Effects of diesel-water emulsion combustion on diesel engine NOx emissions

Material Information

Effects of diesel-water emulsion combustion on diesel engine NOx emissions
Canfield, C. Alan ( Dissertant )
Zhang, Zhuomin ( Thesis advisor )
Proctor, Charles ( Reviewer )
Lear, William ( Reviewer )
Green, Alex ( Reviewer )
Place of Publication:
State University System of Florida
Publication Date:
Copyright Date:


Subjects / Keywords:
Combustion ( jstor )
Combustion temperature ( jstor )
Cylinders ( jstor )
Diesel engines ( jstor )
Emulsions ( jstor )
Engines ( jstor )
Enthalpy ( jstor )
Internal combustion engines ( jstor )
Pollutant emissions ( jstor )
Water temperature ( jstor )
Dissertations, Academic -- Mechanical Engineering -- UF ( lcsh )
Mechanical Engineering thesis, M.S ( lcsh )
bibliography ( marcgt )
theses ( marcgt )
government publication (state, provincial, terriorial, dependent) ( marcgt )
non-fiction ( marcgt )


This study examines the effects of combusting a mixture of diesel fuel, water, and surfactant on the nitrogen oxides, or NOx, emissions from a compression ignition diesel engine. Previous research has attributed the observed reduction of nitrogen oxide emissions to a suppression of flame temperature due to quenching effects from the water, thereby reducing thermal NOx formation. The thesis highlights the relevant theory, operation, and design parameters of diesel internal combustion engines. Experimental procedures conducted using a Detroit Diesel 4-cylinder diesel engine are discussed. Results from testing diesel fuel with varying ratios of water balanced with a surfactant to stabilize the emulsion will be presented and discussed. The data shows significant NOx emission reduction with up to 45 percent water, by volume, in the fuel. These results are correlated with a thermodynamic first law analysis to estimate the adiabatic flame temperature of the standard fuel and fuel--water emulsion cases. Results indicate that thermal NOx is indeed reduced by quenching and flame temperature suppression, confirming reports in the literature. Recommendations are given for further studies, including improving the fuel--water emulsion and considerations for long-term testing. ( , )
KEYWORDS: water injection, emulsion, NOx, nitrogen oxides, diesel engine air pollution control
Thesis (M.S.)--University of Florida, 1999.
Includes bibliographical references (p. 79-84).
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Title from first page of PDF file.
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"x" in "NOx" in title is subscript.
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Document formatted into pages; contains vii, 85 p.; also includes graphics.
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Statement of Responsibility:
by C. Alan Canfield.

Record Information

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University of Florida
Holding Location:
University of Florida
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002456838 ( AlephBibNum )
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I owe thanks to many individuals for supporting me in my graduate studies at the University

of Florida (UF) while working full-time. I extend special and sincere thanks to my advisor and

committee chair, Dr. Zhuomin Zhang, for providing valuable advice in the thesis preparation and

graduate studies. Additional thanks go to Dr. Charles Proctor, III, for guidance in establishing the

thesis proposal and background review, and Drs. William Lear and Alex Green for serving on my


All graduate coursework was conducted through the distance learning program facilitated by

the Florida Engineering Education and Delivery System (FEEDS) offices at UF and the Florida

State University Panama City Campus (FSU/PCC). I sincerely appreciate the dedication and

patience of Professors Hsieh, Zhang, Mittal, and Kurzweg for accommodating off-campus

students in their courses. Ms. Becky Hoover of the UF Mechanical Engineering Graduate Office,

Ms. Joyce Phillips of the UF FEEDS office, and Ms. Pat Lawson of the FSU/PCC FEEDS office

were invaluable in course administration. Mr. Oliver Canaday of the Tyndall AFB Education

Center and Ms. Sheila Ray of the Tyndall AFB Library supported the receipt and viewing of my

course videotapes. Without the support of these individuals, I would not have been able to pursue

or complete this program.

I gratefully acknowledge and value the long-term support and friendship of Dr. Joe Wander

of the Air Force Research Lab, Airbase & Environmental Technology Division. Dr. Aly Shaaban

of Applied Research Associates graciously provided testing data for evaluation. Finally, I

dedicate this thesis to my parents, Charles and Katherine Canfield.


This thesis is the result of graduate studies and research toward a Master of Science in

mechanical engineering at the University of Florida. The coursework and research was

conducted while working full-time for Applied Research Associates, Inc. at Tyndall Air Force

Base, Florida, supporting the Air Force Research Laboratory (AFRL) Airbase & Environmental

Technology Division. Dr. Joseph D. Wander was the Air Force Project Officer.

Most of the test data evaluated in this thesis was collected by Dr. Aly Shaaban in 1996, with

the remainder collected in late 1998 by the author. A project to implement the alternative fuel

described herein for a six-month test is planned for early 1999. A patent application is under

review for the additive package used for maintaining a stable fuel-water mixture, preventing

detailed description of the additive components.


A C K N O W L E D G E M EN T S ................................................................................ ....................

P R E F A C E ................................................................................................................................. iii

A B B R E V IA T IO N S ......................................................................... ............................. vi

A B ST R A C T ............................................................................... .. vii


1. IN T R O D U C T IO N ...................................................... .......................
Background......................................................................... ............... 1
O objective ........................................................................... . . 3
O u tlin e ....................................................................................................... . . 4
History of Internal Combustion Engines ................ ............. .................... 5
Physical Engine Characteristics .......................................................... ......... ...... 6
IC E Therm odynam ic C ycles.......................................................... ............... 8
IC E M mechanical Cycles ..................................... ................ ......... ............... 12
Fuel/A ir M fixtures ............................................... ....... ................ 15
First Law of Thermodynamics Analysis ............................................. ............... 17
H eat T transfer E effects in IC E s ................................................................................... 19
3. INTERNAL COMBUSTION ENGINE EMISSIONS ..................................................... 26
P ollutant F orm action ................................................................................ ............... 2 6
C ontro l T ech n o log ies ................................................................................................. 3 1
4. EXPERIMENTAL METHODS AND PROCEDURES........................................................39
A /M 32A -86 P aram eters ............................................................................................. 39
Fuel Properties ...................................................................... .......... 41
S am p lin g P rocedu re ............................................................ .................................. 4 3
D ata Collection ...................................................................... ......... 45
5. RESULTS AND DISCUSSION .......................................................... ...............48
E xperim mental R results ........................................................................... .................... 48
First Law of Thermodynamics Calculation ......................................................... 60
Equilibrium Code Calculation .......................................................... ......... ...... 64
N O x Form ation R ate Calculation .................................................................... ....... 69
E effects of W ater Injection........................................................................... ............ 72


6. CONCLUSIONS AND RECOMMENDATIONS ....................... ..........................74
R e su lts .............. ...... ............ .................. ...................................................... 7 4
C conclusions .............. ........................................... ....................................... 75
Recommendations for Further Study ..... ................................. ......... ......................75

N O M EN C LA T U R E ........... .................. ..................... .............................. 77

REFERENCES ................................... ......................... 79

B IO G R A PH IC A L SK ET C H ......... .................. ......................................... .......................... 85


Air Force Base
aerospace ground equipment
Air Reserve Base
bottom dead center
brake specific fuel consumption
compression ignition
cetane number
carbon monoxide
compression ratio
Environmental Protection Agency
internal combustion engine
military specification
nitrogen oxides
nitric oxide
nitrogen dioxide
spark ignition
semi-volatile organic compound
top dead center
unburned hydrocarbons
United States Air Force
volatile organic compound


Abstract of Thesis Presented to the Graduate School
of the University of Florida in Partial Fulfillment of the
Requirements for the Degree of Master of Science



C. Alan Canfield

May 1999

Chairman: Zhuomin Zhang
Major Department: Mechanical Engineering

This study examines the effects of combusting a mixture of diesel fuel, water, and surfactant

on the nitrogen oxides, or NOx, emissions from a compression ignition diesel engine. Previous

research has attributed the observed reduction of nitrogen oxide emissions to a suppression of

flame temperature due to quenching effects from the water, thereby reducing thermal NOx

formation. The thesis highlights the relevant theory, operation, and design parameters of diesel

internal combustion engines. Experimental procedures conducted using a Detroit Diesel 4-

cyclinder diesel engine are discussed. Results from testing diesel fuel with varying ratios of

water balanced with a surfactant to stabilize the emulsion will be presented and discussed. The

data shows significant NOx emission reduction with up to 45 percent water, by volume, in the

fuel. These results are correlated with a thermodynamic first law analysis to estimate the

adiabatic flame temperature of the standard fuel and fuel-water emulsion cases. Results indicate

that thermal NOx is indeed reduced by quenching and flame temperature suppression, confirming

reports in the literature. Recommendations are given for further studies, including improving the

fuel-water emulsion and considerations for long-term testing.




McClellan Air Force Base (AFB), California, identified mobile diesel engines as contributing

nearly as much total NOx emissions as aircraft and stationary sources combined. Hourly-rated

diesel engines contributed 75 percent of this NOx, with the remainder emitting from gasoline and

diesel engines rated in miles (Canfield et al., 1997). Hourly-rated diesel engines include non-road

mobile equipment such as forklifts, bulldozers, air-conditioning units, and generators. McClellan

AFB is located in Sacramento, California, classified as Extreme for ozone attainment status. This

classification recognizes ozone pre-cursors, NOx and VOCs, as immediately dangerous to human

and environmental health, and places the highest restrictions on emissions and fines for exceeding

these limits.

In addition to air quality compliance concerns at McClellan AFB, March Air Reserve Base

(ARB) in southern California faced immediate reduction in training and operational use of AGE

generators due to South Coast Air Quality Management District (SCAQMD) Rule 1110.2. This

Table 1: Regulatory emission limits and emissions from A/M32A-86

Rule 1110.2 A/M32A emissions by load, ppm
Air pollutant limit, ppm 18 kW(e) 36 kW(e) 72 kW(e)
NO, 36 761 1171 2118
CO 2000 54 47 95

rule placed regulatory limits on emissions from non-road mobile sources over 37 kW. These

limits, and the corresponding emissions from an A/M32A-86 diesel generator at varying electric

generator loads, are listed in Table 1. The A/M32A-86 generator, the subject of the present study,

is described in detail in Chapter 4. Please note in Table 1 the limit of 36 ppm for NOx emissions,

while the A/M32A-86 emits over 20 times this at 18 kW, only 25 percent of full load.

AGE does not currently meet these emission limitations and will not without emission

controls and/or combustion system or fuel modifications. All AGE units over 50 hp were

required to comply with Rule 1110.2 emission limits by 31 Dec 94 or the facility would: 1)

remove the AGE units from service, 2) pay daily fines, 3) or submit a proposal for complete

electrification by Dec 1999.

AGE at March ARB fell under three general categories:

1. Gasoline-fueled internal combustion engines (ICEs)

2. Diesel-fueled ICEs

3. JP-8-fueled turbine engines

There were 81 total pieces of AGE over 37 kW, excluding turbine-powered AGE, at March ARB

in 1996. The 168 units of AGE under 37 kW are not presently regulated by Rule 1110.2.

The Air Force surveyed its laboratories for potential solutions to air pollutant emissions from

AGE. A team was formed to investigate commercial off-the-shelf (COTS) products or

developing technologies for emission controls or alternate power sources applicable to AGE.

Laboratories with promising technologies were tasked to prepare preliminary and final proposals

for demonstration, validation, and implementation of their technologies. A request for

information was also posted in the Commerce Business Daily to survey COTS for AGE emission


Several technologies proposed by the Air Force laboratories to control NOx emissions from

AGE included selective catalytic reduction, increasing 02 in the intake air, a mobile filter cart, a

nonthermal plasma reactor control system, and water injection with the fuel, described in this


Dryer (1976) reports a reference to water injection in a combustion system dating back to

1791, in which water was used to cool the blades in early gas turbines. In the last 20 years water

has received varying levels of interest as a means to improve combustion efficiency and reduce

air pollutant emissions from ICEs. These reports in the literature will be discussed and compared

with the current research.

Early work demonstrated that with fuel:water volumetric ratios from 1:1 to 9:1, the addition

of 1-2 volume percent surfactant allowed a stable emulsion to be maintained with the diesel and

water mixture. Using this formula and testing varying ratios of fuel and water for combustion

properties developed the data sets that will be evaluated in this thesis. Various injector timing

setting were also evaluated. For a given injector timing setting, data will be presented for the

baseline case of the generator operating on standard military diesel, and for various fuel mixtures

of different water volume ratios.


The purpose of this research is to investigate the use of a stabilized diesel-water fuel

mixture as a drop-in replacement in U.S. Air Force (USAF) mobile aerospace ground equipment

(AGE). The USAF is interested in lowering the emissions of nitrogen oxides (NOx) and other

emissions from AGE during training and non-warfighting missions to comply with air quality

regulations governing facilities in the continental United States.

The tasks involved in this research involved (1) evaluating the performance and behavior of

the fuel-water mixtures and (2) measuring the effect of the fuel-water mixture on AGE power

and air pollutant emissions. Previous unpublished work at the author's laboratory determined the

optimal diesel-water emulsion preparation techniques and mixture ratios. Research reported

herein focused on evaluating varying ratios of water in the fuel and the resultant effect on air

pollutant emissions, especially NO,.

This study was conducted as part of an Air Force initiative to reduce emissions from

aerospace ground equipment (Akridge et al., 1997, and Canfield et al., 1997). Of six technologies

evaluated for AGE NO, reduction, the fuel-water emulsion promises the easiest "drop-in"

solution. NOx emissions can be lowered with little or no modifications required to the diesel



This thesis begins with a review internal combustion engines in Chapter 2, including history,

physical components and geometry, and mass and heat transfer aspects. Chapter 3 will discuss

internal combustion engine air pollutant formation patlhia\\s and existing control technology

options. The test engine, experimental procedures, fuel properties and variables, and data

collection will be described in Chapter 4. Chapter 5 will present the experimental data and

results, with a comparison to first law of thermodynamics and equilibrium products code

predictions. NOx formation rate equations will be presented in Chapter 5, and the potential

effects of water injection on combustion will be described. Chapter 6 will discuss the findings,

conclusions, and recommendations from the research, including confirmation of research

published in the literature.



History of Internal Combustion Engines

From the first application of open fire to provide heating, lighting, and cooking, combustion

science has evolved to providing distributed electricity generation and mechanical energy for

most modes of transportation. Combustion is described as either external or internal. External

combustion is defined as combustion in which the process fluid is external to, or different from,

the mechanical energy-producing fluid. For example, coal-fired power plants operate as external

combustion because the coal is combusted to generate steam, and the steam then turns a turbine to

generate electricity. If the fluid undergoing combustion also generates mechanical energy in the

system, the process is defined as internal combustion. For example, in reciprocating internal

combustion engines the gas expansion from combustion of the air and fuel mixture moves a

piston, which turns a crankshaft, generating mechanical power for propulsion, electricity

generation, etc. Gas turbines and rocket engines are also defined as internal combustion, since, in

both cases, the air and fuel mixtures after combustion and compression provide mechanical power

through thrust.

Advances in the understanding of the thermodynamic cycles of combustion have improved

the design of combustion engines. Improvements in materials science also support improvements

in engine performance and durability. For example, advances in materials and design have

improved the compressor ratio of gas turbines from 3:1 to 30:1, and increased the efficiencies

from 3.5% to 30%. Similarly, computer modeling and simulation has supported advanced intake,

Exhaust valve


m Exhaust manifold

Spark plug
Coolant Cylinder


Connecting rod


Figure 1: Spark-ignition engine cross-section

combustion cylinder, and exhaust systems design on modern automobiles resulting in higher

thermal and mechanical efficiencies with reduced air pollutant emissions.

Physical Engine Characteristics

Before discussing the operation, design, and analysis of internal combustion engines (ICEs),

it is worthwhile to review the specific engine components involved in and affecting the

combustion process. Figure 1 shows a cross-section of a spark ignition, two-stroke engine.

Important characteristics of the internal combustion engine are listed in Table 2.

Critical to evaluating the performance of an ICE are quantities derived from the geometry of

the combustion cylinder and the motion of the piston in the cylinder. A simplified piston,

Table 2: Physical engine components




Combustion chamber

Intake manifold

Exhaust manifold

Inlet valves

Exhaust valves

Spark plug

Description and Function
Channel of circular cross-section bored into the engine
block in which a piston moves linearly in a
reciprocating motion
Cylindrical component riding back and forth in the
engine cylinder converting the thermal energy released
by the combustion process into mechanical energy
Shaft with offsets to hold the piston connecting rod and
translate linear piston motion to circular motion
Portion of the cylinder enclosed by the piston and the
head of the cylinder. When the piston reaches the top
or extent of motion into the cylinder and combustion of
the fuel-air mixture occurs, the themal energy released
raises the pressure in the combustion chamber and
forces the piston back.
The collection of pipes carrying air (fuel injected) or
air and fuel (carbureted) to the engine inlet valves.
Collection of pipes carrying the combusted air/fuel
mixture away from the engine.
Poppet valves that control the introduction of air or air
and fuel into the engine combustion chamber.
Poppet valves that control the release of combusted
air/fuel mixture to the exhaust system.
Electrode protruding into the combustion chamber. A
high-voltage arc is passed across an electrode to
provide ignition in the lower-compression spark
ignition engines.

cylinder, connecting rod, and crankshaft are shown in Figure 2. The compression ratio, CR, is the

ratio of the maximum cylinder volume to minimum cylinder volume,

CR +

where Vc is the cylinder clearance volume, cm3, and V, is the cylinder swept volume, cm3. The

cylinder volume at any crank position 0 is given by

c ------------ TDC


b = cylinder bore, cm
V, = clearance volume, cm3
V, = swept volume, cm3
I, = stroke length, cm
d TDC = top dead center
BDC = bottom dead center
d, = distance from piston pivot
axis to crank axis, cm
S1,I = connecting rod length, cm
8 = crank angle, rad
r, = crankshaft radius, cm


Figure 2: Reciprocating ICE geometry (not to scale)

v v + nb2 (+ dp). (2)

Please refer to Figure 2 for additional ICE geometry nomenclature. The stroke length, ls, and

the cylinder bore, b, are critical in determining the power output of the combustion process.

ICE Thermodynamic Cycles

Combustion in reciprocating piston ICEs is commonly assumed to operate either as a constant

volume or constant pressure process. The Otto cycle, a constant volume heat addition

thermodynamic process, closely models combustion in spark ignition (SI) ICEs. The diesel cycle

is a constant pressure, slower-speed cycle depicting combustion in compression ignition, or

diesel, cycle engines. A combination of the Otto and diesel cycles is referred to as the mixed,


a b

b a5
b a

a b

1 5

b a

5 1: Suction, air intake
1 2: Adiabatic compression
2 3: Constant volume heat addition
3 4: Adiabatic expansion
4 1: Constant volume heat rejection
1 5: Exhaust to atmosphere

Otto Cycle

5 1: Suction, air intake
1 2: Adiabatic compression
2 3: Constant pressure heat addition
3 4: Adiabatic expansion
4 1: Constant pressure heat rejection
1 5: Exhaust to atmosphere

Diesel Cycle

Figure 3: Ideal air-standard Otto and diesel cycles

limited pressure, or combination cycle. The Otto and diesel cycles are shown in Figure 3 to

demonstrate the differences. As mentioned, the constant volume heat addition of the Otto cycle

and the constant pressure heat addition of the diesel cycle are shown in step 2 -> 3 of the

respective set of graphs. The primed points (i.e., 2' and 3') depict the non-isentropic expansion

points attained due to irreversibility. The label s = c and v = c refer to the ideal constant specific

entropy and specific volume for the respective step.

The graphs shown in Figure 3 depict the theoretical, adiabatic compression and expansion of

the Otto and diesel cycles. Actual combustion processes vary from theoretical due to various

losses. Some of these losses are depicted in Figure 4 for the diesel, or compression ignition (CI),


Differences in Figure 4 for actual verses theoretical cycle end-points are explained below:

5' -> 1: The suction stroke provides intake air to the engine cylinder for fuel oxidation. Slight

losses can occur from restrictions and bends in the intake air hose, or clogging of the intake air

filters. Poor intake airflow can also contribute to poor mixing of the air and atomized fuel.

1 -- 2': This is the compression stroke, required for auto-ignition of the air-fuel mixture.

Compression does not reach the pressure indicated at point 2 due to crevice losses and gas

slippage past the piston rings.

2' -- 3': This is the combustion process in which the fuel-air mixture is auto-ignited. The

ideal constant-pressure heat addition shown from 2 -> 3 in Figure 3 is not attained due to finite

(vs. instantaneous) burning time for combustion of the fuel-air mixture.

3' -- 4': The expansion process is the power stroke for the engine. Finite burning time and the

effects of incomplete combustion lower the pressure attained at point 3'. Incomplete combustion

Finite burning time

P 2 3
Heat transfer losses

Low volumetric Early exhaust
efficiency 5'

Exhaust losses 5 1
a b
5' 1: Suction, air intake
1 2': Compression
2' 3': Ignition and combustion
3' 4': Expansion
4' -1: Exhaust opens
1 5': Exhaust to atmosphere

Figure 4. Compression ignition (diesel) cycle with losses

is a large source of losses in a CI engine, since combustion often continues during up to half of

the expansion stroke Heat transfer to the cylinder wall removes heat from the combustion gases,

lowering the thermal efficiency.

4' -> 1: The exhaust valves are required to open slightly before bottom-dead center (BDC) to

allow sufficient time to remove all combustion by-products from the cylinder. Due to this early

opening of the exhaust valves, the power stroke is complete at a lower pressure of point 4' than

ideal, shown by point 4.

1-- 5': Exhaust system back-pressure is the primary source of losses during exhaust. Engine

power is consumed to overcome the pressure drop in the exhaust system to expel combustion

products from the engine cylinder.

Spark Ignition Engines

SI automobile engines use higher-volatile gasoline as fuel and operate with compression

ratios of 6:1 to 12:1, limited on the upper end by the tendency for the volatile fuels to cause

"knocking." An ignition source such as a spark plug is required to initiate combustion.

Compression pressures range from approximately 1000 kPa to 2000 kPa. Load and speed of the

engine are controlled by throttling the fuel charge added. Several advantages include a low cost

and relative weight, low starter cranking energy required, large range of speed and load, relatively

high mechanical efficiency, and low specific fuel consumption. The primary disadvantages

include a lower thermodynamic efficiency and high levels of air pollutant emissions across all

load levels.

Compression Ignition Engines

CI engines use fuels of lower volatility, with compression ratios from 11:1 to 22:1 and

compression pressures between approximately 2700 and 4800 kPa. As the name implies, the high

compression pressures of the CI engine ignite the fuel/air mixture, so no ignition source (e.g.,

spark plug) is required. Advantages of the CI engine over the SI engine include a lower specific

fuel consumption, slightly higher thermal efficiency, relatively cheaper fuel costs, lower CO and

hydrocarbon emissions at low and medium loads, lower capital costs, and higher durability.

Disadvantages include higher noise of operation, higher engine weight required to withstand the

higher pressures, and excess oxygen in the exhaust preventing use of standard catalysts for air

pollutant control.

CI engines can be characterized by the injection type-either direct injection (DI) or indirect

ignition (IDI). The Detroit Diesel 4-71N test engine is DI, implying that the fuel is injected

directly into the combustion cylinder to mix with the intake air. IDI engines mix the fuel and air

prior to entering the combustion cylinder in an attempt to improve mixing and therefore

combustion. CI engines can additionally be characterized by the number of strokes required per

power cycle, discussed below.

ICE Mechanical Cycles

Reciprocating internal combustion engines (both spark ignition and compression ignition) can

be broadly characterized as four-stroke or two-stroke, depending on the number of piston strokes

required for one power cycle. While the subject engine is a two-stroke compression ignition

engine, it is worthwhile to describe both cycles and the inherent differences.

Four-Stroke Engine Cycle

The four-stroke engine cycle requires four piston strokes for one power cycle, which occurs

with every two revolutions of the crankshaft. The step-by-step power cycle for a four-stroke ICE

is shown in Figure 5. The power cycle shown is divided into intake, compression, expansion, and

exhaust cycles, corresponding to steps 5 1, 1 -> 2, 3-> 4, and 5 1, respectively, from the

Otto cycle in Figure 3.

The inlet valve opens, allowing intake air in to the cylinder, to begin the four-stroke power

cycle. The piston is moving down in the cylinder and the exhaust valve is closed. As the piston

passes BDC, the inlet valve closes and the compression stroke begins. Please note Figure 5

S Closed Closed Closed Closed Closed Closed

Intake -' Exhaust
air II. I_ _, gases-

J I 1: U I

Intake Compression Expansion (Power) Exhaust

Figure 5: Four-stroke spark-ignition ICE cycle

depicts a spark ignition four-stroke cycle. A spark plug is shown applying a spark in the cylinder

near the completion of the compression stroke at TDC. As mentioned earlier, spark-ignition

engines require an ignition source to initiate combustion of the fuel-air mixture. A four-stroke

compression ignition engine is fundamentally identical except for the lack of a spark plug, and

more robust to withstand higher compression ratios.

The expansion stroke produces mechanical power from the expanding combustion gas. The

inlet and exhaust valves remain closed during the compression and expansion strokes. After the

piston passes BDC, the exhaust valve opens and the piston is used to drive out the combustion

gas byproducts from the cylinder. This piston clearing of exhaust gases is the fundamental

difference between four-stroke and two-stoke engine cycles, which are discussed next.

Two-Stroke Engine Cycle

The two-stroke engine power cycle only requires two piston strokes, or one revolution of the

crankshaft. The two-stroke combustion process for the Detroit Diesel 4-71N engine, used in this

study, is shown in Figure 6. During the scavenging process shown Figure 6, the exhaust valve is


F el

Air Air Air

Scavenging Compression

Air Air Exhaust

*0.... -

Power Exhaust

Figure 6. Detroit Diesel 4-71N two-stroke combustion
process (adapted from SM-ALC/TISEA, 1986)

open and the piston is at BDC. A blower operates continuously on mechanical power from the

engine. A blower is not required but improves combustion product exhaust from the cylinder.

During scavenging, the blower is pushing air through slots near the bottom of the cylinder, in

preparation of the compression stroke. The exhaust valve remains open to complete removal of

combustion products from the cylinder.

When the compression stroke begins and the piston is rising in the cylinder, the exhaust valve

is mechanically closed. Please note the lack of an intake valve: as the piston rises in the cylinder

it passes and closes the intake openings. At a point designated by the injector timing setting,

atomized fuel is added to the compressing air by the fuel injector. The process is approaching the

thermodynamic point 2' in Figure 4.

When the piston reaches TDC, the fuel-air mixture is fully compressed and the mixture is

auto-ignited, corresponding to step 2' -> 3' in Figure 4. The expansion (power) stroke begins, in

which the expanding gas is pushing the piston down the cylinder, generating mechanical energy

that is transferred to the crankshaft. The expansion (power) stroke occurs in step 2' -- 3' in

Figure 4.

As the piston approaches BDC, it passes the intake slots in the bottom of the cylinder,

opening the cylinder to intake air. The exhaust valve is also mechanically opened. Burned bas

by-products are displaced through the exhaust valve by blower-forced intake air. The piston

passes BDC and the scavenging process begins, repeating the power cycle. Note in Figure that

the crankshaft makes only one revolution per power cycle. Recall Figure 5 depicted two

revolutions and four strokes of the piston for a single power cycle.

Fuel/Air Mixtures

For practical applications combustion air can be approximated as 21% oxygen (02) and 79%

atmospheric nitrogen (N2). Thus, for each mole of 02 in air there are =3.76 moles of N2.

The fuel/air equivalence ratio, 0, is defined as the ratio of actual fuel/air mass ratio, (F/A)actual, to

the stoichiometric fuel/air ratio, (F/A),:

(F /A)V,, (3)
(F /A),

For = 1, combustion is stoichiometric, for 1 < 1, combustion is fuel-lean (excess air is used in

combustion), and for 1 > 1, combustion is fuel-rich. Diesel engines operate significantly fuel-

lean, with typical values of = 0.8. F/A is usually expressed on a mass basis as kg fuel per kg

air, but a molar basis can also by used, kmol fuel per kmol air. The simplified combustion

equation for a hydrocarbon fuel CxHy can be represented as

CHy +a(2 + 3.76N2) bCO2 +cCO+dH20(g)+eH2 + fO2 + 3.76aN2 (4)

For stoichiometric and fuel-lean conditions, 50 1, Equation (4) becomes

CxH, +a(2 + 3.76N2) -bCO2 + dH20(g)+ fO2 + 3.76aN2, (5)

while for fuel-rich conditions, 0 > 1, Equation (4) becomes

CHy +a(O2 +3.76N2) -bCO2 +cCO + dH20(g)+eH2 +3.76aN2, (6)


a = 4 (7)

Equations (4) through (6) are expressed for 1 kmol of fuel. Examining Equation (5) we see

the coefficients for CO and H2, products of incomplete combustion, are zero, since the equation

represents the fuel-lean combustion case with excess air present to ensure complete combustion.

The presence of 02 as a product in Equation (5) further represents the excess air present in

combustion of the fuel. Additionally, for the case of = 1, Equation (5) simplifies to

stoichiometric combustion without 02 present as a product. Similarly, the inclusion of CO and H2

as products of incomplete combustion in Equation (6) demonstrates that fuel-rich combustion is

represented, in which insufficient oxygen is present to drive the reaction to completion.

Since compression ignition engines, the subject of our study, operate with typical equivalence

ratios of 0.8 (0 = 0.8), we will evaluate Equation (5). For 5 < 1, C, H, and O-atom balances on

Equation (5) yield:

b =x (8)

c = 0 (9)

d = (10)

Heat transfer


Reactants:_ 1 cv Products:
I---" -- --- ---1
TR =298K Fuel T, = 2000K
R = 100kPa hf,hf, QHHV Exhaust gas p = 100kPa
nR Engine w- np
-h Air rf +rha,he h,
hR a'h Ahp
A a a -- ---_______________ P -
Control volume

Figure 7: Simplified control volume around internal combustion engine

e=0 (11)

f ={( x+ )=(1-)a (12)

These coefficients are used to balance the combustion equation for a first law of thermodynamics

analysis of the reacting combustion system.

First Law of Thermodynamics Analysis

A control volume is depicted around a combustion cylinder in Figure 7 for demonstration of a

first law analysis. The first law of thermodynamics relates changes in internal energy on a

control volume to heat and work interactions by the control volume with the environment. A first

law analysis deals with end states only-chemical reactions or other changes in the species during

reaction are not evaluated. For this reason, the zero internal energy, or enthalpy, reference state

of species acting on the control volume must be at the same conditions. Temperature and

pressure will define the reference state for the pseudo-constant pressure compression ignition

process. Tabulated values of enthalpy for reactants and products will be used, and the

combustion system's internal energy decreases because the process is exothermic.


Control volume
r -------------
Se- Exhaust
Engine Shaft work
Cooling Cooling
water water

Lubricating Wall
oil radiation

Figure 8: External control volume around ICE

Figure 7 shows fuel and air entering the control volume with enthalpies of hf and ha, and mass

flow rates of mf and mha, respectively. Exhaust gas leaves the control volume with enthalpy he

and mass flow rate of m f + mh Enthalpies of formation for reactants are balanced with the

change in enthalpy of products in the form

Qcv + nhn, =Wcv + n,h, (13)

where Qcv is heat transfer from the control volume, Wcv is shaft or other work done by the

control volume, and summations are of n moles and h enthalpies of reactants and products.

Enthalpies are evaluated at reactant and product temperatures, TR and Tp. Moles of reactant and

product species, nR and np, are described by the governing combustion equation, such as Equation

(5). Equation (13) can be used to find the heat transfer from the control volume given the flame

temperature, or the flame temperature can be estimated using an iterative process demonstrated in

Chapter 5.

The control volume can be expanded to include losses from the internal combustion engine.

Figure 8 depicts a control volume around an ICE including heat inputs and losses, and shaft work.

The heat input to an ICE comes from the heat value of the fuel combusted, cooling water inlet

temperature, and lubricating oil inlet temperature. Energy leaves the control volume of the ICE

in the form of heated lubricating oil and cooling water, shaft work, exhaust gases, and radiative

heat transfer to the crankcase and cylinder. Other losses can include incomplete combustion,

blowdown, and pumping. Blowdown occurs during the heat addition process: as the fuel-air

mixture ignites and begins expanding, gases can slip past the cylinder rings into the crankcase.

Pumping losses are incurred through power used to pump the lubricating oil and cooling water.

Heat transfer modes and effects in ICEs are discussed in the following sections.

Heat Transfer Effects in ICEs

Conduction, convection, and radiant heat transfer are significant to the design, operation, and


Tgas hcl
c, cool

SGas Coolant
w, gas
D Tw, cool
I -e,hc, gas CO

qco'v + qqD qcoAiv

0 b x

Figure 9: Heat transfer from combustion gas to coolant

evaluation of internal combustion engines. Figure 9 depicts the gas and coolant temperature

versus distance from the combustion cylinder centerline in an ICE. Temperature, T, is plotted on

the y-axis, which corresponds with the combustion cylinder centerline at x = 0. The temperature

of the gas in the cylinder falls as it approaches the inner wall of the cylinder, at 2, where b is the

cylinder bore. Both convective heat flux, qco v, and radiative heat flux, q D, occur between

the combustion products and the inner cylinder wall. The gas-side convective heat transfer

coefficient, h,gas, and the emissivity, e, at the inner cylinder wall are needed to calculate qcoNV

and q'AD Heat is then conducted through the cylinder wall, shown as q'oND, with cylinder wall

thickness tw. Temperature through the cylinder wall is reduced from T, on the gas side to Twcoo

on the coolant side. Temperature drops through the thermal boundary layer on the coolant side

near the cylinder wall, as governed by the convective heat transfer coefficient, ho,coo, between the

coolant and outer cylinder wall. The individual modes of heat transfer, governing equations, and

relevant effects on ICEs are discussed below.


Conduction heat transfer refers to the transfer of heat by molecular motion and interaction in

solids and non-moving fluids. Heat flux (q") is governed by the well-known Fourier's Law for

steady heat conduction,

qCD = -kVT (14)

where k is the thermal conductivity, W/m-K, and VT is the gradient of temperature, defined as

a2T a2T a2T
T+ T+ T In addition to conduction heat transfer through the cylinder wall, as shown
ax2 ay2 az2

in Figure 9, heat is also immediately conducted from the combustion gases through the piston,

valves, and cylinder head.


Heat transfer by convection occurs between a fluid in motion relative to a fluid or solid

surface. For steady flow-forced convection, or convection driven by forces other than gravity, the

convective heat flux from a fluid at temperature T to a solid surface at Tw is given by

qcoNv = h(T-Tw) (15)

where he is the convective heat transfer coefficient, W/m2-K. Actual flows in the combustion

cylinder are turbulent and unsteady, and detailed analyses require the use of the conservation,

momentum, and energy equations. Convection is the primary form of heat transfer in the ICE,

and Table 3 summarizes the processes in which heat is transferred by convection from the

combustion gases in an internal combustion engine.


Radiation heat transfer occurs through the emission and absorption of electromagnetic waves

in the visible (0.4 pm to 0.7 pm) and infrared (0.7 p.m to 40 p.m) range (Eckert and Drake, 1987).

Although all substances radiate energy, it is only at elevated temperatures that the heat flux

radiated becomes significant. This is demonstrated by the radiative heat flux equation

qRD =E (T,4 Tr) (16)

Table 3: Summary of convection heat transfer in an ICE

Heat Source Convective Heat Flux To:
Combustion gases (compression, Cylinder heads, cylinder valves,
expansion) cylinder walls, pistons, and O-rings
Outside cylinder walls, cylinder heads Engine coolant
Piston, O-rings Lubricant (or piston coolant, if
Combustion gases (exhaust) Exhaust valves, exhaust ports, and
exhaust manifold
Intake manifold Intake air
Coolant Radiator, environment

where e is the emissivity of the surface, a is the Stefan-Boltzmann constant, T, is the maximum

blackbody temperature of the surface, and Ts,, is the temperature of the surrounding environment.

Shape factors are applied to account for fractions of radiation from a source interacting with the

recipient surface. In combustion engines the radiant heat sources are primarily particulate (soot),

water vapor, and CO2 (Zhou et al., 1987). This radiant heat falls on the cylinder walls, piston

surface, and valve body.

ICE radiant heat transfer is difficult to estimate and measure. Radiant heat transfer is also

more significant in diesel engines than spark ignition engines, due to higher soot content of the

diesel combustion products. High-temperature intermediate soot particles and high flame

temperatures are the source of radiant heat to the cylinder walls, piston, and combustion gases.

The radiant heat transfer due to soot depends on the particle size distribution, number density,

flame geometry, and, as shown in Equation (16), the soot temperature. Emissivity reportedly

ranges from 0.8 to 0.9 during periods of peak radiation. The average radiant heat flux reportedly

ranges from 5 to 50 percent of total heat flux, rising with increases in engine load (Kays, 1989).

The discussion here has focused on the heat transfer and thermodynamic effects of the

combustion process, but engine materials require cooling, and the design of this cooling system is

a significant aspect of overall ICE design. Kays (1989) evaluated the design of air and liquid

cooling systems for ICEs according to fundamental heat exchanger principles. He demonstrated

that engine head design, radiator manufacturing methods and costs, and coolant pumping losses

often control radiator design, verses optimal waste heat exchange from the cylinder to the coolant.

Engine Variable Effects

The ICE is a thermodynamic heat engine, so any changes to components, parameters, or work

per unit time will affect the heat flux through the engine to the environment. Table 4 summarizes

how changes in important engine variables affect heat flux from the combustion gases.

Several trends listed in Table 4 are significant to the present study. Increasing the load or

speed of an engine, thus increasing the power output, requires increasing the fuel/air mixture

flow. With an increase in Mi f, the fuel flow rate, the heat transfer rate q" through the engine is

increased from the additional heat input. Conversely, reducing Thf will lower q". This effect

correlates to the present study when considering the if, diesel fuel component of the diesel-

water mixture is decreased when total ui*f remains constant through the engine for a given load or

speed. In other words, if the total fuel flow rate is not changed when adding water to the fuel,

then the combustible fuel flow rate is reduced, lowering the heat transfer rate in the engine.

From the discussion of convection heat transfer and Equation (15), if the convective heat flux is

Table 4: Engine variable effects on heat flux

Variable Direction Effect Discussion
Speed or load 1 Greatest effect, increases Increases i f, fuel flow rate
heat flux, q"
Equivalence ratio, 1 q" increases to 0 = 1.1, Maximum thermodynamic
(AF/AF, ratio) then decreases gas temperature reached
Compression ratio, r 1T Decreases q"until Lower expansion and
r, = 10, then q"increases exhaust temperatures
Injection timing (CI) Delay Decreases q" Combustion occurs during
larger cylinder volume, Vc
Spark timing (SI) Delay Decreases q" Combustion occurs during
larger cylinder volume, Vc
Swirl, mixing T Increases q" Gas velocities increased
Inlet temperature 1 Increases q" Linear increase in
temperature throughout heat
exchange processes
Coolant temperature T May reduce q" Increases component
Cylinder wall thermal Decreases q"N through May cause heat flux to
conductivity, k cylinder wall combustion gases from wall,
resulting in pre-ignition

lowered, then the temperature difference between the combustion gases and the cylinder wall will

also be lowered. Given a constant convective heat transfer coefficient and constant coolant

temperature, we can theorize that the flame temperature will be lower to result in this lower heat


Reducing cylinder wall thermal conductivity k decreases the conductive heat transfer through

the wall. This can result in a higher inner cylinder wall temperature and lead to pre-ignition of

the fuel-air mixture. The effect is similar that of insulating the cylinder, which can improve

compression ignition engine performance when properly designed (Heywood, 1988).

Efficiency and performance gains in ICEs have been significant, and combustion

modifications and control devices have substantially reduced air pollutant emission rates.

Modeling and experimental analysis of fuel spray, atomization, and vaporization has improved

the understanding of the onset and progression of diffusion combustion in the diesel engine.

Spray formation is increasingly important as advanced swirl designs are incorporated into the

combustion chamber to improve mixing.

Researchers are also improving numerical flow simulation methods for internal combustion

processes. Some models have been under development and refinement for nearly 30 years.

Boundary value solutions to the differential equations using finite difference methods are

valuable for general solutions. Finite element analysis offers a solution to correlate predicted and

experimental results, because the three-dimensional physical constraints can be used as

boundaries for the heat transfer and momentum equation solutions (Campbell, 1979).

Fuel ignition and resultant flame quenching is another area of focus in internal combustion

engines. Research includes modeling and understanding the rates and patterns of fuel ignition,

including the spray pattern effects, and the passive or active quenching of the flame. Active

quenching of the flame is often accomplished by introducing water, alcohols, or other lower

heating value additives to the fuel, which is the premise of the present research. Passive flame

quenching occurs at the engine combustion chamber walls. Flame quenching causes incomplete

combustion and thus results in higher emissions of unburned hydrocarbons.

Improvements in measurement techniques are important to all aspects of ICE analysis. An

experimental study of ICEs involves temperature measurements somewhere in the combustion

process. As shown in the convective and radiant heat equations, Equations (15) and (16), the

cylinder wall temperature, Tw, is fundamental to ICE heat transfer analysis. The high combustion

temperatures can interfere with thermocouple signals, and the unsteady nature of the combustion

field makes direct measurement of radiant flame temperature difficult. Spectroscopic techniques

using optical fibers to transfer a flame image from the combustion chamber to a photoelectric

transducer can be used to measure the spectral illuminance, which is then correlated to the flame

temperature (Nagese and Funatsu, 1990).



The review thus far has focused on the physical and thermodynamic properties of internal

combustion engines. The chemical parth as\\ of formation, mitigation methods, and engine

factors affecting air pollutant emissions from ICEs are discussed in the present chapter. This

review is important in predicting and evaluating how the ICE air pollutant emission levels

respond to the alternative fuel.

Pollutant Formation

The air pollutant emissions from internal combustion engines vary with the operational and

design parameters discussed earlier in Chapter 2. Other factors affecting air pollutant emissions

include ignition and valve timing, fuel types and additives combusted, lubricants employed in the

engine, and exhaust gas treatments employed (Abdel-Rahman, 1998). These air pollutant

emissions can be broadly characterized as gaseous and non-gaseous emissions.

Gaseous Emissions

The major gaseous pollutants emitted include oxides of nitrogen (NOx,) and carbon monoxide

(CO). NOx refers to mixtures of nitric oxide (NO) and nitrogen dioxide (NO2). Small amounts of

sulfur dioxide (SO2), also a criteria pollutant regulated under the CAAA, are emitted. SO2

emissions from ICEs are solely a result of fuel-bound sulfur and are readily reduced by limiting

sulfur in the fuel.

Table 5: Relative levels of ICE air pollutant emissions

Pollutant ppm g/kg fuel
NO, 500-1000 20
CO 1000-2000 200
HC (as C1) 3000 25

Non-Gaseous Emissions

The primary non-gaseous pollutants are unburned or partially burned hydrocarbons (HCs).

Relative levels of both gaseous and non-gaseous emissions from ICEs are shown in Table 5.

These are only ranges, since actual emissions vary greatly on engine design, fuels combusted, and

combustion and post-combustion controls involved. HCs are divided into reactive and

nonreactive categories, based upon their role in photochemical smog formation. The simplest

such breakdown is categorizing HCs as methane and nonmethane hydrocarbons (NMHCs), since

all HCs except for methane (CH4) will react given sufficient time.

NOx Formation

NOx is referred to here as mixtures of nitric oxide (NO) and nitrogen dioxide (NO2). NOx

emissions are controlled because NO and NO2 contribute to the formation chemistry of low-level

ozone, or smog, an environmental and human health hazard. NO2 is also directly of concern as a

human lung irritant.

NOx can also be defined to include other oxides of nitrogen, including N20, NO3, N204, and

N205. These additional nitrogen oxide species are insignificant in the emissions from ICEs, and

readily react to NO and NO2. NO generally accounts for over 90 percent of the total NOx

emissions from fossil fuel combustion, with the remainder being NO2. The formation of NO can

be explained by three different mechanisms (Turns, 1996):

1. The Extended Zeldovich mechanism, or thermal NO, in which O, OH, and N2 are in

equilibrium concentrations

2. Other mechanisms with NO formation rates above that predicted by the Extended

Zeldovich mechanism, including

a. Fenimore CN and HCN path\ a\ s

b. NzO-intermediate route

c. "Super-equilibrium" concentrations of O and OH in combination with the Extended

Zeldovich mechanism

3. Fuel nitrogen mechanism, in which fuel-bound nitrogen is oxidized to NO.

The primary pathway for NO formation is oxidation of atmospheric molecular nitrogen (N2)

through the thermal or Zeldovich mechanism:

O+N,2 NO+N (17)

N+O2 = NO + O (18)

Extending the thermal NO formation mechanism to include the hydroxyl radical reaction with

nitrogen was proposed by Lavoie et al (1970):

N+OH > NO + H (19)

Thermal NO formation rate is slow relative to combustion and is considered unimportant below

1800 K. Thermal NO formation attributed to Equations (17) through (19) is considered formed

in the post-combustion exhaust gases.

Prompt NO, also referred to as the Fenimore mechanism, is NO that is quickly formed in the

premixed laminar flame before thermal NO has formed. Hydrocarbon radicals react with

molecular nitrogen to create hydrogen cyanide as an intermediate to NO formation in the

following steps:





NCO + H NH + CO (22)

NH + H N +H2 (23)

N + OH NO + H (24)

Prompt NO formation is also considered insignificant in internal combustion engines due to the

thin flame fronts, short residence times, and high pressures in the combustion chamber.

The formation of NO through an NzO intermediate mechanism is important in fuel-lean (0 <

0.8), lower temperature conditions (T < 1800 K). The three steps are:

O+N +M N2O+M (25)

H+ NO NO + NH (26)

O+N20 O 2NO (27)

The M in Equation (25) represents a third body collision molecule. The significance of the NzO

intermediate mechanism can be seen in Equation (27) where two moles of NO are formed per

mole of N20. While our subject engine operates at 0 = 0.8, the flame temperature is slightly

higher than 1800 K, so NzO intermediate pathway formation of NO is probably not significant.

Fuel-bound nitrogen is another source of combustion NO emissions. This process is

significant in coal combustion, where bituminous coal contains up to 2% by mass bound nitrogen.

Nitrogen in the fuel is quickly reacted to HCN or ammonia, NH3, and follows the reaction steps

beginning with Equation (21) for prompt NO formation. Kerosene and gasoline fuels contain

trace to zero quantities of nitrogen, so fuel-bound nitrogen contribution to NO formation is not

considered significant in internal combustion engines.

The final reaction mechanism considered here for NO, formation is production of NOz.

Reactions contributing to the formation and destruction of NOz are as follows:

NO + HO,2 NO2 + OH (formation) (28)

NO2 + H NO + OH (destruction) (29)

NO2 + O NO + O2 (destruction) (30)

The HO2 radicals form in low-temperature regions, leading to NO2 formation through Equation

(28). NO2 destruction via reaction with the H and O radicals are active at high temperatures

[Equations (29) and (30)]. Thus NO formation from NO2 would be preferred at high

temperatures and NO2 would only survive during low-temperature cooling of exhaust gases. This

validates the previous statement that most of the NOx emitted from internal combustion engines is


Heywood (1988) reports the following relationship for NO formation rate based upon

empirical data and the assumption of equilibrium concentrations of O, Oz, OH, H, and N2, by

decoupling the NO formation from combustion (i.e., assuming NO formation in post-combustion

gases always dominates NO produced in the flame):

d[NO] 610 e6 (31)
dt e [IN e (31)

where [ ]e denotes equilibrium concentrations. The significant dependence of temperature on NO

formation rate in Equation (31) is evident. Equation (31) will be used later in Chapter 5 to

calculate the relative NO formation rate from estimated flame temperature changes for

comparison with measured percent NO reductions.

CO Formation

Carbon monoxide (CO) emissions from ICEs are a concern from toxicological effects on

humans.. The formation of CO from the combustion of the hydrocarbon radical, R, is as follows:

RH -- R RO2 RCHO RCO CO (32)

Once formed, CO is slow to oxidize to COz, with water providing the primary oxidant source

through the following steps:

CO + O CO2 +0 (33)

O+HO -20H (34)

CO + OH CO2 + H (35)

H+O2 OH+O (36)

The reaction in Equation (33) is slow, with primary oxidation of CO occurring through Equation

(35), with Equation (36) producing OH radicals feeding back to Equation (35). Diatomic

hydrogen (H2) can also provide oxidation of CO through formation of HO2, but H2 is not formed

in sufficient quantities in ICE combustion to contribute to CO oxidation.

Control Technologies

Technologies to control air pollutant emissions from ICEs can be categorized as process or

post-combustion controls. Process controls include changes and improvements to the combustion

chamber, fuel/air delivery, and engine components aimed at reducing air pollutant emissions. For

example, CFD analysis and design has been used to improve and air intake valve ports to enhance

fuel-air mixing. Post-combustion controls include catalytic converters and other technologies

applied to react with combustion exhaust constituents including NOx, CO, HCs, and particulate.

Combustion Modifications

Exhaust gas recirculation (EGR) is commonly employed to reburn combustion by-products,

especially CO and particulate. EGR also dilutes the intake air oxygen concentration, increasing

the heat capacity of the combustion products per unit of heat release, lowering the combustion

flame temperature (Turns, 1996). Larsen et al. (1996) reported 50% NOx reduction at 20% EGR,

but CO emissions doubled and fuel economy decreased 8%. Exhaust gas recirculated on diesel

engines must be well-filtered to prevent fuel sulfur and exhaust particulate from eroding and

corroding engine intake valves, cylinders, and pistons.

Adjusting fuel injection timing is an effective method for decreasing NOx emissions.

Traditional practice is to delay fuel injection into the combustion chamber to lower the final

flame temperature, but this generally results in higher unburned hydrocarbon emissions.

Yanagihara (1997) demonstrated reduced NOx production by shortening the injection duration

while advancing the fuel injection. His results are attributed to improved fuel-air mixing prior

to combustion, which both improves combustion efficiency and reduces unburned hydrocarbons.

Selective and Non-Selective Catalytic Reduction

Selective catalytic reduction (SCR) involves the use of a catalyst generally requiring an

additive, such as ammonia, to initiate NOx reduction chemistry. A common application of SCR

in internal combustion engines is the Pt, Rh, and Pd three-way catalytic converters used on spark

ignition engines combusting gasoline. Unburned hydrocarbons act as the selective reducing agent

for the catalysts. Non-selective catalytic reduction (NSCR) also involves the use of a catalyst but

without the need for an additive to reduce NOx. The application of SCR or NSCR catalysts in a

diesel exhaust is severely complicated (1) primarily by the higher excess oxygen content of diesel

exhaust, resulting in a net oxidizing environment, and (2) by the presence of sulfur in diesel fuel

and resulting catalyst poisoning by sulfur dioxide in the exhaust gases. Significant research is

ongoing using secondary injection of small quantities of fuel in the exhaust stream to act as the

reducing agent, with demonstrated NOx reductions of approximately 45 percent at reasonable

space velocities and high fuel metering rates (Nakatsuji et al., 1998).

Water Injection

Four major approaches for introducing water into the combustion zone have been reported in

the literature:

1. Fumigating the water into the engine intake air

2. Direct injection into the engine through separate injectors

3. In-line mixing of water and fuel prior to injection (unstabilized emulsion)

4. Mixtures of stabilized emulsions treatable as a single-phase drop-in replacement fuel

Urbach et al. (1997) demonstrates water mist injection into the bell housing of diesel-fueled

turbine engines with promising results. Water mist introduced to the intake of reciprocating

compression ignition ICEs, particularly two-stroke engines with the intake air passing through the

crankcase, poses significant corrosion potential. Separate water injecting valves in the engine

avoids intake system contact with the mist (Yoshihara et al., 1996). Several authors have

evaluated all or most of the options for introducing water into the combustion process and have

primarily determined that water-in-fuel emulsions, stabilized or unstabilized, are most effective in

reducing NOx, BSFC, and result in lower increases of CO and UHC emissions (O'Neal et al.,

1981; Greeves et al., 1976).

An emulsion is defined as a mixture of two or more generally insoluble liquids. A permanent

emulsion exists when sufficient droplet sizes have been reached to prevent the separation of the

insoluble materials. Unstabilized emulsions are generated through the high-speed, high-shearing

of particles and solids in a liquid. A limiting concern with emulsions is the high capital costs of

emulsification mixers and pumps, which are used extensively in the food and agriculture


Unstabilized emulsions require high shear to suspend small droplets of water in the fuel

(Greeves et al., 1976). An advantage of unstabilized emulsions are reduced fuel costs, due to lack

of additives needed, and reduced emissions from not combusting surfactants or other emulsifying

agents (De Vita, 1989). Ulrich and Kessler (1992) propose a complex fueling system including a

vortex chamber to provide in-line mixing of water and diesel fuel without requiring the addition

of an emulsifying agent. Diesel fuel pumps, including the unit on our Detroit Diesel 4-71 engine,

operate at high volume and high pressure, with a recirculation loop back to the fuel tank. This

serves several purposes:

1. A high-volume pump can create the high-pressure needed for the fuel injectors at less


2. Recirculating warms the fuel and helps resist gelling at low temperatures

3. The warmed fuel improves combustion

All of these factors contribute to the effectiveness of an in-line fuel-water emulsifying system,

assuming that retrofitting the fuel system is acceptable.

Stabilized emulsions use an emulsifying agent to suspend the water in the fuel and reduce the

energy required for a permanent emulsion. The Air Force preferred a drop-in replacement fuel,

without the requirement for modifying the engines or fueling system. The current research thus

uses a surfactant to create a permanent, stabilized emulsion that can be treated as a single-phase


A drawback to water-fuel emulsions is the amount of air bubbles reportedly contained in the

emulsion mixture. Sawa and Kajitani (1992) evaluated the effect of water-fuel emulsions on

diesel engine performance and emissions under transient conditions. They conclude that air

bubbles in the fuel and its variability contributes to fluctuation in the injection timing and a

poorer performance under transient conditions. They recommend removal of bubbles from the

water-fuel emulsion.

Research has also been conducted extensively on the use of additives to improve the

lubrication, reduce the corrosive effect of water in the fuel, and improve the emulsion stability.

Nitrate-containing ignition improvers are recommended to reduce exhaust emissions (Schwab,

1997). Lubricity additives composed of dimer or trimer acides, phosphate esters, sulfurized

castor oils are recommended by Peter-Hoblyn et al. (1998), and catalysts can also be used in situ

in the fuel to reduce NOx (Peter-Hoblyn et al., 1996). In addition to providing lubrication

improvement, additives to water-fuel emulsions can be employed for antifreeze characteristics

(Marelli, 1995), obviously important when significant volume percent of water are present in a

fueling system in freezing climates. Montagne et al. (1987) demonstrated that surfactants added

to diesel fuels can both clean up fuel injectors and prevent further injector deposits. They also

reported a slight increase in NOx from the combustion of the surfactants, as we will experience

without also adding water to the fuel.

Crookes et al. (1990) attributes water-fuel emulsions with improved combustion and lower

particulate and NOx emissions to the secondary atomization of the water, often designated as

microexplosions. Yoshimoto et al. (1989) extensively examined the microexplosions of

emulsified fuels and determined that there is a minimum percent water content in the emulsion

required for microexplosions to occur, and that the percent increases with the kinematic viscosity

of the fuel.

Table 6 provides a summary of water injection and water-in-fuel emulsion research in the

literature related to the present study. Note that Montagne et al. (1987) reported a 5 percent NOx

increase when adding surfactants only to diesel fuel for cleaning fuel injectors. This would lead

us to eliminate the possibility that surfactants may solely contribute to the reduction of NOx in our

study. Crookes et al. (1990) were comparing diesel and vegetable oil fuels, both dry and as an

emulsion with 10% water, by volume. Their results were included to demonstrate that small

ratios of water provide negligible reductions in NOx emissions. Small quantities of water are

effective in improving BSFC, which could have merit for fuel savings. O'Neal et al. (1981)

tested macro- and micro-emulsions to determine if water droplet size has a significant effect on

NOx reduction, BSFC, and other combustion parameters. They define micro-emulsion as

Table 6: Summary of diesel engine NOx reduction using water and other additives

Reference Method Water, vol % ANOx, -%
Afify, 1985 Stabilized 45 83.3
Afify et al., 1987 Stabilized 40 65
Andrews et al., 1989 Stabilized 25 70
Crookes et al., 1990 Unstabilized 10 4
Fujita et al., 1987 Stabilized 30 65
Greeves et al., 1976 Unstabilized 80 60
Hsu, 1986 Stabilized 30 22
Montagne et al., 1987 Surfactant only -5
Murayama et al., 1978 Stabilized 50 45
O'Neal et al., 1981 Macro-emulsion 20 25
Micro-emulsion 20 23
Sawa and Kajitani, 1992 Stabilized 40 64
Valdmanis and Wulfhorst, 1970 Stabilized 50 72
Vichnievsky, 1975 Stabilized 55 59

emulsions with water droplet sizes smaller than the wavelength of visible light, approximately

555 nm, whereas macro-emulsions are characterized with water droplet sizes larger than the

wavelength of visible light. Thus, micro-emulsions are reported to visually appear clear, while

macro-emulsions appear cloudy. Our fuel mixture would be characterized in this manner as a

macro-emulsion. As shown in Table 6, the reported NOx reductions were 25 and 23 percent for

macro- and micro-emulsions, respectively. They reported significantly-higher (unspecified)

ratios of surfactant were required to establish a micro-emulsion, although they reported longer

stability lifetime for the micro-emulsion. From our experience and the reports of O'Neal et al.

(1981) we feel justified in using the minimum surfactant required to establish a stable macro-


The data from Table 6 are plotted in Figure 10, except for that of Montagne et al. (1987). A



.2 60 *

-7 40


0 I I I I I
0 20 40 60 80 100
Water content, vol. %

Figure 10: NOx reduction as a function of fuel-water emulsion
water content reported in the literature

trend is visible for reduced NOx emissions with increased water content in the fuel. As shown in

Table 6, this data represents 11 independent research programs conducted across a span of 22


Other Control Technologies

Canfield et al. (1997) described a filter cart designed for capture of NOx, CO, VOCs, and

particulate from the A/M32A-86 diesel generator. The device is a series of sub-systems,

including a vermiculite filter to capture particulate, air-to-air heat exchanger and demister for

cooling and dewatering, and granular activated carbon (GAC) filters to adsorb NOx, CO, and

VOCs. The filter cart is shown on the right of Figure 11, attached to an A/M32A-86. This device

requires a large footprint.

After the GAC filters are saturated and adsorption rates begin to decline, the filters are

thermally regenerated. Adsorbed gases are desorbed, and can either be compressed, bottled, and

reused, or destroyed on-line via selective catalytic reduction. The vermiculite particulate filters

are replaced and discarded after excessive increase in pressure drop. The filter cart requires a

Figure 11: NOx filter cart for diesel exhaust capture

large footprint and would not be feasible to mobilize to a war-fighting theatre. Advantages

include that one filter cart can service multiple generators, depending on the capacity of the filter

cart, and it can be used to control emissions from other combustion sources.

Nonthermal plasmas have also been applied to diesel exhaust, and specifically applied to

reduce NO, emissions from the A/M32A-86 (Ackridge et al., 1997; Rolader et al., 1997; Federle

et al., 1998). This application also required the use of a series of subsystems, including a ceramic

particulate filter, nonthermal plasma discharge (NTPD) reactor tube with alcohol injection, and a

wet gas scrubber. The particulate filter captures particulate and would be cleaned in-line using

the hot exhaust gases. The NTPD essentially uses high-voltage, low amperage, high-pulse rate

electrical discharges to generate reactive, oxidative species in the exhaust gases. The addition of

alcohols is reported to increase the reaction efficiency, lowering electron volts required to oxidize

NO to NO2. The wet scrubber is then used to adsorb and react the NO2 with water to form nitric

acid, HNO3, and then with sodium hydroxide, NaOH, to form sodium nitrate, NaNO3, useful as a

fertilizer. This system would also require a large footprint.



A/M32A-86 Parameters

The diesel engine used in the study is a Detroit Diesel 4-71N 2-stroke diesel engine with

forced induction. Engine data are presented in Table 7 and a schematic of the generator is shown

in Figure 12. The manufacturer engine designation corresponds to a 4-cylinder engine with 71 in3

(1163.5 cm3) per cylinder.

The A/M32A-86 generator was originally supplied with the N65 fuel injector, having an

orifice diameter of 1.651 mm (0.065 in) and a recommended injection timing setting of 3.708 cm

Table 7: A/M32A-86 parameters (from SM-ALC/TISEA, 1986)

Engine Model Detroit Diesel 4-71N
Combustion chamber type Direct injection
Number of cylinders 4
Displacement, cm3 4,653.9 (284 in3)
Bore, mm 107.95 (4.25 in)
Stroke, mm 127 (5 in)
Compression ratio 18.71 : 1
Rated power at 2100 rpm, kW 110.4
Power factor 0.8
Maximum electric power output, kW 72
Air box pressure, kPa abs. 34.1 (10.1 in Hg)
Air inlet restriction, kPa abs. 2.86 (11.5 in H20)
Exhaust back-pressure, kPa gauge 13.5 (4.0 in Hg)
Compression pressure, kPa gauge 3,900 (565 psi)






Figure 12: A/M32A-86 diesel generator (from SM-ALC/TISEA, 1986)

(1.460 in). This corresponds to the height of the piston from BDC when the fuel is injected. A

larger dimension for fuel injection delay setting corresponds to a longer time delay before fuel

injection. The dimensions are The N65 injector historically has a problem with plugging and

fouling at low load levels. Thus, scheduled maintenance includes running the engine with

generator at full load (72 kW) once a month to "clean out" the injectors. This results in

significant NOx and hydrocarbon emissions in a short period of time. Experimental results will

be presented for both the N65 and N90 injectors. The N90 injector, with an orifice size of 2.286

mm (0.090 in.) was tested to increase the fuel flow rate delivered to the engine with the diesel-

water emulsion.

Fuel Properties

Military specification (MIL-SPEC) diesel fuel was primarily used in the test, with properties

shown in Table 8. We also conducted tests using JP-8, MIL-SPEC jet fuel. JP-8 is used in

mobility applications of the generators. The Air Force has since standardized to the use of JP-8 in

diesel engines to reduce the need to manage and maintain two fuel types and fueling systems.

However, military specification diesel was in widespread use during the period of this test, and

was therefore the focus of this study. Future project plans are discussed in Chapter 5 that include

long-term testing of a fuel-water emulsion using JP-8.

The diesel-water fuel composition used in these tests is listed in Table 9. Experimental

Table 8: Common values for military diesel (from Avallone and
Bauemeister, 1996, and U.S. DoD, 1995)

Property MIL-F-16884J
Density @ 150C, kg/m3 876
API gravity, deg 40
Total sulfur, percent 0.5
Boiling point, C 357
Endpoint, C 385
Flash point, C 60
Pour point, C -6
Hydrogen, wt % 12.5

Cetane number:
Acid number, mg KOH/100 ml
Kinematic viscosity, cSt, @ 37.80C:
Specific gravity
High heating value, QHHV,, kJ/kg
Low heating value, QLHv kJ/kg



results shown later indicating baseline exhaust temperature, [NOx], and [CO] correspond to the

standard fuel (diesel or JP-8) without water or additives. A corrosion inhibitor was selected to

help offset potential corrosive effects of the water in the fuel during idle storage in the fuel tank.

Lebedev and Nosov (1980) reported reduced wear on the upper area of the cylinder and piston

(near TDC) in a Caterpillar 3304 engine with fuel-water emulsions, but also reported increased

wear in the lower regions of the cylinder liner. They attribute the increased wear to sulfur in the

diesel fuel blends. MIL-SPEC diesel and JP-8 traditionally contain little to no fuel-bound sulfur.

For this study we were concerned with the perception of water in fuel as "corrosive," particularly

to the fueling system, and included the corrosion inhibitor as a precaution. As discussed in

Chapter 3, the emulsion improver allows a permanent emulsion to be stabilized more readily.

Both additives are expected to reduce injector plugging, engine wear, and corrosion (Herbstman

and Virk, 1989, Estefan and Brown, 1990, and Liu et al., 1993). Post-testing materials

evaluations of the effect of the fuel-water emulsion on engine components would be valuable in

determining and quantifying the necessity for a corrosion inhibitor. Eliminating the corrosion

inhibitor would obviously lower the cost of the alternative fuel mixture. Intensive exhaust

characterization using GC/MS would also be important to demonstrate that components of the

corrosion inhibitor are not emitted from the engine. The corrosion inhibitor contains xylene and

ethylbenzene, both targeted as potential carcinogens and targeted for extremely strict controls on

their release from diesel and gasoline engines.

Table 9: Diesel-water emulsion properties

Fuel component Percent, by volume
Water 30-45
Corrosion inhibitor 0.6
Emulsion improver (surfactant) 0.4
MIL-SPEC diesel 69-54

Sampling Procedure

NOx emissions were measured using an Energy Efficiency Systems, Inc. ENERAC 3000

multigas analyzer with electrochemical cells measuring NO, NOz, CO, Oz, SOz, and percent HCs.

The ENERAC 3000 meets the requirements and specifications of the U.S EPA's conditional

reference method (U.S. EPA, 1995). The specifications and parameters for the ENERAC 3000

are listed in Table 10. The NO, NO2, and CO sensors relied upon for this study are listed with an

accuracy of 2.0%. As shown in Table 10, the ENERAC 3000 has three ranges for NO, CO, and

SO2. The unit is able to maintain a 1 ppm resolution across these broad ranges by including

distinct electrochemical cells dedicated to the respective ranges. Systems that use a dilution

factor will lose either accuracy or resolution when scaling to higher ranges. The calibration

procedure for the ENERAC 3000 follows the EPA conditional test method and was performed

prior to testing and data collection. The ENERAC 3000 includes a heated sampling probe with

desiccant to absorb moisture in the exhaust to prevent damaging the electrochemical sensors.

An Omega thermocouple and digital thermometer were used to measure exhaust gas

temperatures, traceable to National Institute of Standards and Testing (NIST) calibration

standards. A fast-response thermocouple was used. A Fisher Instruments thermometer-

hygrometer was used to read ambient relative humidity and dew point.

The A/M32A-86 is governed to 2100 rpm when under load. Load power settings listed are

applied by a load bank, shown in Figure 13, used for engine routine maintenance. The load bank

is an array of resistors that can be configured to mimic the varying loads of electric current drawn

by an aircraft connected to the generator. Resistors are applied in three phases in 0-3 (variable)-,

3-, 6.5-, and 13-amp increments, with power measured on each phase with an analog meter.

Total power is the sum of all three phases.

Water-fuel emulsion mixtures were hand-prepared in the laboratory in batch. An industrial

high-speed clarifier was used to blend the water, fuel, and additives in desired concentrations. As

mentioned in the Preface, a patent application is under review on the additive package, preventing

further disclosure of the constituents. The goal of the additive package, in maintaining a stable

Table 10: ENERAC 3000 specifications (from Energy Efficiency Systems, Inc., 1995)

parameter Range Resolution Uncertainty, %
02 0-25.0% 0.1 % 0.2
NO 0-300 ppm 1 ppm 2
0-1000 ppm 1 ppm 2
0-3500 ppm 1 ppm 2
NO2 0-500 ppm 1 ppm 2
CO 0-500 ppm 1 ppm 2
0-2000 ppm 1 ppm 2
0-20000 ppm 1 ppm 2
SO2 0-500 ppm 1 ppm 2
0-2000 ppm 1 ppm 2
0-7000 ppm 1 ppm 2
HC (as CH4) 0-6.00% 0.01% 10

Figure 13: Load bank for adjusting power load on generator

mixture, is to provide practical application of the water-in-fuel emulsion. Our results showed

similar NOx reductions to unstabilized emulsions, as also reported in the literature (Coon,

1981;Valdmanis and Wulforst, 1970; Sawa and Shuichi, 1992).

Data Collection

Results reported in the following figures and discussion are the averaged values of three

sample points per indicated value. The NOx, CO, and exhaust temperature readings were taken

several minutes apart, tabulated, and averaged. The data are reported in X-Y scatter as functions

of power load setting, in kW, on the generator.

Sampling pipe

Total PM filter

Dilution tunnel
l 'i total PNI

PN lilici

- A/M32A-86
diesel generator

Dilution tunnel

' Method sampling
train controls

combustion gas analyzer

Manometer for flow rate

Figure 14: Exhaust gas sampling setup

The sampling system used in recent tests is shown in Figure 14. The diesel generator is

visible in the rear behind the scaffolding. On top of the scaffolding is the dilution tunnel system

used for capturing total particulate, and the "doghouse" structure houses pumps for filters to

capture particulate fractions less than 2.5 tim. In front of the doghouse is a manometer to

measure the pressure-drop of a static-pressure pitot tube, to calculate exhaust flow rate. A Z-

shaped exhaust pipe is supported inside the scaffolding, from which all the exhaust samples are

measured. To the right of the doghouse is a box housing a pump system for an EPA Method 5

sampling train to measure water content of the exhaust. The ENERAC 3000 used to measure

...... ...


NOx and CO is slightly visible to the immediate right of the doghouse. Attached to the left of the

scaffolding is a helium tank for a portable GC/MS (Viking Instruments) used to characterize

VOCs and SVOCs (not shown in the picture).

The ENERAC 3000 and Omega thermocouple were inserted through holes in the Z-shaped

exhaust pipe immediately after the diesel generator exhaust pipe. The dilution tunnel sampling

system requires a straight vertical run with 8 diameters of pipe length after the bend, which in

turn required the construction of the Z-shaped exhaust extension and scaffolding support. The

PM25 filters feed off the flow through the dilution tunnel, which captures total particulate.

Finally, the pitot tube for calculating exhaust gas flow rate is inserted in the upper horizontal run

after the dilution tunnel.



Experimental Results

Experimental results are given in Figure 15 through Figure 21, where data are graphed for the

NOx and CO concentrations and exhaust temperature, Te, as functions of power load on the

generator. Power load is graphed, with the exception of Figure 19, from idle to 80 kW. Recall

full electric load on the generator is 72 kW with diesel or JP-8 fuel. The idle condition is given

an arbitrary value left of zero. Latter graphs are interpretations of the experimental data used to

demonstrate the correlation between exhaust gas temperature, power setting, injector size, fuel

water content, and exhaust gas composition.

Baseline Emissions

Figure 15 shows the NOx, CO, and exhaust temperature for the diesel engine combusting

diesel fuel, with the N65 and N90 injectors at various injection timing delay settings. Significant

reductions in NOx emissions are shown simply delaying the fuel injection for both injectors, with

the N90 injector displaying significantly lower NOx emissions. In fact, the Air Force is

considering replacing the N65 stock injector with newer swirl injectors, based upon these and

similar emission testing results elsewhere.

Comparing the NOx and CO graphs in Figure 15, we see the customary tradeoff of higher CO

emissions when NOx is reduced. Note the N90 injector at 3.861 cm displays both the lowest NOx

and highest CO emissions. This is due primarily to the effect of fuel quenching causing

incomplete combustion and increased HC emissions, including CO. These results are consistent

with trends described in the literature (Borman and Brown, 1992).

N65 Injectors

Figure 16 presents results with the N65 injector at 3.769 cm using the diesel-water

emulsion. Note in Figure 16 the significant reductions in NOx emissions using 30 and 43% water.

Data for 30 and 43% water are not shown out to 72 kW because the generator could not develop

that power output with the corresponding levels of water in the fuel. Exhaust temperature is a

factor of 1.088 lower at 50 kW for the 30% diesel-water blend versus the standard diesel. This

may be significant because the NOx emissions were 1.75 times lower for the same range. From

this, we can hypothesize that the relationship between NOx and flame temperature is either

nonlinear or chemical kinetic effects are responsible for the greater reduction in NOx. Recalling

the equilibrium NO formation rate, Equation (31), we would expect a nonlinear relationship

between NO and temperature. Substituting T1 = 2000 K and T2 = = 1838 K into Equation

d[NO], d[NO]2
(31), one obtains L = 2.01 d or that the rate of NO production should double with
dt dt

only a 1.088 factor increase in flame temperature, assuming exhaust gas composition is in

equilibrium. Our data deviates from this estimate, since the slope d[ ] of the baseline NOx

curve is only 1.53 times the slope of the 30% water curve. A likely source of the discrepancy is

our use of exhaust temperature, versus in-cylinder flame temperature measurements. The excess

water may continue to cool down the exhaust gases, giving a greater temperature difference

between the baseline and 30% water than the NOx levels would suggest. In other words, the

flame temperature change is likely to be less than the exhaust temperature change.

[NOx] decreasing with
increasing timing delay

-*-N65 at 3.769 cm -E-N65
-A-N65 at 3.861 cm ---N90
G--N90 at 3.861 cm

at 3.830 cm
at 3.830 cm

0 10 20 30 40 50 60 70 80
0 10 20 30 40 50 60 70 80


Figure 15: Baseline exhaust properties












800 --









2500 [NOx] =28.085 P + 502.62


1500 [NO] = 18.336 P + 135.14


[NOx] = 2.3704 P + 33.078



-4- Baseline

400 -Ea- 30% Water
-- 43% Water

300I 0 1
Idle 0 10 20 30 40 50 60 70 80


Figure 16: Exhaust properties with N65 injectors at 3.769 cm












500 -4- Baseline
500 30% Water
400 -A-- 40% Water
45% Water
Idle 0 10 20 30 40 50 60 70 80

Power, kW

Figure 17: Exhaust properties with N90 injectors at 3.830 cm

S 300

0 1500







300 I
Idle 0

-4- Baseline
-E- 40% Water
-A- 45% Water

10 20 30 40 50 60 70 80

P, kW

Figure 18: Exhaust properties with N90 injectors at 3.861 cm

-- Baseline (JP-8)

-E- JP-8/water/methanol

0 10 20 30 40 50 60 70 80


Figure 19: JP-8/water/methanol with N65 injectors at 3.769 cm

2,500 -

2,000 -

1,500 -

1,000 -

500 -




800 +

600 -

400 -

200 -
















1000 +






Te, K

Figure 20: Baseline [NOx] versus exhaust temperature for N65 and N90 injectors
at various fuel injection delay settings

-- *-- N65 at 3.769 cm
A N65 at 3.769 cm (JP-8)
- E-- N65 at 3.830 cm
-A-N65 at 3.861 cm
-*-N90 at 3.830 cm
S--N90 at 3.861 cm




1,600 [NOJ = 2.60 Te + 187



1,000 -- 2
[NOJ =-0.0391 Te -44.78 Te-11486
-B- JP-8/water/methanol
400 I
400 450 500 550 600
T,, K

Figure 21: [NOx] versus exhaust temperature for JP-8
and 65% JP-8 / 30% water / 5 %methanol (by volume) mixture

N90 Injectors

Figures 17 and 18 present data for the N90 injectors at 3.830 and 3.861 cm, respectively. The

N90 injector was not tested at 3.769 cm because tests with the N65 injector at delayed timing

settings resulted in significant NOx reductions. Figure 17 shows baseline NOx emissions of 659

ppm at 60 kW are reduced to 272 ppm with 45% water, a reduction factor of 2.42. Note that the

generator would produce slightly less than 40 kW with the N65 injectors and 43% water (Figure

16), demonstrating the significance of increasing the injector size.

Comparing the exhaust temperatures at 60 kW, the baseline exhaust temperature of 683 K

reduced to 625 K, a factor of 1.09. This is very comparable to the results using the N65 injector

(Figure 16), except we were able to attain 1.5 times the power output (60 kW versus 40 kW) with

1.05 times more water in the fuel (43 to 45%) using the N90 injector.

Figure 18 shows similar reductions in NOx and exhaust temperature, although intermediate

data were not taken at 30% water, only 40 and 45%. The exhaust temperature for the 45% water

case is 1.028 times higher than the 40% water case, although lower NOx, was measured. The

significantly-lower CO for 40% water suggests that a variability between fuels, other than water

content, may have occurred between the 40 and 45% mixtures reported. For example, a poorly-

mixed 45% emulsion, in which the water ratio decreased with time, would account for the slightly

higher exhaust temperature.

JP-8/Water/Methanol Tests

Data shown in Figure 19 was taken using a blend of 64% JP-8, 30% water, 5% methanol, and

1% additives, by volume. A new testing protocol was followed, in compliance with the

International Organization for Standardization (ISO) testing procedures for compression ignition

2-stroke engines. The test cycle called for testing at load settings of 10, 25, 50, 75, and 100% full

load. Based upon experience and preliminary testing, we established 56 kW as full load and then

tested at the prescribed percentages. Please note that the injector and timing setting for data

reported in Figure 19, N65 at 3.769 cm, corresponds to the diesel/water mixture injector and

timing settings reported in Figure 16.

The NOx reduction of 78% at 50 kW is substantially higher than the 44% reduction reported

in previous tests (Figure 19). The CO emissions at low loads are also significantly lower in

Figure 19 versus Figure 16. For example, CO emissions at 15 kW were 623 ppm for the JP-

8/water/methanol mixture in Figure 19, versus 2826 ppm for the diesel/water mixture in Figure

16, a factor of 4.53. Possible explanations for this high variation in CO emissions include engine

variabilities, differences in JP-8 versus diesel combustion, testing condition variables, and CO

monitoring equipment calibration. The most likely cause, demonstrated in recent studies, is the

variability of NOx and CO emissions across these diesel generator units. Data in Figure 16 were

taken in late 1996 using a diesel generator no longer in service, whereas data in Figure 19 were

recorded in late 1998 from a diesel generator fresh out of six-month maintenance. Several other

engines tested in late 1998 also displayed high CO emissions and slightly higher NOx emissions

and were suspected to have poor cylinder compression. Testing conditions also varied: data in

Figure 19 were recorded in mild Florida winter ambient temperature of 289 K, whereas data in

Figure 16 were recorded in humid Florida summer with ambient temperatures of 308 K. Higher

intake temperatures affect the volumetric efficiency and ignition delay of the compression

ignition engine cycle.

NOx reductions at lower loads are slightly higher. Exhaust temperatures for this fuel and

engine test were 4.25% lower with the JP-8-water fuel. Measurable discrepancies exist across

engines and future test programs will require the use of multiple engines.

Shown in Figure 20 are NOx emissions versus temperature for the baseline case of all

injector/timing settings tested. We must be careful interpreting results like those in Figure 20. As

shown earlier, both NOx and exhaust temperature increase with load setting on the generator. We

have demonstrated through practice that thermal NOx formation is significant in the subject diesel

engine, but further correlation is required to prove the hypothesis that diesel-water emulsions

lower the combustion flame temperature and suppress thermal NOx formation.


Error bars are plotted in Figure 19 to demonstrate the low level of instrument error in our

tests. The [CO] and [NOx] data accuracy is 2 percent of indicated value, plotted in Figure 19 as

positive and negative errors. The Omega thermocouples used to measure exhaust temperature, Te,

are accurate to 1.0 K. These errors are also plotted in Figure 19.

NOx Versus Exhaust Temperature

Combustion products and exhaust temperature to this point have been presented as a function

of power load on the generator. Baseline NO, emissions are plotted against the corresponding

exhaust temperature in Figure 20. A strong trend exists for increased NO, with increased exhaust

temperature as fuel injector size increases and timing is delayed. Note the higher values and

slope for N65 injectors versus N90 injectors at timing setting of 3.830 cm. Please also note the

overlap for N65 injectors at 3.769 cm, with both diesel and JP-8 fuels, demonstrating the

similarity between the fuels. The line labeled "JP-8" is repeated in Figure 21.

Figure 21 presents [NOx] as a function of exhaust temperature for the recent investigations

using JP-8 and the JP-8/methanol/water mixture with 30% water and 5% methanol, by volume.

Trend line "JP-8" is repeated from Figure 20, along with data for the mixture, to compare how the

NOx formation as a function of temperature varies with the addition of water (and in this case

methanol). Figures comparing each of the baseline trends in Figure 20 with its corresponding

NOx versus exhaust temperature trend have been not been included for the sake of brevity. Error

bars are plotted in Figure 21 depicting the positive and negative standard error for each data

series. The standard error for the JP-8 data set is lower due to smaller differences between data

points. A clear trend is shown for increased NOx emissions with increased exhaust temperature,

both for the baseline case of JP-8 fuel and the JP-8/water/methanol mixture.

The linear curve-fit equations were plotted in Figure 21. It is interesting to note that while

the y-intercept for the mixture is significantly lower, at -2280 ppm, than that for JP-8, 187 ppm,

the slope for the mixture is 2.56 times that of the JP-8. The slope of the line can be interpreted as

an empirical NO, formation rate based upon exhaust temperature, with units of ppm/K.

First Law of Thermodynamics Calculation

The stoichiometry for fuel-lean combustion (0 < 1) was described by Equation (5) in Chapter


CxHy +a(02 +3.76N2) -> bCO + dHO2(g)+ fO2 +3.76aN2 (5)

We will apply Equation (5) first to determine the adiabatic flame temperature of the baseline case

for diesel fuel combustion:

Case 1: Diesel combustion

Given: Diesel fuel

Equivalence ratio, 0 = 0.8

Temperature of reactants, TR = 298 K

Pressure of reactants, R = 100 kPa

Find: Temperature of products (adiabatic flame temperature), Tp

Assumptions: Steady-state steady-flow process

Pressure of products, R = 100 kPa

Model diesel fuel as C144H249 (Sonntag et al.,1998)


Please refer to first-law control volume around the combustion engine in Figure 7. Modeling

diesel fuel as C144H249 and solving for the coefficients of Equation (5) per Equations (7) through

(12) yields the following:

x+ 14.4+ 24.9
a= 4 4 -25.8 (37)

b = x = 14.4


v 24.9
d = 2- = 12.45 (39)
2 2

f = (1 -)a = ( 0.8)(25.8) = 5.16 (40)

Inserting these coefficients into Equation (5) yields:

C144H249( )+25.8(02 +3.76N, )(
14.4C2 + 12.45H20(g)+ 5.1602 + 96.9N2

Using Equation (13), with Qcv = 0 and Wcv = 0, gives the following equality for the enthalpies of

reactants and products:

nh, = n h, (42)

The total reactant enthalpies can be evaluated according to the sum of the respective species'

enthalpies of formation, h;, and the species' enthalpies, Ah at the reactant temperature and


HR = n,(h + Ah), (43)

Evaluating Equation (43) for reactant molar balances in Equation (41), with enthalpies evaluated

at TR = 298K, R = 100 kPa, and enthalpies of formation from Sonntag et al. (1998), Table A.8,


HR=- h' + 25.8h 0 +96.9h2
HR =hC144H249 2 ,o2 9 fN2 (44)
= -174,000 + 0 + 0 = -174,000 kJ/kmol

Please note that the respective Ah, from Equation (43) are zero because we chose the reactant

temperature and pressure equal to the reference state for enthalpy of formation.

Similarly, the equation for product enthalpy, substituting enthalpies of formation from

Sonntag et al. (1998), Table A.8, is

Hp = n (h + Ah) = 14.4(- 393,522 + Ah2)
P (45)
+ 12.45(- 241,826 + AH2O (g)+5.16Ao2 + 96.9AhN

The flame temperature is now calculated using an iterative procedure to equate HR and Hp using

the product enthalpies Ahco AhH2o, A AhN 2. These results are listed in Table 11 for

product temperature from 1600 K to 2400 K.

Table 11: Adiabatic flame temperature iteration for diesel combustion

Tp, K





Hp, kJ/kmol

HR, kJ/kmol

Note that the product enthalpy, Hp, passes the value of reactant enthalpy HR between Tp

and 2200 K. Interpolating for the product temperature between 2000 and 2200 K yields

]Tp =2102 K.



Calculation of the adiabatic flame temperature for the diesel-water combustion case follows

the same basic procedure, except the addition of water as a reactant must be included in the

combustion equation, demonstrated below:

Case 2: Diesel-water combustion

70% diesel, 30% water (by volume) fuel

Equivalence ratio, 0 = 0.8

Temperature of reactants, TR = 298 K


Pressure of reactants, R = 100 kPa

Find: Temperature of products (adiabatic flame temperature), Tp

Assumptions: Steady-state steady-flow process

Pressure of products, R = 100 kPa

Model diesel fuel as C144H249 (Sonntag et al.,1998)

All water reactant goes to water product


Determine the molar ratio of diesel and water for a 70/30 volumetric ratio, on a 1 m3 basis, using

density and molecular weight of water and diesel at reactant conditions, with diesel again

modeled as C14 4H249:

0.7m3 C144249 kg 1 kmol
Ndesel l m3 m3 198.06 kg 1 kmol C144H249
x x (47)
Nwater 0.3m3 H20 997kg 1 kmol 5.36 kmol H20
lm3 m3 18.015 kg

Thus, for every 1 kmol of diesel there are 5.36 kmol of water in our fuel blend. Adding liquid

water as a reactant to Equation (41) and balancing yields:

C144H249(0)+5.36HO( )+25.8(02 +3.76N2) )
14.4CO2 +17.8H20(g)+5 .1602 +96.9N2

Summing the reactant enthalpies:

+ +5.36 +25.08h;0 + 96.9h;'
HR = hf,C14 4H24 9 +5.36h 'H2(0) + 25.08hfo +96.9hN2 (49)
= -174,000 + 5.36(-285,830) + 0 + 0 = -1,706,049 kJ/kmol

Summing the product enthalpies:

Hp = nj (h + h), =14.4(- 393,522 + A/ )
(p (50)
+17.8(-241,826+ A o(g) )+5.16A/o2 + 96.9A/2

The iterative results are listed in Table 12 for product temperature from 1600 K to 2200 K.

Table 12: Adiabatic flame temperature iteration for diesel-water combustion

TpK hc2 A H2 A2 AhN2 Hp, kJ/kmol HR, kJ/kmol
1600 67569 52907 44267 41904 -3767566 -1706049
1800 79432 62693 51674 48979 -2698760 -1706049
1900 85420 67706 55414 52549 -2158070 -1706049
2000 91439 72788 59176 56137 -1613848 -1706049
2200 103562 83153 66770 63362 -515492 -1706049

The product enthalpy, Hp, passes the value of reactant enthalpy HR between Tp = 1900 and 2000

K. Interpolating for the product temperature yields

Tp =1983K (51)

Comparing the results from Cases 1 and 2, the adiabatic flame temperature dropped 119 K,

from 2102 K to 1983 K, a factor of 1.060 or 5.7 %, when including the enthalpy required to

convert liquid water in the fuel to water vapor. Before discussing these results from Cases 1 and

2 we will first examine the results provided by a computer code to determine the equilibrium

products of combustion.

Equilibrium Code Calculation

The diesel and diesel-water combustion cases were also evaluated using the computer

program HPFLAME provided in Turns (1996), based upon the equilibrium products of

combustion code by Olikara and Borman (1975). HPFLAME calculates the adiabatic flame

temperature, equilibrium products of combustion, and mole fractions of the primary product

species for adiabatic constant-pressure combustion. As described in Chapter 2, the diesel

0.9 -- OFirst Law
0 -
CO CO2 H20 N2 NO 02 OH

Figure 22: Diesel combustion product mole fractions X for
first law and HPFLAME calculations

combustion process can be approximated as adiabatic constant-pressure combustion. Turns

(1996) also provides programs UVFLAME, for adiabatic constant-volume combustion (gasoline

spark-ignition), and TPEQUIL for calculations when the combustion temperature is known.

Results for the analysis of the diesel fuel combustion, with the diesel fuel again modeled as

C144H249, are presented in Table 13. The program requires the input of carbon, hydrogen,

oxygen, and nitrogen atoms in the fuel, as well as the equivalence ratio, initial combustion

temperature guess, pressure, and enthalpy of reactants. These inputs correspond to the previous

evaluation of the reactant enthalpies in the Case 1.

Please note the additional inclusion of H, O, N, H2, OH, CO, and NO beyond the first law

evaluation in Case 1. OH, CO, NO, O2, H2O, COz, and N2 account for over 99.9 % of all

products, so H, O, N, and H2 products will be neglected. HPFLAME estimated the adiabatic

flame temperature at Tp = 2073 K, which is 29 K or a ratio of only 1.014 less than the first law

result of 2102 K. The major mole fractions for the first-law analysis and HPFLAME are shown

in Figure 22. Equilibrium mole fractions of for the first law calculation are simply the fraction of

individual product moles to total product moles from Equation (41) in Case 1. For example,

summing product moles from Equation (41) yields 14.4 + 12.45 + 5.16 + 96.9 = 128.9 kmol, and

mole fraction of N2 is then = 0.752 kmol/kmol. Figure 22 demonstrates there is a good

correlation between the first-law analysis and the equilibrium products of combustion code for

fuel-lean combustion diesel fuel combustion.

Shown in Table 14 are results for HPFLAME calculation of adiabatic flame temperature and

equilibrium products of combustion for the case of diesel-water combustion, as described in

Case 2. Note the inputs correspond to the enthalpy and reactant conditions from Case 2. Please

also note that HPFLAME also allows input of O and N fuel molecules. As described earlier in

Chapter 3, liquid fuels such as diesel, kerosene and gasoline contain little to no fuel-bound

nitrogen. Nitrogen molecules in the fuel might be included if calculating the flame temperature

for a coal-fired boiler. The inclusion of O molecules in the fuel would be treated by HPFLAME

as an oxygenated fuel, so including additional H and O molecules for H2O() in the fuel would

not be modeled properly. Turns (1996) included the source code for HPFLAME, in which the

constituents of air are specified (79% N2 and 21% 02). An interesting project would be to modify

the routines in HPFLAME to include kinetic pathway modeling for high concentrations of

H20() in the fuel.

For the diesel-water combustion case, HPFLAME estimated the adiabatic flame

temperature at Tp = 1799 K, which is 184 K, or a ratio of 1.10 less than the first law result of

1983 K determined in Case 1. This variation is significantly greater than the 1.014 reduction ratio

for the diesel fuel combustion calculations, further demonstrating the variation between the two

methods for calculating diesel-water mixture combustion characteristics.

Table 13: HPFLAME results for diesel combustion

Inputs (Reactants)
Carbon atoms 14.4
Hydrogen atoms 24.9
Oxygen atoms 0.0
Nitrogen atoms 0.0
Equivalence ratio, 0 0.800
Flame temperature guess, K 2000.0
Pressure, kPa 101.325
Enthalpy of reactants, kJ/kmol fuel -174000.0
Results (Products)
Adiabatic flame temperature, K 2073.61
Mixture enthalpy, kJ/kg -46.54
Mixture specific heat, cp, J/kg-K 1419.97
Specific heat ratio, cj/cy 1.2232
Mixture molecular weight, kg/kmol 28.9537
Moles of fuel per mole of products 0.00774456
Mole Fractions of Product Species
H 0.00004084 NO 0.00410114
O 0.00022365 02 0.03810987
N 0.00000000 H20 0.09524323
H2 0.00024079 CO2 0.11018881
OH 0.00183075 N2 0.74868802
CO 0.00133290

Comparing results in Tables 13 and 14, HPFLAME calculates the adiabatic flame

temperature to decrease from 2073 K to 1799 K, a factor of 1.15, or 13.2 %. This result is 1.08

times greater than the 1.060 factor flame temperature decrease calculated with the first law

analysis from Cases 1 and 2. In other words, the equilibrium products calculated by HPFLAME

predicts a lower flame temperature and greater percent flame temperature reduction than the first

law analysis. The reason for this discrepancy can be seen by comparing the stoichiometric

combustion products listed in Case 2, Equation (50). By estimating that all of the water in the

fuel as a reactant goes to steam as a product, we were able to include the enthalpy of formation of

Table 14: HPFLAME results for diesel-water combustion

Inputs (Reactants)
Carbon atoms 14.4
Hydrogen atoms 24.9
Oxygen atoms 0.0
Nitrogen atoms 0.0
Equivalence ratio, 0 0.800
Flame temperature guess, K 2000.0
Pressure, kPa 101.325
Enthalpy of reactants, kJ/kmol fuel -1706049.0
Results (Products)
Adiabatic flame temperature, K 1799.31
Mixture enthalpy, kJ/kg -456.3
Mixture specific heat, cp, J/kg-K 1419.97
Specific heat ratio, cj/cy 1.2540
Mixture molecular weight, kg/kmol 28.9883
Moles of fuel per mole of products 0.00775383
Mole Fractions of Product Species
H 0.00000181 NO 0.00186311
O 0.00002369 02 0.03899801
N 0.00000000 H20 0.09628909
H2 0.00002602 CO2 0.11154139
OH 0.00043824 N2 0.75070493
CO 0.00011370

the H20(g) on the RHS of the first-law balance. This accounts for the enthalpy required to

convert the water to steam and slightly lowers enthalpy available for combustion. However,

HPFLAME is treats the reduced reactant enthalpy, from -174,000 KJ/kg to -1,706,049 KJ/kg as a

very low-heating value fuel, without estimating water in the products.

One benefit of applying HPFLAME to this study included the calculation of equilibrium

concentrations of NO. Comparing equilibrium mole fractions for NO, designated XNO, from

Tables 13 and 14, XNO decreased from 0.00410 to 0.00186 kmol/kmol. Thus, HPFLAME predicts

a 2.20 factor decrease in equilibrium NO production with the reduced reactant enthalpy. This

suggests that lower heating-heating value fuels (or diluted high-heating value fuels) lower the

flame temperature and reduce total NO production.

Figure 23 depicts the diesel-water combustion product molar fractions for the first law and

HFPLAME calculations. Data for HPFLAME were taken from Table 14 and first law product

mole fractions were calculated from Equation (48). The greater molar fraction for N2 is likely the

primary source of discrepancy between the two calculations. Recall that 5.36 kmoles of H2O(O)

was added to the reactant in Equation (48), whereas in HPFLAME the sole input to account for

the reactant composition, other than the fuel, was a reduced reactant enthalpy.

NOx Formation Rate Calculation

Please recall the NO formation rate relationship, Equation (31):

I ----------------|

0.9 I First law
0.8 13 HPFLAME
0 ,
CO CO2 H20 N2 NO 02 OH

Figure 23: Diesel-water combustion product mole fractions X for first
law and HPFLAME calculations

d O1016 -69,090
d -6x106 e [N 0 (31)

It is more convenient to work with the mole fraction, X,. The molar concentration [X,] is related

to the mole fraction as

[X,]-= (52)

where R is the universal gas constant, 8.3145 kJ/kmol-K, and P and T are the adiabatic flame

pressure and temperature. Substituting Equation (52), with P = 100,000 Pa, into Equation (31)


dXN 6.58x1018 x e 6990
dt T e XN2e (53)

Listed in Table 15 are the formation rates for diesel and diesel-water combustion using the

NO formation rate relationships from Equation (53). Results are based upon adiabatic flame

temperature and mole concentrations from the first law analysis, HPFLAME calculations, and

correlations to exhaust temperature variations. First law product temperatures Tp are taken from

Cases 1 and 2, and HPFLAME Tp were listed in Tables 13 and 14. Equilibrium product mole

fractions ZO2,e and ZN2,e are taken from the data graphed in Figures 22 and 23. Please note that

these results depict NO mole fraction production rate, which differs from the equilibrium NO

mole fraction reported earlier for the HPFLAME calculation. Results depicted in Table 15 are

Table 15: NO production rate calculations

dXNO kmol ArdZo
Fuel Model Tp, K Z2,e XN2,e dt 'kmol- s dt
Diesel First law 2102 0.0400 0.752 2.50
Diesel-water First law 1983 0.0384 0.722 1.10 2.27
Diesel HPFLAME 2073 0.0381 0.749 1.56
Diesel-water HPFLAME 1799 0.0390 0.751 0.011 142
Diesel Exhaust 2102 0.0400 0.752 2.50
Diesel-water Exhaust 1932 0.0384 0.722 0.151 16.6

based upon flame temperature and equilibrium.

In summary, Equation (53) predicts NO mole fraction formation rate reduces 2.27, 142, and

16.6 times for diesel-water fuels, based upon data from the first law analysis, HPFLAME, and

exhaust temperature calculations, respectively. The HPFLAME NO mole production rate is

effectively zero for the diesel-water mixture. The greater reduction based upon HPFLAME data

is a result of the calculated 274 K flame temperature reduction, versus 119 K and 170 K for the

first law and exhaust temperature analyses, respectively. Additionally, the HPFLAME NO mole

fraction formation rate for diesel combustion (1.56 kmol/kmol-s) is initially 1.6 times less than

the first-law NO mole fraction formation rate (2.50 kmol/kmol-s). This is a result of the lower

diesel combustion flame temperature and equilibrium 02 and N2 mole fractions for HPFLAME.

Correlating these results to our experimental data could be conducted by either correlating

flame temperature to the exhaust temperature, or using relative exhaust temperature reductions, to

evaluate NO production rate reductions via Equation (53). Taking the latter approach, we refer

again to Figure 16 to compare the baseline diesel exhaust temperature to the diesel-water

mixture exhaust temperature. At 50 kW, the exhaust temperature decreases from 620 K to 570 K,

a -8.06% or 1.088 factor decrease. Applying this ratio to the adiabatic flame temperature

calculated in Case 1 for diesel combustion, 2102 K, yields an estimated flame temperature for the

diesel-water mixture of 1931 K. This result is also tabulated in Table 15. Please note in Table

15 that the first law and HPFLAME flame temperature reduction factors were 1.060 and 1.152,

respectively. Again, our first law analysis appears to be more accurate, correlating to our

experimental factor of 1.088 very closely. Equilibrium product mole fractions for oxygen and

nitrogen, O 2,e and N 2,e are of secondary effect to temperature in Equation (53). Thus 0O2,e and

XN2,e from the first law calculation were used for the exhaust temperature calculation. As shown

in Table 15, using the experimentally observed exhaust temperature reduction, correlating it to an

equivalent flame temperature reduction, and applying the empirical relationship in Equation (53)

yields a 16.6 factor decrease in NO mole fraction production rate.

Effects of Water Injection

Thermodynamic, physical, and chemical effects of water injection on the combustion process

discussed here and in the literature can be summarized as follows:

1. Microexplosions accelerate the diffusion of combustion through the cylinder, decreasing

the time required for combustion and increasing combustion efficiency

2. Water in the fuel decreases the heat content of the fuel, decreasing the energy output per

mass of total fuel.

3. The partial pressure of water may accelerate the water-gas reaction (Sawa and Kajitani,


The effect of microexplosions of liquid water present in fuel is considered in the literature to

be the primary mechanism for increasing combustion efficiency (Tsenev, 1983). Hsu (1986)

conclusively measured slight ignition delay resulting from water added to diesel fuel, and

attributed improved combustion to the delayed ignition improving the evaporation and mixing of

the fuel.

The high heat of vaporization of water in the fuel has been demonstrated here to lower the

adiabatic flame temperature. For individual engines and fuels there is a practical upper limit for

percent by volume water in the fuel, after which combustion would be sufficiently slowed to

significantly reduce the combustion efficiency. De Vita (1989) recommended an upper limit of

20 mass percent water in diesel, which corresponds to 45.5 volume percent water, to prevent

increasing the brake specific fuel consumption (BSFC) of the engine. If slight increases in BSFC

can be accepted, higher water ratios could be used.

Consideration is given in the literature to the extent that water added to the fuel modifies the

chemistry of combustion, possibly through increasing OH, H, and O radical formation and


oxidation chemistry. Combustion chemistry is undoubtedly affected by the contribution of water

to more-complete combustion, thus reducing the amount of unburned hydrocarbons and reducing

eventual soot formation. Greeves et al. (1976) and others theorize that water vapor is generally

present during combustion and that kinetic effects are distantly secondary to thermodynamic

effects in reducing NOx and soot formation. Following this line of reasoning can lead us to

conclude that the theoretical results calculated from a first law of thermodynamics combustion

balance and extended equilibrium products of combustion (HPFLAME) are representative of

actual results, and that chemical effects can be neglected for global approximations.




The summary in Table 15 shows a lower flame temperature and resulting lower production

rate of NO is estimated by theory and was demonstrated by our experimental data. Our results

confirm the reports in the literature by many other researchers (Table 6), that water addition to

fuel lowers the flame temperature and suppresses thermal NOx formation. using a four-stroke

diesel-powered generator in common use by the U.S. Air Force,

A First Law of Thermodynamics calculation estimated a 1.06 factor decrease in adiabatic

flame temperature from the baseline case of diesel only and a diesel-water mixture of 30 percent

water, by volume. Equation (53) predicts a resultant 2.27 factor decrease in NO formation rate

for the diesel-water mixture. An equilibrium products of combustion code corroborated the

trend demonstrated by the first law calculation, with a 1.15 factor decrease in adiabatic flame

temperature and a corresponding 142 factor decrease in NO formation rate. Experimental data

shown in Figure 16 demonstrates this trend to decreased temperature, with exhaust temperature

decreasing a factor of 1.09 at 50 kW with 30 percent water in the fuel.

Figure 16 and Figure 19 demonstrated similar results for diesel-water and JP-8-water

mixtures, and Figures 17 and 18 demonstrated lower exhaust temperatures and NOx emissions

when delaying the fuel injection to the cylinder.


Our primary conclusion, which confirms the findings of researchers listed in Table 6 and the

References, is that water added to fuel lowers the flame temperature and suppresses the formation

of thermal NOx in internal combustion engines. Our calculations estimated factor decreases of

flame temperature of 1.06 and 1.15, with a corresponding NO formation rate decreases of 2.27

and 142, when adding water to fuel.

Our results also lead us to conclude that NOx reduction from fuel injection timing delay and

improved injector design, as demonstrated in Figure 15 and Figure 20, is also probably due to a

reduced flame temperature. Figure 20 shows a trend to lower exhaust temperature and lower NOx

emissions with increased timing delay, which corresponds to the trend for water added to the fuel.

In general, we can conclude from the reported data that fuel-water mixtures are an effective

option to reducing NOx emissions from diesel engines without requiring modifications to the

engine, if a lower full load is acceptable. By installing larger fuel injectors, the diesel engine can

attain the original load level, as shown by Figure 17.

Recommendations for Further Study

Planned research includes further testing and refinement of the fuel emulsification additives.

As shown by Montagne et al. (1987), adding surfactants to fuel can increase emissions of NOx.

To minimize both increased emissions and fuel costs due to surfactants, one should use the

minimum level of surfactant necessary to stabilize the fuel-water mixture. A factorial analysis

should be conducted varying fuel, water, surfactant, and other additive concentrations to

determine first- and second-order effects.

The mixing time required to develop a stable emulsion is best determined empirically.

Again, a factorial analysis varying the water and fuel ratios versus mixing power and time should

be conducted to establish a matrix of mixing times required to guarantee a stable emulsion.

The need for a corrosion inhibitor should be quantified through visual and wear-metal testing

on engine components after extended use of fuel-water blends both with and without a corrosion

inhibitor. The added cost of the corrosion inhibitor should be firmly established and

demonstrated as a necessity to protect the engine and components, not simply a precaution to

subdue suspicions of corrosion potential. If a corrosion inhibitor is demonstrated as necessary,

GC/MS analysis of exhaust emissions across all load ranges should be conducted to determine if

corrosion inhibitor components are emitted from the engine. Worker safety concerns require

monitoring to demonstrate that the xylene and ethylbenzene components common in corrosion

inhibitors are fully combusted.

Ongoing investigations indicate difficulties in cold-start with the fuel mixtures containing

water. Several factors likely contribute to this problem, including the lower heating value of the

fuel when combined with water. When cooled, the water in the fuel raises the auto-ignition

temperature of the fuel mixture. Initial tests indicate pre-warming the fuel is an effective option

to enhance auto-ignition of the diesel-water emulsions. We recommend a test program to

evaluate the temperature relationship, model the thermodynamic effects, and design and install a

fuel tank heating unit on a diesel engine to improve cold-start.



b = cylinder bore (diameter), mm

c, = specific heat at constant pressure, J/kg-K

cv = specific heat at constant volume, J/kg-K

h = enthalpy, kJ/kg

he = convective heat transfer coefficient, W/m-K

Ah = change in enthalpy from reference state, kJ/kmol

h; = enthalpy of formation, kJ/kmol

k = thermal conductivity, W/m-K

kN,f = forward rate coefficient for reaction N, m3/kmol-s

kN,r = reverse rate coefficient for reaction N, m3/kmol-s

M = molecular weight, kg/kmol

rif = mass flow rate, kg/s

P = power, kW

p = pressure, kPa

QLHVV = lower heating value at constant volume, kJ/kg

QLHV = higher heating value at constant pressure, kJ/kg

q" = heat flux, W/m2










Greek Symbols

e = emissivity, 0 e 1

p = density, kg/m3

r = Stefan-Boltzman constant, 5.67 x 10-8 W/m2K4

v = specific volume, m3/kg

[X,] = molar concentration, ppm

X, = mole fraction

dX, = mole fraction production rate, kmol-s /kmol

conductive heat flux, W/m2

convective heat flux, W/m2

radiative heat flux, W/m2

universal gas constant, 8.3145 kJ/kmol-K

compression ratio, m3/m3

specific entropy, kJ/kg-K

temperature, K

exhaust temperature, K

work done on control volume, kJ


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Charles Alan Canfield was born at Chanute AFB, Illinois, and raised in Independence,

Missouri. While attending the University of Missouri-Columbia, he was active in the Engineer's

Club, Pi Kappa Alpha fraternity, and held an engineering internship with the 3M Corporation.

Mr. Canfield received his Bachelor of Science in mechanical engineering in December 1991.

Mr. Canfield has been employed by Applied Research Associates, Inc. at Tyndall AFB,

Florida, since 1993, supporting research programs for the Air Force Research Laboratory Airbase

& Environmental Technologies Division. He is presently conducting catalytic studies with an

advanced plug-flow annular reactor system. Currently the President of the Gulf Coast Chapter of

the Florida Engineering Society, a state society of the National Society of Professional Engineers,

he is a registered Professional Engineer in Florida. He is also a member of the Board of Advisors

of the Gulf Coast Community College Civil Engineering Technologies Department, the Board of

Directors of Girls Inc., and active in the Bay County Chamber of Commerce.

Upon completion of the Master of Science degree, Mr. Canfield will pursue studies in

coordination with his in-house catalytic reaction engineering research. The research reported

herein will be continued if sufficient funding and time is appropriated. Additional courses in

chemistry and kinetic modeling will be pursued to enhance the understanding of the combustion


Mr. Canfield's past-times include road cycling and distance running, beach volleyball,

working on his house and yard, keeping his old truck running, and collecting engineering

references and textbooks.