kinetic mechanism by
which hydrocarbons facilitate the conversion of NO to NO2 at
temperatures between 600 and 1200 K [24]. As in-cylinder
temperatures are generically higher than engine exhaust
temperatures, it can be concluded that the NO to NO2
conversion via unburned hydrocarbons occurs within the
cylinder during the expansion stroke. Though hydrocarbon
assisted NO conversion chemistry has been studied with light
hydrocarbons like methane and ethane, ethanol has not been
investigated.
The work presented here provides a thorough set of
performance and emissions data for an add-on hydrous ethanol
port injection dual fuel diesel engine covering a larger range
of FEF and hydrous ethanol water content than our previous
work. It also investigates the effect of unburned hydrocarbons
on in-cylinder NO to NO2 conversion through comparison of
experimental data to a single zone kinetic model.
EXPERIMENTAL
The objective of the experimental work was to investigate
a hydrous ethanol dual fuel PFI system over a large range of
engine operation using varying hydrous ethanol water content
and FEF. A John Deere 4045HF475 Tier 2 diesel engine was
used in the experiments. The specifications of the engine and
PFI system are shown in Table 1.
The same custom PFI fuel rail from our previous work
was used in this study [20]. The PFI rail was integrated into
the existing intake manifold, and incorporated two
automotive-grade fuel injectors aligned to spray directly in
between the intake ports of the cylinders.
Table 1: Engine Specifications
Manufacturer/Model John Deere 4045HF475
Engine Type 4-Stroke DI Diesel
Cylinders 4, in-line
Displacement (L) 4.5
Bore x Stroke (mm) 106 x 127
Compression Ratio 17.0:1
Maximum Power
(kW/rpm) 129/2400
Aspiration Turbocharged & After
Cooled
Diesel Injection System Common Rail
Ethanol Injection System Port Fuel Injection
Ethanol Heating System None
Emissions Certification EPA Tier 2 (Off-Highway)
IVO (CAD ATDCF) 339
EVC (CAD ATDCF) 380
Hydrous ethanol injections were controlled using the
signal from the manufacturer-installed camshaft sensor.
Analysis of the signal provided engine speed and the location
of cylinder one top dead center (TDC) to a National
Instruments (NI) cRIO controller. Hydrous ethanol injection
pulse width and timing were then output by the cRIO to each
injector. Each injector injected twice per two engine rotations
starting at 360 CAD ATDCF, 21 CAD after IVO, to partially
mitigate fuel bypass from positive valve overlap inherent to
this engine. The hydrous ethanol was stored in a secondary
container and pumped to the PFI rail at constant flow. A
digital scale was used to determine the time rate change of
mass during a given testing duration for hydrous ethanol while
diesel fuel flow was measured using a CUB5 series
mechanical fuel flow meter.
A laminar flow element (LFE) was used to measure
intake airflow rate, and after-cooler outlet temperature was
maintained between 40 and 50 °C using an air-water heat
exchanger. Heated intake air was required to ensure complete
combustion of the charge due to ethanol’s high latent heat of
vaporization.
Gaseous emissions were measured using an AVL Fourier
Transform Infrared Spectrometer (FTIR), while soot emissions
were measured using an AVL Micro-Soot Sensor (MSS).
Engine exhaust was first diluted at a ratio of 5-7 in a residence
chamber with compressed air before being measured by the
MSS. The FTIR sampled both raw exhaust and diluted
exhaust, where the ratio of CO2 emissions before and after
3 Copyright © 2016 by ASME
dilution was used as a measurement of dilution ratio at every
testing point. The experimental setup is given in Figure 1.
In addition to engine performance and emissions data,
high-speed in-cylinder pressure data was collected at each
testing condition. Kistler Type 6065A pressure transducers
were mounted in custom Kistler 6542Q128 glow plug adapters
for all four engine cylinders. A 0.1 CAD resolution BEI H25
incremental optical encoder was mounted to the engine crank
to trigger data acquisition using a NI BNC-2110 and PXI-6123
DAQ. A NI LabVIEW interface was programmed to sample
100 cycles for a total of 720,000 data points per cylinder.
Apparent heat release rate was calculated using a custom post-
processing code and a first law analysis as outlined in ref.
[25].
Figure 1: Diagram of engine test setup
Hydrous ethanol with 10% (180 Proof) and 20% (160
Proof) water by volume were used in this study. Both proofs
of hydrous ethanol have a lower distillation energy to LHV
ratio during the refining stages and do not require dehydration,
thereby significantly increasing the renewability of the fuel
[26]. Hydrous ethanol blends were mixed by volume with lab
grade non-denatured anhydrous ethanol and distilled water.
The primary direct injected fuel used was non-oxygenated #2
ultra-low sulfur diesel (ULSD). The experimental testing plan
consisted of operating the engine over a modified type C1 off-
road vehicle ISO 8178 eight point testing cycle with and
without hydrous ethanol PFI [27]. The testing modes are
shown in Table 2.
Table 2: Modified ISO 8178 engine operation conditions
Mode Engine Speed
[rpm]
Engine Load
[N-m]
BMEP
[bar]
1 2400 450 12.6
2 2400 350 9.77
3 2400 250 6.98
4 2400 50 1.40
5 1400 450 12.6
6 1400 350 9.77
7 1400 250 6.98
8 1000 0 (idle) 0.00
At each testing mode, the engine was first allowed to
reach steady state diesel fuel only combustion. The PFI system
was then toggled “on” for PFI of hydrous ethanol and data was
taken once emissions, temperature, and pressure data reached
steady state. The stock engine ECU was not modified in any
way; all diesel injection parameters followed the OEM
calibration. Data was collected at intervals of two minutes at
steady state operation and then averaged for reported results.
Ethanol injector pulse width was varied to increase FEF,
while engine load was held constant by varying the engine
pedal position, effectively decreasing diesel fuel flow to
accommodate increased load during PFI operation. After data
collection for the selected testing mode was completed, the
PFI system was toggled “off” and cycled to the next testing
mode under conventional diesel combustion (CDC).
Experiments were conducted over a four-day period, where
the eight-point test cycle was conducted in its entirety on each
day for 160 proof, 180 proof, selected repeats (160 and 180
proof), and CDC respectively. FEF was determined using the
time rate change of mass of hydrous ethanol and measured
diesel fuel flow rate at each testing condition in conjunction
with respective LHV values. The ratio of ethanol energy input
over total energy input was then calculated as the FEF, where
maximum FEF corresponds to the maximum pulse width
achieved at each testing condition.
An uncertainty analysis was conducted using standard
deviations of measurements during steady state and between
repeated data sets. Error bars on result figures are based on the
root mean square value of two times the standard deviation,
representing the 95% confidence interval. Propagation of error
calculations were estimated using the numerical sequential
perturbation approach [28]. Systemic error was small
compared to standard deviation error and was only used for
hydrous ethanol time rate change of mass.
RESULTS AND DISCUSSION
Performance and Emissions
The diesel engine equipped with the PFI hydrous ethanol
dual fuel system was operated over a range of injector pulse
widths for each operating mode with 160 proof, 180 proof,
and diesel-only modes. Performance results for the max FEF
achieved at each condition are given in Table 3. All values
were calculated on a diesel equivalent basis. Combustion
efficiency (CE) and air/fuel ratio (AFR) decreased very
slightly with increasing FEF for all conditions. The decrease
in CE can be attributed to charge cooling effects from the
latent heat of hydrous ethanol vaporization and the increased
amount of water being introduced to the engine. BTE
decreased with increasing FEF for most cases, but increased
slightly for a few 180 proof cases. This behavior is coupled
with the brake specific fuel consumption (BSFC), which
increased with increasing FEF for most cases, but decreased
for the same 180 proof cases with increased BTE.
All testing modes were stability limited, defined as when
an increase in injector pulse width resulted in unstable
combustion or audible engine knock. 160 proof hydrous
ethanol testing modes were able to reach higher injector pulse
widths as compared to 180 proof, before the onset of knock.
4 Copyright © 2016 by ASME
The higher knock tolerance of 160 proof ethanol can be
explained by the increased charge cooling from higher water
content. A significant barrier for diesel fuel replacement at
low engine speeds exists as illustrated in Table 3 especially at
high load (Modes 5 and 6) where the stability limit was
reached between an FEF of 23-25%.
Figure 2 gives the in-cylinder pressure and calculated
apparent rate of heat release (RoHR) for Mode 3 (2400 rpm, 7
bar BMEP) for selected FEF and 180 proof hydrous ethanol.
As FEF increased, the premixed heat release event and peak
pressure increased in magnitude. Combustion phasing, as
measured by CA50 advanced slightly with increasing FEF,
while burn duration, defined as CA90 – CA05 decreased with
increasing FEF. For example, for the cases shown in Figure 2,
CA50 advanced from 16.4 DATDC to 13.5 DATDC, while
burn duration decreased from 27.5 CAD to 20.0 CAD. These
trends are primarily due to the increased premixed portion of
combustion with increased ethanol. The bimodal shape of the
RoHR reflects amplified premixed and diffusion portions of
combustion. This has been previously explained by the
combustion of directly injected diesel fuel providing ignition
energy for the combustion of the premixed fumigant [29].
Unlike with RCCI that uses advanced diesel injection timing
to generate a primarily premixed combustion event, the
mixing controlled mode of heat release remained present with
increasing FEF.
Table 3: Engine performance parameters at maximum
FEF achieved
Mode Operation
Max
FEF
[%]
BSFC
[g/kW-hr]
BTE
[%]
CE
[%]
A/F
Ratio
1
160 Proof 41.7 252 34.0 99.6 28.5
180 Proof 39.1 210 40.8 99.8 32.6
Diesel 0 220 39.0 99.9 32.2
2
160 Proof 61.8 274 31.2 99.3 28.5
180 Proof 60.0 220 39.0 99.5 34.9
Diesel 0 226 38.0 99.9 35.2
3
160 Proof 51.6 288 29.8 99.3 32.1
180 Proof 49.6 229 37.4 99.5 39.9
Diesel 0 234 36.6 99.9 39.0
4
160 Proof 41.2 630 13.6 98.9 51.3
180 Proof 46.0 590 14.5 98.8 55.7
Diesel 0 435 19.7 99.9 71.0
5
160 Proof 23.8 214 40.1 99.9 22.0
180 Proof 21.6 201 42.6 99.9 23.1
Diesel 0 206 41.6 99.9 23.4
6
160 Proof 27.6 223 38.5 99.9 23.2
180 Proof 28.7 205 41.8 99.9 25.0
Diesel 0 209 41.0 99.9 25.6
7
160 Proof 33.4 230 37.2 99.8 27.4
180 Proof 26.0 215 39.9 99.9 29.5
Diesel 0 213 40.3 99.9 31.1
8
160 Proof 48.9 980 8.74 99.1 80.1
180 Proof 53.0 848 10.1 98.9 90.5
Diesel 0 604 14.2 99.9 131
Figure 2: In-cylinder pressure traces and apparent RoHR
for Mode 3 and 180 proof hydrous ethanol
Figure 3: Brake specific CO emissions as a function of FEF
for 160 and 180 proof hydrous ethanol
Brake specific CO emissions are shown in Figure 3 for
160 and 180 proof hydrous ethanol injection over a range of
FEF for each testing mode. The dotted lines depicted on
emission plots represents diesel only combustion. CO
emissions increased uniformly with increasing FEF, and were
largely independent of ethanol proof. Incomplete combustion
is the primary cause of increased CO emissions, while low in-
5 Copyright © 2016 by ASME
cylinder temperatures prevent the oxidation of CO to CO2.
The trend also shows CO emissions reaching a horizontal
asymptote at high FEF where the increased A/F ratio prevents
the formation of CO.
Increases in CO have been shown to directly correlate
with an increase in HC emissions for combustion with
alcohols [13]. Further, excess HC emissions in dual fuel
modes are known to arise from the fumigant and not the
directly injected diesel fuel [9]. Figure 4 shows the light HC
distribution for 160 and 180 proof hydrous ethanol at Mode 3
for a range of FEF as measured by FTIR. Methane (CH4) and
ethylene (C2H4) emissions increase with FEF indicating
incomplete combustion becomes more significant. Using 160
proof hydrous ethanol leads to higher light HC emissions
overall. Increased water injection at the same FEF level leads
to additional cylinder charge cooling that lowers in-cylinder
temperature and leads to more incomplete combustion. This is
reflected in the lower CE values for higher ethanol proof given
in Table 3.
Figure 4: Selected HC emissions on a brake-specific basis
as a function of FEF for 160 and 180 proof hydrous
ethanol at Mode 3
Although injected hydrous ethanol proof had a significant
impact on light HC emissions, it did not change brake specific
ethanol emissions as shown in Figure 5. Similar to our
previous work, high engine load cases exhibit the lowest
amount of unburned ethanol due to higher engine temperatures
allowing sufficient heating to combust the ethanol completely.
Unburned ethanol emissions increase with FEF mainly
because it arises from areas in the combustion chamber
uninfluenced by the diffusive combustion event. These areas
include the squish and crevice regions. With increasing FEF,
the concentration of ethanol is greater in these regions leading
to higher emissions.
Further, at low engine load conditions, unburned ethanol
emissions increase much more rapidly as a function of FEF
due to fewer sufficiently hot regions in the combustion
chamber. Unburned ethanol concentration is lower for the
1400 rpm modes because low engine speed conditions operate
at higher temperatures than high engine speed conditions for a
given load. Fang et al. have also shown that ethanol delays
ignition and combustion phasing, also resulting in increased
unburned ethanol in the exhaust at higher engine speed [9]. In
addition, the injected ethanol is premixed with intake air,
causing any overlap between exhaust valve close (EVC) and
intake valve open (IVO) events to increase unburned ethanol
emissions through short-circuiting [20]. Although the PFI
injection strategy used in this study injected 21 CAD after
IVO to mitigate this effect, some short-circuiting is still
expected. In addition, gases from crevice volumes are known
to increase unburned ethanol emissions.
Figure 5: Brake specific unburned ethanol emissions as a
function of FEF for 160 and 180 proof hydrous ethanol
Figure 6 shows soot concentration in mg/m3 of exhaust
for all tested conditions. There was no uniform trend in soot
emissions as a function of FEF. High load modes (Modes 1, 2,
5, 6) showed increasing soot emissions with FEF. Of those
that achieved high FEF (Modes 1 and 2), the soot eventually
decreased. Initial increases in soot could be due to higher
temperature diffusion combustion and richer mixture around
the flame zones due to premixed ethanol. At high FEF
however, significantly less diesel is injected and combustion is
shifted more to a premixed mode.
6 Copyright © 2016 by ASME
Figure 6: Soot concentration as a function of FEF for 160
and 180 proof hydrous ethanol
For lighter load modes at high speed (Modes 2 and 3),
premixed ethanol had a more immediate effect in decreasing
soot emissions. At these conditions, shorter residence time and
diesel injection duration allowed soot to decrease more rapidly
with FEF. At idle, soot emissions were negligible and
independent of FEF due to primarily premixed low
temperature combustion. Ethanol proof also had no impact on
soot emissions for any mode indicating that any benefits
gained were due to ethanol replacement of diesel and not
through water influence on soot formation.
Ethanol may also chemically play a role in decreasing soot
formation at higher FEF. Previous work has shown that dual-
fuel combustion reduces soot concentrations, especially when
the secondary fuel is ethanol where increased OH radicals lead
to greater post-combustion soot oxidation [2,30,31]. Ethanol
(C2H5OH) consists of C-H, C-C, C-O, and O-H bonds. During
combustion, the C-C and C-O bonds can be readily broken due
to their lower bond energies. This chemical reaction causes an
increased concentration of OH radicals within the combustion
chamber [14]. If the theory is accurate, higher OH radical
concentration at higher FEF may result in decreasing exhaust
soot concentration.
NOx Emissions Results
Figure 7 depicts the brake specific NO emissions for 160
and 180 proof hydrous ethanol. For every operating mode,
NO emissions decreased with increasing FEF, as might be
expected due to the increase in charge cooling lowering in-
cylinder temperatures. However, there was no discernable
change in NO emissions between the use of 160 and 180 proof
hydrous ethanol at a given FEF indicating that NO formation
is independent of water content in the fuel. This discrepancy is
evidence that other factors besides charge influence NO
emissions for dual fuel combustion, such as the propensity of
unburned hydrocarbons facilitating the conversion of NO to
NO2 during the expansion stroke. The chemical kinetics and
mechanisms responsible for this conversion are discussed in
the single zone combustion modeling presented later in this
work.
Figure 7: Brake specific NO emissions as a function of FEF
for 160 and 180 proof hydrous ethanol
As in previous work, increasing FEF had very little
impact on overall NOX emissions as shown in Figure 8. This
indicates that fumigation with hydrous ethanol for all cases,
except for the very highest FEF conditions, does not have an
impact on NO formation during diffusive diesel combustion
and that NO is converted to NO2 at some point during the
closed cycle. Only at very high FEF does charge cooling play
a role in mitigating formation via the thermal Zel’dovich
mechanism, especially noticeable during Mode 4.
The experimental data imply that NO is converted to NO2
while overall NOX concentration nearly constant. As a
measure of the conversion process, Figure 9 shows the
NO2/NOX ratio for all experimental data points collected in the
study as a function of unburned ethanol measured in the
exhaust normalized by the diesel-only NOX emissions per
mode respectively. The NO to NO2 conversion process occurs
rapidly as a function of unburned ethanol until reaching a
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horizontal asymptote around 0.72. At this point, no further
conversion of NO occurs with increasing FEF. Reasons for
this trend are not apparent from the experimental data and
require additional kinetics modeling to investigate.
Figure 8: Brake specific NOX emissions as a function of
FEF for 160 and 180 proof hydrous ethanol
Figure 9: NO2/NOX ratio as a function of unburned
ethanol for all modes and ethanol proofs
Single Zone Combustion Modeling
Our experimental results indicate that NO to NO2
conversion increases with increasing unburned ethanol in the
exhaust. In previous research, Hori et al. [24] used an
adiabatic constant pressure single zone reactor model and a
constant temperature quartz flow experimental reactor to
illustrate the mechanism by which hydrocarbons influence the
NO to NO2 conversion at different temperatures, with ethylene
and propane being very effective as compared to methane and
ethane. However, work conducted in Hori et al. did not
consider ethanol.
A single zone constant pressure reactor model was created
in the open-source thermochemistry and kinetics code Cantera
to compare ethanol’s effectiveness to other hydrocarbons for
converting NO to NO2. The C1-C4 Hydrocarbon with NO
Addition mechanism from Lawrence Livermore National
Laboratory (LLNL) was used in the model [24]. Initial
conditions for the model were the same as for the Hori et al
paper including atmospheric pressure and concentrations 20
PPM NO in N2, 50 PPM of HC in N2, and the remainder air.
The constant pressure model was run over a sweep of
operating temperature between 600 and 1200 K at 50 K
increments, where the gases were controlled to a 1.5 second
residence time within the reactor. Ethanol, ethylene and
methane were used as hydrocarbons because these had the
highest concentrations in the measured engine exhaust.
Concentrations of NO and NO2 were exported and the ratio of
final NO2/NOX ratio for each operating temperature can be
seen in Figure 10.
Figure 10: NO2/NOX ratio as a function of reaction
temperature for selected HC’s using a constant pressure
reactor
Figure 10 shows that ethanol has a peak NO2/NOX ratio
similar to ethylene but over a smaller temperature window.
Hydrocarbons predominantly oxidize NO to NO2 through the
NO + HO2 ↔ NO2 + OH mechanism. Hori et al. have shown
that the effectiveness of a hydrocarbon at converting NO to
NO2 is dependent on its ability to simultaneously produce
radicals like OH to sustain fuel oxidation and HO2 for NO to
8 Copyright © 2016 by ASME
NO2 conversion [24]. According to the model, methane does
not readily promote the conversion of NO to NO2. Because the
oxidation of methane is relatively slow, there is a limited
amount of HO2 produced, in addition to methyl radicals
reducing NO2 to NO via CH3 + NO2 ↔ CH3O + NO [24].
Ethanol reacts with OH radicals to make CH3CHOH, and
CH3CH2O, which can be oxidized to produce HO2 in the
following reaction scheme.
𝐶𝐻3𝐶𝐻𝑂𝐻 + 𝑂2 ↔ 𝐶𝐻3𝐻𝐶𝑂 + 𝐻𝑂2 (1)
𝐶𝐻3𝐻𝐶𝑂 + 𝑂 ↔ 𝐶𝐻3𝐶𝑂 + 𝑂𝐻 (2)
Reaction 1 forms HO2 radicals to promote NO to NO2
conversion, while reaction 2 produces OH radicals to feed the
ethanol consumption reaction. The model shows that ethanol
has a high tendency to convert NO to NO2 at temperatures
between 800 - 1200 K. This temperature range is greater than
engine out exhaust temperatures; however temperatures during
the expansion stroke would fall within this temperature range.
To more closely compare the model to experimental
results, a variable pressure single zone model was created.
This model uses the recorded in-cylinder pressure data versus
time to model NO conversion kinetics occurring during the
expansion stroke. The model assumes that NO formation is
complete and the burned gases are mixed by the crank angle
location of 90% gross heat release (CA90). Therefore, CA90
was chosen as the initial condition for the variable pressure
model. The pressure data starting at CA90 was used in
conjunction with a range of initial local in-cylinder
temperatures and a sweep of unburned ethanol concentrations
to predict NO to NO2 conversion. A range of temperatures at
CA90 was used because local in-cylinder temperatures were
unknown and can vary within the cylinder. The in-cylinder
temperatures were calculated using the polytropic relations for
each initial CA90 temperature and measured pressure data.
Figure 11 illustrates the conditions under which the model
was run. The symbols on the plot are for clarity between the
trends, and are not indicative of actual data. The TCA90 surface
represents the isentropic temperature curve fits for initial
temperatures from 1000 to 2000 K at 50 K increments. A
larger lower range of temperatures was chosen because the
introduction of water is well known to decrease in-cylinder
temperatures. The mean in-cylinder temperature curve starting
at CA90 was calculated to validate the TCA90 range used for
the model. The initial exhaust composition points were taken
from a CDC mode (1400 rpm, 250 N-m), and a range of
ethanol (500 - 4000 PPM) was added to the mixture at the start
of the model. Similar results were obtained when running the
model at different engine testing modes, and were not
included for brevity.
Figure 11: Apparent RoHR, in-cylinder pressure, mean in-
cylinder temperature and CA90 temperature range as a
function of CAD
The NO conversion trajectory as a function of crank angle
for two different CA90 temperatures is shown in Figure 12. At
lower temperatures, NO is readily converted to NO2 early in
the expansion stroke, and then “freezes” when the temperature
becomes too low to promote conversion. At higher
temperatures, NO has a sudden decrease followed by an
increase before settling as in-cylinder temperatures do not
enter the right temperature range to promote conversion. The
decrease in NO at high temperatures is caused by the
oxidation of ethanol, while the increase right after is because
the reversion of NO2 to NO is more favored at high
temperatures. The model suggests that unburned ethanol will
not be evident in regions of high in-cylinder temperatures,
preventing the conversion of NO to NO2. Instead, unburned
ethanol evident in low local in-cylinder temperature regions
will readily convert NO to NO2.
Figure 12: NO conversion trajectory as a function of CAD
starting at CA90
9 Copyright © 2016 by ASME
Figure 13: NO2/NOX contour as a function of local in-
cylinder temperatures at CA90 and normalized unburned
ethanol concentration
Figure 13 shows the NO2/NOX ratio at the end of
expansion, plotted as a contour against the initial temperature
at CA90 and the ratio of unburned ethanol concentration
normalized against initial NOX concentration. A clear island
can be seen for TCA90 between 1150 and 1250 K where
unburned ethanol in the exhaust promotes the conversion of
NO to NO2. At local in-cylinder temperatures greater than
1400 K, complete ethanol conversion is predicted, preventing
the conversion. In addition, higher local temperatures favor
the reduction of NO2 to NO, while the consumption of ethanol
yields daughter radicals. Hori et al. have shown that daughter
radicals resistant to oxidation by O2 will reduce NO2 to NO by
the reaction R + NO2 ↔ NO + RO. In contrast, at local in-
cylinder temperatures lower than 1150 K, limited OH radicals
reduces the production of HO2, slowing down the conversion
of NO to NO2. It is also important to note that this model only
takes into account unburned ethanol in the expansion stroke.
As seen in Figure 4, other unburned hydrocarbons such as
ethylene (C2H4), and methane (CH4) increase with increasing
FEF, promoting additional NO to NO2 conversion.
CONCLUSIONS
In this study, a comprehensive dataset was presented
characterizing an add-on dual fuel PFI system using hydrous
ethanol as the secondary fuel. Data was collected over a range
of FEF for each point along a modified ISO 8178 eight point
testing plan. The results show that 160 proof hydrous ethanol
can achieve up to 61.8% FEF, while 180 proof reaches up to
60% FEF. CO, THC, and unburned ethanol emissions all
increase with increasing FEF, while NOX emissions initially
show no change, but begin to decrease at high FEF where
significant charge cooling lowers the diesel combustion
temperature. It was also found that both 160 and 180 proof
hydrous ethanol follow the same emissions trends, with few
significant different between similar FEF values.
Single-zone reactor models were created using Cantera to
investigate the conversion of NO to NO2 due to unburned
hydrocarbons in the exhaust. The first model found that
different hydrocarbons have different propensities for
promoting the NO to NO2 conversion, a finding in validation
with work conducted by Hori et al. [23] Results from this
model indicated that unburned ethanol concentrations promote
the conversion of NO to NO2 at a temperature range from 800
– 1200 K, temperatures likely to be encountered during the
expansion stroke of an engine. The second single zone model
used high-speed in-cylinder pressure traces and CA90 as
initial conditions to estimate NO to NO2 conversion during the
expansion stroke. The model found that in-cylinder
temperature regions between 1150 and 1250 K at CA90 have a
high NO to NO2 conversion rate and that conversion mostly
occurs near the beginning of the expansion stroke.
Overall, our findings indicate that aftermarket dual fuel
systems that directly introduce the secondary fuel into a fixed
calibration engine cannot achieve the emissions reductions
possible with low temperature RCCI combustion modes. To
effectively use aftermarket dual fuel strategies for NOX
emissions reduction, mitigation of NO formation in the diesel
diffusion flame must be achieved by lowering combustion
temperature. Strategies such as exhaust gas recirculation or
hydrous ethanol reformation could be used to increase the heat
capacity of the unburned gas and reduce overall combustion
temperature while maintaining high FEF and thermal
efficiency.
ACKNOWLEDGMENTS
This research was conducted with funding from the
Minnesota Corn Growers Association, The Agricultural
Utilization Research Institute and the University of Minnesota
Institute for Renewable Energy and Environment under grant
AIC209. We wish to acknowledge our colleagues at the
Thomas E. Murphy Engine Research Laboratory at the
University of Minnesota, especially Darrick Zarling for
technical guidance, Andrew Kotz for assistance with
developing the high speed in-cylinder pressure trace data
logging system, and Wei Fang for high speed data processing
assistance.
NOMENCLATURE
AFR – Air/Fuel Ratio
ATDCF – After Top Dead Center Firing
BMEP – Brake Mean Effective Pressure
BSFC – Brake Specific Fuel Consumption
BTE – Brake Thermal Efficiency
CA90 – Crank Angle location of 90% gross heat release
CAD – Crank Angle Degree
CDC – Conventional Diesel Combustion
CE – Combustion Efficiency
CI – Compression Ignition
DATDC – Degrees After Top Dead Center
ECU – Electronic Control Unit
EGR – Exhaust Gas Recirculation
EVC – Exhaust Valve Close
FEF – Fumigant Energy Fraction
FTIR – Fourier Transform Infrared Spectrometer
GHG – Greenhouse Gas
HC – Hydrocarbon
IVO – Intake Valve Open
10 Copyright © 2016 by ASME
LFE – Laminar Flow Element
LHV – Lower Heating Value
LLNL – Lawrence Livermore National Laboratory
MSS – Micro-Soot Sensor
NI – National Instruments
PM – Particulate Matter
PFI – Port Fuel Injection
RCCI – Reactivity Controlled Compression Ignition
RoHR – Rate of Heat Release
SCR – Selective Catalytic Reduction
TDC – Top Dead Center
THC – Total Hydro Carbons
ULSD – Ultra-Low Sulfur Diesel
REFERENCES
[1] EPA, 2013, “Nonroad Compression-Ignition Engines -
- Exhaust Emissions Standards” [Online]. Available:
http://www.epa.gov/otaq/standards/nonroad/nonroadci
.htm.
[2] Olson, A. L., 2010, “The Effect of Ethanol-Water
Fumigation on the Performance and Emissions from a
Direct-Injection Diesel Engine,” (September).
[3] Bowman, C. T., 1992, “Control of combustion-
generated nitrogen oxide emissions: Technology
driven by regulation,” Symp. Combust., 24(1), pp.
859–878.
[4] Lilik, G. K., Zhang, H., Herreros, J. M., Haworth, D.
C., and Boehman, A. L., 2010, “Hydrogen assisted
diesel combustion,” Int. J. Hydrogen Energy, 35(9),
pp. 4382–4398.
[5] Lavoie, G. A., Heywood, J. B., and Keck, J. C., 1970,
“Experimental and Theoretical Study of Nitric Oxide
Formation in Internal Combustion Engines,” Combust.
Sci. Technol., 1(4), pp. 313–326.
[6] Dempsey, A. B., Walker, N. R., Splitter, D., Wissink,
M., and Reitz, R. D., 2012, “Characterization of
Reactivity Controlled Compression Ignition ( RCCI )
Using Premixed Hydrated Ethanol and Direct
Injection Diesel in Heavy-Duty and Light-Duty
Engines.”
[7] Splitter, D., Hanson, R., Kokjohn, S. L., and Reitz, R.
D., 2011, “Reactivity Controlled Compression Ignition
(RCCI) Heavy-Duty Engine Operation at Mid-and
High-Loads with Conventional and Alternative Fuels,”
SAE Tech. Pap. Ser., 1(Ci), p. 0363.
[8] Reitz, R. D., and Duraisamy, G., 2015, “Review of
high efficiency and clean reactivity controlled
compression ignition (RCCI) combustion in internal
combustion engines,” Prog. Energy Combust. Sci., 46,
pp. 12–71.
[9] Fang, W., Huang, B., Kittelson, D. B., and Northrop,
W. F., 2012, “Dual-Fuel Diesel Engine Combustion
with Hydrogen, Gasoline, and Ethanol Fumigants:
Effect of Diesel Injection Timing,” ASME Intern.
Combust. Engine Div., pp. 1–9.
[10] El-hagar, M. M. E., 2014, “Exhaust Emissions of a
Single Cylinder Diesel Engine with Addition of
Ethanol,” 3(1), pp. 74–81.
[11] Rahman, M. M., Stevanovic, S., Brown, R. J., and
Ristovski, Z., 2013, “Influence of Different
Alternative Fuels on Particle Emission from a
Turbocharged Common-Rail Diesel Engine,” Procedia
Eng., 56, pp. 381–386.
[12] Cheng, C. H., Cheung, C. S., Chan, T. L., Lee, S. C.,
and Yao, C. D., 2008, “Experimental Investigation on
the Performance, Gaseous and Particulate Emissions
of a Methanol Fumigated Diesel Engine.,” Sci. Total
Environ., 389(1), pp. 115–24.
[13] Imran, A., Varman, M., Masjuki, H. H., and Kalam,
M. a., 2013, “Review on Alcohol Fumigation on
Diesel Engine: A Viable Alternative Dual Fuel
Technology for Satisfactory Engine Performance and
Reduction of Environment Concerning Emission,”
Renew. Sustain. Energy Rev., 26, pp. 739–751.
[14] Surawski, N. C., Ristovski, Z. D., Brown, R. J., and
Situ, R., 2012, “Gaseous and Particle Emissions from
an Ethanol Fumigated Compression Ignition Engine,”
Energy Convers. Manag., 54(1), pp. 145–151.
[15] Code of Regulations, C., and ARB, Regulation to
Reduce Emissions of Diesel Particulate Matter,
Oxides of Nitrogen and Other Criteria Pollutants from
In-use Heavy-Duty Diesel-Fueled Vehicles.
[16] Hwang, J. T., and Northrop, W. F., 2014, “Gas and
Particle Emissions from a Diesel Engine Operating in
a Dual-Fuel Mode using High Water Content Hydrous
Ethanol,” ASME Intern. Combust. Engine Div., pp. 1–
9.
[17] Papagiannakis, R. G., and Hountalas, D. T., 2004,
“Combustion and exhaust emission characteristics of a
dual fuel compression ignition engine operated with
pilot diesel fuel and natural gas,” Energy Convers.
Manag., 45(18-19), pp. 2971–2987.
[18] Alperstein, M., Swim, W. B., and Schweitzer, P. H.,
1957, “Fumigation Kills Smoke Improves Diesel.”
[19] Chaplin, J., and Binjanius, R., 1987, “Ethanol
fumigation of a compression-ignition engine using
advanced injection of diesel fuel,” Trans. Asae, 30(3),
pp. 610–614.
[20] Nord, A. J., Hwang, J. T., and Northrop, W. F., 2015,
“Emissions from a Diesel Engine Operating in a Dual-
Fuel Mode Using Port-Fuel Injection of Heated
Hydrous Ethanol,” ASME Intern. Combust. Engine
Div., pp. 1–9.
[21] Soloiu, V., Muinos, M., and Harp, S., 2015,
“Investigation of Dual Fuel PCCI (PFI of n-Butanol
and DI-ULSD) Compared with DI of Binary Mixtures
of the Same Fuels in an Omnivorous Diesel Engine,”
SAE Int., (X).
[22] Morsy, M. H., 2015, “Assessment of a direct injection
diesel engine fumigated with ethanol/water mixtures,”
Energy Convers. Manag., 94, pp. 406–414.
[23] Greeves, G., Khan, I. M., and Onion, G., 1977,
“Effects of water introduction on diesel engine
combustion and emissions,” Symp. Combust., 16(1),
pp. 321–336.
[24] Hori, M., Matsunaga, N., Marinov, N., William, P.,
and Charles, W., 1998, “An experimental and kinetic
calculation of the promotion effect of hydrocarbons on
11 Copyright © 2016 by ASME
the NO-NO2 conversion in a flow reactor,” Symp.
Combust., 27(1), pp. 389–396.
[25] Heywood, J. B., Internal Combustion Engine
Fundamentals.
[26] Saffy, H. A., Northrop, W. F., Kittelson, D. B., and
Boies, A. M., 2015, “Energy, carbon dioxide and
water use implications of hydrous ethanol
production,” Energy Convers. Manag., 105, pp. 900–
907.
[27] 1997, “ISO 8178 Emission Test Cycle,” DieselNet
[Online]. Available:
https://www.dieselnet.com/standards/cycles/iso8178.p
hp.
[28] Figliola, R., and Beasley, D., 2000, Theory and
Design for Mechanical Measurements, John Wiley
and Sons, New York.
[29] Northrop, W. F., Fang, W., and Huang, B., 2013,
“Combustion Phasing Effect on Cycle Efficiency of a
Diesel Engine Using Advanced Gasoline Fumigation,”
J. Eng. Gas Turbines Power, 135(3), p. 032801.
[30] Bika, A. S., Franklin, L., Olson, A. L., Watts, W., and
Kittelson, D., 2009, “Ethanol Ultilization in a Diesel
Engine.”
[31] Can, Ö., Çelikten, İ., and Usta, N., 2004, “Effects of
Ethanol Addition on Performance and Emissions of a
Turbocharged Indirect Injection Diesel Engine
Running at Different Injection Pressures,” Energy
Convers. Manag., 45(15-16), pp. 2429–2440.
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